Potential of diesel electric system for fuel saving in fishing vessels: a case study on a bottom longline fleet of Brazil This study concerns a model of the fuel consumption and emissions from the bottom longline fishing fleet at Rio de Janeiro State, Brazil with data comprised vessel and engine characteristics, fuel consumption, landings and fishing trip duration of seventeen longliners operating in Rio de Janeiro between 2015 and 2016. The operational pattern was investigated for this fleet, based on an empirical approach to model the propulsion power requirements of the typical longliner fishing vessel. The current propulsion and power plants of the longliner fishing vessels under consideration are compared with alternative designs by estimating fuel consumption for the different models of propulsion systems. When the operational pattern evaluation is taken into account, it indicates the future potential for diesel-electric propulsion for longliner fishing vessels, as a result of a reduction in fuel consumption. Optimal stowage on Ro-Ro decks for efficiency and safety The roll-on/roll-off (Ro-Ro) ships are true workhorses of coastal and deep-sea shipping. They are valued for their versatility to transport heterogeneous cargo and short turnaround times in ports. However, the optimal utilisation of cargo space has been inherently problematic with the Ro-Ro concept. In view of the existing attempts to contrive optimal stowage plans, the paper proposes three practical improvements with respect to the state of the art. The improvements lead to a finer approach to ship stability, fire safety, and cargo handling efficiency when optimising cargo stowage on Ro-Ro decks. Formally, we express the stowage problem as a mixed-integer linear programming (MILP) problem and solve it to optimality. The paper outlines the mathematical formulation, provides a numerical example, and studies practical application aspects. Robust backstepping ship autopilot design A ship is always under the influence of matched as well mismatched uncertainties and external disturbances owing to its operational environment and, therefore, designing an autopilot that offers uniform performance in the presence of uncertainties and disturbances poses a challenging task. In this work, to address the issue, control design based on adaptive backstepping robustified by Uncertainty and Disturbance Estimator (UDE) is proposed. The design does not need knowledge of bounds of uncertainties and disturbances and is able to effectively deal with the uncertainties and disturbances. Simulation results show that the performance of the backstepping autopilot is enhanced by combining the backstepping design with the UDE approach. Application of quaternary ammonium compound and aliphatic amine as biocides in the mitigation of microfouling adhered to the internal surface of heat exchangers condensers refrigerated by seawater. Impact on the efficiency of different control strategies applied The present article compares the effectiveness of two chemical compounds to reduce the growth of biofilm adhering to the tubes of a heat exchanger-condenser refrigerated by seawater. The compounds tested in the study were: a quaternary ammonium compound (QAC) and an aliphatic amine (triethanol amina, TA). Each biocide was applied intermittently in a first shock stage and in a second stabilising stage. Prior to biocide dosing, a biofilm was created in the inner surface of the heat exchanger tubes under study (AISI 316L and 316Ti stainless steel). The evolution of the biofilm during growing and treatment phases was followed by indirect measurements, i.e. fluid frictional resistance (f) and heat transfer resistance (Rf ). The article also presents results of the impact of the different control strategies applied to the heat exchanger condenser efficiency. Both compounds removed the biofilm, and the tubes were practically restored to their clean condition. Ecotoxicology studies classified both biocides as environmentally harmless under the testing conditions. A numerical study on flow field and maneuvering derivatives of KVLCC2 model at drift condition This paper investigates the flow field around the KRISO Very Large Crude Carrier (KVLCC2) model at different drift angle using the RANS method in combination with the Re-Normalization Group (RNG) k−εk−ε and Shear Stress Transport (SST) k−ωk−ω turbulence models. The hydrodynamic forces and yaw moment acting on the KVLCC2 model are calculated under the different turbulence models and drift angles, and the numerical results are compared with previous experimental results. The accuracy of the numerical results is found to be acceptable, especially under the SST k−ωk−ω turbulence model. Three different mesh conditions are employed to perform the verification and validation based on the methodology and procedures suggested by ITTC. The surface pressure distribution, vorticity distribution, and surface flow features around the hull are observed and analysed. The maneuvering derivatives are calculated based on numerical results to create the mathematic model of ship maneuvering motions in the marine simulator and predict the ship maneuvering motions. A fuzzy-based occupational health and safety risk assessment framework and a case study in an international port authority Ports are complex systems that play a crucial role in the transportation of maritime-related products and services. Since ports face various safety issues in design, installation, operation and maintenance processes, there are many aspects open for improvement in Occupational Health and Safety risk assessment and it is required to deal with risks, reduce them to an acceptable level and control. To struggle with emerging risks and manage this process better, a risk assessment framework is proposed for ports in this study. A case study is performed in a Turkish international port authority using the proposed framework. Fuzzy Analytic Hierarchy Process (FAHP), which is a commonly used fuzzy multi-criteria decision-making method, was used in weighting two risk factors: likelihood and severity. The orders of priority of hazards were then determined by a fuzzy-based VIKOR method based on the output of FAHP. The proposed framework overcomes some drawbacks, including a lack of assignment on the weight to two risk assessment factors and lack of the high dependence of the prioritisation on the experts' opinions. To validate the results, a sensitivity analysis was conducted. In conclusion, the study contributes to the applicability of the proposed framework to assess the risks in ports and provides a distinct risk prioritisation. Microstructual and mechanical studies of the dissimilar tabular joints of Incoloy alloy 825 and AISI 316 stainless steel In this work, dissimilar welding of Incoloy alloy 825 and AISI 316 stainless steel tubes is experimentally investigated using gas metal arc welding process. For this purpose, different filler metals such as stainless steel 308, stainless steel 309, stainless steel 316 and Inconel 82 are employed for joining of Incoloy alloy 825 and AISI 316 stainless steel tubes. The results show that during tensile tests, fracture occurred in all weldments in the heat-affected zone (HAZ) of stainless steel 316. Also, it is concluded that dissimilar welds produced by stainless steel 309 filler metal have the maximum tensile strength while the lowest strength is obtained with Inconel 82 filler metal. The results of average hardness prove that the dissimilar joint with filler metal say stainless steel 309 and filler metal say Inconel 82 have maximum and minimum values of average hardness in the weld metal, respectively. The results show a considerable unmixed zone (UMZ) in the form of laminar layers between filler 82 weld metal and the Incoloy 825 base metal along the welding line. In addition, considerable UMZ is seen in the interface of the weld metal/AISI 316 base metal, but more grain growth occurred in the AISI 316 HAZ due to the temperature increase during various welding passes. It is indicated that there is a very narrow UMZ between the 309 weld metal and Incoloy 825 base metal; however, no considerable UMZs can be observed in the AISI 316 base metal. The use of acoustic emission elastic waves as diagnosis method for insulated-gate bipolar transistor The article describes research into the use of acoustic emission elastic waves (AE) for diagnosing insulated-gated bipolar transistors (IGBTs). Currently used in many modern shipboard power electronic systems, IGBTs are crucial components required to operate with extreme reliability. The acoustic emission elastic waves method was used to determine the acoustic signals of switching transistor and can be used to monitor early stages of damage. Transistor diagnostics makes use of selected frequency descriptors of AE and appropriate signal analysis. The effects of different cross sections on the hydrodynamic behaviour of sandglass-type FPSOs exposed to regular waves This paper presents a numerical investigation on the hydrodynamic performance of sandglass-type FPSOs with four different cross sections. In order to estimate the hydrodynamic performance and utilise the results in the design stage of FPSOs, a frequency-domain numerical simulation program, ANSYS/AQWA software package, has been used. Numerical results were compared with experimental data and good agreement has been achieved in small amplitude regular wave cases. Based on the simulation results, it is concluded that polyhedral cross sections (especially 10-sided cross section) provide similar hydrodynamic performance compared with circular cross section in heave and pitch motions for all ranges of wave frequency. Therefore, it is possible to use these types of cross sections for FPSOs because of manufacturing simplicity. The effects of different cross sections on the hydrodynamic behaviour of sandglass-type FPSOs exposed to regular waves This paper presents a numerical investigation on the hydrodynamic performance of sandglass-type FPSOs with four different cross sections. In order to estimate the hydrodynamic performance and utilise the results in the design stage of FPSOs, a frequency-domain numerical simulation program, ANSYS/AQWA software package, has been used. Numerical results were compared with experimental data and good agreement has been achieved in small amplitude regular wave cases. Based on the simulation results, it is concluded that polyhedral cross sections (especially 10-sided cross section) provide similar hydrodynamic performance compared with circular cross section in heave and pitch motions for all ranges of wave frequency. Therefore, it is possible to use these types of cross sections for FPSOs because of manufacturing simplicity. Temperature characteristic and compensation algorithm for a marine high accuracy piezoresistive pressure sensor Pressure sensor used in marine hydrographic survey is an important device for the deep sea detection, the tsunami forecast and the marine engineering. Owing to its mature technology process and low production cost, silicon piezoresistive sensor is widely applied in the field of industrial pressure measurement. However, its temperature effect can result in non-negligible dynamic error, which cannot meet requirement of deep sea environment. The silicon-on-sapphire pressure sensor based on stress cup structure is presented. The output voltage values show a high deviation in the range of −6∼50°C due to temperature drift. In order to minimise this temperature effect, a temperature compensation algorithm is proposed to realise 0.03% accuracy in the full scale 60 MPa range. Compared with a reference commercial sensor in deep sea experiment, the test pressure sensors exhibit minor error, excellent similarity and coherence. This numerical method provides a new research direction for environmental self-adaptive sensor. It can be emphasised that this sensor will have a good application prospect in unattended long-term ocean observation system. A novel ship energy efficiency model considering random environmental parameters Energy efficiency management is becoming increasingly important in the trend towards decarbonisation and intelligentisation of future ships. Establishing a verified energy efficiency model is essential in realising reliable assessment of various energy efficiency strategies. Based on a 53,000-tonne bulk carrier, modelling and verification of a ship's energy efficiency with consideration of multiple factors is carried out. First, the existing ship's energy efficiency regulation and its evaluation methods are introduced. Second, the onboard data collection system is introduced with the features of the measured data detailed. A ship energy efficiency model is developed from four main aspects, namely ship energy efficiency operational indicator, ship fuel consumption, ship main engine power, and ship resistance characteristics. Based on the Monte Carlo simulation method and utilising Matlab/Simulink, the energy efficiency model for the selected ship is simulated and measured fuel consumption data is used to verify the model. Finally, the simulation results are presented and discussed. The research results show that the devised model provides good enough accuracy to simulate ship energy efficiency with consideration to cargo loading, ship speed and the random impact of multiple natural environmental parameters. This study not only helps the ship manager assess the projected energy efficiency, but can also provide decision support for the optimisation of ship energy efficiency. Velocity tracking controller for simulation analysis of underwater vehicle model A method for dynamics analysis of underwater vehicles is presented in this paper. The strategy consists of two stages. At the first stage, we propose an adaptive velocity tracking controller for the vehicle. The control algorithm is expressed in terms of transformed equations of motion in which the inertia matrix is diagonal. Consequently, it takes into account the dynamics of the system and it can be applied for fully actuated systems in the presence of disturbances and uncertain dynamics. Using this controller, we are able to track the velocity, and also at the same time, to evaluate some dynamic properties of the vehicle. At the second stage, we give some procedure for the dynamics analysis. The advantage of the approach is that the algorithm serves simultaneously for control purposes and dynamics investigation. It can be done before real experiment which is always expensive and time consuming. The effectiveness of the strategy is validated via simulation on a full 6-DOFs underwater vehicle model. Slurry erosive wear and microhardness characteristics with coarse silica sand of dual reinforced particles ADC12 alloy composites The current research was a comparative study of reinforced ceramic particles and mixing proportion on the microstructural and slurry erosive characteristics for fabricated composites. Dual reinforced particle composites were fabricated by adding combinations of numerous wt% of titanium carbide (TiC) and zircon sand (ZrSiO4) to aluminium die-casting (ADC12) alloy matrix composites. ADC12 alloy composite with titanium carbide and zircon sand samples were also prepared for comparison of dual reinforcement particles on microstructural features and wear behaviour in various slurry concentrations. Titanium carbide (44 - 60 um) and zircon sand (20 - 33 um) particles are reinforced in the alloy by the two-step stir-casting method. Slurry erosion wear study reveals that the dual particle reinforcement enhances the wear resistance as compared to single-particle reinforcement if mixed in a definite proportion. The study also indicates that a combination of 15% reinforcement of titanium carbide and zircon sand particles in the ratio of 10 and 5 wt% into the composite exhibits better microhardness and wear resistance as compared to other combination. An improved control-limit-based principal component analysis method for condition monitoring of marine turbine generators The safe operation of marine turbine generators is a crucial concern in industries and academics. It is always important to monitor the health status of marine turbine generators. The lubricant oil usually carries abundant information on the turbine operation conditions. Various oil parameters of the turbines have been used in the existing monitoring systems. However, many of them conflict with each other by contrary detection results. Hence, it should eliminate the redundant oil parameters for efficient condition monitoring. Although many research studies addressed the redundant feature reduction issue using principal component analysis (PCA), PCA is designed for features with a linear relationship, which is not the case in marine turbine generator monitoring. This paper proposes a new nonlinear analysis method, the improved control-limit based PCA, to extract distinct failure indicators from the oil parameters of marine turbine generators. The contribution of this method is that the Hotelling statistic and Q statistic are combined to calculate a fixed control limit for PCA. The ability of the improved PCA to dealing with nonlinearity has been significantly enhanced by the proposed method. Experimental validation demonstrates that the extracted failure indicator using the proposed method is more effective than existing monitoring indexes with respect to fault detection accuracy. Spectral fatigue analyses comparison study: Suez Canal vs. Cape of Good Hope Arab Academy for Science, Technologies and Maritime Transport (AASTMT) Suez Canal is one of the shortest navigational routes between east and west. Vessels that transit the Canal should normally save distance, time and operating cost. During the current economic recession and low fuel prices, longer shipping routes may be chosen to avoid the Canal transit fees. Harsh weather conditions may be encountered along these lengthy routes resulting in higher stresses and structural fatigue damage on ship's hull, which would impact structural safety and affects the cost of ship maintenance. This paper quantifies ship structural fatigue damage along the routes: Suez Canal transit versus the Cape of Good Hope. Several voyages of an Aframax tanker along these routes are selected and evaluated for fatigue damage using a route-specific spectral fatigue damage assessment approach. Fatigue accumulation during the tanker lifetime, for each trade route, is computed as the sum of all the encountered sea states having caused fatigue damage along that route. Results are presented as fatigue damage and fatigue life for each trade route. Results show that in addition to saving in distance and time, less accumulated fatigue is achieved along the Suez Canal route. An innovative ship salvage concept and its effect on the hull structural response In the European research project `Surfacing System for Ship Recovery' (SuSy), gas inflated balloons are envisaged to be used for providing reserve buoyancy to damaged ships for the purpose of preventing ship capsizing and/or sinking, along with lifting wreckages from the seabed. This paper presents the proof of concept tests of the prototype salvage units applied on a full-scale demonstrator section of a double bottom structure, together with measurements of its structural response. Two scenarios of internal and external deployment of the inflatable rescue units were examined. In both cases, the demonstrator was successfully salvaged. Data from the structural response of the demonstrator revealed certain operational aspects of the salvage system which directly affected the response of the salvaged structure. The monitored structural response of the demonstrator during operation of the salvage system was well within the elastic regime of its material (Grade A steel). Accordingly, finite element simulation of the salvage system supporting structure was conducted. The results of these simulations were found to be well correlated to the corresponding experimental measurements. Modelling and performance prediction of a centrifugal cargo pump on a chemical tanker In this paper, a single-stage, horizontal type centrifugal pump, which can be used in a chemical tanker's cargo operations, was modelled with MATLAB/Simulink software. The modelled pump was run with seven different fluids handled in chemical tankers which are ethyl alcohol, N-Propyl alcohol, phenol, chloroform, castor oil, 55% nitric acid and water. Therefore, the pump's performance curves and data sets were obtained for each situation. After these, a neural network was created with MATLAB/ Neural Network Fitting Tool application. Inputs of the network were volumetric flow, head, shaft power, torque, and net positive suction head. The output was the pump efficiency and it is estimated for each fluid from the numeric data. Mean squared error was very close to zero (1.1817e-6) and R [2] provided a prediction accuracy of 99.996%. According to these results, artificial neural network (ANN) had a satisfactory performance to predict the efficiency of a chemical tanker's centrifugal cargo pump. Oxygen depletion in enclosed spaces The aim of this research project was to determine the rate at which oxygen depletes in enclosed spaces. Cargo holds, chain locker and double bottom tanks are only some of the enclosed spaces found on a ship. The project was conducted using oxygen depletion experiments on scrap metal in sealed and open vented containers. In addition, the speed of oxygen depletion in a replicated chain locker is investigated. It is necessary to highlight the speed of oxygen depletion in cargo hold situations and other enclosed spaces and prevent loss of life in these dangerous spaces.In the oxygen-depleted atmosphere within an enclosed space, there is no sensory indication to cause alarm regarding the dangers within that space. Therefore the results of the experiments must be seen as a significant step in raising awareness of the dangers within an enclosed space, in improving health and safety standards and preventing loss of lives at sea. The formulation of epoxy-polyester matrix with improved physical and mechanical properties for restoration of means of sea and river transport The physical and mechanical properties of matrix based on epoxy resin ED-20 were investigated by adding unsaturated polyester resin ENYDYNE H 68372 TAE at different concentrations. The contents of polyester resin in the epoxy one by changing its concentration in the range of q = 10 - 120 mas.fr. were analysed. It was experimentally determined that matrix is characterised with its maximal rates of physical and mechanical properties with q = 10 mas.fr. of polyester resin ENYDYNE H 68372 TAE per 100 mas.fr. of ED-20. The received material is marked by the following indexes of physical and mechanical properties: fracture stresses during the flexion - σ fl = 56.2 MPa, the modulus of elasticity during the flexion - E = 4.2 GPa, impact toughness - W = 12.8 kJ/m[2]. It was proved that indexes of modulus of elasticity and impact toughness during the flexion of developed composite are higher in 1.5 - 2 times, in comparison with the indexes of matrix which is based on the epoxy resin ED-20. The images of fracture of the composite materials were analysed through optical microscopy. It was determined that the results of investigation correlate with the obtained indexes of physical and mechanical properties, confirming their authenticity. The clearly defined heterogeneity of given polymers was not observed on the photos of fracture, which indirectly points on the compatibility of the ingredients of the binder, chosen for the study. Featured risk evaluation of nautical navigational environment using a risk cloud model In this paper, a risk cloud model (RCM) based on cloud model theory is proposed for evaluating risks in the nautical navigational environment (NNEt). To validate the method, the proposed strategy is applied to a risk evaluation of NNEt of the waters around the Hangzhou Bay Sea-Crossing Bridge and the results are compared with those of a conventional fuzzy theory-based evaluation approach. The results show that risk evaluations based on the proposed RCM are robust, can accommodate uncertainty, can leverage multi-channel information, and provide a full-featured analysis of NNEt. In addition, plots of the RCM droplets provide an intuitive depiction of the risk level, which is extremely beneficial in analyses of the results. Thus, the proposed RCM and implementation method is well-suited for use in conducting risk evaluation of NNEt. Hydrodynamic analysis of a deep-sea pressure equaliser Deep-sea hydrothermal source investigation utilising water and gas samples to characterise ocean resources requires water sampler, both large and small volume. How to preserve the original captured samples in the sampler during its lifting process to the surface remains challenging. Our research develops a piston type deep-sea pressure equaliser installed in the sampling bottle to solve the pressure problems using pressure self-adaptive principle. This paper describes a new mathematic analysis of the pressure equaliser and carries out its numerical computation accordingly. Two different scenarios i.e. sampling water only and sampling water as well as gas, are selected to evaluate the performance of this pressure equaliser. The result shows that this pressure equaliser is capable of adjusting the pressure properly in the lifting process of the deep-sea water sampling bottle and thus helps to secure the sampling quality. Therefore, the design of pressure equaliser is acceptable. An energy-efficient nonlinear robust track keeping control algorithm for the Maritime Silk Road This paper proposes a nonlinear robust track keeping control algorithm for energy efficiency from the perspective of ship motion control, based on the great significance of the twenty-first Century Maritime Silk Road. Replacing the course error ee with the sin(ωe)sin⁡(ωe) driven by the sine function in course keeping loop to serve as the input of the controller, the simulation results tested on the ship autopilot simulation platform show that the controller output (maximum rudder angle) decreases, and the control effect of track keeping is better than the previous one in aspects of energy efficiency and rudder shake inhibition. Otherwise, the proposed nonlinear track keeping control algorithm still works in relatively severe sea states. Substituting the nonlinear feedback for the linear feedback has the generality of control effect improving. Exergoeconomic and air emission analyses for marine refrigeration with waste heat recovery system: a case study Nowadays, shipping industry faces challenges of energy efficiency and reducing of fuel consumption. Moreover, Waste Heat Recovery Systems (WHRS) regarding energy efficiency have a larger focus to utilise the heat energy lost from all thermal processes from ship engines. WHRS is one of the best methods to reduce fuel consumption and implicitly emissions. Refrigeration system, which can be evaluated as one of these systems, has a high energy efficiency potential on ships. In this study, exergoeconomic and air emission analyses of a case study ship named M/V Ince Ilgaz have been performed by comparing Vapour Compression Refrigeration System (VCRS) and VCRS with WHRS on exergy destruction and Second Law Efficiency in case of variable sea water temperature with 15 different refrigerants. Furthermore, a novel proposed WHRS is used for preheating of an accommodation water which leads reducing of exergy destruction about 9.31 - 10.60% while using R134A refrigerant. The fuel consumption due to refrigerant compressor has a 36% increase with the 10°C increment of sea water temperature. The increase of CO2, SO2, NO x and Particular Matter (PM) emissions is found about 183.40, 3.10, 4.65 and 0.47 tonnes, by the increase of sea water temperature from 20°C to 30°C for the fleet of 15 ships, respectively. In conclusion, using waste heat recovery on refrigeration system could directly reduce fuel consumption and air emissions. Super-capacitor-based inverter control of wind energy system connected to weak grid This paper proposed an inverter control scheme for wind energy system under weak grid conditions. Wind-based hybrid energy system (HES) is often having the issue of poor transients and slow dynamic response due to the unpredictable nature of the wind speed and inertia of the generator who cannot respond the changes instantly. The proposed control scheme is designed to improve the transients and dynamic responses of the wind energy system. The control scheme introduces super-capacitor (SC) fast charging/discharging characteristics in the inverter control using a bidirectional buck - boost converter to improve the transients and the frequency response of the system. The inverter control scheme under the weak grid conditions is designed and analysed. The proposed control can be applied in the marine electrical system where wind system is installed in the marine to reduce fuel consumption and improve electric power quality. Wind-based HES is modelled and simulated in the MATLAB to demonstrate the effectiveness of the proposed control scheme under different conditions. The results are presented and compared to the conventional control scheme to justify the improvement in the power quality and the frequency response of the system. Comparing flow cytometry and microscopy in the quantification of vital aquatic organisms in ballast water The ability to quantify vital aquatic organisms in the 2 - 50 um size range was compared between five different flow cytometers and several different microscopes. Counts of calibration beads, algal monocultures of different sizes as well as organisms in a Wadden Sea sample were compared. Flow cytometers and microscopes delivered different bead concentrations. These differences between the instruments became larger for algal monocultures and were even higher for the Wadden Sea sample. It was observed that the concentration differences were significant between flow cytometer and microscope counts, and that this difference increased with the size of the objects counted. Microscope counts were more accurate for larger (50 um) objects because cytometers struggled with bigger particles that clogged the instruments. Contrary to microscopy, the flow cytometers were capable of accurately enumerating cultured cells in the 2 - 10 um size range and cells in the lower size range of the 10 - 50 um size class. Flow cytometers were also well-suited to assess low abundance samples due to their ability to process larger volumes than microscopes. The results were used to indicate which tools are suitable for ballast water monitoring: flow cytometry is a suitable technology for an indicative and real time analysis of ballast water samples whilst only microscopy would be robust enough for detailed taxonomical analyses. System-based calibration of offshore Malaysia environmental action factors for the ISO 19902 PETRONAS offshore fixed platforms Previously, the action factors, which are generally used in the region of Malaysia is based on the data from the Gulf of Mexico (GoM) or the North Sea, where the environmental condition is much harsher than the offshore Malaysia. Direct application of the high load factors in these regions can result in a waste of economic resources. Based on a reliability-based Load and Resistance Factor Design approach, the current study is carried out for the provision of system-based approach for the calibration of Offshore Malaysia Annex Extreme Environmental Action Factors based on the data from Malaysian waters. With the accessible SEAFINE and HYCOM hindcast database, long-term statistics of extreme environmental load have been derived for six platform locations in Malaysian waters, following the response-based Load Statistics Module (LSM) process. The Generic Load Model (GLM) including the wave-in-deck forces, which has been employed for transferring the wave motion statistics to structural response statistics, is developed in the paper. The structural resistance uncertainty is characterised by the minimum reserve strength ratios (RSRs) derived from a theoretical model that relates with environmental load factor. A structural system reliability analysis based on the FORM method has been performed for a range of minimum values of theoretical RSR for jacket structures fully designed to ISO Code 19902. Partial action factors associated with extreme environmental loads are calibrated to achieve a target reliability level in terms of reliability index. A new lifting pump for deep-sea mining A lifting motor pump for deep-sea mining was proposed as a researching object. First, the structure and design method of this pump was introduced. Second, the simulating computations of the impeller in this pump were conducted at different working processes. Finally, the experiment of a two-stage lifting motor pump was conducted in the process of work to prove the applicability of this new lifting motor pump for deep-sea mining. Simulation-based investigation of a marine dual-fuel engine Recent developments have rendered the Dual Fuel (DF) engines an attractive alternative solution for achieving cost-efficient compliance to environmental regulations. The present study focuses on the safety investigation of a marine DF engine in order to identify potential safety implications. This investigation is based on an integrated engine model, which was developed in GT-ISE(TM) software and is capable of predicting both the engine steady-state behaviour and transient response. The model includes the engine thermodynamic simulation module as well as the engine control system functional module; the latter is responsible for implementing the ordered load changes and the operating mode switching. The developed model is first validated against available published data and subsequently used to simulate several test cases with fuel changes, from gas to diesel and diesel to gas with rapid and with delayed wastegate valve operation. The derived simulation results are used to investigate the potential safety implications that can arise during the engine operation. The results demonstrate that the engine - turbocharger matching as well as the wastegate control are critical parameters for ensuring the compressor surge free operation during gas to diesel modes transition. Robustness analysis of dual actuator EGR controllers in marine two-stroke diesel engines Exhaust Gas Recirculation (EGR) was recently introduced in large marine two-stroke diesel engines to reduce NOxNOx-emissions. Controlling EGR flow during accelerations, while keeping good acceleration performance is challenging, due to delays in the scavenge receiver oxygen measurement and upper limits on fuel for avoiding black smoke. Previous oxygen feedback controllers struggled during accelerations, but a new EGR-controller based on adaptive feedforward (AFF) has been successful. Nevertheless, further analysis and tests are required before deploying the controller to more EGR ships. A simulation platform is a great asset to test controllers before expensive real-world experiments are conducted. A new EGR flow controller is proposed and tested in a complete ship simulation model. Several acceleration scenarios show that the low load area is most challenging. Controller robustness is analysed in this area, showing that pressure sensor bias in the EGR flow estimator is the most critical factor, which could lead to black smoke formation. This can be prevented with sensor calibration or by using a differential pressure sensor. Errors in the parameters of the flow estimators are not as important. This is a useful result because the right parameters of the flow estimators might be difficult to obtain, on a new engine. Distributed energy management for ship power systems with distributed energy storage Electric systems for naval applications create a challenge for the power system associated control. When incorporating loads with a high-power ramp rate within what is essentially an islanded microgrid, energy sources that supplement generators must be used due to the ramp rate constraints of the generators; this is where energy storages play a key role. A higher-level control layer is needed to effectively coordinate the distributed energy storage in order to ensure that the effect on the generators is minimised. The Energy Management layer is responsible for maintaining the desired state of charge for the distributed energy storage and ensuring that load demand is met while minimising ramp rate violations. In this paper, a distributed Energy Management scheme for a 4-zone ship power system is presented. Experimental results validate the proposed control technique using hardware controllers interfacing with a real-time simulator. Assessing complex failure scenarios of on-board distributed systems using a Markov chain Vulnerability reduction is an important topic during the design of naval ships because they are designed to operate in hostile environments and because their on-board distributed systems are becoming increasingly complex. The vulnerability needs to be addressed in the early design stages already, in order to prevent expensive or time-consuming modifications in later, more detailed design stages. However, most existing methods for assessing the vulnerability are better suited for more detailed design stages. Furthermore, existing methods often rely on pre-defined damage scenarios, while damage - or system failure in general - may also occur in ways that were not expected beforehand. This paper proposes a method that addresses these gaps. This is done by incorporating several additions to an existing vulnerability method that has been developed by the authors, using a Markov chain. With this method, there is no longer a need for modelling individual hits or failure scenarios. The additions are illustrated by two test cases. In the first one, a notional Ocean-going Patrol Vessel is considered, and damage is related to physical locations in the ship. The second test case considers a chilled water distribution system in more detail, with failures modelled independent from the physical architecture. The quantitative nature of the results provide an indication of the generic, overall vulnerability of the distributed systems, which is meant to be used in the early design stages for identifying trade-offs and prioritising capabilities. A collision avoidance algorithm for ship guidance applications The paper presents a collision avoidance algorithm for ship open sea navigation, based on an ad hoc modified version of the Rapidly-exploring Random Tree (RRT*) algorithm. The proposed approach is designed to act as the high level layer of the navigation control structure for an autonomous ship. Collision and grounding still represent the primary source of sea accidents, thus an automatic system able to detect static and moving obstacles and plan an evasive route could significantly improve safety during navigation, especially in crowded areas. Focusing on the maritime field, a review of the scientific literature dealing with collision avoidance is presented, showing potential benefits and weaknesses of the different approaches. Among the several methods, details about the RRT and RRT* algorithms are given. The ship path planning problem is introduced and discussed, formulating suitable cost functions and taking into account both topological and kinematic constraints. The algorithm is able to manage multiple moving obstacles with variable speed and course. Eventually, a time-domain ship simulator is used to test the effectiveness of the proposed algorithm over a number of realistic operation scenarios. The obtained results are presented and critically discussed. Adaptive steering control for an azimuth thrusters-based autonomous vessel The proposed paper presents the design and development of the combined guidance and control strategies for the autonomous navigation of unmanned vessels characterised by azimuth-based thrust architecture. Autonomous marine vehicles (AMVs) are consolidated technological tools commonly employed for different tasks such as exploration, sampling and intervention. With the final aim of autonomous shipping, the AMVs capabilities have to be migrated and adapted towards the reliable and safe control of commercial-like unmanned vessels. These last are spreading thanks to a number of technological research projects. The employment of unconventional hull shapes combined with propulsive layout based on azimuth thrusters requires robust guidance techniques to provide precise and reliable motion control during navigation. The paper introduces a dual-loop guidance and control scheme able to provide advanced navigation capabilities. An inner control loop, devoted to the actuation of the azimuth thrusters, allows the tracking of reference course angle (namely the autopilot). Such a control loop is characterised by a modified PD regulation scheme, where a novel adaptive derivative component is inserted in order to improve the convergence curve towards the required course reference. The outer guidance loop, based on Lyapunov and virtual-target approach, allows the vessel to track generic desired paths, thus enhancing the autonomous navigation capabilities. The paper will provide a deep design and analysis approach for the developed techniques, as well as simulation results of the combined guidance and control scheme, proving the reliability of the proposed approach in different operative conditions. Energy storage design considerations for an MVDC power system The U.S. Navy is investing in the development of new technologies that broaden warship capabilities and maintain U.S. naval superiority. Specifically, Naval Sea Systems Command (NAVSEA) is supporting the development of power systems technologies that enable the Navy to realise an all-electric warship. A challenge to fielding an all-electric power system architecture includes minimising the size of energy storage systems (ESS) while maintaining the response times necessary to support potential pulsed loads. This work explores the trade-off between energy storage size requirements (i.e. mass) and performance (i.e. peak power, energy storage, and control bandwidth) in the context of a power system architecture that meets the needs of the U.S. Navy. In this work, the simulated time domain responses of a representative power system were evaluated under different loading conditions and control parameters, and the results were considered in conjunction with sizing constraints of and estimated specific power and energy densities of various storage technologies. The simulation scenarios were based on representative operational vignettes, and a Ragone plot was used to illustrate the intersection of potential energy storage sizing with the energy and power density requirements of the system. Furthermore, the energy storage control bandwidth requirements were evaluated by simulation for different loading scenarios. Two approaches were taken to design an ESS: one based only on time domain power and energy requirements from simulation and another based on bandwidth (specific frequency) limitations of various technologies. Enhancement signal detection in underwater acoustic noise using level dependent estimation time-frequency de-noising technique In sonar and underwater digital communication, optimal signal detection is imperative. In many applications, additive white Gaussian noise (AWGN) is assumed; thus, a linear correlator (LC), which is known to be optimal in the presence of AWGN, is normally used. However, underwater acoustic noise (UWAN) affects the reliability of signal detection in applications in which the noise originates from multiple sources and doesn't follow the AWGN assumption. As a result, an LC detector performs poorly in tropical shallow waters. Accordingly, this study aims to develop a detection method for improving detection probability (PDPD) by using a time - frequency denoising method based on the S-transform with multi - level threshold estimation. The UWAN used for the validation is sea truth data collected at Desaru beach on the eastern shore of Johor in Malaysia with the use of broadband hydrophones. The performances of four different detectors, namely, the proposed Gaussian noise injection detector (GNID), a locally optimal (LO) detector, a sign correlation (SC) detector, and a conventional LC detector, are evaluated according to their PDPD values. For a time-varying signal, given a false alarm probability of 0.01 and a PDPD value of 90 percent, the energy-to-noise ratios of the GNID, LO detector, SC detector, and LC detector are 8.89, 10.66, 12.7, and 12.5 dB, respectively. Among the four detectors, the GNID using the S-transform denoising method achieves the best performance. Underwater thrust vectoring based on inflated surface The traditional way to underwater thrust vectoring is mainly based on a specially designed mechanical system which provides the propulsion system additional degree of freedom. This method is effective in field applications. However, there are several drawbacks such as complex structure and large dead weight. In this paper, a new type of thrust vectoring method, based on inflated surface is proposed. There are six inflatable surfaces uniformly distributed around the axis of the propulsor. The mechanism of this method is the Coanda effect. Through the control of the curvature of the surface, the deflection angle can be adjusted. The effect of the depth, the size and the curvature of the surface on the deflection angle is numerically studied. The deflection angle decreases with the depth, but increases with the surface size. The deflection angle first increases then decreases with the curvature due to flow separation. A maximum deflection angle of 14° is realised. The pressure and velocity distribution is given to illustrate the mechanism behind the variation of deflection angle in different cases. Strength and stability studies of ring-stiffened circular cylindrical shells using self-organising maps (SOM) and data mining analysis Ring-stiffened circular cylindrical shells are commonly used in many engineering applications. In the present study, the strength and stability characteristics of ring-stiffened circular cylindrical shells are investigated using self-organising map (SOM) technique and multivariate approaches. Stress and critical pressure analysis of ring-stiffened circular cylindrical shells under hydrostatic pressure are obtained from the design guide. Design of experiment (DOE) technique is employed for generating design samples. The information behind these design samples can be studied by the self-organising map (SOM) approach. Multivariate analysis is adopted for post- processing of SOM's results. The clustering of samples can be obtained from the codebook of SOM with the help of multivariate analysis. The samples in different clusters would have their own characteristics. The variable significance in each cluster is examined. The dimensionality reduction approach is used for finding the structure of the data set. A useful visualisation approach, parallel coordinate plot with clustering information, is proposed for visualisation of design samples. Correlations and clustering of variables are also studied by the hierarchical method. The ranking of samples is conducted by minimising the distance to an `ideal' solution in the data set. A final compromises sample is obtained. A numerical example is provided for the illustration of the present analysis. A new risk-based inspection methodology for offshore floating structures The Risk-Based Inspection (RBI) programme is a part of the overall Asset Integrity Management (AIM). The RBI programme offers benefits such as providing a rational basis for inspection plans and efficient tools for updating those plans. The ultimate goal of the RBI programme, coupled with a more generic approach to asset integrity management, is to verify the integrity of the facility and its fitness for service. The present study attempts to address a step by step approach which can be followed for RBI planning inspection. Guidelines regarding the various steps of the RBI process are also provided in order to assist in preparing and maintaining the RBI programme. An advanced tool for semi-quantitative risk assessment of offshore floating structures is developed, where the user guide is presented. Big data analysis of port state control ship detention database This study uses big data analysis to examine the relationships between detention deficiencies and external factors as well as between detention deficiencies themselves. Data are taken from the ship detentions database that has been accumulatively published from Port State Control inspections, which have been executed for many years in the member authorities of the Tokyo Memorandum of Understanding. Each factor of the PSC detention database is analyzed, with additional preprocessing, to explore the potential regularity of ship detention deficiencies. The results show that using association rule mining techniques in big data analysis can accurately and objectively mine the regularity correlation between ship detention deficiencies, as well as between these deficiencies and related factors. The techniques can provide countermeasures and be used as a reference by ship management personnel during the corresponding PSC inspection, reducing the detention rate of ships. This can provide a more targeted method to be adopted by the maritime authority in the practical work of inspections. By using this method, the working efficiency of staff members can be significantly improved, reducing the adverse influences brought to navigation safety and the marine environment due to sub-standard vessels. Effect of uncertainty on techno-economic trade-off studies: ship power and propulsion concepts Results of trade-off studies aiming to compare different ship power and propulsion configurations inherently contain uncertainty. This is true for both the technical part and for the financial part of the trade-off study. The technical part typically includes the system characteristics of the vessel as inputs, resulting in predicted fuel consumption and emissions at various ship speeds. Fuel consumption numbers subsequently feed into the financial part of the analysis which typically includes prices of equipment, fuel prices or fuel price scenarios, the discount factor and other aspects such as the operational profile which plays an important role in the trade-off study. Finally, financial KPI's such as Net Present Value and payback period can be compared between different power and propulsion concepts, thereby supporting decision makers in the selection of a specific configuration or retrofit. In this paper, the effect of uncertain input parameters on both the intermediate technical output and on the financial KPI's is demonstrated by means of a case study. The study shows that the uncertainty associated with relevant KPI's is sufficiently large to warrant further investigation beyond accepting model predictions as completely accurate, particularly when conducting techno-economic trade-off analysis of ship power and propulsion configurations. In the broader context, consideration of uncertainty is a must for statutory and regulatory authorities in the formulation of policy. A numerical investigation on grouted connections for offshore wind turbines under combined loads Grouted connections have been widely used in offshore wind support structures to connect upper structures and lower foundations. Their mechanical behaviour is critical to the integrity of the entire support structure. In this study, numerical models were established and verified to investigate the stress distributions of grouted connections under combined compression and bending loads. It was found that the locations of shear keys had a significant influence on the maximum Mises stress in the steel tubes and Tresca stresses in the grout. The following parametric study of geometric configurations indicated that the thickness of a tube had a reverse relationship to the maximum stress. Meanwhile, the maximum Tresca stress was not sensitive to any geometric configurations, except grout thickness. Considering the gap between the detailed solid modelling and design process, based on beam modelling, a nominal average stress was computed over the compressed side of the steel tube within grouted connections. Consequently, stress correlation factors were obtained between the maximum Mises stress in a steel tube and the nominal average stress of a grouted connection. This method can be used in the preliminary design to evaluate the general behaviour of grouted connections under combined actions. A modified AD-TRIZ hybrid approach to regulation-based design and performance improvement of ballast water management system This paper presents a novel methodology in the design and performance enhancement of a regulation-compliant Ballast Water Management (BWM) System, where three methodologies were integrated in the process. The application of the multi-functional framework of classical Axiomatic Design (AD) in developing a design matrix was firstly modified using the influence of the Software - Hardware - Environment - Liveware interaction concept to factor all the system's interacting elements into the solution design. The BWM Convention was used as a guide to identify the requirements for the proposed system design. The identified AD couplings in the design matrix were then analysed using Sufield technique; a concept of Altshuler's Theory of Inventive Problem Solving. The design's most promising performance enhancement pathways were subsequently determined and prioritised. A new method of the top-down parametric design for quick subdivision based on constraints Currently, the subdivision method of bottom-up is used widely in the ship industry, which is time-consuming and tedious as the ship consists of compartments of several tens to several hundreds. To improve the speed of subdivision, a new parametric subdivision method of the top-down approach based on constraints is proposed. The proposed method is based on space subdivision as if we cut a box into many small pieces. For this, several parameters (bulkhead and inner knuckle positions) are set to generate the subdivision polyhedrons to cut the hull and Non-manifold modelling technology is applied to generate solid model. Then the subdivision constraints are linked with the solid model and the subdivision scheme is optimised by constraints' knowledge, and the hold capacities can be calculated rapidly by this method. This method not only generates and modifies the solid model quickly, but can also reflect the designer's ideal vividly. A fuzzy neural network combined with technical indicators and its application to Baltic Dry Index forecasting To enhance the technical analysis prediction of freight rate trend in the dry bulk shipping market, a fuzzy neural network combined with technical indicators is developed. Firstly, five technical indicators often used in the financial market including %R, RSI, MACD, CCI, and MA are chosen as input signals, and the accuracy rate for forecasting Baltic Dry Index (BDI) trend is about 62%. Secondly, a traditional fuzzy neural network is applied to the forecasting of BDI. The RMSPE of forecasting BDI by the traditional fuzzy neural network is 24.76. Finally, the integrated fuzzy neural network combined with technical indicators is applied to forecasting BDI. The accuracy rate for forecasting BDI trend is 83%, and the RMSPE of forecasting BDI is 1.89. The results show that the integrated fuzzy neural network combined with technical indicators has a higher forecasting accuracy rate than the technical indicator approach or the fuzzy neural network approach. Optimising design and power management in energy-efficient marine vessel power systems: a literature review The conventional marine vessel power systems typically have the potential to improve their fuel consumption and their emissions. This can be done by redesigning the system configuration, the machinery and the power management strategy. The addition of options in power management allows for the running of individual power sources closer to their optimal operating point. However, this immediately raises questions about how to redesign the system and how to operate it to maximise the benefits. The information needed to answer these questions is often scattered around separate sectors of the marine industry. The system integrator needs to be able to combine the complex dependencies of these individual sectors to formulate the big picture that describes the whole power system. Numerical optimisation algorithms provide solutions to develop methodologies to solve multi-variable and potentially multi-objective problems. This literature review presents the authors' findings of design and power management optimisation cases in marine vessel power systems. Combustion and emissions of a glycerol-biodiesel emulsion fuel in a medium-speed engine Emulsion fuels are one option currently being explored to reduce powerplant and maritime emissions. Emulsification enables hydrophilic, typically low-value, molecules to be incorporated into traditional hydrocarbon fuels. Energetic oxygenated molecules, such as glycerol, are biorenewable and have the potential to reduce refueling costs and carbon emissions. When properly formulated, emulsions improve diesel combustion characteristics and reduce particulate matter (PM) and oxides of nitrogen (NOx). This paper explores the utilisation of glycerol-biodiesel emulsion (23 wt% glycerol) in a one-megawatt, six-cylinder, medium-speed diesel engine at constant speed to determine impacts on combustion dynamics, emissions and overall suitability. Fuel performance is compared to ultra-low-sulphur diesel (ULSD) and 380 cSt. heavy fuel oil (RMG 380). Idle emissions are shown to be comparatively poor due to low combustion stability and longer combustion delay. As load is increased to 25% of peak output, reductions in carbon monoxide (CO) and PM are observed. The lower energy density, however, restricts peak engine power which achieved 90% of full load at maximum fuel consumption. Despite engine maximum power derating, glycerol emulsion fuels require no engine modification and show promise as a powerplant fuel with a low-carbon footprint. Structural damage of ship - FPSO collisions The focus of this study is paid to general methodology and design of accident scenarios for floating production, storage and offloading systems (FPSOs) using more sophisticated tools such as the nonlinear FEM so that structural responses during and after a collision can be more precisely predicted. The extent of FPSO hull damage due to the different collision scenarios and impact energy levels are determined using the guidelines contained in the NORSOK Standard [2004a. Design of steel structure N-004, Rev. 2 October 2004; 2004b. Materials selections M-001, Rev. 4 August 2004]. The collision scenario is considered for accidental limit state with five-year onsite environment in this study. The various collision scenarios are defined such as supply vessel collision bow on, supply vessel collision side on, supply vessel collision stern on, and off take tanker collision bow on. Nonlinear finite element analysis, using large deformation FEM, is applied to analyse the damages to FPSO hull structures. Conservatively the impacting vessel is considered non-deformed during the collision (whole energy absorbed by the FPSO). Damage evaluation to hull parts of FPSO is checked, including flare and flares foundations, aft muster station, offloading reel and piping and green water protections above cargo deck. The results and insights derived from the present study are summarised. Analysis of lift, drag and CX polar graph for a 3D segment rigid sail using CFD analysis A rigid sail is an alternative source of propulsive power for marine vessels and they have been installed on ocean-going powered ships in the past. To determine the potential power that could be harnessed by a rigid sail its specific lift and drag characteristics must be known. For this purpose, a segment rigid sail was studied utilising computation fluid dynamics analysis and a virtual wind tunnel. This analysis enabled lift and drag force tables to be produced and based on these data a CX (thrust coefficient) polar graph for the sail was derived. Analysis of velocity magnitude and static pressure contours was also undertaken and the aerodynamic characteristics of the sail including airflow separation and wake size studied. On improving the earthing quality in ship electric energy systems This paper presents a contribution in improving the earthing system of ship electric grids, which affects the phenomena developed in highly unbalanced operating conditions like the single - and two-phase faults. Furthering previous work, an alternative to ungrounded earthing scheme is introduced, namely the grounding via a conditionally varying resistance. Within this framework, the mathematical expressions with respect to zero sequence impedance, of voltages and currents under study are yielded. Furthermore, as a means to easily evaluate the quality of earthing, appropriate indices are introduced. Thus, it is shown that the effect of the earthing scheme in fault conditions can be adjusted based upon certain design limits. The application of hybrid photovoltaic system on the ocean-going ship: engineering practice and experimental research The constant development of electronic inverter technology has played a key role in promoting the exploration and development of solar ships. For the large-scale ocean-going ship platform, the critical issue of applying solar photovoltaic (PV) system is integrating PV equipment into the ship power system (SPS) without changing its original structure. This paper compares the existent technical differences for applying the off-grid and grid-connected PV system in the SPS and proposes the basic design principles for marine integration applications. The 5000 PCTC ro-ro ship is set as the application object, on which a hybrid PV system with large-capacity lithium battery storage device is designed and installed as an independent subsystem. The typical feature of this hybrid PV system is that it can implement operation mode switching between off-grid and grid-connected, according to the evaluation on solar radiation resource, power load requirement and state of charge in the lithium battery. The test results show that this PV system has a stable operation characteristic under different operation modes. In addition, this ship-based PV power system has automatic and reliable operation management capability, which could effectively reduce manual control frequency and maintenance workload of a marine engineer. Improved LJF equations for the uni-planar gapped K-type tubular joints of ageing fixed steel offshore platforms The distribution of fixed steel offshore platforms around the world reveals a global fleet that has exceeded or is approaching the end of the design life of the facility. In many operating areas, there is an attraction to continue using these ageing facilities due to continued production or as an adjoining structure to facilitate a new field development or expansion. To justify continued life extension of the fixed platform, various integrity assessment techniques are often used. One of the major techniques used is based on the phenomenon of local joint flexibility (LJF). While the phenomenon of LJF has been well known in the offshore industry since the early 1980s, there has been little experimental data available. In 1983, Amoco conducted an experimental study primarily to determine stress concentration factors associated with gapped K-type steel tubular joints. The LJFs calculated were based on the effects of in-plane bending, out-of-plane bending and axial compression and tension. The derivations of the existing LJF equations have evolved in many ways including use of finite element (FE) methods to predict the joint behavior. There has been no benchmarking exercise to large-scale experimental data. This paper provides an improvement on existing LJF equations by benchmarking the Amoco K-joints test results to a FE model and through a detailed parametric study. Improved formulations are provided for local joint flexibilities for gapped uni-planar K-type tubular steel joints. Diver deployed autonomous time-lapse camera systems for ecological studies Photographic time-lapse techniques are especially useful in the marine realm for visualising long-term processes and remote monitoring of sites/objects/organisms where the presence of researchers might cause some study bias, or access is limited or impossible. With rapid advances in technology development there is easy access to new tools for time-lapse photography and setting up systems is relatively inexpensive. The essential requirements for low-cost autonomous time-lapse camera systems to be self-sufficient and reliable enough to withstand the extended periods of deployment (up to one year) on the sea floor at up to 50 m depth are presented. In this example a custom-made system developed originally for monitoring the activity of filter/suspension feeders and scavenging fauna in the polar conditions is described. The major issues encountered during the preparation and deployment which should be of benefit to users involved in underwater time-lapse photography are considered. Analysis of drag, airflow and surface pressure characteristics of a segment rigid sail Rigid sails have been used on powered ships in the past to reduce fuel consumption and recently there has been renewed interest in this technology. Rigid sails could potentially be used on a wide variety of ships and vessels; therefore, it is necessary to better understand the performance and characteristics of these devices. The present work investigated the airflow and drag characterises of a segment rigid sail using a virtual wind tunnel. Sail surface pressures were also observed and recorded. It was determined that the most effective angle of attack for this type of rigid sail was between approximately 10° and 30°. It was also noted that at larger angles of attack, the presence of a significant wake would limit the effectiveness sails mounted downwind. Utilising elastic waves of acoustic emission to assess the condition of spray nozzles in a marine diesel engine This article presents an original method of using elastic waves of acoustic emission (AE) to assess the technical condition of nozzles in marine diesel engines. The developed method makes it possible to detect the early stages of wear by friction of the needle-nozzle pair. The phenomenon of AE and the possibility of its use in evaluating the condition of marine diesel-engine injectors are discussed. Experiments are presented in which the AEs were used to evaluate the condition of both a new injector and an injector where galling of the needle-nozzle pair is developing. The measurements were processed by time - frequency analysis and wavelet transformation. Application of the microscopic and Mössbauer studies to the analysis of a marine diesel engine crankshaft In the paper, the issue of the material quality of the large-size crankshaft is discussed. In order to ensure the reliability and wear resistance of the crankshaft, its chemical composition was checked in detail, revealing also the presence of trace elements omitted in Standards. The mechanical characteristics of the material were measured using the samples taken from the upper and lower parts of the forging. The novelty of the quality analysis was the additional measurement with the conversion electron Mössbauer spectrometry method. High phase purity, homogeneity and isotropy of the investigated material were confirmed. The carbon-saturated ferrite was detected as dominating phase of about 95% content. The observed martensite structure appeared to be correct, and the material met the specification requirement. Only slight differences in atomic ordering of alloying elements in ferrite structure were observed between the upper and lower part region, which was not considered a threat. It was assumed that the confirmed chemical purity and homogeneity contributed to the high straightness of the tested material. Using computational fluid dynamic and artificial neural networks to predict the performance and cavitation volume of a propeller under different geometrical and physical characteristics In the current study, simulations of hydrodynamic characteristics of a propeller under different geometrical and physical characteristics are conducted by the computational fluid dynamic (CFD). Then, by designing appropriate artificial neural networks (ANNs), the hydrodynamic performance and cavitation volume of propellers are predicted under intended conditions. For this accomplishment, finite volume-based Navier - Stokes equations associated with incompressible large eddy simulation turbulence model are used. In order to verify the computational procedure, mesh sensitivity analysis and validation study are conducted and appropriate accuracy is observed. In the CFD simulations, propeller thrust, torque and cavitation volume are computed under different pitch ratio (P/D), rake angle (RA) and skew angle (SA), advance velocity ratio (J) and cavitation number (σ). By the CFD results, a significant increase in propeller thrust and torque is observable by enhancement of P/D and positive value of RA. Moreover, maximum mean square errors of ANNs output in the prediction of propeller thrust, torque and cavitation volume achieved are 0.000111, 7.4206E−5 and 0.000667, respectively. Also, related to ANNs' weights and bias, four set of equations are proposed to predict the performance and cavitation volume of propellers. Increasing energy efficiency in passenger ships by novel energy conservation measures To achieve increasing emission requirements, the cruise ship industry is working to develop higher efficiency ships. Cruise ships are different from other ship types in their relatively higher consumption of electrical power, steam and hot water. Several novel high-efficiency system concepts are possible for on-board electrical power generation and other utility services, each with differing impacts and first costs. Low-emission concepts novel to the cruise ship industry include combinations of exhaust gas heat recovery, heat pumps, steam turbines and organic Rankine cycles (ORCs). Yet, evaluation of these concepts is difficult given the different operating modes of cruises, and overall efficiency is dependent on the dynamic operational sequences. In this paper, we compare alternative energy efficiency concepts for cruise ships through simulation studies of the ship operations when equipped with different novel power generation systems. We find that the dual pressure steam systems and ORC offer the greatest potential for energy efficiency improvements in the cruise ship industry. We also find that relatively conventional technologies enable cruise ships to comply with planned upcoming higher ship energy efficiency requirements. Power profile for segment rigid sail Rigid sails have the potential to lower fuel consumption on powered ships by using the wind as a supplementary source of propulsive power. The amount of propulsive power that can be provided by each sail is dependent on a number of variables, which include the sail type, sail area and number of sails deployed. Since each type and size of rigid sail has its own performance characteristics, it is difficult to estimate the propulsive power that could be provided by a particular rigid sail and to make comparisons between different rigid sails. To overcome this problem, a rigid sail power profile is proposed. This power profile will outline key performance characteristics and allow comparisons to be made between different types of rigid sails. RANS simulation of the tip vortex flow generated around a NACA 0015 hydrofoil and examination of its hydrodynamic characteristics Understanding the flow behaviour of the tip vortex around hydrofoils, the corresponding pressure distribution and generated tip vortices are the key factors for prediction of cavitation inception and acoustic noise generation. Accordingly, numerical simulation of tip vortex flow generated around a National Advisory Committee for Aeronautics' hydrofoil (NACA 0015) is hereby investigated in the present study. To this end, the commercial software Ansys-CFX is applied, and Navier - Stokes equations are solved. Numerical simulations are performed using k-ω Shear Stress Transport (SST) model turbulence model. Validity of the solution is examined by comparing the obtained results against available experimental data for pressure coefficient distribution and lift coefficient (CL). Moreover, the lift and drag coefficients are computed and presented at different angles of attack. The velocity vector fields along with generated tip vortices are also presented. Two different vortices are observed, which are combined in the neighbourhood of trailing edge. On the other hand, the pressure distribution at the tip section is reported and effects of tip vortex on pressure variations are displayed and analysed. Ultimately, the streamlines are illustrated, and flow behaviour in the tip area and at the wake of the NACA hydrofoil is investigated. VIV-induced fatigue damage study of helical wires in catenary unbonded flexible riser in time domain Helical wire is a key component of unbonded flexible riser, and is vulnerable to fatigue failure. The present study simplifies the flexible riser into a beam element with constant axial and bending stiffness, and then investigates the vortex-induced vibration (VIV)-induced fatigue damage of the helical wire in a catenary flexible riser using a time domain VIV approach. The simplification of the flexible riser is based on the equivalence of axial and bending stiffness. The former can be iteratively calculated based on the integrated axisymmetric formulation in which the interlayer contact and separation are taken into account, while the latter is simply taken as the combination of all layers' bending stiffness and keeps constant since helical wire may not slide (i.e. full-sticking) under intact outer sheath due to high external pressure. Based on the simplification, this study compares the VIV characteristics of the flexible riser under full-sticking and full-sliding conditions of helical wires since the latter associated with conservative results is often applied, and then parametrically investigates the fatigue damage of the helical wire. The results indicate that the critical position is located near touchdown point, and seabed stiffness, helical wire lay angle and the top end position have a significant effect on the fatigue damage. A systematic simulation methodology for LNG ship operations in port waters: a case study in Meizhou Bay With the increment for liquefied natural gas (LNG) demand, LNG carriers are becoming larger in size. The operational safety of the carriers and the associated terminals is increasingly attracting attention. This is particularly true when a large LNG vessel approaches a terminal, requiring a detailed investigation of ship handling in port waters, especially in certain unusual cases. A full mission simulator provides an effective tool for research and training in operations of both port terminals and ships. This paper presents an experimental design methodology of the full mission simulation. The details as to how the simulation is achieved are described, and the simulation strategies applicable to LNG ships are specified. A typical case study is used to demonstrate and verify the proposed design methodology. The proposed methodology of the full mission simulation provides guidance for port safety research, risk evaluation and seafarer training. Simplified fatigue analysis of structural details of an ageing LPG carrier As the world gas carrier fleet is ageing, it is required to make sure that the transportation of oil and gas is carried out by quality vessels. Ship's conditions can remain of a good level throughout their service lives if they are operated in an appropriate manner, the current state of construction and equipment of each structure are clearly understood and inspections and maintenance are carried out regularly and properly. Fatigue is an important phenomenon that causes damage failure around the weld at member connections of ship structures. Fatigue happens due to high stress concentrations in the long term since loading conditions are time-dependent and continuous in dynamic sea environment. A fatigue crack starts at a localised spot and will with cyclic stress gradually increase over the cross section of the member. This study aims to calculate the fatigue life of longitudinal members amidships. These fatigue calculations are theoretical calculations that should be used as guidance for close-up inspections when ships are surveyed periodically to verify that they are maintained in an acceptable condition in accordance with international conventions, the Rules of Classification societies, etc. Energy efficiency analysis of a ship's central cooling system using variable speed pump Energy efficiency is getting more important due to energy cost and environmental effects. Pumping systems are important because they consume almost 20% of the total energy in land power installations. Comparing to the land installations mentioned above, the energy consumed by pumps in marine installations is much greater. In this study, it is shown that the energy consumed by pumping systems onboard ships is almost 50% of the total energy consumption. These pumps work at a constant speed, in other words constant mass flow. However, some of the systems are working under variable load. For example, seawater cooling system load depends on some parameters such as seawater temperature and engine load. Therefore, the consumed energy of this system can be reduced depending on the seawater temperature. A variable speed pump is one of the solutions to save the energy in this system, which will change the pump speed depending on seawater temperature. In this paper, the energy-saving potential of a central cooling water system is analysed. The saved energy and related annual cost of the pump occupied in the system are calculated for different seawater temperatures, and its environmental results are also discussed. A modified exponential score function for troubleshooting an improved locally made Offshore Patrol Boat engine In this paper, we present an Intuitionistic Fuzzy Technique for Order Preference by Similarity to the Ideal Solution model which is based on a modified exponential score function for detecting early failure in a locally made Offshore Patrol Boat engine, with special regard to component interaction failure, using groups of experts' opinions to detect the root cause and the engine systems most affected by the failures in the Boat engine. The study is aimed at providing an alternative method for the traditional product development failure mode identification and analysis methods which hitherto are limited when it comes to component interaction accidents and failure analysis in the machine system. The results from the study show that although early detection of failures in engines is quite difficult due to the dependency of machine systems and components on each other, using an intuitionistic fuzzy multi-criteria decision-making method which is based on experts' opinions these faults/failure can easily be diagnosed and detected. Studying the marine accidents of the Aegean Sea: critical review, analysis and results This paper aims at discovering characteristics, aspects and statistical patterns of recorded maritime accidents within specific zones of the Aegean Sea. A focused database with 10 years' accidents data has been studied on the special characteristics of the ships that had been involved. Based on this database, the paper presents results derived by statistical analysis that are able to provide valuable information about the accidents under examination in terms of type, size and other parameters of the involved ships. Two models are developed and implemented to establish a risk perspective view of maritime transport within the Aegean Sea; a stochastic Poisson function is used to determine the probability of accident occurrence for each one of the three sea zones of study. Next, a fault-like approach is explored as the basis for an alternative way of studying marine safety; in this context, a seismic-driven model is applied for the calculation of the probability of accident occurrence. To do so, the concept of energy release is used so as to develop a novel modelling approach for maritime accidents. Both models present in a systematic way the hazardous profile of maritime transport in the Aegean. The paper concludes with valuable insights. The parameterisation of turbulence in the marine environment here are many problems in the fields of flow modelling around structures and tidal stream energy yield analysis which require a thorough understanding of the turbulent and time-averaged flow speeds in marine environments. In this paper we examine the relationship between the turbulence intensity and mean tidal flow speed at a potential tidal stream power site. We report data from the Humber Estuary wherein an Acoustic Doppler Current Profiler were used to capture vertical profiles of the high frequency and mean tidal flow speeds throughout Spring and Neap, Flood and Ebb cycles. We show not only that our results extend earlier work but also suggest that the turbulence intensity, IT, can be described parametrically in terms of the mean flow, U, by an inverse power functionwhere the coefficient appears to be dependent upon the anisotropic nature of the turbulence. For the data reported here, the coefficient has value of about 17 - 18 and the exponent lies between −0.6 and −1.0. Confirmation of this relationship should not only improve engineering design work and energy yield analyses in turbulent tidal flows but also be applicable to other problems such as the prediction of sediment mass transport and pollution dispersal in estuarine management studies. Evaluation of safety exclusion zone for LNG bunkering station on LNG-fuelled ships With increasing interests in using LNG as a marine fuel, safety issues for LNG bunkering have brought about global discussion on establishing a safety exclusion zone around LNG bunkering areas. However, international consensus has yet to be reached in determining an appropriate extent of the zone to ensure safe liquefied natural gas (LNG) bunkering. The purpose of this study is to identify potential risks of LNG bunkering and to present a statistical method for determining the safe exclusion zone around LNG bunkering station with the help of a purpose-built computer program, integrated quantitative risk assessment (IQRA). A probabilistic risk assessment approach was adopted in this study to determine the safety exclusion zone for two case ships: one, a 300,000 DWT very large ore carrier (VLOC) and the other a 32,000 DWT bulk carrier. The results are then compared with those obtained by a deterministic approach and the discrepancies are discussed. It was found from this study that the frequency of bunkering is one of the key factors in determining the extent of safety exclusion zone. Thus, a somewhat surprising result of 36 m radius safety exclusion zone for the 32,000 DWT bulk carrier compared to 6.4 m radius for the 300,000 DWT VLOC was obtained. It was also found that the deterministic approach produced a much more extensive safety exclusion zone for the 300,000 DWT VLOC subjected to infrequent large-scale LNG bunkering operations compared to the probabilistic approach, while it was reasonably consistent with the probabilistic approach for the 32,000 DWT bulk carrier which uses frequent small-scale bunkering. Experimental study on electrical properties of deep-sea cobalt-rich crusts The electrical properties of cobalt-rich crusts are selected as the research object, which are necessary to study breaking cobalt-rich crusts by high voltage. A special testing platform is designed for testing this mineral to determine dielectric constant, breaking voltage and breaking electric field strength. Different specimens of cobalt-rich crusts and substrate rocks are tested on this platform. The experimental results are calculated and analysed, and the conclusions are as follows: Under different frequencies of inputting voltage, the relative dielectric constant of cobalt-rich crust is not a constant, usually the lower frequency and the larger dielectric constant. When the voltage frequency is 1 or 4 MHz, the values of dielectric constant of dry and wet specimens are close, and when the voltage frequency is within 1 - 100 kHz, the values of dielectric constant of dry and wet specimens have huge differences. From the calculated results, the breaking and fluctuating electric field strength of cobalt-rich crusts is much smaller than that of substrate rocks, which is the theoretical basis for breaking deep-sea cobalt-rich crusts by high voltage. Analysis of the operational ship energy efficiency considering navigation environmental impacts In this study, a remote collecting system for ship operation energy efficiency was designed by using a tourist ship. According to measured data, the navigation environment factors of the ship were analysed statistically. The Gray correlation analysis was used to study the correlation degree among wind, water depth, water velocity and ship operation energy efficiency. On this basis, the Wuhan - Nanjing section of Yangtze River (upstream and downstream) was selected to analyse the law of wind speed effect on the energy efficiency of ship operation. The data analysis demonstrates that in different navigation environments, the ship energy efficiency can be significantly improved by optimising the main engine speed. Pilotage services in Turkey; key issues and ideal pilotage Previous studies have identified fundamental problems of pilotage organisations: pressure on commerce, improper working conditions, increasing traffic volume, and vessel size. Questioning the needs of pilots, the most important actors in pilotage services, is a key issue to ensure the safety of navigation in restricted waterways. This study, which conducts questionnaires and interviews with 71 pilots, reveals the structure of pilotage organisations and the profiles of pilots in Turkey. The survey examines pilot training infrastructure, professional experience, working conditions, tug services, pilot boats, accommodation facilities, and opportunities from operational, economic, and environmental constraints. The results indicate positive and negative aspects of existing pilotage organisations in Turkey in order to develop an ideal organisational model for pilotage in Turkey. Using physiological signals to measure operator's mental workload in shipping - an engine room simulator study Mental workload (MWL) is one of core elements of human factor construct reflecting arousal level, and its optimisation is crucial to maintain favourable operator functional state. However, sensible, reliable, and diagnostic measurement of MWL is essential for applications of adaptive aiding system design, usability testing, and seafarers' training. To develop robust MWL measures, 10 participants voluntarily participated in a simulator-based experiment study. During this study, the participants carried out standard four-level calibration tasks and simulated four-level maritime operation tasks, their heart rate and electroencephalogram (EEG) were continuously measured using a heart rate sensor and an ambulatory EEG device, which includes an accelerometer to distinguish signal corruption epochs induced by body movement artefact. After each task, NASA Task Load Index was collected as subjective measurement. One-way analysis of variance was used to test the sensitivity of MWL measures and Pearson's correlation coefficients were calculated. The significantly sensitive indices for n-back task and simulator-based maritime operation task were different, supporting the limited cognitive resource pool theory. EEG features showed higher sensitivity than heart-rate-related measures. An experimental evaluation of the effects of sea depth, wave energy converter's draft and position of centre of gravity on the performance of a point absorber wave energy converter In this paper, the effects of sea depth, wave energy converter's (WEC's) draft and position of centre of gravity on the performance of a point absorber WEC are examined using an experimental study. By considering the general cubic shape for the WEC, the tests are implemented in the experimental wave tank. In this context, by creating waves with different amplitudes and periods, evaluation is conducted in three different magnitudes for each parameter. WEC oscillations in the directions of heave, surge and pitch are measured and compared. Furthermore, an optimised model for building a full-scale WEC for operating in the Caspian Sea environment is studied and proposed. Experimental test results show that the model with least draft and least distance between the position of centre of gravity and still water level and also installed in the deepest sea is the best model to be installed in the Caspian Sea. Study on fuzzy GIS for navigation safety of fishing boats Maritime ship collisions involving fishing boats have often been reported, and the major factors have been human error and unpredictable conditions. This paper aims to create a set of collision-avoidance measures for operating fishing boats or a vessel traffic service by way of time and space management. It proposes a novel safe collision-avoidance safeguarding ring for navigational safety between operating fishing boats and merchant ships with a dummy automatic identification system (AIS) to enable fishing boats to establish a dynamic state when approaching ships under operation (fishing). The approach method utilises the maximum relative speed, sea state, and escape time as three linguistic input variables for fuzzy logic control application. After the radius of the safeguarding ring is obtained and shown on a marine geographic information systems (MGIS) platform, the danger index of two ships is evaluated based on the area difference calculated through the MGIS. The software was installed in the dummy AIS receivers of fishing boats for experimentation in a Taiwan harbour. It was able to enhance the collision-avoidance ability of the operating fishing boats effectively and reduce marine disasters involving fishing boats. Conceptual design of a 5 MW OTEC power plant in the Oman Sea Developing ocean energy extraction, including ocean thermal energy conversion (OTEC), has been of interest to researchers for many decades. OTEC is a free fuel technology and could be used as a baseline power generation. These advantages plus the engineering challenge have resulted in researchers striving to design and construct prototype devices. Making use of worldwide experience, all sections of a conceptual design including site selection, technical specifications and cost estimation were carried out for an Iranian OTEC power plant. A 5 MW closed cycle floating plant with an annual average temperature difference of 22°C was chosen at a 33 km distance from Chabahar harbour. Deep seawater would be extracted from 1000 m depth and would result in 3.52 MW of net power. According to cost calculations, the levelised cost of electricity of the plant has been estimated to be approximately 0.117 $/kWh, which is an acceptable level compared to other renewables. The conceptual OTEC design presented in this paper demonstrates a thermal potential in the Oman Sea which could assist with meeting the power demand for the southern coast of Iran. Nonlinear robust control of marine diesel engine Diesel engines are widely used for the propulsion of marine vehicles and are exposed to uncertain working environment. In this paper, a robust nonlinear controller is proposed for a marine diesel engine that employees sliding mode control theory. The robust controller aims to maintain the desired diesel engine speed performance under harsh sea environment. The sliding surface has been carefully chosen that minimises the error in both angular velocity and acceleration. Robust control algorithm development and its tuning are also discussed. The performance of the proposed nonlinear robust controller is investigated thoroughly and is compared with a classical Proportional - Integral - Differentiation controller with integral windup scheme. Simulation results show that the proposed super-twisting-algorithm-based sliding mode controller can effectively improve the speed performance of the marine diesel engine in transient and steady operating conditions. Multi-constrained fuzzy intelligent control for uncertain discrete systems with complex noises: an application to ship steering systems In this paper, a performance constrained fuzzy controller design is investigated for the nonlinear uncertain discrete-time ship steering system with complex noises. The performance constraints considered in this paper include individual state variance constraint and passivity constraint. The nonlinear discrete-time ship steering system considered in this paper is represented by a discrete-time Takagi - Sugeno fuzzy model with perturbations, additional noise and multiplicative noise. According to this complex nonlinear discrete-time stochastic system, a fuzzy controller design approach is investigated based on Lyapunov theory, passivity theory and covariance control theory. Some sufficient conditions are derived to satisfy stability constraint, individual state variance constraint and passivity constraint, simultaneously. These sufficient conditions are constructed as linear matrix inequality forms that can be solved by the convex optimal programming algorithm. Finally, some simulation results are provided to verify the validity and applicability of the proposed fuzzy control approach with multiple performance constraints. Study of a solitary wave interacting with a surface piercing square cylinder using a three-dimensional fully nonlinear model with grid-refinement technique on surface layers This study introduces a three-dimensional fully nonlinear wave model, which solves the Laplace equation of the velocity potentials. A base-grid system is adopted in view of a transient curvilinear coordinate transformation along the vertical direction, while the horizontal coordinates remain uniformly distributed. To capture the variations of the free surface more accurately and effectively, the necessary denser grids can be specified within the vertically divided multi-plane layers near the free surface. In addition, fine grids can be arranged around the structures included in the flow domain. The model is first validated for the case of a solitary wave propagating along a narrow uniform-depth channel to investigate the level of solution improvement by the grid-refining technique. Smoothing is applied to remove the weak-fluctuation waves produced by the interpolation. In addition, the initial condition of the solitary wave is adjusted from the long-wave analytical solution to fit the presented fully nonlinear wave model. The results indicate that both efficiency and accuracy increase greatly by applying the fully nonlinear wave model as compared with a model using a uniformly non-refined grid system. Second, this model is applied to study a coastal problem related to a solitary wave diffracted by a surface-piercing square cylinder with or without a separating distance between the seabed and the bottom of the structure. An unmanned marine vehicle thruster fault diagnosis scheme based on OFNDA In recent years, there has been a growing interest in the use of fault analysis techniques in unmanned marine vehicles (UMVs) owing to their significant impact on marine operations. This study presents a novel approach to the diagnosis of unbalanced load (blades damage) faults in an electric thruster motor in UMV propulsion systems based on orthogonal fuzzy neighbourhood discriminative analysis for feature dimensionality reduction. The diagnosis approach is based on the use of discrete wavelet transforms as a feature extraction tool and the optimal number of mother wavelet function and levels of resolution by analysing the vibration and current signals. As a result of analysis and comparisons, the Deubechies 12 (db12) wavelet and level 8 were chosen. A dynamic recurrent neural network was chosen for fault classification and level of fault severity prediction was implemented. Four faulty conditions were analysed under laboratory conditions and these were recreated by damaging the blades of a motor. The results obtained from the simulation demonstrate the effectiveness and reliability of the proposed methodology in classifying the different faults with greater speed and accuracy compared to existing methods. An optimisation design method for cryogenic pipe support layout of LNG-Powered ships For LNG-Powered ships, cryogenic pipe bears ultralow temperature load; the stress of LNG pipe is very sensitive to the layout of pipe supports. In order to improve the safety of LNG pipe under complex loads, an optimisation design method for cryogenic pipe support layout (ODMCPSL) of LNG-Powered ships is proposed in this paper. The objective of ODMCPSL is to deduce the structure stress, and both the pipe supports number and their positions are taken as variables. An improved genetic algorithm named Genetic Assimilation - Genetic Algorithm (GA - GA) is adopted in ODMCPSL, in which a new factor named magnetic factor (MF) is introduced in mutation operation and the number of pipe supports is optimised by MF. In the algorithm, the number and positions of the pipe supports can be optimised at the same time. ODMCPSL is applied to a dual-fuel harbour tug, and the maximum stress of the LNG pipe is significantly decreased by the optimisation of pipe supports. The engineering application example proves that ODMCPSL has high convergence and efficiency. Game Control of maritime objects The paper presents the application of the theory of differential games and multi-stage static and multi-matrix static games for the automation of the process of controlling maritime objects - merchant ships, ships or unmanned vessels. It has been shown that through the appropriate formulation of the plotter quality control, consisting of an integrated and final pay out, it is possible to implement collision and collision games as well as non-operational and cooperative games. On the example of safe control of one's own ship in situations of passing with many ships encountered at sea, algorithms of defining a safe trajectory of a ship supporting the manoeuvring decision of the navigator were presented. Considerations are illustrated by examples of computer simulation algorithms in Matlab/Simulink software for safe ship trajectories in real-time maritime situations. Vibration diagnostics of common rail injectors This article presents the possibility of evaluating the technical condition of common rail fuel systems injectors using the vibration method. The other most successful methods of diagnosing the common rail injectors were described. Marine diesel engine common rail injection installations had been characterised. Results of tests obtained during vibration measurements were presented as well. The results confirm the possibility of determining the technical condition of modern diesel engines injectors without the need of engine stopping. The results are especially important for marine engines without indicating cocks or thermoelements allowing measurements of exhaust gases temperature on each cylinder. Concept of `Sail by Wire' controller for a ship's propulsion system from an unmanned ship perspective The intention of this paper is to present the idea of total ship control called the Supervisory `Sail by Wire' Control System. As a part of it, there is an optimal controller of the ship's propulsion plant presented in this paper. The paper focuses on the integration of four control subsystems of ships: navigation, propulsion, auxiliary and deck machinery systems, hull structures and cargo handling equipment. Such a controller can be build also on the base of conventional, fuzzy, or neural principles. Critical situations are detected, warned in advance is given about them, and they are reported to the ship's operator. The analytical results and recommendations as well as real vessel conditions will be recorded and post-processed. The intelligent integration of acquired data from subsystems will ensure that the control commands and actions compromise the multi-effects of a vessel's dynamics and operation. While performing accurate regulation and governing functions, this control system should have the capability of identifying and compensating failures during the ship's operation. The `Sail by Wire' Supervisory Controller delivers a comprehensive study and a new idea to common European strategies and decisions for tracking issues in shipping in critical situations. Modelling of the operating process in a marine diesel engine The paper presents elements of a mathematical model of a marine diesel engine. The purpose of developing the model is to enable diagnostics of fuel supply and charge exchange system in a marine engine (simulation diagnostics). The authors have assumed the option allowing to modify geometric parameters of the crank-piston system, charge exchange, heat exchange and chemical composition of fuel. This option offers simulation of selected defects in an engine. The focus of the paper is both the process of model development and its implementation in an object-oriented programming language. A fractional model of supercapacitors for use in energy storage systems of next-generation shipboard electrical networks The reduction of energy consumption and carbon dioxide emission has forced designers and manufacturers to develop new, more efficient ship propulsion systems. Traditional mechanical systems are replaced by hybrid or completely electric ones. In addition to the combustion, the electromechanical sources are considered. But in such small electric grids, in order to reduce the impact of load fluctuations on the power system, energy storage devices are proposed. Such devices are able to supply a peak power. As examples the batteries, flywheels, but primarily supercapacitors can be considered. Unfortunately, a very low voltage causes supercapacitors to work in stacks. This paper presents the results of analysis of the influence of changes in selected supercapacitor parameters changes on its terminal voltage during the cycles of charging and discharging. The model parameters are estimated based on the test measurements, while the differ-integral calculus of fractional order is used to derive the model output. It turns out that such an approach provides a very good fit for model responses even for its very simple structures. In addition, a very precise mathematical model allows to accurately estimate an amount of stored energy and estimate the supercapacitor reliability. Radio over fibre technology for shipboard antenna links The article shows the use of radio over fibre (RoF) analogue fibre links in numerous shipboard radio communication applications, which creates a demanding competition for coaxial lines and waveguides. The structure and properties of fibre optic links used for transporting microwaves in maritime engineering have been described. The article presents the use of RoF technology in satellite communications and fibre optic antenna link for global positioning system. The RoF link parameters essential for obtaining proper quality parameters of signal transmission have been determined. Moreover, the parameters of the RoF link have also been measured and evaluated in terms of the needs and requirements of ship installations. Using fuzzy logic expert system for the estimation of the ARS-880 radar range The article presents a fuzzy expert system designed to determine the possible radar range of the ARS-880. The aim of the project is to develop a fuzzy controller which will allow generating the range extent depending on the sea conditions and the radar cross-section of the tracked object. The project uses Matlab and Fuzzy Logic Toolbox software. The authors present the performance of the system based on 18 samples for research. They perform an automatic simulation, which takes into account various weather conditions at sea as well as different size floating objects. Hardware and low-level control of biomimetic underwater vehicle designed to perform ISR tasks This paper conveys some of the results of a research project involving the design and tests of biomimetic underwater vehicle which is intended for implementation of intelligence, surveillance and reconnaissance tasks. The research presented in this paper is based on the theoretical and practical study. All major subsystems of the robot are specified and briefly described. The mechanical design is based on modular structure. Two types of modules, dry and wet, are specified. The major parts of the tail module were checked with the use of the finite element method. Moreover, hardware and software parts of robot's control subsystem and Simulink model of static depth regulator are briefly presented. The final part presents main parameters and basic functionality of the built biomimetic underwater vehicle. Dynamics of shaft lines of the landing ships The paper presents an analysis of the dynamics of the shaft line of landing ship manoeuvre during grounding at a landing area. The work characterises the specific features of shallow water interactions on the propeller operation. There is a characteristic structure of the ship's propulsion system. Potential operational hazards arising from the problem of grounding and descent from shore shoals of a landing ship have also been identified. A model of sandy water interaction on propellers' torque was proposed for the dynamic calculations of shafts' loads. The results of the simulation of torsional vibrations on the shoal area are presented as well. MPTCP protocol misbehaviour in high-speed, uncongested network Multipath transmission control protocol (MPTCP) protocol is an extension of the well-known TCP transport protocol that enables transmission over multiple paths. The MPTCP is an answer to multihoming, where a single host (here an MPTCP sender or receiver) is connected to multiple networks, using different network interfaces (network cards). The paper is focused on the problem of misbehaviour of the MPTCP protocol in high-speed, uncongested networks. The problem was discussed on the basis of experiments that were carried out in a private test cloud. In the experiments, a large file of 10 GB was transmitted between components of the cloud. Transmission was performed using MPTCP protocol conveying data through high-speed (900 Mbps) 802.11ac links. Results show that the reduced aggressiveness of the MPTCP caused problems with utilisation of unloaded paths and problems with fair resource sharing with the TCP protocol. The central server of the Border Guard's distributed multimedia system for monitoring and visualisation of ongoing and archival events The paper presents the architecture and functionalities of the central server (CENTER) of the distributed system for the Polish Border Guard (BG) for monitoring maritime areas. The overall system has been extended to incorporate, apart from map data, also different multimedia elements such as video from cameras or audio from telephone connections operated by BG units. This requires new system elements: Archive Servers for storing new types of data and Events Visualization Post for the presentation of new media along with the presentation of ongoing or archival tactical situations on maps. The paper focuses on new functionalities of the CENTER and integration of new elements with the rest of the system. Local image features matching for real-time seabed tracking applications Real-time seabed tracking applications play an important role in underwater systems. A lot of them use computer vision for servoing, positioning, navigation, odometry and simultaneous localisation and mapping. They are mostly based on local image features, therefore feature detection, description and matching are crucial for their efficient operations. The aim of this study was to investigate the most popular feature detection and description algorithms such as SIFT, SURF, FAST, STAR, HARRIS, ORB, BRISK and FREAK. Additionally, the image correction technique was presented and image enhancement methods were analysed in order to increase efficiency of image features matching. The matching algorithm was based on the homography matrix and random sample consensus technique. Our results indicate that the combination of the histogram equalisation technique and ORB detector and descriptor enables real-time seabed tracking with sufficient efficiency. Influence of cathode stoichiometry on operation of PEM fuel cells' stack supplied with pure oxygen As a part of ongoing research into the use of fuel cells to power underwater platforms, studies have been conducted to investigate the effect of change in oxygen flow through the proton exchange membrane fuel cell (PEM-FC) cathode channel, to develop a method for controlling this flow. Due to the lack of specific information or studies on the effect of pure oxygen flow through the PEM-FCs cathode channel on the correctness of its work, a series of experiments have been carried out within the framework of in-house research to determine this dependence. The experiment was conducted in series, during which the only change in the operating parameter of the PEM-FC stack was the stoichiometric flow rate of oxygen supplied to the cathode channel of the fuel cells' stack. The research shows that for the PEM-FCs stack fed with pure oxygen, the cathode stoichiometry can be kept at a much lower level comparing to the systems supplied with air. It has been found that the fuel cells' stack fed with pure oxygen will perform well in steady state while reducing the cathode stoichiometry by as much as 40% compared to systems utilising oxidant from atmospheric air. Analysis of propulsion systems of unmanned aerial vehicles One of the most important tasks in the phase of making a design of an unmanned aerial vehicle (UAV) is to properly select its propulsion system. The primary initial data should include application and requirements for a UAV. The type of a propulsion system is directly linked with tasks and aims which the designed unit will execute. Moreover, it is necessary to take into account the required payload, essential for carrying equipment, including munitions. The type of propulsion denotes selecting such a system, among a wide range of the available ones, which allows the best exploitation of the airframe characteristics as well as generating proper thrust in order to comply with all the assumed requirements. This study analyses various propulsion systems currently used in UAVs, paying particular attention to their characteristics which are essential to conduct a specific mission, including geological and photogrammetric research missions. The most common task of UAVs is reconnaissance, and its most significant parameters are indurance, speed, flight altitude and payload taken on board. All of these are largely linked to the type of used propulsion. In conclusion, the above analysis indicates that a future propulsion system for an UAV should be designed on the basis of a rotary piston combustion engine or a flat engine (boxer type). It should be digitally controlled, so that the operator receives information about the engine's working parameters, in real time. Take-off and landing magnetic system for UAV carriers Aircrafts play a main role in modern naval wars. The navy increasingly uses unmanned aerial vehicles (UAVs). It can be assumed that in the future navies can largely resist the use of airplanes and helicopters. This will reduce the risk of crew loss and will also reduce the cost of having an airborne device on board. A new class of war ship has come up. This kind of ship is equipped with the automatic take-off and landing system as well as planning, navigation, and supervising of an UAV system. The article presents the magnetic take-off and landing system for UAV carriers. The system was elaborated as a demonstrator of technology in the GABRIEL project. Method of generating self-supporting linear antennas for marine radiocommunication systems The aim of this paper is to present a new method for generating self-supporting structures of a linear wire antenna. Antenna structure resembles monopole and consists of several sections connected in series. This type of antenna is suitable for marine radiocommunication systems. The generation method proposed in this paper is called TGA (tuned genetic algorithm) and it is based on two algorithms. First one is a typical genetic algorithm, while the second one allows us to modify antenna structure founded by the first algorithm. Mentioned modification involves rotating antenna sections. In this way, antenna parameters such as VSWR can be improved. What is more, with TGA, antenna susceptibility to changes in its geometry can be determined, which is helpful in antenna manufacture. To predict the performance of antennas, we used FEKO. Robotic system for off-shore infrastructure monitoring In this paper, a novel modular reconfigurable mobile inspection robotic system is presented. The system consists of two underwater robotic platforms, including a tracked mobile robot for moving on rough terrain and a remotely operated vehicle robot for free underwater operation. Applications of the system include structural monitoring of off-shore infrastructure made of steel, concrete or other materials. The system is equipped with a 3D sonar, high-definition cameras and laser sensors to navigate underwater and locate structural defects. Mathematical models of the tracked robot are derived for operation underwater. Operation and verification of the ROV vision system is shown. The system sensing equipment is capable of performing simultaneous localisation and mapping tasks, thus an algorithm for navigation and localisation of defects based on sensor fusion is proposed and discussed. Evaluation of numerical modelling application for the crash test planning of the catastrophic Flight Data Recorder Catastrophic Flight Data Recorders (FDRs), the so-called black boxes, belong to the on-board equipment, the task of which is not only to operate properly while recording selected parameters under all flight conditions but also to protect the recorded data against the loss in the aircraft crash conditions. The requirements related to catastrophic resistance of FDRs are defined in the European standard EuroCAE ED-112 and the Polish standard NO-16-A200:2015. One of the requirements includes the catastrophic recorder's resistance to the acceleration overload of 3400 g. The compliance with the standard is checked during crash tests. In this paper, a procedure of planning and carrying out the mentioned tests for S2-3a-K protected unit of S2-3a catastrophic FDR manufactured by the Air Force Institute of Technology was described. It is an iterative procedure involving consecutive experimental tests and both analytical and numerical calculations. The experimental tests were performed using a DPZ-250 pneumatic gun. The testing probe was shot at the target of sand deposits of modified geometry and mass characteristics. Effectiveness of the developed procedure was shown by comparing the experimental results, namely recordings of the probe deceleration captured with a fast camera, with results of numerical simulation of the probe terminal ballistic. Implementation of the assessment method of the lead - acid battery electrical capacity in submarines Lead - acid batteries are still being used in conventional submarines as the primary source of power. Due to their properties, lead - acid batteries require a lot of attention during their operation, including charging according to fairly rigorous rules. Because of their special operating conditions in submarines, it is not possible to maintain optimal conditions for such batteries. The battery's integrated management system allows for remote and online monitoring. However, it gives only general information about how much energy remains in the battery for use in operating conditions. The submarine's operational capabilities, such as range and operation time, depend to a great extent on battery capacity, so information of its condition and status is very important. The proposed method of evaluating the capacity of lead - acid batteries allows them to be processed during their exploitation, without the need for a time-consuming and expensive capacity test. The measurements needed to implement the proposed algorithm are already implemented in the existing battery monitoring system - no additional equipment is required. Based on the results of the proposed method, it will be possible to easily determine how much energy is available to the submarine, which is crucial in planning combat missions at the sea. Quad-Tiltrotor - modelling and control Aircrafts with VTOL capabilities announced a new era in marine aviation. Their aerodrome requirements, which are as low as for helicopters, make them excellent candidates for many missions required by the navy, such as transport or SAR. To successfully exploit their potential many engineering challenges must be overcome. One of the areas with major problems to be solved is connected to the development of a robust control system. The presented article deals with this topic. It describes the methods and concepts, which can be used to model and control these kinds of aircraft. Results of impulse response measurements in real conditions Article presents the results of researches conducted in real conditions concerning the measurement of impulse response of the hydroacoustic channel. For measurements, signals modulated by pseudo-random sequence were used. Presented examples of impulse responses were determined in the shallow water. Researches were conducted on the lake under static conditions. The parameters of excitation current of ship synchronous generator as the diagnostic symptoms of the propelling IC engine For ships, the total loss of electrical power, `blackout', is a critical failure. The failure of the marine power supply system not only results in loss of ability to perform tasks, but also poses a threat to human life and often to the environment. No less serious emergency is `partial blackout', i.e. the loss of electricity in parts of the electrical system. Designers are introducing technical and system solutions to minimise the likelihood of power failure and to support its recovery in a timely manner. One of the methods to minimise the likelihood of power failure is to identify and then monitor diagnostic symptoms that are susceptible to approaching the failure of key components of the power system. This article discusses the use of recording and analysis of changes in voltages and excitation currents in parallel synchronous generators to determine the threshold above which their parallel operation can lead to an emergency power failure. Automated control system with process of the use of ship's power station This paper describes the automated control system of the exploitation process of a ship's power station. The solution, which incorporates the use of local automation systems to control power sets, the use of power grid and several levels of control units to monitor and control the operation of the entire power plant, ensures high reliability of power supply to marine electrical devices. The implemented solutions enable the introduction of new logic elements into the system without the need for system modifications. Advanced reporting is the basis for preparation of navigating ship documentation, and the logic behind the created solution ensures use as a part of the command system. Assumptions of the optimisation of a ship power station use and application in a simulator The article presents assumptions of optimising the process of a ship's power station use and a description of a ship power station simulator with generating sets driven by SW400 engines. Selected issues of designing and testing of a HALE-class unmanned aircraft This work presents the research connected with designing an unmanned aerial vehicle with long mission endurance. Specification of HALE aircraft as well as the possible operating scenario of HAPS system is presented. Issues connected with operating the aerial vehicle at stratospheric altitude are discussed. A mission profile itself is also presented. A construction project of the aerial vehicle and its possible equipment is suggested. Regarding the fact that the aircraft is to perform high-altitude flights, it faces airfoil operation at very low Reynolds numbers. That is why, research concerning quality operation airfoil of low drag airfoil at a stratospheric altitude was carried out. For this reason, the influence of flights at subcritical Reynolds numbers on aerodynamic characteristics is presented. A series of simulation studies and model-based research in a hydrodynamic tunnel were carried out. It follows from the conducted numerical analysis that the decrease in aerodynamic excellence of the airfoil operating at 15 km altitude is small in relation to the airfoil operating at the height of 0 m. The research carried out in the aerodynamic tunnel confirms the efficiency of the airfoil operation at low Reynolds numbers. The work will be continued in order to check the influence of a turbulator on the shift of the range of Reynolds number, and consequently on the improvement of the aerodynamic quality of the airfoil. Algorithms for passive detection of moving vessels in marine environment In this paper, an investigation on the development of a low-cost passive hydroacoustic system for passive detection of moving vessels to counteract possible collision with an unmanned underwater vehicle is presented. The main goal of this paper is to determine if moving vessels generating hydroacoustic signals/signature are present in the space being searched, and if so, to determine the time delay ΔT between two signals (V1(t) and V2(t)) and consequently to estimate the bearing on the source of the hydroacoustic signals, for example, screw propellers of a moving vessel. The acoustic signals V1(t) and V2(t) have been recorded by a two hydrophones mounted in an unmanned underwater vehicle. In practice, signals V1(t) and V2(t) are heavily corrupted by the additional noise. The noise comes from surrounding environment and from the measurement system errors. Moreover, real signals are often unsteady (nonstationary) and random (stochastic). That is why the different methods have been taken under consideration and the received results have been compared. An analysis has been made for time and frequency domain as well. Due to the planned application of the obstacles detection system in the unmanned underwater vehicle, the algorithm had to be feasible for implementation in digital signal processor. Proposition of the vibroacoustic diagnostic methodology of testing toothed gears of marine drives The proposal of the Authors of the hereby paper is based on measuring vibrations of the gearbox housing and on the application of the generalised coherence analysis. The diagnostic model proposed by Klekot [1992. Simplified kinematic model of a spur gear with involute teeth elaborated for analysis of gear vibroactivity. Mach Vib. 4(1):261 - 267] was accepted as the standard, which is relatively simple, but takes into account the influence of the mesh geometry evolution during its operations on this mesh dynamics. In order to identify the proposed model, the authors performed the active diagnostic and model experiments for the toothed wheels pair. As a result, you can see how selective and practically insensitive to the interferences the measure of the evolution of gear tooth damage can be. Hybrid power supply system model for offshore floating platforms This article consists of research results regarding the use of energy profiles, as a tool for the configuration of offshore floating platforms' hybrid power supply system. Given solution supports renewables, as a platform's exploitation economy improvement. As a main current source in the proposed system, a fuel cell with high energy density to mass ratio (Wh/kg) is provided; however, it has disadvantages regarding its dynamic response during the load changes. The proposed solution minimises the disadvantages of the fuel cell by designing a hybrid system consisting few energy sources and storages with a logic system to control its elements. In order to optimise the hybrid system's configurations the author's concept has been applied. This concept models the set of energetic and dynamic characteristics called the energy profile in the considered floating platform's electric loads. Use of energy profiles in the system's configuration process allows minimisation of its size with constant power availability in all the floating platform's exploitation states. Asynchronous phase-location system This paper presents concept and implementation of digital phase-location system, designed as a navigational aid for marine applications. The main feature of the proposed system is the ability to work in both synchronous mode, with one master station and set of slave stations synchronised with master, and in asynchronous mode with independent clocking of all stations. Data transmission in the hydroacoustic channel - experimental researches The paper presents the results of experimental researches of data transmission in water using Direct Sequence Spread Spectrum technique and the RAKE receiver. The signal with the QPSK modulation was spread by factor of 4, 8 or 16 chips. Researches were made to verify that spread spectrum technique can be used to underwater voice communications. Static studies were conducted in a small laboratory water tank and in a harbour. Propagation properties were evaluated based on the estimated impulse response of the hydroacoustic channel. Polish USV `EDREDON' and non-European USV: a comparative sketch Comparative sketch of polish Unmanned Surface Vehicle (USV) `EDREDON' and non-European USV constructions are presented. Review of outside Europe USV is given. USV classification and simplified comparison of EDREDON with similar non-European USV are listed. The comparison of non-European USV's with USV `EDREDON' was aimed to show how the product developed in Poland is presented on the background of USV world leaders. Passive fault tolerant control allocation for small unmanned underwater vehicle The paper presents methods of control allocation in a multi-thruster propulsion system of a small torpedo-shaped underwater vehicle. It concentrates on finding an optimal thrust distribution among thrusters for demanded values of propulsive forces and moments. The approach is directed towards minimisation of energy expenditures necessary to obtain required control. Special attention is paid to the unconstrained thrust allocation for both a faultless and faulty work of the propulsion system. The developed algorithms are designed to be modular, interchangeable and highly computationally efficient. Illustrative examples are inserted to demonstrate correctness and simplicity of proposed thrust allocation methods. Shipboard operating and maintenance procedures and the knowledge gap This paper advocates the use of shipboard procedures as a strategy for bridging the knowledge gap. It also recognises that an over-reliance reflects a failure to equip seafarers with the appropriate prior knowledge. Maritime accident reports frequently cite poor procedures as contributing factors towards maritime accidents. This may suggest a nexus among technical systems, seafarer competence, shipboard procedures and accidents at sea. It is suggested in this paper that a holistic approach is needed to ensure that the approval of technical systems, seafarer training and written procedures are inextricably linked. Shipboard written procedures are described in terms of a negotiation of meaning between writer and seafarer, where successful execution of a shipboard task requires an accurate interpretation, by the seafarer, of the writer's intentions. It is suggested that this may only be achieved by ensuring that seafarer prior knowledge is sufficient, writer's anticipation of this knowledge is accurate and communication of new knowledge is explicit. This paper does not offer specific advice on the design of shipboard procedures; however, the process of acquiring a mental action plan is discussed. The factors that support or impede this acquisition are described and supported by a number of accident reports and academic studies. Study on the contactless data transmission for ocean underwater vertical sensor structure Ocean underwater vertical sensor structure is a kind of modern marine observation platform. But most of the current ocean sensors work independently, and it lacks a suitable method to assemble them in a control system. Having good expansibility and convenient structure, the ocean buoy platform with contactless data transmission technology is possible to organise a long-term, real-time and autonomous working observation network. The contactless data transmission model with a two-stage electromagnetic coupler is investigated. A prototype of communication system using the differential phase shift keying method is constructed. Due to attenuation in seawater, it is difficult to realise the long distance and low bit error rate transmission. To solve this problem, an equivalent two-terminal network model with compensation capacitor is proposed to make this contactless transmission system work at the resonance frequency. In 1000 m deep sea experiment, the sensors equipped with this vertical structure based on buoy platform show a perfect similarity of conductivity points like a commercial conductivity, temperature and depth device. It can be predicted that this underwater contactless data transmission method will be widely used to construct a multipoint distributed marine data network, and it can bring large scientific and economic values in deep sea exploration. Analysis of marine solar power trials on Blue Star Delos In October 2014 the high speed car and passenger ferry Blue Star Delos was fitted with a marine solar power system. This was done as part of a project to study the use of renewable energy on large ships. In May 2015 while the ship operated in the Aegean Sea, the performance of system was checked and a range of data collected. This paper is focused on the analysis of these data and the evaluation of two days of system trials. Corrosion assessment of infrastructure assets in coastal seas The seas play an essential role for the peoples living on their coastal regions, since the marine infrastructure is located in the coasts. Seawater is a corrosive environment that affects infrastructure particularly in polluted seawater. Corrosion and pollution are pernicious chemical, physical processes that impair the quality of the environment and the durability of the marine structures and materials. They are aggravated by the discharge into the sea coast of municipal, industrial and agricultural effluents, which contain and produce toxic and highly corrosive components by biological and chemical degradation. Reinforced concrete and carbon steel are the main engineering materials used for the construction of marine installations and equipment but other metals and alloys: aluminium; copper, stainless steels are applied, too. Laboratory and field corrosion tests in seawater were carried out applying gravimetric, electrochemical and surface examination methods, based on American Society of Testing and Materials (ASTM) and National Association of Corrosion Engineers (NACE) standards. This work is the result of a cooperation between academic institutions in Mexico and Israel. The data generated advance the management of sea corrosion prevention and mitigation, and provide a guide for marine infrastructure maintenance and corrosion control. Several cases of corrosion in the sea coasts based on the authors experience and knowledge are presented. Co-operative control of a team of autonomous underwater vehicles in an obstacle-rich environment This paper presents the cooperative control of a team of autonomous underwater vehicles (AUVs) in the presence of obstacles under environmental disturbance. A leader - follower formation control based on optimisation algorithms using communication topology is designed to navigate towards the target in the presence of obstacles. CLONAL selection optimisation algorithm may be employed as team controller by providing the optimal position to hierarchical control strategy for each AUV incorporating learning, memory and affinity. But unfortunately CLONAL selection optimisation algorithm fails to avoid obstacles because some of the lymphocytes become memory cells of the system. The above prematurity may be overcome by using artificial potential fields and ant colony optimisation technique combined with CLONAL selection optimisation algorithm. The efficacy of the proposed optimisation is verified through MATLAB simulation and the result confirmed the robustness and proficiency of proposed technique over CLONAL selection optimisation algorithm. Wide-ranging radar simulation data generation method based on multi-scale electronic charts in a maritime simulator In maritime simulators there are problems of discontinuity and saltation on radar simulation images caused by the limited range of radar simulation data generated from a single electronic chart. Furthermore, this kind of radar simulation is overlapping, discontinuous and redundant, and thus could lead to simulation error and training interruption. To address this problem, a method is presented to generate a wide range of radar simulation data based on a multi-scale chart data. The method uses automatic splicing on same-scale maps and variable-scale visualisation splicing on multi-scale maps. In this paper, the S-57 standard electronic navigational charts are filtered from the S-57 database and then radar simulation data are generated based on charts of different scales. From clipping, sorting, deletion, combination and matching, a wide range of continuous radar simulation data are generated. Taking the waters around Ningbo Port, China as an example, 27 chart sections are filtered out from S-57 database, with a longitude range from 121°33' to 122°56.8' and a latitude range from 29°3' to 30°21.8'. The wide-range and continuous radar simulation data are generated for a maritime simulator. The results validate this method as being highly accurate and able to meet the practical needs of maritime simulation. Frequency control during transients in offshore wind parks using battery energy storage systems The main objective of this paper is to study the implementation of a battery energy storage system (BESS) in order to improve the dynamic behaviour of an offshore wind farm. Special attention is given to the frequency excursions limitation in the sequence of a fault on the onshore grid. This goal is accomplished by setting up a case study in which an offshore wind farm is connected to an onshore grid through a high voltage direct current line commutated converter (HVDC-LCC) link. The Power System Simulator for Engineering (PSS/E) simulation tool is used to implement the whole system, including the onshore and offshore grids and the load flow and dynamic models of the conventional generators, the wind turbine generators (WTGs), the HVDC-LCC link and the BESS. With the BESS out of service, some protective actions of the DC link lead to the uncontrolled frequency increase of the WTG, which are consequently tripped out. The BESS proves to be essential, as it allows the offshore wind farm to keep running by maintaining the frequency of the WTG at acceptable levels and therefore bringing the entire system to a steady state operating point. A novel hybrid MCDM model based on fuzzy AHP and fuzzy TOPSIS for the most affected gas turbine component selection by the failures Gas turbines are usually the core elements of numerous mechanical systems. Subsequent to the advancement of high efficiency and clean energy, gas turbine has a significantly growing role in different areas, such as aviation and marine propulsion systems, electric power stations, and natural gas and petroleum transportation. However, gas turbines are decisive in operating many industrial plants, and their accompanying maintenance costs are likely to be exceedingly high. A number of failures in the gas turbine components have been detailed and discussed in this paper, with a view to thwarting future breakdowns by suggesting specification for notable maintenance and utilisation of gas turbine components. This study puts forward Fuzzy Analytic Hierarchy Process (AHP) and Technique for Order Preference by Similarity to Ideal Solution (TOPSIS) methods applied in failure detection of gas turbine components. This paper attempts to handle gas turbine components, which contain hydraulic - pneumatic equipment, electronic control equipment, and bearing equipment. Thanks to the assessment of specialists, heavily influenced by failures, gas turbine components have been decided. The results have proved the bearing equipment to have been the most effective alternative, as being ensued by hydraulic - pneumatic equipment and electronic control equipment. Review of marine icing and anti-/de-icing systems The aim of this work is to review the phenomenon of icing in marine operations. The focus is on two main sources of icing, namely atmospheric and sea spray. The literature reveals that sea spray icing is the main contributor to marine icing. This work discusses the available ice accretion prediction models on ships and offshore structures. It also reviews the anti-/de-icing technologies that can be implemented on ships operating in cold climate regions. The significance of ice detection is acknowledged, and a brief review of various ice detection technologies is provided. State evaluation of a corroded pipeline A submarine pipeline is mainly used to transfer liquid or gas between two land masses and may thus suffer corrosion from the severe subsea environment. Hence, the safety of a corroded pipeline must be evaluated by professionals. In this study, stress analysis of a corroded pipeline is conducted for three situations (i.e. simple defect, complex surface, and interacting defects). A simple defect is modelled using the 3D drawing software CATIA, and ANSYS simulates the stress of the corrosion defect; the effects of the defect were determined using different defect widths and depths. A complex surface involves the use of AUTOCAD VBA to establish the 3D surface based on random data and to generate an Initial Graphics Exchange Specification file. CATIA is used for file opening and geometric modelling and is finally keyed into ANSYS for analysis. ANSYS then is used for the stress analysis. The analysis covers a single rectangular defect, a spherical defect, and interacting effects of both types of defects. Finally, the stress analysis results are explored. Coupled oil analysis trending and life-cycle cost analysis for vessel oil-change interval decisions Extending lube oil-drain intervals is a practical way to save money on engine operating costs. At the same time, there may be increased risk of engine damage with increased drain intervals. The risks and benefits can be difficult to weigh and, as a result, starting an oil-drain interval extension programme can be an intimidating endeavour. This paper combines oil-analysis programme data interpretation and life-cycle cost analysis (LCCA) into a methodology for weighing oil-drain interval options. LCCA has shown decreasing incremental cost savings as oil-drain intervals are further extended, indicating that highly extended drain intervals may not be necessary to reap the majority of available cost benefits. Using oil-analysis property trending with life-cycle cost savings can allow an engine operator to set an hours-based extended drain interval based on projected risks and benefits. This interval is set as the baseline under the protocol, but continual oil analysis is used as a precaution to protect against unpredictable events. This methodology helps to reduce the risk of engine damage and ensure that the most easily attainable cost savings are still captured, making the decision about how to extend oil-drain intervals more approachable. Research on the shear stress of single-lap joints using a variational method A great amount of research has been conducted on adhesive joints as they are widely used in industry. In this study, in order to simplify the calculations and obtain more accurate results for the potential energy of adhesive joints, a variational method was used to investigate the stress distribution of adhesive single-lap joints with variable geometrical characteristics and material properties. A two-dimensional finite-element model, Volkersen method and experiments were used to validate the presented model. It was found that if the overlap length is long enough, increasing the length does not reduce the maximum shear stress. It was also found that the maximum shear stress occurs at the loaded side of the thinner plate when the material is the same, and occurs at the loaded side of the plate with the smaller elastic modulus when the two plates have the same thickness. Container hub-port vulnerability: Hong Kong, Kaohsiung and Xiamen Because of the extreme climate and environmental changes, container hub-ports in East Asia focus on port vulnerability rather than competitiveness. To sustain their competitive advantage, port operators adopt appropriate operational strategies based on various factors. We found that the strategies which container carriers adopt at hub ports have the most significant influence on hub-port vulnerability. The primary strategies include route diversity, sufficient cargo sources and transshipment containers, stable collaboration and incentives in the form of subsidies and rewards. These strategies demonstrate the dependence of container carriers on ports, and implementing such measures could reduce port vulnerability. Compared with the neighbouring Hong Kong and Xiamen ports, Kaohsiung Port is the most vulnerable and Hong Kong the least. Regarding all operating conditions that can be offered to carriers, Hong Kong Port possesses a competitive advantage that combats vulnerability. A dimensionless analysis of shear modulus and stress distribution for anisotropic materials The effect of various off - axis angles and lamina material properties on the shear modulus and the stress distribution of a circular cut-out in an orthotropic composite plate under shear loadings are presented. Based on the generalized Hooke's law, the generalized plane stress condition and the complex variable method, a dimensionless analysis is used to evaluate the influence of various elastic moduli E1, E2, G12 and ν12 on the laminate shear modulus and the stress distribution along the boundary of the circular cut-out within the laminate under shear loading. The results obtained from this dimensionless analysis provide a set of general design guidelines for structural laminates with high-precision requirements in engineering applications. A backstepping approach for the formation control of multiple autonomous underwater vehicles using a leader - follower strategy This paper presents a Lyapunov-based backstepping approach for developing cooperative motion control for multiple autonomous underwater vehicles (AUVs). The cooperative motion is achieved using a leader - follower (LF) formation strategy in the presence of discrete data transmission between the leader AUV and the follower AUVs. To address the problem of sampled data transmission between the leader and the followers, an estimator is designed for each follower to estimate the states of the leader AUV. The proposed backstepping control algorithm enables the follower AUVs to successfully follow the desired virtual frames provided by the leader AUV. Thus, a formation shape is maintained by the followers while traversing the desired path. The simulation results suggest that the proposed Lyapunov-based backstepping controller is effective for achieving successful cooperative motion control of multiple AUVs in the presence of discrete data transmission. The unsteady hydrodynamic characteristics of a partial submerged propeller via a RANS solver The motivation of the present study is to contribute to the knowledge of partial submerged propellers (PSPs) in the search for a reliable performance prediction method with regard to the ventilation flow around PSPs using a hybrid grid-based Reynolds-averaged Navier - Stokes (RANS) solver. To achieve accurate results and obtain the correct behaviour extraction of the ventilation zone, a fine mesh was generated around the propeller and the free surface. The hydrodynamic coefficients of the PSP, the ventilation pattern and other results were calculated for a range of advance coefficients. The predicted ventilated open-water performance of the PSPs as well as the ventilation shape are in good agreement with the experimental measurements and observations. Finally, discrepancies between the numerical results and the experimental data are quantitatively evaluated in terms of the relative percentage error for the PSP characteristics. On resolving reactive power problems in ship electrical energy systems This paper provides the main concept of reactive power compensation onboard ship electrical energy systems mainly as a retrofitting means towards assisting the generator plant and improving the ship efficiency in terms of fuel consumption and emissions. The alternative means are presented and discussed via the support of actual case studies. A mixed H2/passivity performance controller design for a drum-boiler system The mixed performance control problem of drum-boiler systems is discussed and investigated in this paper, subject to H2 and passivity performances. In order to minimize output energy and guarantee asymptotical stability, an H2 control scheme is employed to solve the stability problem of system. Passivity theory is applied to achieve disturbance attenuation performance. Based on the Lyapunov function, some sufficient conditions are derived into linear matrix inequality (LMI) form to apply a convex optimization algorithm. Based on the derived sufficient conditions, the asymptotical stability and mixed H2/Passivity performance of the system is achieved by the designed controller. Finally, some simulation results are proposed for a drum-boiler system with external disturbance in order to verify the validity and effectiveness of the proposed design method. Evaluation of shipyard selection criteria for shipowners using a fuzzy technique There are various stages to the process of acquiring a vessel. First, a decision must be made on what vessel is needed. Next, an appropriate shipyard is selected; this decision is very significant, as it will determine the workmanship quality and delivery date of the vessel to be manufactured. From the point of view of a shipowner, receiving the vessel on time and manufactured to specification is crucial. From the point of view of the shipyard, it is vital to take as many orders as possible; therefore, one of the highest priorities of shipyards is to be the first choice for as many shipowners as possible in order to increase their requirement for steel fabrication and consequently their profits. In this study, the shipyard selection criteria for shipowners are presented and the degree of importance of each criterion is determined using a fuzzy analytic hierarchy process (AHP) technique. The aim of the study is to provide shipyards with information about the perspective of shipowners with regard to the shipyard selection process, thus allowing shipyards to strengthen their weaknesses and increase the number of shipowners which choose them to manufacture their ships. Study on the mixing performance of static mixers in selective catalytic reduction (SCR) systems Selective catalytic reduction (SCR) is a promising technique for reducing nitrogen oxide (NOx) emissions from diesel engines. Static mixers are widely used in SCR systems before reactors to promote the mixing of ammonia and exhaust streams. This work aims to investigate the effects of the location of static mixers and the volume ratio of two species on mixing quality using the computational fluid dynamics (CFD) method. The simulation results show that a more homogenous ammonia distribution can be achieved at the exit of the pipe if static mixers are placed close to the ammonia injection point or if more ammonia is injected. Another phenomenon found in the study is that the mixing performance of an identical static mixer may behave discrepantly under different flow conditions if using B and C as the evaluating indexes for mixing homogenization. Structural integrity management system (SIMS) implementation within PETRONAS' operations Since 2008, PETRONAS Carigali Snd Bhd (PCSB) has embarked on developing a structural integrity management system (SIMS) within its Malaysian operations, which is compliant with the recently balloted American Petroleum Institute Recommended Practice for the Structural Integrity Management of Fixed Offshore Structures (API RP 2SIM) and International Standards Organization (ISO) 19902:2007. Increasing demand to extend the life of fixed steel offshore structures due to further oil and gas discoveries has resulted in the platform being subjected to higher loading due to modifications, upgrading and the demands of additional loading due to new drilling campaigns for which the platform may not have originally been designed. In addition, PCSB platforms are also be faced with other challenges and events such as major damage anomalies identified from the inspections performed, an increase in environmental metocean loading, the presence of shallow gas and seismic/earthquake loading for which the structure may have not originally been designed. As such, PCSB must be able to manage these additional integrity triggers and justify the ongoing integrity and fitness for purpose (FFP) of its platforms, beyond the design life of the structure. Making shipping greener: comparative study between organic fluids and water for Rankine cycle waste heat recovery The largest source of energy loss in ships is found in the propulsion system. This study focuses on the concept of managing waste heat energy from the exhaust gases of the main engine. Using waste heat recovery systems (WHRSs) to make shipping more efficient represents a good area of opportunity for achieving the shipping industry's green objectives. Organic Rankine cycles have been applied in land-based systems before, showing improvements in performance when compared with the traditional Rankine cycle. As marine environmental rules requiring greener vessels and engine thermal efficiency continue to increase, thus reducing the available energy in the exhaust, organic Rankine cycle WHRSs become a more attractive option.The proposed WHRS was modelled using MATLAB for a typical ship installation with a slow speed diesel engine and a WHRS installed after the steam boiler in the exhaust gas system. The energy recovered from the exhaust gas flow is transformed via the thermodynamic cycle - coupled with a generator - into electricity, which helps to cover the ship's demand. The MATLAB code found the highest electric power output, hence the maximum fuel and CO2 emission savings possible, by v varying the WHRS HP. Water and four organic fluids were considered and their performance was compared over a range of different engine operating conditions. A representative ship operating profile and a typical marine generator were used to measure CO2 emission reductions. The implications of having flammable organic fluids on board are also briefly discussed. This work demonstrates that a simple organic Rankine cycle can be more effective than a steam cycle for the same engine operating conditions. On electric warship power system performance when meeting the energy requirements of electromagnetic railguns This paper investigates the ability of a gas turbine alternator to provide power to an electromagnetic (EM) railgun in surface combatants whilst simultaneously maintaining acceptable levels of power quality for other consumers. Following the justification of the investigation and a description of the research methods, which includes a proposed EM railgun charging control system, this paper discusses simulated results obtained from a validated power system model for both single shot and salvo operations and identifies factors that limit the rates of fire in each case. The findings from the investigations suggest that a large gas turbine alternator is able to maintain quality of power supply within acceptable limits for the majority of EM railgun operations. However, the transient quality of power supply limits was found to be exceeded for heavy firing rates but would remain within tolerable limits as defined for exceptional loads by North Atlantic Treaty Organization (NATO) standardization agreement (STANAG) 1008 Edition 9. A method to increase the maximum rate of fire whilst maintaining the quality of the power supply within acceptable limits is investigated by increasing the size of the energy storage device and retaining a residual charge. The paper concludes by suggesting that the EM railgun system could be integrated into the ship's electrical system without the need for any additional prime movers to achieve rates of fire commensurate with a surface combatant's requirements. Improving electric power quality in ships via surge protection devices (SPDs) Power supply quality (PSQ) problems in ship electrical systems are a major concern nowadays due to the electrification of most equipment on board. Among PSQ problems, voltage transients and spikes are probably the most important as they can cause a range of malfunctions and even damage. The most efficient method of protection against spikes and transients is the installation of surge protection devices (SPDs). This paper presents a holistic methodology of selecting the most appropriate installation points in the ship's electricity grid on the one hand while not resulting in new problems on the other. Alternative, indirect measures of ballast water treatment efficacy during a shipboard trial: a case study A shipboard study was conducted aboard the cruise ship Coral Princess during a scheduled cruise from San Pedro, CA, USA to Vancouver, British Columbia, Canada. The investigation involved three members of the global TestNet group, with experience in certification testing of ballast water treatment systems (BWTS) designed to eliminate entrained invasive species. A UV-based ballast water treatment system had been employed aboard the vessel for more than 10 years. A variety of established and experimental assessment techniques were employed, both aboard the ship and following shipment of samples via road (5 days) and air (7 days) to remote laboratories. The study was designed to compare the performance of different techniques in assessing BWTS compliance with international regulations, and to test the feasibility of compliance assessment by Port State Control internationally using different laboratories. Overall, biological end-points showed effective treatment of ballast water as judged by the percentage removal (mortality) of organisms in treated samples. Sample transport indicated generally good potential for `off-site' sample analysis and displayed a possible latent effect of treatment as judged by a decline in photosynthetic yield associated with delayed analysis. Subsea cable tracking in an uncertain environment using particle filters Localization of subsea cables is a demanding and challenging task. Among the few methods reported in the literature, magnetic field detection is the most promising one, as the cable does not require to be seen visually. Magnetic noise and a quick attenuation of the magnetic field propagating in sea water often make available methods unreliable. The authors propose a novel method of using particle filters for estimating the position of a subsea cable in a highly uncertain environment. The method was tested on data collected from a buried cable in the Baltic Sea, Denmark and shown to have a close approximation to the true location of the subsea cable. The method can be used to localize a subsea cable in an offshore noisy and uncertain environment and provides an inexpensive alternative to the use of a diver or a remotely operated platform. Integration of navigation systems for autonomous underwater vehicles An autonomous underwater vehicle (AUV) requires a precise navigational system for localization, positioning, path tracking, guidance and control. The main navigational device for an AUV is an inertial navigation system (INS) because high-precision navigational devices such as the Global Positioning System have a limited usage in the underwater environment. In this study, based on the dynamic mathematical model of AUV, we develop two types of low-cost integrated navigational system for AUVs based on error models of INS and its aiding devices such as Doppler velocity logs, compasses and pressure depth sensors. Nine- and 15-state INS error models and corresponding measurement models of aiding devices are derived for the Kalman filter (KF). We compare the performance of those two integrated navigation systems. The simulation results confirm that low-cost inertial measurement unit sensors produce a notable amount of noisy measurements, but KF-based integrated navigation system models for AUV can effectively mitigate those drawbacks. Design and development of an autonomous underwater vehicle - robot dolphin This paper aims to describe the design of a robot dolphin with a voluntary movement function. First, a motion mechanism is described based on 3D motion analysis to determine the length of each link and the swing angle of each joint in the robot dolphin. Two microchips are used to control the swing angle of the actuator in each joint. In order to understand the motion characteristics of the robot dolphin, a microcomputer is installed in order to obtain various motion data. The experimental data for three-axis accelerations and three-axis angles are found to be the same as the oscillatory frequency of the robot dolphin in the swing forward motion. The robot dolphin is designed to possess an avoidable collision function and artificial intelligence, artificial intelligence consists of the preprocessing image and back-propagation learning to implement specific motion identification. Comparison of ship plant layouts for power and propulsion systems with energy recovery Currently the increase in fuel costs and the need to reduce Carbon Dioxide (CO2) emissions have encouraged the search for even more efficient solutions to be adopted in energy conversion systems for ship installations. These systems generally include thermal prime movers consisting mainly of two-stroke or four-stroke diesel engines, a waste heat recovery (WHR) plant, a steam turbine and possibly a gas turbine, as well as electric machinery. These components may be used in various ways and through different plant configurations, whose optimization is under investigation by researchers and diesel engine manufacturers. In this paper four schemes of ship propulsion plants, using a two-stroke diesel engine equipped with waste heat recovery system are presented, analysed and compared by simulation. Some of the considered layouts can include also an electric motor to support the main engine in providing power to the propeller. On the other hand the electric power can be generated from both the diesel generators and the WHR plant. The considered propulsion plant schemes are compared in order to identify the configuration that meets in the best way the request of propulsive, electrical and thermal power of a 158000 dead weight tonnes (DWT) crude oil tanker ship, belonging to the Premuda Company, taken as a reference unit in this study. The comparison is carried out taking into account the payback time of the installation, the annual saving in the fuel outlay and the CO2 emissions. This last parameter is considered through the evaluation of the Energy Efficiency Design Index (EEDI). Implications of ice class for an offshore patrol vessel The importance of the Polar Regions, both in the Arctic and Antarctic, is likely to increase as environmental and geopolitical developments (e.g. as a result of global warming and increasing shortages in fossil fuels) increases international interest in these areas. With this in mind many nations, are increasingly considering the operation of naval vessels in these areas. This paper identifies the aspects of vessel design that must be addressed to enable an offshore patrol vessel (OPV) OPV sized vessel to operate in ice conditions and estimates the cost of adding this capability. A baseline 90 m OPV design was developed that had no requirement for ice operation and a unit platform cost (UPC) calculated for the vessel using a simple, weight based, cost model. The changes that would need to be made to this vessel in order to achieve increasing levels of ice capability were then identified and the ship was redesigned to accommodate these changes. The resulting UPC cost estimates for the ships give an indication of the whole ship cost implications of adding a requirement to operate in ice for a vessel of this size. Two variants were developed, the first to ice class IC (Lloyds Naval ship rules) and the second to Ice Class IAS. These represent light ice conditions (with icebreaker assistance where necessary) and difficult ice conditions, first year ice up to 1.0m, (without the assistance of icebreakers) respectively. The structural and ship system requirements of Lloyds Naval ship rules together with the relevant recommendations on winterisation were incorporated in both designs. Study of de-noising techniques for SNR improvement for underwater acoustic communication A wide variety of underwater applications, such as warning systems and deep water positioning systems used to track autonomous underwater vehicles etc, warrant underwater wireless communication in both shallow and deep water. These systems do, in the main, use underwater wireless acoustic communication modems and hence they play a vital role in underwater networking systems. However, the acoustic signals, particularly in shallow water suffer severe disturbances due to ambient noise from the sea. The ambient noise underwater is very unique, location specific and deterministic. This paper aims at developing a de-noising algorithm to improve the Signal to Noise Ratio (SNR) using the Gabor Wavelet and compares this with another important Shannon wavelet by way of study. The simulation was carried out using the MATLAB Simulink tool. The ambient noise used for the validation is sea truth data collected in the Bay of Bengal using broad band hydrophones. The results obtained were very encouraging as, in the range of input SNR-15 dB to 0 dB, the improved output SNR was in the order of 8 dB. This shows the efficiency of the algorithm simulated. DC: Is it the alternative choice for naval power distribution? Historically, the vast majority of electrical power distribution designs in marine applications have been Alternating Current systems. This has continued for recent platforms and is arguably due to commercial inertia, conservatism and the manufacturing and technology supply base. however whilst it is widespread in use, it may not allow fully optimised power systems to meet tomorrow's demand for more efficient ships with greater flexibility. This paper on the back of a recent programme of work, questions the future use of fixed frequency AC generation and power transmission on board ships, and asks if maybe Thomas Edison was on the right lines over a century ago? Estimation of uncertainty in ship performance predictions Both for a shipyard, and its customers, it is important that a newly delivered ship meets the contracted performance criteria. To test whether the contractual requirements are met, the actual ship performance is measured during sea trials and compared against the contracted values. If the contracted performance is not met, this can lead to penalties for the shipbuilder and dissatisfaction for the customer In extreme cases this can even lead to non-acceptance by the customer On the other hand, an under predicted performance can lead to non-competitive design and a lower income for the shipbuilder Contracted performance values are based on predictions and experience. Unfortunately accurate predictions of performance are often difficult due to the uncertainties that are involved in the design and build of the ship and its (propulsion) installation. Furthermore, uncertainty is introduced by the full scale trials during which the various contracted performance aspects are to be demonstrated. This paper describes two methods to estimate the uncertainty in performance predictions. In particular a ship speed prediction and a bollard pull prediction are discussed. Both methods are explained, compared and strengths and weaknesses are discussed. Water, Nitrogen and Phosphorus Budgets of Lake Manzalah Water, nitrogen and phosphorous budgets of Lake Manzalah have been calculated using data collected from January-December 2012. The main positive contributor in the water budget of Lake Manzalah during this period was the drainage water being 6692.6x 106 m3/yr. The annual rainfall was 26.53x106 m3/yr and the evaporation 1075.6x106 m3/yr. The net annual amount of outflowing water from Lake Manzalah to the Mediterranean Sea was 9503.23x 106 m3/yr and the net amount of seawater invading the Lake estimated is 3895.76x 106 m3/yr Accordingly, the net water exchange between Lake Manzalah and the Mediterranean Sea was 5607.47x 106 m3/yr seaward negatively affects the water budget of the lake. The nutrient salt load of the nitrogen load in the lake from the drainage water is estimated to be 1381.09 tonnes per year and that from the rain 3.703 tonnes per year. The amount of nitrogen load transported to the sea through Boughaz El-Gamil is around 1025.753 tonnes per year; while the inflowing nitrogen load is estimated to be 180.147 tonnes per year The phosphorus load from the drainage water is estimated to be 2212.67 tonnes per year and that from the rain 8.546 tonnes per year. The amount of phosphorus load transported to the sea through Boughaz El-Gamil is 680.668 tonnes per year while the inflowing phosphorous load is estimated to be 101.687 tonnes per year. Developing a new methodology for evaluating diesel -- electric propulsion In this paper a methodology for the design of ships with a diesel-electric propulsion system, according to the All Electric Ship concept is presented. Taking into account that there is no long and extensive relevant experience for all ship types, diesel-electric propulsion should be selected after careful consideration and thorough evaluation of many technical and economic parameters. This paper offers a step-by-step procedure towards this direction; the discussion is enriched by a case-study of a ship type, where the selection of the proper propulsion system is not straightforward as many parameters have a substantial impact. Stability and transient-behavioural assessment of power-electronics based dc-distribution systems. Part 4: Simple compensation to improve system performance Ideas and developments in electrical systems are being stimulated by progress in power electronics; nowhere is this more apparent than in the marine industry. Indeed, it may be that such advances will lead to dc-distribution systems appearing once again in vessels. Concomitant with these developments is the need for new methods, or adaptation of known methods, to be found to deal with design and analysis problems that will inevitably arise. The problem considered in this paper is an extension of that considered in Sudhoff's original paper and our first three parts in the current series. Parts 1, 2 and 3 have discussed the response of the system when attempting to stabilise it with a capacitor C. Root-locus, Bode diagrams and Kharitonov polynomial tools have been used in these investigations. In this part of the paper consideration is given to further improving the stability and transient behaviour. Attempts to improve the behaviour by adding compensating components into the main dc path lead, generally, to loss of energy efficiency or, in some other way, unacceptable conditions being manifested. Crank-angle resolved NOx measurements from a specific engine cylinder of a large marine two-stroke engine The ability to measure NOx in the exhaust gases immediately as they exit the cylinder requires special instrumentation, located at the vicinity of the exhaust valve and capable of measuring with extremely fast response times. The information obtained from such a measurement is important to engine designers and combustion engineers with regards to the causes of formation of NOx emissions. This information can be used for the calibration of Computational Fluid Dynamics (CFD) models for engine development purposes, and also, during the development of the combustion system of an engine where engine tuning could lead to an overall compromise between engine's performance and emission behaviour. The current research work involved the execution of crank-angle resolved NOx emission measurements in the exhaust port of an individual cylinder of a large twostroke marine engine, using a custom designed sampling system. The measurement campaign was performed on the MAN 4T50ME-X test engine, at the MAN Diesel & Turbo test centre in Copenhagen. The extremely fast response time of the sampling system and the analyser provided important details about the NOx profile, as soon as it exits the cylinder of a large marine engine. The observations were further evaluated using CFD analysis of the flow conditions inside the exhaust port. Load resistance factor design (LRFD) calibration of load factors for extreme storm loading in Malaysian waters Petronas Carigali Sdn Bhd (PCSB) performs its operations in three regions in Malaysian waters. Currently, fixed offshore structures are designed in accordance with ISO 19902, which is meant to provide guidance for the design of fixed offshore structures internationally, but the load factors in ISO 19902 are often derived from studies performed in the Gulf of Mexico and the North Sea environments. One such load factor is the 1.35 for environmental loading that has been prescribed within ISO 19902. Site specific metocean studies in Malaysian waters reveal environmental conditions that are more benign than the Gulf of Mexico and the North Sea. This study presents the results of a component-based approach and investigates the suitability of the 1.35 load factor for environmental loading in Malaysian Waters to achieve the same target reliability for a similar structure designed consistent with the provisions of the API RP 2A-WSD. First-order reliability methods (FORM) have been undertaken using a database of tubular components with representative geometries, bending-to-axial stress ratios, dead-to-live load ratios and environmental to gravity load ratios. The analysis results show that if the target were to be based on the weighted average probabilities inherent in the API RP 2A-WSD 21[st] Edition, an environmental action factor of 1.245 (Region A), 1.295 (Region B) and 1.255 (Region C) with the ISO provisions will on average achieve similar reliability. A semi-analytical method of free vibration of fluid loaded ring-stiffened stepped conical shell The motion differential equation of fluid loaded ring-stiffened stepped conical shell is derived by using Flügge classical thin shell theory and equivalent method of ring stiffeners, in which force and moment of a ring stiffener is apportioned in the spacing. In the analysis, the conical shell is divided into many narrow cylindrical strips, the fluid loading of which are available, to consider fluid loading of conical shell approximately. The semi-analytical solution in terms of power series to the equation is then obtained. The free vibration frequencies of a ring-stiffened stepped conical shell in air and in water are respectively solved using the proposed method. The results are compared with the ones obtained from the finite element software ANSYS and they are well agreed. On future ship safety - people, complexity and systems This paper considers the characteristics of risk and the important influences of the human element, innovative solutions and complex engineered systems on the future of maritime safety. Residual risk remains once all practicable steps have been taken to manage the risks associated with any undertaking. For complex engineered systems, and these include modern merchant ships, there will always be a limit on the attainable level of safety, where human performance and technical issues such as complexity and novelty will dominate the residual risk and the causes of incidents. Recent work has also shown a divergence between the historical record and the perception of risks held by a range of maritime professionals. This divergence may explain some of the maritime incidents that appear to be the consequence of apparently inexplicable acts. The paper sets out some of the issues relating to residual risk. It concludes by considering the future of ship safety, and the effective regulation of safety for future ships, taking into consideration people, systems and the management of risk. Experimental and numerical modeling of the high speed planing vessel motion In this paper, the motion of a high speed vessel with planing hull has been investigated using experimental and numerical methods. The aim of this investigation was to compare the results of the model towing test with the numerical results obtained in the Savitsky method. The most important issue in numerical modeling of a planing vessel is to determine its continuous motion conditions, such as trim and draught. Comparing the results, it is concluded that they are in a non-variant range in spite of uncertainties of continuous motion and hence modeling of the positions. Therefore, it could be expected that with more complete modeling, the study of the vessel motion conditions becomes possible, especially for high speed vessels. Fuzzy AHP and Fuzzy TOPSIS integrated hybrid method for auxiliary systems of ship main engines Although significant technical precautions have been taken in marine diesel engine and auxiliary systems, unexpected failures can be observed during operation. These failures can sometimes lead to irreversible losses. The purpose of the study is to present fuzzy Analytic Hierarchy Process (AHP) and TOPSIS (Technique for Order Preference by Similarity to Ideal Solution) methods that can be applied for failure detection in auxiliary systems and marine diesel engine determined by group of experts. By evaluating the decision-making groups, the system most affected by failures was determined. A wind tunnel investigation of the aerodynamics of sailing dhows This paper presents the results of experimental tests conducted to study the aerodynamic performance of sailing dhows. The investigation emerged from the interest shown by designers and sail-makers to understand how these sails perform, since they have never been studied before. The 43ft/13.1m and 60ft/18m classes have been tested in a wind tunnel where the effects of varying several parameters were investigated. These parameters were: heel angle, apparent wind angle, bending and stiffness of the yard, optimisation of the trimming and the influence of the mizzen on the mainsail. This is the first investigation where the aerodynamic performance of the lateen sails on sailing dhows has been investigated in a wind tunnel study. Dynamic analysis of composite marine structures using full-field measurement techniques Composite materials are increasingly used in structural applications within the marine industry. Due to the geometric complexity of marine structures, there is a practical requirement that they are assembled by joining smaller component pieces using either mechanical fasteners or adhesive bonding. In this paper Digital Image Correlation (DIC) is used to provide full-field analysis of the complex strain fields generated within an adhesively bonded composite single lap joint. Tests are undertaken quasi-statically and at high rate, demonstrating a significant change in the assembly response between laboratory testing conditions and dynamic loading events typical of the marine environment. The work demonstrates the potential of applying full-field experimental technique to provide detailed analysis of complex structural problems, typical of large marine structures. Application of the Monte Carlo method to marine system maintenance studies considering failure modes and spare part stock control This paper presents a methodology using Monte Carlo methods to analyse factors affecting the operational efficiency of marine systems. The aim is to show how Monte Carlo methods can take into account a wide range of factors to facilitate the decision making process for maritime operations. Monte Carlo methods have been used to model the varying effects of different failure modes and the items which contribute to failure for a given system. The methods used have been applied to a case study to examine system failure probability, downtime and maintainable item contributions for a marine cooling system. Anchor selection study for ocean current turbines This paper compares anchoring systems suitable for ocean current turbines, specifically those proposed to be installed off the coast of Southeast Florida near latitude 26° 5' N. This location boasts one of the most energy dense ocean current resources, with a mean kinetic energy flux of approximately 2.34 kW/m[2]. Seafloor types ranging from unconstrained sediment to high relief ledges were observed during regional benthic surveys for which applicable anchor types would include deadweight, plate, pile, and drag embedment. Numerical simulations of single point moored marine hydrokinetic devices were used to extract anchor loading at a likely deployment location for mooring scopes from 1.25 to 2.0 and turbine rotor diameters between 3 - 50m. These anchor loading data were used for preliminary sizing of deadweight and driven plate anchors on both cohesionless and cohesive soils. Finally, the capabilities of drag embedment and pile anchors relevant to ocean current turbines are discussed. Multiple types of drag embedment anchors can support all of the predicted loads if adequate sediment exists and the loading is horizontal, while pile and H-type anchors can support all of the evaluated loading scenarios and chain-in-hole anchors can support turbines with rotor diameters up to 30m. LNG as a marine fuel: Likelihood of LNG releases Liquefied Natural Gas (LNG) presents additional hazards compared with traditional marine fuels, such as oil and diesel. Hence, the safety risk is also different to that from traditional fuels. To ensure the safety of those onboard it is necessary to consider different and/or additional safeguards when using LNG. In the absence of prescriptive requirements, risk analysis techniques are being used to identify such safeguards. However, it is not always apparent if these safeguards address those areas that contribute most to the risk. To help with this, a risk assessment model has been developed. A critical step in this model is the determination of release likelihood and this provides guidance on the selection of appropriate safeguards in leak prevention. This paper summarises the release likelihood data used and provides an example of its use for a simplified LNG fuelling system. On certain peculiarities in the simulation of marine power plants This article reviews the possibilities for complex technical systems simulation enhancement due to unification of procedures of representation and statistical sampling of random parameters realisation, as well as the reduction of the necessary number of simulations. A generalised fitting criterion is presented for consequently obtained computational distribution laws of a target function, which may be used for a formalised evaluation of the number of simulations needed for developing a rational engineering strategy. Using several characteristics of marine power plants as an example, the regularities in changes in their distributions' principal parameters, found in process of simulation are shown. A technique is developed for prognosis of computational distribution characteristics, which allows reducing the simulation length significantly. Evaluation of emission sensors for four-stroke marine diesel engines Marine engine manufacturers are developing adaptive engine control systems that utilize the output of various sensors to run algorithms and diagnostic routines in order to operate the engine at optimal conditions, with respect to fuel consumption and emissions. The selection and evaluation of such sensors is a very important process and it involves extensive hardware testing on the test bed. The current study describes the selection and evaluation of NOx sensors for possible adaptation in the control system of marine diesel engines. The most promising sensors for such an application were installed in the exhaust system of two different test engines. The results from both test beds were comparable and the analysis revealed the main areas of concern with regards to the applicability of the sensors in marine diesel engines operating with HFO fuel. Effects of high sulphur content in marine fuels on particulate matter emission characteristics The effect of high sulphur level in marine fuels on diesel exhaust particulate number distributions and mass concentrations was investigated. A strong correlation between increase of emitted particle mass and number and increase in fuel sulphur content was found, with the most affected particles found in the nanoparticle size range. These particulates consist of nucleated/condensed hydrocarbon and sulphur compounds, which are potentially the most hazardous to human health. Nucleation mode particles were found to be greatly affected by primary dilution temperature (PDT), an increase in which cause the vapour pressures of volatile species to rise, considerably slowing down the nucleation process. However, even at PDT=400°C this mode was not entirely eliminated, since with high-sulphur fuels the concentrations of nucleation-prone vapour-phase volatile components are very high, and higher primary (and overall) dilution ratios are therefore required. The soot fraction, which made up most of the particle mass, was only slightly affected by sulphur content and is unaffected by PDT, although it was influenced by engine operating parameters such as load and speed. Prediction of the fuel saving and emissions reduction by decreasing speed of a catamaran The maritime environment faces many challenges due to growing amounts of ships, exhaust gases, especially carbon dioxide which accounts about 3% of global production in addition to the expected increase of marine fuels prices in the coming years. Reducing vessel speed may contribute to overcoming the previous problems. Although speed reduction has been a part of the solution for some vessel types, for example container ships, its applicability for certain ships type, such as high speed crafts, may face some constraints. Based on the results of previously published experiments, regression equations, and data acquired using software packages, this paper analyses the various methods that can be used to estimate the effect of speed reduction strategy on the amount of both consumed fuel and released emissions, in case of high-speed, hard-chine displacement catamarans. As a case study, this concept has been applied on a vessel operating in the Red Sea region to be evidence for applicability and benefits of this approach. Thin plate buckling mitigation and reduction challenges for naval ships Thin plate buckling or distortion on ship structures is an ongoing issue for shipbuilders. It has been identified that a significant number of factors can be put in place based on prior knowledge and good practice. Additionally, research work aimed at reducing thin plate distortion has been relatively prolific, particularly in the area of simulation modelling. However, the uptake in the research findings by industry has been relatively low. A number of these findings are discussed and their application considered. To achieve further reductions in thin plate distortion, there is a clear need for better interaction between research institutes and industry. A proposed methodology for assessing the reduction of a seafarer's performance with insufficient recuperative rest Risk factors in maritime transportation have been analysed by many researchers. After analysing many accident reports and by taking the vessels' flag, ownership, type, age, type of cargo and location of accident into account, human factors were found to be the prevalent causes. A seafarer's performance depends upon many variables and alteration of a criterion value will ultimately alter his (or her) performance. This paper makes use of a Bayesian network and a `symmetric method' by exploiting a conceptual and sound methodology for the assessment of a seafarer's performance. The methodology developed has been applied to a case study in order to demonstrate the process involved. By exploiting the proposed model, an equation for the reduction of a seafarer's performance with insufficient recuperative rest is formulated. Furthermore, the proposed methodology enables decision makers to assess the performance of a seafarer before his designation to any activities and during his seafaring period. A genetic algorithm based nonlinear guidance and control system for an uninhabited surface vehicle There is an increasing drive to develop uninhabited surface vehicles (USV) as cost effective solutions to a number of naval and civilian problems. In part, the resolution of these problems relies upon such vehicles possessing robust guidance and control (GC) systems. Furthermore, the vehicles need to be operated under tight performance specifications satisfying multiple constraints simultaneously. This requires vehicle nonlinearities and constraints to be explicitly considered in the controller design. Nonlinear model predictive control (NMPC) is well suited to satisfy these requirements. This paper reports the design of a novel GC system based on NMPC for use in a USV named Springer which is benchmarked against a linear proportionalintegral-derivative counterpart. The NMPC combines a recurrent neural-network model and a genetic-algorithm-based optimiser. Common to the two GC systems is a waypoint line-of-sight (LOS) guidance subsystem. The control objective is to guide the vehicle through different waypoints stored in a mission planner. The performances of the guidance and control systems are evaluated and compared in simulation studies with and without appropriate disturbances. From the results presented, it is concluded that the GC system based on NMPC is more efficient and more capable to guide the vehicle through LOS waypoints particularly in the presence of disturbances. An integrated methodology for the optimal thermal design of an ocean turbine pressure vessel: A soft-computing approach This paper presents a generally applicable approach to numerical thermal design. A novel thermal design procedure for the prediction of heat transfer inside a pressure vessel of an ocean current turbine using integrated procedure by finite element method of heat transfer analysis, artificial neural network and genetic algorithms is presented. Numerical heat transfer analysis was done using commercial software ANSYS for two-dimensional heat transfer in simplified domains. The calculation was limited to only heat conduction. The ANSYS simulations results were used for training and approximating the unknown functional behaviour of heat transfer by using artificial neural networks (ANN). The trained ANN serves as the nonlinear objective function of the optimisation procedure. Genetic algorithms (GA) were used as the optimisation tool. The optimum results obtained from the GA were verified against ANSYS and ANN results. Both the ANN and GA were implemented in MATLAB, while the overall methodology proposed could be applied to other engineering design problems as well. A fuzzy genetic algorithm approach for analysing maintenance cost of high risk liquefied natural gas carrier systems under uncertainty A Fuzzy Genetic Algorithm (FGA) is used to treat uncertainties associated with unit costs of maintenance of Liquefied Natural Gas (LNG) carrier systems such as a containment system and a transfer arm. A Fuzzy Rule Base (FRB) is established to identify the unit costs of maintenance of the LNG containment system and transfer arm. It includes 125 LNG carrier maintenance cost rules, with technical consultancy cost, maintenance duration, and spare part cost as the antecedents and maintenance cost as the consequent. The outcome from the FRB is used to optimise a risk model using a Genetic Algorithm (GA) approach to find the optimal maintenance cost of each system with provided information on their respective time of interest, failure probability, failure frequency and maintenance cost of the whole LNG carrier system. Dynamic stability analysis of pipeline based on reliability using surrogate model The surrogate model is applied to a reliability analysis of the stability of untrenched pipelines. Experience has shown that the stability of a pipeline is sensitive to a number of variables associated with soil properties, structural modelling, and hydrodynamic loads. Because uncertainties are unavoidable in these key design parameters, it is vital to evaluate pipeline instability failure using a probabilistic approach. However, a reliability analysis of on-bottom stability in the time domain is often a very computationally expensive process. The stability of a pipeline exposed to time-varying environmental loads is a highly nonlinear phenomenon with pipe - soil - fluid interactions. In this work, surrogate model technology was utilised to realise the reliability analysis which is also quite time consuming. A Fourier model was used to calculate the hydrodynamic forces acting on a pipeline. The surrogate model was constructed using experiment sampling and, in conjunction with a Monte Carlo simulation, was employed to perform the reliability evaluation. Moreover, the influences of various random variables on pipeline instability are discussed through a sensitivity analysis. An accurate assessment of the conservatism in the different variables could lead to fewer conservative stabilisation requirements. The results show that using a surrogate model for reliability analysis not only reduces the computational cost significantly but also produces a highly accurate evaluation of pipeline stability. Free running tests of a waterjet propelled unmanned surface vehicle Free running manoeuvring trials of an unmanned surface vehicle were performed to investigate the effects of cross flow at the inlet of the waterjets powering the vehicle. Circle, straight-line and turning manoeuvres were performed. A data acquisition system was developed to simultaneously measure the thrust developed by the waterjet and the vessel's position, speed, and heading. The longitudinal and cross flow at the inlet of the waterjet was estimated from the speed and heading measurements. It was found that the presence of cross flow correlates with a drop in the thrust developed at the waterjet. Submarine power cables: laying procedure, the fleet and reliability analysis This paper investigates the possible problems of laying submarine power cables (SPC). It provides information about the power cables themselves, the laying procedure, and the respective fleet. It also describes the medium voltage (MV) transmission grid in Greece and determines the correlation of different parameters that affect the operation of the network. Furthermore, a reliability analysis, based on a statistical model, is conducted; the model is used to evaluate the operational characteristics of cables, such as the rate of failures from natural or human causes, and is also used to choose between different candidate cable routes in order to determine the one with the maximum reliability. The permeability parameter of curtain-wall breakwaters A numerical model has been developed that can predict the interaction of regular waves on a single curtain-wall pile breakwater (CP B). The model is based on an eigenfunction expansion method and utilises a boundary condition close by the vertical piles that accounts for energy dissipation. Energy dissipation is modelled through the boundary condition via a permeability parameter. This parameter is a complex constant, so that the real part of it corresponds to the frictional resistance of the breakwater and the imaginary part is associated with the inertial resistance of the breakwater. The permeability parameter can be calculated by two different methods: one is advantageous because all the frictional and inertial related variables are known; the other requires the frictional resistance to be calculated by comparing the mathematical and experimental results The purpose of this study is to compare the accuracy of the two methods in the case of CP Bs, and express the frictional resistance of the CP Bs as a function or a constant for the second method. Modelling is validated by comparison with previous experimental studies. The appropriate range for the porosity and relative draft of the breakwater is discussed and it is suggested that CP Bs can operate both effectively and efficiently in the draft range 0.15 - 0.75, although to achieve maximum energy, the porosity between 0.15 - 0.2 is recommended. Time-frequency feature analysis of naval vessel impact response using the Hilbert Huang transform method The empirical mode decomposition (EMD) methodology for analysing the impact response features of a naval vessel is proposed and investigated. Records collected from the vessel structure subjected to underwater explosions (UNDEX) are used in this study. The ability of EMD to extract the two oscillation modes, the intrinsic mode functions (IMFs) and the frequency energy mode, is demonstrated. The spectrum obtained using the Hilbert Huang transform (HHT) method can provide more precise and effective time-frequency energy representations than those obtained using the short time Fourier transform (STFT) or wavelet transform (WT) based methods. It is concluded that HHT is a useful tool from which to gain an in-depth understanding of the impact response of UNDEX and be able to quantify it. SNR and MSE analysis of KLMS algorithm for underwater acoustic communications Adaptive filters are used to alleviate the degradation caused by wind-driven ambient noise in shallow water. Underwater acoustic signals are greatly affected by ocean interference and ambient noise disturbances propagated through underwater channels. Therefore, an effective adaptive filtering system is necessary for denoising signals degraded by noise. In this paper, adaptive algorithms, such as Least Mean Square (LMS), Normalised LMS (NLMS) and Kalman LMS (KLMS) are analysed in terms of their performance with the aid of performance measure characteristics, such as Signal to Noise Ratio (SNR) and Mean Square Error (MSE) for various wind speeds ranging from 2m/s to 6m/s. Simulation results show that KLMS algorithms achieve remarkable performance, even in the very low SNR region, compared to LMS and NMLS algorithms. Moreover, it is observed that the output convergence is also very fast for KLMS and the performance of the KLMS algorithm is unswerving for different wind speeds. Modelling the yaw dynamics of an uninhabited surface vehicle for navigation and control systems design Marine and Industrial Dynamic Analysis Research Group, Advanced Engineering Systems and Interactions, School of Marine Science and Engineering, Plymouth University, UK In many navigation and control systems a key element in both their software architectures is a model which describes the dynamic characteristics of the given vessel. Thus, to ensure uninhabited surface vehicles are capable of successfully completing missions, it is essential that the marine control systems designer has available bona fide dynamic models that are accurate and possess good prediction capabilities. Since hydrodynamic modelling is usually very expensive, time-consuming and requires the use of specialist equipment in the form of a tank testing facility, it was considered more appropriate to model the vehicle yaw dynamics using Black-box identification techniques. In this paper five new models of the yaw dynamics of the Springer uninhabited surface vehicle are developed and presented. Two of the models that are based on auto-regressive with exogenous variable and autoregressive moving average with exogenous variable techniques are linear in composition and expressed in their state space formats; while the others are nonlinear and described by multi-layer perception, radial basis function and recursive neural networks. In all cases, the models were optimised using a steady-state genetic algorithm. The models are compared using a Mean Square Error criterion and residual correlation analysis. From the evaluation undertaken, it is shown that the yaw dynamic model based on the recursive neural network outperformed the others in terms of predictability and exactness. Therefore, it is concluded the recursive neural network model is the best candidate for navigation and control systems design studies involving the Springer vehicle. Tidal effects on pattern-based marine surface operations Coverage planning is a well-known problem in motion planning research and marine pattern-based operations, such as minesweeping, are traditionally performed within a defined quadrilateral area. Minesweepers are deployed to clear areas such as an entrance to port before capital ships are allowed through; such locations generally experience high variation in tidal flow throughout the day as the tide changes direction. Ever-increasing energy costs, combined with limited energy storage onboard ships, prompted this study, evaluating the marine coverage process in further details, such that the vessel can maximize the environmental condition to minimise the operation time or energy expenditure, reducing the overall cost of the operation. This paper presents a preliminary study on the effects of tidal flow on pattern-based marine surface operations. Evaluating navigation safety for harbours in Taiwan:An empirical study The purpose of this study is to develop a model to evaluate navigation safety for vessels sailing to Taiwanese commercial harbours, based on a dataset of marine casualties. In order to devise the practical model, linear programming and Data Envelopment Analysis (DEA) are applied to develop a DEA model algorithm. Firsly, a simplified DEA model structure is constructed with one input variable, eleven output variables and four decision-making units (DMUs). Then, a dataset of marine casualties are employed to obtain all variables, as well as to appraise the superiorities of all DMUs. Finally, utilising a DEA solver, the highest navigation safety for the DMUs can be evaluated. Furthermore, an empirical survey of navigation safety in Taiwanese commercial harbours is performed to appraise the systematic approach, ie, the DEA model. The results of this study show that: (1) Keelung and Kaohsiung are the most dangerous harbours; (2) Taiwan's safest commercial harbour is Haulien, followed in ranking by Taichung, Kaohsiung and Keelung; (3) Based on the algorithm of the proposed DEA model, the results can reflect reality. Adjustable bolted propellers: mitigating uncertainty One of the activities during the design of a ship propulsion system is the matching of the ship, propeller, gearbox and driving machine. The theory behind this matching process is well defined and documented. An important choice in the process is the range of operating conditions for which the system is to be designed. Good practise is to design with some margin for seastate, fouling and displacement growth. This margin is traditionally taken into account for by the empirical factor called Service Margin. Besides the operational factors that are included in the Service Margin, there are many design uncertainties involved in the matching process that are normally not taken into account. This paper shows that these uncertainties can result in a significant uncertainty of the operating points of the propulsion plant, which can easily lead to overloading of the driving machine or to early onset of cavitation. As a possible mitigation of this uncertainty, the benefits of an adjustable bolted propeller (ABP) are analysed. As will be shown, an ABP can be used to compensate for uncertainty during the design process, as well as for increase of ship resistance during the ship's life. An analysis of shipboard waste heat availability for ballast water treatment Heat treatment of ballast water is one of the many treatment options being explored. This analysis has tried to assess the heat availability from the cooling water, exhaust gases of the engines and steam condensers based on design and operational data obtained from an existing crude oil carrier. Time requirements for ballasting and treatment using the seawater and condenser circulating pumps are projected. Heat balance exercises were also carried out on a testbed engine to verify attainable heat recoveries. It is seen that, although considerable heat is available, a longer time than that available during ballasting and normal ballast passages will be required for the treatment process. Research on pre-controllable safety factors of a wind power gearbox transmission based on reliability optimisation design The reliability of a wind power turbine gearbox is crucial but the safety factors are hard to pre-control quantitatively by the traditional matching method. Therefore, a complete reliability optimisation model of a wind power gearbox transmission is presented. Parameters that affect gear safety factors are classified according to the ISO6336 standard and then the safety factors are expressed as explicit functions, with tooth design parameters as variables. Based on this work, the objective function is established. Common and strength reliability constraints are considered, and Sequential Quadratic Programming (SQP) is applied to the optimisation solution. An example of a gear transmission after reliability optimisation shows that it can better reflect the designer's expectations and pre-control the level of each safety factor; therefore, this approach can effectively avoid the limitations of the more traditional method. Design and parameter estimation of a remotely operated underwater vehicle The current study aims to design a remotely operated underwater vehicle (ROV) and estimate the parameters of a systematic model. The entire vehicle comprises the framework, propulsion instrument, sensor, and communication module. The systematic model is very complicated thus making it very difficult to obtain a transfer function of the entire system. Therefore, an experimental method is used to measure the acceleration of the vehicle when traversing underwater along a straight line. Moreover, the method is used to verify the transfer function of the vehicle's model for the entire system based on auto-regressive exogeneous (ARX) and auto-regressive moving average exogeneous (ARMAX) models, and a cost function. Preview of global ballast water treatment markets As full ratification of the 2004 Ballast Water Management Convention approaches, the size of the world Ballast Water Treatment System (BWTS) market has become a subject of intense scrutiny and speculation. Twelve months following full ratification BWTS will have to be installed aboard all qualifying vessels according to a timetable depending on their ballast water capacity and age. BWTS manufacturers and vendors are interested in the commercial opportunities presented by this market, and shipowners are concerned about the logistics of installing treatment system aboard vessels within the proposed timetable. In this paper, the world commercial fleet has been sorted according to flag country, vessel type, number and deadweight tonnage in order to assess the effort required to comply with the convention when it comes into force. The information includes some current equipment and installation costs, designed to gauge the market size, which appears larger than earlier published estimates. Logistics of compliance assessment and enforcement of the 2004 ballast water convention When the 2004 Ballast Water Convention comes into force it is estimated that approximately 70 000 vessels will require functional certified Ballast Water Treatment Systems (BWTS). Certification testing to IMO D-2 regulations has involved both shipboard and land-based trials by a small number of test facilities world-wide. Compliance testing for enforcement purposes under the auspices of Port State Control, will include live/dead counts of residual organisms of different size classes in treated ballast water. However, technical problems in making counts of rare organisms, and difficulties in making live/ dead assessment of smaller non-motile organisms mean that comprehensive testing for full D-2 compliance will be a complex, time-consuming operation. Given the large numbers of commercial ships visiting several hundred ports world-wide and the limited resources for comprehensive testing, it is inevitable that more limited tiered approach to compliance enforcement will be required. An investigation into the properties of V-based catalysts for Selective Catalytic Reduction of NO with NH3 Selective catalytic reduction (SCR) is the most potential method for removing marine NOx emissions by up to 95%. This paper presents the results of an investigation using vanadium (V) based catalysts doped by metals W (tungsten), Ce (cerium) and Ni (nickel) to improve the low temperature performance of an SCR for NOx reduction. The temperature range studied was between 100 - 450°C with intervals of 100°C. A honeycomb ceramic substrate was modified by composite sol and coated with the catalysts. The study also included preparation, characterisation and activity tests of the selected catalysts, including studying the activity of the catalysts in ammonia selective catalytic reduction reaction with an excess of oxygen concentration. The investigation verifies that a mixture of the above metallic catalysts exhibit high activity for selective catalytic reduction of NOx with ammonia. Among all the catalysts studied, the mixed catalyst of Ni-Ce-W-V is the most potential one, not only because of its higher NOx conversion rate at low temperature, but also because of its wider window of operation temperature. Optimisation of the crash-stop manoeuvre of vessels employing slow-speed two-stroke engines and fixed pitch propellers The combination of a slow-speed two-stroke diesel engine and a fixed pitch propeller is the most favourable propulsion system for large seagoing vessels, since both components feature outstanding simplicity and efficiency. However, they are disadvantageous where manoeuvrability is concerned. The crash-stop performance of such vessels is characterised by relatively long stopping distances. As this affects ship safety, a research project identified the engine as the origin of the problem, so a methodology has been developed for simulating engine behaviour during the manoeuvre, incorporating simplified models for the reproduction of the relevant boundary conditions, such as a propeller model as the connection between the engine and the vessel. A variation of the engine parameters having a possible influence on the manoeuvre has shown not to be of sufficient impact or not even technically feasible. Consequently, a novel method of applying a braking torque to the propeller by means of water injection into the cylinders during unfired operation is proposed. The impact of navigation safety in Kaohsiung harbor The International Safety Management Code (ISM Code) has been in force for twelve years. Its main purpose is to ensure safety at sea, avoid human injury and death and protect the marine environment. This paper aims at improving navigational safety in Kaohsiung waters by applying Data Envelopment Analysis (DEA) based on the cases of marine casualties in Kaohsiung harbour and reviewing the implementation of the ISM Code on ship management. DEA has been widely applied to evaluate the system of multi-criterion decision-making and provide solutions for improvement. This study is designed to evaluate the relation between the marine casualties in Kaohsiung waters and the accident sites; furthermore, the disadvantages and solutions for improvement of the ISM Code are proposed. A simplified DEA model is first presented to evaluate navigation safety with a view to improving the ISM Code and reduce marine casualties. Experiments with point absorbers for wave energy conversion Although many existing and proposed wave energy devices operate as point absorbers, for which the overall efficiency depends upon the frequency and magnitude of the buoy's displacement, there is little experimental data presently available with which to optimize power outputs. A harmonic model is presented for the resonant frequency and absorption efficiency of bodies that oscillate in heave mode. The model is tested with a series of still water experiments and with monochromatic wave experiments in a laboratory flume incorporating an axi-asymmetric buoy. The results support the theoretical analysis. Empirical relationships are derived for the added mass and damping of the system. The laboratory experiments demonstrate an empirical absorption efficiency in excess of 80% at resonance. Adaptive navigation systems for an unmanned surface vehicle This paper reports the design of two potential navigation systems for use in an unmanned surface vehicle (USV) named Springer. The approaches adopted are based on fuzzy multisensor data fusion (MSDF) and multiple model adaptive estimation (MMAE) algorithms with adaptive capabilities. A general description of the Springer USV is given along with details of its navigation sensor suite. Of particular interest are the three different types of electronic compass used to supply heading information. Using a system identification technique, state space models of the compasses are derived for use in a simulation study to compare the navigation systems. From the results presented, it is concluded the fuzzy MSDF algorithm is better than the MMAE methodology in terms of heading (yaw error) accuracy and robustness. Influence of accelerations on Compander foundations on modern LNG carriers Fixing heavy machinery such as Companders (compressor and expander arrangement) on modern LNG carriers requires careful examination of holding-down arrangements. Currently there are no specific rules regarding this, and the methods for estimating accelerations differ widely between Classification societies and the International Gas Code (IGC). In highlighting the need for a detailed review of holding-down arrangements, a comparison of methods of acceleration calculations is made. A new method was found using kinematics of rigid bodies and its impact on the location of equipment is analysed. Weather routing optimisation - challenges and rewards The existing practice of weather routing has been analysed, as well as the problems of creating a rational ship route optimisation technique. The principal preconditions for successfully solving this complex optimisation problem are explained. A formalised technique has been developed for creating rational ship routes based on a local climatic database, intended to reduce the time needed to formulate weather scenarios during a voyage simulation exercise. It is based on the combined use of analytical geometry and features of computer graphics. An algorithm is presented for optimising a ship's route based on using principles of dynamic programming. The results of field trials of using software based on the algorithm, which were conducted during voyages of a medium-tonnage tanker, are described. Numerical analysis and visualisation of energy flow by intensity vector method for a fluid-loaded structure The vibration energy flow in a fluid-loaded stiffened plate and the structural-acoustic coupling from the energy flow point of view are investigated using the intensity vector technique. The spatial distribution of the vibration and acoustic energy flow is visualised to show the position of energy source, the direction of flowing energy and the amount of radiation sound energy. The numerical results show that fluid loading changes the vibrational and sound energy flows. A spectrum plot shows that the structural-acoustic energy flow under air loading condition is generally at a smaller rate than that for the under water loading condition. From graphical plots, it is the sound energy flow, not the structural vibrational energy flow, that shows clearly the structural vibration mode shape. Furthermore, significant discrepancy is observed between the sound energy distribution near the surface of the plate and the vibration energy in the stiffened plate. An external damper significantly influences the vibrational and sound energy flows and the damping control strategies are investigated. Visualisations of the energy show that the damper should be placed close to the energy input source and more damping resulted in more energy dissipated. Application of delay-time analysis via Monte Carlo simulation This paper presents a methodology for the application of delay-time analysis via Monte Carlo Simulation. The aim is to demonstrate the worth of delay-time analysis and how its application can provide engineers with more information when making maintenance decisions. The methodology has been applied to two case studies. A subjective risk management approach for modelling of failure induced ship vibrations Performance degradations (failures) onboard ships have a significant contribution to ship hull/machinery vibration (SHV) which may lead to marine accidents. Due to the complexity of risks of SHV, conventional Quantitative Risk Assessment (QRA) techniques may not often be capable of providing sufficient risk management information to minimise the risks of SHV. In this study a subjective novel approach is developed by combining discrete fuzzy sets with Analytical Hierarchy Process (AHP) to deal with management of SHV induced risks. The causes of each failure event are compared with each other in terms of failure likelihood, failure capability, failure recovery incapability and failure consequence probability to achieve the relative importance and overall risk estimation of each cause. Finally, relevant Risk Control Options (RCOs) are introduced and the effectiveness of each RCO is evaluated to minimise the risks of major causes which create SHV. The results of this paper reveal that the marriage of discrete fuzzy sets and AHP is capable of producing the information to manage the risks of SHV. Stretching the Future Surface Combatant: Examining the affordability benefits of a twin-variant ship A ship concept design is presented that utilises a single common hullform across two distinct variants to meet the perceived needs of the Royal Navy's Future Surface Combatant, while satisfying the wide separation in target cost between the variants. This concept emerged from commonalities between the requirements of the Force Anti-Submarine Warfare Combatant (C1) and Stabilisation Combatant (C2) vessels. Both share some requisite capabilities, leading to specific areas where common systems can be adopted. However, other capabilities lead to major differences in the technical solutions that must be employed. While a single base hullform has significant cost advantages, to successfully fulfil both roles while maintaining commonality a number of design innovations are necessary. This paper also estimates the cost savings that could be achieved to adopt the design features proposed. These savings and the flexibility inherent to the design make the vessel well suited to the export market. A final discussion highlights the beneficial role of the twin variant concept in providing part of the affordable future fleet. A quiet revolution in Hunt-class MCMV operation and support Since 2007, a quiet revolution has been taking place in Hunt-class MCMV operation and support. At the beginning of that year, the Surface Ship Support Project established a Joint Project Team (JPT), which brought together a small group of people from both industry and the MoD with the aim of driving down the cost of support while improving availability. After achieving success with HMS Middleton, this team expanded its horizons to take responsibility for the whole Hunt-class, and formed the nucleus of the Hunt-class Output Management (COM) Team, which produces efficient engineering to deliver capable ships to support the operational programme. This paper describes the changes that have taken place and the operational circumstances that have driven their implementation. Development of multi-fuel, power dense engines for maritime combat craft The NATO and UK Single Fuel Policies outline the intent for a single battlefield fuel (JP8) to minimise fuel logistic requirements in theatre. This policy also establishes targets for the removal of petrol (gasoline) from operations by 2015. In response to this policy, the Royal Navy established its own objectives for the removal of petrol-reliant equipments within the maritime domain. The Diesel Engines Group, with cooperation from EP Barrus Ltd, developed a 44hp outboard engine prototype, which is capable of operating on naval fuels. This 44hp project is part of an overall development programme that aims to provide a family of outboard engines that will operate on fuels that will continue to be supported by platforms at sea, while still meeting the necessary performance requirements for small boat operations. Engineers remain challenged to provide robust, non-petrol burning units at the required powers while still meeting the necessary levels of equipment power-density for each application -- standards that were historically simple to meet with conventional, carburetted, two-stroke, petrol outboards. This paper discusses the 44hp project and the challenges facing the team for the continued development of a family of power-dense, non-petrol outboard engines. The nuclear propulsion of merchant ships: Aspects of engineering, science and technology This paper first considers the underlying nuclear physics and then explores the potential application of that science to the propulsion of merchant ships. It then examines the options for the exploitation of nuclear technology and considers some of the engineering implications of deploying the technology. Consideration is then given to the application of nuclear propulsion to a series of ship types, including tankers, container ships and cruise vessels. In each case two sizes of ship are chosen, one of fairly conventional size and the other much larger. Demonstration and evaluation of a retrofit urea-SCR after-treatment system for NOx reduction in marine diesels A stand-alone retrofit urea-SCR after-treatment system has been developed employing a pre-SCR NOx sensor, exhaust flow measurement, wire mesh mixer, clean-up catalyst, open-loop feed-forward control and stoichiometric NOx reduction logic. Demonstration of the urea-SCR system on a 1991 Detroit Diesel Corporation 12L engine in a West Virginia University (WVU) transient test cell achieved a 50% weighted NOx reduction and zero ammonia slip over the ICOMIA E5 marine cycle. The cost-effectiveness of the urea-SCR system was determined, including a $50 000 capital cost and $2705/t NOx reduced annually, which is consistent with Carl Moyer program estimates. Forced commutation controlled series capacitor (FCSC) circuit applied to stand-alone wave energy conversion buoys Stand-alone buoys are used as navigation aids and for scientific marine research, including the measurement of seismographic movements of the seabed as part of a tsunami warning system. Solar power is commonly used to power up the electronic devices of these buoys. However, due to limitations in delivering constant power from solar panels, linear electric generators have been employed as an alternative/additional power source to convert the energy of oceanic waves into electricity. The variable AC amplitude and frequency of the generated voltage waveform is usually converted to DC using a simple, low cost diode bridge rectifier resulting in a low operating power factor (PF) and low power transfer capacity because of the high inductive reactance of linear electric generators. A Forced Commutation Controlled Series Capacitor (FCSC) technique is employed in this paper to improve the PF of these variable amplitude and very low and variable frequency devices. FCSC circuits have been employed in the past in power transmission networks where the voltage and frequency are fixed, but never in a wave power application where the amplitude and frequency of the AC voltage are variable. This paper provides an analytical description of the proposed FCSC converter. The method is experimentally demonstrated and evaluated using a 2.25kVA test circuit. A comparison of the performance of the FSCS circuit with a standard uncontrolled single-phase diode bridge rectifier circuit is included. A robust blade element momentum theory model for tidal stream turbines including tip and hub loss corrections Blade Element Momentum Theory (BEMT) performance models for wind turbines lead to a robust BEMT model of marine current or tidal stream turbines. It is shown that numerical convergence methods for models reported in the literature are problematic when away from the normal operating range and this paper reports a robust numerical scheme using combined Monte Carlo and sequential quadratic optimisation. The model is extended by Prandtl corrections for losses at the blade tip and hub. Results are validated against an industrial code, Garrad Hassan's Tidal Bladed Software (GH Tidal Bladed) evaluation version, and a lifting line theory model. The necessity of reactive power balance in ship electric energy systems Electric load balance is a most significant study performed at the design stage of a ship outlining her electric power plant. However, this analysis is confined only to the active power part. In this paper, on the one hand the importance of extending the electric load balance in terms of including also the reactive power is highlighted, while on the other, a relevant methodology is presented. The discussion is enriched by case studies obtained from actual ship electric energy systems. A Hybrid Diagnostic System for a Small Turbojet Engine A diagnostic system is one of the key components securing the safety of operation of a turbojet engine. There are many different engineering approaches how to design engine diagnostic systems applied in real-world conditions. We propose a novel approach for a diagnostic system design to be applied on the object of a small turbojet engine using thermal diagnostics combined with a more traditional model based approach. The combination of these two approaches together with application of methods from the area of computational intelligence create a hybrid diagnostic system, which can be integrated with a control system of the engine. The article also deals with experimental validation of the designed diagnostic system on the object of a small turbojet engine iSTC-21v in laboratory conditions. Turbojet engines can generally be considered as complex systems operating in a broad environmental external parameters’ range as well as under extreme internal thermodynamic conditions. Safety of operation of aircraft and their engines are directly interconnected. The approach to increase the level of safety has always lied in increased redundancy of its systems. Current trend in control of aircraft engines lays in the use of full authority digital engine control systems (FADEC). The use of digital processors and elements in a control system allows its miniaturization, but also puts higher demands on reliability due to new possible elements, like electro-magnetic disturbances, High speed radial marine diesel engine suitability maintenance model For the application of the general theory of maintenance suitability (reparability) on high speed radial marine diesel engines research maintainability results are presented for the engine type M 504 B2, which are the most important part of the propulsion subsystem for the missile gunboat Navy (RTOP-11 and RTOP-12). A large number of empirical data is taken from the practice and statistically processed and analysed scientifically. The aim of this study was to determine the function of motor reparability special purpose in demanding operating conditions of the Adriatic archipelago. It is scientifically proven that the normal distribution best approximates the empirical function of maintenance suitability respectively the reparability, and the mean time of the maintainability suits the mean time of the repair engines subsystems. The research results are applicable to other similar marine engines, as well to other platforms diesel engines. which can cause failures. Increased reliability is achieved mainly through redundancy in such control systems. The electronic engine control system usually consists of two control units (channels); certification requirements may however demand as much as three redundant channels. Increased safety in engine control systems is usually achieved by the following means [1]: -- Sensors backup/redundancy (dual sensors, single ended sensors, shared sensors).The presented article will specifically deal with the area of small turbojet/turbo-shaft/turbo-prop engines [2] with a conceptual design of a highly redundant control system with increased redundancy of sensors by application of virtual channels and increased redundancy of its main control algorithm channels by utilization of intelligent selectors. The designed control system and redundant elements/backup will be tested on a laboratory object represented by a small turbojet engine iSTC-21v derived from the turbo-starter TS-20. The presently used dual channel architecture with redundant sensors is shown in the figure 1. 2. INCREASED ENGINE SYSTEMS’ REDUNDANCY – THE GENERAL CONCEPT To improve safety of the two channel architecture two measures are proposed to be taken: -- Increase redundancy of sensors through virtualization of selected channels -- Solve the problem of diagnostics and intelligent selection of the active EEC channel Solution of both these problems is connected with diagnostics or fault detection in sensors or electronic engine control unit. Diagnostics of sensors is done usually through a system that validates outputs of individual sensors and selects sensors channels entering the control system as shown in the figure 2 [3]. The proposed approach in sensors validation and fault detection expands the concept by application of a quorum element method, which will act as a diagnostic system for sensors and back-up system at the same time, creating a highly redundant sensor network. The concept will be described in the following chapters and is illustrated in figure 3. The second problem, which lies in selection of the active channel of EEC, which is directly interfaced with fuel flow supply to the engine, can be solved by application of intelligent selectors creating integrated control/diagnostics architectures. By intelligent selector a decision making element based on neural networks is proposed. The resulting architecture is presented in figure 4. Source: [5] Figure 3 A highly redundant sensors network with quorum elements connected to an EEC The main advantage of using such arrangement is increased reliability and resulting safety without needing to implement a third monitor channel, because the intelligent selector is able to select the correctly operating EEC channel based on outputs from the sensor network as well as output and state parameters computed by both EECs [4]. 3. INTELLIGENT SUPERVISORY CONTROL SYSTEM The intelligent supervisory system for small turbojet engine can be decomposed into four basic parts. -- Pre-start diagnostic system. -- Online diagnostic system. -- Backup system. -- Start-up control system. All the systems operate with mathematical models of the engine as described in the previous chapters. The following figure illustrates how the intelligent supervisory system is designed on the highest level of situational control for the small turbojet engine. The diagnostic module is composed of two basic modules that utilize different inputs from the engine’s sensors (figure 6). The first one is the model based diagnostics using computational models of the selected parameters in order to compute error residuals and the second one is based on thermovision diagnostics using a thermal image of the engine and a neural network. Both systems are described in more detail in the following chapters. The error residuals are further processed by expert systems which are rule based systems producing diagnostic signals, which can be further utilized by the control system. SMALL TUBROJET ENGINE DIAGNOSTIC/ BACKUP SYSTEM The small turbojet engine iSTC-21v also serves us for testing purposes of redundant backup/diagnostic systems. The designed architecture of such system has been tested for a single engine parameter the speed of the engine. This parameter is crucial as it defines thrust and power output of the engine and is the primary controlled parameter. The main way to measure the speed of the engine is the optical sensor, while the other ways are synthetic model values: -- successive integration dynamic model, -- a neural network. The basic designed architecture of a simple single parameter diagnostic system is shown in the figure 7. Reliability of the model is secured through independence of input parameters while utilizing virtual engine models to compute the speed. There are two basic errors that can occur with the optical sensor: -- A random value – caused by electro-magnetic environment disturbance, -- Sensor failure – it can be caused by a loss of power, loss of communication channel, loss of reflex area on the compressor blade. The designed backup/diagnostic system is utilizing adaptive voting majority methods and its principal implementation is shown in the figure 8. The system can exclude faulty speed computation/measurement from its output and can also indicate its own total failure utilizing precise dynamic engine models. The resulting speed of the engine is represented by the average value of all means of speed computation/measurement Ic. The designed diagnostic/backup system has been experimentally tested with the iSTC-21v engine during its operation within running on speed of 43500 RPM. During the test all input had simulated errors that are shown in the figure 9. During the test even real error of the optical sensor occurred at time of 30 seconds. The output of the diagnostic/ backup system however was not influenced and has operated as desired. Further expansion of the system will lead into a highly redundant diagnostic/backup system utilizing the presented concept, where all important engine parameters (temperatures, pressures, fuel flow, thrust) will be mutually backed up, thus creating a highly redundant backup system. 5. APPLICATION OF NEURAL NETWORK IN CLASSIFICATION OF THERMAL IMAGES In order to use neural networks for classification of pixels in a thermal image, it was necessary to prepare a training set for the network in order to create an applicable classifier. Five training images were selected for this reason in individual phases of the engine’s operation and were deemed satisfactory. This number of images was sufficient to demonstrate pixel’s classification into the individual levels, which may be critical in terms of safety for the operation of the engine and which pose no threat. As a part of the research three types of graphic images were chosen to find, which type of an image will be the most appropriate for the application of neural network: binary, grayscale, RGB. Because the thermal images captured also a part of exhaust gasses from the engine iSTC-21v, images were cropped from its original size 320x240 pixels into the size of 284x240 pixels. It was done to achieve more accurate results by application of a neural network on the pixels of the exhaust nozzle only, otherwise neurons would also cover an area, which is not important in classification of the thermal image. The individual steps of application of a neural network were realized and programmed in Matlab. The steps included the following ones: -- image read, -- digitalization, Source: [6] Figure 9 Diagnostic/backup system test results -- colour conversion -- matrix transformation, -- preprocessing, -- normalization. Topological structure of the Kohonen map with five neurons Kohonen self-organizing map (SOM) neural network was selected to classify individual image pixels after all previously described steps were applied to every thermal image [7, 8]. A topological grid size of 1x5 neurons was found in pilot experiments and topological structure of the network was a chain with Euclidian distance between neurons. The chosen type of grid can classify each image pixel into one of five clusters. In this way, we can obtain five different categories, which can be further labeled. The weights of the network were found during adaptation of training images and the corresponding five clusters indicate the criticality of the classified pixel. After transformation into a visible image, it means that the network will classify colours and areas on the thermal image, where its brightest parts include critical areas because of high temperature and darkest parts are the least critical areas. Three neural networks have been created as classifiers for an individual type of training image with the following input layer. 1. Black and white training image, SOM with two inputs (x, y position of a white dot) 2. Grayscale image, SOM with three inputs (x, y, grey) 3. RGB image, SOM with five inputs (x, y, R, G, B) Because of the distribution of individual weight vectors (neurons), it was not possible identify which pixel points on the image belong to the individual levels (weights), so each weight was assigned a color. Red colour represents the most critical area, blue color the least critical area. The following part of the article presents results of the application of the Kohonen neural network with previously described structure on the individual types of graphic images. Classification of pixels on the binary type of image (figure 11), it is possible to observe that the activity of neurons gradually goes through each segmented areas of the exhaust nozzle, which do not mean classification areas on the nozzle according of criticality level, network takes into reasoning only their spatial distribution. Size of the areas depends on the thresholding algorithm parameters and thus the resulting diagnostic information will be defined only by the size of the area classified by a particular neuron. The results of classification on a grayscale image (figure Source: authors Figure 11 Classified binary image 12) show that the image is classified according to the level of grey and spatial disposition. The colour image shows red areas, which are critically heated, while blue colour represents areas, which are colder and not critical. These areas can be considered as safe in terms of nozzle activity. The third classifier was trained on a RGB image set and the results of classification of individual areas by activating different neurons can be seen in the figure 13. The RGB image shows that the critical area of the exhaust nozzle is larger. We consider the RGB classifier as the most perspective and precise as it has larger input information in classification. The article presents a novel method in turbojet engines’ diagnostics using thermal vision and a model based approach, integrating them both in a single diagnostic system. Further Source: authors Figure 12 Classified grayscale images Source: authors Figure 13 Classified RGB image 92 R. Andoga et al: A Hybrid Diagnostic System for a Small... integration is done by interconnecting this system to the engine’s control system allowing the concept of engine control in all its operational states including control during failure states. This has a potential to increase the safety of operation of turbojet engines and this is paramount to flight safety. As the design is modular any number of diagnostic approaches, like vibro diagnostics, can be integrated in the presented way, further increasing redundancy and safety as a result. Gear of Hydraulic Pumps for Ship Equipment The aim of this article is to analyse the involute gear of a ship´s gear pump. The gear applied in a direct gear is usually designed to comply with requirements for sustainability, reliability and operability under unfavourable working conditions. These requirements can be achieved by adjusting the axis distance, angular correction, as well as by rearranging the pressure angle. These modifications are directly interconnected. Modern ship equipment contains hydraulic pumps that offer considerable advantages over mechanical energy transfer. For consistencies sake, basic pumps and standard sizes have been introduced. Producers are mainly interested in the price, technical parameters, noise level, reliability, and sustainability. The design and production of modern hydraulic pumps requires large amounts of capital investment in advanced technologies. To determine whether a possible capital investment is worthwhile it is appropriate to assess the financial situation of the company, for example by using financial ratios or different benchmarking methods [1]. In order to make returns on this investment modern hydraulic pumps are therefore produced in long series. During the design process close attention is paid to the results of basic and applied research, which have been proven in practical experiments. In practice, gear pumps are much more affordable in price than axis pumps, although the parameters are more difficult to regulate. The parameters are influenced by the engine revs. Gear pumps are the most commonly used rotary hydrostatic pumps. The first mention of a gear pump dates back to 1600 and is ascribed to J. Keppler [2]. The next reference dates to 1604 and is ascribed to a Prague clockmaker, J. Buergi. However, the gear pump first started to be effectively used by Pappenheim in 1636. Gear pumps are popular for their simple construction, sustainability and reliability. They do not carry sensitive distribution elements or even inflow. Of importance is their reversibility (the pump may function in the pump and/or engine mode). Salemm [3] dealt with the influence of the geometrical parameters of cogging and their output. As a result, an adjusted gear mechanism can cope with a smaller pulsation inflow than a standard non-adjusted one. An alternative method for the analysis of cogging with inclined cogs was introduced by Kapelevich and Kleiss [4]. The layout included non-symmetrical cogs and a gear mechanism whereby the number of cogs amounted to . Manring and Kasaragadda examined the influence of a different number of cogs and came to the conclusion that the ratio could be smaller when , and the pulsation inflow of hydraulic oil is reduced by half. The ratios of strength change on the cogging shaft bearings. The number of cogs, angle meshing, and modulus of the maximum hydraulic liquid inflow were optimized by Huang and Chen. Also of importance is the pulsation pressure of the output. The issue was investigated by Klarecil, Rabsztyn and Hermaòczyk [5]. They carried out an experiment to measure the dependency of the pulsation pressure on the shaft revs on the basis of frequency analysis. The parameters were verified in a mathematicalphysical model created by G. Dalpiaz [6]. The conclusion was drawn that the model functioned as a cog pump noise-level and vibration indicator. The basic inflow parameters for a gear pump are dependent on the revs and geometrical parameters. For a standard gear this can be expressed as follows flow rate of the pump number of teeth module addendum module bottom depth coefficient pressure angle on the reference circle rotation speed of the pump The inflow parameters for an adjusted non-standard gear can be expressed as follows Two types of corrections can be made in practice to improve the parameters. Firstly, angle correction. Secondly, rearrangement. By rearranging an adjusted gear mechanism to a positive or negative direction the following could be achieved [2,5]: -- exact axis distance; -- prevention of cog wear - cog wear is often the result of a small number of cogs, whereby the coefficient reduces contact ratio and reduces cog sustainability; -- prevention of cog angle being too sharp; -- prevention of production and assembly interference with the cog; -- contact ratio enhancement; -- specific slide enhancement; -- noise-level and cogging vibration reduction; -- increase in mechanical efficiency; -- enhancement of cogging touch, flection, seizure, wear and sustainability. Cog wear may be prevented by the enlargement of the pressure angle. The minimum number of cogs may be reduced as part of the angle adjustment. As a result, the cogs should not wear. As a matter of fact, in order for the angle of the cogs to not be too sharp it is limited to 32°. On the other hand, the disadvantage of the angle adjustment is the necessity for a special tool with an abnormal meshing angle. Such an adjustment can be used on cogwheels for which the production is on a large-scale [7]. -- Interference means a limited condition where the profiles of the cog meshing simultaneously engage a certain part of the space – overlap and collide with each other. -- Interference with a transition curve comes when the head of the cog collides with the transition curve of another cog (this curve represents a side of a cog without a cut). -- Head Interference means a cog head collision of the cogwheels. -- Operation Interference happens when a cogwheel does not delineate its trajectory; however it follows the transition curve of the cooperating cogwheel. The interfering cogwheels carry a different transmission ratio. Hence, an additional dynamic strain occurs. -- Production Interference means a head collision of the cogwheel with the transition curve of the tool. In fact, the production tools are designed in order to rule out any possible production interference. On the other hand, production interference may occur as long as the size of the cogging heads has been significantly changed. On the whole, the interference collision does not seriously endanger the functioning of the gear mechanism; it only in effect reduces the head size. -- As a matter of fact, there is no possibility to insert one cogwheel into another during assembly interference. This sort of interference may be carried out as long as it is possible to install the sprocket axially. However, in case of assembly interference, it is advisable to remove the adjustment. curve a – gear mechanism with the maximum acceptable cut of the sprocket cog curve b – gear mechanism endangered by the interference on the transition curve of the cogs curve c – gear mechanism with limited value of the meshing coefficient curve d – gear mechanism on the brink of interference with the transition curve of the sprocket cogs curve e – gear mechanism of which the sprocket borders with sharpness curve s – gear mechanism with equal values of the specific slides within the marginal points of the meshing line segment point A – suggested gear mechanism requires increased contact strength of the cog sides (resistance to pitting) point B – suggested gear mechanism requires increased flexibility strength of the cogs point C – suggested gear mechanism requires increased resistance to the wear of the cog sides; point C is situated on the curve of the equal specific slides. The equal specific slide may be achieved by the usage of a VN gear mechanism point D – cogwheels carry equal major adjustments and hold equal specific slides; VN gear mechanism. the gear mechanism is VN. For such gear mechanisms, the rolling circle is identical to the distance circles. The coefficient of the meshing duration of two united wheels defines the meshing continuity and is defined as the ratio of the length of the meshing segment line to the distance of the basic The coefficient of the meshing duration of two united wheels defines the meshing continuity and is defined as the ratio of the length of the meshing segment line to the distance of the basic circle for the adjusted gear mechanism V The non-standard g.m. consists of the maximum meshing coefficient For the theoretical values of the meshing coefficient, the radii of the cooperating wheels would have to be interminably large. Table 3 shows the results for two strained racks for the angles given. The minimum number of cogs necessary for the production of an external cogged tool in the shape of a cogwheel The specific slide of the wheel is given The specific slide of the sprocket base is given by The specific slide of the wheel head is given by:The specific slide of the sprocket head is given by Specific slide of the wheel base is given by The total of the specific slides is the total of the single absolute values of the specific slides as given by: The theoretical inflow of a hydraulic pump for a standard cogging – non-adjusted cogging N with an equal number of cogs within the wheel and sprocket is given by The theoretical inflow of a hydraulic pump for a nonstandard cogging – adjusted cogging N with an equal number of cogs within the wheel and sprocket is given by The pulsation inflow factor for a non-adjusted cogging N with an equal number of cogs within the wheel and sprocket is given by The pulsation inflow factor for an adjusted cogging N with an equal number of cogs within the wheel and sprocket is given by A standard gear with parameters ? = 20°, x1 = x2 = 0 will be considered in the analysis. The non-standard gear has the following parameters: ? ? 20°, x1 ? x2, x? ? 0, z2 = 12, x1 = x2 = –0.08. The calculation and graphic interpretation of the parameters for the inflow of the hydraulic liquid Q: the number of cogs, modulus, the width of the cogging and revs, has been depicted in the Figures 3 and 6. The inflow Q = 30 l/min complies with the number of cogs – 13. The following calculations consider cogging adjusted by x1 = x2 = –0.08 and maximum revs n = 3000 rpm. The inflow of 30 l/min complies with the modulus m = 2.5 mm where the inflow dependency Q is non-linear (see Figure 4). The inflow Q = 30 l/min corresponds to the width of the cogging b = 20 mm (see Figure 5). The sprocket revs n = 3000 rpm corresponds to the hydraulic liquid inflow Q = 30 l/min (see Figure 6). When comparing the inflow dependency Q of the hydraulic liquid to the parameters z, m, b, n, it is evident that the standard gear mechanism N has a heavier inflow than that of the nonstandard one. This statement is further testified by the negative adjustment V- (machine tool is inserted into the workpiece during production). Resources claim that the minimal hypothetical number of cogs for x1 = x2 = 0 and ? = 20° equation of a non-adjusted gear mechanism N amounts to z1min = 17 where the minimal actual number amounts to z'1min = 14. However, in this case the number amounts to z1 = 13 < z'1min otherwise undesirable wear of the cogs is likely to happen causing a reduction in the base strength of the cog. The dependency of the meshing angle on the hypothetical number of cogs z1min and the actual number z'1min = 14 with the acceptable wear is depicted in Figure 8. The specific slides in the base of the sprocket , in the head of the wheel , in the head of the sprocket in the base of the wheel and the total of the specific slides are depicted in Figure 9. The specific slides in the base of the sprocket and the wheel according to the number of cogs of the sprocket z1 on the invariable number of cogs of the wheel z2 = 12 are depicted in Figure 10. Source: author Figure 7 Meshing duration dependency on the number of cogs Source: author Figure 8 Dependency of min. number of cogs on meshing angle Screw cutters or rack-shaped machine tools are used for cogging production. Since the tools for the production of external cogging are commonly used in practice, the minimum number of cogs must be determined according to this kind of production. Figure 11 shows the dependency of the number of cogs of the machined wheel on the number of cogs of the wheel (Fellow) for pressure angle ? = (15°, 20°, 25°, 30°). The figure asserts that it is possible to reduce the number of cogs of the production wheel by increasing the meshing angle. However, it is even possible to reduce the number of cogs of the sprocket. The pump is characterised by high noise levels and a pulsating variable inflow. Figure 12 shows the dependency of the pulsation inflow factor on the number of cogs for varying meshing angles. The procedure confirms that the number of cogs can be doubled. Afterwards, the pulsation inflow factor decreases by 20% - 30%. According to the methodology and the analysis, rotary front cogging can be applied to a gear pump in order to obtain optimal parameters. The analysis was processed using Excel. The results have been presented numerically and graphically. It is therefore an effective tool with which to achieve good results, and in which alternative parameters can be processed. The aim of this article was to define the optimal parameters for cogging. The analysis was carried out on the basis of accurate calculations in the DS TECHNIK program. The test proportions of the cogging, as well as the limit deviation of the sprocket and wheel were prepared in accordance with ISO 1328-1 and -2, measuring cogs, pellets, cylinders. The following elements of the stress on the sprocket and wheel were calculated: touch stress, touch stress within the one-time, heaviest strain, flection stress within the one-time, heaviest strain. When taking these parameters into consideration, the cogged gear mechanism meets the requirements. Experimental Method for Marine Engine’s Emissions Analysis The world is considering various solutions to reduce exhaust emissions, the use of alternative fuels and the development of more efficient marine engines. IMO the company gave guidance for future borders of individual exhaust components, taking into account a variety of applications and conditions for the operation of marine engines. Engine manufacturers have been forced to develop new technologies, to meet the existing IMO rules and regulations that are yet to come, for the emission of harmful exhaust emissions.This paper describes the experimental procedure that support the modeling of the operating parameters of marine internal combustion engines in order to diagnose the condition, optimization of engine and reduce exhaust emissions. Special attention was given to verify the adequacy of the investigation, its accuracy and relevance. It outlines experiments that support the selection procedure tests during sea trials and later in the exploitation of the ship. Pollution of the environment is largely applied to the shipping industry in general and on the boat as transportation for navigation at sea. The ship in the fleet, is a major polluter of the environment, namely the atmosphere, which is saturated with harmful gases acting at their releasing. The combustion of fuel in the ship’s engine creates exhaust gases that contain harmful compounds such as nitrogen oxides, carbon dioxide and monoxide, and sulfur oxides and soot (Figure 1). According to the provisions of the Protocol of 1997, which is added to the International Convention for the Prevention of Pollution from Ships (ICPPS 1973), to any new type of marine engine, with standard test bench is dedicated to testing exhaust emissions, carried out by qualified companies, according to the order and the way that is given and approved according to IMO NOx Technical Code. These data and protocol tests are entered in the ‘’technical document’’ boat engine (Technical File, IMO TF Supportive document) which must meet the IMO requirements for greater reduction of exhaust emissions. If the ship’s engine meets the IMO criteria, issued by the authority: international certificate for prevention of air pollution from marine engines, or ‘’International Air Pollution Prevention Certificate’’ (IAPP Certificate) [1]. Tests were performed according to standard procedures, tests reduced data using this calculation codes for exhaust emissions, engine manufacturer’s approved method notified to IMO [1]. Data emissions were corrected according to the standard ISO environmental conditions, using the formula developed for ambient corrections. Report is submitted to an authorized company that issues EIAPP Certificate. Also, this report contains the necessary documentation for the method of performance measurement of exhaust emissions of boat engine on the test bed. The documentation: A Guideline to the Unified Technical File,Regulations for the Prevention ofAir Pollution from Ships, Nov. 2004, contains the procedures that are standard for all marine engines: -- NOx functions – prediction model of NOx emissions, -- Using the NOx function to determine tolerance, -- Ambient correction factor for marine engines, -- Procedures for measuring emissions IMO certification of marine engines, -- Calculation Code for exhaust emissions’ approved by IMO, -- Correction factors for marine engines, Once the engine is tested on test bed according to the above-mentioned procedure and when the EIAPP Certificate is issued, it is considered that the engine meets the regulations on exhaust emissions, and checks are made only through the Technical file that must register all changes to the relevant parts of the engine, important for the combustion process. During engine life all operating parameters have to be within the permissible limits. The problem is we do not measure the exhaust emissions, it is indirectly concluded that the emissions are within acceptable limits. This paper will analyse the possibilities of obtaining data and propose experimental analysis to support the selection procedure tests during sea trials the ship in operation in order to determine exhaust emissions and influencing parameters on the optimal engine performance, and to propose a method for future research in order to reduce exhaust emissions and optimization of marine engines. Examining on test bed gives the access to certain operating parameters at loads of 25%, 50%, 75%, 100%. When defining the experimental method more parameters must be included, perform research and identify the most important parameters to be used for the modelling and optimization of the engine. In medium speed and slow speed marine diesel engines, modelling the working process is far more complicated in the light engine, research and measurement are more complex, more expensive and a small number of scientific papers. The experiments are difficult, and the processes are complex, because every engine is different, every single cylinder chamber in the engine has variations in the performance which gives a variety of parameters. These engines are most commonly found in commercial shipping traffic and they practically produce the majority of exhaust emissions in maritime transport [2]. It is necessary to find an effective and convenient method for the determination of emissions and optimization. There are different programs for commercial development of the working process of the engine. Most are based on Vibe model for heat release and Woschni heat transfer model[3,11,12], but the coefficients that are proposed, have not been tested in such large engines. Also Zeldovich’s starting mechanism in nitric oxide production [3], should be checked at low speed engines. A numerical-experimental method should be used to determine the relevant data for emission analysis. All parameters important for the process should be used in modelling the process and the results should be verified with obtained measurements. This paper will cover the parameters which are given by the manufacturer and the recommended literature during questioning and assessing the performance characteristics of the engine, and also recommend the procedures and parameters that are important in the experiment to determine the exhaust emisions. This paper processes sensors which are used for enabling the implementation of experiment on board during the sea trials and later in exploitation. The proposal of the experiment is given by the protocol for conducting the experiment. Finally, the results are given to preliminary experimental analysis. 2.TESTING PERFORMANCE OF MARINE DIESEL ENGINES Each test consists of procedure where circumstances and operating conditions are determined and included in the Figure 1 The impact of emissions on climate change and the environment [1] Slika 1. Utjecaj emisija na izmjenu klime i okoliš 26 T. Juriæ et al: Experimental Methods for Marine ... calculation of engine performance. This procedure is essential for correct measurements and diagnosis of engine’s performance. The knowledge of methods of measurement, characteristics of measuring sensors and techniques for diagnosing are important for the design and optimal operation complex technical system in ship. During engine operation following parameters are observed: fuel quality and viscosity, cylinder oil, circulating oil and turbine oil quality, pressure and temperature; for each cylinder: cylinder pressure, and exhaust gas temperature; Fuel pump index and VIT index, cooling water temperature, piston outlet lubrication oil temperature; cooling water temperature, inlet and outlet for main engine and air cooler; sea water temperature, turbine exhaust gas temperature on inlet and outlet; exhaust pressure in receiver and on turbine outlet; turbocharger speed, Scavenge air pressure difference through air filter and cooler, as well as in receiver; scavenge air temperature: inlet blower, before and after cooler; engine lubricating oil pressure and temperature: system oil, camshaft, turbocharger, thrust segment; fuel oil temperature and pressure before and after filter. Important parameters to be calculated and corected according to ISO conditions are: fuel consumption, indicated pressure and power, efficiency of the engine and , turbochargers. According to the observation, engine diagnostic is performed:fuel combustion conditions, the general state of the cylinder’s chambers and the general condition of the engine. If there are disturbances, they can be found at an early stage and thus prevent further development and the emergence of failures. The main parameters being monitored during engine operation and important to determine the performance of the engine [8] are: atmospheric pressure, engine speed, draft, the middle indicated pressure, compression pressure, the maximum combustion pressure, the index of fuel pumps, exhaust gas pressure, exhaust gas temperature, pressure turbocharged air temperature turbocharged air, the speed of turbochargers, exhaust gas back pressure in the exhaust pipe after the turbocharger, the air temperature before the filter turbochargers, the difference in air pressure through a filter (if installed pressure gauge), the difference in air pressure through water cooling, air temperature and cooling water before and after drilling cooler air. During examining records will be kept and monitored whether the parameters are within the permissible limit values set by the the engine manufacturer and the classification society that accompanies the engine during the test. 3. DETERMINATION OF EXHAUST GAS EMISSIONS The most appropriate tools and principles on which the sensors converts the required physica size into the data have to be analysed and determined. The exhaust emissions of internal combustion engines depend on the process of fuel combustion, the engine status and systems to control exhaust emissions. The products of incomplete combustion of hydrocarbon particles in the fuel make up a small portion fraction of the exhaust gas [5]. The majority of NOx emissions comes from the high temperature reaction of atmospheric nitrogen with oxygen present in the combustion process. A secondary source of NOx emissions comes from nitrogen-related mixture of fuel. Proces combustion produces, almost entirely, NO, that during the expansion process further oxidates to NO2. Depending on the retention of the exhaust system (depending on the volume of the exhaust system) and the temperature of the exhaust gas, NO is typically transformed into NO2, in amounts of 5 to 7% of the total amount of NO. Emitted NOx continues oxidize in the atmosphere, getting the characteristic yellow-brown color. Depending on the formation of the jet of mixed and injected fuel, it creates the combustion chambers of different combustion temperatures for each particular engine, and due to the great influence of temperature on the formation of NO and produce various results of the NOx emissions. Furthermore, the different fuels used for combustion, burn at different temperatures, and therefore also vary in the amount of generated NOx[2]. Hydrocarbon emissions have a number of sources. Great portion of HC emissions comes from unburned lubricating oil cylinder liners and valves for leakage of fuel. For motors with a crosshead there is a possibility of lubricating oil to enter into the cylinder through the scavenging channels. Forming CO is mainly a function excess air from the air/ fuel mixture. The formation of CO strongly affected by local conditions in the combustion chamber. For this reason, a good process of mixing of fuel and air and air surplus, which is possible to achieve in tubrocharged marine engine adjusted for minimum CO emissions. Standard HC fuels are all organic origin and therefore different fuels contain different amounts of sulfur coming into the combustion chamber. During the combustion process, the sulfur is oxidized into various sulfuric oxides (SOx), principally SO2 and SO3. For this reason, SOx emissions from marine engines is a function of the sulfur content in the fuel. Furthermore, a smaller proportion of SOx has a sulfur content in the lubricating oil, which is also combusted in the cylinder. SO2 and SO3connect on themselves part of the water content in the exhaust gases. SO2 and SO3 are condensed in sulfuric acid which appears on all the cooler places in the exhaust system where it reaches the point of condensation, which is usually around channels. For these reasons, to prevent corrosion of the engine, the lubricating oil is fed with various additives to neutralize acid. SOx can be controlled either by removing sulfur from fuel or removing SOxfrom exhaust gases [5]. The content of particulate matter in the exhaust gas is made up of several different components. Solid particles are formed, apart from the solid material contained in the fuel, and liquid condensed material from nucleic remains. During the combustion process, soot is formed and the process of decomposition of carbohydrates after oxidation of the products of decomposition. Connected with conglomerates soot and ash is composed of several metal oxides. During the expansion process, and later in the exhaust system (or atmosphere) different carbohydrates and metal oxides continue to condense into particles, forming a final solid particles [1] friendly. In order to analyse the particles from exhaust gases that endanger people and the environment, adapted dilution method ISO 8178 for measuring the mass of particulate matter. This method is dilution of exhaust gases collected in the filter material which is maintained at a temperature of up to 51oC and subsequently analysed. Several methods based on the measurement of soot using a degree of opacity white filtring paper through which a fixed amount of gas. Finally, we have the optical measurement method. It is based on the percentage of light that breaks through a certain amount of exhaust gas to the instrument for measuring the brightness. The advantage of this method is the speed of obtaining results, or at any time the known state of the exhaust gases. 3.1. Detecting composition of exhaust gases and solids To determine the composition of the exhaust gas ise regulated by the norms and standards that describe measuring devices which are used to make such measurements, and measurement methods used for the measurement. The measuring method is based on the sampling and can be divided into: -- non-extractive (measuring probe and the devices are located within or on the exhaust duct andanalyze the composition of gases directly or indirectly), -- extractive (exhaust gas sample is taken from the exhaustgas and water in the device where analyses) [6]. Depending on the regulations and the size of the stationary source emission measurements are carried out continuously or intermittently. The sensors with which to determine the composition of the gas are divided according to the mode. For the detection of concentrations of certain gases there can be used different types of measuring sensors, but for the detection of certain gases there can be used only certain measuring principles. Electrolytic sensors are used for the determination of O2 in the exhaust gases, as well as harmful gases CO, CO2 or H2S. Smoke gases passing through the cathode and the chemical reaction occurring OH- ions traveling towards the anode. The flow of current is proportional to the concentration of oxygen in the exhaust gas [6]. These types of sensors commonly used in portable devices for measurement of emissions, as they are very compact, robust and does not require special conditions of work [5]. Sensors on the principle of infrared absorption are used to determine the emissions of CO, SO2, CO2 and NOx. It works on the principle of selective absorption of infrared light by the gas. Gas at the specific wavelengths absorbed by infrared light, in proportion to the concentration of the detected gas. The exhaust gases pass through a chamber through which the gas transversely through the IR lamps emit infrared rays of a specific wavelength. On the basis of the absorbed ligh it is determined by the concentration of a particular gas in a mixture of exhaust gases. This type of sensor is used to determine the concentration of CO2 in the exhaust gas [5]. The optical signal can be a carrier of measurement information and a carrier signal is depending on the measuring physical values and transfer the media, because we have two types of optical radiation. The first group consists of optical sensors based on photoelectrical effects that operate on the principles of photo emission, photoconductivity, photovoltage, photoniductive and photoionization [9].The second group of sensors absorb the photons being absorbed energy leads to a change in the material temperature sensors. The most famous optical sensors are: thermistor, bolmetar, bimetal and piezoelectric sensor. This type optical sensors for light source used LEDs and laser diodes. Solid particles represent all components of the exhaust gases that are in the solid state, the composition of which can vary depending on the composition of the fuel and thecombustion conditions. The composition of the particles is not subject to analysis, but only determine the mass concentration ofparticles in the exhaust gases[9], [2]. Determination of particulate matter (PM) in the exhaust gas canbe performed in several ways using different methods: -- gravimetricis okinetic methods-determination of the mass of particles accumulated on the filterpaper, the method determines the specific conditions of sampling, -- optical methods-using the properties of absorption and/or reflection of light particles -- electrical methods-statical electricity. Advantages of these types of devices are sensitivity to high concentrations of gases and acid compounds, and precision, as they can be calibrated continuously. These kinds of devices are mainly used for continuous monitoring of emissions from large power plants. Extractive methods are used mostly for periodic measurement, but also for continuous measurement.The extractive measurement method is characterized by sampling gas from the exhaust and analysing in a separate device. The sample gas is rapidly cooled and separated from the condensate (water vapour). The processed gas is delivered to 28 T. Juriæ et al: Experimental Methods for Marine ... sensors for detecting particular gases (O2, NO, NO2, SO2,…) In order to properly use the emission measuring device it is necessary to conduct a thorough preparation for measurement. Important parameters are: diameter of the exhaust pipe, distance from the source device, and a way of connecting devices to the exhaust pipe, measuring time and to prevent leakage of exhaust gases during measurement through the measurement hole. In conjunction with mentioned parameters, it is essential to incorporate in analyses also: ambient air temperature, humidity, sea state and the whole system of marine plants, characteristics of engine and engine process, to avoid sudden fluctuations in the system. When all of these conditions are met, then the experiment with sampling at different engine loads can be performed. While the sample is taken, exhaust gas analyser processes the data and provides insight into the state of the sample taken. Parameters can be obtained in the percentage, mg / m 3, ppm, and°C [9]. All parameters and theirs dimension must be initially determined, as well as the proper method for measuring each parameter, to avoid mistakes such as wrongly adjusted fuel or engine process, or engine characteristics. 3.3. Equipment and methods for measuring emissions on the test bed The most common procedures for measuring exhaust gas emission on the test bench is according to MAN B&W document ”Emission measurement procedures for IMO certification of MAN B & W two stroke engines’’. Table 2.specifies example of exhaust gas analysers and their working range for the measurement obtained on test bed. One of the most important parameteris the quality of fuel used in engine and fuel must be specified according to following fuel specification: ISO code: DM; Density: 0.8462 (g/mlat 15 oC); Heat value: 42.60 (MJ/kg); Viscosity: 3.13(cStat 40 oC); Sulphur content:0.49 (%); Nitrogen: 0.24 (%); Carbon content: 85(%); Hydrogen: 13.4 (%); Oxygen: < 0.1(%). The testing program described in IMO E3 cycle includes certification test at 25%,50%,75% and 100% load. The raw data was obtained directly from the engine on test bench, with the atmospheric conditions and environmental that was at that time on test bench. Duration of each test is at least 10 minutes, and the average value for each exhaust component is calculated to last for about 3 minutes of measured time. The actual value is calculated as a percentage (%) or parts per million (ppm), using zero and the calculated values for each component [4]. The most important engine characteristics used in calculation are:• pmax - maximum cylinder pressure; -- Tcool – cooling water temperature; • Tscav – scavenging air temperature; • p back – turbine back pressure -- pcom - compression pressure in the cylinder For the purpose of comparing measured values with standard characteristic for tested engine, measured values is corrected to the standard ISO environmental conditions, using the equation for the correction as it is given in the ‘IMO Technical Code’ ‘ 3.4. Exhaust gas emissions analyse during sea trial / Proper engine monitoring and optimisation of the exhaust emission imply proper measurement of exhaust gas during sea trial and during exploitation of the engine.It is necessary to experimentally determine the method that could give relevant parameters for process modelling and getting results to decide about engine condition and possibility for optimisation. MRUVario plus SE, exhaust gas analyser use extractive method for sampling gas from the exhaust-gas analysis. The processed gas is lead away to the individual sensors. A list of sensors and their accuracy are given in Table 3. 4. EXPERIMENTAL METHOD FOR DETERMINING THE OPTIMAL OPERATION OF THE ENGINE IN ORDER TO REDUCE EXHAUST EMISSIONS In order to start with experimental method, the system must first be calibrated i.e., measured and tested in order to obtain the characteristics of the system on which the experimental method must be implemented. When conducting tests, it is important to find out whether the engine obtains all the tasks it is designed for. Therefore, testing is carried out according to the purpose and location. Control tests are used to determine the values of the most important engine parameters such as power, fuel consumption, technical condition etc... Tests are conducted on new or serviced engine. Additional test, related to the consumption of fuels and lubricants, developing strength, speed and thermal state are performed for different engine working load.[ 5]. Experimental method, described in this paper, is used to verify the assumption upon the model is chosen for the calculations. This type of data obtained allows more accurate engine parameters calculations and improvement of engine design. The main targets of the above mentioned experimental method are the fuel injection pump, injection timing and fuel type and therefore the exhaust emissions. This experimental method can obtain a new achievements for the improvement of the engine construction [5], and can reduce the time required for additional mathematical analysis and interpolation whenever it is possible. Engine parameters obtained by this experimental method must be compared to the real engine parameters, which are measured, and thus able to perform the most accurate engine construction. In addition to this method, time necessary for engine tests is shorter and numbers of real time sensors are reduced without any unnecessary test repeating and excessive variation of parameters. The following parameters are measured and analysed:Tgas, Tamb, O2, CO2, CO, NO, NO2, SO2. The sensors which are measuring above mentioned parameters are described in Chapter. 3. Experimental method is performed for different working conditions of ship engines, and these conditions of operation may be linear, stepped and oscillating [5]. 4.1. An example of emissions measuring for medium speed marine engine using measuring device MRU / The measuring device MRU Vario plus SE (Figure 3.), that meets the requirements, was used for this experiment. The mentioned device processes the data obtained from the medium speed marine engine MaK and compares them with the regular data for this type of engine. This type of data obtained allows more accurate calculations and development of engine design. Due to the complexity of the problems it is necessary to compare the results obtained from simulations and evaluate measurements. Values must be in limits according to EU and IMO regulations The measured values are obtained by the experimental method which is performed on four stroke marine engine Mak 9M 32C, with sensor placed in exhaust pipe after turbocharger .Resulting data can be used to better describe the working values of measured engine and it can be summarised in tables to obtain better understanding of engine working status in particular moments. These data show how the engine behaves under different working loads, thus enabling the improvement of simulation systems that will take place in one of the simulation programs, allowing the improvement of engine performance and reduce emissions.The characteristics of the engine exhaust emissions at 50% working load are shown in Figure 5. with the smoke number which was 3 and at 100% load smoke number was 2 (measuring range is from 0 to 9). Characteristic values (O2, CO2, CO, NO, NO2, SO2) were measured in ppm unit, and it is taken 15 measurements every 3 sec. The test has found that the measured parameters reach steady values only after some time which is clearly seen on the graphs in Figure 6. It was concluded that after 30 sec. the steady values of measured parameters can be taken with sufficient accuracy. The next experiment was performed at 100% working load. The exhaust gas temperature was slightly higher at 100% load. After the working parameters obtained steady value, smokenumber was 2. Characteristic emission values (O2, CO2, CO, NO, NO2, SO2) were measured in ppm units, and again it is taken 15 measurements every 3 sec. (Figure 7). Emission levels of this tested medium speed marine diesel engineare lower than the levels required by the regulations of MARPOL 73/78 Annex VI. Comparing the characteristic values of exhaust gases following is concluded: -- carbon dioxide emission is lower at 50% load (approx. 1%) comparing it to 100% load, -- nitrogen monoxide was lower by 50 ppm at 50% working load, -- sulphur dioxide was lower for 40-60 ppm at 50% working load. During experimentengine working parameters needed certain time to reach steady values, and minimum time to perform measurement is determined. All measuring parameters and influencing parameters on the measured values has to be properly determined and then corrections has to be performed to the standard conditions according to ISO correction method. When using a calculation method for emission analysis and for modelling of engine working process [10,11,12], it is necessary to take into account all influencing parameters such as the size of the combustion chamber, the injection timing and the injection angle, the characteristics of the fuel pump and other parts of the engine, as well as internal and external parameters which affect the characteristics of emissions. For efficient implementation of future tests based on the results of the above experimental method, it is necessary to make the development of important characteristic values which will determine the emission quantities, and then that important values must be included in standard engine characteristics diagram as shown on Figure 4. This paper describes the experimental method for analysing exhaust emissions of internal combustion engines. Using the measuring device, emission level of medium speed diesel engine is analysed and it is determined that the value levels obtained by this testing are not higher than the requiring limit. Further research will be implemented and the impact of the various parameters on the working characteristics of the engine will be analysed in order to reduce emission levels. The main target of this research is to get a reliable method for obtaining information of the engine working parameters which could be used to determine the optimal operating point of the engine with regard to fuel consumption and exhaust emissions. Emergency Generator Design for the Maritime Transport Based on the Free-Piston Combustion Engine The paper concentrates on the challenge of creating a system of emergency power supply in maritime transport on the basis of a linear electrical generator operated by the free-piston internal combustion engine. Reluctance-flux electrical machine is used as an electromechanical converter. The aim of this paper is to study the interaction of the linear electrical machine and the internal combustion engine. The electrical generator’s main parameters are determined. Linear electrical generator reluctance-flux electrical machine free-piston combustion engine mathematical modelling Application of modern life-saving appliances on ships is one of the main components of the crew and passengers safety. Nowadays the lifeboat is a vessel with a variety of communication, control and determination of geographic co-ordinates systems, which requires a sustainable and long-term emergency electrical power. The use of mobile generators or batteries is the solution to this challenge. Classical electrical generators have large dimensions and weight parameters, which makes them difficult to operate in small boats. The batteries are not good to use because they do not have sufficient energy performance. At the same time the use of additional equipment that controls the charge and capacity of elements is a problem. For these reasons, the development of a compact and economical source of electrical energy for use in lifeboats is a relevant scientific and technical challenge. It is expedient to use linear electrical generators driven by internal combustion piston engines as auxiliary power source with a capacity of more than 4 kW in small vessels [1]. Application of the reciprocating electrical machines combined with engine’s cylinder improves weight, size and power of the generator. A positive feature of the reciprocating generators is the absence of crank mechanism. It reduces the system inertia and accelerates the system response to fluctuations in the power consumed by the load. The cylinders do not have a mechanical connection and can be made independently. It allows building a generating system of electricity on a modular principle. Possibility of independent operation allows connecting and disconnecting separate modules, to optimize power generation system, to increase the capacity of autonomous power supply system [2-4]. The following goals of this research were formulated by analyzing the standby generator application and the demands for his energy, weight and size parameters: • choosing the optimal type of electrical machine and variant of coupling of the combustion engine used in the generator; • elaboration of the electrical machine construction used in the generator; • elaboration of the control system for the generator to reach high energy, weight and size parameters. GENERAL REQUIREMENTS FOR THE LINEAR ELECTRICAL MACHINE USED AS A POWER GENERATOR Using the linear electrical machine as a generator has some peculiarities. Firstly, the power converter, which is used with electromechanical energy converter, must be able to work in both motoring and generating operation. Secondly, the battery shall be provided in the system. It must be able to accumulate enough energy to ensure the work of electromechanical energy converter during motor mode. It is necessary to take into account some factors complicating the use of such generators: • discreteness of magnetic fields in both stator and moving element (as an edge effect) reduces the energy transfer from the edge zone; • weight of the movable construction is restricted, that limits the amount of available active material and the size of available area of the active air gap. It can limit the power of machine for given dimensions. The electrical machine must be able to work in motor mode to ensure the start of the internal combustion engine. Switching modes should be realized dynamically at runtime. We can mark the following main variants of coupling of the linear electrical machine with the free piston engine (fig.1) [4-6]. The simplest variant is the connecting of the linear generator with the cylinder of engine on single side with one free end (Fig. 1.1, a). Its advantage is the possibility of implementing both twostroke and four-stroke engine of any type (Otto, Diesel, Trinkler) with controlled compression ratio and ignition moment. The peculiarity of the free end system is that forces of linear electrical machine operating in the motoring mode move the piston during all strokes except power stroke. It is necessary to develop the force to overcome the inertia of moving parts of the system and opposing force generated by the compression stroke. To implement these strokes, energy must be stored in the energy accumulator. During a power stroke, linear electrical machine works as a generator. It must provide for the transfer of energy as a load and recharge energy storage to provide power at the next stroke. The limit of such system is the internal energy exchange between the system “electromagnetic field - power converter - energy accumulator” and moving part. To ensure system performance linear electrical machine must develop a major effort, that increases its weight and dimensions and, in some cases, make it unrealizable. The energy accumulator must have a significant capacity and be able to accept and give it during the strokes. It excludes the use of batteries and requires the use of supercapacitors. The power converter must work with high currents and must be optimized for the capacity, which is bigger than the capacity supplied to the load. As a result, the cost of the system increases, its technical and economic parameters deteriorate. One of the ways to improve the efficiency of generators is the exclusion of the energy exchange by electromagnetic field between the energy accumulator and the moving element and its realization within the mechanical system. It can be achieved by connecting the free end of the linear electrical machine with the elastic element, which will accept the kinetic energy of the moving part and will store the part of the energy for the next compression stroke. In this case, the exchange of energy through the electromagnetic field is missed, which reduces the requirements for their efforts and storage capacity of electricity. It is advisable to use a gas spring as a mechanical energy accumulator (Fig. 1, b). The advantages of the gas spring are mechanical forces balance, which facilitates the design of supports as well as the reducing of the weight and dimensions of the linear electrical machine. The further development of design with a gas spring is the disposition of engine cylinders at both ends of the linear electrical machine according to the oppositional scheme (fig. 1, c) The double-cylinder engine of the oppositional scheme has a piston block composed of two pistons connected by a rigid rod. However, the using of such structure evolves difficulties connected with the implementation of the algorithm of energy storage and control the generator frequency. STRUCTURE VARIANTS OF LINEAR REACTIVE RELUCTANCE-FLUX ELECTRICAL MACHINES Analysis of possible implementations of the linear generator connected with combustion engine shows that the structure with free end has some disadvantages associated with the complexity of the control over the distribution of power and high norms of converter components subsystem. In the case of opposite cylinders, placement generator has simpler inverter, but it has a rather complex system and the control algorithm. The most efficient system of coupling the combustion engine and linear electrical generator is the gas spring structure because of the realization simplicity of mechanical subsystem and the simple control algorithm [4]. We will use it during our research. a) b) c) Figure 1 Electric generator based on free-piston internal combustion engine: a) single side with one free end; b) single side with Gas Spring; c) both sides with opposite cylinders placement Slika 1 Elektrièni generator pokretan motorom s unutrašnjim izgaranjem sa slobodnim klipom: a) jedna strana s jednim slobodnim krajem; b) jedna strana s oprugom za plin; c) obje strane s cilindrima na suprotnim krajevima 80 P. Kolpakhchyan et al: Emergency Generator Design for the Maritime Transport... It is necessary to choose the type of the electrical machine when designing electrical reciprocating machines for heavy conditions. The main criteria for choosing are structure simplicity and high relative indices. Comparative analysis of electrical machines shows that synchronous machines with permanent magnets and reactive inductor machines correspond to these criteria. Electrical machines with permanent magnets have the best specific parameters, but the use of permanent magnets limit the area of application of this type of electrical machines under heavy duty conditions. The permanent magnets retain their magnetic properties at temperature below 200°C. During the work of the reciprocating electrical machine with the combustion engine the temperature of the rotor can exceed 200°C. It reduces the reliability of the synchronous electrical machine with permanent magnets. In addition, shock stress and vibrations provide demagnetizing effect on the permanent magnet. Moreover, the high cost and complexity of mounting the permanent magnets impose additional economic and technological constraints. In spite of the fact that reactive inductor machines have lower power density, we decided to design a heavy-duty reciprocating electrical machine based on reactive inductor machine. This choice was made due to the simplicity of design, high reliability and the ability to work in difficult conditions. Two types of linear reactive reluctance-flux electrical machines were examined: the cylindrical stator-to-rotor gap electrical machine and the plain stator-to-rotor gap electrical machine. The flat stator-to-rotor gap electrical machine’s disadvantages are the considerable consumption of active materials and the construction complexity. The use of cylindrical stator-to-rotor gap electrical machine is optimal. However, it is necessary to consider that increase of the length of the translator needs to strengthen its flexing resistance because the air gap between the stator core and the moving element core must be minimum and it must be bordered only by the electrical machine’s structure to improve efficiency of the reactive inductor machine. The air gap is 0.2 mm in most reactive inductor machines. In cylindrical stator-to-rotor gap electrical machine the bush bearing provides the air gap. THE MATHEMATICAL MODEL OF THE SYSTEM PROCESSES IN QUESTION The correct choice of structure type, parameters and principles of the control system determines the reliability and efficiency of the electric generator on the basis of a free-piston engine [7- 9]. We should determine the operational frequency, the peak displacement of the moving element from the equilibrium state and the generated force. For the purpose of this research due consideration of the mechanical and dynamic processes in the system “free-piston combustion engine - linear electrical machine – gas spring” [8-10] is required. Since the performing analysis is preliminary and its purpose is to define the requirements for the electrical machines, the system in question is represented as a single-mass system with one degree of freedom. The piston force, the spring stiffness and the electromagnetic force act on the moving element. The dynamic equation has the following form: POWER STROKE MODEL AND DETERMINING THE PISTON POWER To determine the piston power operating on the moving element it is necessary to concentrate on the dynamic pneumatic process in the engine. In the power generator with one cylinder and gas spring, the use of only a two-stroke working cycle is possible. In that case full cycle ends in one stroke of the piston. The description of dynamic pneumatic process was based on the approach outlined in the [8-10]. When the piston moves toward the combustion chamber after exhaust valve overlap, an adiabatic compression of the fuel-air mixture occurs. The following expression describes the dependence of the cylinder pressure from the piston moving in this mode: h – the piston position (piston position farthest from the combustion chamber is taken as zero); P0 – the ambient pressing; Vcc – the combustion chamber volume; S?, hp – piston surface and piston-stroke, ka – heat capacity ratio. After the position corresponding to the desired compression ratio is reached the compressed air–fuel mixture in a gasoline engine is ignited and the power stroke starts. Isochoric heat supply occurs in the working medium and the pressure increases in the cylinder. The piston starts to move in the opposite direction. In this mode the pressure in the cylinder is described as: cylinder pressure at the start of the expansion process During the exhaust stroke, the exhaust valve is open. This action expels the spent fuel-air mixture through the exhaust valve. The cylinder pressure can be described as: the exhaust valve position; Gsc – the exhaust intensity parameter. The force acting on the piston during the power stroke is determined using the expressions (1) - (4) as the product of the cylinder pressure by the surface of the piston in dependence on its position and direction. POWER CONVERTER SUPPLYING RELUCTANCE-FLUX ELECTRICAL MACHINE AND ALGORITHM OF RESULTANT ELECTROMAGNETIC FORCE / The winding power of reluctance-flux electrical machine under the current is carried out by the single current. Therefore the converter shown on the figure 2 is used as its power. It consists of two singlephase half-bridge autonomous voltage invertors with independent connection of windingsconnected in parallel to the DC link . Both the motor and generator mode of each module of the electric machine can be realized with the use of such converter. The power converter of the two-phase reluctance-flux electrical machine CONTROL SYSTEM STRUCTURE Control system of linear electrical generator based on the free-piston engine is made according to the principles of descendant control (fig.3). Its structure and principles correspond to [12,13]. The loop controlling the position of the moving element is external. Required position of the moving unit is given to its input in the form of a sinusoidal signal with the working frequency and amplitude corresponding to the half-length of the power stroke. The velocity supplied as a reference to minor loop is the output loop position. Information from the sensor position of the movable element and its derivative is used as a feedback for these circuits. The force from the output goes on the input of the control system of the inductor electric machine. Depending on the position of the moving element, the value of the required winding currents to obtain a predetermined force is determined. The required winding current value is generated using the pulse-width modulation. TECHNICAL SPECIFICATION FOR THE LINEAR ELECTRICAL MACHINE COUPLED WITH THE FREEPISTON ENGINE The ratio between the cylinder and piston stroke cannot vary widely, because it leads to the deterioration of the engine characteristics in the combustion engine with crank mechanism. This ratio can vary over a considerable range in the free-piston engine. That is why three variants of the ratio of the diameter of the cylinder and the piston stroke with an equal amount of work are considered. For the electrogenerating system of 10 kW the free-piston engine with working volume 400 cm3 and a compression ratio of 8.8 was adopted. The pressure at the beginning of the power stroke was determined so that the mechanical power was 14 kW at the operating frequency of 50 Hz. The dimensions and the figures of variants under the question are given in the table 1. Calculated parameters of the free-piston engine The high degree of interaction and mutual influence processes in the subsystem consisting of converter power, linear reactive inductor electric machine and its control system, is represented in the form of an electromechanical transducer and is replaced by a periodic link of the first order. The time constant of this link is equal to uncompensated time constant of the force control loop. Its value determines the speed of the force control loop and it is one of the electro-mechanic converter parameters. Taking into account the parameters of modern semiconductor devices the converter modulation frequency in the selected power gains 10 kHz. Therefore, in subsequent calculations it was assumed that the time constant of the electromechanical converter is 0.5 ms. The controller synthesis of linear electrical generator control system has been performed for the electromechanical converter variant described above and for the parameters of free-piston engine. PID and PI -regulators are used as speed and position controllers respectively. Calculations were performed for each of the variants under consideration. The moving parts mass of the system is 5 kg, the gas spring stiffness was determined from the resonance condition at the operating frequency. Figure 4 shows the results of the calculation for a piston stroke of 90 mm. The calculation of processes developed by the limited electromagnetic force was made to optimize the electromechanical converter. It was found that the system retains functionality while limiting forces to +/- 3000 N. However, the situation when electrical machine cannot ensure the power take-off from the system can emerge if the parameters of the process in the engine cylinder or other factors are changed. In this case, there is a risk of uncontrolled growth of fluctuations and emergency operation of the system. Therefore, to ensure reliable operation the electromagnetic force limit must be set to 20-30% over the limit value. Fig. 5 shows the results of calculation of processes developed by the limited electromagnetic force of +/- 3000, and figure 6 shows the results of calculation of processes developed by the limited electromagnetic force of +/- 4000 N. Results of the calculation for a piston stroke of 90 mm a) displacement and speed of the moving element; b) the force acting on the piston and the electromagnetic force; c) instantaneous mechanical power As evident from these results, the linear electric machine must develop force in the range from 8600 N to 2600 N with used free piston engines parameters. Electric machine, designed for such efforts will have considerable weight and dimensions. The similar calculations were made for a piston stroke of 60 mm and 120 mm. As a result, it was determined that the electrical machine must ensure the force from -18200 N to 8000 N with the piston stroke of 60 mm. The limited system maintains the availability to the force of +/- 8000 N. The limited force must be +/- 10000 N to ensure the correct work The results of calculation of processes developed by the limited electromagnetic force of +/- 3000 and a piston stroke of 90 mm: a) force acting on the piston and the electromagnetic force; b) instantaneous mechanical power. The results of calculation of processes developed by the limited electromagnetic force of +/- 3000 and a piston stroke of 90 mm a) force acting on the piston and the electromagnetic force; b) instantaneous mechanical power. of electrical generator. When the piston stroke is 120 mm the developed force is in the range of -4300 – 1500 N. The minimum and operational limited force read +/- 1800 N and +/- 2200 N respectively. Analysis of the results shows that the increase of piston stroke reduces the force developed by the electric machine for the same operational volume of the cylinder and the oscillation frequency. As the mass and the size of linear reluctance-flux electrical machine depend on realized force, it can be rational to increase the piston stroke. The piston stroke of 90 mm is the most efficient in terms of the mass and size minimization of the electrical machine. The reluctance-flux electrical machine was designed; the figure 7 shows its 3D model. This model consists of two equal units. This electrical machine has a complex configuration. Consequently, it is appropriate to use the 3D modelling to determine the electromagnetic field distribution and the developing forces. To verify whether the electrical machine size and the electromagnetic load have been determined correctly, the calculation of electromagnetic field distribution was made using the software program COMSOL. Figure 8 shows the displacement distribution in the active elements of one of two units of the electrical machine model. There are two positions of the moving element. The positive feature of the reciprocating generators is the absence of the crank mechanism. It reduces the system inertia and accelerates the system response to fluctuations in the power consumed by the load. The most efficient coupling of the free-piston engine and the linear electrical generator is the gas spring structure because of the realization simplicity of mechanical subsystem and the simple control algorithm. During the work of the reciprocating electrical machine coupled with the combustion engine the temperature of the rotor can exceed 200°C. It reduces the reliability of the synchronous electrical machine with permanent magnets. The hits and vibrations provide demagnetizing effect on the permanent magnet during the reciprocating movement of the moving element. Therefore, the use of reluctance-flux electrical machine is efficient. The augmentation of the piston stroke allows reducing the requirements of the electrical machine force. The 400-sm2 system with the piston stroke of 90 mm has the best mass and size figures for the electrical generator output power of 10 to 12 kW. The developing force must be 3500 – 4000 N. CONTROL SYSTEM STRUCTURE The displacement distribution when the moving element is a) in the extreme position; b) in the intermediate position. The graphs of the electromagnetic force were made after a series of calculations for different combinations of the moving element positions and the current in the windings (fig.9). The analysis of 3D modelling results shows that the electrical machine can obtain the force sufficient for the stable work of system provided the amount of current does not exceed the limit value. The positive feature of the reciprocating generators is the absence of the crank mechanism. It reduces the system inertia and accelerates the system response to fluctuations in the power consumed by the load. The most efficient coupling of the free-piston engine and the linear electrical generator is the gas spring structure because of the realization simplicity of mechanical subsystem and the simple control algorithm. During the work of the reciprocating electrical machine coupled with the combustion engine the temperature of the rotor can exceed 200°C. It reduces the reliability of the synchronous electrical machine with permanent magnets. The hits and vibrations provide demagnetizing effect on the permanent magnet during the reciprocating movement of the moving element. Therefore, the use of reluctance-flux electrical machine is efficient. The augmentation of the piston stroke allows reducing the requirements of the electrical machine force. The 400-sm2 system with the piston stroke of 90 mm has the best mass and size figures for the electrical generator output power of 10 to 12 kW. The developing force must be 3500 – 4000 N. The use of the electrical generators based on free-piston combustion engines improves the weight, size and energy parameters of the emergency power supplies on the rescue vehicles. Large Bore Low Speed Marine Diesel Engine Cylinder Corrosion Action Analysis Current trends in the development of maritime transport and technology in general, and the ship owners’ pressures themselves, have forced the producers of large marine two-stroke low-speed diesel engines ( by means of the structural modifications) to use low quality fuels oils. These fuels contain substances which under certain conditions can become very corrosive. Today, generally speaking, for the propulsion of merchant ships the use of low speed two-stroke diesel engines in the power range between 5.000 and 80.000 kW, which use very poor quality fuel oil. Combustion of these fuels in the engine cylinders inevitably leads to the formation of electrochemical and dry corrosion, friction and wearing of cylinder elements and exhaust pipes. This paper will analyze the emergence of corrosion and the protection used today against the effects of corrosion in the cylinders of large marine two-stroke low-speed diesel engines. Onboard Switching of Fuel in the context of Safety of Navigation Requirements for reducing air pollution from ships and limiting of SOx emissions from ship’s exhaust system in some areas resulted in the use of ‘low-sulphur fuels’ (instead of the previously common ‘high-sulphur’ fuel), which allows ships compliance with legislation. However, it has caused difficulties in the operation of diesel engines as a consequence, particularly when fuel switching. As the fuel switching occurs when entering or leaving emission controlled areas, which are usually loaded with high traffic, the difficulties are reflected on the safety of navigation. The aim of this paper is to analyse the causes of onboard fuel switching problems, and based on the results obtained and published researches reviewed to propose solutions to prevent their occurrence. Keywords: safety of navigation, marine fuels sulphur content limit, fuel switching. Failure Diagnostics of Marine Engine Fuel Supply System A marine engine fuel system is described as one of essential, yet we have to be aware of its complex role on today’s marine market, economically, ecologically and energetically. Improvement is developed in many different ways. Sulphur oil amount gravitate to zero with constant regulation of TBN oil number, fuel is better purified, emulsion of the fuel and water is injected into the cylinder, SCR catalytic exhaust gas reduction is developed, hybrid dual fuel engine is developed and finally today’s different engines are called intelligent. That type of engine is capable to analyse, anticipate and resolve the problem in real time without human influence. The results of experimental determination of the actual flow rate external gear pump show that increasing pressure reduces the value of the actual flow rate and volumetric efficiency of the pump. On the basic of hydrodynamic characteristics of the pump, mathematical expressions for the actual flow rate and volumetric efficiency have been given. The applied measuring system is described and measurement results have been presented. In accordance with Poiseuille’s law, assuming that the physical properties of liquid are constant, a little pressure on the suction side of the pump is ignored. By the method of approximation of measurement results using linear regression defined model of the actual flow rate of the pump. On the basis of a defined experimental model for different values of pressure of the pump, values of volumetric efficiency of the pump have been calculated which differ less than 1% of the experimentally determined values, what justifies the model application. The resulting model allows the determination of the actual flow rate of the pump and volumetric efficiency as a function of pressure at the pressure side of the pump. Keywords: external gear pump, actual flow rate, volumetric efficiency, linear regression Definitions of the theoretical and actual flow rate of the external gear pump Experimental determination of the flow rate Equipment used for measurements and measurement data analysis The modern approach of the exploration of dynamic characteristic of elements or system of some device has been enabled by system dynamics. System dynamics is an excellent software system for studying the dynamic of the behaviour of natural, technical and societal realities i.e. system of various natures and features in which there exist relatively high analogy rate. The methodology of its work, including even the use of digital computers has shown to be an efficient mean for solving the problems: management, behaviour, sensitivity, flexibility and studying of dynamics of the behaviour of the systems of high complexity by computer simulation, in laboratory what means without danger for the realities observed. For the usage of system dynamics in the solution of complex problems, various complex knowledge from various scientific disciplines are required, but good knowledge of the nature of the problem itself, which is being analyzed. The paper has analyzed dynamic characteristic of cooling system in marine diesel engines by means of system dynamics. The cooling system of propulsion of MDE is a complex system onto which various disorders act and due to their changeable act result into the changes of working temperature of MDE. System dynamic approach in the analysis of MDE enable us to obtain the answer to the question “What if...” in other words we arrive at the solution what the reaction would be at the changes of any variable in the cooling system. The paper has shown: dynamic and static regime of work, mathematic model of the cooling system of MDE, simulation model in POWERSIM programme language and diagrams of the flows of water. Key words: simulation, diesel engine propulsion, cooling system a) Relative change of the power of diesel engine b) Relative value of the changes: - temperature of gases Y1 - openess of the walve at the entrance in water-cooling - openess of the walve at the entrnce in central-cooling plant Speed of relative changes of water temperature: - at the entrance from diesel engine (dX1_dt), - for cooling of water cooling plant (dX2_dt) exit from cooling plant - at the entrance of water cooling plant for cooling of the shell of marine diesel engine (dX3_dt). b) Relative values of water temperature: - at the exit from diesel engine - for cooling water cooling plant (X2) (exit from cooling plant), - at the entrance of cooling water plant for cooling shell of marine diesel engine (X3) Simulation of Performances of the Diesel Engine while Cutting off Fuel Delivery in one Cylinder Availability of propulsion diesel engine such as power or any other characteristics, under different exploitation conditions is determined by calculations, and is confirmed by empirical methods. Computer simulation models are very important for such analyses. The computer-simulation model of ship propulsion diesel engine based on mathematical model has been presented in the work [1]. The model is applied in the analysis of the static and dynamic operating conditions of the turbo-charged slowspeed diesel propulsion engine and the fixed blade propeller. The model was created by applying computer program MATLAB 7.0.4.-SIMULINK, and according the detailed mathematical zero-dimensional model. In this paper, the aimed model is used for analyzing the influence of the power changing of one of the cylinder on the engine performance. The power changing has been simulated by cutting of fuel delivery to one cylinder under different loading conditions and by the different limit of fuel rack. By analyzing the results of the simulation, the limits of engine working parameters and safety boundary conditions have been determined. Keywords: diesel engine, simulation, fuel delivery Shift of fuel rack depending on engine speed, with limits Change of engine characteristics while cutting off fuel delivery in one cylinder Changes of operation point in the charger during cutting off and restoring again fuel delivery in one cylinder under 100 % of engine load (Limit set up at 75 mm) Voith-Schneider Propeller The paper shows Voith-Schneider propeller, as one of the most convenient propulsion devices installed on ferries, tug-boats and other vessels where sophisticated manoeuvrability peformance is required. Beside the description of the device, it is given the overview of the forces acting on the blades, resulting in the propulsion of the vessel and review of materials for blade. Key words: propulsion device, blade, propeller, forces acting on a blade, path of a blade, materials for blade. Fuel Supply Systems in Marine Diesel Engines The paper analyses fuel supply systems in modern marine diesel engines where engine operation is controlled electronically. The basic elements of fuel supply system in MAN-B&W marine engine are analyzed. Unlike traditional method of engine operation by camshaft. MAN-B&W company has developed the system of electronic operation of working process, the main task being the control of fuel injection, opening and closing of exhaust valves and cylinder lubrication. Electronically controlled MAN-B&W engine is based on separately driven fuel injection pump and exhaust valves so that exhaust valve lifters and oil fuel atomizers are controlled electronically by means of a definite numbers of control units making together “engine control system”. Besides the analysis of fuel supply system in MAN-B&W engine, common fuel oil supply in Wartsila engines, so called Common rail system, which is also used in Sulzer RT-flex engines, has been presented. Finally traditional and electronic engine operation have been compared. Key words: injection, fuel, electronics, control, supply, efficiency Crankcase Explosion in Marine Diesel Engine This paper deals with a problem which occurs in the operation of marine diesel engine, because of unprofessional handling, or sometimes in bad course of events, it often results in the total damage on the engine. In serious incidents it may result in bad injuries and even death of the ship's crew. Therefore, more attention must be paid to the phenomenon of crankcase explosion in maritime diesel engine. Attention has been paid to this problem and the way to avoid it successfully has been found to ensure the safe work of the engine. Special attention should be paid to the safety equipment on the crankcase all of which for the purpose of reducing risk of accidents during the work in the engine-room. The oil mist monitoring and detection is also a part of this paper. Key words: marine diesel engine, crankcase explosion, safety in crankcase, detection and monitoring. Method of Correction of Ship's Propeller The paper has described the examination of the main engine of a training vessel on testing desk after the general overhaul. After the installation into the vessel, the examinations were carried out on the system of propulsion on trial voyages which discovered the impossibility of achieving nominal number of revolutions and at the same time, overstraining of the main engine. The technological model for the diagnosis of breakdowns of the system of propulsion was established. It was revealed that it was caused by inadequate vessel's propeller. Namely, the existing vessel, which was redesigned into a training one, has been exploited for three years with wrongly chosen vessel's propeller. By means of a special approach, the calculation of the correction of the vessel's propeller was carried out as well as its adaptation with adequate technological operations at disposal. After the description of the adaptation, the *mr. sc. Željko Kurtela, Sveuèilište u Dubrovniku **dr. sc. Vedran Jelaviæ, Sveuèilište u Dubrovniku *** Ante Markoviæ, dipl. ing. brodogradilište Viktor Lenac, Rijeka (u mirovini) results of the trial voyage have been displayed which proved the application of diagnostic model, the calculation of the correction of adaptation and the correctness of the operations carried out on the vessel's propeller. Key words: vessel, vessel's propeller, main engine, correction. Status and Trends of Heat Pumps Exploitation This paper deals with historical development of heat pumps from 1834 till today. Present applications of heat pumps in residential and commercial buildings, and in industry are specified. Application in the area of heating/cooling is presented through dependence of energy efficiency of different systems and their working temperatures. Comparison of heat pumps use in various European countries is also presented. Ecological and energy reasons for their use is analysed while advantage of described technology is presented on example of realized pilot project from national energetic program. Potential sources of heat are considered with criteria of their applicability. Convenient new working fluids are reviewed. Key words: heat pumps, heat sources, new working fluids, energy efficiency Generally for ships it is said that they are complex technical systems, because they are composed of a large number of subsystems and elements. The ship is a complex technical system, open and realistic system with its dynamic and deterministic properties. It consists of tens of thousands of subsystems, assemblies and elements of different exploitation reliability and maintenance suitability, but also of different importance for the operational availability [1], [3]. Technical system operational availability stands for probability that the system, when used under the specified conditions, will function satisfactory at any point of time, where time observed includes the usage time tk and malfunction time tz of the technical system. Operational availability Or is determined with a following relation for which it is intended, including failures and problems that occur in the repair and maintenance of engines. System time usage tk is actually the time when the exploitation reliability of technical systems is expressed, while malfunction time tz is the time when suitability maintenance of the technical system is expressed. Warships, as specific marine ships, are normally presented in two conditions: operational and non-operational status, and therefore under the maintenance suitability imply actually probability that the warship which appeared with malfunction, to be returned to the operating state during scheduled malfunction which makes active repair time, logistics time and administrative time. Operational availability respectively the return of warship to the operating state depends primarily on the type of warship and marine subsystems state; therefore, in extension the attention will be aimed to the particular category of warship and ship subsystems and their mutual interconnections, which basically can be serial or parallel. Analogously, according to (1), the engine maintenance suitability imply actually probability that the engine, on which a failure appeared, is to be returned to the operating state estimated time of motor malfunction tzm. Therefore, the maintenance suitability of the engine M 504 B2 also designed the engine’s ability as reparability system to maintain (through preventive maintenance) or to return (through corrective maintenance) engine operational availability Orm, which is determined by the aforementioned relation (1). It is absolutely clear, that to return the projected capacity after the failure, during corrective maintenance tkodr must be less or equal to the engine malfunction time tzm, ie. that the following relation is valid Engines logistic time tlm includes the time of purchase and completion of spare parts, engines administrative time tam includes time in which the person responsible for the maintenance makes administrative activities necessary for the successful implementation of maintenance, and engines maintenance active time tavkm includes: time during the dismantling and preparation tdmp, the time required for defecation perceived failure tdk, the time of repair and assembly tokm and the time required for functional engine testing after assembly. Accordingly, the following relation is valid: Following the foregoing, the table 1 is made showing the corrective maintenance active medium time tavkm sr of individual subsystems engine M 504 B2. Data on corrective maintenance active time are shown in table 1 and were collected during research of reliability high speed radial diesel engine M 504 B2, in cooperation with the staff of the NCP from Šibenik Group (NCP – Repair shipyard “Šibenik” d.o.o. – repairs, maintenance, modifications, new yachts up to 75 m/1,500 t), which are directly involved with the organization, supervision and implementation of all maintenance engine M 504 B2 [13]. On the basis of the table 1.1 a diagram is made of cumulative values of corrective maintenance active medium time tavkm sr i i = 1,2 ... 10 per individual engine subsystems M 504 Dismantling the engine from the boat – 130 hours, Overhauling the engine – 7494 working hours, Engine service – 1260 operating hours, Buckle revision – 748 hours, Turbocharger service – 340 hours, Bench test – 550 hours, Assembly on board – 240 hours. This paper describes the developed quasi-dimensional numerical model, implemented in the 0D numerical model for direct injection diesel engine simulation. Quasi-dimensional model uses direct solution of equations for cylinder pressure and zone temperatures, without numerical iterations. Numerical model validation has been performed on measured working parameters of the diesel engine with direct fuel injection. After a successful validation, in which simulations have been examined, the movement of various operating parameters in the engine cylinder has been performed. Operating parameters movement for the whole cylinder and the zone without combustion is shown. Except the displayed operating parameters, numerical model monitors thermodynamic processes that occur in spray fuel packages (volumes), from the beginning of the fuel injection into the cylinder, until the opening of the exhaust valves. The developed numerical model goal is to monitor a large part of the engine operating parameters, which have a major impact on the engine working process. Some of them would be difficult or impossible to measure with the existing measuring equipment. Numerical model offers accuracy and precision in the engine operating parameters prognosis. Calculation of single engine process takes less than a minute on a conventional personal computer. Quasi-dimensional model development starts from the initial division of the space inside the cylinder into two zones – a zone of combustion products and a fresh mixture zone [1, 2]. Progress in quasi-dimensional modeling occurs at the moment when the cylinder volume division is performed in a manner that during the fuel injection, packages (volumes) that accompany each fuel spray are created and outside the fuel sprays there is a zone without fuel (zone without combustion) [3, 4, 5]. Fuel spray packages are annular in shape, spatial creations and in the spray core they have a form of a truncated cone, Figure 1. As injectors can have a plurality of nozzles, separate volumes are created for each of the fuel sprays, which may be mutually identical or different. The basic assumption of these models states that between fuel spray packages any exchange of mass and energy is not allowed. The only allowed mass exchange is air entrainment from the zone without combustion into spray packages [6]. The division of combustion space into zones enables a more accurate monitoring of temperature distribution. Knowing the exact temperature in certain cylinder zones is essential for the calculation of nitrogen oxide formation. Kinetics of these reactions is very dependent on temperature and requires exact knowledge of temperature fields. While choosing the appropriate model assumptions, it is necessary to be very responsible and aware of the possible consequences. The consequences are reflected not only on emissions, but also on other engine process parameters. In quasi-dimensional, as well as in the other numerical models, the most difficult task is to properly define pressure and temperature changes that occur in the engine cylinder. The changes of these two values affect every zone in the cylinder space volume, which is not known in advance, and this fact leads to a complex mathematical model. Consequently, a lot of quasi-dimensional models developed so far do not have a direct, but an iterative calculation of these two values. In this paper, the quasi-dimensional model developed in [7] and implemented in the existing 0D model [8], to the model presented in [6], has been used. The numerical model by its specificity and complexity is reflected in the fact that equations were developed for a direct solution of pressure and temperature changes in the cylinder, without the necessity for time consuming numerical iterations. Such a numerical model allows the analysis of all diesel engine process parameters. With the indexes which are related to each package (i = axial index, j = radial index), Figure 1, it was necessary to use an additional index k for each fuel spray when the fuel sprays are not mutually identical. The mathematical quasi-dimensional model is based on the presumptions of the multizone model [9]. With mathematical excerpt, the following differential equations of pressure and temperature changes in the cylinder have been obtained: The variables E, F, G, K3 and K4 in the equation (1) are substitutes for differential expressions, and marks S1, S2 are the replacement for the sums that need to be inserted into the equation for the pressure change (2). Detailed mathematical excerpt are presented in [7]. The index i is an index for any observed volume (for each package of each fuel spray as well as for the zone without combustion). For the fuel spray packages, it should be noted that all of the displayed equations are related to the thermodynamic volume of the package (volume of gases and vapours). Thermodynamic volume of the package is the geometric package volume reduced for the liquid fuel volume. Fuel vapour in this model is considered as an ideal gas in gaseous mixture with other species. Several measurement sets have been carried out and for numerical model simulation, the measurement set shown in Table 2 has been chosen. The parameter for this numerical model calibration is the change of cylinder pressure, as opposed to some other authors who take the rate of heat release as a calibration parameter. The reason for this selection is that the change of cylinder pressure has been obtained experimentally, and the rate of heat release has been calculated from the measured cylinder pressure changes by using the adjusted 0D model which uses the linearized submodel for calculating the properties of the operating substance. Quasi-dimensional model uses a different method for calculating the operating substance properties, which is more accurate and which deviates from the linearized submodel. Details of the numerical model validation and of the used validation parameters are presented in [7]. As a numerical model follows a large number of different engine parameters, in this paper the results of a numerical model for only one operating point, and that is operating point 3, have been shown. Figure 2 shows the temperature of certain areas in the engine cylinder. The greatest value is the average temperature of fuel spray packages, which is a result of combustion and the most intense heat release. The average cylinder temperature is slightly lower than the fuel spray packages average temperature, which is caused by low temperature in the zone without combustion. This has been expected, because the zone without combustion is a large zone around fuel spray packages where combustion does not occur. The lowest temperature is the average cylinder walls temperature, which is essential for an accurate calculation of heat exchange. Changes in the total mass of different fuel phases in the cylinder are shown in Figure 3. The mass of the injected fuel increases during the fuel injection process. At the same time, fuel spray packages are formed. When the entire fuel amount per process has been injected, injection stops, and the mass of the injected fuel does not change from that moment. Combusted fuel, observed at the level of the entire cylinder, since the start of the fuel injection has a certain mass. The reason for this is the combusted fuel that remains in the zone without combustion from the previous process, so that, at the time of exhaust valves opening, the mass of the combusted fuel in the entire cylinder is greater than the mass of the injected fuel exactly for the residual mass in the ZWC. The specific heat at constant volume and pressure in the entire cylinder and in the ZWC show the same trends, as seen in Figure 6. Throughout the cylinder, specific heat rapidly grows during the start of combustion, and after that it is continuously declining until exhaust valves opening. In the zone without combustion the same trend is evident, with the difference in that the initial increase at the beginning of combustion for both specific heats is much lenient when compared to an entire cylinder. This is an expected phenomenon because combustion does not occur in the ZWC volume. The total volume change in the zone without combustion is presented in Figure 4. At the beginning, the operating media flow from the ZWC into fuel spray packages and the zone without combustion volume reduces proportionally to that flow volume. The flow from the ZWC into fuel spray packages becomes more intense as the packages are progressing through the cylinder, while the cylinder expansion increases the volume of the zone without combustion. The result is that the cylinder expansion has a greater impact on the ZWC volume than the flow into packages, so the overall volume of the zone without combustion increases in that area. At the end of the cylinder expansion, the flow into packages becomes more intensive than the ZWC volume increase by cylinder expansion and the volume of the zone without combustion decreases. The total ZWC volume reduction in this area is also caused by the increasing volume of fuel spray packages, which are becoming bigger and bigger. Shortly before 470 °CA minimum mass of the ZWC is achieved, and the volume which corresponds to that minimum mass remains constant until the exhaust valves opening. At the time of the exhaust valves opening, the entire cylinder content will be completely mixed and so mixed it will be exhausted out from the cylinder. Changes of pressure gradient in the cylinder are presented in Figure 5. At the beginning of the fuel injection, pressure gradient shows some fluctuations, which are caused by small packages creation. The basic assumption of this model says that at a certain moment (or crank angle) the pressure field is homogeneous for the entire cylinder. Figure 7 and Figure 8 show the change in the total number of moles for various species in the cylinder operating media. In the entire cylinder H2O and CO2 closely follow each other. Diatomic oxygen O2, observed at the level of the entire cylinder, begins to decrease with fuel vapour combustion. Upon combustion completion, the O2 amount is stabilized and, with that stabilized amount, is ejected out from the cylinder during exhaust, as shown in Figure 8. The increased OH emission is evident at the start of combustion, and in the middle of combustion it reaches its maximum, Figure 7. During expansion, the OH emission is reduced, and appears on exhaust with a very low amount. A large number of parameters monitored with the quasi-dimensional model have been the reason that the paper has presented the simulation results for only one operating point, with remarks that the model was also tested for other operating points and in them all it has shown acceptable deviations. Measuring some of the presented simulation results would be very difficult on the real engine, even with the highest quality measuring equipment. This numerical model achieved the desired goal: to monitor different engine operating parameters by using the numerical simulations with equal or smaller deviations in comparison to other quasi-dimensional or CFD models in the available literature. In this paper, the analysis of the ship piping installation effect on the possibility of mounting and on the operation of pumps was performed. Smaller or greater deviations from the project dimensions which complicate the mounting of the equipment, such as pumps or valves, appear frequently during the piping installation. Additional works on ship piping inevitably mean high additional costs and overrun of time limits scheduled for the piping installation. On the other hand, incorrectly mounted pump will have significantly shorter working life and if pump malfunction happens in an unfavourable moment, the consequences could be serious. Methodology, by which it could be verified if the measured piping deviation is acceptable before the equipment mounting, could yield in significant savings during the piping installation and exploitation. Loads that affect on the pump housing upon incorrect mounting could be calculated precisely by the application of the finite element method. In this paper, the finite element method analysis was performed on the example of a chosen segment of the ship piping where the pump housing loads were calculated as the function of the piping deviation from the project dimensions. On the basis of the obtained results, it is possible to evaluate whether the preexisting condition of piping could be acceptable or not. Directives for the continuation and systematization of this research according to the needs and requirements of a specific technological piping production and installation process are given in the conclusion. Key words: ship piping, pumps, stress analysis, finite element method Piping has been developing in parallel with the development of shipbuilding. It is considered that ship piping dates from the beginning of the use of engines for the ship propulsion, although simple piping was installed on board the oldest ships. By using steam engines and internal combustion engines and other devices afterward, the need for the installation of piping which were necessary for their operation was emerging. The first pipelines were feed water, pressurized steam and condensate pipelines, and thereafter fuel oil, lubricating oil and cooling water pipelines. With the installation of propulsion engines, shipbuilding experienced a true boom. Ships became more equipped, international conventions and classifying societies required higher safety of ships at sea, while the rising number of passengers wanted a higher comfort. Such development leaded to the evolution of different ship services with associated piping. Thereby, ships were equipped with piping for bilge, ballast, fire fighting, drinking water, hot and cold fresh water etc. The growing use of automation and the needs for a special cargo transport caused the appearance of hydraulic piping, refrigerant piping and diverse special purpose piping [1, 2]. Pumps and piping represent the ship “blood circulation” and are of vital importance for navigation and ship exploitation. Every deviation during the piping production, whether concerning piping dimensions or location, results with difficulties in the pump mounting and its earlier damage or failure. Pump repair or its replacement during navigation are not easy and simple tasks and could be risky for the safety of the crew and the ship. Therefore, it is necessary to provide conditions in which pumps and piping faultlessly perform their functions during the estimated working life while their replacement and repair are scheduled during docking and regular maintenance of the ship auxiliaries. Deviations during the ship piping production and installation are common and they do not represent a major problem in work accomplishment. Problems arise in the decision-making process on the mounting of pump on a specific piping segment according to the observed deviations. The key factor for such a decision are loads that occur by forced pump mounting or by connecting on pipes that deviate from the required dimensions. The calculation of such loads can be performed even for the most complex piping by applying an adequate numerical method where the finite element method is particularly suitable for those cases. The analysis of the piping section is presented in this paper and loads that affect pump housing due to unmatched position of the connecting flanges are calculated with the dependence on the deviation magnitude. The ship piping systems are used for the transport of different fluids in ship services such as: fuel oil, lubricating oil, seawater and fresh water, compressed air and other different fluids (refrigerants, inert gas, hydraulic oil etc.) thus providing a normal operation of the main and auxiliary engines, pumps, fans, compressors, heat exchangers and other devices. Nowadays, the design of ship piping is completely performed by computers and a complete manufacturing documentation is made according to the computer models. Such documentation contains drafts with an adequate number of views and sections for all piping in the cargo space, engine room, ship superstructure, fore and aft peaks and it is usually made for pipes over DN 32 mm [3]. Piping is frequently made in assemblies and as such the units are installed on board a ship (Figure 1). Regarding relatively large dimensions and technological processes in manufac turing which include plastic deforming and considerable heat input, deviations from the project dimensions could be expected. Parts which are going to be connected with devices like pumps should be installed with adequate tolerances in order to provide their correct mounting and reliable operation. 3 CONNECTING A PUMP WITH A PIPELINE Pumps are an inevitable part of the equipment in almost all ship systems. The most frequently used pumps are centrifugal pumps and they comply with the requests which are set to certain ship systems. They consist of the spiral housing and of an impeller fixed on the rotating shaft. An example of vertical centrifugal pumps which are frequently used on board ships is shown in Figure 2. The pump housing has flanges on inlet and outlet ports which are used for connecting a pump with the pipeline. Figure 3 shows the possible pump position regarding the pipeline described by values A, B, C and D. Values A, B, C and D could be positive or negative and they are independent of each other. Such deviations can be the consequence of inaccuracy during part production, inaccuracy upon welding pipes and flanges, deformations due to heat input, irregularly mounted supports etc. In practice, these deformations are not great and in some cases the pump can be mounted on the piping by the use of improvised equipment (lever, hand hydraulic press, chain hoists etc) by which the piping can be temporary elastically deformed so holes on flanges align in the position suitable for bolts inserting. Depending on the piping size, its material, fixing and other relevant factors, forces for such intervention could be relatively great. By removing the improvised equipment used for the temporary piping deformation, the pump housing takes over the load that was previously acting on the piping. As a consequence of such a pump mounting, different problems or even permanent damage could occur during the pump exploitation. 4 EXAMPLE OF THE SHIP PIPING SECTION ANALYSIS In this paper, a section of the ship piping for sea water suction with several branches connected to adequate pumps was analyzed. The model of the piping segment is shown in Figure 4 and it is clearly visible that the piping consists of several branches of a smaller diameter connected to two pipes of a larger diameter. The analysis covered one of the branches on which a centrifugal pump of the type CGB 100 V48 (Figure 5) is fitted, with a nominal nozzle size of DN 100, produced by Croatia Pumpe Nova d.d. The pump is fitted on the support with three bolts. According to the producer’s instructions, the misalignment between the pump and the piping axes can be so large that the bolt can easily pass through the holes on the flanges when connecting the pump with the pipeline flange. Furthermore, the flange surfaces should be spaced so much to enable the gasket to be placed between them during the pump mounting. The support surface for the pump mounting should be sufficiently rigid to prevent vibrations during the normal pump running. To prevent problems during the pump running, the pump producer requests that the propositions prescribed by the API 610 standard [4] must be complied with during the pump mounting. This standard contains information on the requirements for pumps as well as for forces and moments which might occur on pump connecting flanges. The values of forces and moments on pump flanges, depending on the nominal flange size, are given in Table 1. Pumps with flange size DN 400 (16”) and less, with housing made of steel should be able to work normally under loads shown in Table 1. The same standard regulates the orientation of the coordinate system for a particular housing construction which is the reference for values referred to in the above-mentioned table. The centrifugal pump of the CGB 100V48 type with the associated coordinate system is shown in Figure 5. According to the API 610 standard, the values from Table 1 should be multiplied by factor 2 for this pump housing construction. For the purpose of this analysis, the piping segment shown in Figure 4 was divided into smaller parts. The part of the piping in front of the pump consists of a larger and of a smaller pipe connected with the pump (Figure 6). The whole assembly will be named as pipe 1 in the following text. The other part of the piping that continues after the pump is shown in Figure 7. That part of the piping is rigidly welded on the engine room bulkhead because it is connected with the piping that is rising vertically on the upper deck of the engine room. The support of this piping segment holds the weight of other pipes which are added in a vertical direction. This pipe is named pipe 2 in the following text. The misalignments of flanges that are connected with the pump might occur during the installation of those parts of the piping. If those misalignments are relatively small and bolts cannot pass through the holes in the flanges, it is necessary to decide what method of the pump mounting will be used: an elastic deformation of the piping or the dismantling of the piping segment and an additional processing. If the elastic deformation of the piping requests forces that are smaller than forces allowed by the API 610 standard, such a procedure would save considerable resources for dismantling and additional pipes processing. On the contrary, if the pump housing is loaded with forces higher than allowed by the standard, problems during pump running or permanent damages might occur. By calculating the forces and the moments which occur under elastic piping deformation, it is possible to find out which are the maximum allowable deviations where the values prescribed by the standard will not be exceeded. The calculation of the forces and moments was performed by a finite element method and by the Autodesk Inventor computer software [5]. The same software was used for building a three dimensional model of piping shown in Figure 4 and its segments were meshed with finite elements meshes and prepared for analysis (Figure 8). Two types of boundary conditions were used in the model where the boundary condition 1 represents clamping while boundary condition 2 represents displacement in the chosen direction. Due to simplicity, the bolted joints of the piping flanges are modelled as rigid connections. The influence of the piping supports was not included in this analysis what affects on the stiffness of the whole analyzed piping segment. Thermal dilatations due to the piping heating or cooling were not taken into account because the analyzed piping deals with the seawater suction and its temperature does not significantly differ from the ambient temperature. The pipe models are meshed with an unstructured mesh of three dimensional elements where the tetrahedral finite elements of the second order were used [6]. 4.1 Results analysis – Pipe 1 The analysis is performed in a way that the free end of the flange or the boundary condition 2 position were set with the displacements in the directions of the x and z axes. The displacements varied in range from 0 to 4 mm in steps of 0.5 mm. The equivalent stresses on the model of the pipe 1 for displacement of 1 mm in the direction of the z axis is shown in Figure 9, while values of the forces and moments on the pipe flange for the whole range of displacements are shown on graphs in Figure 10. According to the obtained results, it can be concluded that he deviation in the direction of the z axis must not exceed 0.5 mm during the pump mounting, because loads on the pump housing will be higher than the values prescribed by the standard. This part of the analysis is performed only for the deviation in the direction of the z axis, while deviations in other directions were equal to zero. The results for the case of the flange deviation in the direction of the x axis are shown in Figure 11. The analyzed cases are for the displacement form 0 to 4 mm in steps of 0.5 mm, identically as in the previous example. The values of the forces and the moments which occur in this case indicate that the flange deviation must not be greater than 0.5 mm in the direction of the x axis if the requirements of the API 610 standard are followed [7]. 4.2 Results analysis – Pipe 2 This part of the analysis refers to pipe 2 which is connected on the exit pump flange. In the same way as in the previous example, the flange deviation is altered from 0 to 4 mm in the steps of 0.5 mm. The influence of the flange deviation is analyzed separately for directions of the x axis, z axis and x and z axes simultaneously. The equivalent stresses on the pipe 2 model for the displacement of 4 mm in the direction the x axis are shown in Figure 12, while the forces and the moments which occur on the pipe flange depending on the deviation in the direction of the x axis are shown in Figure 13. The results obtained due to deviations in the direction of the z axis are shown in Figure 14, while the results for the simultaneous deviation in the direction of the x and z axes are shown in Figure 15. On the basis of the obtained results, it can be concluded that a pump can be mounted if the deviations in the direction of the x axis or in the direction of the z axis are 0.5 mm or less. In the case when the deviations occur simultaneously in the directions of the x and z axes, the loads on the pump housing with a nominal flange size DN 100 would be too high even for the deviation of 0.5 mm. In this paper, the analysis of the ship piping installation effect on the mounting and operation of pumps was carried out. The problems and specifics of the piping installation within the ship systems are presented. Deviations in the piping installation are common, what is reasonable if different factors are taken into account like production technology, applied methods of joining, working conditions on board a ship etc. On the basis of the measured deviation, it is difficult to evaluate if the pump mounting with an elastic deformation of the piping will cause too high loads of the pump housing. Loads that affect on the pump housing upon incorrect mounting could be calculated precise ly by applying the finite element method and a right decision can be easily taken whether the measured deviation is acceptable or not. The results obtained by the analysis of the presented example have shown that very small deviations could cause too high stresses in the pump housing if the pump is mounted irregularly. The analysis performed in this paper opens additional questions that the authors intend to research too. The final aim of such a research is the formation of a basis of knowledge with all data concerning the possibility of mounting pumps and other equipment depending on the measured deviations at a particular place of installation included. But before forming a basis of knowledge, it is necessary to set up a system of relevant variables which have an influence on the final result and to systemize them in a suitable way. Some of them are as follows: pipe diameter, pipe length, pipe wall thickness, pipe material, support type, distance between the support and the flange, type and size of the pump etc. It must also be pointed out that such an analysis should be carried out in collaboration with the production entity (e. g shipyard) and should be adapted to a specific approach in the piping production. The benefit of a basis of knowledge so formed is beyond any dispute, because it lowers the cost of the shipbuilding and decreases the possibility of the equipment failure or damage during exploitation. HEAT TRANSFER INFLUENCE ANALYSIS ON COMBUSTION PRESURE IN TWO – STROKE SLOW – SPEED MARINE DIESEL ENGINES In the presented paper, the effect of the heat transfer coefficient on the cylinder combustion pressure has been analyzed. The analysis has been performed by using a simulation model that has been used for the research into the influence of multiple injections on the combustion processes and products in the cylinder of slow-speed marine diesel engines. For the analysis purpose, the combustion nul-dimensional mathematical model has been applied, which was validated by comparing the actual data measured on board a ship and the results obtained from simulations. The paper presents the results of the heat transfer influence comparison obtained by applying formulas for the heat transfer used by Annand, Eichelberg and Woschni. The results are statistically analyzed and graphically presented. The best choice is explained and suggested. Key words: two-stroke slow-speed marine diesel engines, heat transfer coefficient, combustion pressure HEAT TRANSFER IN INTERNAL COMBUSTION ENGINES The maximum temperature2 in the combustion chamber (cylinder) of internal combustion engines is now moving around 2500 oC, while the maximum temperatures of metals in contact with such temperatures are limited to much lower values3 and, therefore, cooling is necessary. Due to such a high temperature differences, very large heat fluxes occurred, which, during combustion, can reach around 10 MW/m2 [2]. During the other processes, the heat flux is small or close to zero. So heat flux varies in intensity, direction, time and space. The greatest cylinder heat flow is in the area where the highest combustion gases temperatures and velocities are. On these areas, it is necessary to maintain the cooling thermal load within acceptable limits. The gas side cylinder liner wall temperature should be maintained below 180 oC, in order to maintain a sufficient oil film thickness4. The heat transfer affects the engine performance, namely, the efficiency and the emissions. For the same fuel amount brought into the cylinder, a greater heat transfer to the cylinder liner wall (cooling increased) means the pressure and average combustion gas temperature drop, thus reducing the performance and the efficiency. The heat is transferred to cooling water and oil depending on which part of the process is in question. In the scavenging process, the cylinder liner walls are usually warmer than the scavenging air that has a temperature of around 40 0C and has a relatively high speed so there is a heat transfer from the cylinder liner to the scavenging air. This reduces the engine turbocharging efficiency that influences again the engine performances. In the compression stroke the scavenging air temperature rises and at some points is beyond the temperatures of the cylinder liner wall, cylinder head and piston (the exhaust valve seat temperature is still higher because there is no time to cool it), the scavenging air velocity decreases and the heat transfer from air to wall occurs. In the combustion and expansion strokes, the combustion gases reach the highest temperatures of the process, gas velocities are high and turbulent and at that time there is the largest heat transfer (flux) from the working fluid to the cylinder walls. As the expansion progresses, the gas speed and the temperatures are dropping thus reducing the heat flux. When the exhaust valve opens, due to the pressure difference in the cylinder and exhaust gas manifold, the gas velocity is increasing again and there is a sudden cylinder pressure and temperature drop, and a large part of heat, that contains combustion gases, is discharged into the exhaust gas manifold. A part of this heat is transferred to the exhaust valve walls and to the exhaust valve channel which is cooled to maintain the temperature limits, and the remaining part of the heat drives the rotor of the turbo-charger. The exhaust gas manifold is isolated and the heat transfer (lost) through it depends on the insulation material quality. It should be as small as possible in order to get more heat to be used in the turbo-charger. The combustion gases temperature change, caused by heat transfer, affects the pollutants formation process, both in the area of combustion and later in the exhaust gas system5. The heat transfer affects the turbo-charger exhaust gas thermal energy efficiency that has a significant impact on the engine power, efficiency and performance. The friction between the piston, piston rings and cylinder liner also contributes to the generation of additional heat load, and the coolant is needed to maintain the permitted limits. The heat transfer into the cylinder is applied by convection and radiation. For diesel engines, unlike gasoline, radiation can have a significant part in the total heat transfer, and it should be taken into consideration. In practical calculation and simulations, the heat transfer coefficients, experimentally obtained for each type of the engine and which contain the convective and radiation part of the heat transfer, are commonly used. The heat amount, which is transferred to the cooling water through the cylinder liner wall, ranges from 1/3 to 1/4 of the total chemical energy brought into the cylinder by the fuel. Approximately half of this heat is transferred to the walls inside the cylinder while most of the remaining half goes to the wall of the exhaust gas channel, in a case when the channel is not isolated. During the working cycle the heat flux varies in time and space. The method and selection of the mathematical description, and thus the complexity of the simulation depends on the research matter. It can be said that there are two extremes. The first, the simpler one, applies the total cooling water heat transfer calculation by using the medium heat transfer coefficient over the entire surface in time and space. For this purpose empirical or semi-empirical equations are the most common in use. The second, for the calculation a significantly heavier choice, is the case when certain engine parts6 thermal loads have to be examined.6 In such cases, the assessment of the heat flux spatial distribution and time change should be as accurate as possible and, therefore, much more complicated algorithms with multidimensional models are used. The most common cases of the power prediction, efficiency and emissions fall within the above-mentioned two extremes, i.e. the QD (quasidimensional models) give satisfactory results. A detailed effect on certain engine parts thermal loads is not observed in this paper, so that the medium heat transfer coefficient has been used. The heat time change estimation within the cylinder is important but the surface heat distribution can be much more important. The required accuracy of the assessment is not great because of the 10% error in estimating the total heat transfer during a cycle usually generates an error within ±1% when power and efficiency has toue of 10%, the influence on the CO production prediction is negligible, for NOx it is around 1%, while for CH it is around 2% [3]. HEAT TRANSFER COEFFICIENTS Studies [4] and [5] have found that the cylinder liner walls, piston and cylinder head temperature in a stationary mode is a constant one, so that the mean surface temperature can be counted on. The heat transfer coefficients difference over the cylinder liner surface can be ignored, for the purpose of this study, and the mean heat transfer coefficient accepted. The wall surface where the heat transfer takes place is equal to the cylinder liner wall exposed surface (AC,i), increased for a portion of the piston surface up to the top piston ring: where ?? is the height from the piston top up to the top piston ring. In the above expression the greatest unknown is ??. Since the gas behaviour (flow) inside the cylinder is a little known, therefore, for calculating ?? empirical formulas obtained by the experimental measurements of various cases, are commonly in use. All equations for heat transfer coefficients calculation are based on the Nusselt’s heat transfer theory. This dimensionless number is defined by the term: The choice of ?? depends on the gas velocity at the cylinder liner surface, i.e. the heat exchange between the working fluid and the cylinder liner inner surface is carried out mainly by forced convection. Thus, related to this case, in the literature a large number of expressions for ?? can be found in literature, and some of the best known are as the following ones: pressure, temperature and volume at the inlet valve closing time (in the two-stroke slow-speed marine diesel engine that is the exhaust valve closing time), where The Woschni’s expression modifications have been proposed by various authors, depending on whether it is the case of petrol or diesel engines, of high or middle speed engines, and depending on the type of process and other conditions as well. To mention only some of them: Hohenberg, Asley-Campbell, Kolesa. Schwarz, Huber, Vogel and Gerstle. In large middle-speed four stroke and slowspeed two stroke diesel engines the above-mentioned equation application gives the deviation between the measured and calculated values of the exhaust gas temperature of about 20 K [8]. In the simulation results, this is presented as less enthalpy gas supply to turbo charger which results in a lower scavenging air pressure. For large engines this is important because they are usually optimized to stationary operating point at which they most often operate. In 1999, Gerstle modified the Woschni’s equation for the case of scavenging and charging of fresh air into the cylinder. The old term for C1 is increased for the constant k = 6.5 ÷ 7.2, and is valid for a period from the exhaust valve opening till the scavenging ports closing; The heat transfer coefficient has a major effect on the combustion process simulation results and there is a need to choose the best one for a particular research. It should be noted that the determination of the most appropriate term for the heat transfer is not easy because it depends on a number of parameters. The same heat transfer equations for different types of engines do not always provide the best results. Various authors have dealt with this research, mostly with the low-power and small diameter IC engines and low piston speed which has a significant impact on the research topics. Large slow-speed diesel engines have their own peculiarities that affect, both the combustion and the heat transfer process. Therefore, when modeling the engine processes, a careful attention should be paid in choosing the right heat transfer coefficient(s), as well as the entire heat transfer formula. This paper aims at presenting the results of the research carried out on the two most common types of slow-speed marine diesel engines of the MAN B&W ME 60 and Wärtsilä RT Flex 50 new generation. It should be also noted that the local cylinder temperature are not taken into consideration, but the model is simplified by taking mean exhaust gas temperatures, mean cylinder liner wall temperatures and mean cooling water temperature for one combustion process. The best results were the ones obtained by using the Woschni’s fomula. APPLICABILITY OF DUAL FUEL ENGINES FUELLED ON PRODUCER GAS FROM THE ASPECT OF AIR POLLUTION REDUCTION AS AN ALTERNATIVE TO CLASSIC DIESEL ENGINES DUAL-FUEL TECHNOLOGY IN GENERAL Dual fuel engines operate on the so called lean burn principle2 that provides for a rather high air and fuel mixture ratio (appx. 2,1:1). The advantage of this concept is that the engine produces significantly lower emissions of nitrogen oxide NOx (<1 g/kWh) because the released thermal energy created during fuel combustion is used for additional heating of air limiting thus the combustion temperature. It must be pointed out that the air fuel ratio is kept within a narrow range: neither too high (below 1,9:1) – there will be detonation; nor too low (above 2,1:1) – there is a danger of misfiring. The engine is run on diesel fuel by injecting diesel as pilot fuel. When stable combustion is achieved the engine is switched to operation on gas. This process takes about one minute during which time gas is gradually admitted into fuel system. The engine can be switched from gas to diesel fuel during operation at any load. If the engine stops working due to insufficient supply of gas or for some other reason, the engine is automatically switched to diesel fuel operation. Once the mentioned faults are corrected the operator can switch the engine back to gas. This can be done at working loads up to 80%. Air intake and gas injection During intake gas is supplied via gas inlet valve and mixed with air admitted through an open air intake valve due to under-pressure in the cylinder when piston is descending, much like in standard diesel engines. Gas and air mixture compression The mixture will not ignite through auto-ignition under the effect of compression because the gas has high ignition temperature. Injection of pilot fuel and ignition When the piston is in a TDP a small quantity of diesel (about 1%) is injected through the nozzle on the injector. Pilot diesel fuel is autoignited under the influence of compression and thus the mixture of gas and air is ignited. Keeping up the correct value of fuel and air ratio under all working modes is of extreme importance in order to prevent detonation and delayed ignition of the mixture. The basic idea of this technology is to utilise electronic control of diesel fuel injection, which serves as pilot fuel, to enable ignition of an exact quantity of natural (or some other) gas and air mixture. Dual fuel engines are usually four-stroke diesel engines working with high compression ratio of lean mixture of natural gas and air. The adaptability of this technology opens the possibility for these engines to work fully on diesel fuel in case of shortage of natural gas. High compression ratio in diesel engines can be maintained due to high auto-ignition temperature of methane (352 °C more than diesel), which is the basic component of natural gas. While the gas SI engines there are problems in ignition of “light” mixtures, in dual fuel engines these problems are solved by introduction of pilot fuel. Ignition by pilot fuel enables creation of more ignition points, the so-called multipoint ignition, than in diesel engines where there is auto-ignition o diesel fuel which is injected into compressed air. This results in dual fuel engines being able to work with Lambda ? ? 2. As pilot diesel fuel the usual and well known diesel fuels are used, that are also utilised to run conventional FIE diesel engines. The same as in burning diesel fuel, by introduction of pilot fuel there is less need for mixing fuel and air in pre-chambers before the very ignition and combustion phase. In dual fuel engines a small quantity of pilot fuel enables minimum diesel combustion phase and consequently the fuel consumption is significantly reduced as well as exhaust gas emission. This is the first step contributing to reduction of nitrogen oxide emissions and noise during combustion process which further reduces additional thermal and structural loads. A variety of consistency and endurance tests as well as lube oil quality and wear of moving parts tests brought to the conclusion that dual fuel engines in operation show less wear and longer intervals of regular oil change (SAE972664). Since the system of combustion in these engines is based on the principles of diesel engine operation, this system represents an upgraded version of diesel technology. Reduced exhaust gas emission by reducing the quantity of pilot fuel, changing injection pressure and compression ratio, represent the procedures regularly carried out nowadays, but we shall elaborate on this later on. Presently with the advanced diesel FIE there is the possibility of delivering a certain micropilot that enables adequate gas emission reduction as well. Also, the dual fuel system is combined with the system for reduction of exhaust gas via exhaust gas recirculation (EGR) system and considerable reduction in emission of NOx has been indicated. According to available US DOE research the level of NOx of 0,5 g/bhphg was achieved through testing on US heavy-duty cycles with dual fuel C-12 engines [1]. Dual fuel engines retain the efficacy of diesel engines and at the same time emission of NOx is reduced. If approximately the same levels of thermal efficacy and compression ration are kept the total heat transfer onto the cooling medium and exhaust gas temperature values remain the same as in the regular diesel engine [1]. 2.1. Comparison of exhaust gas emissions in diesel and dual fuel engines Figures 2.1.-1 i 2.1.-2 show how lower emissions of NOx are achieved at approximately the same efficacy of engine operation by testing operation of dual fuel engine with the so called back-to-back engine test under common conditions, as well as comparison of exhaust gas emission in a classic diesel engine when operated on diesel and on heavy fuel and dual fuel engine when operated on diesel and gas. PRODUCER GAS AS ENGINE FUEL In the recent years, due to more frequent oil crises and oscillations in raw oil prices per barrel at the global market, the centre of attention is brought to the possibility of utilisation of alternative fuels to propel internal combustion engines. The interest for utilisation of such fuels is growing also due to the increased number of governmental and non-governmental organisations and associations involved in environmental protection and storage of crude fuels. Producer gas and hydrogen are two potential alternative fuels. Engines utilising hydrogen as fuel have many advantages, but there is the problem of early ignition, especially under high load [2]-[3]. Early ignition is a more significant problem in engines propelled by hydrogen than other types of internal combustion engines, due to smaller quantity of energy required to ignite hydrogen, wider ignition area and shorter flame extinguishing area. The problem is considerably smaller in engines propelled on producer gas, although this gas contains 12-20% of hydrogen. Much like any other type of gaseous fuel, producer gas can be used to proper internal combustion engines since it has the required purity of gas so that the combustion products do not leave deposits on combustion chamber walls and engine cylinder. Nevertheless, this gas has not been sufficiently used due to certain prejudices and ideas on such type of gas, such as auto-ignition tendencies at high compression ratio, too long period of time required for the engine to reach the highest output power which causes reduction in gas energy. These assumptions require additional elaboration. Following arguments can be stated against the standard approach to improved resistance against detonation: the first argument is higher laminar combustion speed that increases in the presence of hydrogen, which could reduce the risk of detonation. Argument number two is the presence of inert gases in raw gas (CO2 and N2) which could significantly reduce early ignition reactions, and this is one of the main causes of detonations. Also, the highest flame temperature that can be achieved by usage of producer gas is lower than with conventional fuels, which indicates that better resistance to detonation can be expected. An insight into available Croatian and international literature proves that the effect of producer gas on creation of detonation has not been subject of studies. Furthermore, there is a general impression that producer gas is low energy density fuel and that a considerably longer period is required to reach the highest output power of an engine by using this gas in comparison with utilisation of high energy density fuels such as liquefied natural gas and liquefied petrol gas. It is to be kept in mind that when making comparisons energy density6 of the mixture is taken into account and not fuel energy density. When compared to methane the energy density of producer gas is some 23% lower, as shown in Table 2. There are two types of plants in which producer gas is used: turbo generator system (boiler- turbine-generator – BTG) and internal combustion gas engine system. The turbo generator system is used in plants with extremely high outputs, however, when this system is used for smaller outputs the starting price is relatively high a thermal efficacy is small. Producer gas is suitable for relatively smaller gas engines due to the fact that their thermal efficacy is higher. In the past period several gas engines have been invented propelled by this type of gas and most of them have SI combustion system. SI engines are not suitable for this type of fuel under excessive loads due to problems in reaching stable combustion process due to fluctuation of components contained in the producer gas. Producer gas has been used to propel dual fuel engines in several designs [5]; however, those are naturally suction engines, i.e. engines without turbochargers installed. Starting of these engines for operation on gas or dual fuel operation will lead to reduction of power output due to reduced thermal value of the burning mixture. Thermal value of stoichiometric mixture of gas and air is approximately 2,5 MJ/m3. When this value is compared with thermal value of stoichiometric mixture of petrol and air, which is approximately 3,8 MJ/m3, the difference in power output obtained between the power of petrol engines and the power of producer gas engines is clear and obvious. The loss of power of approximately 25% can be expected especially as a result of lower thermal value of the gas and air mixture [4]. As it is well known, the main method to test engine power output is to increase the content of burning medium in the burning mixture in SI engine cylinder, or to increase fuel supply in CI engines. Nevertheless, higher inlet pressure of the mixture is of significant importance in obtaining the higher engine power output, which is possible by pre-charging cylinders. Pre-charging of cylinders does not only increase the engine output power, but also creates the possibility of lean burn. Namely, by increasing the pressure of the mixture by pre-charging cylinders with air the required quantity is reduced and consequently the quantity of gas in it, which enables reduction of nitrogen oxide (NOx) emissions. By avoiding detonation higher efficiency and engine power are achieved. Producer gas created by gasification usually contains the following components (volumetri cally given in%): carbon monoxide 18-22%, hydrogen 12-20%, methane 2%, hydrocarbon 0,2- 0,4%, nitrogen 45-55%, water vapour 4%. The lowest heating value is within 4,5-4,9 MJ/kg, with stochiometric ratio air and fuel of 1,255+0,05 (mass share). The comparison of this gas and methane is rather interesting especially from the aspect of internal combustion engine operation. Namely, most engines run on gaseous fuels can also run on pure methane (natural gas) or diluted methane (biogas). The appropriate ratio of fuel and air, i.e. stochiometric ratio of fuel and air at the ignition limit is closely compared for both gases in [6]. However, laminar combustion velocity of producer gas at the level at which in the mixture of fuel and air there is much more air than needed for full combustion (lean limit) is much higher. Laminar combustion velocity of producer gas is approximately 0,5 m/s, which is some 30% higher than methane. 5. OPERATION ASPECTS OF ENGINES PROPELLED ON PRODUCER GAS Engines operating on Otto principle, that usually use petrol or kerosene as fuel, can work independently on producer gas as well. Engines working on diesel principle can work on producer gas with some modifications, which involve primarily reduction of compression ratio and installation of spark ignition system. Another possibility for operation of diesel engines without these modifications is dual fuel operation mode, when the engine works up to 90% on producer gas and the rest is on diesel fuel used to ignite the mixture of gas and air. The advantage of dual fuel operation mode is in its high flexibility shown in the capability when in case of fault in gas operation or lack of gas for any reason the engine can normally run on diesel fuel. However, not all diesel engines can be modified for dual fuel operation. Compression ratios in combustion chambers in diesel engines are too high to perform satisfactorily the dual fuel operation mode and utilisation of producer gas in such engine greatly increases the risk of detonation as a result of too high pressures in engine cylinders due to delayed ignition. Diesel engines with direct injection of fuel have smaller compression ratios and in general they can easier be modified for producer gas operation. The output reached by the engines propelled on producer gas is determined on the basis of the same parameters as in the engines propelled on liquid fuels, such as: heating value of the mixture of fuel and air burning in the engine cylinder during one combustion stroke, the quantity of mixture burning during one combustion stroke, degree of power efficiency at which thermal energy of the fuel and air mixture is transformed into mechanical power, number of combustion strokes in a set time interval (number of rpm). Conversion of an engine to operate on producer gas or on dual fuel mode will inevitably result in reduction of power output. Heating value of the mixture Heating value of producer gas depends on the relative quantity of various fuel components contained in it: carbon monoxide, hydrogen and methane. Heating values of these gases are given in the table 3. However, to obtain full combustion it is necessary to mix the gas with an adequate quantity of air. The mixture to burn in the engine cylinder will have lower heating value per unit of volume than the gas itself. The portion of oxygen required for full combustion (stoichiometric combustion) of each burning component of gas is also given in the table. Heating value of such stoichiometric mixture can be calculation on the basis of the following: where: Hig – heating value of stoichiometric mixture of producer gas and air Vco – volume fraction of carbon monoxide in the gas (prior mixing with air) VH2 – volume fraction of hydrogen in the gas (prior mixing with air) VCH4 – volume fraction of methane in the gas (prior mixing with air) Heating value of the gas depends on the design of the gasification system as well as on the characteristics of the gas created by the process. It is important to obtain minimum thermal losses in the gasification process in order to reach the highest possible heating value of the produced gas. Two most important characteristics of the fuel are water content and the area of its distribution. Additional reasons for the loss of power can be found in the procedure of mixing gas and air, which changes the content of gas, as well as the loss of pressure in the gasification system, due to which it is extremely difficult to maintain constant stoichiometric mixture of gas and air. The only way to achieve stochiometric state of the mixture is installation of manually controlled valve at the air inlet side of the engine to provide for the highest engine output power by regulating the valve. If the highest output power is not required in most cases it is better to have the engine work with a minor excess air to prevent exhaust gases to be drawn back from the exhaust collector. The quantity of mixture supplied to the engine cylinder The required quantity of mixture supplied to the engine cylinder is determined on the basis of the cylinder volume and gas pressure in the cylinder at the time of closing the inlet valve. The cylinder volume is a constant value determined by the engine construction. The pressure of the mixture in the cylinder at the beginning of compression stroke depends on the design of the inlet collector and inlet port on the engine head, number of revolutions – the higher the number of revolutions the lower the pressures in the cylinder, and finally it depends on the pressure of gas in the inlet air collector. The first two characteristics belong to the so called “volumetric efficiency” of the engine, defined as a ratio between the actual pressure of gas in the cylinder and atmospheric pressure (1,013 bar). During normal operation the volumetric efficiency is 0,7 to 0,9. Gas pressure in the inlet air collector depends on the drop of pressure in the gasification system, i.e. its components such as cooler, purifiers and gas and air injector. This drop in pressure results with repeated reduction of inlet pressure to its value in volumetric factor of 0,9. It is to be concluded that the quantity of mixture entering the engine cylinder will be only 0,65 to 0,80 times smaller than the maximum theoretical value as a result of the pressure drop in the mixture entering the cylinder. Obviously the result will be reduction of the highest power output the engine can reach. Apart from the regulation of pressure drop in the gasification system, purification and cool ing systems, the quantity of mixture burning in the cylinder can be increased in two ways: • By increasing the volumetric efficiency of the engine by installing a larger diameter inlet air collector which will result in reduced resistance to gas flow and smaller pressure drop. The importance of properly installed inlet air collector is often undermined. The tests carried out by Finkbeinger in 1935 show that a properly installed inlet air collector can increase the highest output power of an engine up to 25% [4]. • By pre-charging of cylinders and installation of a turbo-blower. As it has been mentioned before, it is obvious that by increasing the highest pressure of the mixture on the inlet side the highest power output of the engine will be increased. The development of turboblowers operated on engine exhaust gases has made such solutions the most frequent choices on many types of engines, especially marine diesel engines. However, due attention must be given to proper and efficient operation of the cooling system on the turboblower casing as well as the bearing lubrication system to prevent creation of high temperatures and explosion of combustible mixture. Engine efficiency The efficiency at which the engine converts thermal energy into mechanical energy depends primarily on the engine compression ratio. The effect of increased engine compression ratio can be calculated as follows: where: ?1 – thermal efficiency of the engine for compression ratio value 1 ?0 – thermal efficiency of the engine for compression ratio value 0 ?1 –compression ratio value 1 ?0 – compression ratio value 0 k – constant value of 1,3 if producer gas utilised. In petrol engines the compression ratio is limited by octane number of petrol, which represents the measure of the value of compression ratio at which detonation occurs. Since the mixture of producer gas and air has a larger octane number than the mixture of petrol and air, in engines operating on producer gas it is possible to obtain higher compression ratio value and thus better thermal efficiency and higher power output of the engine. Increased efficiency of engines operating on Otto principle can be achieved by increasing the value of compression ratio from 1:10 to 1:11. Gas engines have a constant value of compression ratio in the mentioned area and for that reason they are very applicable for operation on producer gas. Number of revolutions per minute Since the highest output power of an engine is determined according to a unit of time, it will depend on the number of revolutions per minute. Diesel engines have the highest output power that is almost linearly dependent on the number of revolutions due to changes in different efficiency factors. When calculating the power of a four-stroke engine it is to be taken into consideration that during one (out of total two) revolution of the crankshaft the compression and combustion strokes take place. The value of the highest number of revolutions of the engine propelled on producer gas is limited by the combustion rate of the mixture of gas and air in the engine cylinder. Since this rate is smaller in comparison with the combustion rate of the mixture of petrol and air, the efficiency of the engine can drastically decrease if the combustion rate of the mixture equals the rate at which the piston moves in the cylinder. This phenomenon is not rare in commercial engines when working in the area of about 250 rpm, consequently engines operating on producer gas should work in the lower rpm area. Due to the low combustion rate of the mixture of gas and air the time of ignition of the mixture in engines propelled on producer gas, that are basically engines working on Otto principle, must be changed in general. Optimum ignition time in Otto engines depends on the load and rpm, and this applies to producer gas engines as well. In this mode of operation of diesel engines the problem of detonation sometimes occurs. In addition to the engines with very high compression ratio values, which can reach up to 1:16, this problem usually occurs when efforts are made for the loss of power of the engine to be substituted by increased supply of diesel oil. Depending on the producer gas system and portion of gas in the gas and air mixture, the excess pilot fuel can lead to detonation. For this reason the quantity of pilot diesel oil, in dual fuel operating mode, must be below the upper limit required for detonation to occur. In general, this upper limit of pilot fuel quantity should be set to the value at which the engine reaches 30% of its highest output power, to prevent detonation. The quantity of pilot diesel oil in dual fuel mode also has the bottom limit as well. Depending on the number of rpm a certain minimum quantity of pilot fuel should be injected in the engine cylinder for ignition. This value is somewhere between 3 and 5 m3 per stroke. In practice a larger quantity of fuel is sometimes injected per stroke for safety reasons. It is recommended that the quantity of pilot fuel injected in the engine cylinder per stroke is kept at 8-9 m3 per stroke [4]. Taking into consideration the facts that fuels obtained from biomass have played an important role in the past, when liquid fuels for internal combustion engines were expensive or unavailable, and that recently an increased interest into this kind of alternative to fossil fuels is being noticed, it is safe to say that countries investing into this technology will be able to considerably reduce their dependence on oil import. Delayed ignition of gas-diesel engines of dual fuel type is strictly dependant on the quantity and quality of the pilot fuel used. Performances of dual fuel engines can be improved by using pilot fuel with higher cetane number. Their utilisation is argumented by introduction of smaller quantity of pilot fuel and it can improve the performance of the engines operating on gaseous mixture with low heating value when compared to the existing diesel fuels whose cetane number is relatively low. A balanced operation of the engine without fumes can be achieved by utilisation of low energy gas in dual fuel engines, using at the same time different ratios of fuel and air and multi ple fuel injection points. Two-phase combustion is an indicator of the conditions under which the engine reaches the highest output power. For the main combustion in two-phase combustion less than half the time required for the usual combustion process is needed. Loss of power of 20-30% in dual fuel engines propelled on producer gas is compensated by the fact that these engines release significantly smaller quantities of toxic gases in comparison with the classic diesel engines. The release of nitrogen oxide and in particular sulphur oxide is significantly reduced as well. THE OPTIMIZATION OF THE STEAM PLANT BY MEANS OF AN ENGINE ROOM SIMULATOR The optimization of the steam propulsion plant on LNG tankers is dealt with in this paper. The operational parameter analysis was performed and the situations influencing the plant efficiency were simulated on the engine room simulator. The performed simulations on the LNG simulator module enabled monitoring of the operational parameters change in undesirable and unexpected conditions for which there is a probability to occur within the operational system. In such a way an enhanced diagnostics of failures within a real system and the determination of the optimal performance of individual components are possible. The boiler analysis has concentrated not only on the influence of soot deposit on the economizer and superheater piping to the fuel consumption but also on the influence of burning different fuels to NOx emission. Another analysis has been carried out in the main turbine in which the influence of worn out thrust bearing and low-pressure turbine blades to the turbine output, axial displacement and vibrations was analysed. The results have shown that the simulator can serve the shipping companies and maritime educational institutions to train the students and seafarers for working in various situations and to prepare them to respond to emergencies promptly and efficiently. Key words: optimization, steam plant, simulation, failure analysis The steam plant optimization is divided in two main parts in this paper. Firstly, the analysis of the main operational parameters and losses that affect the overall steam propulsion plant efficiency is dealt with. In the second part, the simulation of the impact of undesirable operational conditions in main boilers and the main turbine is analysed. The system is considered to be optimal when operating perfectly, regardless the load, and achieves the highest efficiency at minimum maintenance. The engine operational parameters analysis combined with the usage of a marine steam plant simulator gives rise to a better plant control, a more thorough understanding of the overall process and a higher overall efficiency. The satisfactory safety level and response to the possible actual situations can be achieved by constant training. The simulation of the various occurrences of a steam turbine plant on a liquefied natural gas tanker (LNG tanker) is shown in the paper. The simulations were taken on the marine engine room simulator Transas ERS 4000 ver. 7. 3. at the Faculty of Maritime Studies in Rijeka. The simulator represents a true model of the actual system allowing the execution of various tests, checks and the prevention of undesirable situations within the plant. The operational parameter adjustment and the optimization of the plant operation in various conditions are enabled by means of the simulator. Faults and failures on the individual system components and their influence on the plant operation and efficiency can also be simulated. In this way, the students are trained in a better understanding of the propulsion system operation and of the optimum operating parameters setting. Besides, experienced seafarers are prepared to respond promptly and efficiently in specific situations on board a ship. 2. ANALYSIS OF THE STEAM PLANT OPERATION The steam plant scheme, taken out from the Transas ERS 4000 ver. 7.3. simulator, is shown in Figure 1. This plant is typical for the ships carrying liquefied natural gas (LNG tankers). On these types of ships, the gas is transported in specially designed tanks at the temperature of approximately -160 °C. Despite today’s trend of the reliquefaction of vaporized cargo and the installation of various propulsion machinery types, the steam plant has still been retained on board such ships. The reason can be sought in its high reliability and in the safe use of vaporized cargo used as fuel. The plant consists of the Kawasaki UA-400 propulsion turbine of 29 450 kW output power at steam inlet conditions of 57.4 bar and 515 °C at M.C.R. There are also two Mitsubishi MB-4E boilers with dual fuel burners. The steam generating capacity of each boiler is 63 500 kg/h at M.C.R. at steam conditions of 61.5 bars and 515 °C. There are also two 3150 kW turbo generators and steam driven feed pumps in the system. The condenser vacuum is 0.963 bar. The theoretical closed cycle with regenerative feed water heating is shown in Figure 2. The heat in the steam plant is produced by means of the fuel combustion within the boiler and the mechanical work is gained on the turbine shaft. The overall plant efficiency represents the ratio between the gained work and the produced heat, as well as in all thermal engines: where P is the power on the propeller and Qg is the thermal power produced in the boiler furnace by means of fuel. If all losses (efficiencies) within the steam plant are considered, then the overall efficiency can be calculated as follows [3]: where ?t is the thermodynamic closed cycle efficiency, ?T is the turbine efficiency as a result of irreversible losses in the turbine, ?m is the turbine mechanical efficiency, ?r is the gearbox mechanical efficiency, ?GP is the thermal boiler efficiency, ?c is the pipeline efficiency. The specific fuel consumption dg [kg/kWh] of a steam plant is gained from the heat balance: where ? is the overall steam plant efficiency, Dg is the mass fuel consumption in the boiler [kg/s], Hd is the calorific value [kJ/kg]. The above equations show that the plant efficiency, thus the specific mass consumption of fuel, depends on the calorific value and on the efficiency of individual plant elements. The plant efficiency is increased and the specific fuel consumption is reduced by increasing the efficiency of individual plant elements achieved with proper maintenance and operation. There are several basic methods to increase the marine steam plant efficiency, thereby reducing the specific fuel consumption. This is accomplished by increasing the inlet steam temperature and the pressure of the turbine, by reducing the exhaust steam pressure from the turbine and by the regenerative feed water heating. Figures 3a and 3b in T-s diagram show the influence of the increased steam temperature and pressure at the turbine inlet. The diagram shows that the overall thermal cycle efficiency is increased by increasing the average temperature of the process part where the heat is brought. Increasing the input steam temperature, the wetness of steam, and thus the erosion in the turbine last stages, is reduced. As the inlet steam pressure increases, the wetness of steam, and thus the erosion in the turbine last stages, is increased. In order to avoid erosion, it is necessary to increase both the inlet steam pressure and temperature. This can be made up to the limit corresponding to the material structural features. The influence of the condenser pressure drop is shown in Figure 4. As it can be seen, the low exhaust steam pressure from the turbine i.e. the pressure drop in the condenser causes a lower temperature in the process part where the heat is rejected in the environment. There- fore, the thermal cycle efficiency is higher. The exhaust steam pressure from the turbine and the condenser pressure are affected by the size and cleanliness of the condenser cooling surfaces, ambient temperature i.e. cooling water temperature, turbine load and the steam flow to the condenser. The regenerative feed water heating increases the average temperature of the process part where the heat is externally brought, whereby the thermal cycle efficiency is also increased. The amount of steam that enters the condenser and the heat which is irreversibly rejected into the environment by means of the cooling water are also reduced. The additional effects of the installation of the regenerative feed water heaters are that the steam flow through the first turbine stages is increased and the steam flow through the last stages is decreased. This increases the turbine efficiency and reduces the losses due to whirling in the turbine last stages. The steam flow into the condenser, its dimensions and the required amount of cooling water are also reduced, but the plant requires a greater number of feed water heaters [2]. As it can be seen from the above mentioned equations, the conversion of the thermal energy into the mechanical work in the steam plant occurs with some losses. Those are the losses in boilers, turbine, pipelines, friction losses and other losses which cause the reduction of the overall efficiency below the value achieved by the diesel or combined propulsion plants. The thermal losses in the boiler operation are due to the unutilised heat from the exhaust gas, heat radiation and incomplete combustion. The exhaust gas leaving the boiler has always a higher temperature than the ambient, so the loss due to the unused heat occurs. This is the largest loss occurring in the boiler, and it depends upon the exhaust gas temperature and the excess air. This loss can be reduced by using the exhaust gas heat for the feed water and / or combustion air heating. Furthermore, care must be taken when using fuel oil with high sulphur content in order to avoid the too low exhaust gas temperature and low-temperature corrosion. The loss due to heat radiation into the environment is caused by temperature differences between the boiler external surface and the environment which causes the part of heat to transfer by convection and radiation. This loss is reduced by means of the quality insulation and marine boiler construction which are frequently built with double walls where heat is used for additional combustion air heating. The loss due to incomplete combustion depends upon the fuel oil content, fuel and air mixing efficiency and the furnace load. This loss is reduced by the proper adjustment of the excess air, fuel viscosity and by the proper handling and maintenance of burners. The loss due to soot appearance occurs mainly because of the irregular boiler handling and incomplete combustion. Poor fuel oil and air mixing and the low temperature of boiler heating surfaces enhances carbon deposition in the form of soot. These losses are reduced by proper handling of the boiler and associated equipment and by the regular soot blowing. The internal and external losses occur during the steam turbine operation. Frictional losses in nozzles, inlet passages and blades, the losses due to friction and whirling in the rotary turbine parts, the losses due to steam leakage through the turbine stages and the energy losses due to the exit velocity of the exhaust steam are considered as internal losses [3]. Mechanical losses due to bearing friction and the losses due to heat radiation from the pipeline into the environment are considered as external losses. The majority of losses which occur during the turbine operation are the consequence of the construction arrangement. Those losses can be minimized by proper maintenance. One of the main threats to the turbine rotor is the axial displacement and vibration. The rotor vibrations occur due to its imbalance which is a direct consequence of the inadequate operation, critical revolutions, wear of the turbine parts and the excessive thermal load or the rotor distortion due to uneven cooling. 3. SIMULATION OF THE IMPACT OF UNDESIRABLE OPERATIONAL CONDITIONS TO THE PLANT EFFICIENCY A marine engine simulator allows the display of various scenarios within the real system, enabling the simulation of common faults or breakdowns without any consequences to the plant. This represents an advantage over the real systems where it is technically impracticable to carry them out. The simulator can simulate a large number of different conditions on each individual system component and monitor their impact to the whole system. In such way fuel consumption, turbine output, emissions, fire in engine room, failures in automation equipment and etc. can be observed. The results can be used for the education of both the students and the seafarers who should be prepared to respond to emergencies promptly and efficiently. For the purpose of this paper, four states on the LNG tanker steam plant at full load were simulated: the influence of soot deposit on the economizer and of the superheater piping to fuel consumption, the influence of burning different fuels to NOx emission, the influence of worn out thrust bearing and low-pressure turbine blades to the turbine output, axial displacement and vibrations. The influence of soot deposit on the economizer and of the superheater piping to fuel consumption of boilers in parallel operation is shown in Figure 5. The increased fuel oil consumption (line 1) due to dirty boiler pipes is visible from the figure. The difference in the consumption shows the importance of proper maintenance, inspection and timely diagnosis. The simulator does not have the option of measuring or changing the thickness of soot deposits, so the differences in the fuel consumption are predetermined for educational purposes. The influence of burning liquid and gaseous fuel to NOx emission is shown in Figure 6. One boiler burns heavy fuel oil (1) and the other burns boil-off (2). It can be seen that it is more appropriate, from the ecological point of view, to burn methane because it emits less nitrogen oxide (approximately 35 mg/m3 ). Nevertheless, the use of boil-off as a fuel depends upon the navigational conditions, the amount of boil-off, terms of contract, the cost of gas and oil on the market, the environmental regulations etc. The influence of the worn out thrust bearing to the turbine output (1), the vibrations of the high pressure (2) and low pressure turbine (3) and the axial displacement of the high pressure (4) and low pressure turbine (5) is shown in Figure 7. The simulation below shows the reduction in the turbine output at constant steam consumption. On top of all, increased turbine vibrations and axial displacement are developed which could lead to the tripping out of the turbine. To maintain the turbine output at a constant level, it is necessary to increase the production of steam, thereby increasing the fuel consumption. Consequently, this will result in the reduction of the overall plant efficiency. The influence of the worn out low-pressure turbine blades to the turbine output (1) and the low-pressure turbine vibrations (2) is shown in Figure 8. The simulation shows that there is a rapid reduction in the turbine output and a gradual increase in the low-pressure turbine vibrations in case of the worn out blades. In practice, this phenomenon does not occur so quickly as simulated, but through out a longer period of time. If the turbine operates in such an irregular way, it is possible that the vibrations could cause the turbine to trip after some time. The phenomenon of the low-pressure turbine blade wear is usually caused by the high concentration of water in the steam (wetness of steam), leading to the blade erosion. The other possible causes of this phenomenon are the worn out blades and improper operation. The turbine blade wear can be avoided by maintaining a high boiler water quality and the appropriate steam turbine operation. 4. CONCLUSION The paper analyses the steam propulsion plant and the methods of optimizing such a propulsion system on board LNG tankers. The plant efficiency and the methods of improving the economic operation, depending on the operating parameters, were considered by the analysis of plant characteristics. The main parameters that affect the overall plant efficiency are the pressure and the temperature of the inlet steam and the pressure or temperature at which the exhaust steam condensation occurs. The efficiency of the plant substantially depends on the design i.e. on the regenerative feed water heating. The overall efficiency of the steam plant also depends on the efficiency of its individual accessories. The engine room simulator enables the analysis of various scenarios, errors, failures and adjustments of the working parameters aiming at the plant optimization and professional training. In this paper four scenarios were simulated and their influences on the plant efficiency were analysed. The influence of the soot deposit on the economizer and of the superheater pipes to the fuel consumption, the influence of burning different fuels to the NOx emission and the influence of the worn out thrust bearing and of the low-pressure turbine blades to the turbine output, axial displacement and the vibrations were simulated. The optimization, by using a simulator, enables a successful control and maintenance of the plant i.e. of those operating parameters whose adjustment results in the maximum plant efficiency and in the minimum fuel consumption. The results have shown that the simulator can serve the shipping companies and maritime educational institutions to train the students and the seafarers for working in various specific situations. THE NECESSITY OF USING MDO AND OF CHANGING OVER FROM HFO TO MDO (AND VICE VERSA) ON THE EXAMPLE OF TWO REEFER VESSELS Using the HFO1 on board merchant ships has an advantage in terms of voyage expenses where fuel costs have an important role. However, there are special circumstances in which the use of MDO becomes necessity and leads to the increased total costs. Such circumstances are recognized on the example of two ships of the same owner2 who operate in the system of one Pool3 , and that the Pool’s normative tables are shown only through the use of HFO. In this way the result of the differences in costs that would be prescribed by the rules of the participants, should be charged only to the owner. Taking into account the information found in available sources, the main characteristics of the machines in which the fuel burns on board both the ships were pointed out in relation to the usage of MDO and to the changing over from HFO to MDO (and vice versa), and thus highlighted the special circumstances for each of them, laying the foundation for the adoption of the concluding positions. After analyzing all available data, and examining the professional literature a uniquely synthesized definition for the specified case is given. Conclusive result is explained the difference in accordance with the manufacturer’s instructions and operating conditions of ships, which is made a precondition of accepting the necessity of recognition of these increased costs by Pool. Key words: HFO, MDO, changing over from HFO to MDO and vice versa. MAIN CHARACTERISTICS OF THE MACHINERIES BURNING THE FUEL O/B EACH VESSEL REGARDING THE USAGE OF MDO AND OF THE CHANGINGOVER BETWEEN HFO AND MDO (AND VICE VERSA) Main engine of Vessel 1 Referring to the data found in the Mitsui B&W Volume 1 ‘Operation and data’ manual, Chapter 4.: Other operational aspects, It. 4.2: ‘Fuel change-over’ – Pg 705.08 – 705.10, it is to be noted that the mentioned main engine is designed for the constant operation on preheated heavy fuel oil, even during standstill. It is also highlighted that the manufacturer’s recommendation is ‘not to use diesel oil for the operation of the engine’ (applying it to all loads) due to latent risk of diesel oils and heavy fuels of marginal quality forming incompatible blends during fuel change-over. There is to be noted that such blends, as well as too rapid temperature changes, can evoke problems such as: – fuel pump and injector sticking/scuffing, – poor combustion, – fouling of the gas ways. It is not mentioned in that chapter, but it is well known, how this can affect the other components of the engine (for ex. exhaust valves, piston rings, piston crowns, cylinder liners, cylinder covers). However, the manufacturer permits ‘special circumstances’ when the change-over to diesel oil becomes necessary. This is the case when ‘the vessel is expected to have a prolonged inactive period with cold engine’, for ex. due to: – a docking, – more than 5 days stop, – a major repair of the fuel oil system etc.; or – when the vessel is approaching the areas where environmental legislation requires the use of low-sulphur fuels. The owner has presented the records of the vessel’s ESP (end of sea passage) and SSP (start of sea passage) respecting the voyages completed during 2009. (Voyage Nr. 012009 – 082009), showing the average standstill of approx. 6.9 days, with notably the highest stoppage when approaching West African ports. (See Table 1) The durations of stoppage at ports estimated by the operator were not usually in accordance with the durations of the actually stoppage (usually under estimated). It is also to be noted, that part of the voyages made in that period were to the North Sea ports that are included in the SECA areas, so the engines had to be changed-over to LSHF oil (or MDO if the first one wasn’t available o/b). In those circumstances it is understandable that the vessel’s main engine should be ready for the manoeuvering when approaching the ports and the crews were obliged to follow the manufacturer’s instructions and environmental legislation required. So, for those ‘special circumstances’ (when stoppage was more than five days long or when entering/leaving the SECA areas was without LSHF oil o/b) it was absolutely necessary to change-over the fuel from HFO to MDO at the end of the sea passage and from MDO to HFO at the beginning of the sea passage (or when entering/leaving the SECA areas) to end manoeuvering in a safety manner. Main engine of Vessel 2 Referring to the data found in the Kobe Diesel – Mitsubishi UE Diesel Engine (UEC 45 LA) instruction book, chapter 2 – Engine operation, page 023-01-02, it is stated that the engine is to be started using diesel oil only. Due to that fact, it is understandable that at the end of the voyage the fuel should be shifted from HFO to MDO before manoeuvering when approaching the port, as well as from MDO to HFO at the beginning of the sea passage when leaving the port. Regarding entering/leaving the SECA areas the changing-over should be considered in the same way as above mentioned. Auxiliary engines on both vessels The usage of MDO and the changing-over from MDO to HFO (and vice versa) for auxiliary engines should be considered in a different way, but equally for both the vessels. The auxiliary engines on board both the vessels can use MDO and HFO equally. For starting of the engines, the manufacturers recommend the usage of MDO and when the engines are warmed up the changing-over from MDO to the preheated HFO can be performed. Furthermore, before stopping the engines that were run on HFO, the fuel should be switched to MDO. This is to be done due to the fact that, when stopped, there is no circulation of the fuel provided through fuel system components of the engines themselves, so any HFO left in the system will cool down clogging the system and preventing future starting. The risk when changing over from one to another fuel can cause damages which are the same as the ones mentioned for the main engine. It is to be noted that ‘engine performance optimum’ and ‘best combustions’ are achieved when the engine runs at an optimal load which is usually above 75-80% of MCR. When the engine runs at low load, its components (e.g.: fuel valves, internal combustion spaces, exhaust valves, exhaust gases lines, T/C etc.) are exposed to deposits of residuals left from the decreased efficiency of the combustion process. This occurs when using MDO and it is especially high when using HFO. Due to this fact, the manufacturers do not permit the usage of the HFO when the load applied is too low (e.g.: 30% MCR). So, if the engines are to be run with care and in good manner the operator should consider those facts respectively to obtain safely running and avoid excessive maintenance/repair costs. The deciding factor which fuel is to be used for A/Es at the port or during the voyage mainly depends on the load applied to them, that respectively depend on: the operation of the vessel, type of cargo loaded/carried and on the usage of the machineries installed (e.g.: reefer compressors, hold fans and derricks, etc.). There is a significant difference between loads on auxiliary engines when the vessel is loading/carrying frozen or chilled cargoes or cargo that doesn’t require any cooling. It is a common thing that one generator will not be enough to maintain the entire load applied, and that another one is to be run ‘in parallel’. The parallel run of the generators on both the vessels is available in ‘equal’ mode only (there is no possibility of changing the mode (e.g. ‘optimal’ or ‘cyclic’ load). That means that when the generators are ‘in parallel’ they should carry equal load (approx. 1 each). This might lead to the most common situations that neither one of them runs at the optimal load. The cases of intermittent load peaks and downs to be maintained (which are so common on the reefer vessels having three reefer compressors, plenty of hold fans, derricks, etc.), require two generators and mostly when running in parallel they carry slightly above 50% of MCR. This load is far away from the optimum level (above 75% MCR) and for a prolonged use it will cause accumulation of the deposits mentioned above, reducing the power capacity of the engine and raising the costs of maintenance. Those effects are much higher when HFO is in use. The ‘peaks’ can shift the load to optimum levels for both, which is favorable, but the ‘downs’ can reduce the load to less then 30% MCR when HFO is not allowed for use. Knowing this, it is up to the Chief Engineer o/b to decide which the best solution at the current situation is (e.g. if frequently changes of load during loading or discharging and the necessity for several starting and stopping of another auxiliary engine are to be expected then it will be reasonable to run them on MDO. Otherwise, when the load is expected to be pretty constant and near to the optimum 75-80% of MCR, the change-over to HFO will be appreciated, reducing the costs of the fuel consumed.). Regarding the pre-purchased survey reports, it is to be noted that the figures for that purpose are usually obtained from the Engine Log Book and consulting the C/E in charge o/b during the survey. Therefore, supposing that the surveys were done in the same way, and taking into account the above mentioned considerations, the values might be realistic respecting the cargo o/b in the time being conducted (usage of 3-5 t/day of MDO). Oil fired boiler on both vessels Generally, in the starting of the boiler when ‘heating up’, it is favorable to use MDO even if there is a possibility of using HFO (e.g. if there is an electrical heater installed comparing to systems with a steam heater - when there is no possibility of starting on HFO from the cold state). But, once heated up, the use of HFO is recommended. The consumption of the oil burned largely depends on the climate zones and seasons, and on the quantity of HFO in tanks that needs to be pre-heated. The tropical zone requires less fuel to consume and the North/South winter season requires higher quantities of fuel to be burned to heat up HFO tanks, accommodation spaces or even, occasionally, cargo spaces. Both vessels have the possibility not to use an oil fired boiler during the voyage (due to the exhaust gas economizer installed: on Vessel 1 composite boiler and on Vessel 2 separated exhaust gas economizer). But, approaching the port when there will not be enough exhaust gases for heating purposes due to the maneuvering procedure or during standstill at the port when M/E is stopped - the start of the oil fired boiler is a necessity because, even in the tropical areas, there is a constant need for heating the HFO tanks if HFO is to be used. Anyway, there is no possibility that, for example A/Es were run on preheated HFO at the port without the consumption of oil at the boiler. Because, to heat up the HFO for the A/Es there are steam heaters installed and to maintain the temperature of the HFO tanks there are steam heating coils to be used – so, steam is necessary, and for its production the part of the oil should be used for oil fired boiler when the vessels are in the port. The consumption, as mentioned above, depends on the climate zones and on the seasons, as well as on the quantity of HFO having o/b that need to be heated. 3. CONCLUSION In spite of the fact that o/b each vessel there are engines and boilers that are designed for the use of HFO almost permanently, there are several occasions when changing-over to MDO becomes absolutely recommended by the manufacturer or required to comply with environmental legislations. In such occasions (e.g.: prolonged standstill of the vessel - more than 5 days, entering/leaving the SECA areas without having LSFO o/b, running the auxiliary engines on low load during intermittent peaks and downs, etc.) the changing-over to MDO is recommended if the machineries are to be operated in good, safely and responsible manner. Otherwise, it will result in high maintenance/repair costs and in a reduced reliability of the machineries. The Pool normative in Ships presentation tables show A/E’s ‘in port’ consumption figures of HFO, and in the same time no consumption of the fuel for the boiler, simply cannot be valid as explained in the previous chapter under ‘Oil fired boiler on both vessels’ items. Such ‘special circumstances’ may raise the differences in fuel oil consumption figures, but they are to be taken in consideration when calculating the fuel being used in order to obtain the correct and acceptable figures for the owner and the Pool. FAILURE ANALYSIS OF THE DIESEL ENGINE SHIP PROPULSION SYSTEM The data, collected within the thirteen-year-long period, on the failures and maintenance activities when the propulsion system failed during the twenty-four hour navigation are taken from the engine-room log book. The significant components that usually failed during exploitation are defined by a delay analysis (of the registered failure cycles) and include the exhaust valve, the injection valve and the fuel pump. The most endangered component of the propulsion system is the exhaust valve having the highest failure rate. In order to increase the ship’s safety and her operational level, the authors suggest decreasing the failure impact on the significant components of the ship propulsion system which mostly failed during their exploitation. This can be achieved by choosing significant components of much higher quality and by using a better maintenance procedure that is by applying an adequate maintenance concept. Key words: propulsion system, diesel-engine, failure analysis, significant components, exhaust valve, maintenance. GEAR TRANSMISSION IN THE SHIP PROPULSION SYSTEM The paper provides a survey of gear transmissions in marine propulsion system gearboxes which transmit a torque and reduce the rotational speed of the main propelling engine into the rotational speed of the ship’s propeller. Gear transmissions in marine propulsion system gearboxes are divided into standard gear transmissions and multiple mesh gear transmissions. Multiple mesh gear transmissions in marine propulsion system gearboxes are divided into ordinary multiple mesh gear transmission, planetary and combined multiple mesh gear transmissions. Due to the merger or division of the mechanical energy in these transmissions, the problem of equal torque (load) distribution onto individual gear couples appears. An equal torque distribution is achieved by a flexible shaft, flexible couplings, hydraulic cylinder etc. All gear transmission models are designed on the basis of the relevant literature in the computer program Autodesk® Inventor® 2010 Professional. Key words: propulsion system, multiple mesh gear transmission, load equalization, equal torque distribution A CONTRIBUTION TO THE DEVELOPMENT OF THE MARINE STEAM TURBINE IN LOAD CONDITIONS The aim of this paper is to present the successful application of the system dynamics simulation modelling in the research carried on the performance dynamics of the marine steam turbine in load conditions, and in this very example at load of a marine synchronous generator. The marine steam turbine at load of a synchronous generator is a complex non-linear system, which needs to be systematically researched into as a unit consisting of a number of subsystems and elements, which are linked by cause-effect (CE) feedback loops (FBL), both within the propulsion system and with the relevant environment. The authors of this paper aim at presenting the efficient application of the scientific methods used in the research of the complex dynamic systems called System dynamics qualitative and quantitative simulation methodology, which will make the production and use of a larger number and a variety of simulation models of the observed elements possible, thus enabling the continuous computer simulation, which will significantly contribute to the acquisition of new information about the non-linear character of the turbine and generator system performance dynamics in the designing and education process. The marine steam turbine will be presented by a set of non-linear differential equations, after which mental and verbal structural models and flowcharts in the System dynamics symbols [1 and 2] will be worked out, and the performance dynamics in load conditions will be simulated in the POWERSIM simulation language [5]. Key words: Steam turbine, synchronous generator, simulation modelling, simulation and heuristic optimization Comparison of COGES and Diesel-Electric Ship Propulsion Systems Abstract Diesel-electric ship propulsion is a frequent shipowners choice nowadays, especially on passenger ships. Despite many diesel engines advantages, their primary disadvantage is emission of pollutants. As environmental standards become more stringent, the question of optimal alternative to diesel-electric propulsion arises. COGES (COmbined Gas turbine Electric and Steam) propulsion system is one of the proposals for alternative propulsion system, primarily due to significant reduction of pollutant emissions. On the other hand, gas turbines have higher specific fuel consumption in comparison with diesel engines what represents their noticeable disadvantage. However, some analyzes suggested that COGES propulsion system could be still cost-effective in comparison to diesel-electric propulsion, particularly on passenger ships where higher initial investment can be compensated by increasing the number of passenger cabins. This paper shows a comparison of abovementioned propulsion systems, which can be useful for the optimal ship propulsion system selection. Keywords: COGES, diesel-electric propulsion, emission of pollutants 1. Introduction For a large number of today’s passenger and cruise ships, diesel-electric propulsion was selected as optimal solution. On the other hand, other combined propulsion systems are developing rapidly and they have a goal to reduce the share of diesel-electric propulsion on these types of ships. COGES propulsion system is one of the potential candidates which must be taken into consideration during selection. Some of its features and characteristics are significantly better in comparison to diesel-electric propulsion. The main disadvantage of COGES propulsion system is operating cost and designers made some efforts of cost compensation on passenger or a cruise ships. The general advantage of combined cycles with electric transmission is that they allow the adaptation of high-speed turning turbines to the slow turning propellers without the need of a heavy reduction gear. In addition, with all the generators supplying power to a common distribution system, one prime mover can easily provide power to two or more shafts and ship service. Another advantage of combined cycles with electric transmission is that the location of the main engine is less constrained compared with designs featuring a direct transmission between the engine and the propeller [1]. Further improvement can be achieved if prime mover is more compact and it’s engine room use less space so additional passenger cabins could be accommodated. This paper presents a comparison of COGES and diesel-electric propulsion systems according to their most important operating characteristics and the exploitation costs. COGES propulsion system 2.1. COGES propulsion system history Due to concerns of exhaustion of world petroleum reserves and the will to make vessels more effective, the combined cycle propulsion has been considered as an option for naval applications in the past. The naval ships are in most cases powered by gas turbines and thermal efficiency of the propulsion plant is improved by introducing heat recovery of the exhaust and a steam cycle. COGES propulsion system for ships was firstly proposed by Mills [2] and Brady [3]. A detailed analysis of the US Navy Rankine Cycle Energy Recovery (RACER) proposal by Halkola [4] also shows COGES as an attractive propulsion system. Ahlqvist [5] compared five different combined propulsion systems for application on a 2500 passenger cruise ship. In Shipping World and Shipbuilder [6] a number of different gas turbine plants and combined cycles plants powering a fast ferry and cruise ship were compared. More recently, installation of COGES propulsion systems in Celebrity Cruise’s ships Millennium and Infinity as well as Royal Caribbean’s ship Radiance of the Seas was reported by McKesson [7] at the Seatrade Miami conference. 2.2. COGES propulsion system overview COGES propulsion system is based on electric propulsion motors and alternators driven by both gas turbines and steam turbine(s). Heat recovery steam generators are fitted in the gas turbine exhaust lines and generated superheated steam (at approximately 30 bar) is led to a steam turbo-alternator. COGES completely change the properties of conventional propulsion system, whereas gas turbine efficiency decreases at low load, the steam turbine recovers the loss of power. The result is constant fuel consumption over a wide operational range. Heat for ship’s services is taken directly from the steam turbine extraction or at the steam turbine exhaust and thus there is normally no need to fire-up auxiliary boilers. This is an important aspect in calculating total fuel consumption and overall emissions (prime movers plus boilers). Attention must be paid on the choice of steam-cycle details, such as boilers, condensers and type of cycle (condensing or back-pressure) in order to design an economically optimal plant. Cruise ships have high heat demands and thus the boiler fuel price impact is significant. In general, consumption of heat energy needs greater attention on large cruise ship installations due to ever-increasing heat requirements, mainly for fresh water production. Compact and simple machinery as well as lower fuel consumption are the main advantages, but the sensitivity to fuel price changes remains as disadvantage. Some authors give specifications of gas turbines which are usually installed in COGES propulsion plants [8], [9]. Moreover, it is difficult to obtain realistic data for built-in steam turbines, because COGES steam turbines are used exclusively to produce necessary ship electricity and steam turbines participate in total propulsion power output only through actual demand on the ship electrical grid. Therefore, it can be concluded that gas turbines are prime movers in the COGES propulsion systems while steam turbines are auxiliary machinery. Some examples of gas turbines installed in the COGES propulsion systems and their characteristics are given in Table 1: Table 1.: Working parameters of gas turbines used in the combined gas and steam turbine cycle Examples of different COGES propulsion systems layouts can be found in many papers and basic system layout is shown on Diesel-electric propulsion system 3.1. Diesel-electric propulsion system history After the rather experimental applications of battery driven electric propulsion at the end of the 19th century took place in Russia and Germany, the first generation electric propulsion was taken into use in the 1920’s as a result of the strong competence of reducing transatlantic crossing times for passenger liners [12]. At that time, the high propulsion power demand could only be achieved by turbo-electric machinery. With the introduction of high efficient and economically favorable diesel engines in the middle of the 20th century, steam turbine technology and electric propulsion more or less disappeared from merchant marine vessels until the 1980’s. The development of variable speed electric drives, first by the AC/DC rectifier (Silicon Controlled Rectifier – SCR) in the 1970’s and the AC/AC converters in the early 1980’s enabled the power plant based electric propulsion system, which is typical for the second generation of electric propulsion. A fixed voltage and frequency power plant consisting of a number of generator-sets feeding to the same network was supplying the propulsion as well as the hotel and auxiliary power. The propulsion control was done by speed control of the fixed pitch propellers (FPP). These solutions were firstly used in special vessels like survey ships and icebreakers, but also in cruise vessels. In direct driven diesel propulsion, thrust can also be controlled by a hydraulic system varying the propeller pitch angle, what is denoted as controllable pitch propellers (CPP). Podded propulsion was introduced in early 1990’s where the electric motor is installed directly on the fixed pitch propeller shaft in a submerged, rotatable pod. While this concept was originally developed to enhance the performance of icebreakers, it was early found to have additional benefits on hydrodynamic efficiency and maneuverability. 3.2. Diesel-electric propulsion system overview This is “traditional” machinery, which in case of very high power calls for large engines. The prime movers, e.g. diesel engines, supply a power to the electric generator shaft. The electric motor, which could be the main propulsion motor, is placed in submerged, rotatable pod or drives ship’s screw by line shaft [13]. In any isolated power system, the amount of generated power must be equal to the consumed power including losses. For an electric system consisting of an electric power generation plant, a distribution system, including distribution transformers and a variable speed drive, the power flow can be illustrated as in Fig. 3. If electric propulsion is chosen, diesel electric propulsion system is still the most frequent shipowner's choice for different ship types. Orders of this system at the global level are visible in Table 2 for the year 2012 [17]. Table 2 provides an order review according to world’s regions, which shows that the biggest customers of these propulsion systems come from the Far East and from Western Europe. Shipowners often choose Azipod propulsion because of higher operation flexibility, better ship manoeuvrability as well as the higher possible financial savings in ship’s exploitation. Presented systems of diesel-electric propulsion (Fig. 5 and Fig. 6) bring numerous benefits compared to diesel-mechanical propulsion systems, which can be seen primarily in the specific fuel consumption reduction Comparison of COGES and diesel-electric propulsion systems 4.1. Air pollutant emissions There are four principal air pollutants produced by prime movers which use fossil fuels: nitrogen oxides (NOx), sulfur oxides (SOx), carbon dioxide (CO2) and particulates emission [19]. The production of NOx is unavoidable because of the nitrogen presence in atmospheric air. However, the quantity of NOx is a function of the fuel nitrogen content and a strong function of the combustion temperature in engine cylinders [20]. Higher temperatures result in more NOx and combustion of natural gas produces less NOx than combustion of HFO. Peak temperatures in diesel engines are significantly higher than in gas turbines so they produce more NOx. Production of SOx is governed by the fuel sulfur content. HFO typically contains significant amount of sulfur, but there is low sulfur HFO available at higher cost. Natural gas contains only a trace of sulfur so any engine burning natural gas produces very little SOx. CO2 production depends primarily on the ratio of hydrogen to carbon in the fuel and the efficiency of the propulsion system. Liquid fuels such as HFO and MDO have a ratio of hydrogen to carbon about 2:1 while methane, the dominant constituent in natural gas has a ratio of 4:1. As a result, combustion of liquid fuels produces significantly more CO2 than combustion of natural gas. Particulates are primarily carbon particles or soot. The intermittent nature of diesel engine combustion is inclined to produce more particulates than the steady combustion in a gas turbine. Table 3 represents pollutant emissions from various propulsion systems where shown values are based on kWh of released heat. As Table 3 shows, a COGES plant is at or near the lowest value for each of the four emissions. Since COGES plants will be shown to have the highest cycle efficiency of the various alternatives, they look very attractive in terms of emissions production. Fig. 8 shows that diesel-electric propulsion has lower CO2 emissions compared to COGES propulsion system but the other emissions are 6 to 11 times higher, which only confirms the advantage of COGES system in terms of emissions. Steam turbine in COGES system increases emissions, but not enough to be significantly closer to emissions from an internal combustion engine. . Weight and size Gas turbines are lighter and need less space than a diesel engine of the same power output [19], [7]. Diesel engine’s size and weight are undisputable disadvantage in many applications. As regards the space savings of gas turbines, this potential cannot be fully utilized because gas turbines have approximately 15% larger air intake and exhaust ducts as comparable diesel engines and their starting devices also occupy significant space [21]. Relative investment costs for COGAS and diesel-electric propulsion systems Fig. 9 shows relative investment costs for various propulsion systems. COGES system would be slightly cheaper if installation costs were included. Compared with COGES, investment cost for diesel-electric propulsion system is almost half the price when engines run on HFO (Heavy Fuel Oil) or almost third the price when engines run on MDO (Marine Diesel Oil). Maintenance costs Gas turbines represent the pinnacle of engineering in the quality of used materials, in the accuracy, precision and demanding processing procedures. Hence, all of these characteristics increase COGES maintenance costs compared to diesel engines [21]. The maintenance costs of diesel engines usually fall proportionally to the engine size, but in general, the costs of diesel engines maintenance are lower in comparison to COGES propulsion system of the same power output. 4.5. Fuel consumption Diesel engines use significantly cheaper fuel compared to gas turbines. Very important fact is that specific fuel consumption of diesel engines is smaller in whole working range compared to gas turbines [10]. Efficiency of the COGES propulsion system decrease dramatically at low loads what results with significant increase in fuel consumption. Diesel engine’s tolerance to load change is considerably better. 4.7. Vibration and noise As regards vibration and noise, multiple cylinder reciprocating engines with their intermittent combustion are at a disadvantage. By direct-resilient mounting of diesel engines, their structure-borne vibration transmitted into a ship’s foundation is reduced to a level of approximately 50 dB at frequencies of 1 000 Hz. Although resiliently seated gas turbines still might reach lower values. Regardless of the type and number of silencers, diesel engine can not compete with gas turbines in this area. Air-borne engine room noise of gas turbines is claimed to be less than 85 dB, whereas the noise emission of a large-bore medium-speed diesel engine varies between 102 and 108 dB at full load [10], [21]. 4.9. Lube-oil consumption The specific lube-oil consumption of modern gas turbines is typically only 1% of the diesel engine’s figure, but high priced synthetic lubes have to be used in comparison to the low priced mineral oils for the diesel engines. The annual lube-oil costs of gas turbines are only about 6% of that for diesel engines [10], [21]. 4.10. Relative total annual cost for COGES and diesel-electric propulsion system Relative total annual costs (over 10 years) including investment, fuel and lubricating oil costs are shown on Fig. 10. It is also assumed that selective catalytic reduction (SCR) units are installed and run on diesel-electric propulsion systems. Relative total annual costs for diesel-electric propulsion systems are approximately 25 - 35% lower than the COGES system, regardless of whether diesel engines are running on HFO or MDO. Relative total annual cost for COGES and diesel-electric propulsion system Relative total annual costs (over 10 years) including investment, fuel and lubricating oil costs are shown on Fig. 10. It is also assumed that selective catalytic reduction (SCR) units are installed and run on diesel-electric propulsion systems. Relative total annual costs for diesel-electric propulsion systems are approximately 25 - 35% lower than the COGES system, regardless of whether diesel engines are running on HFO or MDO. Overall performance comparison of COGES and diesel-electric propulsion systems The overall comparison performance of diesel-electric and COGES propulsion systems are shown in Table 4. Table 4.: Overall performance comparisons of diesel-electric and COGES propulsion system COGES and diesel-electric propulsion for large cruise ship Diesel-electric propulsion system is installed in most modern large cruise ships because this system proved to be more acceptable and more economical. Passenger vessels, cruise ships and ferries have high requirements for on-board comfort regarding noise and vibration [12]. Additionally, reliability and availability is very critical for the safety of the passengers and the vessel. Consequentially, electric Case study of COGES propulsion system installed on large cruise ship Cruiser Millennium is one of the few cruise ships which do not use diesel-electric propulsion, but COGES propulsion system. COGES system of this cruiser ship consists of two General Electric gas turbines, 25 MW each, and one steam turbine with power of 8 MW. COGES propulsion system generates enough electricity to drive two RollsRoyce/Alstom Mermaid azimuth thrusters (with power of 19 MW each). COGES system is designed in a manner that the gas turbines are main propulsion devices, and the steam turbine is used to produce additional electricity. The steam for the steam turbine operation is obtained from the gas turbines exhaust gas steam generator. Depending on the amount of steam required for the electricity production, the entire COGES propulsion system has efficiency between 45 and 50%. Electricity produced with steam turbine is used for a variety of ship services: for additional drive of azimuth thrusters, for heating and cooling throughout the ship, and finally for ship lighting and ventilation. Also, the electricity generated by steam turbine fully supplies the ship’s main galley and laundry. 5.2. Exploitation costs of COGES propulsion system compared to equivalent diesel-electric propulsion system for large cruise ship Millennium Diesel-electric propulsion system could be a major competitor to COGES system when shipowner chooses a propulsion system. COGES propulsion plant installed on cruiser Millennium, with the total power output of 58 MWel is compared with dieselelectric propulsion plant which consists of five medium-speed diesel engines (and the corresponding generators and electric motors) with total power output of 61 MWel. To compare these two systems, it is necessary to take into consideration the annual fuel costs. The calculation of annual fuel cost is based on a typical weekly load scenario of 60 hours per week in ports (power requirement 10 MWel) with one medium-speed diesel engine or one gas turbine with a steam turbine in operation [21]. This scenario includes 3840 working hours per year for each of the five diesel engine or 6150 hours for each of the two gas turbines. For this scenario, the annual fuel costs have been calculated at average prices of marine fuel markets of north-western Europe for year 2000 and obtained values are given in Table 5. Conclusion This paper presents a comparison of diesel-electric and alternative COGES propulsion system. Due to lower exploitation costs, diesel-electric propulsion system remains the primary choice for the majority of shipowners. On the other hand, the exploitation costs of COGES propulsion system are much higher. It is also shown that additional passenger cabins built on ship with COGES propulsion system cannot compensate the difference in costs of these two systems. The performance characteristics of these systems show that neither has a distinct advantage over the other. The only aspect where COGES propulsion system has a significant advantage is emissions of pollutants into the environment. COGES propulsion system could become very interesting in the future if the emission regulations become more rigorous. Marine Diesel-generator Model for Voltage and Frequency Variation Analysis During Fault Scenarios Abstract Various faults in ship’s electrical power system, primarily those connected with diesel-generators, governors and automatic voltage regulators, may lead to oscillations in generator load, voltage and frequency. If those transients are large enough, a partial or total blackout may occur. In order to prevent such events, it is very important to get an insight into dynamic behaviour of electrical energy sources on board vessel. In this paper, a dynamic model in the time domain of marine diesel-generator is presented. The model is realized in the MATLAB/SIMULINK environment and consists of three main parts: the synchronous generator model, diesel-engine governor model and voltage regulator model. Simulation of a sudden loss of one generator when two generators are running in parallel is performed. Simulation results show that the presented model is fully applicable for the purpose of analysis of the marine diesel-generator dynamic behaviour during transient periods. Key words: Diesel-generator, ship’s electrical system, diesel-engine governor, voltage regulator, time domain simulation 1. Introduction Brushless self-excited synchronous generators driven by a medium speed diesel engine are the primary electrical power sources on board modern vessels. Stability of electrical power supply depends on many factors related to their proper operation but voltage and frequency are the most important ones. Stable power supply system is the one in which energy sources are returning to their original state after having been exposed to disturbances, or periodically accept a new steady state without any loss of synchronism. In other words, voltage and frequency should stabilize within the time interval specified by class society regulations. Disturbances in the system, which generally occur due to dynamic phenomena related to faults and sudden load changes, cause the oscillatory swinging of the rotor and the load due to effects of the resulting torques that accelerate or break the rotor. These transients cause changes in the generator load angle, and consequently in voltage and frequency. If the specified torques are large enough, the loss of synchronism in the generator may occur, which will most certainly lead to total or partial blackout in the system. Diesel-generator dynamic response to sudden load changes is a result of the combined interaction of all the system components. It is impossible to quantify all influential factors which determine the dynamic response of diesel-generators. The most important factors usually considered in practice are: • diesel-engine governor response, • type of generator, • type and characteristics of excitation system, • voltage regulator response, • diesel-generator torque characteristics, • diesel-generator moment of inertia. One of the worst case scenarios that may occur in ship’s power network is sudden loss of one or more generators running in parallel, which will cause instantaneous load increase on remaining generators [1][2]. It is very important, both for vessel’s electrical system designers and operators, to be familiar with the diesel-generator response during such events. Very often, it is not possible or may even be dangerous to test all possible failure scenarios on board vessel. For that reason, a dynamic model of diesel-generator which includes all the above mentioned influential factors is developed. The main purpose of the model is simulation of the diesel-generator voltage and frequency dynamic response during sudden load changes. 2. Synchronous generator model Modelling the dynamic behaviour of synchronous generator is quite complex because of the rotor winding movement relative to the three phase armature windings, so the magnetic coupling coefficient changes continuously with the rotor position. Such model of synchronous machine can be described by differential equations with timevarying values of mutual inductances, the solution to which is not always easy to find. In order to build a synchronous generator model that is independent of timevarying terms, and therefore suitable for computer simulation, it is necessary to express the rotor and stator variables in separate coordinate systems. Mathematical model of synchronous generator is based on Park’s transformation. Voltage equations of imaginary armature circuits in the direct (d) and quadrature (q) axis are obtained by the linear transformation of voltage equations in three phase abc into dq coordinates [3][4]. Diesel-generator is a rotating system consisting of the diesel motor prime mover and a synchronous generator connected via the common shaft. The equation of the motion describing the system is: Diesel-engine governor model Block diagram of the diesel engine governor model is shown in Figure 3 [8][9]. The generator actual speed ?r is compared with the reference speed ?ref . Error signal is applied to the input of controller, which is modelled as a second order system: The output of the actuator block is the prime mover mechanical torque command which is limited to 110% of the nominal power according to class society rules [lit]. The mechanical torque command is sent to the engine delay block with delay time and is multiplied by the rotor angular speed . The result is the mechanical power signal which is used as an input for the mechanical part of the synchronous generator model. Voltage regulator model The voltage regulator model is based on the IEEE recommended practice for excitation system models for power stability studies. The base of the model is the IEE AC5A brushless excitation system where the regulator is supplied from a permanent magnet generator, which is the most common arrangement for marine generators [10]. Block diagram of the voltage regulator model is shown in Fig 4. Reference value of the generator output voltage is compared with its actual value which is obtained from the output of the electrical part of the synchronous generator model. Error signal is applied to the input of the voltage regulator with the following transfer function: The exciter is modelled as a PI regulator with the gain and time constant . The output of the exciter is the generator field voltage , which is applied to the corresponding input of the synchronous generator model. The exciter saturation function represents the demands for the field voltage increase due to the effects of saturation, and can be satisfactorily defined by two values of field voltage and . The procedure for determining the saturation function can be found in [10] [11]. The model is realized in Simulink using the standard block from the SimPowersystem library. Generator parameters that are normally available in generator data sheets, and are also used in this model, are: synchronous reactance in d and q axes Xd and Xq, generator leakage reactance Xl, transient reactance Xd’, sub transient reactance in d and q axes Xd’’, and Xq’’, stator resistance Rs, transient short-circuit time constant Td’ and sub transient short-circuit time constants Td’’ and Tq’’. The equations connecting the above mentioned parameters with parameters used in Eqs. (1) to (17) can be found in [5][6][7]. The following scenarios are simulated: 1. Two generators are running in parallel, each loaded with approximately 55% of the nominal load and one generator is suddenly disconnected from network. 2. Two generators are running in parallel, each loaded with approximately 85% of the nominal load and one generator is suddenly disconnected from network. Electrical protections used in the simulation model are set as follows: • The under frequency and over frequency protection disconnects the generator circuit breaker if frequency drops under 90% or rise above 110% of the nominal frequency, with time delay of 5 seconds. • Under voltage protection disconnects the generator circuit breaker if voltage drops below 70% of its nominal value, with time delay of 2 seconds. With respect to simulated scenarios, the effect of the turbocharger lag is ignored and is assumed that diesel-engine is able to take a 55% rated load in one step. Simulation results for the first scenario are shown in Figure 5. At t=3s generator number 2 is suddenly disconnected from the network and generator number 1 is instantly taking over the entire load. The frequency drop undershoot equals 6%, and the voltage drop undershoot 9% of the nominal value. Both, the voltage and frequency are stabilized after the transient period of 4.8 seconds, while the voltage is stabilized within ± 3% of its nominal value in 1.2 seconds, which is accordance with class requirements. Voltage and frequency do not exceed the limits set by corresponding electrical protections, and they maintain their nominal values after the transient period. Generator number one is loaded with 110% of its nominal power (which is the maximum allowable load for marine generators). In this case, there is no need for fast load reduction, but no further increase of power consumption is possible until a stand-by generator is connected to the network. In Figure 6, simulation results for the second case are shown. In t=3s, generator number 2 is suddenly disconnected from the network, causing a step load increase on the remaining generator. The value of 85% of the nominal load is chosen because the specific fuel consumption in g/KWh of modern medium speed diesel engines gets close to its minimum value near this load point. The maximum voltage undershoot is 13.5% of its nominal value. Although the voltage value is all the time above the limits set by the under voltage protection, it still fails to stabilize at ± 3% of its nominal value within the period of 1.5 seconds after the start of the transient. Frequency continuously drops and at t=10.66s the under frequency protection disconnects the generator circuit breaker and a blackout occurs. It is clear that in this case a proper fast load reduction method is required to prevent blackout. Conclusion In this paper, a dynamic model of the marine diesel generator is presented. Its main purpose is simulation of a diesel-generator dynamic behaviour during various fault scenarios. The focus has been put on frequency and voltage variations. The model consists of three separate parts (synchronous generator model, diesel-engine governor model and voltage regulator model) and is realized using the Matlab/Simulink software. Its modular structure allows it to be easy adapted to various system configurations. Also, it is not limited to single generator operation and can be used for the simulation of two or more diesel-generators behaviour when running in parallel. Diesel-generator parameters used in the presented model are chosen in a way making them easily obtainable from manufacturer’s data sheets. Simulation has been performed for the worst case scenarios when two generators are running in parallel and one of the generators is suddenly disconnected from network. In the first case, the load on each generator was 55% and in the second case 85% of its nominal power. Simulation results show that the presented model is fully applicable for the analysis of different fault scenarios. Beside the analysis of voltage and frequency variations during the transient period, it can also be used for simulation and analysis of the generator protection circuits operation. MAN B&W, Wärtsilä or Mitsubishi Summary The paper presents the review of modern two stroke slow speed marine diesel engines with electronically fuel oil injection control and exhaust valve actuating control. The comparisons were made between two biggest manufacturers of marine two stroke diesel engines; MAN B&W and WÄRTSILÄ as well as UEC – ECO engine made by Mitshubishi Heavy Industries. Since all of them produce new generation of electronically control diesel engines as well as mechanically controlled diesel engines, the idea was to compare those three systems. In the paper, “Wärtsilä – Sulzer” RT-flex series engines, and “MAN B&W” ME series engines have been analyzed and compared. The comparison has been made between working principles and main working parameters. The working principle of new UEC ECO engine has been also explained and compared. The new concepts of electronically controlled marine two stroke diesel engines bring new and wide possibilities. Key words: marine propulsion, marine two stroke diesel engine, electronic control, fuel injection, exhaust valve Ship’s Propulsive Fuel Saving Analysis Summary Notwithstanding the development technology and business operation quality system, it seems hard to find the manoeuvring area for capital savings within a logistic system. This paper deals with an analysis of ship’s propulsive fuel saving possibility to be achieved by changing fuel viscosity before injection, with a view to achieving polyvalent savings and an overall quality improvement. Key words: savings, engine, viscosity, fuel THERMAL DISPLACEMENT OF CRANKSHAFT AXIS OF SLOW-SPEED MARINE ENGINE The paper presents analysis of displacement of a crankshaft axis caused by temperature of marine, slow-speed main engine. Information of thermal displacement of a power transmission system axis is significant during a shaft line alignment and a crankshaft springing analysis. Warmed-up main engine is a source of deformations of an engine body as well as a ship hull in the area of an engine room and hence axis of a crankshaft and a shaftline. Engines' producers recommend the model of parallel displacement of the crankshaft axis under the influence of an engine heat. The model gives us the value (one number!) of the crankshaft axis displacement in the hot propulsion system's condition. This model may be too simple in some cases. Presented numerical analyses are based on temperature measurements of the main engine body and the ship hull during a sea voyage. The paper presents computations of MAN B&W K98MC type engine (power: 40000 kW, revolutions: 94 rpm) mounted on 4500 TEU container ship (length: 290 m). Propulsion system is working in nominal, steady-state conditions; it is the basic assumption during the analyses. Numerical analyses were preformed with usage of Nastran software based on Finite Element Method. The FEM model of the engine body comprised over 800 thousand degree of freedom. Stiffness of the ship hull (mainly double bottom) with the foundation was modelled by a simple cuboid. Material properties of that cuboid were determined on the base of separately performed calculations. Key words: temperature deformation; marine propulsion system; shaftline alignment; crankshaft springing; slow-speed main engine Proper shaftline alignment is one of the most important procedures during marine propulsion system designing, installation and exploitation. The axis of journal bearings of shaftline should be displaced (mainly in vertical direction) to the proper position [1, 2, 3]. Usually, the crankshaft axis is a baseline for shaftline alignment. Even bearings loads and proper interaction between a shaftline and a crankshaft is the aim of this procedure. Measurements of crankshaft springing give information about the proper engine foundation as well as the proper loads coming from shaftline. During the shaftline alignment and crankshaft springing analyses, knowledge of the thermal displacement of the crankshaft axis is essential. Engines' producers proposed the model of crankshaft axis thermal displacement but it is very simple. They recommend the model of parallel displacement of a crankshaft axis under the influence of engine heat. Such a model give us one number - a value of crankshaft displacement between cold and hot propulsion system (in steady state condition), different for each type of engine. Sometimes (but not always!), the displacement value is depended on the temperature difference between not running and running propulsion system. In such a case the model might be too simple. Couplings between thermal (with influence of electromagnetic field) and mechanical loads of a marine propulsion system are very complicated and cannot be omitted during analyses of a main engine mounting [4]. Practically, a relatively simple model of the thermal-mechanical coupling is sufficient: the engine temperature is a source of the body, main bearings and crankshaft displacement; and so the displacements are a source of additional mechanical loads of propulsion system's bearings. For example, problems with shaftline alignment and crankshaft springing are befalling especially for high powered engine mounted on optimised (light - with elastic hull) ships [5, 6]. The aim of the presented analyses has been evaluation of displacements of the crankshaft axis under a steady-state thermal load [7, 8]. Up to now in the shaft line alignment typical methodology, an interaction of the crankshaft and the shaft line was considered in a simplified way. The crankshaft was modelled as a linear system of cylindrical beam elements, while its displacements due to working temperature and its foundation stiffness were evaluated based on a simple data supplied by the producer, which did not address the type of the ship on which the engine is mounted [2]. Presented research goal has been a better representation of the boundary conditions of the marine power transmission system. It is especially important for the high power propulsion systems, as in the literature there may be found numerous examples of the damage of the first three (counting from the driving end) main bearings of the main engine [9, 10]. One of the causes might be the imprecise mathematical model of a crankshaft proposed for the shaft line alignment analysis. Marine power transmission systems are modelled as isolated from ship hull and engine body [11, 12, 13, 14]. There are several reasons for this methodology. Difficulties with the proper oil film calculation (Reynolds's equation) and the need for a detailed crankshaft, engine body and ship hull FEM model are ones of the most important. That is why I apply this methodology in my research. Within the research there have been carried out a number of analyses of MAN B&W K98MC engine mounted on a container ship (~4500 TEU). The computation of the engine’s body deformation due to the gravity and its natural eigenvectors has been performed as well as the analysis of its thermal deformation in nominal work conditions. The thermal analysis requires an accurate temperature distribution on the engine’s body. Wide temperature measurements on the ship and her main engine supplied the appropriate data. The temperature measurements were performed during a ship sea voyage. The main data of the analysed ship are as follows: the total length - 292 m, the width - 32 m, the maximal draught - 13 m, the maximal caring capacity - 58000 ton, the ship's speed with 90% MCR - 24 knots. The main data of the analysed main engine are as follows: the power - 40000 kW, the nominal revolutions - 94 rpm, the crankshaft journal diameter - 1062/400 mm, the crankshaft pin diameter - 1062/531 mm. The main data of the analysed power transmission system are as follows: the intermediate shaft diameter - 735 mm, the propeller shaft diameter - 845 mm, the propeller diameter - 8.20 m, number of blades - 5, the propeller mass - 72200 kg. 2. Model verification All analyses were performed on the base of Finite Element Method. The commercial software: Patran - Nastran was used for modelling and numerical calculations. The FE model of the B&W K98MC main engine’s body has been presented in Fig. 1. Fig. 2 presents a part of the model with details of the engine’s main bearing foundation. The foundation of crankshaft in the main bearings is the most important region in the presented type of analysis. FEM model of main bearings is realised by 3-D solid elements (8-nodes), the rest part of engine body is modelled by 4-node plate elements. The whole FEM model of the engine’ body has over 812 thousands degrees of freedom and over 170 thousands elements. Engine model is 8 times (!) greater than model of the ship hull (see Fig. 3). It is the main reason for separate calculations of the engine temperature deformations and the ship hull stiffness. The detailed model of the engine body has to be analysed as separated from the ship hull. Three types of engine foundation model (boundary conditions) were analysed. First one is classical - known from literature: foundation arms are completely blocked (fixed deformation). In the second way the ship hull stiffness was modelled by beam elements. This method does not take into account couplings between supporting points of the ship hull (the ship hull is treated as a continuous beam). In the third method the foundation arms are modelled by continuous cuboid (with the cross section 0.468x0.5 m) with special material properties. On the base of the separate analyses, the stiffness of the ship hull in the engine room area (with the fundaments) was estimated and its value is equal to 1.1x109 N/m. FEM model of the container ship for those analyses is presented in Fig. 3. Area of all propulsion systems bearings' foundation was distinguished and loaded by unitary pressure [15]. Displacements of the bearings give me the value of the ship hull local stiffness. During separate calculations, the properties of the vicarious cuboid were determined in the way that the local stiffness of the cuboid was equal to the local stiffness of the ship hull with the engine foundation. The Young's modulus was determined The model of the engine body was verified by eigenvalue vectors (natural vibrations) determinations. The main target of that kind of analysis is model coherence checking. In my opinion, each FEM model (even made up for static type analysis) should be checked by natural modes analysis. It was assumed that dynamic characteristics of engine main bearings should be performed in the range of 0-30 Hz because the main force harmonic component of the engine is equal to 10.97 Hz (94 rpm, 7 cylinders) and the propeller's is equal to 7.83 Hz (5 blades). Examples for the most interesting natural modes of the engine body with the vicarious model of the ship hull are presented in Fig. 4-6. The names of the shape modes are commonly used by marine engine producers. H-mode describes transverse vibrations of the engine body (Fig. 4). X-mode describes torsional vibrations of the engine body (Fig. 5). The obtained frequencies values are confronted with my own experience in measurements and calculations and are assessed as correct (up to an order of magnitude). Values of natural frequencies for each type of boundary conditions (the modelling method of the ship hull and the engine foundation) were compared. While the boundary conditions have not very important influence on natural frequencies of the main bearings foundations, these conditions affect the global engine eigenvalues very much. The modelling method of the boundary conditions (the ship hull stiffness with the engine foundation) is essential during engine body analyses. Fixed nodes in the foundation arms area give us too stiff model but hull stiffness modelled by beams gives us too elastic model (because of not taking into account couplings between hull areas). The values of calculated natural vibration frequencies are shown in table 1. Model with cuboid foundation is the best and it is consistent with author's experience (comparison with measurements onboard of typical main engine body natural frequencies). During further calculations cuboid model will be analysed. On the base of natural vibrations analyses it may be observed that stiffness of the engine body (especially of the main bearings foundations) is high. It is much higher than primary excitation frequencies of propulsion system. Therefore, the dynamic stiffness of the engine bearings is not much different to the static stiffness. Temperature distribution of engine body Before the start of the thermal deformation analysis of the engine body it is necessary to determine the engine temperature distribution. The temperature map has been created on the basis of the measurements carried out on a marine main engine during sea trials. The temperature determination on the base of measurements is much more accurate in comparison to calculation analysis of heat transfer. Calculation has to be based on several values which are difficult for determination, e.g. the power of heat source, the temperatures of oil and cooling water, and the coefficients of heat transfer. The engine load was under stable parameters of nominal working condition. The meter used for measurement was pyrometer made by AlfaTech type Rytek MT 4. The temperature was measured in 60 points located around whole engine body. Ten points were located regularly around the cylinder heads; twenty points were located regularly on each lateral side of the engine body and ten points were located regularly on the fore and aft sides of the engine body. There is a big difference between cylinder heads and other part of engine body. The temperature distribution was estimated (with using linear interpolation between measured points) and included in the numerical model. The temperature distribution of cylinder heads and other part of the engine body is presented in Fig. 7 and Fig. 8. Of course, the temperature of the inner parts of engine body had to be estimated. The estimation was performed on the base of lubricating oil temperature measured at the inlet and outlet as well as on the base of cooling water inlet and outlet and temperature of exhaust gases. The temperature distribution of the inner part of the engine body is presented in Fig. 9. Thermal analysis of engine body deformation The coefficient of thermal expansion of the engine’s body has been assumed as ?=1.6x10- 5 . The temperature distribution has been applied to the engine’s body analogous to the one obtained from the measurement (see Fig. 7-9). A thermal deformation of the main engine’s body is presented in Fig. 10. On this and all next figures the SI standard unit is compulsory (e.g. m, Pa). From a point of view of the propulsion system and the main engine – shaft line cooperation, the most important are the displacements of the main bearings of the engine. The thermal displacements of the main bearings foundation in engine body is presented in Fig. 11. The values of the displacements are presented in table 2. The bearing numbering begins from the crankshaft free-end (opposite to shaft line). A diagram of the vertical thermal displacement of the crankshaft axis is presented in Fig. 12. The yellow circles show the places of the main bearings. According to the producer's information: "all main engine bearings are placed in hot condition higher than in cold condition". Total vertical thermal displacement of the engine bearings in hot condition, when comparing with cold condition, is equal to: he=0.37 mm - according to MAN B&W drawing No. 0793023-4, when the engine temperature is raised from cold (20?C) to normal running temperature (55?C). The numerically computed (tab. 2) average value of the translation of the crankshaft's axis (0.46 mm) is greater than the one recommended by the producer (0.37 mm). The difference is not particularly big (bellow 20%), but the displacement is of a hogging type. It seems that the producer’s assumption about the parallel translation of the crankshaft’s axis is incorrect. The hogging type deformation of the crankshaft results can have significant influence on the moment load coming from the shaft line. The effect seems to be considerable in the precise shaft line alignment analysis. Presented conclusion should be treated with caution because the thermal deformation of ship hull is not taken into account. Heat is flowing from the engine to the ship hull and locally it may be a source of other deformation of engine foundation. An analysis of the temperature deformation of ship hull in the engine room area is planned by the author. Horizontal deformations of the crankshaft axis under the heating are negligible, in spite of the temperature differences between left and right sides of the engine body. This is in accordance with the marine engines manufacturers' recommendations. They recommend that the shaft line alignment is performed only in vertical plane. Influence of crankshaft axis thermal deformation on shaft line alignment and crankshaft springing Influence of crankshaft axis thermal deformation on shaft line alignment was analysed with usage of author's specialized software [15, 16]. The software is based on finite element method. The main advantages of the software are: modeling of stern tube bearing as a continuous support and taking into account elasticity of propulsion system foundation. Several calculations were performed for the analysed propulsion system accounting for the analyses of the oil film distribution (non-linear Reynolds equation) in the sliding bearings (especially the stern tube bearing with a variable clearance) [17]. Finally, two models (results) were compared: with crankshaft axis thermal deformation recommended by the engine's producer and the other one determined on the base of the analysis presented in chapter 4. An example of the shaft line alignment analyses (the deformation and the bending moments and shear forces distribution) performed by the author are presented in Fig. 13 and 14. Two values of shaft line deformation (Fig. 13) are important. First one: absolute linear deformation (left axis) given usually in millimeters; and second one: angular deformation - shaft line rotation, given in shipyards practice in millimeters per meters (right axis). Both results have been achieved for the most complicated model with the numerically estimated stiffnesses of the propulsion system [18] and the thermal deformations. Taking into consideration detailed thermal deformation of crankshaft axis has not big influence on calculation results of global parameters of shaft line alignment. The Fig. 13 and Fig. 14 correspond to the inclusion of the thermal load deformations. The analogous curves for the case of neglecting this deformation look qualitatively similar. The quantitative differences between both cases are as follows. Maximal bending stress was changed by 2.2%. Shaft line bearings' (except the fore intermediate bearing) reactions value was changed less than 9%. But the hogging type of the thermal deformation of the crankshaft axis has big influence on the mutual loading between the crankshaft and the shaft line. The reaction of the fore intermediate bearing (closest to the crankshaft) was changed by 24% (bearing loading was increased by 64 kN). Also, loadings of first three main bearings of crankshaft were changed significantly. Main bearing No. 1 was relieved by 173 kN (71%), No. 2 was more loaded by 27 kN (10%) and No. 3 was also more loaded by 96 kN (23%). Other main bearings are not sensitive to model changes. Also, bending moment and shear force acting on crankshaft flange was changed significantly. Discussed values of the shaft line alignment are presented in table 3. Influence of crankshaft axis thermal deformation on its springing was analysed with usage of Nastran-Patran software. The crankshaft of eight-cylinder engine was calculated. The analyses were performed for two propulsion system's working conditions: cold (just after shaft line and crankshaft connection) and hot (engine body is thermally deformed after relatively long continuous working). The crankshaft foundation elasticity was taken into account as well as the bending moment and the shear force coming from shaft line. Changes of the distance between the crank arms were calculated for 8 different crankshaft positions (45? increments). The deformed crankshaft in the hot working condition was presented in Fig. 15. Cranks springing values in relation to limits data given by the engine producer - MAN B&W, for each cylinder in hot condition (with thermal deformation of crankshaft axis) are presented in Fig. 16. The influence of the shaft line on the crankshaft loading (the bending moment and the shear force), just after connection (cold condition) is inconsiderable. The springing difference before and after connection do not exceed 1% for most cranks. Only for the second and third crank it is greater but still does not exceed 3%. The thermal deformation of the crankshaft axis has big influence on two first cranks: the change of those cranks springing in relation to the limit values achieved 14%. Influence on the rest of cranks is much smaller ? the difference does not exceed 2.5%. It means that the loading values of the first three main bearings may be higher than theoretical. The first significant natural modes (eigenvalues) of engine body have natural frequencies above the range of excitation frequencies of the propulsion system. What's more the significant natural modes are quite few and those of interest are of the whole engine's body. It speaks well about the right design – the rigid structure of the engine's body. Therefore, dynamic stiffness of the engine bearings should not be much different to the static stiffness. While the boundary conditions (the modelling method of the ship hull and the engine foundation) have not very important influence on natural frequencies of main bearings foundations, the global engine eigenvalues are completely different. The modelling method of boundary conditions is essential during engine body analyses. The thermal displacement of the crankshaft computed by numerical analysis is greater than the value recommended by the producer. The difference is not particularly big (bellow 20%) but the displacement is hogging type and is a source of the additonal bending moment and shear force acting between the crankshaft and the shaft line. It seems that the producer’s assumption about the parallel translation of the crankshaft’s axis is incorrect. The effect seems to be considerable for the precise shaft line alignment analysis. Presented conclusion should be treated with caution because the thermal deformation of ship hull is not taken into account. Horizontal deformations of the crankshaft axis under the heating are negligible, in spite of the temperature differences between left and right sides of the engine body. The detailed thermal deformation of the crankshaft axis has not big influence on calculation results of the global parameters of the shaft line alignment. But the hogging type of the thermal deformation of the crankshaft axis has significant influence on the mutual loading between the crankshaft and the shaft line. Loadings of the intermediate bearing close to the crankshaft as well as loadings of three first engine's main bearings can change significantly. The thermal deformation of crankshaft axis has also big influence on the crankshaft springing for two first cranks. It also means that the loading value of the first three main bearings may be higher than theoretical. This direction of research looks very promising. It may allow improvement in installation of high power propulsion systems and avoiding failure of the engine's main bearings. The worked out methodology may be used for more advanced and complete numerical computations for multiple main engine types together with specific ship's hulls. As a further step, the propulsion system analysis methodology should be elaborated, which incorporates more complex crankshaft representation including in full its 3D characteristics. The effect of crankshaft's springing on the shaft line alignment should also be examined further. Compared to the other marine engines for ship propulsion, turbocharged two-stroke low speed diesel engines have advantages due to their high efficiency and reliability. Modern low speed ”intelligent” marine diesel engines have a flexibility in its operation due to the variable fuel injection strategy and management of the exhaust valve drive. This paper carried out verified zerodimensional numerical simulations which have been used for MLP (Multilayer Perceptron) neural network predictions of marine two-stroke low speed diesel engine steady state performances. The developed MLP neural network was used for marine engine optimized operation control. The paper presents an example of achieving lowest specific fuel consumption and for minimization of the cylinder process highest temperature for reducing NOx emission. Also, the developed neural network was used to achieve optimal exhaust gases heat flow for utilization. The obtained data maps give insight into the optimal working areas of simulated marine diesel engine, depending on the selected start of the fuel injection (SOI) and the time of the exhaust valve opening (EVO). Key words: Marine two-stroke diesel engine; MLP neural network; Numerical simulation; Utilization; Start of fuel injection; Time of exhaust valve open; Two-stroke diesel engines are the main component of ship propulsion. They are applied for propulsion of different ship types and classes due to their low price (regarding other propulsion machines), reliability, high efficiency and their very simple maintenance and servicing [1]. Turbocharging provides an increase in engine power and a modest reduction of specific fuel consumption [2]. Turbocharging causes an increase of medium effective pressure and maximum temperature of the in-cylinder process. This has an influence on the strain of engine components (as a result of differing thermal expansions) and also on the emissions of pollutants [3]. The diesel engine, as the main ship propulsion device, has to maintain very high reliability of its operation, even with the allowed degradation of performance when a failure occurs [4]. Precisely for this reason it is necessary to continuously monitor all the engine major operating parameters. Intelligent control system of the engine must have access to all diagnostic data and be able to adapt the engine to the optimal mode for desired operation [5]. In this paper, the main observed points were the engine steady states, although the numerical simulation model was not limited to steady state engine operation only. Standard engine simulations rarely include an analysis of engine transients and engine behaviour in exchanged working conditions. Simulation models based on neural networks in marine propulsion systems can achieve a number of objectives, such as the optimization of the propulsion system by changing the configuration or customizing the engine control settings [6], [7]. Two-stroke low speed marine diesel engine 6S50MC MAN B&W, whose data were used for numerical simulations, Table 1, is originally not designed for variable settings in fuel injection and exhaust valve opening. This can be done with the same manufacturer modified engine design, which has a new designation MCE for "intelligent" engine variant (electronically controlled electro-hydraulic drives for exhaust valves and fuel injection). Manufacturer set the basic angle settings for the start of fuel injection and the opening of the exhaust valve, so different settings of these angles may worsen or improve the engine operating parameters. Engine available data from test bed The main data of the marine diesel engine are obtained by measurement [8]. Such measurements are performed during the testing of the new engine on the test bed. Table 2 presents the measured values for the selected engine steady operation points The examination was performed at the following environment state: ? Ambient temperature 30 °C, ? Ambient pressure 1005 mbar, ? Relative humidity 50%. The engine was tested on diesel fuel D-2, whose features are, according to a supplier report: ? Density 844.7 kg/m3 , ? Kinematic viscosity 3.03 mm2 /s, ? Sulfur content 0.45%, ? Net caloric value 42.625 MJ/kg. Modern marine diesel engines with electro-hydraulic control of fuel injection and exhaust valve opening allow a very large area of engine customization in various modes. This entire area is usually too large for complete engine testing, and detailed measurements are not publicly available Effective power shows the same trend, regardless of whether it is at full load or at 50% engine load, Figure 3 and Figure 4. SOI has a decisive influence on the engine effective power. The maximum effective power was reached with the start of fuel injection just before the engine factory settings. Moving the SOI for later reduces the effective power, and later injection causes the proportional decrease in effective power. EVO has an almost constant effect on the engine effective power for the selected SOI. The specific fuel consumption has been the lowest for the SOI just before the factory settings, and the same is optimal for engine operation regarding the specific fuel consumption and effective power. Also in this situation, for specific fuel consumption, EVO has an almost The exhaust gases thermal flow at the turbine outlet leads to similar conclusions as for the exhaust gases temperature after the turbine, regardless of the engine load, Figure 7 and Figure 8. Although it is not explicitly visible in Figure 8, this picture also points to the engine operation instability at a very early SOI and late EVO at 50% load, which is reflected in the changes of the engine operation area borders. At full engine load, Figure 9 shows that the maximum pressure in the cylinder is obtained for very early SOI shift and very late EVO shift. This working area at this engine load is stable and there is no danger of falling out of operation. Also, in this same area the effective power decreases and the specific fuel consumption increases, and surely this working area is not preferred for selection. The maximum cylinder pressure rapidly decreases for later SOI shift, and earlier EVO shift regarding to the engine referent values. Maximum cylinder pressure at 50% load shows the same trend as for a full engine load, Figure 10. The only exception is the area of early SOI shift and late EVO shift, which also at this engine load shows unstable engine operation due to worsened scavenging process. The highest temperature in the engine cylinder, for both of the observed loads was achieved at a very early SOI shift and proportionally early EVO shift, Figure 11 and Figure 12. However, it should be noted that too high temperatures in the engine cylinder cause high emissions, primarily emissions of nitrogen oxides, and under these conditions the engine certainly could not provide the required environmental standards. Regarding the maximum temperature of the engine process, usually a compromise between the achieved emissions and the produced heat necessary for utilization has to be found. The optimization of SOI and EVO for maximum thermal flow of exhaust gases for utilization The simulation results indicated that the SOI shift for 3.5 °CA later, and EVO shift for 20 °CA later, ensure the highest exhaust gas thermal flow, 9.5% higher than the reference one, Figure 13, with an increase in specific fuel consumption, Figure 14. The increase in specific fuel consumption is small enough, that increased fuel cost will be very quickly paid off by using higher obtained exhaust thermal flow for the utilization process. At the same time the engine power (ie. engine torque at the constant engine speed) decreased by approximately 7%, Figure 15. Presented simulation results show justifiability for SOI and EVO shifts, in order to obtain a sufficient additional thermal flow, which can be effectively applied in the utilization process. In that way, it is possible to achieve significant savings in the ship propulsion plant with such a diesel engine and with the possibility of achieving multi-criteria optimization. In this mode of engine operation, excessive pressures and temperatures in the engine cylinder can be avoided, Figure 16 and Figure 17. This fact proves that the displayed change of operating parameters would not lead to significant thermal load increase, or to high increase in emissions. Such operating parameters change during the actual engine operation surely will result in a substantial impact on the entire propulsion plant efficiency. Satisfying the required thermal capacity at constant torque and engine speed In this case, the engine operating point was given with engine speed nM = 118.5 min-1 and the required torque MM = 465 kNm. The minimal required exhaust thermal flow after turbine Qmin ? = 3200 kW was also given. The limits on the position of the fuel rack were between 40 and 75 mm. Additionally, the limits for the highest maximum temperature of the engine process Tmax,lim = 1900 °C and the maximal allowable pressure in the cylinder pmax,lim = 14 MPa were also set. The pressure values from the simulation can exceed the limited value of 14 MPa but those working points are constrained in optimisation algorithm because they are not used in real engine operation. The initial idea was to find the engine operation point at which all given conditions are met. For referent settings of SOI and EVO the given value of exhaust thermal flow was not achieved. In order to achieve the operating point where the parameters are equal or the nearest possible to default ones, SOI and EVO shifts were allowed. In order to achieve and maintain the engine torque, since it varies by SOI and EVO shifts, it was necessary to make a fuel rack position correction. After this step, the engine has reached a working point where the desired thermal flow was satisfied. In that working point, the desired engine torque was also reached. Simulation passes the entire field of SOI and EVO shifts, and for each of the shifts a new fuel rack position was calculated. With the new position of the fuel rack, the predefined engine torque and engine speed (nM = 118.5 min-1 , MM = 465 kNm) must be satisfied. Then it is checked if the new operation point satisfies a predetermined minimum exhaust thermal flow on the turbine outlet. Finally, the simulation checks if the newly determined operation point has a lower specific fuel consumption than the previous one. If at least one point satisfies all these conditions, the system has a solution. At given operating conditions, engine torque shows almost constant value, but stable change for almost the entire working area, for all SOI and EVO shifts, Figure 18. The only exceptions are the areas of large SOI shift later than the reference value, with intense loss of engine torque, and thus engine power, due to the worsened scavenging conditions. A major SOI shift to later shows late fuel injection, so in this area incomplete combustion can be expected, which results in a huge loss of engine torque. It is necessary to avoid the area where these phenomena occur, because it is impossible to achieve a stable operating point. Even in this operation mode, specific fuel consumption was the lowest at the reference (factory) engine settings, Figure 19. Large increases in specific fuel consumption occurred only at intense shift of SOI for later, where a huge exhaust thermal flow at the turbine outlet was available, Figure 20, but it is necessary to avoid this operating area due to the large reduction in engine torque and highly probable fall-out from the drive, Figure 18. Maximum exhaust thermal flow, in the engine stable operation area, is presented in Table 7. Maximum cylinder pressure occurs in areas of very early SOI shift, and very late EVO shift, Figure 21. Therefore, for the maximum cylinder pressure it is optimal to hold SOI and EVO parameters to reference values, with the recommended EVO shift to earlier, in order to avoid excessively high pressures. In this area, other operating parameters do not indicate a sudden or unexpected change, so this engine operating area would be advisable for given conditions. EVO shift to earlier, while retaining the referent SOI, would be recommended also for maximum engine process temperature, Figure 22, because the maximal temperature would be optimal for utilization, and thermal load of engine working parts or emissions remain acceptable. The optimal solution could be achieved also with more complex methods of optimization (multi-criteria optimization, multi-objective optimization, etc.). In this paper, the changes in the characteristics of "intelligent" marine two-stroke diesel engine were studied, when crank angles for the start of fuel injection (SOI) and for the opening of the exhaust valve (EVO) were shifted. The fuel injection strategy (fuel injection flow) and the exhaust valve opening curve did not change, which was left for future research. The investigations have pointed to the great potential that provides electro-hydraulic control of fuel injection and exhaust valve drive to bring the modern marine diesel engine in the desired working conditions. Some aexamples of the described neural network applications in optimization of marine diesel engine were presented, in order to achieve the desired exhaust heat flow for the utilization purposes, along with minimum specific fuel consumption, as well as to maintain maximum engine process temperature as low as possible in order to reduce NOx emission. The developed neural network model is fully prepared for the reception of new data, measured during the engine operation. With comparisons of measured data and data obtained by the neural network, it will be possible to evaluate the quality of measured data and the entire measuring system. This was already proven on various sets of measured data. The neural network model was developed using data obtained from numerical simulations, for the engine steady state operation, with verification from available data measured on the test bed. Therefore, the existence of high-quality numerical simulation model in neural network development was very important. Also, the resulting neural network model has limitations (eg. the model is valid for engine steady state operation, for the same type of engine, the same selected turbocharger etc., but the obtained structure can effectively learn on the data for a new engine type). LEAST INERTIA APPROACH TO LOW-SPEED MARINE DIESEL PROPULSION SHAFTING OPTIMUM DESIGN In this study, a novel approach to the low-speed marine diesel propulsion shafting design is proposed and examined. The proposed approach is based on the shafting least inertia principle, in which the design task is formulated and solved as a constrained nonlinear optimization problem. The core of the approach is a cost function, which is defined as a weighted sum of the shafting, turning wheel, and tuning wheel inertias, because it is a suitable proxy of the propulsion shafting material and production costs. The constraint set is composed of the three mandatory constraints, where the crankshaft, intermediate shaft, and propeller shaft torsional vibration stresses should be lower than the corresponding stress limits, as well as a few additional constraints that help ensure that the plant behavior complies with applicable regulatory and operational requirements. For optimization purposes, a Recursive Quadratic Programming method is utilized, while the shafting torsional vibration response is determined using a standard vibration analysis program with slight modifications. Numerical experiments have shown that fast convergence can be achieved. Compared to the classically obtained solution, the proposed approach provided more than 8 % reduction in cost function as well as significantly reduced design time. The designs of modern merchant ships tend to maximize the cargo space, which thus reduces space in other parts of the vessel. When looking for space to minimize, the engine room comprising the main engine, auxiliary engines, boilers, and other utilities is the prime candidate. However, there are clear limitations to reducing the engine room space. Firstly, there are physical limitations in terms of placing all equipment into limited smaller sized engine room. Secondly, a limitation is imposed by the minimum length of the propulsion shafting that is needed to retain its torsional vibration stress levels within acceptable limits. In general, the main engine location is usually selected as the aft-most position allowed by the ships aft body geometry; see, for example [1]. This means the least propeller shaft and intermediate shaft lengths are involved, Fig. 1. The unfavorable consequence of this approach is the high torsional vibration stress generated on the propulsion shafting [2]. Since modern engines are characterized by maximized combustion pressures and therefore, increased vibration excitations, the correct design of the propulsion shafting design emerges as one of the most challenging tasks during the process of designing the ship's machinery. The correct design of the marine diesel propulsion shafting was reviewed in the classic work [3], while other important aspects have been covered in greater detailed elsewhere [4-8]. Optimization methods in the design of marine diesel propulsion shafting have mostly been applied in the field of shaft statics [9-11], whereas dynamic responses have rarely been taken into account. All these papers have dealt with the number and optimal positions of the journal bearings, implying the already known shafting, turning wheel, and tuning wheel dimensions. However, no one paper has researched the selection of these dimensions. The optimal selection of shafting dimensions has been the theme of a small poster [12], but no practical realization was provided at that time. The aim of the present study was to propose a procedure for the selection of the lowspeed marine diesel propulsion shafting dimensions, in which the design task is formulated and solved as a constrained nonlinear optimization problem. The proposed procedure takes into account the torsional vibration response of the propulsion plant (because this is the most influential factor that determines the shafting dimensions), an approach which has not hitherto been reported in literature. The applicability of the proposed procedure was examined using a case study of the design of a Suezmax tanker propulsion shafting, which was originally optimized using a classic trial-and-error approach, [2]. The remainder of this paper is organized as follows. In Section 2 the basics of the least inertia design approach are provided. Then, in Section 3, the torsional vibration response of the propulsion shafting is outlined. In Section 4, the specific characteristics of the numerical experiments are reported in more detail. Finally, in Section 5, some conclusions arising from the study are discussed. Least inertia approach The marine diesel propulsion shafting of a typical merchant ship usually contains the following components: engine crankshaft, intermediate shaft, and propeller shaft, Fig. 1. Regarding the dynamic behavior of the shafting, important additional components include the turning wheel (connected to the crankshaft aft side), tuning wheel (connected to the crankshaft fore side), and the propeller. In cases when a favorable torsional vibration response is able to be obtained, the tuning wheel can be removed from the system. For analysis purposes, the propulsion shafting is usually simplified to a torsional scheme, as shown in Fig. 2 Shafting inertia and stiffness are the most influential factors that determine the overall torsional vibration behavior of the propulsion system. The constant diameter shaft element inertia, J, and stiffness, k, are defined with: where G and ? are the shaft material shear modulus and density, and l and d are the shaft element length and diameter, respectively. According to Eq. (1) and (2), the lower shaft inertia implies a smaller diameter and lower stiffness. The opposite is also valid. In general, the smaller shafting inertia systems possess a number of advantages: smaller shaft diameters imply lower material costs in terms of smaller bearings, stern tubes and oil glands, smaller propeller hubs, simplified shaft alignments [13], and reduced propellerinduced variable thrusts [14], that provoke unfavorable engine room and overall ship hull vibrations. However, smaller diameters also imply higher vibration stresses, and thus vibration analyses should be carried out with the utmost care and accuracy. In addition, it should be clearly realized that the inertia of the turning wheel and tuning wheel has a completely opposite influence on the propulsion plant torsional vibration behavior. Specifically, the smaller inertia wheels generally increase the vibration torque, and hence enlarge the vibration stress. Therefore, it is essential to consider the inertia of the whole system, where the shaft inertia and wheel inertias are the constituents. Applied in this way, the least inertia approach ensures a balance between the opposite shaft diameter and wheel size influences. The rationale behind the least inertia design approach is to select those shafting dimensions, and thus the accompanied components, that minimize the overall inertia of the propulsion system and, at the same time, satisfy all the imposed constraints (e.g. vibration stresses should be within the allowable limits). More formally, the least inertia design approach for the low-speed propulsion shafting system can be defined as follows. Figure 1 shows that the intermediate shaft and propeller shaft are actually stepped shafts composed of various sections of distinct diameters. However, it can be shown that these stepped shaft diameters are not independent, because they are defined by particular relations [6]. Therefore, with no substantial loss in accuracy, both shafts can be satisfactorily defined by using two unique diameters only, denoted here as IS d and PS d . More generally, the set of design variables can be expanded by introducing both shaft lengths, the required material properties, and other parameters that define the propulsion plant in greater detail. 2.2 Cost function The definition of the cost function is a cornerstone of the least inertia propulsion shafting design approach. The cost function is defined as the sum of the inertias of the shafting components. Because some components are predefined, such as the engine crankshaft and propeller, they can be omitted from this definition. In addition, wheel inertias are usually an order of magnitude greater compared with the shafting inertias. Therefore, some kind of inertia scaling is desirable. Taking these considerations into account, the cost function can be finally set as: where S J is the shafting inertia, and ww w S FW TW , , and are the weight factors assigned to the shafting, turning wheel, and tuning wheel inertias, respectively. It is reasonable to set the weight factor of the shafting inertia equal to unity (wS = 1), and assign smaller values to the other weight factors (values in the range of 0,01 to 0,1 seem to be appropriate). The cost function then implies the corresponding or equivalent shafting inertia. Because the shafting production costs are usually strongly correlated with the shaft dimensions and the sizes of the accompanied equipment, the proposed cost function becomes a suitable proxy of the actual shafting costs. It is important to realize the true meaning of the selected weight factors. They are not simply pure numbers that scale the contributions of various inertias to the cost function. On the contrary, they are an expression of design intent, namely a designer's will, expressed in a few simple numbers. Lower weight factors minimize the relative importance of the respective property and thereby allow for higher values of it. By contrast, higher weight factors stress the relative importance of the respective property and thus support its reduction. Further insights into the benefits of these weight factors can be obtained by analyzing the results provided in Section 4. 2.3 Constraints Constraints are set of conditions that, when met, ensure that design feasibility complies with the pre-agreed classification rules. Furthermore, they help ensure the plant behavior complies with operational requirements. The three mandatory constraints are: where j ? is the peak vibration stress encountered in the whole engine speed range, 2, j ? is the corresponding stress limit, j s is the stress limit factor, and j is the index denoting the crankshaft, intermediate shaft, and propeller shaft, respectively. The vibration stress limits, Equation (5), applicable to the crankshaft are defined in [5], or in the enginebuilders documentation, while the corresponding stress limits for the intermediate shaft and propeller shaft are defined in [6]. If more specific requirements are called for, the set of constraints can be easily expanded through additional constraints. Typical candidates are the peak vibration stress limit in the event of irregular firings in one of the engine cylinders, the angular displacement or angular velocity limit of the engine crankshaft, and the position of the so-called barred speed range (Section 4.3). In addition to these explicit constraints, Eq. (5), a number of implicit constraints can be expressed in the form of variable bounds. For instance, the diameters of both the propeller shaft and the intermediate shaft should be greater or equal to the minimum shaft diameters prescribed by classification bodies, [6]. Furthermore, the dimensions of the turning wheel or tuning wheel should also be within the bounds stipulated or approved by the enginebuilders. 3. Shafting torsional vibration response The analysis of the torsional vibration response of the propulsion shafting is nowadays a well-developed field that has multiple sources that thoroughly treat its computational aspects [15-18]. The main equation governing the system response is: where J is the inertia matrix, C is the damping matrix, K is the torsional stiffness matrix, ? is the displacement vector, and f is the vibration excitation vector. Equation (6) is a non-homogenous system of linear ordinary differential equations of second order with constant coefficients, which after applying the proper substitution [17], transform into a system of algebraic equations with complex coefficients: where ? is the excitation frequency, i is the imaginary unit, and ?^ and ^ f are the vectors of the complex angular displacement and excitation amplitudes, respectively. The number of equations in Eq. (7) corresponds to the number of lumped masses in the torsional vibration scheme, Fig. 2. Vibration analysis starts by calculating the natural vibration (eigenvalue problem), [17], where the plant's natural frequencies, mode shapes, and critical speeds are determined. Then, a forced vibration response is calculated, including working out the angular displacements of all system masses. After that, the remaining process is straightforward: the vibration torques are calculated using the shaft stiffness properties and finally the vibration stresses are determined. This analysis process refers to one harmonic of the vibration excitation. Because the actual vibration excitation is of the periodic form, a series of responses should be calculated and synthesized until the total response is obtained. In addition, the whole process should be repeated multiple times for various engine speeds within the operating speed range. A Suezmax tanker case study The proposed approach was used during the development of a ShaftOpt v1.0 design tool that provides the optimum dimensions of a propulsion shafting in terms of its torsional vibration behavior. The case study of a 166300 dwt Suezmax oil tanker is used to demonstrate the utility of this program. The basic design data are provided in Tables 1 and 2. Other engine-specific data can be obtained from the enginebuilders documentation. 4.1 Optimization problem The studied tanker [19] is characterized by a relatively small engine room (22,95m in length), with the engine located in the foremost position given the limited engine room space available, Fig. 1. The design of the propulsion shafting can be defined as an optimization problem where the vector of the design variables, Eq. (3), should be determined to minimize the cost function, Eq. (4), such that a set of constraint functions, Eq. (5), is satisfied. 4.2 Optimization method For optimization purposes, we use a variant of the Recursive Quadratic Programing method (RQP; also known as Sequential Quadratic Programing, SQP), [20]. This method was originally developed by Pshenichny [21], and further improved by Lim and Arora [22], and Belegundu and Arora [23]. The method possesses a number of favorable properties, including a sound theoretical foundation, proof of global convergence with a potential constraint strategy, and a good track record of reliability and efficiency [24]. Optimization runs In order to better understand the influence of the weight factors related to the cost function, Eq. (4), it was decided to perform a series of optimization runs, each with different weights prescribed and/or different starting points. Table 3 comprises 12 series of the weight factors that were applied to the cost function. In all trials, the shafting weight factor was set to unity, while the weight factors for the tuning wheel and the turning wheel were varied. Table 3 has four weight factor blocks, each containing three options. The first one gives equal weight to both wheels, whereas the second and third options give preference to reducing the tuning wheel or the turning wheel, respectively. The relative difference between the two weights and the basic one is set to ±20%. Since the optimization problem is not unimodal (at least two local minima are expected), for each combination of weight factors three starting points were adopted (Table 4), thereby raising the total number of optimization runs to 36. The Mini starting point (Table 4) resembles the minimum dimensions stipulated by the classification bodies [6], or allowed by the enginebuilders. This starting point is deeply infeasible, as the maximum constraint violation of 0,485 indicates, Fig.3. In addition to the stress response of the shafting vibration, this figure includes two curves that represent the stress limits imposed. The lower one (denoted by 1 ? ) represents the stress limit for the continuous running of the engine, whereas the upper one (denoted by 2 ? ) represents the stress limit for the transient running of the engine. The 1 ? stress limit can be violated for a limited time only, whereas 2 ? should not be violated under any circumstances. The engine speed range where the actual vibration stress exceeds 1 ? stress limit is referenced as the barred speed range. The Maxi starting point (Table 4), by contrast, includes the maximum turning wheel and tuning wheel allowed by the enginebuilders. The maximum starting diameters of the propeller shaft and intermediate shaft are selected as 20% higher values compared with the corresponding minimum one. This starting point is feasible, as indicated by V ( ) x =0, Table 4. Finally, the Midi starting point (Table 4) is assembled of values that are in the middle of the previous two. The corresponding maximum constraint violation is equal to 0,106, indicating the moderate infeasibility. During the evaluation of the constraints, Equation (5), the following stress limit factors are used: 1s = 0,95 and 2s = 3s = 0,92. These values provide a safety margin for cases when the peak vibration stresses are additionally elevated owing to irregular firings in some of the engine cylinders. In order to calculate the torsional vibration response, a TorViC [26] computer code is used, although the basic code is slightly modified to enable analysis calls by the optimization package. At the source-code level, the analysis program is transformed into the subroutine and the subroutine arguments are used to communicate between the optimization and analysis programs. All analysis program outputs are suppressed. The ordinary analysis includes a vast number of evaluations of the forced vibration response within the whole engine speed range. For optimization purposes only, a subset of these response evaluations is carried out, depending on the resonances determined in the natural vibration analysis phase. The proposed procedure is based on the shafting least inertia principle, when shafting dimensions are used that minimize the inertia of the propulsion plant while fulfilling the torsional vibration stress constraints. The benefits of the proposed approach are threefold. First, it can efficiently determine feasible designs that fully comply to classification society rules and other design requirements. Second, it provides design solutions that have a series of favorable features such as lower material and production costs, a simplified shaft alignment, and reduced propeller-induced variable trust remedies. Finally, by varying the cost function weight factors, it can generate different optimal solutions depending on the designer's preferences. Numerical experiments have shown that fast convergence of the resulting nonlinear optimization task can be achieved. They have also shown that the proposed procedure provides cost-effective designs and reduces design time. Infl uence of Low-Speed Marine Diesel Engine Settings on Waste Heat Availability The low-speed marine diesel engine is the most effective of all the ship propulsion systems. On every ship there is a need for thermal energy besides mechanical power to drive the propeller. It is possible to install a heat exchanger in the exhaust system that makes use of waste heat of the exhaust gasses of the diesel engine. Such a combined mechanical and thermal energy generation is called cogeneration. Modern engines allow the variation of the fuel injection timing and the variation of the exhaust valve timing, which results in a great usage fl exibility. In the current work a computer simulation model of a low-speed marine diesel engine is presented. The exhaust gas heat energy available to power a heat exchanger was calculated. The time of the beginning of fuel injection and the time of the opening of the exhaust valve was varied. It was analyzed how these parameters infl uence the power, the fuel consumption, the engine ef- fi ciency, the exhaust gas temperature, the heat energy available in the exhaust gasses, the overall effi ciency of the cogeneration system, and the power to heat ratio. Keywords: cogeneration, computer simulation model, low-speed marine diesel engine Ship transport is the most effi cient mode of transport, especially if the fuel consumption per cargo mass is considered. Of all ship propulsion systems, the most effi cient is the one with the low-speed marine diesel engine. There were periods when steam turbines were the fi rst choice for large ships, but only because such powerful diesel engines were not available [1]. Modern marine engines achieve the fuel conversion effi ciency of around 0.5, which means the lowest specifi c fuel consumption. Besides this, which is very important today, it means the lowest specifi c emission of greenhouse gasses. Advanced fuel injection systems have been developed, and they allow the variation of the start and the duration of fuel injection. On certain engine models, the hydraulic exhaust valve drive allows variable valve opening [1]. Besides mechanical energy for ship propulsion, on ships there is the need for heat energy. The amount of such energy depends on the ship type. Some examples of heat applications are accommodation heating, steam generation, or cargo heating. A great amount of heat is needed on oil tankers. In fact, crude oil at lower temperatures changes from liquid to a state similar to fat, so it is necessary to heat it up during transport. Before the arrival to the destination port, the crude oil temperature has to be raised to allow pumping of the fl uid. Furthermore, another possible application of waste heat is to use it as the power for an absorption refrigeration system [2] or air conditioning device. A tendency towards pollutant emissions and fuel consumption reduction can be recognized in the world. One of the many strategies to achieve this goal is cogeneration or CHP (Combined Heat and Power) [3]-[7]. The comparison between the effi ciencies of a separate production of heat and power and of a cogeneration system is shown in Figure 1. Cogeneration can be performed with a steam turbine, with a gas turbine or with a piston internal combustion engine. A low-speed marine diesel engine uses a big part of the chemical energy stored in the fuel and converts about a half of it to mechanical work, and dissipates the rest in the form of waste heat. Moreover, it operates with a low equivalence ratio (regarding fuel/air ratio), which results in lower exhaust gas temperatures in comparison with other engine types. These facts make it a less suitable choice for cogeneration when compared with other systems. However, if it is considered that its main goal is ship propulsion and that the value of the mechanical energy is higher than that of the heat energy, that it uses cheap heavy fuel oil and that it is often already installed on existing vessels of certain type, the cogeneration with the diesel engine is an option that has to be considered The marine engine settings are usually such that they allow the highest effi ciency, the fulfi lment of pollutant emission limits and ensuring of safe operation. However, modern engines with variable fuel injection systems and variable exhaust valve opening systems allow the change of these settings according to necessities. Hence, it is possible to optimize the operation for different goals: lower specifi c fuel consumption, higher maximum power or lower pollutant emission. If there is a waste heat utilization device installed onboard, there is another possible optimization goal: a larger quantity or a higher temperature of exhaust gasses, or in other words a higher amount of waste heat. Among low-speed marine diesel engine manufacturers, MAN B&W offers the TES (Thermo Effi ciency System) system, which allows the raising of the exhaust gas temperature so their heat could be used for steam production. One of the tools that allow development and analysis of the processes in an engine is numerical simulation [8]-[12]. The mathematical model of engine components is based on conservation laws and on laws of thermodynamics and it consists of differential equations. Modern computers allow fast solution of such systems and a thermodynamic analysis of the infl uence of parameters on engine performances. The engine mathematical model is based on the fi rst law of thermodynamics and on energy and mass conservation laws. The engine processes are described by nonlinear ordinary differential equations. The slow-speed marine engine system can be split in various elements, Figure 2, which can be analyzed separately or all together as a system. The main system elements are: engine cylinders, exhaust receiver, scavenging receiver, air cooler, turbine, compressor, engine governor and ship propeller. Besides these, a waste heat utilization heat exchanger has been added to the exhaust line. The cylinder is confi ned by the liner walls, cylinder head and by the piston which runs between the top dead centre to the bottom dead centre positions. By applying the fi rst law of thermodynamics on the variable mass and composition fl uid in the cylinder, the differential of heat energy is: By inserting equation (2) in equation (1) and deriving by crankshaft angle it follows: (3) which is the basic equation for the heat energy increase in the engine cylinder. The heat energy exchange is defi ned by the energy released from fuel combustion, by the heat exchanged between the fl uid and the cylinder walls, and by the heat that fl ows to the environment: The gas mass exchange in the engine cylinder is defi ned by the mass of the fl uid that fl ows through the scavenging ports and trough the exhaust valve, by the blow-by process, and by the injected fuel mass. The heat energy fl uxes in the two-stroke, unifl ow scavenged diesel engine cylinder is shown in Figure 3 where: Additional details about the engine mathematical model are given in [8] and [9]. The waste heat fl ux contained in the exhaust gases that can be utilized is calculated according to: where mg . is the mass fl ux of the exhaust gasses, cp is the specifi c heat capacity of the exhaust gas, Tout is the supposed outlet temperature of the exhaust gases. The inlet temperature of the gases in the heat exchanger Tin is the outlet temperature of the gases that leave the turbine and is calculated according to: where T4 is the temperature of the exhaust gases after the turbine, T3 is the temperature of the exhaust gases before the turbine, ?T is the turbine effi ciency, ?T is the turbine pressure ratio, and ? is the adiabatic exponent for the conditions in the turbine. The mathematical model was built in MATLAB-SIMULINK environment. The model was applied on a low-speed, two stroke marine diesel engine MAN B&W 6S50MC. The main technical characteristics of the engine are presented in Table 1. A crosscut of the engine is shown in Figure 4. The validation of the model and the comparison with experimental results from the test bench are presented in [8] and [9]. The engine allows the variation of the fuel injection timing. It is assumed that there is a possibility of variation of the start of the exhaust valve opening, which is an option on some engines of similar characteristics. The possibilities of fuel injection timing variation and exhaust valve opening variation are integrated in the model. The parametric analysis was done for the start of fuel injection in the range from ? = -6o to +6o compared to the original value (? = 177o after BDC – Bottom Dead Centre), that is from ? = 171o to 183° after BDC. As far as the beginning of the exhaust valve opening is concerned, simulations were done for cases from ? = -16o to +8° compared to the standard value (? = 290o after BDC), that is from ? = 274° to 298° after BDC. The governor regulates the amount of fuel per process in function of the engine speed and of the position of the throttle handle. The engine load was 100% of the maximum continuous power according to the results obtained on the test bench. It means that the governor leads the engine in the way to obtain 8182 kW at the revolution speed of 121.4 min-1. It was achieved for all the cases except for those with the latest injection timing, with the beginning of the injection at 183 °CA The start of fuel injection (SOI) was varied from ? = 171° to 183°. The moment of the beginning of the exhaust valve lift (EVO) was also varied from ? = 274° to 298°. The duration of fuel injection and the duration of valve lift remained unchanged. It was observed how the change of these parameters infl uences the fuel amount per process (Figure 5), the engine effi ciency (Figure 6), exhaust gas temperature (Figure 7), the heat available in the exhaust gasses (Figure 8), the overall effi ciency of the cogeneration system (Figure 9), and the power to heat ratio (Fig. 10). The governor tries to achieve the selected power by changing the amount of fuel per process. Since the load is determined by the propeller curve, each power level is related to a determined revolution speed. In Figure 5 it can be seen that the fuel mass per process increases for later injection and decreases for earlier. This is an expected behaviour, since for an earlier fuel injection higher cylinder pressures are reached for the same fuel amount, so for an earlier injection a smaller fuel quantity is enough to reach the same power. In the same fi gure it can be noticed that the beginning of exhaust valve lift has a weaker infl uence: an earlier opening and closing of the valve results in the need for a greater fuel mass to produce the same power on the crankshaft. The engine effi ciency is given by the following equation: (13) where Pb is the brake power while Q f . is the heat fl ux released by the combustion of the injected fuel. In Figure 6 the basic trend can be seen and this is that engine effi ciency rises for an earlier fuel injection and for a later exhaust valve opening. Figure 7 shows the exhaust gas temperature. It changes similarly as the fuel mass per process. It rises for later fuel injection and earlier exhaust valve opening. It can be noticed that the temperature decreases for the latest fuel injection checked, the one that occurs 6oCA after the standard setting. For these settings the selected power could not be reached, the regulator cuts off the required fuel mass, and as a consequence the power and exhaust gas temperature drop. The exhaust gas temperature infl uences strongly the heat contained in the exhaust gases, as it can be seen in Figure 8. The highest value of 5.317 MW is obtained for the valve opening at 274oCA and for the beginning of fuel injection at 181oCA. If the heat is used to power an absorption refrigeration system, with an assumed coeffi cient of performance of 0.7, it means that in the optimal case from the waste heat it is possible to obtain the cooling capacity of: The overall effi ciency of a cogeneration system takes into account that the heat from the exhaust gases is being used. It is calculated according the equation In Figure 9 it can be noticed that the highest overall effi ciency is found for earlier injection timing and for earlier exhaust valve opening. The ratio between the mechanical power and the useful heat in the exhaust gases or PHR (Power to Heat Ratio) is calculated according to the equation Usually, it is desirable to have the PHR as high as possible since it is more diffi cult to obtain the mechanical power, and hence it has a higher value. However, it is possible to imagine the situation in an operational cogeneration system in which a greater demand for heat arises. For example, a tanker before arriving to the destination port has to raise the temperature of the carried crude oil in order to pump it. In such a situation it is possible to change the engine settings in order to obtain a higher quantity of heat. In such a case a lower PHR could be welcome. In Figure 10 it can be seen that the PHR rises for earlier fuel injection and for later exhaust valve opening. It means that for the above described situation of higher heat demand, the engine parameters should be set towards later fuel injection and earlier valve opening. A slow-speed marine engine model was developed. The infl uence of the start of fuel injection and the infl uence of the start of the exhaust valve opening on the engine effi ciency, on the exhaust gas temperature, on the amount of heat in the exhaust gases and on the cogeneration system properties: overall effi ciency and power to heat ratio was analyzed. The mechanical power output was constant since the governor dosed the fuel mass per process in order to achieve the selected engine revolution speed. It can be concluded on the basis of the results that a greater amount of fuel is injected as a consequence of a later start of injection. Furthermore, a greater amount of fuel is injected as a consequence of an earlier valve opening. A higher exhaust gas temperature and a grater amount of heat energy in the exhaust gases is manifested for the same engine settings as for those for which a greater amount of fuel was injected. Hence, greater heat and higher temperatures of the exhaust gases result from the larger fuel amounts injected. Since the crankshaft power remains constant, the engine effi ciency decreases for these settings. However, the overall effi ciency of the cogeneration system is the highest for earlier injection and especially for earlier exhaust valve opening. The power to heat ratio increases for earlier injection and for later exhaust valve opening. For a more effective cogeneration application on a low-speed marine diesel engine powered vessel, the key would be a specifi - cally developed governor. The target parameter of a standard governor is the engine revolution speed (or engine power, which is a simple function of the revolution speed), while the input parameters are the current revolution speed and the throttle handle position. A regulator specifi cally developed for the use with a cogeneration system should also have as a target value the exhaust gases heat or the exhaust gases temperature. The variable injection timing and the hydraulic valve drive allow the variation of the beginning of the injection and of the opening of the exhaust valve. However, such systems would allow even a greater fl exibility, such as different injection patterns, split injection and different exhaust valve lift curves. These options are not analyzed in the current work. In turbocharged engines boost air is cooled after raising its pressure in the compressor and before the intake. If the utilization of the heat retrieved from the boost air it is taken into account, the overall effi ciency of the cogeneration system would increase even further. However, it would be necessary to do an economic analysis of the convenience of the recuperation of this heat. Four-stroke diesel engines or other propulsion systems have higher exhaust gas temperatures. Hence it is possible to utilize a greater part of waste heat. However, it has to be taken into account that mechanical energy has a higher value than heat energy and that the main purpose of the here analyzed engine is ship propulsion. Furthermore, low-speed marine diesel engines are fuelled by cheap heavy fuel oil. It can be concluded that the system with a diesel engine is irreplaceable as ship propulsion system on most ship types. The heat exchanger installation in the exhaust pipes reduces exhaust gas temperatures, so it is necessary to take care to avoid condensation and low temperature corrosion. Besides, the heat exchanger confi guration has to be such to avoid soot deposition or to allow its removal. For further investigation it would be interesting to model a different governor which would allow a more fl exible control of the fuel quantity. It would be necessary to analyze the infl uence of greater fuel amounts which the new regulator would allow. Furthermore, the infl uence of different injection patterns, longer injection durations and different exhaust valve curve lifts on engine and cogeneration system performance should be also analyzed. It is supposed that in this way a much higher amount of heat in the exhaust gas could be generated and utilized in a heat exchanger. This could maybe allow savings on a separately fuelled steam generator. It would also be good to develop the engine model to allow the calculation of pollutants emissions and to check what the described variation of parameters means for pollutant emission. Reliability of a Light High Speed Marine Diesel Engine The empirical reliability functions Re (t), failure rate ?e (t) and the density of failures f e (t) of the main marine diesel engine were determined using the empirical data on failures. It was found that the Weibull [1] distribution with parameters ß=2.613 and ?=400 approximated well the reliability of a light high speed marine diesel engine. Serial reliability confi guration of the marine diesel engine subsystem was analysed and the failure frequency as well as the values of the failure rate by subsystems were determined. The further study will also determine the intervals for preventive replacement of the subsystem parts based on the empirical data which will be compared to the recommendations given by the manufacturer of marine diesel engines. For the Croatian Navy (HRM) it is very important to have reliable marine engine systems of multiple uses, such as marine diesel engines. The Croatian Navy naval vessels are equipped with light high speed diesel engines whose strength ranges from 1000-6000 kW. These motors are mostly made in “V” or “star-shaped structure”, and the Croatian Navy missile boats are powered by engines of the world’s two major manufacturers: MTU (Motoren und Turbinen Union, Friedrichshafen) and ZVEZDA, Saint – Petersburg [2]. Operational requirements of modern marine propulsion systems on naval vessels require that marine engines should be, on one hand, as easy and economical as possible with the ability to develop high power, and on the other reliable. These operational requirements include: maximal safety and durability, minimal weight and volume with greatest possible strength, maximum range, greatest possible manoeuvre including high readiness to set off, fl exibility of machinery and optimal effi ciency under small load, damage resistance and minimum possible contamination in operation at sea and minimum possible sounds and vibrations. The fundamental requirement of the reliability of marine engines refers to the number of hours of engine failure-free performance. A light high speed marine diesel engine type M 503 A2 is required to work for 600 hours prior to the major overhaul [3]. Data from the Engine Log Book kept on board were used to determine the reliability of diesel engines. The ship’s service regulations book stipulates keeping diary for all works and all engine starts. Every failure of the engine is kept in the ship’s log, and especially in the log book of engine failures [4]. Failure of the engine means presence of conditions causing the engine out of order according to regulatory parameters. Failures can be divided into the failures which cause the engine breakdown, those which reduce the engine strength and those which do not reduce the engine strength but can be the cause of the engine breakdown. Analyzing empirical data on failures of the marine engine derived from the engine log book, and calculating the reliability function and failure rate function of marine engines it is possible to predict the expected failure-free operating hours and plan the process of preventive maintenance. Reliability of marine diesel engine M 503 Every technical system whose reliable operating depends on each of the subsystems within the system represents a model of serial reliability confi guration. Figure 7 shows a block diagram of a marine diesel engine defi ned with 9 subsystems whose names are given in the caption of Figure 7 [8]. Marine Diesel Engine M 503 A2, produced by ZVEZDA, Saint-Petersburg is a high-speed, multicylindrical, star derived, water-cooled four-stroke engine, supercharged with turbocharger. The engine has 42 cylinders in seven blocks placed star. Each block has six cylinders. The angle between the engine block is 51° 25‘43”. A cylinder block with the cylinder head is cast from aluminum alloy. Maximum engine power while running forward, (the number of engine crankshaft revolutions of 36.67 s with the following atmospheric conditions: air temperature at the inlet 20° C, atmospheric pressure 1013 bar, 70% relative humidity, fuel temperature and seawater at entrance 20 °C) is 2944 kW. Exploitation engine power when running forward and when the number of crankshaft revolutions is 31.67 s is 2429 kW. The RPM of the output shaft of the coupling when running forward and at maximum power is limited to 17.16 10-1 sec. The RPM of the output shaft while running aft is 8.58 10-1 sec. The mass of the diesel engines (dry) with a buckle and gearbox and all the auxiliary aggregates is 7150 kg [3] and [2]. If Tri is a randomly changeable variable representing the period of time by the moment of failure of rth subsystem, then the reliability of a technical system, composed of m serially connected subsystems in one whole, on the basis of equation (5), is defi ned by expression (25): perate without failure during a specifi ed period of time (i.e. time of use), and the maintainability is constructional characteristic specifi ed by diesel engines capability as a technical system to maintain (through preventive maintenance) or return (through corrective maintenance) in the proper operational condition. Time is a major factor and measures the effectiveness of the technical system. Lower downtime increases operational readiness and availability of the system as a whole, so in terms of reliability and maintainability two capabilities are considered: a) capability of staying in operating condition by preventing failures due to aging, wear, corrosion and similar processes (preventive maintenance) and b) capability of rapid return to operational condition after random failures (corrective maintenance). Preventive maintenance (i.e. preventive replacements of parts) is applied when the technical system and its subsystems (parts) have the increasing rate of failures in relation to the use of the system. Specifi cally, in the case of a constant rate of failure, preventive replacements would not affect the security of the achieved level of reliability, and maintenance costs would increase. In this case, when the failures are random, optimal procedure would be replacement of the parts in case of a failure. The conducted research regarding the reliability of the light marine diesel engine showed that one cannot take for granted the assumption about constant failure rate ? (t) = const. Therefore, empirical approximation of functions was taken and it showed that the Weibull distribution with parameters ß = 2.613 and ? = 400 approximates well the reliability of the light high-speed marine diesel engine M 503 A2, and that the expected time of failure-free function E (T) = 348 h. Starting from the established fact that marine diesel engine has an increasing rate of failure, and that failure causes may be different in nature: from overloading the engine and fatigue of material, to wear and corrosion, it was necessary to determine the individual failure rates of diesel engine subsystems (parts) and their contribution to overall reliability and failure rate. It is well known that the mechanical parts of a technical system, unlike electrical parts, have immensely increasing failure intensity in the course of a technical system operation time. This is caused by failure mode of the mechanical parts, and it is manifested as a consequence of the material fatigue, corrosion, material creeps and similar processes which, with time of system operation, only augment the probability of causing the failure of the technical system. Thus, for studying the reliability of a marine diesel engine it was necessary to defi ne the behaviour of so called mechanical parts of that system that have experienced failure, because only with the application of preventive maintenance in the way that the subsystems (parts) of the marine diesel engine are preventively replaced prior to the abrupt increase of the failure rate, the so called state of balance is achieved when it is possible to apply the concept “constant mean time between failures“. The task of preventive maintenance is to prevent the degradation of design characteristics of the marine diesel engine in the way that the planned activities are implemented in order to prolong the material life cycle and prevent the increase of failure rate. Because of that, it is of crucial importance in the continuation of research to defi ne on the basis of empirical data the intervals of preventive part replacements and compare them to the recommendations of marine diesel engine manufacturers. It is a routine procedure that these intervals are to be defi ned in accordance with the previous experiences, and optimised depending on empirical data gathered during the exploitation of the technically complex systems such as marine diesel engines. The empirical data presented in this paper gathered from the marine diesel engine exploitation, which served to defi ne the characteristics of individual subsystem operation and approximation of pertaining functions, are of great importance (together with data on preventive replacement costs) for defi ning optimal intervals of preventive replacement of individual subsystems (parts) of the main system (light high speed marine diesel engine for the propulsion of fast naval vessels) Reaction Propulsors – Basic Concepts and Comprehension The most general theory of reactive propulsors – RP, their characteristics, various types, their principal parts and quantities on which their effi ciency depends are reviewed. Starting from elementary physical laws, a simple and transparent mathematical model of RP, indispensable in ship’s and airplane’s propulsion, is constructed. The specifi c features of RPs compared to other types of propulsors, which are for the sake of unambiguousness called active, are pointed out. Some analogy of RP’s way of action with the process of transformation of heat energy to mechanical work, as described by the second law of thermodynamics, is accentuated too. For the sake of simplicity of the derivation and clearer insight, an isolated RP is considered. Different sources of energy losses are judged, nondimensional parameters of RP loading are reviewed and effi ciencies are defi ned. An original table containing a full set of formulas which are necessary to describe the functioning of RP is presented. Key words: propulsor, reaction propulsion New Stirling Engine Concept (NSC-Engine) with Application of Direct Heat Introduction The effi ciency of thermal engines is directly dependent on the temperatures and temperature differences of their heat reservoirs (heater and cooler). Up to now the improvements of the Stirling engine’s effi ciency have been exclusively achieved through the increase of the engine hot-side temperatures, whereas the possible improvements of the engine thermal effi ciency by lowering their cold-side temperatures, have not been thoroughly researched. In the research work of Wilhelm Servis the infl uence of lowering the engine cold-side temperature down to the deep, cryogenic temperature range on the Stirling engine performance was investigated. The research work results show a regular, very perceptible increase of thermal effi ciency and brake power by lowering the cold-side temperatures of the investigated engines down to the cryogenic temperature ranges. On the basis of this investigation it was possible to defi ne a list of measures to be taken to achieve improvements of the classic Stirling engine. These improvements were achieved through the application of the principle of the direct heat input and output from the engine process, the realisation of engine cooling through the injection and additional evaporation of liquefi ed working medium in the engine cylinder, and the engine cooling down to the cryogenic temperature range. The application of the mentioned measures has led to the development of the “New Stirling Engine Concept (NSC-engine ? NKS-motor)”. These are the engine types PROFIT0, PROFIT1 and PROFIT2, covered by pending and obtained patents. The NSC-engines, realised on the basis of this “New Concept” have the highest possible achievable thermal effi ciency and power production, substantially higher than the presently used thermal engines. These engines, will be able to use nearly any energy source (also porter of heat and cold) and also will be able to operate as real “class zero” or “class ultra low emission” engines (no or very low CO2 - and NOx-emissions). Keywords: aerothermal engine, NSC-engine, Stirling isothermal engine, thermal effi ciency. Current state of the art technology of Stirling engines This paper pleads an opinion that one of possible solutions of the above mentioned problem must start from the facts that large part of actual energy consumption starts with energy conversion in thermal engines from source energy (e.g. energy of diesel fuel) into mechanical work, wherefrom it follows: a) Energy saving and reduction of environmental pollution could be achieved by the application of more effi cient thermal engines than internal combustion engines (like compression (Diesel) - or spark-ignition engines), b) That more energy effi cient engines with a potential for improvement include the engines performing Stirling isothermal process, which belongs to the most effi cient processes for energy conversion from heat to mechanical work. With Stirling engines performing such process, mechanical power is produced from conversion of the heat energy fl ux (by using classical Stirling process) between two heat storages. Temperature difference between heater and cooler (or heat storages) is one of the crucial factors infl uencing the energy conversion effi ciency and intensity. The engine effi ciency increases proportionally with the temperature difference between heater and cooler. Possibilities in increasing the already high classic Stirling engine effi ciency by increasing the heater temperature are exhausted and limited by existing technology. Heater temperatures are as high as the material limits (technological limit). In classical Stirling engines these temperatures are at approx. 1000 K. c) Essential improvements of the Stirling engine effi ciency are possible only by: c1) Imposing the lowering of the cooler temperature and increasing at the same time the temperature difference having the same heater temperature, and c2) By introduction of improved thermodynamic processes and strategies for their performing. In the PhD thesis by Servis [1] the current knowledge on classic Stirling engines is substantially extended to new areas. These new fi ndings and the proposals given in the thesis for the improvements in the Stirling process and strategies for their performing in the so-called “New Stirling Engine Concept (NSCconcept)” need to be presented separately. Therefore, in this 3rd chapter a short overview of “Classical Stirling Engines” will be presented, while the presentation of the “New Stirling Engine Concept (NSC-concept)” will be given in the 4th chapter. Classical Stirling Engine Classical Stirling engine, named after its inventor Robert Stirling (1790-1878) [12] and [13] is a piston heat engine with external fuel combustion (heat addition and extraction are external), which performs isothermal Stirling process cycle. This thermal process is performed through two isothermal and two isochoric changes in a closed cycle with the same working fl uid (which at the time was air). Part of the heat content is recovered in the process cycle by using the heat regenerator. Figure 3 presents the classical Stirling engine from 1816, which is considered by authors as the representative for the class of classical Stirling engines. All main components and the principle of operation of these engines can be seen in the same fi gure. In the time period from the fi rst Stirling engine up to our days (almost 190 years) a large number of Stirling engines were patented and realised, different in their designs and operation modes, but always with the same idea of the isothermal Stirling cycle. To enable the differentiation between various engine designs Stirling engines are classifi ed into different confi guration types; the newest confi guration type was presented by Servis’s new Stirling engine typology [1]. The Stirling engine confi gurations can be classifi ed into three basic designs: ALFA, BETA and GAMMA. These engines with additional design options may be further classifi ed in more details than it was possible so far. According to this new classifi cation, the Stirling engine form 1816 (the fi rst operating engine of that kind, Figure 3) may be classifi ed as BETA(P), which is characterised by having a working piston (06) and a displacer piston (03) in the same engine cylinder (02), with the heat regenerator (03). Engine power was at 1.5 kW. Just for illustration, a few typical designs of the classical Stirling engines will be presented here: 1) Figure 4 presents the photography of the Lehmann engine from 1866 in BETA(P) design, without the heat regenerator. Engine power was at 745 W. This engine is specifi c because of the similarity with the Stirling engine from 1816, but it is without the heat regenerator, resulting in a lower effi ciency. This, horizontal engine was realised 50 years after the Stirling patent, and it was the basis for the fi rst numerical analysis of such engines by Schmidt [1] and [16]. 2) Figure 5 presents the photography of the modern Stirling engine “SES” in the ALFA(P) design, powered by high temperature solar energy. The engine has the heat regenerator and develops a power of 25 kW. This engine was used in the pilot project to research the possible applications of the solar energy. 3) Figure 6 presents the photography of the modern Stirling engine produced by Kockums in the ALFA(P) design with the heat regenerator. The engine has a power of 130 kW. Such engines of higher power are used in submerged submarine propulsion (AIP=Air Independent Propulsion), while for the propulsion on the sea surface diesel engines are applied. Using the numerical analysis of the processes in actual classical Stirling engines and modern diesel engines (presented by the Sabathé cycle), the comparison between various idealised thermal cyclic processes was performed. Figure 7 presents the results of such comparison by using the “Elementary numerical analysis of the zeroed order”. From the results is evident that there is a large potential of the classical Stirling engines when compared with the diesel engines. The advantage of such analysis of idealised thermal cycles is in the remedy of the infl uence of design details in the resulting analysis. It is evident from Figure 7: • in typical conditions of the classical Stirling engine (K1, 348 K/ 900 K) and the diesel engine (D, 348 K/ 2200 K), the thermal effi ciency and power of the diesel engine are higher, but also • at the same operation conditions (K2, 348 K/ 2200 K) the effi ciency and the power of the classical Stirling engines could be even better than that of the diesel engine. From Figure 7 it is evident that by lowering the cooler temperature (K3, 80 K/ 2200 K), signifi cant improvement of the effi ciency and power of the classical Stirling engines is achieved. Servis [1] has presented the reasons why the realised Stirling engines, despite the application of the best isothermal cycle, have not achieved the effi ciency of the diesel engines. Reasons for this are: •Maximum temperatures for heat addition in Stirling engines are lower than and limited to approx. 1000 K (technology limit) and •The compression ratios (up to ?. =2) and pressures in Stirling engines are low (due to the problems in sealing), which reduces the engine effi ciency. In further discussion Servis [1] stresses that the classical Stirling engines, besides the advantages (the most important of which are higher effi ciency and low ecological impact), and besides the already existing technical and technological possibilities for the improvement and competitiveness of such engines, have not achieved the favourable market position. It is now the time to achieve this position. Based on these considerations Servis [1] presents the following list of intentions aimed at realising the goal of improving the competitiveness of Stirling engines: a) Simplifi cation of the engine kinematics, b) Increase of the compression ratio in the engine, c) Possibility to increase the maximum pressure and temperature by introducing the heat into the engine cylinder, d) Application of cryogenic temperatures on engine coolers, e) Simplifi cation of the heat transfer by direct heat transfer, 1 PROFIT0 type cycles of NSC-engines This cycle process is based on the “idealised classical Stirling process cycle” (closed cycle consisting of isothermal compression, isochoric heat addition, isothermal expansion and isochoric heat removal of the working fl uid and with the above cited characteristics of the NSC-concept). Direct heat addition in these engines is achieved by special high- or low-temperature heaters in the engine cylinder or by catalytic fuel combustion, solar heating etc. Direct cooling of the working fluid during isothermal compression (or heat removal from engine cycle) is performed using the in-cylinder cooler, cooled by some cooling fl uid. It is important to point out here that the working fl uid and the cooling fl uid are not in direct contact, so they do not mix together. This is favourable when some aggressive cooling fl uids are applied. PROFIT1 type cycles of NSC-engines This engine cycle is performed as an “idealized NSC-process cycle” (open engine process, presented in Figure 11, with isothermal compression, isochoric and isobaric heat addition, isothermal expansion and isochoric heat removal of the injected and additionally evaporated liquid working fl uid and with the above cited characteristics of the NSC-concept. Direct heat addition in the engine cycle (heating) is performed similarly as in the engines of the PROFIT0 type (e.g. with high temperature solar heating, high temperature catalytic combustion heating etc.). Direct cooling of the working fl uid during the isothermal compression (heat removal) is performed by injection and evaporation of the liquefi ed working fl uid in the engine cylinder (combustible or non combustible). The excess of the working fl uid is exhausted from the engine during the isochoric heat removal part at the end of each engine process cycle. 6.3 PROFIT2 type cycles of NSC-engines This engine cycle is performed as an “idealized NSC-process cycle” (open engine process, presented in Figure 11, with isothermal compression, isochoric and isobaric heat addition, isothermal expansion and isochoric heat removal of the injected and additionally evaporated liquid working fl uid and with the above cited characteristics of the NSC-concept. Direct cooling in the engine cycle is performed (like in the PROFIT1 engine cycle) by direct injection and evaporation of the liquefi ed and combustible working fl uid in the engine cylinder. The direct heat addition to the NSC-process of these engines will be performed in the way that the injected, evaporated and compressed combustible working fl uid is combusted through controlled liquid oxygen injection. The excess of the working fl uid is exhausted from the engine during the isochoric heat removal part at the end of each engine process cycle. The here presented idealised NSC-process cycles (as presented in Figures 7 and 11), are due to their isothermal compression and expansion parts better than those of modern diesel engines (Sabathé) and are also for the identical temperature range the best among the actually used heat engines. The effi ciency and power of the real NSC-process cycle in engines of the PROFIT1 and PROFIT2 types will be improved due to the increased mass of the working fl uid, which is injected in the engine cylinder during the engine cycle. 7 Description of the NSC-engines of PROFIT1 type Based on the above-presented performances, by applying the animations (in property of Servis), Figure 15, it is possible to present and describe the general example of the NSC-engine of the PROFIT1 type (single acting/ multiphase/ non-resonant/ BETA (CR+CH+DHI/O+P+R+I)/ rigid [26]). In Figure 15 it can be seen that the engine consists of: 1) the base engine, here with a crosshead and a cranking mechanism, and 2) the additional part to engine, called the PROFIT1-device, enabling the NSC-process cycle to be performed in the engine. The PROFIT1-device consists of a casing with an exhaust valve and exhausts system, a dynamic displacer, an engine heater (catalytic combustor) with the fuel and oxidiser input and exhaust system, a heat regenerator (here of needle brush type). The working fl uid may be here of combustible or noncombustible type. Figure 16 presents the catalytic combustor for gaseous fuel in commercially available execution for surface temperatures up to approx. 1500 K. Figures 17, 18 and 19 present the PROFIT1 engine type during the idealised NSC-process cycle (presented in Figure 11) in the following positions of the working piston: • Figure 17: at the start of injection of the liquefi ed working fl uid in the cylinder of the PROFIT1-device engine, • Figure 18: at the end of the isothermal compression and start of the isochoric heat addition, • Figure 19: at the end of the isobaric heat addition and start of the isothermal expansion of the working fl uid. At the end of the second stroke the excess of the working fl uid is exhausted from the engine cylinder through the exhaust valve. NSC-Engines of the PROFIT2 type have the PROFIT-device, as described previously (see chapter 6.3). 8 Some possible applications of the NSC-engines of the PROFIT1 type Already during the stage of performing the analyses and experiments of the engines (which were later named NSC-engines based on NSC-concept) the possibilities for their practically unbound applications were evident. Here, some of the most interesting aspects and application possibilities of the PROFIT1 type NSC-engines are presented in Figures 20 to 22 in the form of plant schematic drawing. NSC-Engines, as presented in Figure 20, by schematic drawing of the plant, apply the exergy of liquefi ed gases in LNG pipeline terminals for direct cooling by liquid gas injection during the engine cycle (LNG=Liquid Natural Gas). For direct heat addition to the working fl uid, catalytic combustor is placed in the engine cylinder (small part of combustible gas + air). Excess of the evaporated natural gas (bigger part of it) is exhausted directly to the gas distribution pipeline (under pressure). Engines of such design may be used also for propulsion of LNG-tankers, where the gas provided for evaporative cooling may be combusted completely on catalytic combustor in the engine. Engines of such design are able to reach the maximum power in the MW range (see Figure 23) in the design based on the existing large slow speed two-stroke diesel engines. NSC-Engines, as presented in Figure 21, by schematic drawing of the plant, use injected water (H2 O) as the working and cooling fl uid for direct cooling during the engine process. For direct heat addition, a catalytic combustor using combustible gas (e.g. natural gas, hydrogen etc.), which may be stored in the liquid phase, is applied. Such engines may be also applied for LNG tanker propulsion, where the evaporated gas is used as fuel gas. Engines in such design are provided for small, medium or high power range (see Figure 23). Their design is also based on the existing diesel engines of the same power range. NSC-Engines, as presented in Figure 22, by schematic drawing of the plant, use water for cooling in the same manner as in Figure 21, with the difference that the water for direct cooling is produced by condensed vapour from hydrogen combustion products on the catalytic combustor used for direct heat addition (liquid hydrogen + oxygen). Engines of such design are provided for small and middle power range. They may be of special execution for permanent propulsion of submarines in surface and submerged drive (for the so called AIP=Air Independent Propulsion). Based on the performed experiments and measured NSCengine (here named as PROFIT0) operation parameters, and on the results of numerical simulations of PROFIT1 and PROFIT2 named engines, the following can be concluded: 1. By lowering the temperature of the heat removal, with the idealised NSC-engine process, essentially higher thermal effi ciency is achieved at the same conditions in comparison with the classic Stirling engine process. This effi ciency is for really possible operation of the NSC-engine higher than the effi ciency of the idealised Diesel processes (Sabathé) or of other internal combustion engines. 2. Possible compression ratio increase, increase in process operation pressure, increase in process maximum temperature, direct heat introduction from the heater or heat removal to or from the working fl uid in the engine cylinder space, application of a hybrid displacer and improved heat regenerating device are all contributing to the improvement of the thermal effi ciency of the NSC-engine. The simplifi cation of the engine kinematics (introduction of classic cranking mechanism) enables the introduction of further improvements and lowers the price of the NSC-engine. 3. Energy supply for the NSC-engines is possible from various available energy sources using not only heat, but also freeze (as for example in liquefi ed gases at LNG terminals and onboard ships for liquefi ed gas transport), at heat discharges (for example from thermal power plants, conventional or nuclear, from geothermal sources), from catalytic or other combustors, or from alternative energy sources (for example low- or highconcentrated solar light). This is why the NSC-engines can be applied as basic or utilised drive engines for electric energy production or mechanical drive for ship or vehicle propulsion. Their operation is accompanied by very low or no emission of harmful gases, depending on the engine design and heat power source. As NSC-engines can work without atmospheric air, when they do not need combustion as heat source, they are convenient for space applications with solar heating and heat removal by heat radiation into space with actually three-fold effi ciency when compared with solar cells. They are also convenient for possible application in submarines. 4. The achieved results of R&D activities, presented in this paper, give a reliable basis for further development and application of NSC-engines aiming at rational use of energy resources and prevention of environment pollution for better future life on the Earth.

Adjustable Stroke Kits enable an infinite range of zero to full stroke adjustments from either end, and can also be used to prevent a cylinder rod from rotating. They are designed to be used with most Allenair Cylinders of 1-1/2”’, 2”, 2-1/2”, 3” & 4” bore sizes, except where noted.

TYPE K SHOWN WITH TYPE A CYLINDER

TYPE: K ADJUSTABLE STROKE KIT

Adjustable Stroke Kits for all Cylinders, except Integral Square Head Types and Valve-in-Head Models VAR, VER and VCR consists of Front Foot Mount, Rod Tie Bar, Threaded Adjustment Rod and Guide Rod.

TYPE K SHOWN WITH MODEL SVSR CYLINDER

TYPES: KRE AND KRR AUTOMATIC RETURN

KRE and KRR automatic return Adjustable Stroke Kits are designed to be used with Valve-in-Head Cylinders Models SVA or SVEVA only.

Upon a momentary air bleed (Model SVA) or momentary electrical signal (Model SVEVA), the unit will make one complete cycle. The cylinder rod is normally extended on Types KRE and normally retracted on Types KRR.

These Kits are identical to Type K with the additions of one V2 Valve, Actuating Arm, and all necessary hardware.

For complete description of Cylinders mentioned above, Click Here.

TYPE KRR SHOWN WITH MODEL SVA CYLINDER

TYPE KRE SHOWN WITH MODEL SVEVAE CYLINDER

TYPE: KRVCR FULLY AUTOMATIC RECIPROCATING

This fully automatic reciprocating Adjustable Stroke Kit is designed to be used with the Valve-in-Head Cylinder Model SVA only.

As soon as air pressure is applied, the unit will automatically reciprocate. Because of this it is recommended that a shut-off valve be mounted in the inlet line.

This Kit is identical to Type K with the addition of two V2 Valves, Actuating Arms, and all necessary hardware. For complete description of Model SVA Cylinder, Click Here.

ALLENAIR CORP. 255 East Second St. Mineola, NY 11501-3520 USA Telephone: 1-516-747-5450 ~ Fax: 1-516-747-5481 Email: info@allenair.com

FROM OUTSIDE THE USA and CANADA - CONTACT

MASTEN WRIGHT 280, State Street North Heven, CT. 06473, USA Phone: +1-203-230-4130 ~ Fax: +1 203-248-8093 Email : uc@exportdept.com

OR

W.R. MAGNUS, INC. 20 NORTH WACKER DRIVE SUITE 3333 CHICAGO, IL 60606, USA TELEPHONE: 312-263-6124 ~ FAX: 312-263-4879 Email: rwmagnus@coexport.com

Pneumatic Cylinders, Digital feedback air Cylinders, Air Cylinders, Small Bore Cylinders, Valve In Heads Cylinders, Cyl Check Cylinders, Air Clamp Cylinders, Small Bore SS Cylinders, Position Feedback Cylinders, Valves, 4 Way ¼, 3/8, & ½ NPT Valves, 2 Way, 3 Way Solenoid 1/8 & ¼ NPT Valves, 4 Way Direct Acting 1/8 NPT Valves, Time A Valve, Air Switch Valve, 1/8 NPT Poppet Valve

An adsorption heat pump can be thought of as utilising a chemical rather than a mechanical compressor and is driven by heat rather than mechanical work.

Basic adsorption cycle

The operation of adsorption heat pumps and refrigerators is based on the ability of porous solids (the adsorbent) to adsorb vapour (the adsorbate or refrigerant) when at low temperature and to desorb it when heated. The most basic configuration for an adsorption refrigerator or heat pump is illustrated in the animation below.

This basic system is comprised of two linked containers, one of which contains the solid adsorbent and is termed the generator and the other is the combined evaporator and condenser or receiver in which the refrigerant is evaporated and condensed. Initially the system is at a low temperature and pressure and the adsorbent contains a high concentration of refrigerant, whilst the receiver contains only refrigerant gas.

The generator is then heated which causes refrigerant to be desorbed, raising the system pressure. Refrigerant condenses in the receiver, rejecting heat and producing a useful heat output if the system is to be used as a heat pump.

Cooling the generator back down to its initial temperature completes the cycle and causes the adsorbent to readsorb the refrigerant. The system pressure is reduced and the liquid refrigerant in the receiver evaporates, absorbing heat. This produces the useful cooling effect if the system is to be used as a refrigerator.

Although the heating and cooling provided by a single generator is discontinuous, it can be made continuous by operating two or more generators out of phase.

Advantages

Adsorption heat pumps utilise natural refrigerants such as water, ammonia or methanol, which have no harmful effects on the environment. They can also offer better utilisation of primary energy and be driven by waste heat or solar energy.

Challenges

The basic adsorption cycle has low efficiency, which requires methods for recovery of heat between adsorbent beds, and the inherently low thermal conductivity of available adsorbent materials results in low power densities and machines of high capital cost. Research work is focused on improving performance in these two main areas.

Adsorption chiller designed and built at Warwick -now installed in Italy.

AEROSHELL PISTON ENGINE OILS AEROSHELL PISTON ENGINE OILS For many years the performance of aircraft piston engines was such that they could be lubricated satisfactorily by means of straight mineral oils, blended from specially selected petroleum base stocks. However, demand for oils with higher degrees of thermal and oxidation stability necessitated ‘fortifying’ them with the addition of small quantities of non-petroleum materials. The first additives incorporated in straight mineral piston engine oils were based on the metallic salts of barium and calcium. In highly-rated engines the performance of these oils with respect to oxidation and thermal stability was excellent, but the combustion chambers of the majority of engines could not tolerate the presence of the ash deposits derived from these metal-containing additives. Following extensive operational success in a wide range of civil engines, military specifications based on the general characteristics of AeroShell W Oils were prepared and issued. AeroShell W Oils were in service with the world’s airlines and aircraft operators for many years when they operated big transport piston-engined aircraft, during which time these oils became virtually the standard for all aircraft piston engines. Nevertheless, supplies of straight AeroShell Oils remained available primarily for running-in the aircraft piston engine and for the few operators who required them. Today these oils (both AeroShell W Oils and AeroShell Oils) are still required for the smaller piston-engined aircraft flying in air taxi operations, flying clubs or flown by private pilots. In the early 1980s a semi-synthetic multigrade W oil for piston engines (AeroShell Oil W 15W-50) was added to the range. This grade has become very popular amongst engine manufacturers and operators alike. In order to cater for those Lycoming engines which need improved load-carrying (i.e. those engine models which require the addition of Lycoming Additive LW 16702) AeroShell Oil W 15W-50 was upgraded in 1986 to include an antiwear additive. PISTON ENGINE OILS To overcome the disadvantages of harmful combustion chamber deposits, a non-metallic, i.e. non-ash forming, polymeric additive was developed which was incorporated in blends of selected mineral oil base stocks, to give the range of AeroShell W Oils. 3.1 AEROSHELL PISTON ENGINE OILS In recent years utilisation of piston engine aircraft has decreased, resulting in the aircraft spending more time on the ground. This led to an increase in corrosion being seen inside the engine. In order to combat this, AeroShell Oil W 15W-50 was further upgraded in 1993 to include a very effective anti-corrosion additive package. 3.2 To cater for the demands of operators of light sport aviation piston engines, two new grades – AeroShell Oil Sport Plus 2 (for 2-stroke engines) and AeroShell Oil Sport Plus 4 (for 4-stroke engines) have recently been introduced. With the development of compression ignition (Diesel) piston engines specifically for the aviation market, Shell Aviation has been working closely with the OEMs to develop appropriate lubricants for this new engine type. The result of these co-operative efforts was the development of AeroShell Oil Diesel 10W-40, to be followed by the recent launch AeroShell Oil Diesel Ultra. SPECIFICATIONS Since the 1940s, piston engine operators have relied on two U.S. Military Specifications for defining piston engine lubrication requirements. Beginning with the non-dispersant MIL-L-6082 oils and continuing through the MIL-L-22851 Ashless Dispersant products, the U.S. Military Specifications were the standards for oil performance worldwide. In military circles Grades 1065 and 1100 as well as Type II and III were familiar grade identifications, whilst in civil use Grades 65, 80, 100 and 120 were common. However, that has all changed. The SAE Fuels and Lubricants Technical Committee 8 – Aviation Piston Engine Fuels and Lubricant Committee worked very closely with the U.S. Navy to convert these Military Specifications into SAE Standards. Also involved were oil manufacturers, engine builders, test laboratories and the American FAA. In due course agreement was reached on a new set of performance standards for piston engine oils. These new SAE Standards are J-1966 Lubricating Oil, Aircraft Piston Engine (Non-Dispersant) and J-1899 Lubricating Oil, Aircraft Piston Engine (Ashless Dispersant), both of which have now been adopted for use. The adoption of these new SAE Standards means that the two Military Specifications (MIL-L-6082 and MIL-L-22851) are now obsolete. The most obvious change for users is the move from the old Grade or Type Number system to the more common SAE viscosity classification. Thus products in both SAE specifications are defined as SAE 30, 40, 50 or 60. In addition, for the first time, multigrade aviation oils are included in the new specifications. The U.K. has now cancelled DERD 2450 and DERD 2472 and adopted the SAE specifications. FUNCTION OF PISTON ENGINE OIL A piston engine oil’s function inside a piston engine is to: - reduce friction between moving parts - provide necessary cooling to internal areas - cushion moving parts against shock and help seal piston rings to cylinder walls - protect highly finished internal parts of the engine from rust and corrosion - k eep interior of engine clean and free of dirt, sludge, varnish and other harmful contaminants APPLICATION AeroShell Oils and AeroShell W Oils are intended for use in four-stroke (four-cycle) aircraft reciprocating piston engines. They are not recommended for use in automotive engines converted for use in aircraft, and in these cases the conversion shop should be consulted for proper oil recommendations. The term “ashless dispersant” was given to aviation oils to distinguish them from straight mineral aircraft piston engine oils. Automotive and heavy duty truck engine oils contain ashless dispersants and ash-containing detergents. They were traditionally called detergent oils (some aircraft operators incorrectly refer to ashless dispersant oils as “detergent oils”). Because of the negative effect of ash on aircraft engine performance, it is very important that ash-containing oils are NOT used in an aircraft piston engine. 3.3 PISTON ENGINE OILS PISTON ENGINE OILS For those operators who prefer a single grade but still want the anti-wear and anti-corrosion benefits of the multigrade oil, AeroShell Oil W80 Plus and AeroShell Oil W100 Plus have been added to the range of ashless dispersant oils. These new specifications include upgraded and improved tests and have been designed to meet current technology, and include the latest test methods and precision limits. AEROSHELL PISTON ENGINE OILS Due to differences in metallurgy, operating conditions and fuel specifications, an aircraft oil will not meet all of the automobile/heavy-duty engine’s requirements. In addition, the aviation oils are not qualified for this application and their use could result in voiding the warranty and/or reduction in engine life. ENGINE CONVERSION Thus automobile oils MUST NOT be used in aircraft engines which use or specify SAE J-1899 or J-1966 oils. Similarly aviation oils MUST NOT be used in automobile engines. Experience has shown that AeroShell W Oils do not loosen or affect the hard carbonaceous material already deposited in high-time engines, and may therefore be introduced at any time during the operational life of an engine. Elaborate precautions are not needed when changing from straight mineral oil to AeroShell W Oils, since both types of oil are compatible with each other. 3.4 3.5 For the majority of aircraft piston engines the selection of the right grade is important to maximise engine performance and engine life. Running-in Normal operation use use AeroShell Oils AeroShell W or W Plus Oils The easiest and possibly the best way of converting a fleet of engines to an AeroShell W Oil is to ‘top-up’ with the oil commencing from a given date. The majority of operators use this method following procedures recommended by the engine’s manufacturer. However, other operators have drained engines and refilled them with AeroShell W Oil. If this procedure is adopted, the oil filters should be checked after a ground run and at short intervals during initial operation, because the fresh charge of AeroShell W Oil may disperse ‘pockets’ of partly oxidised straight mineral oil which may have bound together and retained flaky carbonaceous material during previous operation. SELECTION OF CORRECT VISCOSITY GRADE AeroShell Oils and AeroShell W Oils are each available in four grades. The grades differ only by viscosity and thus cover the needs of all reciprocating engines now in airline and general aviation operation. There is no general rule by which the correct grade for every engine type can be chosen, but the following table, based on recommendations from Lycoming, provides approximate guidance for selecting the most suitable grade, based on the average ambient outside air temperature at engine start-up. AeroShell Oil 65 80, W80 and W80 Plus 100, W100 and W100 Plus 120 and W120 Outside air temperature °C Below –12 –17 to 21 16 to 32 Above 26 Corresponding SAE No. 30 40 50 60 Note: This table does not apply to AeroShell Oil W 15W-50. N.B. For large engines the choice depends greatly upon the operator’s preference and past experience. Traditionally the choice seems to be associated with climatic zones: AeroShell Oil W100 or W100 Plus is preferred for temperate regions and AeroShell Oil W120 for warmer climates. OIL CHANGE INTERVAL Almost all oil change recommendations specify not only an engine hour time limit, but also a calendar time limit; typically 4 or 6 months depending upon engine manufacturer. On low usage aircraft the calendar time limit is usually more critical than the engine hour limit. The need for frequent oil changes in aircraft is not caused by the oil wearing out, but rather by the oil becoming contaminated with by-products of combustion, dirt, water (both atmospheric as well as from condensation inside an engine) and unburnt fuel. This contamination can cause corrosion in the oil wetted areas of an engine and thus changing the oil removes these contaminants and helps to minimise corrosion. In order to minimise this corrosion inside low usage engines, calendar time changes are important. PISTON ENGINE OILS PISTON ENGINE OILS SELECTION OF RIGHT GRADE OF OIL AEROSHELL PISTON ENGINE OILS OIL CHANGE EXTENSION RADIAL ENGINES Many operators are interested in extending oil change intervals. As a general rule extensions are not recommended for the following reasons: Radial engines utilise special parts and, depending upon the type of aircraft, application and climate are often subject to specific problems not seen in other types of piston engines. - many engine manufacturers do not approve extended intervals - possibility of losing engine manufacturers’ warranty on the engine - possibility that extended intervals will shorten engine life In a radial engine each bank of cylinders has all of the cylinders in the same plane and transmits power through a single master rod bearing to the crankshaft. This master rod bearing is subjected to high loading and absorbs the shock and vibration from the cylinders and thus requires very good protection from the lubricant. Generally radial engines have greater piston and bearing clearances and thus require a higher viscosity oil. 3.7 The initial enthusiasm in the U.S. for extended intervals has declined due to problems associated with lead sludge found in engines. Many operators have now reverted back to the engine manufacturers’ oil change recommendations and found that these problems disappear. Operators are urged to follow the engine manufacturers’ or rebuilders’ recommendation for oil change interval. BREAK-IN PROCEDURE Some aircraft engine manufacturers and rebuilders/overhaul agencies suggest in their service bulletins the use of straight mineral oil in new or newly overhauled engines for breakin. These straight mineral oils are usually recommended for the first 25 to 50 or even 100 hours of operation, or until the oil consumption stabilises. Other rebuilders or manufacturers, especially for such engines as the Lycoming O-320H and O/LO360E, allow either ashless dispersant or straight mineral oil for break-in, whereas ashless dispersant oils are mandated for break-in for all turbocharged Lycoming engines. Operators should check with engine manufacturers or rebuilders for the correct recommendation for the specific engine and application. STABILITY IN STORAGE AeroShell W Oils are inherently stable and, providing they have been stored and handled correctly, prolonged storage does not have any effect on their quality, properties or performance. As a result of all this heavy duty stress, it is recommended that for radial engines used in normal operation (all operations except agricultural spraying), an oil such as AeroShell Oil W120 is used in moderate to temperate climates and AeroShell Oil W100 in cooler climates (if breaking-in, then AeroShell Oil 120 and 100 respectively). Alternatively AeroShell Oil W 15W-50 could be used in those radial engines for which it is approved. None of these oils contain zinc additives which if used would quickly destroy the master rod bearing. Agricultural operations represent a special problem for an oil used in radial engines. This is because of problems with high dirt and overspray ingestion into the oil. The best way to combat this is proper maintenance, good flying procedures and frequent oil changes. VINTAGE AIRCRAFT Vintage aircraft piston engines, including vintage radial engines, were approved on oils produced at the time the engine was originally manufactured. Many of these oils are no longer available. If the engine was approved on an aviation oil other than a MIL-L-6082 or a MIL-L-22851 oil then operators should consult with either the engine rebuilder or oil supplier. On no account assume that present oils are direct replacements for old vintage aircraft applications. PISTON ENGINE OILS PISTON ENGINE OILS 3.6 AEROSHELL PISTON ENGINE OILS NOTES OIL ANALYSIS It is important to note that the information gained is only as good as the sampling procedure. A single test is not enough to reveal trends and significant changes, it can only tell an operator if there is already a serious problem. Operators should therefore:■ ■ ■ T ake samples properly For best results, take the sample about midway through the draining of hot oil from the sump. A sample pulled off the bottom may be dirtier than normal. The sample should be taken the same way every time. An improperly taken sample can lead to mistaken conclusions about engine problems.  ely on a series of consistent tests over time R Operators should look for significant changes or trends over time, not just absolute values. T ake samples consistently Always take the sample the same way at the same time interval. Always properly label the sample so that its identity is known. It is likely that higher wear metal levels will occur during break-in or following some maintenance procedures. NON-AVIATION USE OF AEROSHELL PISTON ENGINE OILS In selecting an AeroShell piston engine oil for a non-aviation application the properties of the oil must be examined. This will only give an approximate indication as to the expected performance in the specific application. However, such data must be regarded as guidance only. There is no laboratory test that can give a complete prediction of performance in actual use, and the final stage in any decision must involve performance tests in either the actual equipment or in the laboratory/test house under conditions expected in service. 3.9 PISTON ENGINE OILS PISTON ENGINE OILS 3.8 Routine oil analysis is now seen as a valuable part of a good maintenance programme. Increasingly, operators are adopting oil analysis programmes in order to help discover problems before they turn into major failures. Typically these programmes consist of spectrometric wear metal check, together with a few simple oil tests such as viscosity and acidity. Shell Companies can offer this service to operators. AEROSHELL PISTON ENGINE OILS AEROSHELL OILS 65, 80, 100 and 120 AeroShell straight mineral oils are blended from selected high viscosity index base stocks. These oils do not contain additives except for a small quantity of pourpoint depressant (which is added when improved fluidity at very low temperature is required) and an antioxidant. 3.10 AeroShell Oil 65 – AeroShell Oil 80 – AeroShell Oil 100 – AeroShell Oil 120 The suffix for each grade corresponds to the viscosity of the oil at 210°F in Saybolt Universal Seconds. The appropriate grades of these AeroShell Oils are approved for use in four-stroke (fourcycle) certified aircraft reciprocating piston engines (except Porsche) and other aircraft radial engines which use oil to specification SAE J-1966 (MIL-L-6082) and which do not require use of an oil containing a dispersant additive. AeroShell Oils are used primarily during break-in of most new or recently overhauled four-stroke aviation piston engines. The duration and lubrication recommendations for break-in vary, so operators should refer to the original engine manufacturer and/or overhaul facility for specific recommendations. SPECIFICATIONS The U.S. Specification SAE J-1966 replaces MIL-L-6082E. AeroShell Oil 65 80 100 French (AIR 3560/D (AIR 3560/D) (AIR 3560/D Grade SAE 30) Grade SAE 40) Grade SAE 50) Russian - NATO Code O-113 - Obsolete Joint Service Designation OM-107 OM-170 OM-270 Obsolete MS-14 120 MS-20 - O-117 Obsolete - 3.11 OM-370 Obsolete ( ) indicates the product is equivalent to specification. Typical Properties 65 80 100 120 SAE viscosity grade 30 40 50 60 Density @ 15°C 0.879 0.880 0.886 0.889 Kinematic viscosity mm²/s @ 100°C @ 40°C 11.8 - 14.6 140 19.7 230 24.8 - Viscosity index 94 Above 94 Above 94 94 Pourpoint kg/l Although it was planned to replace the British Specification DERD 2472 with a DEF STAN specification this has now been put into suspension and instead the SAE specification has been adopted. °C –20 Below –17 Below –17 –11 Flashpoint Cleveland Open Cup °C 250 Above 240 Above 250 Above 250 AeroShell Oil 65 Total acidity <0.1 <0.1 <0.1 <0.1 U.S. Approved Approved J-1966 J-1966 SAE Grade 30 SAE Grade 40 0.3 0.45 0.48 0.51 Copper corrosion @ 100°C 1 1 1 1 Ash content 0.006 0.006 0.006 0.006 British - 80 Approved J-1966 SAE Grade 40 100 120 Approved Approved J-1966 J-1966 SAE Grade 50 SAE Grade 60 Approved - J-1966 SAE Grade 50 Table continued Sulphur mgKOH/g %m %m PISTON ENGINE OILS PISTON ENGINE OILS APPLICATIONS AeroShell Oils are available in four different viscosity grades: Table continued AEROSHELL PISTON ENGINE OILS AEROSHELL OILS W80, W100 and W120 AEROSHELL W OILS ■ Promote engine cleanliness ■ Help keep engines sludge free ■ Help reduce oil consumption ■ Help engines reach TBO (Time Between Overhaul) ■ Protect highly stressed engine parts against scuffing and wear 3.13 APPLICATIONS AeroShell W Oils are available in four different viscosity grades: SPECIFICATIONS AeroShell Oil W80 – AeroShell Oil W100 – AeroShell Oil W120 The U.S. specification SAE J-1899 replaces MIL-L-22851D. The suffix for each grade corresponds to the viscosity of the oil at 210°F in Saybolt Universal Seconds. Although it was planned to replace the British Specification DERD 2450 with a DEF STAN specification this has now been put into suspension and instead the SAE specification has been adopted. AeroShell W Oils are intended for use in four-stroke (four-cycle) certified reciprocating piston engines, including fuel-injected and turbocharged engines. AeroShell W Oils are not recommended for use in automotive engines. For automotive engines converted for use in aircraft, the specific engine manufacturer or the conversion agency should be consulted for proper oil recommendation. Most radial engine operators use AeroShell Oil W120 in warm weather operations with AeroShell Oil W100 or AeroShell Oil W 15W-50 being used in cooler ambient temperatures. AeroShell Oil W100 or AeroShell Oil W 15W-50 are the common choices for most operators of Lycoming and Continental flat engines but, during colder parts of the year, use of AeroShell Oil W80 in place of AeroShell Oil W100 would be an excellent choice. Although some aircraft engine manufacturers and rebuilders/overhaul agencies suggest in their service bulletins the use of straight mineral oil in new or newly overhauled engines, other rebuilders or manufacturers, especially for such engines as the Lycoming O-320H and O/LO360E, allow either ashless dispersant or straight mineral oil for break-in, whereas ashless dispersant oils are mandated for break-in for all turbocharged Lycoming engines. Operators should check with engine manufacturers or rebuilders for the correct recommendation for the specific engine and application. AeroShell Oil W80 W100 W120 U.S. Approved J-1899 SAE Grade 40 Approved J-1899 SAE Grade 50 Approved J-1899 SAE Grade 60 British Approved J-1899 SAE Grade 40 Approved J-1899 SAE Grade 50 Approved J-1899 SAE Grade 60 French (AIR 3570 Grade SAE 40) (AIR 3570 Grade SAE 50) (AIR 3570 Grade SAE 60) Russian MS-14 MS-20 - NATO Code O-123 Obsolete O-125 Obsolete O-128 Obsolete Joint Service Designation OMD-160 OMD-250 OMD-370 ( ) indicates the product is equivalent to specification. PISTON ENGINE OILS PISTON ENGINE OILS 3.12 AeroShell W Oils were the first non-ash dispersant oils to be used in aircraft piston engines. They combine non-metallic additives with selected high viscosity index base stocks to give exceptional stability, dispersancy and anti-foaming performance. These additives leave no metallic ash residues that can lead to deposit formation in combustion chambers and on spark plugs, which can cause pre-ignition and possible engine failure. AEROSHELL PISTON ENGINE OILS NOTES EQUIPMENT MANUFACTURERS APPROVALS AeroShell W Oils are approved for use by the following engine manufacturers:3.15 Textron Lycoming 301F Teledyne Continental MHS 24B Pratt & Whitney Service Bulletin 1183-S Curtiss Wright Various Service Bulletins – refer to relevant Bulletin Franklin Engines Various Service Bulletins – refer to relevant Bulletin Typical Properties W80 W100 W120 SAE viscosity grade 40 50 60 Colour ASTM 4.0 4.0 5.0 Density @ 15°C 0.880 0.884 0.887 Kinematic viscosity mm²/s @ 100°C @ 40°C 14.5 118 20.2 200 24.8 270 Viscosity index 118 118 115 Pourpoint °C Below –22 Below –18 Below –18 Flashpoint Cleveland Open Cup °C Above 240 Above 260 Above 240 Total acidity <0.1 <0.1 <0.1 0.3 0.38 0.51 1 1 1 0.006 0.006 0.006 Sulphur Copper corrosion Ash content kg/l mgKOH/g %m @ 100°C %m A viscosity/temperature chart is shown at the end of this section. PISTON ENGINE OILS PISTON ENGINE OILS 3.14 AEROSHELL PISTON ENGINE OILS AEROSHELL OIL W 15W-50 The anti-wear additive system in AeroShell Oil W 15W-50 provides outstanding wear protection for critical camshafts, lifters and other high wear components. The anti-corrosion additive package in AeroShell Oil W 15W-50 helps protect low usage engines and engines in high humidity climates against rust and corrosion of critical engine parts such as camshafts and lifters. These results indicate that AeroShell Oil W 15W-50 can provide maximum anti-corrosion protection for aircraft piston engines, when combined with proper maintenance practices and proper operating conditions. Because of the improved flow characteristics of AeroShell Oil W 15W-50, operators may observe slightly lower oil temperatures in some aircraft. On larger aircraft, the oil cooler flap will normally compensate for this change. However, in small aircraft, oil temperature could be reduced slightly. Operators should always check the oil temperature to ensure that they are in the range specified by the manufacturer. Most manufacturers recommend cruising oil temperatures between 82 to 93°C (180 to 200°F). Oil temperatures significantly below this range can result in excessive water and fuel contamination in the crankcase. AEROSHELL OIL W 15W-50 ■ ■ ■ AeroShell Oil W 15W-50 provides superior anti-corrosion protection for all types of certified aircraft piston engines. When used with proper maintenance procedures, the product provides maximum protection and improves the likelihood that aircraft engines will reach TBO. In addition, this product provides outstanding high temperature oxidation protection for hot running engines. It is designed to keep engines cleaner with less sludge and varnish build-up in critical ring belt and other areas. APPLICATIONS AeroShell Oil W 15W-50 is intended for use in certified four-stroke (four-cycle) aircraft piston engines. AeroShell Oil W 15W-50 is superior to single grade oils in almost every application. It offers easier starting, better lubrication after start-up, reduced wear, reduced corrosion and rusting, and improved cleanliness, with oil pressures and temperatures equal to that of single grade SAE 50 oils at fully warmed up conditions. The anti-corrosion additive system is designed to prevent rust or corrosion in all types of aircraft piston engines. In comparative testing of camshaft rusting under high humidity conditions, AeroShell Oil W 15W-50 was almost entirely rust free while camshafts conditioned on other oils showed heavy rusting on some cam lobes and bearing surfaces. ■ ■ ■ ■ Provides excellent rust and corrosion protection for aircraft engines Promotes engine cleanliness, fights wear, offers excellent anti-foam properties Helps reduce oil consumption by up to 50% and provides superior oil flow at low temperatures Compatible with other approved aircraft piston engine oils Functions as an all season oil, no seasonal changes needed Reduces fuel consumption by up to 5% over single grades Provides superior high temperature oxidation stability Refer to General Notes at the front of this section for information on oil change recommendations and engine break-in. AeroShell Oil W 15W-50 is not recommended for use in automotive engines. For automotive engines converted for use in aircraft, the specific engine manufacturer or the conversion agency should be consulted for proper oil recommendation. 3.17 PISTON ENGINE OILS PISTON ENGINE OILS 3.16 AeroShell Oil W 15W-50 is a unique blend of high quality mineral oil and over 50% synthetic hydrocarbon base stocks, plus the AeroShell Oil W ashless dispersant additive system. This semi-synthetic blend offers high performance in a wide variety of applications and conditions. The synthetic base stock performance provides for better cold temperature pumping and protection than single grade oils. In addition, the blend of synthetic and high quality mineral base stocks provide high temperature performance superior to that of other fully approved aircraft piston engine oils. The mineral base stocks help disperse lead byproducts of combustion, thereby keeping engines free of “grey paint” or lead sludge that can be a problem with some fully synthetic oils. AEROSHELL PISTON ENGINE OILS AeroShell Oil W 15W-50 already contains, in the correct proportions, an anti-wear additive equivalent to the Lycoming additive LW 16702; operators who use AeroShell Oil W 15W50 DO NOT need to add this Lycoming additive to the oil. AeroShell Oil W 15W-50 is qualified for use in all Continental Motors’ liquid cooled and air cooled aircraft piston engines. Typical Properties SAE J-1899 Multigrade Typical Oil type - Mixed synthetic hydrocarbon and mineral SAE viscosity grade Multigrade Multigrade Colour ASTM - 4.0 Density @ 15°C Report 0.86 Kinematic viscosity mm²/s @ 100°C @ 40°C - - 19.6 140 Viscosity index 100 min 157 Pourpoint °C Report –39 Flashpoint Cleveland Open Cup °C 220 min 238 mgKOH/g 1.0 max 0.01 %m 0.6 max 0.2 1 max 3 max 1 2 Total acidity kg/l U.S. Approved SAE J-1899 Grade Multigrade British Approved SAE J-1899 Grade Multigrade French - Russian - Copper corrosion 3 hrs @100°C 3 hrs @ 204°C NATO Code O-162 Obsolete Ash content %m 0.011 max 0.006 Joint Service Designation OMD-162 Trace sediment Must pass Passes Foaming tendency Must pass Passes Elastomer compatibility AMS 3217/1 72 hrs @ 70°C swell % AMS 3217/4 72 hrs @ 150°C swell % Must pass Passes Must pass Passes Trace metal content Must pass Passes Compatibility Must pass Passes EQUIPMENT MANUFACTURERS APPROVALS AeroShell Oil W 15W-50 is approved for use by the following engine manufacturers: Textron Lycoming 301F Service Bulletins 446E and 471B Service Instruction 1409C Continental MHS 24A SIL 99-2 Pratt & Whitney Service Bulletin 1183-S FAA Airworthiness Directive 80-04-03 R2 Sulphur A viscosity/temperature chart is shown at the end of this section. This product is made in more than one location and the approval status and typical properties may vary between locations. 3.19 PISTON ENGINE OILS PISTON ENGINE OILS 3.18 SPECIFICATIONS AeroShell Oil W 15W-50 was developed in co-operation with Textron Lycoming and Continental Motors and conforms to their specifications 301F and MHS-24A respectively. This oil is also approved under Military Specification MIL-L-22851 which is now obsolete and has been replaced by the SAE J-1899 specification. AeroShell Oil W 15W-50 is also approved for use in all Pratt & Whitney radial aircraft engines. In addition AeroShell Oil W 15W-50 meets the provisions of Lycoming Service Bulletin 446C and 471, plus Service Instruction 1409A and meets the American FAA Airworthiness Directive 80-04-03 which specifies special anti-wear requirements for certain engine models. AEROSHELL PISTON ENGINE OILS AEROSHELL OILS W80 PLUS and W100 PLUS EQUIPMENT MANUFACTURERS’ APPROVALS AeroShell Oils W80 Plus and W100 Plus are approved for use by the following engine manufacturers: APPLICATIONS The advanced additives in AeroShell Oils W80 Plus and W100 Plus provide better rust and wear protection than conventional single grades. The additives work as a protective barrier to prevent critical parts from being slowly degraded by rust or wear, especially when an aircraft sits idle. This protection helps keep the camshaft and lifters coated, reducing the likelihood of premature damage and helping operators reach TBO. Textron Lycoming 301F Service Bulletins 446E and 471B Service Instruction 1409C Continental SIL 99-2 FAA Airworthiness Directive 80-04-03 R2 AeroShell Oils W80 Plus and W100 Plus Typical Properties W80 Plus W100 Plus Colour ASTM <3.0 <3.0 Density @ 15°C 0.883 0.887 Kinematic viscosity mm²/s @ 100°C @ 40°C 14.0 113 19.5 190 Viscosity index 100 min 124 119 Pourpoint °C –30 –21 Flashpoint Cleveland Open Cup °C 260 288 mgKOH/g 0.02 0.02 %m 0.40 0.44 Copper corrosion 3 hrs @100°C 1A 1B Ash content 0.001 0.002 ■ ■ ■ ■ ■ Blended from selected high viscosity mineral base oils Contains AeroShell’s proven W Oils additive package Additional anti-wear additives (containing Lycoming additive LW 16702) Additional anti-corrosion additives Fully compatible with other approved aircraft piston engine oils SPECIFICATIONS Approved SAE J-1899 SAE Grade 40 (AeroShell Oil W80 Plus) Approved SAE J-1899 SAE Grade 50 (AeroShell Oil W100 Plus) AeroShell Oils W80 Plus and W100 Plus already contain, in the correct proportions, an anti-wear additive equivalent to the Lycoming additive LW 16702; thus complying with FAA Airworthiness Directive 80-04-03. Operators who use AeroShell Oils W80 Plus and W100 Plus DO NOT need to add this Lycoming additive to the oil. AeroShell Oils W80 Plus and W100 Plus are qualified for use in all Continental Motors liquid cooled and air cooled aircraft piston engines. 3.21 Total acidity Sulphur kg/l %m A viscosity/temperature chart is shown at the end of this section. PISTON ENGINE OILS PISTON ENGINE OILS 3.20 AeroShell Oil W80 Plus and AeroShell Oil W100 Plus are new single grade oils that combine the single grade, ashless dispersant performance found in AeroShell Oils W80 and W100 and the anti-wear/anti-corrosion additives of AeroShell Oil W15W-50 Multigrade. They are the oils for pilots who prefer a single grade but who also want the extra protection and performance from the additive package. AEROSHELL PISTON ENGINE OILS AEROSHELL OIL SPORT PLUS 2 Provides full performance with both unleaded and leaded (AVGAS 100LL) fuel types. This oil can be used in all climates. APPLICATIONS AeroShell Oil Sport Plus 2 is intended for use in 2-stroke aircraft piston engines, which have previously relied on general purpose 2-stroke oils originally developed for ground/marine based applications. ■ ■ ■ Suitable for all air-cooled and water-cooled engine types. Can be used in premix and separate oil injection systems. Can be used with unleaded and leaded (AVGAS 100LL) fuels SPECIFICATIONS No Aviation specifications yet defined. FEATURES AND BENEFITS ■ First specific oil for Light Sport and Very Light/Ultra Light 2-stroke aircraft engines ■ High Film & Shear strength formulation specifically designed for strenuous operating conditions experienced by these types of aviation engines ■ Promotes engine cleanliness – protects engine parts such as pistons, rings and exhaust ports from excessive (or harmful) deposits and coking ■ Outstanding performance in regard to ring ‘sticking’ ■ Excellent ‘clean burn’ performance ■ Helps to protect engine parts from corrosion during engine shutdown and storage ■ Helps engine achieve TBO (Time Between Overhauls) ■ Suitable for use in oil injection and pre-mixed oil/fuel systems ■ Protects highly stressed engine parts against scuffing and wear ■ Can be used in any climate ■ Superior performance compared to synthetic 2-stroke products when used in the aviation application ■ Advanced anti-rust and anti-wear package ■ Dyed green for better recognition ■ Can be mixed with other mineral & synthetic 2-stroke oils previously used DO NOT use AeroShell Oil Sport Plus 2 in engines that are designed to use Ashless Dispersant aviation piston engine oils such as AeroShell W oils. This includes aircooled Continental Motors, Textron Lycoming, Jabiru and ROTAX® 4-stroke engines. Fully approved – all ROTAX® 2-stroke series engines, ROTAX® Service Instruction SI-2ST-008 Selection of suitable operating fluids for ROTAX® 2-stroke UL engines (series). Typical Properties Sport Plus 2 Density @ 15°C kg/l 0.88 Meets the requirements of API TC Kinematic viscosity mm²/s @ 100°C @ 40°C 9.0 61.1 Viscosity index 123 Pourpoint °C –33 Flashpoint Cleveland Open Cup °C 65 Please consult Operators Handbook/Manual to confirm the correct fuel/oil mix ratio before use. 3.23 PISTON ENGINE OILS PISTON ENGINE OILS 3.22 Developed in conjunction with ROTAX®, AeroShell Oil Sport Plus 2 is the first oil specifically developed for light sport 2-stroke (2-cycle) engines such as the ROTAX® air and watercooled series engines. These types of engines commonly encounter intense operating conditions, i.e. full power take off, cruise descent and idle conditions. Varying power outputs and higher operating temperatures demand a specific 2-stroke oil formulation which will also reduce the formation of deposits and protect the 2-stroke engine’s inherent exposure to corrosion and potential ring sticking. AEROSHELL PISTON ENGINE OILS AEROSHELL OIL SPORT PLUS 4 APPLICATIONS AeroShell Oil Sport Plus 4 is intended for use in four-stroke (four-cycle) aircraft piston engines that are of an original automotive design and which cannot, therefore, use traditional Ashless Dispersant aircraft engine oil types. These engines include carburetted, fuel-injected and turbocharged types such as the ROTAX® 912 & 914 series. AeroShell Sport Plus 4 can be used in integrated gearbox and wet clutch systems. AeroShell Oil Sport Plus 4 can be used in engines which operate on both unleaded gasoline and Avgas 100LL. The correct choice of additives and good solvent properties allow the oil to handle lead by-products that can form a semi solid sludge in the oil which can restrict oil passages and compromise lubrication. AeroShell Oil Sport Plus 4 is superior in this respect to those oil types intended for automotive/motorcycle application. Please refer to Operators Handbook/Manual for the correct oil drain interval when operating on different fuels. SPECIFICATIONS No Aviation specifications yet defined. Meets or exceeds the requirements of the highest international specifications: API SL JASO MA Fully approved – all ROTAX® 912 & 914 series engines, ROTAX® Service Instruction SI-912-016/SI-914-019 Selection of suitable operating fluids for ROTAX® engine type 912 & 914 (series). Please consult Operating Handbook/Manual to confirm the correct lubricant specification before use. FEATURES AND BENEFITS ■ First specific oil for Light Sport and Very Light/Ultra light aircraft engines ■ Promotes engine cleanliness ■ Helps keep engines sludge and varnish free ■ Helps reduce oil consumption ■ Helps engines reach TBO (Time Between Overhauls) ■ Protects highly stressed engines parts against scuffing and wear ■ Anti-foaming additives to maximise lubrication effectiveness – especially for those engines operating an integrated gearbox ■ Better cold flow characteristics for easier starts and quicker protection ■ High thermal stability for longer-lasting and safer lubrication ■ Can be used in any climate ■ Advanced anti-rust and anti-wear package DO NOT use AeroShell Oil Sport Plus 4 in engines that are designed to use Ashless Dispersant aviation piston engines oils such as AeroShell W oils. This includes air-cooled Continental Motors and Textron Lycoming engines. Typical Properties Sport Plus 4 SAE viscosity grade Multigrade 10W-40 Density @ 15°C kg/l 0.871 Kinematic viscosity mm²/s @ 100°C @ 40°C 14.46 94.2 Viscosity index 159 Pourpoint °C –33 Flashpoint Cleveland Open Cup °C 228 3.25 PISTON ENGINE OILS PISTON ENGINE OILS 3.24 Developed in conjunction with ROTAX®, AeroShell Oil Sport Plus 4 is the first oil specifically developed for light sport aviation piston engines such as the ROTAX® 912 & 914 series. A combination of low cylinder head temperature (compared with air cooled engines), low oil consumption and the engine internals requires a blend of high quality hydrocarbon base stocks, incorporating synthetic technology, which allows full performance with different fuel types. This oil can be used in all climates. AEROSHELL PISTON ENGINE OILS AEROSHELL OIL DIESEL 10W-40 AeroShell Oil Diesel 10W-40 is a fully synthetic, multigrade engine oil designed for use in the new generation of compression ignition (Diesel) Aviation Piston Engines. The formulation has been selected to be suitable in piston engines fuelled by Jet A or Jet A-1 and is designed for use in the latest highly rated turbocharged diesel engines under all operating conditions. U.S. - British - French - Russian - NATO Code - Joint Service Designation - AeroShell Oil Diesel 10W-40 has been developed to be suitable for use in engines burning Jet fuel and its performance has been optimised to cope with the demands of this type of engine. Its key performance features include the ability to sustain high bearing loads, neutralisation of acid build-up from the sulphur present in the fuel and high dispersancy to allow for the relatively high particle loading produced when burning Jet fuel. ACEA E4, E5 equivalent API CF equivalent Typical Properties Diesel 10W-40 During development, AeroShell Oil Diesel 10W-40 has amassed around 40,000 hours in engine- and flight-testing. It has been used throughout the SMA engine development program and during Thielert engine development testing: it is fully approved by both manufacturers. Further approvals are being sought as other engines are developed for this emerging market. Oil type Fully synthetic hydrocarbon SAE viscosity grade Multigrade10W-40 Density @ 15°C 0.859 APPLICATIONS AeroShell Oil Diesel 10W-40 is a fully synthetic engine oil containing a unique additive package to provide superior piston cleanliness, resulting in a clean, efficient and reliable engine. The package includes a powerful surface active additive that bonds to the surface of highly loaded engine parts, protecting the engine from scuffing damage. AeroShell Oil Diesel 10W-40 MUST NOT be used in spark ignition or Avgas powered aircraft engines. ENGINE MANUFACTURERS’ APPROVALS AeroShell Oil Diesel 10W-40 is approved for use in the following engines. kg/l Base oil viscosity mm²/s @ 100°C @ 40°C 14.6 93.0 Viscosity index Above 160 Pourpoint °C –38 Flashpoint Cleveland Open Cup °C 220 mgKOH/g 16.0 SMA Engines SR 305 (Later models yet to be produced) Total base number Thielert Engines 1.7 and 2.0 Centurion® (Other models yet to be produced) Sulphated ash content %m 1.9 NOTE: At the time of writing, AeroShell Oil Diesel 10W-40 was in the process of being superseded by AeroShell Oil Diesel Ultra. 3.27 PISTON ENGINE OILS PISTON ENGINE OILS 3.26 SPECIFICATIONS No Aviation specifications yet defined. AEROSHELL PISTON ENGINE OILS AEROSHELL OIL DIESEL ULTRA AeroShell Oil Diesel Ultra is a fully synthetic, multigrade engine oil designed for use in the new generation of compression ignition (Diesel) Aviation Piston Engines. The formulation has been selected to be suitable in piston engines fuelled by Jet A or Jet A-1 and is designed for use in the latest highly rated turbocharged diesel engines under all operating conditions. U.S. - British - French - Russian - NATO Code - Joint Service Designation - ACEA Meets the requirements of A3/B4 API Meets the requirements of SL/CF Mercedes Benz MB 299.5 SAE Viscosity grade 5W-30 Typical Properties Diesel Ultra Oil type Fully synthetic hydrocarbon AeroShell Oil Diesel Ultra MUST NOT be used in spark ignition or Avgas powered aircraft engines. SAE viscosity grade Multigrade 5W-30 Density @ 15°C kg/l 0.84 ENGINE MANUFACTURERS’ APPROVALS AeroShell Oil Diesel Ultra is approved to Mercedes Benz Specification 229.5, recognised and required by the leading Diesel aero engine manufacturers AeroShell Oil Diesel Ultra is approved for use in the following engines. Whilst this is correct at the time of writing, testing is ongoing to extend this approval listing as new engines are produced. Kinematic viscosity mm²/s @ 100°C @ 40°C 12.2 68.2 Pourpoint °C –39 Flashpoint Cleveland Open Cup °C 215 Thielert/Centurion® Engines 1.7 & 2.0 Centurion® (Other models yet to be produced) HTHS viscosity @ 150°C Austro Engine AE300 APPLICATIONS AeroShell Oil Diesel Ultra is a fully synthetic engine oil containing a unique additive package to provide superior piston cleanliness, resulting in a clean, efficient and reliable engine. This package includes a powerful surface active additive, which bonds to the surface of highly loaded engine parts, protecting the engine from scuffing damage. This oil has been developed to provide excellent component wear protection and engine cleanliness, based on substantial engine and component endurance tests with all the major diesel aero-engine manufacturers, and flight experience with diesel aero-engines in the field over recent years. AeroShell Oil Diesel Ultra has been developed to be suitable for use in engines burning Jet fuel and its performance has been optimised to cope with the demands of this unique type of engine/fuel combination. Its key performance features include the ability to sustain high bearing loads, neutralisation of acid build up from the sulphur present in the fuel, and high dispersancy to allow for the relatively high particle loading produced when burning Jet fuel. mPaS 3.50 NOTE: At the time of writing, AeroShell Oil Diesel Ultra was in the process of replacing AeroShell Oil Diesel 10W-40. 3.29 PISTON ENGINE OILS PISTON ENGINE OILS 3.28 SPECIFICATIONS No Aviation specifications yet defined. 3 AeroShell W Oils Diesel Ultra 15W-50 Sport Plus 4 W80 & W80 Plus W100 & W100 Plus W120 PISTON ENGINE OILS 3.30 4 5 Kinematic viscosity: mm2/s TYPICAL TEMPERATURE/VISCOSITY CURVES OF AEROSHELL W OILS 7 10 15 25 50 100 200 500 1000 3000 10 20 30 40 50 60 70 80 90 100 110 120 130 0 -10 -20 -30 -40 10000 Temperature: °C

Stroke : 200mm;Max Pressure : 1.0 MPa;Screw Hole Dia : 8.5mm (Approx.). Fluid : Air;Power : Pneumatic;Bore : 32mm. Product Name : Pneumatic Cylinder;Model : SC 32 x 200;Action Type : Double Acting;Rod...

This was never used or put into service, just sitting on a pallet, it is in new condition. Liquidating excess inventory, bargain price.

63mm bore and 200mm stroke, double action and single rod,Adjustable cushion on both cylinder end covers to make sure the cylinder works very smoothly, safely and with low noise. With self-lubricating ...

One, used SMC CXSM25-40 air cylinder. Removable custom assembly on cylinder shaft end. Approximate dimensions 4 3/8” x 3 1/8” x 1 1/8”. Most of the ones available have two 1/4” hose fittings installed...

63mm bore and 300mm stroke, double action and single rod,Adjustable cushion on both cylinder end covers to make sure the cylinder works very smoothly, safely and with low noise. With self-lubricating ...

Really mint condt was from a liquidation of a plastics blow molding plant kept in their spare parts invnitory guess it to be 1.5 in dia and 3 in travel made in japan.

Square Mount HD Magnetic Piston Cylinder. A: Magnetic Piston Cylinder with Sensor Mounting Rail. Stroke Length: 4". Bore Size: 1/2". Compact Air. S: Square, End Mount. PDF Specification Sheet for the ...

YOU ARE PURCHASING A USED AIR CYLINDER. WHEN YOU OPEN UP A DISPUTE CASE IT LIMITS US TO HOW WE. TO RESOLVE YOUR ISSUES. WE DO NOT USE AN AUTOMATED SYSTEM AS SOME DO.

YOU ARE PURCHASING A USED AIR CYLINDER. WHEN YOU OPEN UP A DISPUTE CASE IT LIMITS US TO HOW WE. TO RESOLVE YOUR ISSUES. WE DO NOT USE AN AUTOMATED SYSTEM AS SOME DO.

With self-lubricating bearing, the piston rod is lubrication free. 63mm bore and 150mm stroke, double action and single rod,Adjustable cushion on both cylinder end covers to make sure the cylinder wor...

AIR PISTON. PART NUMBER: 164-741. The color of the photo may vary from the actual product due to translation and reproduction limitations of photography. We apologize in advance for any inconvenience ...

Bore : 32mm, Stroke : 200mm; Port size: PT1/8. Action Type : Double Acting; Rod Type : Single Rod. Cylinder is a mechanical devices which utilize the power of compressed gas to produce a force in a re...

KOGANEI PBDAS 16X15-7. Pneumatic Air Pistons. LOT of 4 No exceptions. Included is only what is mentioned. No other cables, parts, power cords, power adapters, software or any other accessories include...

Model: 03.25 J2MAUS14 4.000. Envelope Pressure: 250 psi Air. Air Cylinder Piston. FROM DROP DOWN BOX AND THEN CHOOSE. Made in the USA.

This was never used or put into service, just sitting on a pallet, it is in new condition. Liquidating excess inventory, bargain price.

Stroke 1". Features Sq. Head, Piston magnet, Adjustable Cushions. Bore Dia. 2-1/2". Piston Material Aluminum. Piston Rod Material Chrome Plated Steel. Item Air Cylinder.

3x PBDAS 16X25-7. KOGANEI Pneumatic Air Pistons. 2x PDAS 16X25-7. LOT of 5 No exceptions. Included is only what is mentioned. No other cables, parts, power cords, power adapters, software or any other...

STROKE : 1.1938 METERS. SHAFT : 21.90MM. NEW, in box. NEW, no box. THESE DO NOT AFFECT USABILITY. USED, but in good working condition. Removed from closed plant. NEW, old inventory.

This page was last updated:  Jan-17 05:30. Number of bids and bid amounts may be slightly out of date. See each listing for international shipping options and costs.

4 Internal Combustion Engines Internal combustion engines are devices that generate work using the products of combustion as the working fluid rather than as a heat transfer medium. To produce work, the combustion is carried out in a manner that produces high-pressure combustion products that can be expanded through a turbine or piston. The engineering of these highpressure systems introduces a number of features that profoundly influence the formation of pollutants. There are three major types of internal combustion engines in use today: (1) the spark ignition engine, which is used primarily in automobiles; (2) the diesel engine, which is used in large vehicles and industrial systems where the improvements in cycle efficiency make it advantageous over the more compact and lighter-weight spark ignition engine; and (3) the gas turbine, which is used in aircraft due to its high power/weight ratio and also is used for stationary power generation. Each of these engines is an important source of atmospheric pollutants. Automobiles are major sources of carbon monoxide, unburned hydrocarbons, and nitrogen oxides. Probably more than any other combustion system, the design of automobile engines has been guided by the requirements to reduce emissions of these pollutants. While substantial progress has been made in emission reduction, automobiles remain important sources of air pollutants. Diesel engines are notorious for the black smoke they emit. Gas turbines emit soot as well. These systems also release unburned hydrocarbons, carbon monoxide, and nitrogen oxides in large quantities. In this chapter we examine the air pollutant emissions from engines. To understand the emissions and the special problems in emission control, it is first necessary that we understand the operating principles of each engine type. We begin our discussion with 226 Sec. 4.1 227 Spark Ignition Engines a system that has been the subject of intense study and controversy-the spark ignition engine. 4.1 SPARK IGNITION ENGINES The operating cycle of a conventional spark ignition engine is illustrated in Figure 4.1. The basic principle of operation is that a piston moves up and down in a cylinder, transmitting its motion through a connecting rod to the crankshaft which drives the vehicle. The most common engine cycle involves four strokes: 1. Intake. The descending piston draws a mixture of fuel and air through the open intake valve. Intake Compression Power Exhaust Piston I j I Piston rod Crank c ---L B =0° (top dead center) B = crank angle B = 180° (bottom dead center) Figure 4.1 Four-stroke spark ignition engine: stroke 1. intake; stroke 2. compression; stroke 3. power; stroke 4, exhaust. Internal Combustion Engines 228 Chap. 4 2. Compression. The intake valve is closed and the rising piston compresses the fuelair mixture. Near the top of the stroke, the spark plug is fired, igniting the mixture. 3. Expansion. The burning mixture expands, driving the piston down and delivering power. 4. Exhaust. The exhaust valve opens and the piston rises, expelling the burned gas from the cylinder. The fuel and air mixture is commonly premixed in a carburetor. Figure 4.2 shows how engine power and fuel consumption depend on equivalence ratio over the range commonly used in internal combustion engines. Ratios below 0.7 and above 1.4 generally are not combustible on the time scales available in reciprocating engines. The maximum power is obtained at a higher ratio than is minimum fuel consumption. As a vehicle accelerates, high power is needed and a richer mixture is required than when cruising at constant speed. We shall return to the question of the equivalence ratio when we consider pollutant formation, since this ratio is one of the key factors governing the type and quantity of pollutants formed in the cylinder. The ignition system is designed to ignite the air-fuel mixture at the optimum instant. Prior to the implementation of emission controls, engine power was the primary concern in ignition timing. As engine speed increases, optimal power output is achieved 0.3 'I-, ~ 0' ~ 0.2 u l.L (f) III 0.1 0.0 '---..L_-L.._L...---L_..l.---l_.....l-_.L--..L---' 0.6 0.8 1.0 1.2 1.4 1.6 ¢ Figure 4.2 Variation of actual and indicated specific fuel consumption with equiv- alence ratio and load. BSFC denotes "brake specific fuel consumption. " Sec. 4.1 229 Spark Ignition Engines by advancing the time of ignition to a point on the compression stroke before the piston reaches the top of its motion where the cylinder volume is smallest. This is because the combustion of the mixture takes a certain amount of time, and optimum power is developed if the completion of the combustion coincides with the piston arriving at socalled top dead center. The spark is automatically advanced as engine speed increascs. Also, a pressure diaphragm senses airflow through the carburetor and advances the spark as airflow increases. Factors other than power output must be taken into account, however, in optimizing the engine operation. If the fuel-air mixture is compressed to an excessive pressure, the mixture temperature can become high enough that the preflame reactions can ignite the charge ahead of the propagating flame front. This is followed by very rapid combustion of the remaining charge and a correspondingly fast pressure increase in the cylinder. The resultant pressure wave reverberates in the cylinder, producing the noise referred to as knock (By et al., 1981). One characteristic of the fuel composition is its tendency to autoignite, expressed in terms of an octane rating. High compression ratios and ignition spark timing that optimize engine power and efficiency lead to high octane requirements. The octane requirement can be reduced by using lower compression ratios and by delaying the spark until after the point for optimum engine performance. Emission controls require additional compromises in engine design and operation, sacrificing some of the potential engine performance to reduce emissions. 4.1 .1 Engine Cycle Operation The piston sweeps through a volume that is called the displacement volume, V". The minimum volume occurs when the piston is in its uppermost position. This volume is called the clearance volume, Ve . The maximum volume is the sum of these two. The ratio of the maximum volume to the clearance volume is called the compression ratio, (4.1 ) The efficiency of the engine is a strong function of the compression ratio. We shall see that R e also has a strong influence on the formation of pollutants. The volume in the cylinder can be expressed as a simple function of the crank angle, (), and the ratio of the length of the piston rod to that of the crank, that is, V = Ve + -Vd 2 ( 1 + -l c cos () - (4.2 ) where l is the piston rod length and c is the length of the crank ann as defined in Figure = 0°, commonly referred to as top dead center, TOC. The maximum volume occurs at bottom dead center, BOC, () = 180 0. These positions are illustrated in Figure 4.1. Engine speeds range from several hundred revolutions per minute (rpm) for large 4.1. The minimum volume occurs at () Internal Combustion Engines 230 Chap. 4 industrial engines to 10,000 rpm or more for high-perfonnanee engines. Most automobiles operate with engine speeds in the vieinity of 3000 rpm. At this speed, each stroke in the cycle takes place in 20 ms. As an automobile is driven, the equivalence ratio and intake pressure vary with the engine load. Such changes in engine operation, however, are slow by comparison with the individual strokes. In discussing engine operation, we can assume that in anyone cycle the engine operates at constant speed, load, and equivalence ratio. We begin with a discussion of the thennodynamics of the spark ignition engine cycle and develop a model that has been used extensively in optimizing engine operation to minimize emissions and to maximize performance. The spark ignition engine is one of the few combustion systems that burns premixed fuel and air. Fuel is atomized into the air as it flows through a carburetor and vaporizes before it enters the cylinder. Even though the fuel and air are premixed prior to combustion, the gas in the cylinder becomes segmented into burned and unburned portions once ignition occurs. A flame front propagates through the cylinder as illustrated in Figure 4.3. The fuel-air mixture ahead of the flame is heated somewhat by adiabatic compression as the burning gas expands. Not only are the burned and unburned gases at widely different temperatures, but also there are large variations in the properties of the burned gases. These variations must be taken into account to predict accurately the fornlation and destruction of NO, and CO in the engine. Another important feature that distinguishes reciprocating engines from the systems discussed thus far is that the volume in which the combustion proceeds is tightly constrained. While the individual elements of fluid do expand as they burn, this expansion requires that other elements of fluid, both burned and unburned, be compressed. As a result, the burning element of fluid does work on the other fluid in the cylinder, oW = p dV, increasing its internal energy and therefore its temperature. Whilc the engine strokes are brief, the time is stilJ long by comparison with that required for pressure equilibration. For an ideal gas, the propagation rate for small pressure disturbances is the speed of sound, a, = .JyRT/M (4.3 ) gas Figure 4.3 Flame propagation in the cylinder. Sec. 4.1 Spark Ignition Engines where 'Y is the ratio of specific heats, 231 cilcu ' and M is the molecular weight of the gas; as is of the order of 500 to 1000 m s- for typical temperatures in internal combustion engines. For a cylinder 10 cm in diameter, the time required for a pressure disturbance to propagate across the cylinder is on the order of 0.2 ms, considerably shorter than the time required for the stroke. Thus, to a first approximation, we may assume that the pressure is uniform throughout the cylinder at any instant of time, at least during norn1al operation. 4.1.2 Cycle Analysis The essential features of internal combustion engine operation can be seen with a "zerodimensional" thermodynamic model (Lavoie et aI., 1970; Blumberg and Kummer, 1971). This model describes the thermodynamic states of the burned and unburned gases as a function of time, but does not attempt to describe the complex flow field within the cylinder. We consider a control volume enclosing all the gases in !he cylinder. Mass may enter the control volume through the intake valve at flow rate, ];. Similarly, mass may leave through the exhaust valve and possibly through leaks at a flow rate];,. The first law of thermodynamics (2.8) for this control volume may be written in the general form dU d1 -- - - = ];h i - ];.h" + dQ dW d1 - dt where U is the total internal energy of the gases contained in the cylinder and h; and he are the mass specific enthalpies of the incoming and exiting flows, respectively. Q denotes the heat transferred to the gases. The work done by the gases, W, is that of a pressure acting through a change in the volume of the control volume as the piston moves. If we limit our attention to the time between closing the intake valve and opening the ex~aus~ valve and assume that no leaks occur, no mass enters or leaves the cylinder (i.e.,]; = Ie = 0). The energy equation then simplifies to d _ dt (muT) dQ = d1 - dV P dt where UT is the total mass specific internal energy (including energies of formation of all species in the cylinder), - Q is heat transferred out of the charge, and m is the total mass of the charge. The only work done by the gases is due to expansion against the piston, so the work is expressed as p dV I dt. If we further limit our attention to constant engine speed, the time derivations may be expressed as d d - = wdt de where w is the engine rotation speed (crank angle degrees per s). Thus we have d de (muT) dQ = de - dV p de (4.4 ) Internal Combustion Engines 232 Chap. 4 The total specific internal energy of the gas includes contributions of burned and unburned gases, with a mass fraction (X of burned gas, (4.5 ) where < ) denotes an average over the entire mass of burned or unburned gas in the cylinder. The unburned gas is quite uniform in temperature (i.e., 0, NO tends to form; when (3 > 1 and dYNo/ dO < 0, NO tends to decompose. Equation (4.36) is integrated at each point a' in the charge from the crank angle at which that element initially bums to a crank angle at which the reaction rates are negligible. At this point the quenched value of the NO mole fraction *We use ~ here as this traction to avoid contusion with the traclion burned a. Sec. 4.1 241 Spark Ignition Engines YNO" is achieved. The overall mole fraction of NO in the entire charge is given by )lNO = i~ YNOJa') da' (4.37) Nitric oxide concentrations versus crank angle, computed by Blumberg and Kummer (1971), are shown in Figure 4.6. Both rate calculated and equilibrium NO are shown at three positions in the charge, a' = 0, 0.5, 1.0. The major contribution to the total NO fomled results from the elements that bum first. They experience the highest temperatures and have the longest time in which to react. Considerable decomposition of NO occurs in the first element because of the high temperatures. However, as the first element cools during expansion, the rate of NO decomposition rapidly decreases, so that after about 40 crank angle degrees, the NO kinetics are effectively frozen. We can now summarize the processes responsible for the production of nitric oxide 10,000 First element \/" I 8000 I 6000 E / n. n. o z 4000 I I I I I I I I I _ ... Equivalence ratio = 0.95 Inlet temp = 338 K Inlet pressure = 66.6 kPa RPM = 1200 68c = 10° BTDC to 30° ATDC - - Rate calculated -- - Equilibrium Overall NO ", ~~ ,,~ ~~ Last " ~~element ! "'~~ 2000 , ' "..... ~ -...::.::: Middle element ~Last element 60 70 80 Figure 4.6 Nitric oxide concentration in the burned gas as a function of crank angle for the first, middle, and last element to bum for 1> = 0.97 (Blumberg and Kummer, 1971). Reprinted by permission of Gordon and Breach Science Publishers. 242 Internal Combustion Engines Chap. 4 in the internal combustion engine. During the flame propagation, NO is formed by chemical reactions in the hot just-burned gases. As the piston recedes, the temperatures of the different burned elements drop sharply, "freezing" the NO (i.e., the chemical reactions that would remove the NO become much slower) at the levels formed during combustion, levels well above these corresponding to equilibrium at exhaust temperatures. As the valve opens on the exhaust stroke, the bulk gases containing the NO exit. It is to the processes that occur prior to the freezing of the NO levels that we must devote our attention if we wish to reduce NO formation in the cylinder. 4.1 .6 Carbon Monoxide The compression due to piston motion and combustion in a confined volume leads to very high burned gas temperatures in reciprocating engines. Peak temperatures may range from 2400 to 2800 K, with pressures of 15 to 40 atm. In Chapter 3 we saw that the CH-O system equilibrates rapidly at such high temperatures. It is therefore reasonable to assume that CO is equilibrated immediately following combustion. The equilibrium CO mole fraction at these peak temperatures is very high, greater than 1 %. Work done by the gas in the cylinder on the piston during the expansion stroke cools the combustion products. When the exhaust valve first opens, the pressure in the cylinder is much larger than that in the exhaust manifold. As the gas is forced out through the valve, work is done by the gas remaining in the cylinder, so the temperature drops even more rapidly. Ultimately, this cooling of the combustion products exceeds the ability of the three-body and CO oxidation reactions to maintain equilibrium. The combustion products are rapidly cooled during the expansion stroke and the exhaust process, causing the CO oxidation kinetics to be quenched while the CO level is still relatively high. In Chapter 3 it was shown that CO oxidation proceeds primarily by reaction with OH, CO + OH CO 2 + H ~ E and that the OH can be present at concentrations significantly greater than that at equilibrium in rapidly cooled combustion products. The concentrations of OH and other radicals can be described using the partial-equilibrium model developed in Chapter 3, wherein it was shown that the rate of CO oxidation is directly coupled to the rates of the three-body recombination reactions, primarily, H + O2 + M E ~ H02 + M in fuel-lean combustion. CO levels in spark ignition engines are generally high enough that the influence of the CO oxidation on the major species concentrations cannot be ignored. The direct minimization of the Gibbs free energy is better suited to incorporating this detail than is the equilibrium-constant approach developed in Chapter 3. Heywood (1975) used the rate-constrained, partial-equilibrium model (based on direct minimization of the Gibbs free energy) to study CO behavior in spark ignition engines. His calculations confinn that at the peak temperatures and pressures the equilibration of CO is fast compared to the changes due to compression or expansion, so Sec. 4.1 Spark Ignition Engines 243 equilibrium may reasonably be assumed immediately following combustion. The burned gases are not uniform in temperature, however; so the equilibrium CO level depends on when the clement burned. Furthern10re, the blowdown of the cylinder pressure to the exhaust manifold pressure in the initial phase of the exhaust process lasts about 90 crank angle degrees. Thus the temperature-time profiles of fluid elements within the charge differ depending on the time of burning and on when they pass from the cylinder through the valve into the exhaust manifold. These effects are illustrated by the results of an idealized calculation shown in Figure 4.7. CO mole fractions for individual fluid elements in the burned gas mixture are shown as a function of crank angle. The elements are identified in terms of the fraction of the total charge burned when the element burned, Ct, and the mass fraction that has left the cylinder when the element leaves the cylinder, z. The partial-equilibrium calculations are close to equilibrium until about 50 crank angle degrees after top dead center, when the rapid cooling due to adiabatic expansion leads to partial quenching of the CO oxidation. - - - z =0.01 _ . - . - z = 0.50 - - - z=0.99 -------- Equilibrium CO ep = 1.0 Rc = 7.1 , '\ \ c o N = 3000 rpm \ \ \ += u \ ~ \ /\ '+Q) o E COe " o u ,, , '\,, ,' \ ,, " ' ',z = 0.50,\ \' ,, , '. 'I' Exhaust I ~I valve 'z=0.99 opens z=0.01 TDC 30 o 60 90 120 150 10 5 t (ms) Figure 4.7 Carbon monoxide concentration in two elements in the charge that bum at different times, during expansion and exhaust processes. 0' is the mass fraction burned and z is the fraction of the gas that has left the cylinder during the exhaust process (Heywood, 1975). Reprinted by permission of The Combustion Institute. 244 Internal Combustion Engines Chap. 4 The CO levels measured in fuel-lean combustion are substantially higher than those predicted with the partial-equilibrium model, but agreement is good near stoichiometric (Heywood, 1976). In fuel-rich combustion, the CO levels in the exhaust gases are close to the equilibrium concentrations, as predicted by the partial-equilibrium model. The reasons for the high levels in fuel-lean combustion are not fully understood, but may be coupled to the oxidation of unburned hydrocarbons in the exhaust manifold. 4.1.7 Unburned Hydrocarbons The range of equivalence ratios over which spark ignition engines operate is narrow, typically 0.7 < cf> < 1.3, the fuel and air are premixed, and the flame temperatures are high. These conditions, in steady-flow combustion systems, generally would lead to very low emissions of unburned hydrocarbons. Why, then, are relatively large quantities of hydrocarbon gases present in the combustion products of automobile engines? This question has been the subject of numerous investigations in which hypotheses have been developed and supported with theory and experiment, only to be later challenged with new interpretations that contradict earlier models. In an early investigation of this problem, Daniel and Wentworth (1962) magnified photographs of the flame spread in the cylinder of a spark ignition engine. It was observed that the flame failed to propagate through the mixture located within 0.1 to 0.7 mm of the cylinder wall. They hypothesized that this wall quenching allowed hydrocarbons to escape combustion in spark ignition engines. Figure 4.8 shows the nature of these wall quench regions. In addition to the quench layers at the cylinder walls, the small volume between the piston and cylinder wall above the top piston ring, called the crevice volume, contains unburned hydrocarbons. Experiments were performed in which the quench zone of an operating engine was sampled. It was found that the proportion of the quench zone exhausted is less than that of the total gas exhausted. This observation was attributed to trapping in the boundary layer. Quench layer Figure 4.8 Schematic showing the quench layer and crevice volume where heat transfer to the walls may quench the combustion (Tabaczynski et a!., 1972; © SAE, Inc.). Sec. 4.1 Spark Ignition Engines 245 A fraction of the gas remains in the cylinder at the end of the exhaust stroke. Although this residual gas amounts to a small fraction of the total gas in the cylinder in a normally operating engine, the residual gas hydrocarbon concentration tends to be very high. The recycled hydrocarbons may be a significant fraction of the hydrocarbons left unburned in the cylinder. The trapping effect can be explained as follows. Gases adjacent to the wall opposite the exhaust valve are the farthest from the exit and least likely to be exhausted. Gases along the walls near the exhaust valve have a better chance to be exhaustcd, but viscous drag slows their movement. Some quenched gases do escape, but on the whole the more completely burned gases at the center of the chamber are preferentially exhausted first, with the result that the residual gas has a higher concentration of hydrocarbons than the exhaust gas. It is likely that the quench zone hydrocarbons that remain in the cylinder are burned in the succeeding cycle. In the experiments reported by Daniel and Wentworth (1962), about one-third of the total hydrocarbons were recycled and probably burned in succeeding cycles. Figure 4.9 shows the measured variation in the exhaust hydrocarbon concentration and mass flow rate with crank angle. As the exhaust valve opens and the emptying of the combustion chamber starts, the hydrocarbon concentration in the exhaust manifold increases rapidly to a peak of 600 ppm. The hydrocarbon concentration then drops and remains at 100 to 300 ppm for much of the exhaust stroke. Late in the exhaust stroke, the hydrocarbon level again rises sharply. The hydrocarbon mass flow rate shows two distinct peaks corresponding to these concentration maxima. The early peak in the hydrocarbon concentration was attributed to the entrainment of the quench layer gases near the exhaust valve immediately after it opens. The low hydrocarbon concentration during the middle portion of the exhaust stroke is most probably due to the release of burned gases from the center of the cylinder. Tabaczynski et al. (1972) further observed that, during the expansion stroke, the gases in the crevice volumes are laid along the cylinder wall. As the piston moves up during the exhaust stroke, the layer is scraped off the wall and rolled up into a vortex, as depicted in Figure 4.10. The second peak in the hydrocarbon concentration was attributed to the passage of this vortex through the exhaust valve late in the exhaust stroke. Although the quench layer model does appear to explain many of the observations of hydrocarbons in spark ignition engines, recent studies have questioned the importance of quench layers as sources of unburned hydrocarbons (Lavoie et aI., 1980). The cooling effect of the wall does, indeed, prevent the flame from propagating all the way to the cylinder wall. Hydrocarbon vapors can diffuse from this cool region, however, into the hotter gases farther from the wall. If this occurs early in the cycle when the temperature of the burned gases is high, the hydrocarbons from the quench layer will be burned. We can gain some insight into the quench-layer problem by examining the time scales of diffusion and reaction of the hydrocarbon gases. The characteristic time for diffusion of gases from the quench layer into the bulk gases is Tf) ~ [} / D. Adamczyk and Lavoie (1978) report values of 0 of order 50 to 75 J-tm and diffusion times ranging from 0.1 to 0.3 ms at atmospheric pressure. Inasmuch as this time is short compared to that of the expansion stroke and typical combustion times, a considerable amount of the 246 Internal Combustion Engines Chap. 4 50 I Vl .3' 40 Q) "§ 3 0 <;: 30 Vl Vl 0 E 20 Vl 0 Ol ~ Vl J 0 10 .<:: '" W ,/'"'11 I Q) ~ C 0 '" Q) .<:: I I , ,, 600 I c I E Cl. Cl. 400 () I >- Exhaust valve opens , ~ , p' , I 200 140 180 ,a-"O..-a,..d n,...c:! 220 260 I 0.6 \ , \ 0.5 ,, \ 0.3 ~ 0 3 .2 ..... Vl Vl 0 E \ 0.2 It \ 300 Vl 0"> ~ 0.4 ¢ \ \ \ I \----- \ I 0 100 0.7 \ 800 \ 340 0 I 0.1 0 Crank angle (deg) Figure 4.9 Measured instantaneous mass flow rate exhaust hydrocarbon concentration, and hydrocarbon mass flow rate out of the exhaust valve (Tabaczynski et al., 1972; © SAE, Inc.). quench layer hydrocarbons may be expected to diffuse away from the walls and bum in the cylinder. Some quench-layer hydrocarbons may survive because the thermal boundary layer spreads at a rate comparable to that of the hydrocarbons, preventing the hydrocarbons from reaching high temperatures at which they would rapidly oxidize. The quantities of hydrocarbons that survive by this route, however, are much too small to explain the observed hydrocarbon levels. In one study in which the quench-layer gases were sampled directly, it was estimated that the quench-layer gases could account for not more than 3 to 12 % of the hydrocarbons measured in the exhaust (LoRusso et al., 1983). Hydrocarbons contained in the crevice volume between the piston, piston ring, and cylinder wall account for much of the hydrocarbon release. These vapors expand out from the crevices late in the expansion stroke, so lower temperatures are encountered Sec. 4.1 247 Spark Ignition Engines (a) (b) (e) Figure 4.10 Schematic illustrating the quench layer model for hydrocarbon cmissions. (a) Quench layers are formed as heat transfer extinguishes the flame at the cool walls and in the crevice volume. (b) Gas in the crevice volume expands and is spread along the cylinder wall as the pressure falls. When the exhaust valve opens, the quench layers near the valve exit the cylinder. (c) The hydrocarbon-rich cylinder wall boundary layer rolls up into a vortex as the piston moves up the cylinder during the exhaust stroke (Tabaczynski et aI., 1972; © SAE, Inc.). by crevice gases than by the quench-layer gases (Wentworth, 1971). Adamczyk et al. (1983) examined the retention of hydrocarbons in a combustion bomb that consisted of a fixed piston in an engine cylinder. About 80% of the hydrocarbons remaining after combustion were attributed to the piston crevice, with most of the remaining hydrocarbons surviving in smaller crevices associated with the head gasket and with the threads on the spark plug. The crevice volumes contribute primarily to the peak in the hydrocarbon flux late in the exhaust process, since those gases originate far from the exhaust valve. Other sources must therefore contribute significantly to the hydrocarbon emissions, particularly those that exit the cylinder early in the exhaust process. Haskell and Legate (1972) and Wentworth (1968) suggested that lubicating oil layers on the cylinder walls may adsorb or dissolve hydrocarbon vapors during the compression stroke. These stored hydrocarbons are protected from the flame. As the pressure in the cylinder drops during the expansion stroke and exhaust process, these hydrocarbons desorb into the combustion products. Kaiser et al. (1982) showed that fuel vapors and fuel hydrocarbon oxidation product emissions increase as the amount of oil in the cylinder increases. Carrier et al. (1981) developed a model for cyclic hydrocarbon adsorption and desorption in a liquid film, taking into account thermodynamic equilibrium at the gas-liquid interface and diffusional resistance within the liquid layer. The results from this model are qualitatively consistent with the observed reduction of hydrocarbon emission with engine speed. 248 Internal Combustion Engines Chap. 4 4.1.8 Combustion-Based Emission Controls The equivalence ratio has a strong influence on the formation of nitrogen oxides and on the oxidation of carbon monoxide and unburned hydrocarbons, but the extent to which these emissions can be controlled through fuel-air ratio adjustment alone is limited. Other combustion parameters that can influence emissions include the ignition timing and design parameters. The compression ratio determines the peak pressure and hence the peak temperature in the cycle. The piston and cylinder head shapes and the valve geometry influence the turbulence level in the engine and therefore the rate of heat release during combustion. Temperatures can also be reduced through dilution of the incoming air with exhaust gases. Design and operating variables not only influence the levels of pollutant emissions, but also directly affect the engine power output and efficiency. As we examine various emission control strategies, we must also examine their effects on engine performance. The efficiency of an internal combustion engine is generally reported in terms of the specific fuel consumption (SFC), the mass of fuel consumed per unit of energy output, kg MJ - I or g kW-h - I. The work output per engine cycle is presented in terms of the mean effective pressure (MEP), the work done per displacement volume. If the MEP is determined in terms of the net power output, P, the quantity is called the brake mean effective pressure (BMEP) and is calculated as BMEP = 2~ ~tH (4.38) where n is the engine rotation speed (revolutions per second). Many factors not directly involved in the combustion process influence the BMEP: friction; pumping work associated with the intake and exhaust flows; and work used to drive engine equipment such as generators, water pumps, fans, and so on. The work performed by the gas during the compression and expansion strokes, 3600 Wi i 0° dV p-dO dO (4.39) that is, that that would be indicated by a pressure measurement, is of more concern to us here. The mean effective pressure based on this work, (4.40 ) is called the indicated mean effective pressure (IMEP). It is also convenient to present the specific fuel consumption in terms of the indicated work to eliminate the influences of parasitic losses and loads. This quantity is then called the indicated specific fuel consumption (ISFC). Figure 4.11 shows the influence of engine operating equivalence ratio on the indicated specific NO x emissions (g NOt Mr l ) and fuel consumption for three different values of the combustion duration. NOt emissions are maximum at ¢ = 1 and decrease Sec. 4.1 Spark Ignition Engines 249 1.0 0.9 0.8 0.7 ..., I ~ ~ 0 z 0.6 0.5 0.4 (j) 03 0.2 0.1 75 68 c = 60 0 'I..., 500 ~ 0' 70 0 I.L (j) 65 60L~~~:::=-_L--_--L-_----L._~ 0.6 0.7 0.8 0.9 1.0 11 1.2

= I. The combustion duration influences NO< emissions more strongly than fuel consumption, but even there the effect is small. The influence of the operating equivalence ratio on emissions of carbon monoxide and unburned hydrocarbons is illustrated in Figure 4.12. The CO level is relatively low for fuel-lean operation but rises abruptly, as expected, when the mixture becomes fuelrich. The hydrocarbon emissions, on the other hand, exhibit a minimum and increase for very fuel-lean operation. In lean operation the temperature can be too low for hydrocarbons to bum late in the expansion stroke. Furthermore, the low laminar flame speed at low c/> means that the flame may not even reach all the mixture. 250 Internal Combustion Engines Chap. 4 4~------,.----,-----,-----,----...., 3 o<) 2 >, 350 300 '" u E 00- 250 200 <) I >, 150 100 50 0 0.7 0.8 0.9 1.0 1.1 1.2 Figure 4.12 Influence of equivalence ratio and load on carbon monoxide and hydrocarbon emissions. Solid lines: 2000 rpm, (Ji = - 38 0, and 80 km h- I road load; Dashed lines: 1200 rpm, (Ji = -J0 0 , and 48 km h - l road load. To reduce NO, emissions significantly, it is necessary to reduce the peak temperature significantly. Delaying the initiation of combustion results in the peak pressure occurring later in the expansion stroke, as illustrated in Figure 4.13. The spark is usually fired before top dead center, so that the combustion rate is maximum near top dead center. Delaying the spark results in the energy release occurring when the cylinder volume has increased significantly. The peak pressure and temperature are therefore reduced by this spark retard. At the most extreme level, the spark can be retarded past top dead center so that the gases begin to expand before combustion begins. The influence of equivalence ratio and ignition angle on fuel consumption and NO, emissions has been calculated by Blumberg and Kummer (1971). Their results are shown in a map of BSFC versus BSNO in Figure 4.14. Clearly, if an engine could be operated at very low equivalence ratios, NOt emissions could be reduced dramatically with only a minimal efficiency penalty. Operating at equivalence ratios more typical of premixed combustion 3000r------,-----,-------r----r-.----,---~--~--~-~ 2500 Ii' 2000 -'" ~ ~ Cf) ~ 1500 0- Bj ~ = 10° ----I-JI--~ Q) -0 c >- <-> 1000 500 TDC -180 -140 -100 -60 -20 20 60 100 140 180 B Figure 4.13 Influence of ignition timing on cylinder pressure profiles. ep-'=1.81 ep-' = 1.51 ep-l = 1.41 ep-'=1.31 ep-'=1.18 I nlet temperature = 339 K Inlet pressure = 66.6 kPa RPM = 1200 LlR = 4.0 Compression ratio = 8.5 '"o x 12 I J ~ <-> LL 20° to 60° ATDC 10 (f) 50 0 ATDC co 8 10° BTDC to 30° ATDC 2.0 4.0 6.0 BS NO (g NO Mr') Figure 4.14 Effect of equivalence ratio and ignition timing on efficiency and NO for~ mation for flOc = 40° (Blumberg and Kummer, 1971). Reprinted by permission of Gor~ don and Breach Science Publishers. 251 Internal Combustion Engines 252 Chap. 4 in spark ignition engines and relying on ignition retard to control NO, yield smaller emission benefits and substantially larger fuel consumption increases. Such emissions! performance trade-offs are typical of efforts to control engine emissions and have been the motivating factor behind much of the research into engine emission control technologies. Reducing the compression ratio can also lower peak temperatures, thereby limiting NO, formation. However, the NO, emission reductions achieved by reducing the compression ratio are small compared to those accrued by retarding the spark. Another way to reduce the peak temperatures is by diluting the charge with cool combustion products. In engines, this process is called exhaust gas recirculation (EGR). The use of combustion products for dilution instead of excess air has dual benefits: 1. Dilution of the fuel-air mixture without the addition of excess O 2 that aids in NO, formation. 2. An increase in the specific heat of the gas due to the presence of H2 0 and CO 2 , This reduces the temperature somewhat more than would equivalent dilution with excess air. Figure 4.15 shows how significantly EGR can reduce NOr emission levels. For small amounts of EGR, the theoretical predictions agree closely with experimental ob10 4 ,----..,---,-----,---,---,-----r----, ,.- . , 10% EGR • Data '. ~'.\' Theory 0% EGR • ',:- Data Theory . '&~ , 'A,~ \'\. " ,. o '. " on :::l o ~1O w 28% EGR o Data Theory - - without flame NO - - - with flame NO ""'.7 10 '::;.7,----::0~.8::c---0 g --:-1.'=0--::,-'-:.,,----,,-';.2::----;-~-:-".4 Figure 4.15 Influence of exhaust gas recirculation on NO emissions as a function of equivalence ratio (Heywood. 1975). Re- printed by pennission of The Combustion Institute. 253 Spark Ignition Engines Sec. 4.1 servations; however, at 28 % EGR, the measured NO, emission levels for lean or rich mixtures are significantly higher than those predicted considering only postflame chemistry. The dashed curve presents more detailed chemical mechanism calculations that take into account the nonequilibrium radical concentrations that are present within the flame front (i.e., "prompt NO"). Agreement on the fuel-lean side is very good. On the other hand, even when the flame chemistry of the 0, H, and OH radicals is taken into account, the predictions of NO, fOffi1ation in fuel-rich combustion are significantly lower than those observed. This discrepancy may be due to nitrogen chemistry not included in the model, particularly the reactions of N2 with hydrocarbon radicals. From these results we see the EGR can substantially reduce NO, formation in spark ignition engines, but the degree of control achievable by this method is limited. These gains are not achieved without penalties. Figure 4.16 shows calculations of the variation of fuel consumption and mean effective pressure with equivalence ratio and amount of exhaust gas recirculated. While the fuel consumption penalty is relatively small, the loss of power is significant, so the engine size must be increased to meet a particular power requirement if EGR is employed to control NO, emissions. It is apparent that spark retard and exhaust gas recirculation are effective measures for NO, emission control. The equivalence ratio range that can be employed effectively is limited. Rich mixtures lead to high CO levels. As the mixture becomes too fuel-lean, hydrocarbon emissions rise. Hence control of emissions without the use of exhaust gas cleaning involves compromises. Spark retard and exhaust gas recirculation are usually used in combination to achieve low NO, emission levels. The introduction of strict NO, emission controls in combination with limits on CO and hydrocarbon emissions was accompanied by a substantial increase in fuel consumption of automobiles in the United 550 EGR=O% , "" "0 0..>< t:.e ", ,~, ",' 0.- w Inlet temp = 339 K Inlet pressure = 66.6 kPa rpm = 1500 Rc = 8.5, L/R = 4.0 = 100BTDC to 40 0ATDC :'2' co 350 EGR=20% ,~ ,~, 1.18 1.14 ill 1.095 ----BSFC 1.05 0 x I -, 2 en 0.963 EGR 20 %- . . . __ _ _= _ _- "'iii;~ 250 1.1 --BMEP 1.01 '~~ 1.2 1. 23 I '""',", 1.3 1.27 1.0 ...-- .".. -- ...-~~::io% ----~-- ------EGR=O% 0.9 0.8 0.7 0.6 0.92 0.876 0.83 0.5 ep Figure 4.16 Effect of equivalence ratio and exhaust gas recirculation on power (brake mean effective pressure) and fuel consumption (Blumberg and Kummcr. 1971). Rcprintcd by pcnnission of Gordon and Breach Science Publishers. .>< u u.. co (fJ Internal Combustion Engines 254 Chap. 4 States. Ultimately, exhaust gas treatment was required to achieve acceptable emissions and performance simultaneously. Exhaust gas treatment is discussed in a subsequent section. 4.1.9 Mixture Preparation The spark ignition engine bums premixed fuel and air. In conventional engines, this mixture is prepared in the carburetor, a complex device that controls both fuel and air flows to the engine. The mixture requirements depend on engine speed and load. A richer mixture is required at high load (such as during vehicle acceleration) than at low load. Even though combustion will be incomplete, fuel-rich mixtures have been used to increase the heat release per cycle, thereby increasing the power delivered by the engine. Carburetors have evolved as mechanically activated control systems that meet these requirements. As we have seen in the preceding discussion, emission controls place additional constraints on engine operation that are not readily met with purely mechanical control. To understand the need for and the nature of the new systems for mixture preparation that are being developed as part of integrated emission control systems, it is useful to examine the operation of a conventional carburetor. The power output and speed of a spark ignition engine are regulated by a throttle that limits the airflow into the engine. In conventional engines, the airflow rate is used to control the fuel/air ratio. Part of the difficulty encountered in early attempts to reduce automobile emissions derived from the complex coupling of fuel and airflow rates. A simple carburetor is illustrated in Figure 4.17. The throttle is a butterfly valve, a disk that is rotated to obstruct the airflow, producing a vacuum in the intake manifold. The low pressure reduces the mass entering the cylinders, even though the intake gas volume is fixed. The rate at which fuel is atomized into the airflow is controlled by the pressure drop in a venturi, /:!p, that is, (4.41 ) Gf = CIF .J2Pf /:!Pf where G/ is the fuel mass flux, CIF the flow coefficient associated with the fuel metering orifice, Pf the density, and /:!p/ the pressure drop across the fuel metering orifice. This pressure drop corresponds to the difference between the pressure drop created by the airflow through the venturi /:!Pa and the pressure needed to overcome surface tension at the nozzle exit, /:!Pu = 2a / d, where a is the surface tension and d is the nozzle diameter. The total pressure drop becomes /:!p/ "'" Po + p/gh - PI' - 2 a d (4.42 ) where PI' is the gas pressure in the venturi. The airflows in the intake system involve large pressure drops, so the compressibility of the gas must be taken into account. The pressure drop associated with the gas flow drives the fuel flow, so we need to know the relationship between pressure drop and flow rate. By considering the conservation of energy, we can readily derive such an expression for the adiabatic and thennodynamically reversible (i.e., isentropic) flow of an ideal gas. Sec. 4.1 255 Spark Ignition Engines Main jet Idle vent line Idle air tube 'h'-'-~V£ZLn.--ldle adjusting screw Idle passage Idle well Idle metering orifice Idle jet Throttle plate Figure 4.17 Schematic of a simple carburetor. The flows through real devices such as the venturi or throttle are not perfectly reversible, so the flow rate associated with a given pressure drop is lower than that for isentropic flow. The ratio of the actual flow rate to the ideal flow rate is the flow coefficient for the device, that is, G Cf = (4.43 ) Gs where G denotes the mass flux and the subscript s denotes that for isentropic flow. The flow coefficient for a sharp-edged orifice is 0.61. The venturi is designed to achieve nearly reversible flow so that Cf will be closer to unity. The flow coefficient for the throttle changes as the throttle plate is rotated. It is unity when the throttle is fully open and decreases toward that for the orifice as the throttle is closed. We consider adiabatic flow through the device in question. As the gas is accelerated, its kinetic energy must be taken into account in the fluid energy balance, that is, for the flow at velocities VI and V2' -h 1 2 -h 1 2 -h I + "2 V j = 2 ho is the stagnation enthalpy corresponding to the specific heats are constant, we may write + "2 V2 = V = 0 O. Assuming that the gas is ideal and Internal Combustion Engines 256 Chap. 4 (4.44 ) The mass flux is G = p1vl> so we may write (4.45 ) If the flow is adiabatic and isentropic, the density and temperature are related to the pressure by E.p'I = Po p1, P (4.46 ) Po T'Ih- 1 = T1,h- (4.47) 1 Using the ideal gas relation and these results, the mass flux thus becomes G = P s 0 ~ RT M r1h o 1_2_ (1 - ~ 'Y - 1 r('I-llh) (4.48 ) where r = P /Po is the pressure ratio. At sufficiently low pressure ratio, the velocity at the minimum cross-sectional area will equal the local speed of sound (4.3). Further reduction in the pressure below the throat has no influence on the mass flow rate, so the flow is said to be choked. Substituting (4.3) into (4.44), we find To T* + 1 'Y (4.49 ) 2 where the asterisk is used to denote a property evaluated at locally sonic conditions. Using (4.47) we find the critical pressure ratio, r* = (_2_)'I/('I-ll 'Y + (4.50) I The corresponding mass flow rate is obtained by substituting r* into (4.48), * _ G, (r ) - Po ~ M (_2_)('1+ )/2('11 1 ) (4.51 ) RT 'Y + 1 o The mass flow rate for a real device becomes r > r* (4.52 ) r ::c; r* Sec. 4.1 257 Spark Ignition Engines For a well-designed venturi, the flow coefficient will be nearly unity and the stagnation pressure downstream of the venturi will be close to that at the venturi inlet. Butterfly valves and other nonideal flow devices will have lower flow coefficients. If a subsonic flow separates at the minimum area, the pressure at that point will correspond approximately to the downstream stagnation pressure. Thus, closing the throttle results in the pressure in the intake manifold being substantially below atmospheric pressure. The fuel flow rate is governed by the pressure at the throat of the venturi, so (4.41) can be expressed in tenns of the pressure ratio (4.53 ) > The fuel/air ratio becomes (for r r *) 2a 1 r) + gz - d A C P I' jF 0 ~ RT M o r1h ~_2 __ (I 'Y - I (4.54 ) _ rb-1lh) The complex dependence of the equivalence ratio on the pressure ratio is readily apparent. Examining (4.42) we see that, for 2a gz r2::1--+-pod Po (4.55) the pressure drop in the venturi is insufficient to overcome surface tension and atomize the fuel. These high pressure ratios (low pressure drops) correspond to low engine speeds. A separate idle nozzle supplies the fuel necessary for low-speed operation. This ideal adjustment is coupled to the pressure drop at the throttle valve. Figure 4.18 illustrates the variation of equivalence ratio with airflow that is produced by these metering devices. The pressure in the venturi throat decreases with increasing airflow. Since the difference between this pressure and that of the atmosphere provides the driving force for the main fuel flow, the fuel supplied by the main jet increases with increasing airflow. The idle jet compensates for the precipitous drop in the fuel flow supplied by the main jet. The pressure at the throttle plate provides the driving force for the idle fuel flow, so this flow is significant only when the idlc plate is closed, i.e., at low airflow. As the throttle plate is opened and the airflow increases, the idle fuel flow decreases markedly. The operating equivalence ratio of the engine is detennined by the sum of the two fuel flows, shown by the upper curve. At high engine load, a richer mixture may be required than is supplied by this simple metering system. The power jet shown in Figure 4.19 is one method used to supply the additional fuel. Ideally, the throttle position at which the power jet opens would vary with engine speed. A mechanical linkage that opens gradually as the throttle 258 Internal Combustion Engines Chap. 4 1.6 o 0.2 0.4 0.6 0.8 1.0 1.2 Figure 4.18 Variation of equivalence ratio with airflow rate for a simple carburetor (Taylor. 1966). Reprinted by pem1ission of MIT Press. opens beyond some point is a compromise solution. When the power jet is fully open, the fuel flow is about 10% more than that supplied by the main jet. If the throttle is rapidly opened (as when the gas pedal of a car is quickly depressed), the fuel flow does not respond instantly. To improve the engine response, an accelerator pump may be used to supply fuel at a rate that is proportional to the speed of the accelerator motion. A very fuel-rich mixture is used to start a cold engine, on the assumption that if enough fuel is introduced into the intake manifold, some of it will surely evaporate and start the engine. A butterfly valve called a choke is installed between the impact tube and the venturi, as illustrated in Figure 4.19, to increase the pressure drop and therefore the fuel flow rate through the main metering orifice. The choke is frequently operated automatically, controlled by the exhaust manifold temperature and the inlet manifold pressure. Rich operation during startup leads to high CO and hydrocarbon emissions. As much as 40% of the hydrocarbons emitted during automotive test cycles may be released during the warm-up phase. We have examined only a few of the features that have been incorporated into automotive carburetors. Since the carburetor directly controls the equivalence ratio of the mixture reaching the engine, it plays a central role in the control of automotive emissions. Much more elaborate fuel metering systems have been developed to achieve Sec. 4.1 Spark Ignition Engines 259 Air Cam Power-jet metering orifice Figure 4.19 Carburetor with power jet and choke (Taylor, 1966). Reprinted by pennission of MIT Press. the fine regulation required for emission control. Electronically manipulated valves have replaced the simple mechanically controlled fuel metering, facilitating more precise control of engine operation through the use of computers. Fuel injection is used in place of carburetion in some spark ignition engines because the quantity of fuel introduced can be controlled independently of the airflow rate. Atomization of high-pressure fuel replaces the flow-induced fuel intake of conventional carburetors. Fuel may be injected into the intake manifold (injection carburetion) so that the mixture is controlled by an injector pump rather than being directly coupled to the airflow. Injection into the inlet ports allows cylinder-by-cylinder regulation of the equivalence ratio. Direct injection into the cylinder is also used in some engines, although this method is more sensitive to spray characteristics and may lead to imperfect mixing of fuel and air. Injection systems are becoming more common because they are so well suited to integration into feedback-controlled engine operation. 4.1.10 Intake and Exhaust Processes The flows through the intake and exhaust valves also influence engine operation and emissions. We have seen that the intake flow induces turbulence that, after amplification by rapid compression, governs the flame propagation. The opening of the exhaust valve near the end of the expansion stroke causes a sudden pressure decrease and adiabatic cooling that influence carbon monoxide emissions. Internal Combustion Engines 260 0.8C2:: j Chap. 4 0.6 0- 0.4 0.2 Valve o ~ 0.05 0.10 0.15 0.20 0.25 0.30 035 LID Figure 4.20 Poppet valve geometry and flow coefficient (Taylor. 1966). Reprinted by pemlission of MIT Press. The poppet valves through which the charge enters and the combustion products exit from the cylinder are illustrated in Figure 4.20. The mass fluxes through these valves are also described by the compressible flow relation, (4.53). The discharge coefficient depends on the valve lift, L, as illustrated in Figure 4.20. For large lift, L/ D > 0.25, the flow coefficient based on the valve area approaches a constant value of about 0.65, slightly larger than that for a sharp-edged orifice. For smaller lift, the flow coefficient is proportional to the lift, suggesting the area of a cylinder between the valve and the port could be used to describe the flow with a constant coefficient. Shrouds placed on the intake valve to induce swirl or to increase engine turbulence reduce the open area on this cylinder and therefore the flow rate. The intake and exhaust flows are not steady. There may be a substantial pressure difference between the cylinder and the manifold when a valve is first opened, leading to a brief period of very high flow rate. This transient flow is particularly pronounced during exhaust when the flow is initially choked. After a brief blowdown, the pressure drop decreases and the flow rate is governed by the piston motion. Calculated and measured flow rates from the work of Tabaczynski et al. (1972) are presented in Figure 4.9. Note that the exhaust valve opens about 50° before bottom dead center to allow the cylinder pressure to drop before the beginning of the exhaust stroke. It is also common practice to open the intake valve before the end of the exhaust stroke. This overlap reduces the amount of residual combustion products being mixed with the fresh charge. Improved scavenging achieved in this way increases the engine power output. The exhaust system includes a length of pipe, a muffler, and gas-cleaning equipment through which the combustion products must flow before entering the atmosphere. The pressure in the exhaust manifold must therefore be greater than atmospheric pressure. The pressure of the gas entering the cylinder is lower than atmospheric pressure, due to pressure drops in the carburetor (particularly across the throttle), intake manifold, and inlet valve. The work required to draw the fuel and air into the cylinder and to pump the combustion products from the cylinder is called the pumping work. The pressure in the cylinder at the end of the intake stroke only approaches atmospheric pressure for open-throttle operation at relatively low speed. From the cycle analysis, it should be apparent that the peak pressure and temperature depend on the intake pressure. Heat transfer from the hot engine block to the fuel-air mixture also influences the temperature. The variation of temperature and pressure with throttle po- Sec. 4.1 Spark Ignition Engines 261 sition, engine speed, and engine temperature can be expected to be important factors in the fonnation of pollutants. 4.1 .11 Crankcase Emissions Crankcase emissions are caused by the escape of gases from the cylinder during the compression and power strokes. The gases escape between the sealing surfaces of the piston and cylinder wall into the crankcase. This leakage around the piston rings is commonly called blowby. Emissions increase with increasing engine airflow, that is, under heavy load conditions. The resulting gases emitted from the crankcase consist of a mixture of approximately 85% unburned fuel-air charge and 15% exhaust products. Because these gases are primarily the carbureted fuel-air mixture, hydrocarbons are the main pollutants. Hydrocarbon concentrations in blowby gases range from 6000 to 15,000 ppm. Blowby emissions increase with engine wear as the seal between the piston and cylinder wall becomes less effective. On cars without emission controls, blowby gases are vented to the atmosphere by a draft tube and account for about 25 % of the hydrocarbon emissions. Blowby was the first source of automotive emissions to be controlled. Beginning with 1963 model cars, this category of vehicular emissions has been controlled in cars made in the United States. The control is accomplished by recycling the blowby gas from the crankcase into the engine air intake to be burned in the cylinders, thereby keeping the blowby gases from escaping into the atmosphere. All control systems use essentially the same approach, which involves recycling the blowby gases from the engine oil sump to the air intake system. A typical system is shown in Figure 4.21. Ventilation air is drawn down into the crankcase and then up through a ventilator valve and hose and into the intake manifold. When airflow through the carburetor is high, additional air from the crankcase ventilation system has little effect on engine operation. However, during idling, airflow through the carburetor is so low that the returned blowby gases could alter the air-fuel ratio and cause rough idling. For this reason, the flow control valve restricts the ventilation flow at high intake manifold vacuum (low engine speed) and permits free flow at low manifold vacuum (high engine speed). Thus high ventilation rates occur in conjunction with the large volume of blowby associated with high speeds; low ventilation rates occur with low-speed operation. Generally, this principle of controlling blowby emissions is called positive crankcase ventilation (PCV). 4.1.12 Evaporative Emissions Evaporative emissions issue from the fuel tank and the carburetor. Fuel tank losses result from the evaporation of fuel and the displacement of vapors when fuel is added to the tank. The amount of evaporation depends on the composition of the fuel and its temperature. Obviously, evaporative losses will be high if the fuel tank is exposed to high ambient temperatures for a prolonged period of time. The quantity of vapor expelled when fuel is added to the tank is equal to the volume of the fuel added. Evaporation of fuel from the carburetor occurs primarily during the period just Internal Combustion Engines 262 Chap. 4 Oil filler cap Crankcase ventilator valve Figure 4.21 Crankcase emission control system. after the engine is turned off. During operation the carburetor and the fuel in the carburetor remain at about the temperature of the air under the hood. But the airflow ceases when the engine is stopped, and the carburetor bowl absorbs heat from the hot engine, causing fuel temperatures to reach 293 to 313 K above ambient and causing gasoline to vaporize. This condition is called a hot soak. The amount and composition of the vapors depend on the fuel volatility, volume of the bowl, and temperature of the engine prior to shutdown. On the order of 10 g of hydrocarbons may be vaporized during a hot soak. Fuel evaporation from both the fuel tank and the carburetor accounts for approximately 20% of the hydrocarbon emissions from an uncontrolled automobile. It is clear that gasoline volatility is a primary factor in evaporative losses. The measure of fuel volatility is the empirically detennined Reid vapor pressure, which is a composite value reflecting the cumulative effect of the individual vapor pressures of the different gasoline constituents. It provides both a measure of how readily a fuel can be vaporized to provide a combustible mixture at low temperatures and an indicator of the tendency of the fuel to vaporize. In a complex mixture of hydrocarbons, such as gasoline, the lowest-molecular-weight molecules have the greatest tendency to vaporize and thus contribute more to the overall vapor pressure than do the higher-molecular-weight constituents. As the fuel is depleted of low-molecular-weight constituents by evaporation, the fuel vapor pressure decreases. The measured vapor pressure of gasoline there- fore depends on the extent of vaporization during the test. The Reid vapor-pressure detennination is a standard test at 311 K in which the final ratio of vapor volume to Sec. 4.1 263 Spark Ignition Engines 30 ~----,-------,-----=------ c 60 0 u +- c (j) ~ ... I /'HC I / 80 Q; ;' Chap. 4 40 (j) 0.. 20 I I I I I I I I I I I I I I I I I I I I I I Catalytic Thermal ~ Exhaust temperature II o 400 800 1000 Reactor temperature (K) 1200 Figure 4.24 Comparison of catalytic converter and thennal reaetor for oxidation of CO and hydroearbons. Noncatalytic processes for vehicular emission control can yield significant improvements in carbon monoxide and hydrocarbon emissions. The problem of NO< emission control is not easily alleviated with such systems. Control of NO< emissions through noncatalytic reduction by ammonia is feasible only in a very narrow window of temperature, toward the upper limit of the normal exhaust temperature range, making joint control of products of incomplete combustion and NO< a severe technological challenge. Furthermore, the need to ensure a proper flow of ammonia presents a formidable logistical problem in the implementation of such technologies for control of vehicular emissions. Catalytic converters. By the use of oxidation catalysts, the oxidation of carbon monoxide and hydrocarbon vapors can be promoted at much lower temperatures than is possible in the gas phase, as shown in Figure 4.24. The reduction of NO is also possible in catalytic converters, provided that the oxygen content of the combustion products is kept sufficiently low. In the catalytic converter, the exhaust gases are passed through a bed that contains a small amount of an active material such as a noble metal or a base metal oxide deposited on a thermally stable support material such as alumina. Alumina, by virtue of its porous structure, has a tremendous surface area per unit volume. Small pellets, typically a few millimeters in diameter, or thin-walled, honeycomb, monolithic structures, illustrated in Figure 4.25, are most commonly used as the support. Pellet supports are inexpensive, but when packed closely in a reactor, they produce large pressure drops across the device, increasing the back-pressure in the exhaust system. They may suffer from attrition of the catalyst pellets due to motion during use. This problem can be reduced, but not entirely overcome, through the use of hard, relatively high density pellets. The mass of the catalyst bed, however, increases the time required for the bed to heat to the temperature at which it becomes catalytically active, thereby allowing substantial CO and hydrocarbon emissions when the engine is first started. Mon- Sec. 4.1 267 Spark Ignition Engines Insulation Converter shell ~~~~:::=:;cata Iyst Fill plug Insu~;;:;;;;;:;;;;:;;:;~=====~===::::::::::::.._ B'd'"PP"t,o,,~2N~tg", Catalytic pellet compound l<"igure 4.25 Schematic of pellet-type catalytic converter. olithic supports allow a freer exhaust gas flow, but are expensive and less resistant to mechanical and thermal damage. In particular, the rapid temperature changes to which a vehicular catalytic converter is exposed make them1al shock a very serious problem. Many materials will catalyze the oxidation of CO or hydrocarbons at typical exhaust gas temperatures. The oxidation activities per unit surface area for noble metals, such as platinum, are high for both CO and hydrocarbons. Base metal oxide catalysts, notably CuO and C0 3 0 4 , exhibit similar activities for CO oxidation but are significantly less active for hydrocarbon oxidation (Kummer, 1980). Base metal catalysts degrade more rapidly at high temperature than do the noble metal catalysts. They are also more susceptible to poisoning by trace contaminants in fuels, such as sulfur, lead, or phosphorus. Hence most automotive emission catalysts employ noble metals. NO reduction can be achieved catalytically if the concentrations of reducing species are present in sufficient excess over oxidizing species. CO levels in the exhaust gases of 1.5 to 3% are generally sufficient. Two schemes are employed to achieve catalytic control of both NO, and products of incomplete combustion: (I) dual-bed catalytic converters and (2) three-way catalysts. The dual-bed system involves operation of the engine fuel-rich, to promote the reduction of NOt' Secondary air is then added to facilitate the oxidation of CO and hydrocarbons in a second catalyst. Rich operation, while necessary for the NO reduction, results in reduced engine efficiency. FurthemlOre, it imposes severe restrictions on engine operation. If the exhaust gases are too rich, some of the NO may be converted to NH 3 or HCN. The oxidation catalyst used to eliminate CO and hydrocarbons readily oxidizes these species back to NO, particularly if the catalyst temperature exceeds 700 K. If the engine is operated at all times at equivalence ratios very close to unity, it is possible to reduce NO and oxidize CO and hydrocarbons on a single catalyst bed known as a three-way catalyst (Kummer, 1980). The three-way catalytic converter requires very Internal Combustion Engines 268 Chap. 4 precise control of the operating fuel/air ratio of the engine to ensure that the exhaust gases remain in the narrow composition window illustrated in Figure 4.26. Platinum can be used to reduce NO" but the formation of NH, under fuel-rich conditions limits its effectiveness as an NOr-reducing catalyst. NO can be reduced in slightly fuel-lean combustion products if a rhodium catalyst is used. Moreover, rhodium does not produce NH 3 efficiently under fuel-rich conditions. Rhodium is not effective, however, for the oxidation of paraffinic hydrocarbons. In fuel-lean mixtures, platinum is an effective oxidation catalyst. To achieve efficient control of CO and hydrocarbons during the fuel-rich excursions, a source of oxygen is needed. Additives that undergo reduction and oxidation as the mixture composition cycles from fuel-lean to fuel-rich and back (e.g., Re0 2 or Ce02), may be added to the catalyst to serve as an oxygen reservior. Three-way catalysts are thus a mixture of components designed to facilitate the simultaneous reduction of NOr and oxidation of CO and hydrocarbons. Because this technology allows engine operation near stoichiometric where the efficiency is greatest, and because the advent of semiconductor exhaust gas sensors and microcomputers make feedback control of the fuel-air mixture feasible, the three-way converter has rapidly become the dominant form of exhaust gas treatment in the United States. The extremely narrow operating window has been a major driving force behind 90 80 - i? 70 .'" 60 eQ) '+'+- Q) e o .~ Q) > e o u 50 40- 30 (.) ~ I- 20 10 - 1.0 ... > 0- "'ee 0.8 - 06 Q).- ",0 00. 0'wQ) '" 0.4 0.2 0.98 1.0 1.02 Stoichiometric ratio, ~-1 Figure 4.26 Three-way catalyst conversion efficiency and exhaust gas oxygen sensor signal as a function of equivalence ratio (Hamburg et aI., 1983; © SAE, Inc.). Sec. 4.2 Diesel Engine 269 the replacement of mechanically coupled carburetors with systems better suited to electronic control. Feedback control of engines using exhaust gas sensors results in operation that oscillates about the stoichiometric condition in a somewhat periodic manner (Kummer, 1980). The frequency of these oscillations is typically on the order of 0.5 to 4 Hz, with excursions in equivalence ratio on the order of ±O.Ol equivalence ratio units. In addition to NO, CO, and unoxidized hydrocarbons, catalyst-equipped spark ignition engines can emit sulfuric acid aerosol, aldehydes, and under rich conditions, H 2 S. Unleaded gasoline typically contains 15G to 600 ppm by weight of sulfur. The sulfur leaves the cylinder as S02, but the catalyst can promote further oxidation to SO,. As the combustion products cool, the SO, combines with water to form an H 2 S0 4 aerosol. H 2S formation requires high catalyst temperatures (> 875 K) and a reducing atmosphere. This may occur, for example, when an engine is operated steadily at high speed for some time under fuel-lean conditions and is then quickly slowed to idle fuel-rich operation. HCN formation may occur under similar conditions. During startup, when the catalyst is cold, hydrocarbons may be only partially oxidized, leading to the emission of oxygenated hydrocarbons. Aldehyde emissions, however, are generally low when the catalyst is hot. 4.2 DIESEL ENGINE Like the spark ignition engine, the diesel is a reciprocating engine. There is, however, no carburetor on the diesel. Only air (and possibly recycled combustion products for NO, control by EGR) is drawn into the cylinder through the intake valve. Fuel is injected directly into the cylinder of the diesel engine, beginning toward the end of the compression stroke. As the compression heated air mixes with the fuel spray, the fuel evaporates and ignites. Relatively high pressures are required to achieve reliable ignition. Excessive peak pressures are avoided by injecting the fuel gradually, continuing far into the expansion stroke. The rate at which the fuel is injected and mixes with the air in the cylinder determines the rate of combustion. This injection eliminates the need to throttle the airflow into the engine and contributes to the high fuel efficiency of the diesel engine. As in the steady-flow combustor, turbulent mixing profoundly influences the combustion process and pollutant formation. The unsteady nature of combustion in the diesel engine significantly complicates the process. Rather than attempt to develop quantitative models of diesel emissions, we shall explore some of the features that govern the formation of pollutants in diesel engines. Several diesel engine configurations are in use today. Fuel is injected directly into the cylinder of the direct injection (DI) diesel, illustrated in Figure 4.27(a). In the direct injection diesel engine, most of the turbulence is generated prior to combustion by the airflow through the intake valve and the displacement of gases during the compression stroke. The fuel jet is turbulent, but the time scale for mixing is comparable to that for entrainment, so the gas composition does not approach homogeneity within the fuel jet. 270 Internal Combustion Engines Chap. 4 Fuel injector Cylinder head / ' Cavity /"...,.::0----- Piston (a) Fuel injector Glow plug ~~~~~~~f ~ 1 I Passageway Exhaust valve Prechamber Main chamber (b) Figure 4.27 Diesel engine types: (a) direct injection; (b) prcchamber. The use of a prechamber, as shown in Figure 4.27(b), enhances the mixing of the fuel and air in the indirect injection (IDI) or prechamber diesel engine. As the gases burn within the prechamber, they expand through an orifice into the main cylinder. The high kinetic energy of the hot gas jet is dissipated as turbulence in the jet and cylinder. This turbulence enhances mixing over that of the direct injection engine. Improved mixing limits the amount of very fuel-rich gas in the cylinder, thereby reducing soot emissions. Most light-duty diesel engines are of the indirect injection type because of the reduced particulate emissions afforded by this technology. This benefit is not without costs, how- Sec. 4.2 271 Diesel Engine ever. The flow through the orifice connecting the prechamber to the cylinder results in a pressure drop, thereby reducing the efficiency of the engine. Diesel engines may also be classified into naturally aspirated (NA), supercharged, or turbocharged types, depending on the way the air is introduced into the cylinder. In the naturally aspirated engine, the air is drawn into the cylinder by the piston motion alone. The supercharger is a mechanically driven compressor that increases the airflow into the cylinder. The turbocharger similarly enhances the intake airflow by passing the hot combustion products through a turbine to drive a centrifugal (turbine-type) compressor. Compression of the air prior to introduction into the cylinder results in compression heating. This may be detrimental from the point of view of NO, formation because it increases the peak combustion temperature. An intercooler may be installed between the compressor and the intake valve to reduce this heating. The fuel is sprayed into the cylinder through a number of small nozzles at very high pressure. The liquid stream issuing from the injector nozzle moves with high velocity relative to the gas. The liquid stream fonns filaments that break into large droplets. The breakup of the droplets in the fuel spray is characterized by the Weber number, the ratio of the inertial body forces to surface tension forces, We a where Pg is the gas density, v the relative velocity between the gas and the droplets, and a the surface tension of the liquid. As long as the Weber number exceeds a critical value of approximately 10, the droplets will continue to break into smaller droplets. Aerodynamic drag on the droplets rapidly decelerates them and accelerates the gas entrained into the fuel spray. Evaporation and combustion of the fuel can be described using the model developed in Section 2.7. In some cases, however, pressures and temperatures in the cylinder are high enough that the liquid fuel is raised above its critical point. The fuel spray then behaves like a dense gas jet. The entrainment of air into the unsteady, two-phase, variable-density, turbulent jet has been described by a variety of empirical models, simple jet entrainment models, and detailed numerical simulations. The problem is frequently complicated further by the use of swirling air motions to enhance mixing and entrainment. The swirling air motion sweeps the fuel jet around the cylinder, spreading it and reducing impingement on the cylinder wall. Since combustion in nonpremixed systems generally occurs predominantly at equivalence ratios near unity, combustion will occur primarily on the perimeter of the jet. Mixing of hot combustion products with the fuel-rich gases in the core of the fuel spray provides the environment in which large quantities of soot can be readily generated. (We discuss soot fonnation in Chapter 6.) The stoichiometric combustion results in high temperatures that promote rapid NOt fonnation in spite of operation with large amounts of excess air in the cylinder under most operating conditions. Some of the fuel mixes with air to very low equivalence ratios before any of the mixture ignites. Temperatures in this region may be high enough for some fuel decomposition 272 Internal Combustion Engines Chap. 4 and partial oxidation to occur, accounting for the relative abundance of aldehydes and other oxygenates in the dicsel emissions (Henein, 1976). Thus we see that diesel engines exhibit all of the complications of steady-flow spray flames, in addition to being unsteady. To describe the formation of pollutants quantitatively would require the development of a probability density function description of the unsteady mixing process. While such models are being explored (Mansouri et a!., 1982a,b; Kort et a!., 1982; Siegla and Amann, 1984), the methods employed are beyond the scope of this book. We shall examine, instead, the general trends as seen in both experimental and theoretical studies of diesel engine emissions and emission control. 4.2.1 Diesel Engine Emissions and Emission Control Relatively low levels of gaseous exhaust emissions are achieved by light-duty (automobile) diesel engines without the use of exhaust gas treatment usually applied to gasoline engines to achieve similar emission levels (Wade, 1982). The species mole fractions in diesel exhaust are somewhat misleading because of the low and variable equivalence ratios at which diesel engines typically operate. At low-load conditions, the operating equivalence ratio may be as low as 0.2, so the pollutants are diluted significantly with excess air. Because the equivalence ratio is continually varying in normal use of diesel engines, and to facilitate comparison to other engines, it is more appropriate to report emission levels in terms of emissions per unit of output: g Mr 1 for stationary engines or heavy-duty vehicles or g km- 1 for light-duty vehicular diesel engines. Injection of the liquid fuel directly into the combustion chamber of the diesel engine avoids the crevice and wall quench that allows hydrocarbons to escape oxidation in carbureted engines, so hydrocarbon emissions from diesel engines are relatively low. Furthermore, diesel engines typically operate fuel-lean, so there is abundant oxygen to burn some of the hydrocarbons and carbon monoxide formed in midair in the cylinder. NO, emissions from prechamber diesel engines are lower than the uncontrolled NO, emissions from homogeneous charge gasoline engines (Wade, 1982). The low NO, emissions result from the staged combustion in the prechamber diesel and the inhomogeneous gas composition. Particulate emissions from diesel engines tend to be considerably higher than those of gasoline engines and represent a major emission control challenge. Factors that influence the emissions from diesel engines include the timing and rate of fuel injection, equivalence ratio, compression ratio, engine speed, piston and cylinder design, including the use of prechambers, and other design factors. The influence of the overall equivalence ratio on engine performance is shown in Figure 4.28(a). The brake mean effective pressure increases with equivalence ratio, so higher equivalence ratios correspond to higher engine power output or load. The exhaust gas temperature also increases with equivalence ratio. Fuel consumption is high at low equivalence ratio, but decreases sharply as the equivalence ratio is increased. As the equivalence 1.3 12 1.1 1.0 0.9 Exhaust T x l O > - - - X 0.8 /x x 0.7 0.6 0.5 0.4 0.3 0.2 0.1 0 0.1 0.3 0.4 0.5 ep 0.6 0.7 0.8 0.9 1.0 0.9 1.0 (a) 30 28 <) CO He 26 0 24 + Particulate matter x NO I -, :2: 22 !?) 20 ~'" 18 c 0 ·til .~ E '" .~ ..... u '" -'" '" 0. This assumption, however, is rather tenuous. During the intake stroke, the two mixtures flow through the two intake valves. Because of the large displacement volume, some of the prechamber mixture may flow out through the connecting orifice into the main chamber during intake. In the compression stroke, gas from the main chamber is forced back into the prechamber, so the prechamber will contain a mix of the rich and lean charges. Some of the prechamber mixture will remain in the main chamber. Thus, in spite of using carburetors to prepare the two charges, the two mixtures may not be unifonn at ignition. During combustion, further mixing between the two gases takes place. Analysis of gas samples collected at the exhaust port suggests that no significant stratification of the mixture remains at the end of the expansion stroke. This does not mean, however, that the gas is unifonnly mixed on the microscale. As in the diesel engine, the turbulent mixing process plays a critical role in detennining the pollutant emission levels. The dependence of the emissions from the prechamber stratified charge engine on equivalence ratio, as a result, is somewhat weaker than that of the conventional homogeneous charge engine. The reduction in the peak NO, emissions with this technology is modest, but the ability to operate at very low equivalence ratios makes significant emission reductions possible. Other types of stratified-charge engines utilize direct injection of the fuel into the cylinder to create local variations in composition. The fuels used in such engines generally still have the high-volatility characteristics of gasoline, and a spark is used for ignition. Two types of direct-injection stratified-charge engines are shown in Figure 4.31. The early injection version of the direct-injection stratified-charge engine typically uses a broad conical spray to distribute the fuel in the central region of the cylinder, where the piston has a cup. The spray starts early, about halfway through the compression stroke, to give the large fuel droplets produced by the low-pressure injector time to evaporate prior to ignition. In contrast, a late injection engine more closely resembles a diesel engine. A high-pressure injector introduces a narrow stream of fuel, beginning just before combustion. A swirling air motion carries the spray toward the spark plug. The details of combustion in these engines are not well understood. Mixing and combustion occur simultaneously, so the dependence of emissions on equivalence ratio more closely resembles the weak dependence of poorly mixed steady-flow combustors than that of the conventional gasoline engine. Because pure fuel is present in the cylinder during combustion, the direct-injection stratified-charge engines suffer from one of the major difficulties of the diesel engine: soot is fonned in significant quantities. These are but a few of the possible configurations of reciprocating engines. The 280 Internal Combustion Engines Chap. 4 (a) (b) Figure 4.31 injection. Direct-injection stratified-charge engine: (a) early-injection; (b) late- introduction of emission limitations on automobile engines has led to major new technological developments, some successful, many that failed to meet their developers' expectations. Renewed concern over engine efficiency has once again shifted the emphasis. To satisfy both environmental and fuel utilization constraints, the use of computer control, catalytic converters, and exhaust gas sensors has been introduced. These have diminished the level of interest in stratified charge engines, since NO, reduction catalysts are needed to meet strict emission limits and require operation near stoichiometric. The problems of emission control for heavy-duty engines remain unresolved. Yet, as automobile emissions are reduced, large engines in trucks, railway engines, compressors, and so on, are becoming increasingly important sources of atmospheric pollutants, notably NO, and soot. 4.4 GAS TURBINES The fourth major class of internal combustion engines is the gas turbine. The power output of the gas turbine engine can be very high, but the engine volume and weight are generally much smaller than those of reciprocating engines with comparable output. The Sec. 4.4 281 Gas Turbines original application of the gas turbine engine was in aircraft, where both weight and volume must be minimized. The high power output also makes the gas turbine attractive for electric power generation. Like reciprocating engine exhaust, the exhaust gases from the gas turbine are quite hot. The hot combustion products can be used to generate steam to drive a steam turbine. High efficiencies of conversion of fuel energy to electric power can be achieved by such combined-cycle power generation systems. Clearly, the constraints on these applications differ greatly, so the technologies that can be applied to control emissions for one type of gas turbine engine are not always applicable to the other. Unlike the reciprocating engines, the gas turbine operates in steady flow. Figure 4.32 illustrates the main features of a gas turbine engine. Combustion air enters the turbine through a centrifugal compressor, where the pressure is raised to 5 to 30 atm, depending on load and the design of the engine. Part of the air is then introduced into the primary combustion zone, into which fuel is sprayed and bums in an intense flame. The fuel used in gas turbine engines is similar in volatility to diesel fuel, producing droplets that penetrate sufficiently far into the combustion chamber to ensure efficient combustion even when a pressure atomizer is used. The gas volume increases with combustion, so as the gases pass at high velocity through the turbine, they generate more work than is required to drive the compressor. This additional work can be delivered by the turbine to a shaft, to drive an electric power generator or other machinery, or can be released at high velocity to provide thrust in aircraft applications. The need to pass hot combustion products continuously through the turbine imposes severe limits on the temperature of those gases. Turbine inlet temperatures are limited to 1500 K or less, depending on the blade material. The development of turbine blade materials that can withstand higher temperatures is an area of considerable interest because of the efficiency gains that could result. In conventional gas turbine engines, the cooling is accomplished by dilution with additional air. To keep the wall of the combustion chamber, known as the combustor can, cool, additional air is introduced through wall jets, as shown in Figure 4.33. The distribution of the airflow along the length of the combustion chamber, as estimated by Morr (1973), is also shown. Combustion chamber Fuel Turbine Compressor- ---+- Air Figure 4.32 Products Gas turbine engine configuration. Net work 282 Internal Combustion Engines Liner~~~ Chap. 4 I ~)\ i ~ r-~ \\ cooling ,;, Primary zone Fuel -I-+H-_ Secondary zone I C~(i!1 ~--=---------- Dilution zone I I i I =----- II 100 ~ .2~ '+-0 ~ .- o +

This is a preview of the paper, limited to some initial content. Full access requires DieselNet subscription.Please log in to view the complete version of this paper.

Abstract: Components located after the intake manifold in four-stroke diesel engines serve important functions in managing the air supply to the cylinder. Poppet-type valves control the timing of flow into and out of the cylinder. The intake port design impacts the breathing capacity of the engine as well as the bulk motion of the air as it enters the cylinder.

As the airflow passes various components and stages of the intake system, different properties and characteristic of the intake charge have been modified to achieve the overall goals of the air management system. The intake air filter ensures that air cleanliness is adequate, the charge air composition and oxygen content is controlled by introducing EGR to the intake air and the compressor and charge air cooler ensure that intake manifold pressure and temperature objectives are met and that intake charge density is within design limits. A few final aspects of air management are achieved after the intake charge exits the intake manifold and enters the cylinder. Valves or ports control the timing of air flow to the cylinder. Also, the passage between the intake manifold and cylinder can have a significant influence on the flow as it enters the cylinder and can be used to impart a suitable bulk motion and kinetic energy to the charge to support the mixing of air, fuel and intermediate combustion products in-cylinder.

In four stroke engines, intake gas enters the cylinder through a port located in the cylinder head and past a valve used to open and close the port. In two stroke engines—discussed elsewhere—ports in the cylinder liner that are alternately covered and uncovered by the piston are commonly used.

Gas flow into and out of the cylinder in 4-stroke engines is controlled almost exclusively by poppet-style valves (Figure 1). While other valve designs have been used or proposed, none appear to be able to match the reliability and sealing ability of the poppet style valve. The most common poppet valve construction in automotive use is the one-piece valve where the entire valve is made from the same material. However, other variations are available including:

A welded tip construction has a separate tip welded to the stem above the keeper grove. The tip can be made from a material that is much more wear resistant than the rest of the valve.

A two-piece construction has a separate stem welded above the fillet, Figure 2 left.

An internally cooled construction has a hollow stem containing a coolant such as metallic sodium or sodium-potassium mixture and is commonly used in extreme duty and high performance exhaust valves, Figure 2 center. Valve temperature peaks are reduced due to the “shaker effect” of the molten metal and these designs can withstand thermal loads particularly well. The temperature in the hollow neck can be lowered by about 80 to 130 K, reducing overall wear of the valve and valve seat insert.

Some designs also have a hollow cavity in the valve head that contains metallic sodium, Figure 2, right. This is an extension of the classic sodium-filled hollow valve, with an additional cavity in the valve head. This can further temperature peaks in the valve head and further increase the valve service life.

A welded seat face construction has a valve seat that is welded with a hard overlay to better withstand conditions that would otherwise lead to extreme valve seat wear and/or corrosion.

In addition to different construction styles, valves can have different design enhancements to improve their durability. Seat face strain hardening can be used to moderately enhance seat wear endurance in cases where a welded seat face construction is not necessary. Stem surface treatments can be used to reduce friction and/or wear especially were adhesive wear may otherwise be encountered. Aluminizing the valve seat face and sometimes the combustion face to improve corrosion resistance in lead oxide environments was once popular for engines burning leaded gasoline. Tip caps fitted over the end of the valve stem can be used to improve tip wear resistance where welding of dissimilar metals is a problem.

Shift System Components in Manual Transmissions Automotive Product Information API 09 This publication has been produced with a great deal of care, and all data have been checked for accuracy. However, no liability can be assumed for any incorrect or incomplete data. Product pictures and drawings in this publication are for illustration only and are not intended as an engineering design guide. Applications must be developed only in accordance with the technical information, dimension tables, and dimension drawings contained in this publication. Due to constant development of the product range, we reserve the right to make modifications. The terms and conditions of sale and delivery underlying contracts and invoices shall apply to all orders. Produced by: INA-Schaeffler KG 91072 Herzogenaurach (Germany) Mailing address: Industriestrasse 1–3 91074 Herzogenaurach (Germany) © by INA · September 2003 All rights reserved. Reproduction in whole or in part without our authorization is prohibited. Printed in Germany by: mandelkow GmbH, 91074 Herzogenaurach Table of Contents Page Shift System Components in Manual Transmissions 4 Introduction 4 4 4 4 QFD – Quality Function Deployment CAE – Computer Aided Engineering Tests to Verify Function and Operation MEOST (Multiple Environment Overstress Testing) Modern Manufacturing Technology 5 Manual Transmission Shifting Requirements 5 5 5 Transmission Operation: Driver Requirements Gear Shifting: Design Requirements Ideal Shifting 6 Ideal Shift Lever Moment Curve during the Selection Stroke 6 7 Theoretical Ideal Shift Lever Moment Curve Shift Lever Moment Curves for Shift Systems Supported by Plain Bearings and in Roller Bearings 8 Ideal Shift Lever Force during the Shift Stroke 8 9 Theoretical Ideal Shift Lever Force Curve Shift Lever Force Curve – Comparative Measurements 10 Summary 10 10 Automobile Shift System Component Selection Development Trends 11 Addresses 3 Introduction INA’s expertise in developing shift systems and components is based on many years of experience working closely with automobile and transmission manufacturers. Because of the continuous development of components and the use of cutting-edge technologies in manufacturing our products, INA is a well-known engineering partner and a full service supplier. INA employs the most modern engineering and quality assurance methods currently used. These methods include: ■ QFD – Quality Function Deployment – To establish customer requirements and translates these requirements into a design concept ■ CAE – Computer Aided Engineering – The use of state-of-the-art analysis and calculation methods for component design and function simulations respectively ■ Tests to Verify Function and Operation MEOST (Multiple Environment Overstress Testing) To evaluate a manual transmission in terms of the following: – shifting characteristics – shift forces – under extreme temperatures – friction characteristics – component strength – service life – corrosion resistance – transmission testing under simulated operating conditions to fine-tune components to the desired performance level 4 Modern Manufacturing Technology INA’s technology allows a cost-effective component design by means of the following: ■ high precision machining or cold forming of components ■ extrusion methods ■ heat treatment (e.g. hardening) ■ surface plating (INA Corrotect® plating and DSV thin layer chrome plating) ■ in-house plastic molding ■ fine blanking techniques ■ state-of-the-art welding and bonding technology Manual Transmission Shifting Requirements The customer’s acceptance of a vehicle is greatly influenced by the positive operation of the transmission and how well it is adapted to the vehicle. However, with increasing comfort demands, additional criteria are now being used to evaluate the quality of manual transmissions such as ease of use, shift comfort and positive shift feel. 1 R 2 3 1 5 2 4 Transmission Operation: Driver Requirements – Figure 1 Since the shift system is the only direct connection between the driver and the transmission, the perceived shift quality is important in the driver’s assessment of the vehicle. The driver wants: 1 to know the shift lever position at all times 2 to feel precise resistance when shifting 3 to use minimal and consistent force when shifting gears 4 minimum shift lever throw al F i m s minim al 4 Figure 1 · Criteria for a positive shift feel 1 1 3 5 Ideal Shifting – Figure 2 It is extremely difficult to base technical requirements on a “positive shift feel” since this is necessarily a subjective evaluation. One solution is to evaluate the mechanics of the shift system and plot the shift lever displacement, lever force and shift time in a graph. Observing the time and speed of the gear shift allows technically feasible “ideal shifting conditions.” – + 2 4 R 2 1 3 5 2 4 + – R 134 095 Gear shifting operation Shifting gears involves two orthogonal motions of the shift lever the “selection” stroke and the “shift” stroke: 1 the shift rail is chosen in the “selection” stroke 2 in the “shift” stroke, a gear is synchronized and engaged m ni 134 092 Gear Shifting: Design Requirements The perceived quality of shifting can be achieved through the proper design of the shift system. Gear changes are judged positively if they have the following characteristics: ■ precise ■ quick ■ require little effort ■ smooth 3 Figure 2 · Ideal shifting conditions – Selection and shift strokes 5 Ideal Shift Lever Moment Curve during the Selection Stroke In order for the driver to get the ideal lever feel when selecting a shift rail, the following conditions must be given: ■ The gearshift lever must be in neutral. ■ The motion curve must be smooth across the entire pivoting range. ■ The force required must be minimal and increase gradually. Theoretical Ideal Shift Lever Moment Curve Figure 3 shows the theoretical ideal moment curve when the gear shift lever is pivoted left and right from the neutral position. Positive and negative directions are indicated in the graph. Positive direction When the lever is pivoted to the 5th/reverse gear shift gate, the lever is said to pivot in the positive direction Negative direction Pivoting the shift lever into the 1st/2nd gear shift gate corresponds to the negative direction. Reversing the pivot direction also reverses the moment direction. Interpreting the moment curve 1 The sharp rise in the curve from the horizontal axis results from the clearance-free lever support in the neutral position. 2 The remainder of the curve is smooth and rises gradually. A horizontal curve – corresponding to a constant shift force – would be assessed as being undefined and unstable. 3 The final position of the shift lever is marked by another increase in moment. This final effort spike is favored by the driver. 4 The moment values are on the return stroke, the hysteresis, is from the lever inertia. 3 1 3 5 2 – + 1 4 R 0 2 – Selection moment + 4 1 = Neutral position 2 = Selection motion 3 = Final position (5 th gear/reverse gear shift gate) 4 = Return motion – 0 Pivot angle + Figure 3 · Moment curve during the selection stroke 6 = Return displacement 162 475 = Shift travel Measuring conditions The selection motion was measured in the shift gate neutral position and in 1st/2nd gears to the opposite positions 3rd/4th gears and 5th/reverse gears respectively. The pivoting motion occurred in less than two seconds. The maximum pivot angle of the selector shaft was 12º. The moment was checked at the selector shaft at the transmission entry. Several overlapping motion measurements are given in Figure 4. Shift elements in vehicle transmissions such as selector shafts, shift rails, shift rods and reversing levers must have the best bearing supports possible. To do this, their function in the transmission housing must be considered. The type of bearing support – plain bearing or rolling bearing arrangements – has a significant effect on the shift process, the shift curve and thus the feel the driver has when shifting. Shift Lever Moment Curves for Shift Systems Supported by Plain Bearings and in Rolling Bearings – Figure 4 The shift system of a manual transmission used for comparison is mounted in an aluminum housing and incorporates a selector shaft, shift rails as well as a steel reversing lever. The selector shaft and shift rods have plain bearing supports in machined bores of the transmission housing. Interpreting the moment curve 1 The design containing only plain bearings displays an imprecise neutral position of the shift lever. The friction between the movable components leads to significant losses in aligning force (hysteresis). 2 Due to the significantly lower internal friction of the rolling bearings, the moment curve is much better and hysteresis is lower. Even the return stroke of the shift lever to the neutral position is more precise. Shift system with plain bearing supports The reversing lever has plain bearing supports on steel studs on both sides. Shift system with some rolling bearing supports As a means of comparison, the reversing lever has rolling bearing supports on the steel stud. Reverse lever with plain bearing supports Reverse lever with rolling bearing supports 1 3 5 + – 2 4 R + Selection moment 0 – 0 – – 0 Pivot angle + – 0 Pivot angle + 180 944 Selection moment + Figure 4 · Comparing the moment curves during the selection stroke: reverse lever with plain bearing supports versus rolling bearing supports 7 Ideal Shift Lever Force during the Shift Stroke Positive and negative shift forces occur across the shift curve or shift path. Theoretical Ideal Shift Lever Force Curve The theoretical ideal shift forces curve – see also section entitled Ideal Shifting, p. 5 – for the necessary shift motion when engaging a particular gear is described in Figure 5. Positive shift forces The positive shift forces counteract the motion of the driver’s hand. Interpreting the shift lever force curve 1 The shift stroke is initiated by moving the gear out of the neutral position with the shift lever. 2 The increase to the first force peak – the synchronization of speeds – follows. It is not too high and does not stop abruptly 3 The second force peak characterizes smooth gear clutch teeth engagement. 4 Reversing the force conveys the feeling that the gear has reached the final position on its own. The shift stroke is now complete since the shift lever locks. 5 When shifting the gear back from the final position, a precise force increase occurs followed by a force reverse. Negative shift forces The negative shift forces result when the direction of the shift force is reversed. The driver notices a reduction in resistance. 1 3 5 + 2 – 3 1 4 1 = Neutral position 2 = Synchronization 3 = Gear engagement 4 = Final position 5 = Return displacement 0 Shifting force + 2 4 R – 5 0 Shift travel Figure 5 · Theoretical ideal force curve during the shift stroke 8 + 162 477 – The earlier and more precise the force reversal occurs (i.e. the shift lever is back in the initial position in the shift gate), the higher the driver’s assurance that the gear has been disengaged properly. Shift Lever Force Curve – Comparative Measurements Figure 6 shows force curves for shift strokes from 1st to 2nd gear and from 2nd to 3rd gear. Measuring conditions Several measurements were made on the selector shaft of a manual transmission containing rolling bearings. The measurements are shown in the figure below projected on top of each other. Interpreting the shift force curve The movement to the right in the figure shows the shift curve for 1st and 3rd gears and the movement to the left the curve for the opposite 2nd gear. Since the direction is reversed here, the direction of force also changes. The shift points described in the section entitled Theoretical Ideal Shift Lever Force Curce, p. 8 can clearly be seen here. Although the required force is at different levels depending on the gear, it is not the ideal force. When shifting from 1st to 2nd gear, the force difference between the points “speed synchronization” and “engagement” is still too large to be evaluated as favorable. Shifting from 1st to 2nd gear 1st gear Shifting from 2nd to 3 rd gear 1 3 5 3 rd gear + – 2 4 R + – – Shift force 0 Shift force 0 + – 0 Shift travel + – 0 Shift travel + 162 478 2nd gear 2nd gear Figure 6 · Comparison of force curves during shift travel 9 Summary 2 Series RLF 6 Series ARRE Figure 7 · Selection of products for manual transmissions 10 3 Series RLF 7 Series SYN 123 013 Development Trends Because of the increasing demand for systems solutions, INA also supplies components or assemblies. These products have the following advantages: ■ combine several functions in one assembly ■ fit the mating parts exactly ■ reduce manufacturing complexity at the transmission assembly. 4 Series RLF 8 Gear shift module SE 014 073 Series HK..RS 8 140 104 5 7 134 074 Series PAP 105 103 1 6 Rolling bearings for rotary and oscillating motion for bearing supports in the shift fork, such as drawn cup needle roller bearings (open/closed end) and angular contact ball bearings Detent pins Intermediate rings for multiple-cone synchronization Gear shift modules 123 011 136 156 Automobile Shift System Component Selection 1 Permaglide® plain bearings for rotary and linear motion 2 Rolling bearings for rotary and limited linear motion for round shaft cross sections 3 Rolling bearings for limited linear motion for rectangular cross-sectioned shafts 4 Rolling bearings for limited linear motion with torque transmission 5 123 012 Separate optimization attempts will not bring about the required comfort for the entire shift system, even when expensive and flawless bearing are used. For this reason INA develops and manufactures specific products for vehicle shift systems that are adapted to the entire transmission application. A selection of these products is given in Figure 7. Addresses 100 009 Automotive Division North America Canada INA Canada Inc. 2871 Plymouth Drive Oakville Ontario L6H 5S5 Tel. +1/ 905 /829-27 50 Fax +1/ 905 /829-25 63 Mexico INA Mexico, S.A. de C.V. Paseo de la Reforma 383, int. 704 Col. Cuauhtemoc 06500 Mexico, D.F. Tel. + 52 / 5 / 5 25 00 12 Fax + 52 / 5 / 5 25 01 94 USA INA USA CORPORATION 308 Springhill Farm Road Fort Mill, South Carolina 29715 Tel. +1/ 803 /5 48 85 00 Fax +1/ 803 /5 48 85 94 South America Argentina INA Argentina S.A. Avda. Alvarez Jonte 1938 14 16 Buenos Aires Tel. +54 /11 /45 82 40 19 Fax +54 /11 /45 82 33 20 E-Mail inaarg@ina.com.ar Brazil INA Brasil Ltda. Av. Independência, nr. 3500 Bairro de Éden 18103-000 Sorocaba/São Paulo Caixa Postal 334 18001-970 Sorocaba Tel. +55 /15 /2 35 15 00 +55 /15 /2 35 16 00 Fax +55 /15 /2 35 19 90 E-Mail vendauto@ina.com.br Asian Pacific Rim Australia INA Bearings Australia Pty. Ltd. Locked Bag 1 Taren Point 2229 Tel. +61 /2 / 97 10 11 00 Fax +61 /2 / 95 40 32 99 E-Mail sales@ina.au.com China INA (China) Co. Ltd. 18 Chaoyang Road Taicang Economic Development Area Jiangsu Province 215 400 Tel. +86 /512 / 53 58 09 48 Fax +86 /512 / 53 58 09 95 Asian Pacific Rim Japan INA Bearing, Inc. Square Building 15 F 2-3-12, Shin-Yokohama Kohoku-ku, Yokohama, 222-0033 Tel. + 81/ 45 / 4 76 59 00 Fax + 81/ 45 / 4 76 59 20 Korea INA Korea Inc. 1054-2 Shingil-dong Ansan-shi, Kyounggi-do 425-839 Korea South Tel. + 82 / 31 / 4 90 98 00 Fax + 82 / 31 / 4 90 98 99 Africa South Africa INA Bearings (Pty.) Ltd. South Africa Caravelle Street Walmer Industrial Port Elizabeth 6001 P.O. Box 400 30 Walmer Port Elizabeth 60 65 Eastern Cape Tel. + 27/41/ 5 01 28 00 Fax + 27/41/ 5 81 04 38 E-Mail inquiries@ina.co.za India INA Bearing India Pvt. Ltd. Indo-German Technology Park Survey No. 297, 298, 299 Urawade, Tal: Mulshi Dist. Pune, Pin: 412108 Tel. +91/ 20 / 4 10 10 36 Fax +91/ 20 / 4 00 12 44 USA INA USA CORPORATION 335 East Big Beaver Road Suite 101 Troy, Michigan 48083-1235 Tel. +1/ 248 /5 28 90 80 Fax +1/ 248 /6 19 21 39 Europe Germany INA-Schaeffler KG Industriestrasse 1–3 91074 Herzogenaurach Tel. + 49 / 91 32 /82-0 Fax + 49 / 91 32 /82-49 50 E-Mail info@ina.com France INA France 93, route de Bitche BP 186 67506 Haguenau Cedex Tel. +33 / 3 88 63 40 40 Fax +33 / 3 88 63 40 41 Telex 870 936 Netherlands INA Nederland B.V. Gildeweg 31 3771 NB Barneveld Postbus 50 3770 AB Barneveld Tel. +31/ 342 / 40-30 00 Fax +31/ 342 / 40-32 80 E-Mail info@ina.nl Russia INA Moskau ul. Bolschaja Moltschanovka Nr. 23/38, Building 2 121019 Moskau Tel. + 7/ 095 / 2 32 15 38 + 7/ 095 / 2 32 15 39 Fax + 7/ 095 / 2 32 15 40 E-Mail inarussia@col.ru Slovenia INA kotalni lezaji Maribor Glavni trg, 17/b 2000 Maribor Tel. + 386 / 2 / 22 82-0 70 Fax + 386 / 2 / 2 28 20 75 E-Mail info@ina-lezaji.si Turkey Great Britain Norway Austria INA AUSTRIA GmbH. Marktstraße 5 Postfach 35 2331 Vösendorf Tel. + 43 /1/ 6 99 25 41-0 Fax + 43 /1/ 6 99 25 41 55 E-Mail ina.austria@ina.at Belgium INA Belgium S.P.R.L. Avenue du Commerce, 38 1420 Braine-l’Alleud Tel. + 32 / 2 / 3 89 13 89 Fax + 32 / 2 / 3 89 13 99 E-Mail ina@be.ina.com Czech Republic INA Ložiska s r.o. Průběžná 74 a 100 00 Praha 10 – Strašnice Tel. + 420 / 2 / 67 29 81 40 Fax + 420 / 2 / 67 29 81 10 E-Mail inaloziska@inaloziska.cz INA Bearing Company Ltd Forge Lane, Minworth Sutton Coldfield West Midlands B76 1AP Tel. +44 /121/ 3 51 38 33 Fax +44 /121/ 3 51 76 86 E-Mail ina.bearing@ina.co.uk Hungary INA Gördülöcsapágy Kft. 1146 Budapest, XIV. Hermina út 17. Postfach 229 1590 Budapest Tel. +36 /1/4 61 70 10 Fax +36 /1/4 61 70 13 Italy INA Rullini S.p.A. Strada Regionale 229 - km. 17 28015 Momo (Novara) Tel. +39 / 03 21/ 92 92 11 Fax +39 / 03 21/ 92 9 3 00 INA Norge AS Postboks 6404 Etterstad Nils Hansens Vei 2 0604 Oslo 6 Tel. +47/ 2 /2 64 85 30 Fax +47/ 2 /2 64 54 11 E-Mail ina@ina.no Poland INA Lozyska Spolka z o.o. ul. Stepinska 22/30 00-739 Warszawa Tel. +48 /22 /8 41 73 35 +48 /22 /8 51 36 85 Fax +48 /22 /8 51 36 84 Telex 813 527 omig pl Portugal INA Rolamentos Lda. Av. Fontes Pereira de Melo, 470 4149-012 Porto Tel. +351/ 22 / 5 32 08 90 Fax +351/ 22 / 5 32 08 61 E-Mail marketing@pt.ina.com Rumania CN INDUSTRIAL GROUP S.R.L. Bdul Garii Obor, nr. 8D 7000 Bucuresti, Sector 2 Tel. +40 /1/2 52 98 61 Fax +40 /1/2 52 98 60 E-Mail office@inacn.ro INA Rulmanlari Ticaret Ltd. Sirketi Aydin Sokak Dagli Apt. 4/10 1. Levent 34340 Istanbul Tel. + 90 / 212 /2 79 27 41 Fax + 90 / 212 /2 81 66 45 E-Mailinaturk@tr.ina.com Spain INA Iberia, s.l. Polígono Pont Reixat 08960 Sant Just Desvern Barcelona Tel. + 34 / 93 / 4 80 34 10 Fax + 34 / 93 / 3 72 92 50 E-Mail marketing@es.ina.com Sweden INA Sverige AB Box 41 195 86 Arlandastad Charles’gata 10 195 61 Arlandastad Tel. + 46 / 8 /59 51 09 00 Fax + 46 / 8 /59 51 09 60 E-Mail info@ina.se 11 91072 Herzogenaurach Internet www.ina.com E-Mail info@ina.com In Germany: Phone 0180 / 5 00 38 72 Fax 0180 / 5 00 38 73 From Other countries: Phone +49 / 9132 / 82-0 Fax +49 / 9132 / 82-49 50 Sach-Nr. 005-349-567/API 09 US-D 09031 INA-Schaeffler KG

For over 85 years, the ARO® Fluid Products business of Ingersoll Rand® has developed partnerships with more than 200 original equipment manufacturers and distributors, enabling us to better focus on the unique pumping needs of many industries. It’s a strategic merger of our partners’ application expertise, along with our decades-long legacy of designing and building outstanding piston pumps.

ARO piston pumps are capable of handling a wide variety of viscous fluids. With a wide selection of pressure ratios and displacement rates available, ARO offers a number of piston pump packages that can meet your specific application needs. Offered in multiple configurations, including single-post, two-post and heavy-duty two-post, our piston pump packages ensure you have the right solutions for the following applications as well as many others.

ARO offers a versatile range of piston pump packages capable of meeting the specific needs of your specific application. Available in multiple configurations with a broad selection of pressure ratios and displacement rates, our pumps can help unlock superior performance and increase production across a number a unique applications.

Paint and Coating Applications

Paint and Coating

The abrasive nature of paints and coatings can quickly break down weaker components. That’s why systems that handle the transfer of high-solids materials trust ARO piston pumps to stand up to the pressure.

Lubricant applications demand a pump that can dispense oils and greases accurately and without wasting a drop. That’s why ARO piston pumps are at the heart of the most vital systems.

When your application deals with dispensing highly abrasive materials, you need a system with strength that lasts. Strength that only ARO piston pumps can deliver.

2-Ball piston pumps are designed for uniform, consistent fluid delivery

Simply Versatile

2-ball pumps are among the most versatile in the ARO® line. They are capable of handling applications from simple transfer to the extrusion of low- to medium- viscosity materials up to 100,000 centipoise (cPs) with fluiddelivery up to 18.1 gpm (68.6 L/min).

Applications advantages of 2-balls:

Spray: 2-ball pumps deliver the constant flow rate and pressure needed for consistent and efficient coatings applications

Transfer & Supply: Depending on the model, some 2-ball pumps can achieve flow rates up to 18.1 gpm, providing dependable results

Extrusion: 2-ball pumps generate pressure to handle extrusion of low-to medium-viscosity

Lubrication:  ARO® 2-balls pumps provide the flow and pressure required for moving light oils and grease lubricants

High Pressure Cleaning: ARO® Wash Pump Packages are portable and offer the the right option for industrial cleaning applications

Materials advantages of 2-balls:

Sheer sensitive materials: ARO® 2-ball pumps minimize friction points to prevent fluid shear, when compared to other similar pumping technologies

Abrasive and high solid materials: ARO 2-ball pump minimize friction points and allow smooth passages to maintain fluid consistency

Optional ceramic coating provides an even stronger defense against highly abrasive fluids like paint and glass-filled sealer

4-Ball piston pumps are ideal solutions for higher volume and recirculation applications.

Simply Versatile

4-ball piston pumps are designed to transfer high volumes of low- and medium-viscosity fluids up to 12,500 cPs with fluid delivery up to 32.8 gpm (124.0 L/min), depending on the application.

Transfer & Supply: 4-ball pumps are designed to move high volumes of fluids in order to minimize time spent filling or emptying large containers. The high flow rate is a result of double-acting pumping (draws material on the upstroke and downstroke) and is ideal for the large scale transfer of paints, chemicals, varnishes, enamels, lacquers and other low-to-medium viscosity fluids.

Another common application: circulating fluid from the original container to the point of use, and then back.

Materials advantages of 4-balls:

Corrosive materials: ARO® 4-ball lower ends are constructed with stainless steel to maximize fluid compatibility with water based materials.

Abrasive materials: A ceramic coating on the plunger rod and cylinder tube come standard on ARO® 4-ball pumps, protecting from highly abrasive materials and extending the service life of these parts up to 2x.

Chop-Check pumps are designed to move medium-to-high viscosity fluids in difficult applications.

Simply Versatile

The heavy-hitters of the line, ARO® chop-check pumps are designed to move medium-to high- viscosity fluids ranging from 15,000 cPs to more than 1,000,000 cPs, and at delivery rates up to 12.2 gpm (46.3 L/min).

Extrusion: Chop-Check pumps are workhorses - able to deliver viscous adhesives, sealants, caulking and heavy lubricants to the application point at the desired pressure and flow rate.

Transfer & Supply: Depending on the model, some Chop-Check pumps are capable of achieving flow rates up to 12.2 gpm, ensuring efficient transfer of heavy greases, inks and other viscous materials.

Materials advantages of Chop-Checks:

Chop-Check pumps have a primer piston to ‘feed’ material into the pump. They are often used in conjunction with a follower plate to force material down. This prevents loss of prime due to Channeling - a condition where the pump inlet becomes surrounded by air due to the cavity formed when a high viscosity material does not level off when removed (Slumpability).

Stringy materials: Chop-Checks have mechanical, flat checks to overcomematerials that resist cut-off and adhereto themselves.

Tacky materials: Chop-Check pump optionscome in high ratios to generate the fluidpressure needed to move materials that willadhere to hoses and equipment while wet.

Proper fluid management is about more than just a reliable pump–it requires a comprehensive system of integrated parts working together to enhance your productivity and output. To simplify the ordering process, ARO offers complete hydraulic & pneumatic piston pump packages that provide the right configuration of air motor, piston pimp, mount, follow, controls, and down-stream accessories for your specific application.

An extensive array of piston pump parts and accessories are available through Authentic ARO Parts. Simple, cost-effective, and built directly by ARO these accessories offer a way to run your extrusion or finishing application with precise efficiency and accuracy.

Whether your application requires a 2-ball, 4-ball, or chop-check piston pump, with ARO, you get a pump that’s better engineered, from the inside out. We offer design features and performance enhancements that ensure your pump is as durable and dependable as possible.

With an ARO piston pump, you can be assured of getting reliable equipment with the least amount of effort, calculation, or hassle on your part. There are basically four factors to take into consideration when selecting the right configuration for your application:

Type of fluid Viscosity or thickness Required flow rate Required output pressure

Knowing fluid viscosity and flow rates makes it easier to choose the right pump for your application. Not sure what size air motor or piston pump you need? No problem. You can find out by contacting ARO Technical Support and working directly with our expert pump consultants to be sure you get a motor and pump package that operates efficiently, reliably, and safely.

Automotive Engine Valve Recession This page intentionally left blank ENGINEERING RESEARCH SERIES Automotive Engine Valve Recession R Lewis and R S Dwyer-Joyce Series Editor Duncan Dowson Professional Engineering Publishing Limited, London and Bury St Edmunds, UK First published 2002 This publication is copyright under the Berne Convention and the International Copyright Convention. All rights reserved. Apart from any fair dealing for the purpose of private study, research, criticism, or review, as permitted under the Copyright Designs and Patents Act 1988, no part may be reproduced, stored in a retrieval system, or transmitted in any form or by any means, electronic, electrical, chemical, mechanical, photocopying, recording or otherwise, without the prior permission of the copyright owners. Unlicensed multiple copying of this publication is illegal. Inquiries should be addressed to: The Publishing Editor, Professional Engineering Publishing Limited, Northgate Avenue, Bury St Edmunds, Suffolk, IP32 6BW, UK. Fax: +44 (0)1284 705271. © R Lewis and R S Dwyer-Joyce ISBN 1 86058 358 X ISSN 1468-3938 ERS 8 A CIP catalogue record for this book is available from the British Library. Printed and bound in Great Britain by The Cromwell Press Limited, Wiltshire, UK. The publishers are not responsible for any statement made in this publication. Data, discussion, and conclusions developed by the Authors are for information only and are not intended for use without independent substantiating investigation on the part of the potential users. Opinions expressed are those of the Authors and are not necessarily those of the Institution of Mechanical Engineers or its publishers. About the Authors Before going to university, Dr Roger Lewis worked for a year at the Royal Naval Engineering College in Plymouth, UK. He then studied for his MEng in Mechanical Engineering at the University of Sheffield between 1992 and 1996. During this time he was sponsored by the Ministry of Defence. Dr Lewis went on to do his PhD at the University of Sheffield (1996–1999) as part of the Tribology Research Group. His research involved the investigation of wear of diesel engine valves and seat inserts. This work was funded by the Ford Motor Company. Dr Lewis is now a research associate at the University of Sheffield. He is currently working on railway wheel wear as part of a European project concerned with the design of a new hybrid wheel. He is also involved in a Unilever-funded project to investigate the interaction of abrasive particles and toothbrush filaments in a teeth-cleaning contact. Professor Rob S Dwyer-Joyce is senior lecturer in the Department of Mechanical Engineering, University of Sheffield, UK. He graduated in 1993 with a PhD from Imperial College, London, where he studied the wear of rolling bearings and the effects of lubricant contamination. Before this, he worked for British Gas Exploration and Production. Professor Dwyer-Joyce’s research covers a range of tribology projects. His research group has pioneered the use of ultrasound to look at dry and lubricated engineering contacts, studied the way contaminated oil limits component life, quantified how surface damage effects railway track, and investigated aspects of automotive engine wear. He also teaches a course on the Tribology of Machine Elements to undergraduate students. Related Titles IMechE Engineers’ Data Book – Second Edition Design Techniques for Engine Manifolds – Wave Action Methods for IC Engines Theory of Engine Manifold Design – Wave Action Methods for IC Engines Statistics for Engine Optimization International Journal of Engine Research Journal of Automobile Engineering C Matthews ISBN 1 86058 248 6 D E Winterbone and R Pearson ISBN 1 86058 179 X D E Winterbone and R Pearson ISBN 1 86058 209 5 Eds S P Edwards, D M Grove, and H P Wynn ISBN 1 86058 201 X ISSN 1468/0874 Part D of the Proceedings of the IMechE ISSN 0954–4070 Other titles in the Engineering Research Series Industrial Application of Environmentally Conscious Design (ERS 1) Surface Inspection Techniques – Using the Integration of Innovative Machine Vision and Graphical Modelling Techniques (ERS 2) Laser Modification of the Wettability Characteristics of Engineering Materials (ERS 3) Fatigue and Fracture Mechanics of Offshore Structures (ERS 4) Adaptive Neural Control of Walking Robots (ERS 5) Strategies for Collective Minimalist Mobile Robots (ERS 6) T C McAloone ISBN 1 86058 239 7 ISSN 1468–3938 M L Smith ISBN 1 86058 292 3 ISSN 1468–3938 J Lawrence and L Li ISBN 1 86058 293 1 ISSN 1468–3938 L S Etube ISBN 1 86058 312 1 ISSN 1468–3938 ISBN 1 86058 294 X ISSN 1468–3938 ISBN 1 86058 318 0 ISSN 1468–3938 Tribological Analysis and Design of a Modern Automobile Cam Follower (ERS 7) G Zhu and C M Taylor M J Randall C Melhuish ISBN 1 86058 203 6 ISSN 1468–3938 For the full range of titles published by Professional Engineering Publishing contact: Sales Department Professional Engineering Publishing Limited Northgate Avenue Bury St Edmunds Suffolk, IP32 6BW UK Tel: +44 (0)1284 724384 Fax: +44 (0)1284 718692 www.pepublishing.com Contents Series Editor’s Foreword Authors’ Preface Notation Chapter 1 1.1 1.2 1.3 1.4 xi xiii xv Introduction Valves and seats Valve failure concerns Layout of the book References 1 1 1 3 5 Chapter 2 Valve Operation and Design 2.1 Valve operation 2.1.1 Function 2.1.2 Operating systems 2.1.3 Dynamics 2.1.4 Operating stresses 2.1.5 Temperatures 2.2 Valve design 2.2.1 Poppet valve design 2.2.2 Materials 2.3 References 7 7 7 8 9 12 13 15 15 17 19 Chapter 3 Valve Failure 3.1 Introduction 3.2 Valve recession 3.2.1 Causes of valve recession 3.2.2 Wear characterization 3.2.3 Reduction of recession 3.3 Guttering 3.4 Torching 3.5 Effect of engine operating parameters 3.5.1 Temperature 3.5.2 Lubrication 3.5.3 Deposits 3.5.4 Rotation 3.6 Summary 3.7 References 21 21 21 22 26 28 29 29 31 31 34 34 34 36 36 Automotive Engine Valve Recession Chapter 4 Analysis of Failed Components 4.1 Introduction 4.2 Valve and seat insert evaluation 4.2.1 Specimen details 4.2.2 Profile traces 4.2.3 Visual rating 4.3 Lacquer formation on inlet valves 4.3.1 Valve evaluation 4.3.2 Discussion 4.4 Failure of seat inserts in a 1.8 litre, DI, diesel engine 4.4.1 Inlet seat insert wear 4.4.2 Deposits 4.4.3 Misalignment of seat insert relative to valve guide 4.4.4 Inlet valve wear 4.5 Conclusions 4.6 References 39 39 39 39 40 42 45 45 46 47 48 50 51 51 52 53 Chapter 5 5.1 5.2 5.3 5.4 5.5 Valve and Seat Wear Testing Apparatus Introduction Requirements Wear test methods Extant valve and seat wear test rigs University of Sheffield valve seat test apparatus 5.5.1 Hydraulic loading apparatus 5.5.1.1 Design 5.5.1.2 Test methodologies 5.5.1.3 Experimental parameters 5.5.2 Motorized cylinder head 5.5.2.1 Design 5.5.2.2 Operation 5.5.3 Evaluation of dynamics and loading 5.5.3.1 1.8 litre, IDI, diesel engine 5.5.3.2 Hydraulic test machine 5.6 References 55 55 55 56 56 59 60 60 64 66 67 67 69 69 70 74 79 Chapter 6 Experimental Studies on Valve Wear 6.1 Introduction 6.2 Investigation of wear mechanisms 6.2.1 Experimental details 6.2.1.1 Specimen details 6.2.1.2 Test methodologies 6.2.1.3 Wear evaluation 6.2.2 Results 6.2.2.1 Appearance of worn surfaces 6.2.2.2 Formation of wear scars 6.2.2.3 Comparison with engine recession data 81 81 81 81 81 82 84 84 84 88 92 viii Contents 6.2.2.4 Lubrication of valve/seat interface 6.2.2.5 Misalignment of valve relative to seat 6.2.2.6 Effect of combustion load 6.2.2.7 Effect of closing velocity 6.2.2.8 Valve rotation 6.2.2.9 Effect of temperature 6.3 Seat insert materials 6.3.1 Experimental details 6.3.1.1 Valve and seat insert materials 6.3.1.2 Specimen details 6.3.1.3 Test methodologies 6.3.2 Results 6.4 Conclusions 6.5 References 93 94 96 97 100 102 103 104 104 105 105 106 111 111 Chapter 7 Design Tools for Prediction of Valve Recession and Solving Valve Failure Problems 7.1 Introduction 7.2 Valve recession model 7.2.1 Review of extant valve wear models 7.2.2 Development of the model 7.2.2.1 Frictional sliding 7.2.2.2 Impact 7.2.2.3 Final model 7.2.3 Implementation of the model 7.2.4 Model validation 7.2.4.1 Engine tests 7.2.4.2 Bench tests 7.3 Reducing valve recession by design 7.4 Solving valve/seat failure problems 7.5 References 113 113 113 113 115 115 120 123 124 127 127 129 132 132 136 Index 137 ix This page intentionally left blank Series Editor’s Foreword The nature of engineering research is such that many readers of papers in learned society journals wish to know more about the full story and background to the work reported. In some disciplines this is accommodated when the thesis or engineering report is published in monograph form – describing the research in much more complete form than is possible in journal papers. The Engineering Research Series offers this opportunity to engineers in universities and industry and will thus disseminate wider accounts of engineering research progress than are currently available. The volumes will supplement and not compete with the publication of peerreviewed papers in journals. Factors to be considered in the selection of items for the Series include the intrinsic quality of the volume, its comprehensive nature, the novelty of the subject, potential applications, and the relevance to the wider engineering community. Selection of volumes for publication will be based mainly upon one of the following: single higher degree theses; a series of theses on a particular engineering topic; submissions for higher doctorates; reports to sponsors of research; or comprehensive industrial research reports. It is usual for university engineering research groups to undertake research on problems reflecting their expertise over several years. In such cases it may be appropriate to produce a comprehensive, but selective, account of the development of understanding and knowledge on the topic in a specially prepared single volume. Volumes have already been published under the following titles: ERS1 ERS2 ERS3 ERS4 ERS5 ERS6 ERS7 Industrial Application of Environmentally Conscious Design Surface Inspection Techniques Laser Modification of the Wettability Characteristics of Engineering Materials Fatigue and Fracture Mechanics of Offshore Structures Adaptive Neural Control of Walking Robots Strategies for Collective Minimalist Mobile Robots Tribological Analysis and Design of a Modern Automobile Cam and Follower Authors are invited to discuss ideas for new volumes with Sheril Leich, Commissioning Editor, Books, Professional Engineering Publishing Limited, or with the Series Editor. Automotive Engine Valve Recession The present volume, which is the eighth to be published in the Series, comes from the University of Sheffield and is entitled: Automotive Engine Valve Recession by Dr R. Lewis and Dr R. S. Dwyer-Joyce The University of Sheffield This volume follows closely the topic of the previous volume on Automobile Cams and Followers from the University of Leeds. The coincidence of successive volumes devoted to the topic of valves in automotive engines is a clear indication of current interest in these vital engineering components. In this volume the authors outline the essential features of valve operation and the potentially serious problems associated with wear and valve recession in automobile engines since the introduction of lead-replacement and low-sulphur fuels. The authors then outline an experimental study of valve wear and the development of design tools carried out in the Department of Mechanical Engineering in the University of Sheffield. The control of gas flow into and out of engine cylinders still presents a major challenge to the tribologist. The authors consider the fundamental nature of contact and wear between valves and valve seats and they outline the development of a special apparatus for the simulation of engine operating conditions. Valve wear and its effect upon engine performance will continue to be of concern for some time to come. This latest volume represents a valuable addition to the Engineering Research Series. It will be of particular interest to students of wear, designers and manufacturers of reciprocating engines, valve train specialists, and tribologists. Professor Duncan Dowson Series Editor Engineering Research Series xii Authors’ Preface Valve wear has been a serious problem to engine designers and manufacturers for many years. Although new valve materials and production techniques are constantly being developed, these advances have been outpaced by demands for increased engine performance. The drive for reduced oil consumption and exhaust emissions, the phasing out of leaded petrol, reductions in the sulphur content of diesel fuel, and the introduction of alternative fuels such as gas all have implications for valve and seat insert wear. This book aims to provide the reader with a complete understanding of valve recession, starting with a brief introduction to valve operation, design, and operating conditions such as loading and temperature. A detailed overview of work carried out previously, looking at valve and seat wear, is then given and valve and seat failure case studies are discussed. A closer look is then taken at work carried out at the University of Sheffield, UK, including the development of purpose-built test apparatus capable of providing a simulation of the wear of valves and seats used in automotive engines. Experimental investigations are carried out to identify the fundamental valve and seat wear mechanisms, the effect of engine operating parameters on wear, and to rank potential new seat materials. An important aspect of research is the industrial implementation of the results and the provision of suitable design tools. A design procedure is outlined, which encapsulates the review of literature, analysis of failed specimens, and bench test work. This includes a semi-empirical model for predicting valve recession run in an iterative software programme called RECESS, as well as flow charts to be used to reduce the likelihood of recession occurring during the design process and to offer solutions to problems that do occur. R Lewis and R S Dwyer-Joyce The University of Sheffield, UK This page intentionally left blank Notation Unless otherwise stated the notation used is as follows: A Wear area (m2) b Valve disc thickness (m) Initial valve clearance (m) ci e Valve energy (J) E Modulus of elasticity (N/m2) f Actuator sinusoidal displacement cycle frequency (Hz) h Penetration hardness (N/m2) k Sliding wear coefficient K Empirically determined impact wear constant l Valve lift (m) Actuator lift (m) la L Initial actuator displacement (m) m Mass of valve + mass of follower + half mass of valve spring (kg) n Empirically determined impact wear constant N Number of cycles Peak combustion pressure (N/m2) pp Contact force at valve/seat interface (N) Pc Peak combustion load (N) Pp r Recession (m) RT Room temperature (°C) Seat insert radius as specified in part drawing (m) Rd Initial seat insert radius (m) Ri Valve head radius (m) Rv s Wear scar width (m) t Time (seconds) v Valve velocity (m/s) Actuator velocity (m/s) va V Wear volume (m3) w Seat insert seating face width (m) Initial seat insert seating face width (as measured) (m) wi Seat insert seating face width as specified in part drawing (m) wd W Wear mass (kg) Work done on valve during combustion in the cylinder (J) Wv x Sliding distance (m) y Vertical deflection of valve head under combustion pressure (m) Greek characters α Actuator sinusoidal displacement cycle amplitude (m) β Difference between valve and seat insert seating face angles (°) δ Slip at the valve/seat insert interface (m) θ Camshaft rotation (°) xv Automotive Engine Valve Recession θs θv ν ω xvi Seat insert seating face angle (°) Valve seating face angle (°) Poisson’s ratio Camshaft rotational speed (r/min) Chapter 1 Introduction 1.1 Valves and seats Valves (shown in-situ in an engine in Fig. 1.1) are used to control gas flow to and from cylinders in automotive internal combustion engines. The most common type of valve used is the poppet valve (shown in Fig. 1.2 with its immediate attachments). The valve itself consists of a disc-shaped head having a stem extending from its centre at one side. The edge of the head on the side nearest the stem is accurately ground at an angle – usually 45 degrees, but sometimes 30 degrees, to form the seating face. When the valve is closed, the face is pressed in contact with a similarly ground seat. It is the contact conditions and loading at this interface that will have the largest influence on the rate at which valve and seat wear will occur, so understanding these is a key in determining the mechanisms that cause valve recession. Fig. 1.1 Overhead camshaft valve drive 1.2 Valve failure concerns Valve wear has been a small but serious problem to engine designers and manufacturers for many years. It has been described as ‘One of the most perplexing wear problems in internal combustion engines’ [1]. 1 Automotive Engine Valve Recession VALVE TIP VALVE SPRING RETAINER (KEEPER) GROOVE(S) SEATING FACE ANGLE VALVE GUIDE SE AT IN G VALVE SEAT INSERT SEATING FACE WIDTH FA CE W ID TH VALVE STEM STEM-BLEND FILLET AREA VALVE HEAD SEATING FACE ANGLE HEAD DIAMETER Fig. 1.2 Valve and seat insert Although new valve materials and production techniques are constantly being developed, these advances have been outpaced by demands for increased engine performance. These demands include: ● higher horsepower-to-weight ratio; ● lower specific fuel consumption; ● environmental considerations such as emissions reduction; ● extended durability (increased time between servicing). The drive for reduced oil consumption and exhaust emissions has led to a reduction in the amount of lubricant present in the air stream in automotive diesel engines, and the effort to lengthen service intervals has resulted in an increasingly contaminated lubricant. These changes have led to an increase in the wear of inlet valves and seat inserts. Lead, originally added to petrol to increase the octane number, was found to form compounds during combustion that proved to be excellent lubricants, significantly reducing valve and seat wear. Leaded petrol, however, has now been phased out in the UK (since the end of 1999). As an alternative, lead replacement petrol (LRP) has been developed. This contains anti-wear additives based on alkali metals such as phosphorus, sodium, and potassium. Results of tests run using LRP containing such additives, however, have shown that, as yet, lead is unchallenged in providing the best protection. In several countries where LRP has already been introduced, a high 2 Introduction incidence of exhaust valve burn has been recorded. In Sweden the occurrence of valve burn problems has increased by 500 per cent since LRP was introduced in 1992 [2]. The suspected cause of the valve burn problems is incomplete valve-to-valve seat sealing as a result of valve seat recession (VSR). The occurrence of VSR is blamed on hot corrosion – an accelerated attack of protective oxide films that occurs in combustion environments where low levels of alkali and/or other trace elements are present. A wide range of high-temperature alloys are susceptible to hot corrosion, including nickel- and cobalt-based alloys, which are used extensively as exhaust valve materials or as wear-resistant coatings on valves or seats. Materials used for engine components have always been designed to resist corrosive attack by lead salts. No such development has taken place to form materials resistant to alkali metals or other additive chemistries. It is clear that LRP will not provide an immediate solution to the valve wear issue, which is likely to cause tension between car manufacturers and owners for some time to come. The impending reductions in the sulphur content of diesel fuel and the introduction of alternative fuels, such as gas, will also have implications for valve and seat insert wear. Dynamometer engine testing is often employed to investigate valve wear problems. This is expensive and time consuming, and does not necessarily help in finding the actual cause of the wear. Valve wear involves so many variables that it is impossible to confirm precise, individual quantitative evaluations of all of them during such testing. In addition, the understanding of wear mechanisms is complicated by inconsistent patterns of valve failure. For example, failure may occur in only a single valve operating in a multi-valve cylinder. Furthermore, the apparent mode of failure may vary from one valve to another in the same cylinder or between cylinders in the same engine. An example of such inconsistency is shown in Fig. 1.3. This illustrates exhaust valve recession values for four cylinders in the same engine (measurements taken on the cylinder head). The valve in cylinder 1 has recessed to the point where pressure is being lost from the cylinder, while the other valves have hardly recessed at all. No hard and fast rules have been established to arrive at a satisfactory valve life. Each case, therefore, has to be painstakingly investigated, the cause or causes of the problem isolated, and remedial action taken. In order to analyse the wear mechanisms in detail and isolate the critical operating conditions, simulation of the valve wear process must be used. This has the added benefit of being cost effective and saving time. Based on the wear patterns observed, the fundamental mechanisms of valve wear can be determined. Once the fundamental mechanisms are understood, a viable model of valve wear can be developed that will speed up the solution of future valve wear problems and assist in the design of new engines. 1.3 Layout of the book Chapters 2 and 3 are review chapters outlining valve function, different operating systems, the operating environment, and valve design and materials. Valve failure is also examined in detail, and work on likely wear mechanisms and the effect of engine operating parameters are described. 3 Automotive Engine Valve Recession Fig. 1.3 Recession measurements for exhaust valves in the same 2.5 litre diesel engine cylinder head Chapter 4 details the evaluation of failed valves and seat inserts from tests run on automotive engines. This includes the validation of test rig results, the establishment of techniques for the evaluation of test rig results, and the provision of information on possible causes of valve recession. Chapter 5 outlines experimental apparatus able to simulate the loading environment and contact conditions to which the valve and seat insert are subjected in an engine. Chapter 6 then describes bench test work carried out to investigate the wear mechanisms occurring in valves and seat inserts. This includes studies on the effect of engine operating conditions, the effect of lubrication at the valve/seat insert contact, and the evaluation of potential new seat materials. Finally, Chapter 7 describes the development of design tools that enable the results of the review of literature, analysis of failed specimens, and bench test work to be applied in industry to assess the potential for valve recession and solve problems more quickly. 4 Introduction 1.4 References 1. De Wilde, E.F. (1967) Investigation of engine exhaust valve wear, Wear, 10, 231–244. 2. Barlow, P.L. (1999) The lead ban, lead replacement petrol and the potential for engine damage, Indust. Lubric. Tribol., 51, 128–138. 5 This page intentionally left blank Chapter 2 Valve Operation and Design 2.1 Valve operation 2.1.1 Function The two main types of internal combustion engine are: spark ignition (SI) engines (petrol, gasoline, or gas engines), where the fuel ignition is caused by a spark; and compression ignition (CI) engines (diesel engines), where the rise in pressure and temperature is high enough to ignite the fuel. Valves are used in these engines to control the induction and exhaust processes. Both types of engine can be designed to operate in either two strokes of the piston or four strokes of the piston. The four-stroke operating cycle can be explained by reference to Fig. 2.1. This details the position of the piston and valves during each of the four strokes. INDUCTION COMPRESSION EXPANSION EXHAUST Fig. 2.1 Four-stroke engine cycle 7 Automotive Engine Valve Recession 1. The induction stroke The inlet valve is open. The piston moves down the cylinder drawing in a charge of air. 2. The compression stroke Inlet and exhaust valves are closed. The piston moves up the cylinder. As the piston approaches the top of the cylinder (top dead centre – tdc) ignition occurs. In engines utilizing direct injection (DI) the fuel is injected towards the end of the stroke. 3. The expansion stroke Combustion occurs causing a pressure and temperature rise which pushes the piston down. At the end of the stroke the exhaust valve opens. 4. The exhaust stroke The exhaust valve is still open. The piston moves up forcing exhaust gases out of the cylinder. 2.1.2 Operating systems In engines with overhead valves (OHV), the camshaft is either mounted in the cylinder block, or in the cylinder head with an overhead camshaft (OHC). Figure 2.2 shows an OHV drive in which the valves are driven by the camshaft via cam followers, push rods, and rocker arms. Since the drive to the camshaft is simple (either belt or chain) and the only machining is in the cylinder block, this is a cost-effective arrangement. ROCKER ARM VALVE SPRING RETAINER SPRING COLLET PUSH ROD VALVE SPRING CAM FOLLOWER CAM VALVE GUIDE CAMSHAFT VALVE STEM VALVE HEAD VALVE SEAT INSERT Fig. 2.2 Overhead valve drive 8 Valve Operation and Design In the OHC drive shown in Fig. 2.3 the camshaft is mounted directly over the valve stems. Alternatively it could be offset and the valves operated using rockers. The valve clearance could then be adjusted by altering the pivot height. Once again, the drive to the camshaft is by toothed belt or chain. In the system shown in Fig. 2.3 the camshaft operates on a follower or bucket. The clearance between the valve tip and the follower is adjusted by a shim. This is more difficult to adjust than in systems using rockers, but is less likely to change. The spring retainer is attached to the valve using a tapered split collet. The valve guides are usually press-fitted into the cylinder head, so that they can be replaced when worn. Valve seat inserts are used to ensure minimal wear. The valves rotate in order to ensure even wear and to maintain good seating. This rotation is promoted by having the cam offset from the valve stem axis. This also helps to avoid localized wear on the cam follower. CAM CAMSHAFT CAM FOLLOWER VALVE SPRING RETAINER SHIM SPRING COLLET VALVE SPRING VALVE GUIDE VALVE STEM VALVE HEAD VALVE SEAT INSERT Fig. 2.3 Overhead camshaft valve drive 2.1.3 Dynamics The geometry of the cam and its follower defines the theoretical valve motion. The actual valve motion is modified because of the finite mass and stiffness of the elements in the valve train. 9 Automotive Engine Valve Recession 3 Spring-controlled movement 4 5 Constant velocity 2 Sinusoidal deceleration 1 Sinusoidal acceleration STAGE: Constant velocity Figure 2.4 [1] shows theoretical valve lift, velocity, and acceleration. The motion can be split into five stages. Lift Valve-period Max. valve lift Camshaft rotation Velocity Camshaft rotation Sinusoidal acceleration Sinusoidal deceleration Acceleration Camshaft rotation Spring-controlled deceleration Spring-controlled acceleration Fig. 2.4 Theoretical valve motion [1] 10 Valve Operation and Design 1. Before the valve starts to move, the clearance between the follower and the valve tip has to be taken up. This clearance ensures the valve can seat under all operating conditions and allows for bedding-in of the valve. The cam is designed to give an initially constant velocity to control the impact stresses as the clearance is taken up. The impact velocity is typically limited to 500 mm/s at the rated engine speed. 2. During the next stage the cam accelerates the valve. Rather than designing the cam to give the valve a constant acceleration (which would lead to shock loadings), sinusoidal or polynomial functions, which cause the acceleration to rise from zero to a maximum and then fall back to zero, are used. 3. Deceleration is controlled by the valve spring as the valve approaches maximum lift. As the valve starts to close, acceleration is also controlled in this way. 4. Final deceleration is controlled by the cam. 5. The cam is designed to give a constant closing velocity in order to limit impact stresses. Actual valve motion is modified by the elasticity of the components in the valve train; a simple model is shown in Fig. 2.5. CAM COMBINED STIFFNESS OF VALVE GEAR VALVE MASS VALVE SPRING Fig. 2.5 A simple valve gear model A comparison of actual and theoretical valve motion [1] is shown in Fig. 2.6. Valve bounce can occur if the impact velocity is too high or if the valve spring preload is too low. 11 Automotive Engine Valve Recession Theoretical valve motion Actual valve motion Valve lift Valve bounce Fig. 2.6 A comparison of theoretical and actual valve motion [1] 2.1.4 Operating stresses During each combustion event, high stresses are imposed on the combustion chamber side of the valve head. These generate cyclic stresses peaking above 200 MN/m2 on the port side of the valve head, as shown in Fig. 2.7 [2]. The magnitude of the stresses is a function of peak combustion pressure. The stresses are much higher in a compression ignition engine than a spark ignition engine. 200 Stress (MPa) 400 Fig. 2.7 Tensile stresses on the surface of the port side of the valve head due to combustion loading [2] 12 Valve Operation and Design As the valve impacts the seat insert, cyclic stresses are imposed at the junction of the valve stem and fillet. If thermal distortion of the cylinder head has caused misalignment of the valve relative to the seat insert, seating will occur at a single contact point (as shown in Fig. 2.8). Bending stresses as a result of this point contact increase the magnitude of the valve seating stresses. Ideal Seating Off-Centre Seating Due to Valve and Seat Misalignment Fig. 2.8 Off-centre seating due to valve misalignment 2.1.5 Temperatures A typical inlet valve temperature distribution is shown in Fig. 2.9 [3]. It was not made clear whether these were experimental or theoretical values or whether the valve was from a diesel or gasoline engine. The asymmetric distribution may have been due to non-uniform cooling or deposit build-up affecting heat transfer from the valve head. As shown in Fig. 2.10 [3], exhaust valve temperatures are much higher. Although both inlet and exhaust valves receive heat from combustion, the inlet valve is cooled by incoming air, whereas the exhaust valve experiences a rapid rise in temperature in the valve head, seat insert, and underhead area from hot exhaust gases. As much as 75–80 per cent of heat input to a solid valve exits via contact with the seat insert [4]. The remainder is conducted through the valve stem into the valve guide. Effective heat transfer to the seat insert and into the cylinder head is, therefore, essential. Figure 2.11 [3] shows the thermal gradient existing from the centre of the valve head to the cooling water in the cylinder head. It clearly indicates the large temperature differential at the seating interface. Heat transfer can be affected by valve bounce. However, the effect of seating deposits is much more significant. If deposits are allowed to build up, they may not only lead to an increase in valve temperature, but may also break away locally and create a leakage path, leading to valve guttering and, possibly, torching (see Sections 3.3 and 3.4). 13 Automotive Engine Valve Recession (a) BELOW 200 200 - 300 300 - 400 400 - 500 500 - 600 320OC 500 OC 520OC (b) a axial section; b top of head Fig. 2.9 Typical inlet valve temperature distribution [3] (a) A A G F E H I J B C D L K M N O N M L G G K H I J BELOW 450 B 450-500 C 500-525 D 525-550 E 550-575 F 575-600 G 600-625 H 625-650 I 650-675 J 675-700 K 700-725 L 725-750 M 750-775 N 775-800 O 800-825 J I H (b) G H I J K a axial section; b top of head Fig. 2.10 Typical exhaust valve temperature distribution [3] 14 Valve Operation and Design 700 VALVE Temperature ( oC) 600 500 400 INSERT (Fe BASE) 300 200 INSERT (Cu BASE) CASTING COOLANT 100 10 20 30 Distance From Valve Centre (mm) Fig. 2.11 Thermal gradient from valve centre to coolant [3] 2.2 Valve design The most commonly used valve is the poppet valve. It has several advantages over rotary and disc valves [1]: it is cheap, has good flow properties, good seating, easy lubrication, and good heat transfer to the cylinder head. 2.2.1 Poppet valve design A number of different poppet valve designs are used, as shown in Fig. 2.12 and outlined in Table 2.1. The final choice usually depends on the performance and cost objectives. Inlet valves are usually constructed using the one-piece design, whereas exhaust valves are generally constructed using a two-piece design. Valve seats formed in-situ within the cast iron cylinder heads were originally used in passenger car engines. In order to provide greater wear resistance, hardfacing alloys were developed as well as seat inserts designed to be press-fitted into the cylinder heads. An inlet valve seat insert has been designed that is shaped to induce a swirling motion 15 Automotive Engine Valve Recession WELD INTERNAL CAVITY SEAT FACE WELD One-Piece Two-Piece Wafer or Tip Welded Seat Welded COOLANT Internally Cooled Fig. 2.12 Typical valve designs Table 2.1 Typical valve designs Design Description One-piece This is the most cost-effective design. It is widely used in passenger car applications. Wafer or tip welded To eliminate tip wear or scuffing in one-piece austenitic valves, it is possible to weld on a hardened martensitic steel tip. Two-piece In a two-piece design an austenitic head is welded to a hardened martensitic stem. This increases both valve tip and stem scuffing resistance. Seat welded To increase the wear and/or corrosion resistance of the valve seating face it is possible to apply hardfacing alloys using gas or shielded-arc techniques. Internally cooled Internally cooled valves contain a cavity partially filled with a coolant, usually sodium, which dissipates heat from the valve head through the stem and valve guide to the cylinder head. This reduces the valve head temperature significantly. in the fuel–air mix as it passes through the valve [5]. This swirling motion improves the mixing of the two components and enhances combustion. The seating faces of both valves and seat inserts are usually ground to an angle of 45 degrees. The seat insert seating face width is narrower than that of the valve to reduce the risk of trapping combustion particles and wear debris in the interface between the two. In a few cases, the valve seating face is ground to an angle about half a degree less than the seating face angle, as shown in Fig. 2.13. There are three reasons for this [6]. 1. The hottest part of the valve, under running conditions, is the underside of the head. The additional expansion of this side makes the two seating face angles equal at running temperatures. 16 Valve Operation and Design 2. When the valve is very hot the spring load can cause the head to dish slightly, which can lift the inner edge of the valve seating face (nearest the combustion chamber) clear of the seat insert if the angles are the same when cold. 3. The risk of trapping combustion particles between the two seating faces is reduced. 44.5O 45O Fig. 2.13 Valve seating face angle different from seat insert seating face angle (exaggerated) [6] 2.2.2 Materials Most inlet valves are manufactured from a hardened, martensitic, low-alloy steel. These provide good strength and wear and oxidation resistance at higher temperatures. Exhaust valves are subjected to high temperatures, thermal stresses, and corrosive gases. Most exhaust valves are manufactured from austenitic stainless steels. These can be iron, or nickel, based. Solid solution and precipitation strengthening provide the hot hardness and creep resistance required for typical exhaust valve applications. The 21.4N composition is widely used in diesel engine exhaust valve applications. This alloy has an excellent balance of hot strength, corrosion resistance, creep resistance, fatigue resistance, and wear properties at an acceptable cost [3]. In heavy-duty diesel engine applications higher strengths and creep resistances are attained by using superalloys as valve materials. Valve seating face wear and corrosion can be reduced by applying seat facing materials. Stellite facings are commonly used for passenger car applications. Typical compositions of martensitic and austenitic steels, superalloys, and seat facing materials used in valve applications are shown in Table 2.2. Engine test work carried out using ceramic valves has shown a significant reduction in valve recession compared to results achieved with metal valves [7, 8]. The reduction in mass of ceramic valves results in improved seating dynamics, reducing seating forces and eliminating valve bounce. In addition, the high stiffness of ceramics helps resist flexing of the valve head, reducing sliding between the valve and seat. 17 Automotive Engine Valve Recession Table 2.2 Compositions of typical valve and seat materials [2, 3] Nominal compositions of martensitic valve materials (weight %) Designation C Mn Si Cr SAE 1541 SAE 1547 SAE 3140 SAE 4140 Silchrome 1 Sil XB 0.40 0.47 0.40 0.40 0.45 0.80 1.50 1.50 0.80 0.88 0.80 max 0.80 max 0.23 0.23 0.28 0.28 3.25 2.12 422 SS 0.22 0.75 0.50 max 11.75 – – 0.65 0.95 8.50 20.00 Ni Mo Fe – – 1.25 – 0.50 max 1.35 – – – 0.20 – – Bal. Bal. Bal. Bal. Bal. Bal. 0.75 1.08 Bal. Other W V W V 1.08 0.25 18.00 1.00 Nominal compositions of austenitic valve materials (weight %) Designation C Mn Si Cr Ni 21.2N 21.4N 21.12 23.8N Silchrome 10 Gaman H XCR YXCR TPA 0.55 0.53 0.20 0.33 0.38 8.25 9.00 1.50 2.50 1.05 0.25 max 0.25 max 1.00 0.75 3.00 20.38 21.00 21.25 23.00 19.00 0.52 0.45 0.40 0.45 12.25 0.50 0.80 0.60 2.65 0.50 0.80 0.60 21.25 23.50 24.00 14.00 N Fe 2.12 3.88 11.50 8.00 8.00 0.30 0.46 – 0.32 – Bal. Bal. Bal. Bal. Bal. – 4.80 3.80 14.00 0.45 – – – Bal. Bal. Bal. Bal. Other Mo Mo W Mo 1.08 1.40 2.40 0.35 Nominal compositions of superalloy valve materials (weight %) Designation C Mn Si Ni Fe Others N-155 1.50 1.00 max 21.25 19.50 Bal. 2.25 0.08 16.0 0.50 max 0.50 max 15.50 Bal. Bal. 6.50 7.00 0.10 max 1.00 max 1.00 max 19.50 Bal. 3.0 max 0.04 56.50 Bal. Co 19.75 Mo 3.50 W 2.50 Nb 1.00 Ti 3.05 Ti 2.30 Al 1.22 Nb+Ta 0.95 Ti 2.25 Al 1.40 Ti 2.25 Al 1.25 Mo 2.00 Nb 0.85 0.12 TPM 0.04 Inconel 751 0.06 Nimonic 80A Pyromet 31 Cr 0.20 max 0.20 max 22.60 Nominal compositions of typical valve seat facing materials (weight %) Designation C Mn Si Cr Stellite 6 Stellite F Stellite 1 Eatonite Eatonite 3 Eatonite 6 VMS 585 0.50 0.30 0.50 0.50 0.50 0.75 – 1.20 1.00 1.30 1.00 1.20 1.30 1.00 28.00 25.00 30.00 29.00 29.00 28.00 24.00 18 1.20 1.75 2.50 2.40 2.00 1.75 2.25 Ni 3.00 22.00 1.50 Bal. Bal. 16.50 11.00 Co Bal. Bal. Bal. 10.00 – – – W Mo Fe 4.50 12.00 13.00 15.00 – – – 0.50 – 0.50 – 5.50 4.50 5.50 3.00 3.00 3.00 8.00 8.00 Bal. Bal. Valve Operation and Design Valve seats were formed in-situ within the cast iron cylinder heads in most passenger car engines. These proved inadequate in heavier-duty engines so, in order to provide greater wear resistance, hardfacing alloys were developed as well as seat inserts designed to be press-fitted into the cylinder heads. Nickel- and iron-base alloys are commonly used as hardfacing and insert materials. Sintered seat insert materials incorporating solid lubricants have also been developed to compensate for the reduction of lead and sulphur in fuels [5, 9]. These have found use in gasoline engines, but have been largely unsuccessful in diesel engines because of the higher loads. More recently, work carried out on ceramic seat insert materials has shown that they may offer a solution to valve recession problems [7, 10]. However, while ceramic materials offer unique properties that make them ideally suited for application in passenger car engines, they are relatively brittle, which raises the issue of component reliability. 2.3 References 1. Stone, R. (1992) Introduction to internal combustion engines, Macmillan, Basingstoke. 2. Larson, J.M., Jenkins, L.F., Belmore, J.E., and Narasimhan, S.L. (1987) Engine valves – design and material evolution, Trans ASME J. Engng Gas Turbines Power, 109, 351–361. 3. Beddoes, G.N. (1992) Valve materials and design, Ironmaking and steelmaking, 19, 290–296. 4. Giles, W. (1971) Valve problems with lead free gasoline, SAE Paper 710368. 5. Lane, M.S. and Smith, P. (1982) Developments in sintered valve seat inserts, SAE Paper 820233. 6. Hillier, V.A.W. (1991) Fundamentals of motor vehicle technology, Fourth edition, Stanley Thornes (Publishers) Ltd, Cheltenham. 7. Kalamasz, T.G. and Goth, G. (1988) The application of ceramic materials to internal combustion engines, SAE Paper 881151. 8. Updike, S.H. (1989) A comparison of wear mechanics with ceramic and metal valves in firing engines, SAE Paper 890177. 9. Fujiki, F. and Makoto, K. (1992) New PM seat insert materials for high performance engines, SAE Paper 920570. 10. Woods, M.E. and McNulty, W.D. (1991) Ceramic seats and intermetallic coated valves in a natural gas fired engine, SAE Paper 910951. 19 This page intentionally left blank Chapter 3 Valve Failure 3.1 Introduction Three types of valve failure have been observed: ● valve recession; ● guttering; ● torching. Valve recession, the most common form of wear in diesel engine inlet valves, is caused by loss of material from the seat insert and/or the valve. Guttering is a hightemperature, corrosive process usually caused by deposit flaking. Torching or melting of a valve is triggered by a rapid rise in the temperature of the valve head, which may be caused by preignition or abnormal combustion. Inlet valve wear is a particular problem in diesel engines because the fuel is introduced directly into the cylinder. The inlet valve, therefore, receives no liquid on its seating face and seats under rather dry conditions. Exhaust valve wear is less prominent than inlet valve wear as combustion products deposited on the seating faces provide lubrication. Exhaust valves are more likely to fail due to guttering or torching. Such failures are rarely seen in inlet valves. 3.2 Valve recession Valve recession is said to have occurred if wear of the valve and seat insert contact faces has caused the valve to ‘sink’ or recede into the seat insert, thereby altering the closed position of the valve relative to the cylinder head (as shown in Fig. 3.1). Engines are typically designed to tolerate a certain amount of valve recession. After this has been exceeded, the gap between the valve tip and the follower must be adjusted to ensure that the valve continues to seat correctly. If the valve is not able to seat, cylinder pressure will be lost and the hot combustion gases that leak will cause valve guttering or torching to occur, which will rapidly lead to valve failure. 21 Automotive Engine Valve Recession VALVE GUIDE MEASURE OF VALVE AND SEAT RECESSION VALVE SEAT INSERT VALVE VALVE AND SEAT RECESSED VALVE AND SEAT NORMAL Fig. 3.1 Valve recession 3.2.1 Causes of valve recession Valve recession is caused by loss of material from the seat insert and/or the valve. It occurs gradually over a large number of hours. Sometimes the material loss will be greater from the seat insert, and sometimes the material loss will be greater from the valve. The nature of the material loss is not clearly understood, although it has been suggested that it may occur by the following mechanisms [1]: ● metal abrasion; ● fretting; ● adhesion mechanisms; ● high temperature corrosion. Tauschek and Newton [2] suggested that recession problems were caused by pounding of the valve due to misaligned seating, since seating contact pressure varies inversely with contact area. It was thought that improper seating was caused by cylinder head deformation due to thermal effects. Thermal effects are often associated with nonsymmetric cooling passages in the cylinder head near the valve seat inserts. Work carried out on inlet valves by Zinner [3] began on the assumption that the wear was caused by the ‘hard pounding of the seat by the valve cone’. This led to an initial study of valve motion at increasing engine speeds. However, when the amount of wear under different operating conditions was measured, it became apparent that the effect of mean effective pressure on wear was much greater than that of engine speed. It was, 22 Valve Failure therefore, assumed that the main factor accounting for wear must be sliding friction between the valve and valve seat caused by ‘wedging’ of the valve into the seat under the effect of the gas pressure. The ‘wedging’ action was found to be greater, and hence the sliding motion lengthened, if cylinder head deformation also occurred as a result of uneven cooling. Figure 3.2 diagrammatically represents the effect of thermal distortions in the cylinder head. Figure 3.2(a) shows a valve not subjected to pressure on a non-deformed seat insert in a position flush with the cylinder head. The assumption in Fig. 3.2(b) is that the cylinder head has been deformed due to uneven temperatures and the valve, being under no load, projects slightly from its seat insert and makes only one-sided contact. Figure 3.2(c) shows the valve forced into its seat by the gas pressure, which implies the existence of sliding motion. a) 90O b) c) Fig. 3.2 Deformation of a cylinder head bottom and valve: (a) valve with 45 degree seat angle in plane cylinder head bottom; (b) downward bottom deflection; valve making one-sided contact; (c) bottom deflected upwards and valve disc bent under influence of gas pressure [3] 23 Automotive Engine Valve Recession The presence of sliding motion between the valve and valve seat insert was firmly established by Zinner [3] using a static test rig designed to study the effect of seat insert distortion. Valve stem protrusion was measured at different air pressures applied to the valve head. An increase in valve stem movement resulted in greater sliding at the valve/seat contact area. Work carried out by Marx and Muller [4] on wear of inlet valves in supercharged fourstroke diesel engines started on the supposition that wear was provoked by incorrect closing of the inlet valves (valve bounce). However, it was eventually concluded that wear was caused by small frictional movements between the valve seating face and the seat insert. It was thought that the friction movements were a result of elastic bending of the valve and working face of the cylinder head due to the combustion pressure. It was found that wear was more frequent in higher power engines due to higher engine velocity and, therefore, an increased number of combustion cycles. Lane and Smith [5] studied the force mechanism between a valve and a valve seat insert (as shown in Fig. 3.3). Force P, applied to the seat insert, consists of valve train inertia force due to acceleration when the valve is seating and forces applied by the valve spring and cylinder pressure when the valve is resting on the seat. This force can be resolved into two components – Psinθ parallel to the seating face and Pcosθ perpendicular to the seating face. It was suggested that if the Psinθ component exceeds the shear stress of the seat insert material, plastic shear deformation of the surface material may be induced. This could lead to crack formation and eventually, after repeated loading, to particles of material (asperities) breaking away from the surface. The wear debris may either be blown off by gas flow when the valve opens or remain on the seating surface and: Fig. 3.3 Forces acting on a seat insert seating face [5] 24 Valve Failure 1. prevent complete gas sealing; 2. lead to an increase in valve temperature due to reduced seat contact area (the reduced seat contact area inhibits heat transfer from the valve to the engine coolant through the seat insert, causing possible valve failure); 3. lead to abrasive wear of seating surfaces when the valve slides on the seat insert (if the asperities are hard material). Using a reduced angle for the seating face is often suggested for an approach to reduce surface flow. It was calculated that a reduction in the seating angle from 45 to 30 degrees would reduce the shear force (Psinθ) by 29 per cent. The normal force (Pcosθ), however, would increase by 22 per cent. It was suggested that, if the Pcosθ component exceeds the compressive yield strength of the material, it may fail in a fashion known as ‘hammering’ or ‘brinelling’. Therefore, to compensate for a reduction in the seating angle, material hardness and toughness may need to be increased to withstand increased dynamic loading. Work performed by Pope [6] led to the conclusion that wear of the seating face of air inlet valves was due to relative movement between the valve and its seat when the head flexed under the action of the firing pressure since: 1. the measured valve head stress due to the cylinder firing pressure was about 20 times that due to the valve impact on its seat; 2. a considerable reduction in the valve seating velocity had a negligible effect on seat wear; 3. stiffer valve gear to reduce the possibility of resonant valve gear vibrations produced little effect. Van Dissel et al. [7] concluded that seat recession can occur by the systematic gouging away, or deformation and eventual wearing, of the valve and/or insert material. It was found that the deformation led to the formation of concentric ridges on the seating face of the valve which were described as ‘single wave’ or ‘multiple wave’ formations. It was suggested that the gouging and deformation were generated by the same process, only the severity of the damage was different. It was speculated that the problem was caused by valve misalignment, which resulted in the valve seating face making contact with only a portion of the seat insert seating face. As it was expected that the initial contact would be the most severe, it is at this stage that the surface ridges were thought to be generated. Subsequent ridge formation was thought to occur as the valve bounced to self-centre, driven by the combustion cycles. Valve rotation, resulting in a different portion of the valve seating face impacting the insert for each cycle, was thought to cause the ‘single’ or ‘multiple wave’ appearance. Both Fricke and Allen [8] and Ootani et al. [9], in testing materials for poppet valve applications, assumed that impact of the valve on the seat was the major cause of valve recession. 25 Automotive Engine Valve Recession 3.2.2 Wear characterization Marx and Muller [4] found that after a relatively short running time it was possible to observe concentric rings on the seating faces of inlet valves. Local manifestations also occurred which were reminiscent of sand dunes. These changed into smooth, bright, slightly curved surfaces over time, indicating consistent removal of material. Wear rates of approximately 1 mm/1000 h were observed. Van Dissel et al. [7] observed distinct trenches on the seating faces of recessed valves where seating occurred. Seat recession varied from negligible to 0.56 mm. Worn regions typically exhibited unique topographic features. Pitting, gouging, and indentation of the seating face were all observed. One feature common to most recessed valves was observed. This was a series of ridges and valleys formed circumferentially around the axis of the seating face. In some cases the ridges and valleys were concentric, as waves would appear if a single stone was dropped into a pool. This was described as a ‘single wave formation’ (see Fig. 3.4). In other cases the ridges and valleys overlapped each other, as waves would appear if several stones were dropped into a pool. This was described as a ‘multiple wave formation’ (see Fig. 3.5). Ridge formation was found to vary according to engine type. Fig. 3.4 Inlet valve seating face showing ‘single wave ridge formation’ (Major diameter is towards the bottom) [7] 26 Valve Failure Fig. 3.5 Inlet valve seating face showing ‘multiple wave formation’ (Major diameter is towards the top) [7] Wear of seat inserts ranged from negligible to moderate. The material removal process resulted in a concave seating face configuration. Radial scratches were observed within the concave region, as shown in Fig. 3.6. Fig. 3.6 Pitting of inlet seat insert (Major diameter is towards the top) [7] 27 Automotive Engine Valve Recession Narasimhan and Larson [10] observed reorientation of carbides near the surface of a seat insert seating face, as shown in Fig. 3.7. This provided evidence of sliding contact between the valve and seat insert. Fig. 3.7 Microstructural features of alloys at the valve/seat insert interface: (a) eatonite insert; (b) alloy No. 6 hardfacing [10] 3.2.3 Reduction of recession Zinner [3] concluded that the major cause of valve recession was sliding friction between the valve and seat insert caused by ‘wedging’ of the valve into the seat under the action of the combustion pressure. In order to reduce the effect of the sliding motion, design modifications were introduced. Below is a summary of the design modifications that resulted in a reduction of wear, presented in order of effectiveness: l. decreasing the valve seating face angle; 2. lubrication of the valve seating face; 3. use of valve seat inserts of a suitable material; 4. increased rigidity of valve disc; 5. greater rigidity of valve mechanism; 28 Valve Failure 6. welding material onto valve seats; 7. smaller valve diameter; 8. reducing speed of valve impact; 9. increasing seating face width. Introduction of the 30 degrees seating face angle, more effective cooling of the cylinder head bottom, and greater rigidity of the valve and valve mechanism reduced wear rates by about 50 per cent. It should be noted, however, as mentioned in the previous section, that changes such as reducing the valve seating face angle may not always be possible as they will change flow characteristics and may adversely affect engine performance. In order to reduce the frictional motion, and thereby wear, Marx and Muller [4] reduced the seating face angle from 45 to 30 degrees. They then carried out investigations with over eighty material pairs. They concluded that suitable material pairing and reducing the seat angle was not enough to solve the problem. Tests were then carried out in which a small amount of lubricating oil was introduced shortly before the inlet valve. This was found to be very effective in reducing wear if used in conjunction with the changes previously detailed. Giles [11] used an expression for predicting the depth of material lost through adhesive wear to identify parameters that should be modified in order to reduce valve recession. Tests indicated that the use of lower seat angles and hardened seat inserts showed the greatest promise in reducing valve recession. Table 3.1 provides a summary of design modifications that have been tested in an effort to reduce valve recession. 3.3 Guttering Guttering is a high-temperature corrosive process that usually occurs in exhaust valves. Guttering causes a leakage path to form radially across the sealing area between the valve seating face and the seat insert. In some instances the channel enlarges and, as combustion occurs in the cylinder, it follows the leak path into the valve port, rapidly melting, or ‘torching’ the valve [12]. Under magnification, guttered valve surfaces suffering from intergranular corrosion have a characteristic cobblestone appearance [13]. Guttering valve wear eventually causes excessive leakage of cylinder compression, and misfiring results in loss of power. A common cause for exhaust valve guttering in diesel engines is ash deposit flaking [14]. An initiating leakage path can form across a valve seating face/seat insert interface where a flake is missing. The path then becomes wider and wider. 3.4 Torching Torching, or melting, of a valve has been observed to occur rapidly in just a few engine cycles. It is associated with engines experiencing preignition or abnormal combustion. 29 Automotive Engine Valve Recession Table 3.1 Design changes tested in an effort to reduce valve recession Design change Effect in reducing wear Reduction in seating face angle (45– 30 degrees) Reduced wear. With all other factors unchanged, wear was lowered by 1/3 to 1/4 of that with 45 degree angle [3]. Tests carried out with 80+ material pairs and reduction in seat angle. Concluded that change in seat angle and suitable material pairing alone would not solve problem [4]. Reduced wear [6]. ‘Recession was reduced by approximately 75 per cent when seat angle was reduced from 45 to 30 degrees’ [11]. Lubrication of contact between valve and seat Reduced wear [3]. ‘Very effective method for reducing wear’ [4]. ‘Considerable’ reduction in seat wear obtained by injecting lubricating oil into the inlet manifold [6]. Reducing speed of valve impact Reduced wear [3]. ‘The most powerful means of reducing wear is to keep impact velocities as low as possible’ [8]. ‘A considerable reduction in valve seating velocity had negligible effect on seat wear’ [6]. Use of hardened valve seat inserts Reduced wear [6]. ‘Recession rates were reduced 80–95 per cent by using hardened inserts’ [11]. Positive rotation of valve Reduced wear [3]. Greater rigidity of valve head Reduced wear [3]. ‘Stiffer valve head sections were found to be beneficial in reducing wear rates. Lower valve head deflection or ‘oil canning’ reduces scrubbing distance’ [11]. Improved cooling of cylinder head bottom Reduced wear [3]. Greater rigidity of valve mechanism Reduced wear [3]. Welding material onto valve seating face Reduced wear [3]. Increasing seat width Reduced wear [3]. ‘Increased seat width showed little reduction in wear rates’ [11]. Hardening of seat Flame hardening of seats had little effect in reducing wear [3]. Induction hardening of seats reduced wear rate 25 per cent over that with non-hardened seats [11]. Reducing clearance between stem and valve guide Little effect in reducing wear [3]. Reducing valve head weight Little effect in reducing wear [3]. Using slightly different valve and seat angles Little effect in reducing wear [3]. Use of resilient valve head (tulip valve) Increased wear [3]. Use of resilient seat insert Increased wear [3]. 30 Valve Failure Preignition is described as unscheduled premature combustion. Some part of the combustion chamber becomes hot enough to ignite the fuel–air charge prior to timed ignition from the spark plug. Once initiated, the cycle continues with the combustion chamber becoming hotter as ignition occurs earlier and earlier in the cycle. If unchecked the severe temperature rise literally melts the more susceptible components in the combustion chamber, usually the piston crown or the exhaust valve. Preignition failures of this type are seldom seen in diesel/gas engines because there is no fuel mixed with the air during early compression. With abnormal combustion (‘knock’) there is a high pressure rise before or near piston top-dead-centre. The high pressure rise causes compression heating of burning gases and a more rapid heat release rate, raising the valve temperature which may trigger valve torching. 3.5 Effect of engine operating parameters 3.5.1 Temperature Inlet valve temperatures are not normally high enough to cause significant corrosion or thermal fatigue failures. Such failures are far more likely to occur in exhaust valves. However, a recent study of inlet valve failures [15] led to the conclusion that deposit build-up on the seating face of an inlet valve (formed from engine oil and fuel) had reduced heat transfer from the valve head (a valve transfers approximately 75 per cent of the heat input to the top-of-head through its seat insert into the cylinder head [11]), resulting in tempering and reduced hardness. As a result, some valves had suffered failures due to radial cracking of the seating face induced by thermal fatigue while others had failed due to valve guttering. Cherrie [16] found that when the temperature of a 21.4N steel valve was increased from 704 to 732 °C, the stress that could be sustained to rupture (failure accompanied by significant plastic deformation) in 100 hours decreased by 35 per cent. De Wilde [17] found, however, that at the temperatures experienced in exhaust valves and seat inserts, there was no significant reduction in mechanical properties and thus discounted temperature as a major influence on valve wear. Matsushima [18] investigated the wear rate of valve and seat inserts at elevated temperatures. Several insert materials were tested with Stellite valves. Figures 3.8 and 3.9 show the wear rate of the exhaust valve and seat insert, respectively. Valve seating face wear increased as the temperature exceeded 200 °C and continued to rise as the temperature was increased to 500 °C. At temperatures below 200 °C the wear was almost negligible. The wear rate of the seat inserts peaked at 300 °C, decreased at 400 °C, but rose again as the temperature was increased above 400 °C. The use of superalloy seat inserts reduced the seat insert wear below that of cast iron inserts, but increased the wear on the valve seating face. The powder metal insert containing materials to form lubricious oxides improved both seat insert and valve wear. 31 Automotive Engine Valve Recession Fig. 3.8 Wear of Stellite valve faces when mated with various seat inserts [18] Fig. 3.9 Wear of various seat inserts when mated with Stellite valves [18] 32 Valve Failure Hofmann et al. [19] and Wang et al. [20], however, have shown that the general trend is that wear decreases as temperature increases (as shown in Fig. 3.10). This was thought to be because of oxide formation at high temperatures preventing metal-tometal contact and thus reducing adhesive wear. Fig. 3.10 Seating face wear as a function of temperature: (a) valve seating face scar depth; (b) valve seating face scar width; (c) seat insert seating face scar depth; (d) seat insert seating face scar width [20] 33 Automotive Engine Valve Recession 3.5.2 Lubrication Engine lubricants have been found to both increase and reduce valve and seat wear, depending on the additive composition and the amount of oil that reaches the valve/seat interface. In inlet valves, liquid film lubrication is most dominant as temperatures are not usually high enough to volatilize the lubricant hydrocarbons and additives. Exhaust valves, however, are predominantly lubricated by solid films formed at the higher operating temperatures by oil additive ash compounds such as alkaline-earth and other metal oxides, sulphates, and phosphates (e.g. calcium, barium, magnesium, sodium, zinc, and molybdenum). Very thin metal oxide films have been found to be beneficial in reducing valve wear [21]. Too much solid film lubricant, however, can be detrimental and lead to valve guttering or torching due to flaking (as described in Sections 3.3 and 3.4). 3.5.3 Deposits Extensive work has been carried out investigating the formation mechanisms, effect, and methods of reducing inlet valve deposits in gasoline engines [22–28]. This has shown that deposits are produced from engine oil, fuel, and soot-like particles [24, 27] and that deposits accumulating on inlet valves affect drivability, exhaust emissions, and fuel consumption in gasoline engines. Many experiments have demonstrated that engine parameters, such as oil leakage at valve guides and positive crankshaft ventilation, valve temperature, and exhaust gas recirculation influence the deposit formation. Very little work, however, covers inlet valve deposits in diesel engines. It has been shown that exhaust valve deposits, formed from combustion products, prove favourable in providing lubrication on the seat contact surface (see Section 3.5.2). Their influence on inlet valve wear, however, has not been investigated. Esaki et al. [24] characterized the deposit formation on inlet valves in diesel engines. Figure 3.11 shows this formation, as well as the temperatures of the various parts of the valve. The black deposit was made up mainly of concentrated engine oil, oxidation products, and precarbonization products. The major part of the grey deposit was ash. It mainly comprised calcium sulphate. It was demonstrated that the black deposits accumulate at inlet valve temperatures of approximately 230 to 300 °C. At inlet valve temperatures above 350 °C the deposits or components of engine oil on the inlet valve were converted to ash. 3.5.4 Rotation Valve rotation can be achieved either by the use of positive rotators or by the use of multi-groove collets rather than clamping collets. These allow the valve to rotate under vibrational influences from the valve train or valve spring. This rotation can be promoted if the centre of the cam is offset from the valve axis. Hiruma and Furuhama [29] measured exhaust valve rotation and found that, in the engine under consideration, the valve started rotating after the engine speed exceeded 3000 r/min and then increased rapidly at higher engine speeds (as shown in Fig. 3.12). 34 Valve Failure 250 OC BLACK 350 OC GREY 450 OC Fig. 3.11 Cross-section of accumulated deposits on diesel engine inlet valves as characterized by Esaki et al. [24] Fig. 3.12 Speed of rotation of an exhaust valve [29] It was also found that at low speed operation (3000 r/min) the valve did not rotate constantly and changed its direction occasionally. Beddoes [30] also observed that valve rotation was random and occurred in either direction. It was found that valves stopped and started rotating as engine speed altered, usually beginning rotation at about 50 per cent of maximum engine speed. There is some agreement that valve rotation is beneficial in grinding away deposits. This prevents local hot spots forming and helps maintain good sealing and thermal contact of the valve to the seat [30–32]. The role valve rotation plays in valve/seat insert wear, however, is not fully understood. 35 Automotive Engine Valve Recession 3.6 Summary The review of literature indicated that the majority of the work carried out previously on diesel engine valve wear had focussed on large engines rather than those utilized in passenger cars, and a greater emphasis had been placed on investigating exhaust valve wear than that found in inlet valves. The complex nature of the valve operating environment and the difficulties associated with making a quantitative analysis of the effect of the many variables involved in the valve operating system was highlighted in the work. Most investigations had concluded that valve and seat wear was caused by frictional sliding between the valve and seat under the action of the combustion pressure. Little account was taken of other possible wear mechanisms. Parameter studies had mainly focussed on the effect of engine operating conditions such as temperature and load. Little or no work had been carried out to investigate the effect on wear of design parameters, material properties, valve closing velocity, and the effect of reducing lubrication at the valve seat/interface. 3.7 References 1. Pyle, W. and Smrcka, N. (1993) Effect of lubricating oil additives on valve recession in stationary gaseous-fuelled four-cycle engines, SAE Paper 932780. 2. Tauschek, M.J. and Newton, J.A. (1953) Valve seat distortion, SAE Preprint 64. 3. Zinner, K. (1963) Investigations concerning wear of inlet valve seats in diesel engines, ASME Paper 63-OGP-1. 4. Marx, W. and Muller, R. (1968) Ein βeitrag zum Einlaβventilsitz-Verscheiβ an aufgeladeren Viertakt-Dieselmotoren (A contribution on the subject of the wear of inlet valve seats in supercharged four-stroke diesel engines – its origins and some remedies), MTZ Paper No. 29, in German. 5. Lane, M.S. and Smith, P. (1982) Developments in sintered valve seat inserts, SAE Paper 820233. 6. Pope, J. (1967) Techniques used in achieving a high specific airflow for highoutput medium-speed diesel engines, Trans ASME J. Engng Power, 89, 265–275. 7. Van Dissel, R., Barber, G.C., Larson, J.M., and Narasimhan, S.L. (1989) Engine valve seat and insert wear, SAE Paper 892146. 8. Fricke, R.W. and Allen, C. (1993) Repetitive impact-wear of steels, Wear, 163, 837–847. 9. Ootani, T., Yahata, N., Fujiki, A., and Ehia, A. (1995) Impact wear characteristics of engine valve and valve seat insert materials at high temperature (Impact wear tests of austenitic heat-resistant steel SUH36 against Fe-base sintered alloy using plane specimens), Wear, 188, 175–184. 10. Narasimhan, S.L. and Larson, J.M. (1985) Valve gear wear and materials, SAE Paper 851497, SAE Trans, 94. 36 Valve Failure 11. Giles, W. (1971) Valve problems with lead free gasoline, SAE Paper 710368. 12. Arnold, E.B., Bara, M., and Zang, D. (1988) Development and application of a cycle for evaluating factors contributing to diesel engine valve guttering, SAE Paper 880669. 13. McGeehan, J.A., Gilmore, J.T., and Thompson, R.M. (1988) How sulphated ash in oils causes catastrophic diesel exhaust valve failures, SAE Paper 881584. 14. Tantet, J.A. and Brown, P.I. (1965) Series 3 oils and their suitability for wider applications, NPRA, Tech 65-29L. 15. Pazienza, L. (1996) 1.8 IDI chromo 193 intake valve failures, EATON Report No. 86/96. 16. Cherrie, J.M. (1965) Factors influencing valve temperatures in passenger car engines, SAE Paper 650484. 17. De Wilde, E.F. (1967) Investigation of engine exhaust valve wear, Wear, 10, 231–244. 18. Matsushima, N. (1987) Powder metal seat inserts, Nainen Kikan, 26, 52–57, in Japanese. 19. Hofmann, C.M., Jones, D.R., and Neumann, W. (1986) High temperature wear properties of seat insert alloys, SAE Paper 860150, SAE Trans., 95. 20. Wang, Y.S., Narasimhan, S., Larson, J.M., Larson, J.E., and Barber, G.C. (1996) The effect of operating conditions on heavy duty engine valve seat wear, Wear, 201, 15–25. 21. Wiles, H.M. (1965) Gas engines valve and seat wear, SAE Paper 650393. 22. Bitting, B., Gschwendtner, F., Kohlhepp, W., Kothe, M., Testroet, C.J., and Ziwica, K.H. (1987) Intake valve deposits – fuel detergency requirements revisited, SAE Paper 872117 (SP-725), SAE Trans., 96. 23. Cheng, S. (1992) The physical parameters that influence deposit formation on an intake valve, SAE Paper 922257. 24. Esaki, Y., Ishiguro, T., Susuki, N., and Nakada, M. (1990) Mechanism of intake valve deposit formation: Part 1 – Characterization of deposits, SAE Paper 900151, SAE Trans., 99. 25. Gething, J.A. (1987) Performance robbing aspects of intake valve and port deposits, SAE Paper 872116. 26. Houser, K.R. and Crosby, T.A. (1992) The impact of intake valve deposits on exhaust emissions, SAE Paper 922259. 27. Lepperhoff, G., Schommers, J., Weber, O., and Leonhardt, H. (1987) Mechanism of the deposit formation at inlet valves, SAE Paper 872115 (SP-725). 28. Nomura, Y., Ohsawa, K., Ishiguro, T., and Nakada, M. (1990) Mechanism of intake valve deposit formation: Part 2 – Simulation tests, SAE Paper 900152, SAE Trans, 99. 37 Automotive Engine Valve Recession 29. Hiruma, M. and Furuhama, S. (1978) A study on valve recession caused by non-leaded gasoline – measurement by means of R.I., Bullet. JSME, 21, 147–160. 30. Beddoes, G.N. (1992) Valve materials and design, Ironmaking and steelmaking, 19, 290–296. 31. Heywood, J.B. (1988) Internal combustion engine fundamentals, McGraw-Hill, London. 32. Stone, R. (1992) Introduction to internal combustion engines, Macmillan, Basingstoke. 38 Chapter 4 Analysis of Failed Components 4.1 Introduction This chapter details work carried out to evaluate valves and seat inserts from durability dyno tests run on a 1.8 litre, IDI, automotive diesel engine. It then goes on to report the findings of two investigations carried out concerning the failure of valves and seat inserts. The first relates to lacquer formation on valve seating faces during durability tests on a turbocharged diesel engine. The second investigates failures during durability testing of a 1.8 litre, DI, diesel engine caused by poor valve seating due to uneven seat insert wear. The work was undertaken in order to provide a means of comparison and validation for future bench test work and to establish analysis techniques for use during such work. It was also intended to provide information on possible causes of valve recession, and engine operating conditions and design features that may influence the wear process. 4.2 Valve and seat insert evaluation Recessed valves and seat inserts from durability dyno tests run on a 1.8 litre, IDI, automotive diesel engine were evaluated in order to provide a comparison and validation for later bench testing. Techniques such as profilometry and optical microscopy were used for the analysis. 4.2.1 Specimen details Valves and inserts made from a variety of materials were examined. Seat insert materials S1 and S2 consist of a sintered martensitic tool steel matrix with evenly distributed intra-granular spheroidal alloy carbides. Metallic sulphides are distributed throughout at original particle boundaries. The interconnected porosity in both is substantially filled with copper alloy throughout. Seat insert material S3 is cast and consists of a tempered martensitic tool steel matrix with a network of carbides uniformly distributed. Valve material V1 is a martensitic low-alloy steel and material V2 is an austenitic stainless steel. Some of these materials are currently used in inlet valve applications while the others were being tested to assess their performance. 39 Automotive Engine Valve Recession 4.2.2 Profile traces Profile traces of both valve and seat insert seating faces were taken using a profilometer (Surfcom) (see Fig. 4.1). The valve and seat insert stands shown were clamped in position and the stylus was returned to the same position for each profile taken. Therefore, apart from inaccuracies due to slight differences in the machining of the valves and seat inserts, each profile had the same origin. STYLUS HEAD VALVE SEAT INSERT Fig. 4.1 Use of a profilometer to take profile traces of valves and seat inserts Figures 4.2 and 4.3 show profiles of inlet valves and seat inserts taken from two dyno tests run under the same operating conditions. Comparison of the valve profiles with that of an unused valve, also shown in Fig. 4.2, gives an indication of the amount of wear that has occurred. Valve clearance data recorded during the tests (shown in Fig. 4.4) indicated that in the first test (valve material V1 run against seat insert material S1) major valve recession occurred (0.4 mm in 100 hours against a benchmark of 0.6 mm in 250 hours), whereas in the second (valve material V1 run against seat insert material S3) only minor valve recession occurred (0.15 mm in 130 hours). This is confirmed by comparison of the seat insert profiles (see Fig. 4.3), which clearly shows that the sintered seat insert material S1 has worn more than the cast seat insert material S3. The valve wear, however, was greater in the second test. This can be explained by looking at the seat insert material used in each test. The cast seat insert material used in the second test is tougher and more resistant to impact than the sintered material used in the first test, hence the reduction in seat insert wear and the increase in valve wear. 40 Analysis of Failed Components Fig. 4.2 Inlet valve profile traces Fig. 4.3 Inlet seat insert profile traces 41 Automotive Engine Valve Recession Fig. 4.4 Valve clearance data from engine dyno tests Profilometry clearly provides data that compare well with valve clearance data taken during engine tests. It also provides an indication of relative valve and seat insert wear rather than just giving an overall figure. 4.2.3 Visual rating On the inlet valve shown in Fig. 4.5 it is just possible to see the wear scar on the seating face. It can also be seen that the deposits on the valve head compare well with those characterized by Esaki et al. [1] (see Fig. 3.11). In some cases, when using optical microscopy to examine the valves, a series of circumferential ridges and valleys were observed around the axis of the seating faces (see Fig. 4.6). These correspond to the ‘single wave formation’ described by Van Dissel et al. [2]. The seating face of an unused valve is shown in Fig. 4.7 for comparison. When analysing some of the seat inserts, evidence was found of scratches in the radial direction (see Fig. 4.8) similar to those described by Van Dissel et al. [2]. Circumferential grooves and pitting were also observed. Profiles taken indicated that in some tests uneven wear of seat inserts was occurring. The results of the evaluation of inlet valves and seat inserts from engine dyno tests indicate that they provide a means of comparison and validation for later bench test work. Wear features observed on both valve and seat insert seating faces correspond well with those described in the literature. 42 Analysis of Failed Components Fig. 4.5 Inlet valve (valve material V1 run against insert material S3) Fig. 4.6 Inlet valve seating face (valve material V1 run against seat insert material S3 in a high-speed engine test); major seat diameter is towards bottom of figure 43 Automotive Engine Valve Recession Fig. 4.7 Unused inlet valve seating face (valve material V1); major seat diameter is towards bottom of figure Fig. 4.8 Inlet seat insert seating face (valve material V1 run against seat insert material S1 in a high speed engine test); major seat diameter is towards top of figure 44 Analysis of Failed Components 4.3 Lacquer formation on inlet valves After running a full load durability test on a turbocharged diesel engine, it was found that the seating faces of the inlet valves were coated with what appeared to be a dark lacquer. Inlet valves from a similar test run at part load, in which exhaust gas recirculation (EGR) was able to operate, revealed no presence of lacquer. Lacquer build-up presents a serious problem because a piece breaking away from the seating face can create a channel through which hot gases are able to escape. This causes guttering of the seating face which eventually leads to valve failure. Failures of this type had been observed in similar full load durability tests on this engine. The objective of this work was to investigate diesel engine inlet valve deposits and lacquer formation and propose some reasons for the appearance of lacquer on the inlet valves. 4.3.1 Valve evaluation Valves from both the full load test and the part load test with EGR were evaluated. Valves from each test are shown in Fig. 4.9. The deposit formation on the valve that underwent the part load test with EGR was black and oily around the valve stem fillet area turning into ash approaching the valve seating face. The valve seating face itself was clear of any deposit. The deposit on the valve that underwent the full load test, however, was virtually all ash and the lacquer formation had visibly dulled the appearance of the valve seating face. Fig. 4.9 Accumulated deposits on inlet valves: (a) part load test with EGR – no lacquer present on valve seating face; (b) full load test – lacquer present on valve seating face 45 Automotive Engine Valve Recession 4.3.2 Discussion The deposit formation on the valve from the part load test compares well with that characterized by Esaki et al. [1], see Fig. 4.10. 250OC BLACK 350OC GREY 450OC Fig. 4.10 Cross-section of accumulated deposits on diesel engine inlet valves as characterized by Esaki et al. [1] The high ash content of the deposit found on the valves from the full load test clearly indicates that the temperatures were higher than the 350 °C threshold for such deposit formation described above. The temperature of the valve in the part load test was clearly considerably lower, however, hence the more characteristic oily, black deposit formation. The deposits found on the valves from both tests are clearly composed of lubricating oil that leaked through the valve stem seals onto the valve stem fillets. Lacquer formation is said to occur as a result of oxidation of lubricating oil within the engine [3]. Schilling [4] observed that the unfavourable operating condition in engine parts where lacquer formation was likely to occur was high temperature. This indicates that the lacquer formation on the valves in the full load test could have been caused by the high temperatures experienced at the valve seating face, leading to oxidation of lubricating oil present as a result of valve stem leakage. A reduction in heat transfer from the valve head would have occurred as a result of the lacquer formation, further increasing the valve head temperature and increasing the probability of the lacquer build-up progressing. 46 Analysis of Failed Components The fact that no lacquer was found on the valves from the part load tests can be explained by the lower temperatures experienced. Also, during the part load test when the EGR was in operation, hard combustion particles would have been recirculated in the exhaust gases. These would have provided an abrasive wear medium between the valve seating face and the valve seat insert, further reducing the likelihood of any deposit formation. 4.4 Failure of seat inserts in a 1.8 litre, DI, diesel engine The engine under consideration in this investigation was a 1.8 litre, DI, naturally aspirated diesel engine with a direct acting cam. This engine is an upgraded version of the 1.8 litre, IDI, diesel engine considered previously. One of the major design changes incorporated was the use of direct fuel injection rather than indirect injection, which required the inclusion of holes in the cylinder head between the inlet and exhaust seat ports to accept the fuel injector (see Fig. 4.11). EXHAUST VALVE SEAT INSERT INLET VALVE SEAT INSERT 180O 270O 90O 0O FUEL INJECTOR HOLE Fig. 4.11 Position of profiles taken on 1.8 l DI inlet seat insert It was found that on start-up after durability testing, the inlet valves were not seating correctly and consequently sealing was not achieved at the valve/seat insert interface. As a result, pressure was being lost from the cylinder. After the engine had warmed up the valves began to seat correctly. However, on starting the engine again the valves did not seat correctly. It was hypothesized that this was a wear problem. The seat inserts were deforming (elastically) when hot and being worn unevenly. On cooling they were returning to their original shape. On restarting the engine the inlet valves were not seating correctly as 47 Automotive Engine Valve Recession the insert seating faces had been left severely out of round due to the uneven wear. The objective of this work was to establish whether it was a wear problem combined with elastic deformation of the seat inserts or whether there was another reason for the valve/seat insert failures. 4.4.1 Inlet seat insert wear In order to establish whether, or how much, wear had occurred on the seating faces of the inlet seat inserts, profiles were taken of an inlet seat insert at four positions, each 90 degrees apart, as shown in Fig. 4.11. The seat insert was removed from the cylinder head prior to the profiles being taken. Examples of the resulting plots are shown in Fig. 4.12. The seating face widths were measured to see how much wear had occurred relatively at the four positions (for widths see Fig. 4.13). Fig. 4.12 Profile traces taken on 1.8 l DI inlet seat insert 48 Analysis of Failed Components 2.03mm 180O 2.30mm 270 O 90 O 1.75mm 0O 1.73mm Fig. 4.13 Seating face widths at four positions around inlet seat insert The measurements clearly indicated that there was uneven wear around the seating face which fitted in with the hypothesis outlined above. The maximum wear was found to have occurred in the proximity of the hole for the fuel injector in the centre of the cylinder head. Photographs were taken of the seating face of the seat insert (using an optical microscope) that revealed some evidence of indentations in the radial direction, see Fig. 4.14. This indicated that wear resulted from valve head flexure rather than valve rotation. Fig. 4.14 A 1.8 l DI inlet seat insert seating face (valve material V1 run against seat insert material S3 in a high-speed engine test); major seat diameter is towards top of figure 49 Automotive Engine Valve Recession 4.4.2 Deposits On examining the cylinder head it was found that a thick, black, oily deposit had built up on the inside of the inlet valve seat inserts. As shown in Fig. 4.15, the deposit was also found on the cylinder head material just below the seat insert and on the base of a seat insert removed from the cylinder head. It was noted that the deposit build-up had occurred in the proximity of the hole positioned in the centre of the cylinder head for the fuel injector. The deposit found on the base of the inlet seat insert could have indicated either that it was not inserted correctly or that it moved during the engine test. Movement could have resulted from non-uniform deformation of the cylinder head and seat insert. This could have been as a result of the fuel injector hole in the centre of the cylinder head. DEPOSIT ON BASE OF S.I. DEPOSIT INSIDE S.I. Fig. 4.15 Deposit build-up on 1.8 l DI inlet seat insert 50 Analysis of Failed Components 4.4.3 Misalignment of seat insert relative to valve guide In removing the inlet seat insert from the cylinder head, the cylinder head material around the seat insert was milled away to gain access to the seat insert. The milling machine was centred on the valve guide. During the milling process, it was noted that when the cutter was through to the seat insert on one side, cylinder head material was still present on the other, as shown in Fig. 4.16. Fig. 4.16 Evidence of seat insert misalignment relative to the valve guide (1.8 l DI) It was apparent that on insertion, the seat insert had been misaligned relative to the valve guide. This would have been remedied when the grinding of the seating faces on the seat insert occurred, as the grinders are lined up with the valve guide. However, one side of the seat insert was thinner than the other as a result of the initial misalignment. The thinner section was in the proximity of the hole for the fuel injector between the inlet and exhaust ports. This could have affected the heat transfer from the valve head into the cylinder head in this region, which could have led to deformation of the seat insert. 4.4.4 Inlet valve wear It was thought that if the problem was related to wear, the inlet valve seating faces may exhibit wear of a similar magnitude to that of the inlet seat inserts. In order to establish whether or how much wear had occurred on the seating face of the valves, profiles were taken of a valve seating face at four positions, each 90 degrees apart. The resulting plots are shown in Fig. 4.17. The similarity of the profiles obtained at the four different positions indicated that if wear had occurred it was evenly distributed 51 Automotive Engine Valve Recession Fig. 4.17 Profile traces taken on 1.8 l DI inlet valve around the circumference. There was a slight dip in the profiles, which may indicate that some wear occurred. It is possible that this was the point on the valve at which it made contact with the seat insert. Although even valve wear may be expected as a result of rotation, it was clear, as hypothesized, that the problem was related to seat insert wear. 4.5 Conclusions 1. Three failures of valve/seats from dyno tested automotive diesel engines have been analysed. 2. Techniques such as optical microscopy and profilometry have been established for analysis. 3. The following parameters have been identified as influencing valve recession: ● valve and seat insert material properties; ● valve and seat insert misalignment (caused by cylinder head deformation as a result of non-uniform cooling); ● deposit formation on the seating face of a valve. 4. Wear data and surface appearance of failed components were available for comparison and validation with later bench testing. The work has also further emphasized the unique nature of each valve recession problem. Valve recession is certainly not purely a materials selection issue; it is also affected greatly by the valve train and cylinder head design and manufacturing tolerances. 52 Analysis of Failed Components 4.6 References 1. Esaki, Y., Ishiguro, T., Susuki, N., and Nakada, M. (1990) Mechanism of intake valve deposit formation: Part 1 – Characterization of deposits, SAE Paper 900151, SAE Trans., 99. 2. Van Dissel, R., Barber, G.C., Larson, J.M., and Narasimhan, S.L. (1989) Engine valve seat and insert wear, SAE Paper 892146. 3. Denison, G.H. and Kavanagh, F.W. (1955) Recent trends in automotive lubricating oil research, Section 6/C, Preprint 1, Proc. Fourth World Petroleum Congress, Rome. 4. Schilling, A. (1968) Motor oils and engine lubrication, Scientific Publications (GB) Ltd. 53 This page intentionally left blank Chapter 5 Valve and Seat Wear Testing Apparatus 5.1 Introduction Dynamometer engine testing is often employed to investigate valve wear problems. This is expensive and time consuming and does not necessarily help in finding the actual cause of wear. Since valve wear involves so many variables, it is impossible to confirm precisely individual quantitative evaluations of all of them during such testing. In addition, the understanding of wear mechanisms is complicated by inconsistent patterns of valve failure. For example, failure may occur in only a single valve operating in a multi-valve cylinder. Furthermore, the apparent mode of failure may vary from one valve to another in the same cylinder or between cylinders in the same engine. Each case, therefore, has to be painstakingly investigated, the cause or causes of the problem isolated, and remedial action taken. In order to isolate the critical operating conditions and analyse the wear mechanisms in detail, simulation of the valve wear process must be used. This has the added benefits of being cost effective and saving time. This chapter details the requirements of valve and seat wear test apparatus, wear test methods, and extant valve wear test rigs. It then goes on to describe work undertaken at the University of Sheffield to design and build experimental apparatus that would be able to simulate the loading environment and contact conditions to which the valve and seat insert are subjected in an engine. The apparatus was to be used in future bench test work intended to isolate parameters critical to the valve recession problem. 5.2 Requirements The review of literature and the failure diagnosis carried out indicated that there were a number of critical parameters which affect the rate of valve and seat insert wear: combustion load influences the frictional sliding; high temperatures affect the wear mechanism and the formation of deposits; misalignment leads to uneven seating loads; deposit formation affects heat transfer; and rotation plays a role as yet not understood. The major requirements of valve and seat wear test apparatus could, therefore, be listed as: ● combustion loading; ● impact load on closing; ● valve misalignment; 55 Automotive Engine Valve Recession ● valve rotation; ● temperature control. 5.3 Wear test methods A number of different wear tests developed to evaluate material wear are compared in Table 5.1. They are listed in order of increasing complexity. Table 5.1 Wear test methods [1] Test method Type of test Test conditions Measured quantity Crossed cylinder (ASTM G 83-83) Adhesive wear High-contact stress High-sliding velocity No lubrication Weight loss Block on ring (ASTM G 77-83) Adhesive (sliding) High-contact stress Sliding speed High temperature No lubrication Weight loss Friction Thrust washer Adhesive/abrasive High-contact stress Sliding speed High temperature No lubrication Weight loss Wear depth Wear profile Cycles to failure Bench test rigs General Valve gear lube Speed Temperature Spring load Seating velocity Oil residue analysis Wear depth Wear profile Motorized or fired engine tests General Engine operating conditions Speed Torque Basic wear tests are standard tests. They allow close control of the test conditions such as loads, environment, and dimensions. The ‘crossed cylinder’ and ‘block on ring’ tests are generally used to evaluate adhesive wear resistance. Thrust washer tests, at high temperatures, can be used to simulate valve seat insert wear conditions. Bench test rigs simulate actual valve gear operating conditions. However, they only allow limited control over operating conditions. Engine tests can be motorized fixtures, firing engine dynamometer tests, or fleet tests with actual vehicles. These are expensive and time consuming, however, and it is difficult to isolate the actual cause of any wear that occurs. 5.4 Extant valve and seat wear test rigs A review of extant valve wear test rigs was carried out in order to look at how wear investigations or material ranking could be achieved using the various testing methods outlined in Section 5.3. The test rigs reviewed could be split into three different types: 1. static test rigs used to measure deflection; 2. wear test rigs using material specimens; 3. wear test rigs utilizing actual valves and seat inserts. 56 Valve and Seat Wear Testing Apparatus Brief notes on each of the rigs reviewed can be found in Table 5.2 and diagrams of some of the rigs are shown in Fig. 5.1. Table 5.3 gives a summary of the engine operating parameters that can be simulated using the rigs reviewed. Table 5.2 Details of the valve/seat wear test rigs reviewed Reference Notes Zinner [2] Rig was used to study the effect of seat insert distortion on the wear of diesel engine inlet valves. Static rig in which actual seating condition was simulated. Seat insert was forced from side to artificially deform. Valve stem protrusion was measured at different air pressures applied to the valve head. It was found that seat insert distortion increased valve stem protrusion. It was concluded that partial contact of the mating faces increased valve seat slide, causing excessive valve seat wear. Pope [3] Rig was constructed to check the validity of Pope’s theoretically derived wear factor for inlet valve seats. Valve stem protrusion was measured at different oil pressures applied to the valve head. Results showed that the magnitude of the theoretical deflection was of the correct order. Matsushima [4] Rig was used to study the wear rate of valve face and seat inserts at elevated seat temperatures. Several powder metal seat insert materials were tested. The rig was able to move the valve up and down to simulate the opening and closing motion. The valve was also rotated. The valve was heated using a gas burner. Not clear how or whether the valve was loaded during the opening and closing cycle. Narasimhan and Larson [1] The test rig was used for wear testing valve seat and seat insert materials at high temperatures. The valve head was heated by a gas burner and the temperature of the valve head, seat and seat insert were monitored and controlled during the test. Valve seating velocity and the seating loads were controlled using a servo-hydraulic actuator. Tests were performed to a fixed number of cycles or until failure occurred. Blau [5] The test rig was used to simulate repetitive impact and seating of a valve on its seat. Rectangular ceramic coupons were used and a spherically tipped hammer produced the impact. The rig was able to heat and lubricate the test specimens. Hofmann et al. [6] Hot wear tester designed to assess the hot wear properties of powder metal alloys used in seat inserts. The tester was able to control the test variables of load, temperature, atmosphere and speed. The contact geometry was a rotating cylinder on a stationary block. A positive correlation was found between wear loss data and engine test results. Nakagawa et al. [7] The wear test rig was used to test hard surfacing alloys for internal combustion engine inlet valves. Valve face wear depth was measured for different valve face temperatures. Fujiki and Makoto [8] Valve wear test rig utilized actual valve gear. The valve was heated using a gas burner. It is not clear how or whether the valve was loaded. A degree of success was indicated in correlating valve seat wear measurements from bench testing with engine dynamometer testing. Malatesta et al. [9] Test rig utilized a hydraulic actuator to compressively load a valve against its seat insert. A gas burner was used to heat the valve. It was possible to produce various conditions of misalignment (angular and lateral). The valve was prevented from rotating in the rig. A load cell was used to monitor the magnitude of the seating load, and in conjunction with a control loop the desired load was maintained. Sample and system temperatures were monitored using thermocouples. The test rig was able to achieve seating velocities of approximately 250 mm/s, load the valve to a maximum of 37 810 N and produce valve seat temperatures up to 816 °C. Studies were made of the relationship between the number of cycles and both wear depth and area. As a means to validate the results a comparison was made with several heavy duty diesel valves. 57 Automotive Engine Valve Recession (a) (b) (c) (d) (e) (f) Fig. 5.1 Examples of the test rigs reviewed: (a) Fujiki and Makoto [8]; (b) Malatesta et al. [9]; (c) Blau [5]; (d) Nakagawa et al. [7]; (e) Hofmann et al. [6]; (f) Matsushima [4] 58 Valve and Seat Wear Testing Apparatus Table 5.3 Summary of engine operating parameters that can be simulated by the test rigs reviewed Reference Rig type Impact load on valve closing Combustion loading Valve rotation Valve misalignment Temperature control Zinner [2] Static • Pope [3] Static • Narasimhan and Larson [1] Wear • (components) Hofmann et al. [6] Wear (specimens) • Matsushima [4] Wear (components) • Nakagawa [7] et al. Wear (components) • Fujiki and Makoto[8] Wear (components) • • Wear (specimens) • • Wear (components) • Blau [5] Malatesta [9] et al. • • • • • • • • • • • It was found that wear test rigs using material specimens provided good control over test conditions. However, without using actual components it was difficult to recreate exact contact conditions and it was, therefore, impossible to simulate wear features found on components taken from the field. Such test rigs could not be used for the study of wear mechanisms, but were suitable for ranking of valve and seat material performance. Wear test rigs utilizing actual valve train components were found to provide the best means of replicating wear features and studying wear mechanisms. While they offered less control over operating conditions than rigs using material specimens, it was still possible to isolate important test parameters. 5.5 University of Sheffield valve seat test apparatus Having established the requirements of the apparatus and the methods available for valve gear wear testing (see Section 5.3), and using the results of the review of extant test rigs (see Section 5.4), it was decided that, in order to realistically simulate the contact conditions to which a valve and seat are subjected in an engine, the best approach would be to design apparatus that utilized actual valve and seat inserts. In order to accurately replicate impact load on valve closure it was clear that actual valve gear would have to be used. However, it was determined that producing the combustion loading conditions would be difficult using camshaft and valve train components and that in order to investigate the effect of both it would be necessary to design two test rigs: one designed to fit in a hydraulic test machine, to study the effect 59 Automotive Engine Valve Recession of combustion loading while approximating valve dynamics; and a motorized cylinder head utilizing actual valve gear, to study the effect of impact on valve closure without the application of combustion loading. Different configurations were considered for the two rigs. The final design, however, for the combustion loading rig was mainly determined by the test facilities available for applying the required loading. It was decided to base the impact rig on a motorized cylinder head. 5.5.1 Hydraulic loading apparatus 5.5.1.1 Design The test rig (shown in Fig. 5.2) was designed to be mounted on a hydraulic fatigue testing machine (as shown in Fig. 5.3). The hydraulic actuator on the machine is used to provide the combustion loading cycles required and acts to ‘close’ the valve. A spring returns the valve to the ‘open’ position. The control system allows loading or displacement waveforms and their amplitude to be set, as well as the frequency of the loading or displacement. A built-in load cell and linear variable displacement Fig. 5.2 Hydraulic loading apparatus (Reprinted with permission from SAE paper 1999-01-1216 © 1999 Society of Automotive Engineers, Inc.) 60 Valve and Seat Wear Testing Apparatus Fig. 5.3 Test rig mounted in a hydraulic test machine transducer (LVDT) provide load and displacement measurements. A counter records the number of loading cycles. A removable seat insert holder is used to mount the seat insert in alignment with the valve. An additional holder was designed in which the seat insert was slightly offcentre in order to misalign the valve relative to the seat insert. A valve guide of bronze bushes was also built into the rig. The bushes and seat insert holders are designed to accommodate inlet valves and seat inserts from a 1.8 litre, IDI, automotive diesel engine, the geometries of which are shown in Fig. 5.4. The materials not available as seat inserts are made up into specimens as shown in Fig. 5.5. The use of a seat insert holder and an inserted valve guide allows some flexibility in the size of the valve and seat insert being tested. Heating is provided to both sides of the rig by hot air supplies directed at the valve and seat insert. Temperature measurements are taken using a contact probe placed on the outer edge of the valve head at the top of the seating face. A cooling coil is provided to prevent the load cell from overheating. Valve rotation is achieved using a motor-driven belt and pulley system as shown in Fig. 5.6. The valve collet, clamped around the valve stem, is able to move up and down in the collet guide while rotating (see Fig. 5.7). The valve is not intended to rotate while seated. 61 Automotive Engine Valve Recession Fig. 5.4 Valve and seat insert geometry (Reprinted with permission from SAE paper 1999-01-1216 © 1999 Society of Automotive Engineers, Inc.) Fig. 5.5 Seat specimen geometry 62 Valve and Seat Wear Testing Apparatus Fig. 5.6 Plan view of valve rotation system Fig. 5.7 Valve collet and collet guide Lubrication of the valve/seat insert interface can be achieved via holes in a tube placed around the thrust bearing housing above the valve head (see Fig. 5.8). Flow is controlled using a valve fitted below the lubricant reservoir. 63 Automotive Engine Valve Recession Fig. 5.8 Lubrication of valve/seat insert interface 5.5.1.2 Test methodologies In order to develop test methodologies with which to investigate the likely causes of valve and seat insert wear and establish experimental parameters, testing was carried out on the hydraulic loading apparatus to evaluate the performance of the hydraulic test machine to which it was mounted. Two different test methodologies were developed, the first to investigate the effect of the frictional sliding of the valve on the seat insert under the action of the combustion pressure, and the second to investigate the effect on wear of combining this with the impact of the valve on the seat insert during valve closure. Frictional sliding The first methodology employed a triangular loading waveform to investigate the effect of the combustion loading on valve and seat insert wear (see Fig. 5.9). The intention was to isolate the frictional sliding between the valve seating face and the seat insert, hence the valve seating face was in constant contact with the seat insert. A triangular waveform was used as this was the closest approximation to the cylinder pressure curve for a compression ignition engine (as shown in Fig. 5.10). 64 Valve and Seat Wear Testing Apparatus Fig. 5.9 Frictional sliding test methodology 60 Pressure (bar) 50 40 30 20 10 0 -120 -60 0 60 120 Crank Angle (degrees before tdc) Fig. 5.10 Hypothetical pressure diagram for a compression ignition engine [10] 65 Automotive Engine Valve Recession Impact and sliding The second test methodology used a sinusoidal displacement waveform (allowing the valve to lift off the seat), to investigate the effect of the impact of the valve on the seat insert as the valve closes, in combination with the combustion loading (see Fig. 5.11). A sinusoidal waveform was used as this was the closest available approximation of the motion of a valve in an engine. Fig. 5.11 Impact and sliding test methodology 5.5.1.3 Experimental parameters During initial testing, load, misalignment, rotation, and temperature were varied in order to establish baseline parameters for each of the test methodologies outlined above. Test parameters used are shown in Table 5.4. Table 5.4 Parameters varied during initial testing Frictional sliding tests Combustion load (kN) Load waveform Frequency (Hz) Misalignment (mm) Valve temp. (°C) Rotation (r/min) 13–15 Triangular 5–20 0–0.5 R.T.–130 0–1 Impact and sliding tests Combustion load (kN) Displacement waveform Valve lift (mm) Frequency (Hz) Misalignment (mm) Valve temp. (°C) Rotation (r/min) 13 Sinusoidal 0.6–1 10–15 0–0.5 R.T.–130 0–1 The magnitude of the load to be used, Pp, was calculated by multiplying the maximum combustion pressure, pp, by the valve head area for an inlet valve in a naturallyaspirated (N/A), 1.8 litre, IDI, diesel engine, see equation (5.1) 66 Valve and Seat Wear Testing Apparatus Pp = pp × π × Rv2 (5.1) where Pp is the peak combustion load (N) and Rv is the radius of valve head (m). For a N/A, 1.8 litre, IDI, diesel engine, pp = 13 MN/m2 and Rv = 18×10−3 m. Therefore, from equation (5.1) peak combustion load, Pp = 13.2 kN Other parameters, such as frequency, lift, misalignment, and rotation, were varied in order to optimize the test rig performance at the calculated load. Temperatures were restricted by the maximum temperature to which the load cell within the hydraulic test machine could be exposed. The baseline parameters established for optimum rig performance are shown in Table 5.5. Table 5.5 Baseline test parameters Frictional sliding tests Combustion load (kN) Load waveform Frequency (Hz) Misalignment (mm) Valve temp. (°C) Rotation (r/min) 13 Triangular 20 0.25 R.T. 1 Impact and sliding tests Combustion load (kN) Displacement waveform Valve lift (mm) Frequency (Hz) Misalignment (mm) Valve temp. (°C) Rotation (r/min) 13 Sinusoidal 0.6 10 0.25 R.T. 1 5.5.2 Motorized cylinder head 5.5.2.1 Design The test rig (shown in Fig. 5.12) utilizes a cylinder head from a 1.8 litre, IDI, diesel engine. The cylinder head is bolted to an adjustable bedplate which is mounted on a steel frame. An electric motor (1440 r/min rated speed), housed within the frame, is used to drive the camshaft via a belt and pulley system. This gives the camshaft a rotational speed of approximately 2700 r/min. For each rotation of the camshaft, the inlet valves open and close once. A soft starter is used to provide a gradual increase and decrease of motor torque during start-up and shut-down, respectively. This suppresses explosive start-up conditions and reduces peak loads on drivetrain components. It does not, however, allow the rotational speed of the motor to be varied. The camshaft speed is therefore fixed, but could be varied by using different pulley configurations. An overhead camshaft configuration is employed in this valvetrain system, with the cams acting directly on flat faced followers. The cams are offset from the valve stem 67 Automotive Engine Valve Recession axis (as shown in Fig. 5.13) in order to reduce localized wear on the follower. The use of offset cams, along with split collets, promotes valve rotation. Removable seat insert holders are used in order to speed up analysis during testing. These are clamped into holes machined in the cylinder head around the inlet ports. A gravity-fed oil drip system is used to lubricate the cam/follower interfaces and the camshaft bearings. A collection tray/shield fitted around the camshaft is used to recycle the lubricant. Fig. 5.12 Motorized cylinder head 68 Valve and Seat Wear Testing Apparatus VALVE SPRING VALVE STEM COLLET FOLLOWER CAM CAMSHAFT Fig. 5.13 Cam offset from the valve stem axis With regard to the study of impact on valve closure, the motorized cylinder head has the following advantages over the hydraulic loading apparatus described in Section 5.5.1: ● actual valve dynamics are utilized; ● impact on valve closure is isolated; ● actual valve rotation can be studied; ● valve dynamics can be varied by using different cam profiles. 5.5.2.2 Operation Development of a test methodology and selection of experimental parameters was more straightforward for the motorized cylinder head. In studying the impact of the valve on the seat at valve closure, only variation of the valve closing velocity was required. In order to vary the valve closing velocity, the rotational speed of the camshaft could be changed. In order to change the closing velocity of individual valves, the clearance was adjusted by employing seat insert holders of different thickness. 5.5.3 Evaluation of dynamics and loading In order to assess the extent to which the dynamics of the hydraulic test machine on which the test rig was mounted simulated the dynamics of an inlet valve, it was decided to investigate the dynamics of an inlet valve in a 1.8 litre, IDI, diesel engine with a direct-acting cam and compare them with those for the hydraulic test machine. 69 Automotive Engine Valve Recession 5.5.3.1 1.8 litre, IDI, diesel engine The valve lift curve for the 1.8 litre, IDI, diesel engine was taken directly from the inlet cam lift data (lift versus angle of rotation). The valve velocity, v, was then derived by multiplying the gradient of the lift curve by the rotational speed of the camshaft ω , see equation (5.2). An engine speed of 4800 r/min was assumed (engine speed for durability tests) giving a camshaft rotational speed of 2400 r/min. v=ω dl dθ (5.2) where l is the valve lift (mm) and θ is the rotation of camshaft (degrees). Both valve lift and velocity curves are shown in Fig. 5.14. Ideally the valve should close in the region of constant valve velocity between 145 and 160 degrees of camshaft rotation in order to limit impact stresses. Fig. 5.14 1.8 l, IDI, diesel engine inlet valve lift and velocity 70 Valve and Seat Wear Testing Apparatus The valve lift and valve velocity curves plotted against cam rotation for the region in which valve closure occurs are shown in Fig. 5.15 (magnification of Fig. 5.14). Valve closure does not occur when the lift is equal to zero as a clearance is introduced between the valve tip and the follower. This allows the engine to tolerate a certain amount of valve recession. The recommended maximum and minimum clearances for this type of engine are shown in Fig. 5.15. The valve closing velocities for the maximum and minimum clearances, also shown in Fig. 5.15, are 375 mm/s and 288 mm/s, respectively. These values are well below the maximum recommended value for valve closing velocity of 500 mm/s given by Stone [10]. Fig. 5.15 1.8 l, IDI, diesel engine inlet valve lift and velocity at valve closure Valve misalignment relative to the seat insert will cause the valve to impact the seat insert at a higher valve lift (as shown in Fig. 5.16). For a seating face angle of 45 degrees, the amount of misalignment will equal the increase in valve lift at closure. As shown in Fig. 5.15, a valve at maximum clearance misaligned by 0.25 mm closes off the constant velocity ramp and the closing velocity is increased to 1860 mm/s. 71 Automotive Engine Valve Recession Difference in Valve Lift at Closure Aligned Misaligned Fig. 5.16 Effect of valve misalignment on closing position The impact energy e calculated using equation (5.3) at a closing velocity of 1860 mm/s would be 24 times higher than that at a closing velocity of 375 mm/s – the valve closing velocity at the maximum valve clearance (see Fig. 5.17). e= 1 2 mv 2 (5.3) where m is the mass of the valve added to the mass of the follower and half the valve spring mass (kg). Fig. 5.17 1.8 l, IDI, diesel engine inlet valve energy at valve closure 72 Valve and Seat Wear Testing Apparatus Work done is equal to force multiplied by distance. In order to calculate the work done on a valve during combustion, Wv, the distance was taken as the maximum valve head deflection, ymax, and the force the maximum combustion load, Pp, see equation (5.4) Wv = ymax × Pp (5.4) Assuming that a valve head is a flat circular plate with radius Rv and thickness b (as shown in Fig. 5.18), the maximum valve head deflection was calculated using the equations for the deflection of a simply supported flat circular plate with a uniformly distributed load, see Figure 5.19 and equation (5.5) as outlined by Roark and Young [11] − Pp Rv (5 + ν ) 4 ymax = 64 D (1 + ν ) Eb 3 , where D = 12 1 − ν 2 ( (5.5) ) −3Pp Rv (1 − ν 2 )(5 + ν ) 4 ⇒ ymax = 16 Eb 3 (1 + ν ) −3Pp Rv (1 − ν )(5 + ν ) 4 ⇒ ymax = (5.6) 16 Eb 3 where b is the plate thickness (m), ν is Poisson’s ratio, and E is the modulus of elasticity (N/m2). Rv b Fig. 5.18 Flat circular plate ppp P ymax max Fig. 5.19 Simply supported circular flat plate with a uniformly distributed load 73 Automotive Engine Valve Recession For a N/A, 1.8 litre, IDI, diesel engine: pp = 13×106 N/m2; Rv = 18×10−3 m; b = 8×10−3 m (estimated value as valve head thickness varies); ν = 0.3; and E = 210×109 N/m2. Therefore, from equation (5.6) maximum valve head deflection, ymax = −8.8 µm For a N/A, 1.8 litre, IDI, diesel engine Pp = 13 200 N Therefore, from equation (5.4) work done on a valve during combustion, Wv = 0.12 J per combustion cycle 5.5.3.2 Hydraulic test machine The actuator on the hydraulic test machine acts to move the test rig up to and away from the valve which is held in a ‘fixed’ lateral position by a spring. Calculations were, therefore, based on the actuator motion. The sinusoidal lift curve for an ‘impact and sliding’ test (see Section 5.5.1.2) using baseline parameters (see Table 5.5) was determined using the amplitude a, frequency f, and initial actuator displacement L, see equation (5.7). The velocity curve was then determined by differentiating the lift function, see equation (5.8). Both curves are shown in Fig. 5.20. la = α sin(2πft ) + L (5.7) where la is the actuator lift (m) and t is time (seconds). va = dla = 2π f α cos (2π ft ) dt where va is the actuator velocity (m/s). 74 (5.8) Valve and Seat Wear Testing Apparatus Fig. 5.20 Test rig displacement and velocity in hydraulic test machine (13 kN combustion load) The velocity at closure for an aligned valve is 18 mm/s. The closing velocity for a valve misaligned by 0.25 mm, also shown in Fig. 5.20, is 16 mm/s. The energies at valve closure for an aligned and a misaligned valve (shown in Fig. 5.21) are 0.0081 J and 0.0064 J, respectively (calculated using equation (5.3), where m = test rig mass = 50 Kg). Fig. 5.21 Test rig energy in hydraulic test machine (13 kN combustion load) 75 Automotive Engine Valve Recession In order to calculate the work done on the valve per cycle, the valve head deflection first had to be estimated. The loading data shown in Fig. 5.22 indicate that the total deflection at the baseline ‘combustion’ loading (13 kN) is 0.15 mm (maximum deflection minus deflection at valve closure; readings taken from the LVDT on the hydraulic test machine). The deflection of the rig itself at this load was calculated using the equation for the deflection of a built-in flat circular plate with a central load, as outlined by Roark and Young [11], see equation (5.9) ymax = − Pp Rr 2 16π D , where D = Eb 3 12 1 − ν 2 ( ) −3Pp Rr (1 − ν 2 ) (5.9) 2 ⇒ ymax = 4π Eb3 (5.10) where ymax is the maximum vertical deflection (m), b is the plate thickness (m), ν is Poisson’s ratio, E is the modulus of elasticity (N/m2) and Rr is the plate radius (m). Fig. 5.22 Test rig displacement and actuator load 76 Valve and Seat Wear Testing Apparatus For the hydraulic loading apparatus at baseline ‘combustion’ loading: Pp = 13.2×103 N; Rr = 282×10−3 m; b = 20×10−3 m; ν = 0.3; and E = 210×109 N/m2. (These values were obtained from measurements taken on the rig and from properties of the material used in the manufacture of the rig.) Therefore, from equation (5.10) ymax = −135.7 µm The valve deflection at the baseline ‘combustion’ load is equal to the total deflection minus the test rig deflection, therefore valve deflection = total deflection − ymax = 14.3 µm Therefore, from equation (5.4) Wv = 0.18 J per combustion cycle While the errors apparent in calculating the deflection of the valve and test rig in this manner are potentially high due to the simplifications made, it should be emphasized that the purpose of the calculation was merely to provide a simple comparison between the work done in the test rig with that in the engine rather than provide an accurate assessment of the deflection. For ease of comparison the valve dynamic and loading characteristics for the 1.8 litre, IDI, diesel engine and the hydraulic test machine are shown in Table 5.6. As can be seen, for baseline conditions (see Table 5.5) the impact energy on valve closure and the work done on the valve during the ‘combustion cycle’ in the hydraulic loading apparatus are of the same order of those in the engine. The hydraulic loading apparatus is, however, unable to reproduce valve closing velocities that occur in the engine. In the engine an increase in the valve clearance, because of either poor adjustment or misalignment relative to the seat, can massively increase the valve closing velocity and energy. In the hydraulic loading apparatus, however, due to the operating constraints on the hydraulic actuator, such variation cannot be achieved even by introducing misalignment. 77 Automotive Engine Valve Recession Table 5.6 A comparison of valve dynamics and loading in a 1.8 l IDI diesel engine and a hydraulic test machine 1.8 litre IDI diesel engine Min. clearance Hydraulic test machine Max. clearance 13 kN combustion load Aligned Aligned 0.25mm misalignment Aligned 0.25 mm misalignment Closing velocity (mm/s) 288 375 1860 18 16 Closing impact energy (J) 0.0076 0.013 0.32 0.0081 0.0064 Work done during combustion (J) 0.12 0.91 It is clear that the hydraulic loading apparatus is able to simulate combustion loading of the valve well. As already observed, however, it is not possible to accurately reproduce the dynamics of valve closure seen in an engine over a range of closing velocities. Performance charts intended to simplify parameter selection were produced for the hydraulic loading apparatus at the baseline combustion load (13 kN) (as shown in Figs 5.23 and 5.24, respectively). Fig. 5.23 Valve closing velocity against amplitude for varying frequencies (13 kN combustion load) 78 Valve and Seat Wear Testing Apparatus Fig. 5.24 Impact energy on valve closure against amplitude for varying frequencies (13 kN combustion load) Sinusoidal lift curves for a range of amplitudes and frequencies were differentiated to calculate the corresponding velocity curves, see equations (5.7) and (5.8). The closing velocities taken from these curves were then used to calculate impact energies at valve closure, see equation (5.3). Both closing velocity and impact energy were then plotted against amplitude for a range of frequencies. 5.6 References 1. Narasimhan, S.L. and Larson, J.M. (1985) Valve gear wear and materials, SAE Paper 851497, SAE Trans., 94. 2. Zinner, K. (1963) Investigations concerning wear of inlet valve seats in diesel engines, ASME Paper 63-OGP-1. 3. Pope, J. (1967) Techniques used in achieving a high specific airflow for highoutput medium-speed diesel engines, Trans. ASME, J. Engng Power, 89, 265–275. 4. Matsushima, N. (1987) Powder metal seat inserts, Nainen Kikan, 26, 52–57, in Japanese. 79 Automotive Engine Valve Recession 5. Blau, P.J. (1993) Retrospective survey of the use of laboratory tests to simulate internal combustion engine materials tribology problems, ASTM STP Paper 1199. 6. Hofmann, C.M., Jones, D.R., and Neumann, W. (1986) High temperature wear properties of seat insert alloys, SAE Paper 860150, SAE Trans., 95. 7. Nakagawa, M., Ohishi, S., Andoh, K., Miyazaki, S., Mori, K., and Machida, Y. (1989) Development of hardsurfacing nickel-based alloy for internal combustion engine intake valves, JSAE Rev., 68–71. 8. Fujiki, F. and Makoto, K. (1992) New PM seat insert materials for high performance engines, SAE Paper 920570. 9. Malatesta, M.J., Barber, G.C., Larson, J.M., and Narasimhan, S.L. (1993) Development of a laboratory bench test to simulate seat wear of engine poppet valves, Tribol. Trans., 36, 627–632. 10. Stone, R. (1992) Introduction to internal combustion engines, Macmillan, Basingstoke. 11. Roark, R.J. and Young, W.C. (1975) Formulas for stress and strain, Fifth edition, McGraw-Hill, New York. 80 Chapter 6 Experimental Studies on Valve Wear 6.1 Introduction This chapter details bench test work designed to investigate valve and seat wear and isolate critical parameters, the data from which could be used to develop a model to predict valve recession. The aims of the bench test work were to: ● investigate the wear mechanisms occurring in passenger car diesel engine inlet valves and seat inserts; ● study the effect of engine operating conditions on wear; ● quantify the effect of lubrication at the valve/seat insert contact; ● test potential new seat insert materials and compare the results with those for existing materials. Testing was carried out using bench test apparatus designed to simulate the loading environment and contact conditions to which the valve and seat insert are subjected (as described in Chapter 5). 6.2 Investigation of wear mechanisms 6.2.1 Experimental details 6.2.1.1 Specimen details Valve and seat insert materials characteristic of those currently in use in passenger car diesel engine applications were selected for use in the tests, details of which are shown in Table 6.1. The geometries of the valves and seat inserts used are shown in Fig. 5.4. The ‘hardness’ of selected materials is shown in Table 6.2. 81 Automotive Engine Valve Recession Table 6.1 Valve and seat insert materials Valve material Description V1 Martensitic, low-alloy steel V2 Austenitic stainless steel Seat insert material Description S1 and S2 Sintered martensitic tool steel matrix with evenly distributed intra-granular spheroidal alloy carbides. Metallic sulphides are distributed throughout at original particle boundaries. The interconnected porosity in both is substantially filled with copper alloy throughout. S3 Cast, tempered martensitic tool steel matrix with a network of carbides uniformly distributed. Table 6.2 Hardness of valve and seat insert materials Material designation Hardness (Hv) V1 (Valve) 630 S2 (Seat insert) 490 S3 (Seat insert) 490 6.2.1.2 Test methodologies In order to investigate the effect of the frictional sliding of the valve on the seat insert under the action of the combustion pressure and the effect on wear of combining this with the impact of the valve on the seat insert during valve closure, the two test methodologies developed during initial testing of the hydraulic loading apparatus were employed (see Section 5.5.1.2). These were as follows. Frictional sliding The first methodology employed a triangular loading waveform to investigate the effect of combustion loading on valve and seat insert wear. The intention was to isolate the frictional sliding between the valve seating face and the seat insert, hence the valve seating face was in constant contact with the seat insert. Test parameters used for selected tests are shown in Table 6.3. These were based on the baseline established during initial testing of the hydraulic loading apparatus (see Section 5.5.1.3). Impact and sliding The second test methodology used a sinusoidal displacement waveform, with valve lifts of up to 1.2 mm and similar peak loads to those used in the frictional sliding tests, to investigate the effect of the impact of the valve on the seat insert as the valve closes, in combination with the combustion loading. Test parameters used for selected tests are shown in Table 6.4. Again, these were based on the baseline established during initial testing of the rig (see Section 5.5.1.3). When parameters such as combustion load were varied, the rig performance charts (see Figs 5.23 and 5.24) were used to select suitable frequencies and amplitudes in order to maintain a constant closing velocity. 82 Experimental Studies on Valve Wear Table 6.3 Frictional sliding test parameters Seat insert material Valve temp. (°C) Freq. (Hz) Load (kN) Load waveform Misalignment (mm) Rotation (r/min) No. of cycles S1 R.T. 20 13 Triangular 0 0 500 381 S1 R.T. 20 13 Triangular 0.5 0 506 521 S1 130 20 13 Triangular 0 0 500 016 S3 R.T. 20 13 Triangular 0 0 500 014 S3 R.T. 20 13 Triangular 0.5 0 500 020 S1 R.T. 5 0.6 Sinusoidal 0 1 66 540 Table 6.4 Impact and sliding test parameters Seat insert material Valve temp. (°C) Freq. (Hz) Valve lift (mm) Valve closing velocity (mm/s) Load (kN) Displacement (waveform) Misalignment (mm) No. of cycles Lubn. (Y/N) S3 R.T. 10 0.6 18 13 Sinusoidal 0 25 006 N S3 R.T. 10 0.6 18 13 Sinusoidal 0.25 18 067 N S2 R.T. 10 0.6 18 13 Sinusoidal 0 39 997 N S2 R.T. 10 0.6 18 13 Sinusoidal 0.25 24 009 N S2 130 10 0.6 18 13 Sinusoidal 0 24 371 N S3 130 10 0.6 18 13 Sinusoidal 0 24 011 N S3 R.T. 12 0.6 18 6 Sinusoidal 0.25 24 039 N S3 R.T. 10 0.6 18 18.5 Sinusoidal 0 100 027 N S2 R.T. 20 1.2 59 13 Sinusoidal 0 24 041 N S3 R.T. 10 0.6 18 13 Sinusoidal 0 100 179 N S2 R.T. 10 0.6 18 13 Sinusoidal 0 100 038 N S3 R.T. 10 0.6 18 13 Sinusoidal 0 100 000 Y In order to investigate the effect of impact of the valve on the seat insert as the valve closes on the motorized cylinder head, tests were run using two different seat insert materials, one cast (S3) and the other sintered (S2). In each test, different valve clearances were used in order to vary the valve closing velocity. Different clearances were achieved by using seat insert holders of varying thickness. Having chosen the desired valve closing velocities, the valve clearances required to achieve such velocities were determined using valve lift and velocity curves (such as shown in Fig. 5.14). The thickness of the seat insert holders could then be calculated from these clearances. Details of the clearances used and the closing velocities, energies, and forces at each clearance are shown in Table 6.5. Each test was run for 160 000 cycles (chosen in order to fit the test matrix into the available time). Valve rotation was 83 Automotive Engine Valve Recession measured for each valve during testing. This was achieved by marking the valve head and then timing a set number of rotations. Table 6.5 Motorized cylinder head test parameters Valve clearance (mm) Closing velocity (mm/s) Closing energy (J) 0.215 324 0.0096 0.415 960 0.0845 0.515 1600 0.234 0.615 2100 0.404 1.420 3680 1.239 6.2.1.3 Wear evaluation Wear evaluation was achieved using optical microscopy to study wear scars. Wear features were studied to establish wear mechanisms occurring in the valves and seat inserts. In addition, wear scar widths were measured both during and after the tests. 6.2.2 Results 6.2.2.1 Appearance of worn surfaces When employing the frictional sliding test methodology on the hydraulic loading apparatus to simulate combustion loading, the wear scars achieved on the valve seating faces appeared uneven. The formation of a brown oxide, characteristic of fretting wear (reciprocating sliding wear caused by very small displacements), was observed as well as debris at the edges of the wear scars. The wear scars on valves run against S2 sintered seat inserts were similar to those run against S3 cast seat inserts. Observation of the seating faces of both the sintered and cast seat inserts revealed the presence of scratches in the radial direction (see Fig. 6.1). These were similar to those previously observed [1]. The unevenness of the wear scars was caused by non-uniform contact between the valve and seat insert. The observations characteristic of fretting and sliding wear verified that the frictional sliding caused by the combustion load had been isolated. The consistency of these observations on both cast and sintered seat inserts indicates that the two materials have a similar resistance to sliding wear. When employing the impact and sliding test methodology on the hydraulic loading apparatus, and thus allowing the valve to lift from the seat during a cycle, the wear scars achieved on the seating face of the valves were more even. Examination of the wear scars of valves run against S3 cast seat inserts revealed evidence of deformation and the presence of a series of ridges and valleys formed circumferentially around the axis of the valve seating face (as shown in Fig. 6.2). These correspond to the ‘single wave formation’ previously described [1]. 84 Experimental Studies on Valve Wear RADIAL DIRECTION WEAR SCAR CIRCUMFERENTIAL DIRECTION RADIAL INDENTATIONS Fig. 6.1 Seat insert seating face showing indentations in the radial direction (valve material V1 run against sintered seat insert material S1 on hydraulic loading apparatus). Major seat diameter is towards top of figure. See Table 6.1 for details of materials. (Reprinted with permission from SAE paper 1999-01-1216 © 1999 Society of Automotive Engineers, Inc.) RADIAL DIRECTION WEAR SCAR CIRCUMFERENTIAL DIRECTION Fig. 6.2 Valve seating face showing a series of ridges and valleys formed circumferentially around the axis of the valve (valve material V1 run against cast seat insert material S3 on hydraulic loading apparatus). Major seat diameter is towards bottom of figure. See Table 6.1 for details of materials. (Reprinted with permission from SAE paper 1999-01-1216 © 1999 Society of Automotive Engineers, Inc.) This type of deformation was less prevalent on valves run against S2 sintered seat inserts. Evidence was found, however, of adhesive pick-up from seat inserts (see Figs 6.3 and 6.4). With a cast insert there appeared to be greater surface damage to the valve than the seat insert, whereas with a sintered seat insert there appeared to be greater insert damage. 85 Automotive Engine Valve Recession “PICK-UP” ON VALVE SEATING FACE WEAR SCAR Fig. 6.3 Valve seating face showing evidence of adhesive pick-up from the seat insert (valve material V1 run against sintered seat insert material S2 on hydraulic loading apparatus). Major seat diameter is towards bottom of figure. See Table 6.1 for details of materials ORIGINAL SURFACE HOLES FROM WHICH MATERIAL HAS BEEN PLUCKED WEAR SCAR Fig. 6.4 Seat insert seating face showing evidence of adhesive pick-up (valve material V1 run against sintered seat insert material S2 on hydraulic loading apparatus). Major seat diameter is towards top of figure. See Table 6.1 for details of materials Observation of the seating faces of both types of seat insert revealed that surface damage was more severe when combining impact with frictional sliding. A wear scar was seen to form and grow on sintered seat inserts, whereas on the cast seat inserts no obvious scar formed, although evidence was found of pitting and radial indentations. 86 Experimental Studies on Valve Wear The results achieved using the two different test methodologies, frictional sliding and impact and sliding, indicate that the effect of the two loads imposed on the valve have been isolated. The frictional sliding test was used to simulate the load imposed during combustion in the cylinder and the impact and sliding test was used to simulate the impact load on valve closure in combination with the combustion loading. The radial scratches on the seat inserts found when using the loading waveform are caused as the valve slides against the seat insert as it deflects under loading. The circumferential ridges and valleys found when using the displacement waveform are caused by a deformation or gouging process as the valve impacts against the seat insert. Both impact and sliding clearly have a large influence on valve recession. It is in combination, however, that they have the largest effect. In the tests with impact and sliding run in combination, it took a few thousand cycles to achieve surface damage attained in several hundred thousand cycles in the frictional sliding tests. Tests run on the motorized cylinder head (without any combustion loading) were intended to isolate the impact of the valve on the seat insert on valve closure. Evidence was found on valves run with a high closing velocity, however, that a small amount of sliding was occurring even in the absence of the combustion loading. It was also found on removing the valves, that a film of oil was present on the valve head and seating face. This was formed by lubricating oil leaking past the valve stem seals. As with tests run on the hydraulic loading apparatus using impact and sliding, valve wear was observed to be far more severe with S3 cast seat inserts, and insert wear was more severe with S2 sintered seat inserts. Examination of the wear scars of valves run against S3 cast seat inserts again revealed evidence of deformation. There was also a series of ridges and valleys formed circumferentially around the axis of the valve seating face, similar to those observed during testing on the hydraulic loading apparatus. Observation of the cast seat insert seating faces (see Fig. 6.5), however, revealed the presence of surface cracking and evidence of subsequent material loss, not previously observed. The wear features observed on both valves and seat inserts (deformation and surface cracking) are characteristic of processes resulting in wear loss due to single or multiple impact of particles [2] (see Fig. 6.6). Similar observations made during work on the wear of poppet valves operating in hydropowered stoping mining equipment led Fricke and Allen [3] to use a relationship of the same form as that used in erosion studies to model impact wear. Fricke and Allen [3] justified the use of such a relationship for impact wear of valves, citing work by Hutchings et al. [4] in which it was shown that erosion can be satisfactorily modelled by the impact of large particles. In their work, they used hard steel balls up to 9.5 mm in diameter. It was thus assumed that a relationship exists between impacts on a macroscale (greater than 1 mm) and impacts on a microscale (less than 1mm), such as those found typically in erosive wear. 87 Automotive Engine Valve Recession SURFACE CRACKING MATERIAL REMOVAL Fig. 6.5 Seat insert seating face (valve material V1 run against cast seat insert material S3 on motorized cylinder head). Major seat diameter is towards top of figure. See Table 6.1 for details of materials (a) (b) (c) Fig. 6.6 Processes resulting in wear loss due to single or multiple impact of particles: (a) extrusion of material at the exit end of impact craters; (b) surface cracking (microcracking); (c) surface and subsurface fatigue cracks due to repeated impact [2] 6.2.2.2 Formation of wear scars Unlike the valve wear scars observed when applying the frictional sliding test methodology on the hydraulic loading apparatus, those seen when employing the impact and sliding test methodology were uniform and seen to increase in width as the tests proceeded (it should be noted that during these tests no valve rotation was used). The progression of wear when using a sintered seat insert differed from that when using a cast seat insert (as shown in Fig. 6.7). The observations made indicate that the introduction of impact caused a bedding-in process to occur, improving the uniformity of the valve wear scars and causing them to increase in width. Figure 6.7 shows that, initially, there was a rapid increase in the wear scar width; this progression then slowed until bedding-in was achieved. It also shows 88 Experimental Studies on Valve Wear how the cast and sintered seat inserts responded differently to the introduction of impact. When using a cast insert, a high initial progression of wear was observed compared to the more gradual progression when using a sintered seat insert. Fig. 6.7 Average wear scar width for a V1 valve run against an S2 sintered seat insert and a V1 valve run against an S3 cast seat insert using impact and sliding on the hydraulic loading apparatus. See Table 6.1 for details of materials Recession is the main parameter of interest to engine developers. Wear scar measurement, while providing information on the progression of wear during a test, gave no indication of the actual magnitude of the wear or valve recession that had occurred as no account could be taken of initial contact conditions. In order to obtain a more useful indication of wear, give an improved comparison of test rig data, and provide a means to compare test rig data with valve recession data from engine tests, a method was developed for estimating a recession value from seat insert wear scar data. Two different wear ‘cases’ were observed during testing. 1. The valve and seat insert seating face angles differed slightly and a wear scar was seen to form and grow on the valve and seat insert until full contact with the seat insert seating face was achieved, the width of which then began to increase as the test progressed. 2. The valve initially made full contact with the seat insert seating face, which was then seen to grow as the test progressed. 89 Automotive Engine Valve Recession Figure 6.8 shows a comparison of valve recession for impact and sliding tests run with cast and sintered seat inserts on the hydraulic loading apparatus (replot of data from Fig. 6.7), calculated using equations that give recession and wear volume as a function of the seat insert wear scar width or seating face width [5]. This shows that valve recession is higher when using sintered seat inserts despite the lower wear scar widths observed. Fig. 6.8 Valve recession for a V1 valve run against an S2 sintered seat insert and a V1 valve run against an S3 cast seat insert using impact and sliding on the hydraulic loading apparatus. See Table 6.1 for details of materials Figure 6.9 shows a comparison of valve recession (calculated from wear scar and seating face width data) for tests run with cast and sintered seat inserts on the motorized cylinder head. It can be seen, as with the hydraulic loading apparatus results, that greater valve recession occurs when using a sintered seat insert. The tests run on the motorized cylinder head were intended to isolate the impact of the valve on the seat insert on valve closure. These results, therefore, indicate that the cast seat insert material has a greater resistance to impact than the sintered material. It was clear from the wear scars observed on both valves and seat inserts after tests run on both the hydraulic loading apparatus and the motorized cylinder head, that the wear mechanisms were different for the cast and sintered seat inserts. A sintered seat insert appeared to wear at a greater rate than the valve, i.e. the valve was ‘bedding’ into the seat insert (see Fig. 6.10). When using a cast seat insert, however, it appeared that the valve was wearing more than the seat insert and that the seat insert was ‘bedding’ into the valve (see Fig. 6.10). 90 Experimental Studies on Valve Wear Fig. 6.9 Valve recession for a V1 valve run against an S2 sintered seat insert and a V1 valve run against an S3 cast seat insert with a valve closing velocity of 960 mm/s on the motorized cylinder head. See Table 6.1 for details of materials (a) SINTERED SEAT INSERT (b) CAST SEAT INSERT Fig. 6.10 Sintered seat insert versus cast seat insert. (Reprinted with permission from SAE paper 1999-01-1216 © 1999 Society of Automotive Engineers, Inc.) Both the cast and sintered seat insert materials have similar hardness. It has been shown, however, that with impact wear there is no direct correlation between the material loss and hardness [6]. The fracture toughness of a material has been shown to be one of the important factors controlling impact wear [7]. The cast insert material has a higher fracture toughness than the sintered material and is, therefore, more resistant to the impact load imposed during valve closing, which could explain the two different bedding-in processes observed. It should be noted that the cast and sintered materials 91 Automotive Engine Valve Recession did not have the same composition, which could also help to explain the differences in response to impact. Part of the reason for the development of sintered seat inserts was to allow the incorporation of solid lubricants into the matrix to reduce sliding wear in ‘dry’ running conditions brought about by the reduction in lead in gasoline. The frictional sliding tests have shown that sintered seat insert materials perform adequately under these conditions. However, in order to increase the performance of sintered materials in seat insert applications, the issue of resistance to impact may need to be readdressed. Material choice is clearly critical in addressing valve and seat insert wear problems. In deciding which material combination to use, consideration should be given to deciding which is more preferable: greater valve wear or greater seat insert wear. Clearly, replacing valves is less costly than replacing seat inserts or an entire cylinder head. Therefore, work needs to be focussed on reducing seat wear while keeping valve wear at an acceptable level. Ultimately, however, it would be preferable not to have to replace either and to reduce the adjustment required on valve clearances, as a result of recession, to an absolute minimum. In selecting materials, consideration must be given to the relative resistance required to sliding and impact wear. This can be determined by looking at the engine operating parameters. If a high valve mass or closing velocity is being used then resistance to impact is critical. However, if high peak combustion loads are in use, then resistance to sliding could be more important. 6.2.2.3 Comparison with engine recession data Hydraulic loading apparatus wear scar width data were used to calculate recession values using the equations relating recession to wear volume [5] (see Section 6.2.2.2). A line was fitted to the recession data using an exponential relationship. This was then extrapolated in order to allow a comparison with engine test data. Figure 6.11 compares extrapolated hydraulic loading apparatus recession data for a test run with a V1 valve against an S2 sintered seat insert and engine test data for a V1 valve run against an S1 sintered seat insert, alongside hydraulic loading apparatus and engine test data for tests run with a V1 valve against an S3 cast seat insert (hydraulic loading apparatus data points were calculated to correspond with available engine test data points, hence the extrapolated line is not a smooth curve). The engine tests were run at a speed of 4800 r/min. Good correlation was achieved between the hydraulic loading apparatus and engine test rig data, which further established the validity of the test methodology and indicates the suitability of calculating recession values from wear scar data. The differences between the hydraulic loading apparatus test operating conditions and those found in an engine, however, should have given rise to higher recession rates in the test rig (hydraulic loading apparatus tests were run unlubricated and service 92 Experimental Studies on Valve Wear temperatures were not replicated). The fact that this did not occur could have been because of the differences in valve dynamics in the hydraulic loading apparatus and an engine, as explained in Section 5.5.3. It should also be noted that the two sintered materials used differ in composition. S1 contains a solid lubricant while S2 does not. Fig. 6.11 Comparison of hydraulic loading apparatus results with engine test data (solid line – engine test data; broken line – extrapolated test rig data. See Table 6.1 for details of materials 6.2.2.4 Lubrication of valve/seat interface When running the hydraulic loading apparatus with a steady supply of lubricant to the valve/seat insert interface, evidence of wear on the valve or seat insert seating faces was barely visible. Figure 6.12 compares valve recession (calculated from wear scar data) for tests run with and without lubrication. Although the lubricant was supplied in a larger amount and at a lower temperature than would be experienced in an engine, this provides an approximate quantification of the effect of lubrication on valve and seat insert wear. For this particular case, recession is approximately 3.5 times higher for the test run without lubrication. It is thought, due to the relatively slow sliding velocity and high load, that the lubrication mechanism occurring during the sliding at the valve/seat interface is that of boundary lubrication. In such a lubrication regime the lubricant film thickness is too small to give full fluid film separation of the surfaces, and surface asperities come into contact. 93 Automotive Engine Valve Recession One of the reasons for the work under discussion is the impending reduction of oil in the air stream of diesel engines which will reduce the amount of lubricant at the inlet valve/seat interface. These results give an indication of the increase in wear likely when this occurs. Fig. 6.12 Valve recession for V1 valves run lubricated and unlubricated against an S3 cast seat insert using impact and sliding on the hydraulic loading apparatus. See Table 6.1 for details of materials 6.2.2.5 Misalignment of valve relative to seat It was found that the magnitude and uniformity of the wear when running valve and seat inserts in an ‘aligned’ position on the hydraulic loading apparatus were principally affected by the initial contact conditions. Slight misalignments caused by small differences in roundness or seating face angles led to uneven wear. This has also been observed in engine testing. Differences in contact area around a valve seating face can also lead to non-uniform heat transfer from the valve head to the seat insert, causing hot spots or thermal distortions that worsen the problem. Misaligning the valve relative to the seat insert in the hydraulic loading apparatus produced a crescent-shaped wear scar on the valve seating face (as sketched in Fig. 6.13). The widest point of the scar was at the point of initial contact with the seat insert. At this point, the wear was more severe than was achieved when the valve was aligned. When using impact and sliding to simulate impact and combustion loading, the deformation observed on the valve run against a cast insert was more severe at the point 94 Experimental Studies on Valve Wear of initial contact than when the valve was aligned with the seat insert. The equivalent point on the seat insert also showed evidence of severe wear. Plastic deformation (not observed with an aligned valve) was also found on the valve seating face when misaligning the valve relative to a sintered seat insert (as shown in Fig. 6.14). The comparison of wear scar widths for both aligned and misaligned valves run against S2 sintered seat inserts, shown in Fig. 6.15, gives an indication of the increase in the wear scar width, at the point of initial contact, when misaligning the valve. (a) ALIGNED VALVE (b) MISALIGNED VALVE WEAR SCARS Fig. 6.13 Aligned versus misaligned valve wear scars WEAR SCAR Fig. 6.14 Valve seating face (valve material V1 run misaligned against sintered seat insert material S2 on hydraulic loading apparatus). Major seat diameter is towards bottom of figure. See Table 6.1 for details of materials 95 Automotive Engine Valve Recession Fig. 6.15 Valve wear scar width for a V1 valve run aligned and misaligned against an S2 sintered seat insert using impact and sliding on the hydraulic loading apparatus. See Table 6.1 for details of materials The deformation caused by impact on valve closure is increased, as misaligning the valve decreases the initial contact area between the valve and seat insert. The effect of the frictional sliding is increased and, as a result, the wear scar width is increased, since misalignment leads to an increase in the sliding distance as the valve head flexes under the action of the combustion loading. The magnitude of misalignment was an arbitrary value taken to investigate the effect of valve misalignment (0.25 mm). Actual values have not been measured. Depending, however, on the accuracy of the seat machining and the flexibility of the engine head, it is conceivable that such misalignment may occur in an engine and, therefore, be a source of valve recession. 6.2.2.6 Effect of combustion load When using the hydraulic loading apparatus, increasing the applied combustion loading while maintaining the closing velocity increased the severity of the wear on both the valves and the seat inserts. It was also observed that at higher loads there were a greater number of radial scratches present on the seat insert seating faces. Figure 6.16 shows recession rates (calculated from wear scar widths) for valves run against cast seat inserts for combustion loads of 6 kN, 13 kN, and 18.5 kN. The increase in wear as the combustion load rises is caused by the increasing effect of frictional sliding as a result of the increase in the combustion loading. The lack of recession at 13 kN for 10 000 cycles was an anomaly. Wear damage was observed during this period, but no measurable wear scar. The recession observed from 10 000 cycles compared well with other tests run at 13 kN. 96 Experimental Studies on Valve Wear Fig. 6.16 Valve recession for V1 valves run at three different combustion loads against S3 cast seat inserts using impact and sliding on the hydraulic loading apparatus. See Table 6.1 for details of materials 6.2.2.7 Effect of closing velocity Tests run on the motorized cylinder head – varying the valve closing velocity, clearly indicated that increasing valve closing velocity increased both valve and seat insert wear. As shown in Figs 6.17 and 6.18, valve recession (calculated from wear scar and seating face width data) rapidly increased, when using both cast and sintered insert materials, as the closing velocity was increased. 97 Automotive Engine Valve Recession Fig. 6.17 Valve recession for V1 valves run with different closing velocities against S3 cast seat inserts on the motorized cylinder head. See Table 6.1 for details of materials Fig. 6.18 Valve recession for V1 valves run with different closing velocities against S2 sintered seat inserts on the motorized cylinder head. See Table 6.1 for details of materials 98 Experimental Studies on Valve Wear Figure 6.19 shows wear scars for valves run against cast seat inserts at three different closing velocities. Deformation and ridge formation was visible as well as fine pitting at higher velocities. The appearance of the wear scar at the lowest velocity shown [Fig 6.19(a)] may have been a due to waviness of the unworn surface. It is interesting, however, that two ridges formed at 1600 mm/s [Fig. 6.19(b)] and only one at 2100 mm/s [Fig 6.19(c)]. Fig 6.19 Seating faces of valves run with closing velocities of: (a) 960 mm/s; (b) 1600 mm/s; (c) 2100 mm/s. Valve material V1 run against cast seat insert material S3 on motorized cylinder-head. Major seat diameter is towards bottom of figure. See Table 6.1 for details of materials Figure 6.20 clearly shows that the cast seat insert wear also became more severe as the valve closing velocity was increased. At 960 mm/s [Fig. 6.20(a)] the original machining marks are still visible. As the velocity was increased, however, these became less visible and at 2100 mm/s [Fig. 6.20(c)] they have been completely worn away and surface cracking and subsequent material loss is visible. Figure 6.21 (a replot of data from Figs 6.17 and 6.18) shows the relationship between recession and valve closing velocity for both sintered and cast seat inserts (at 160 000 cycles). Valve recession is roughly proportional to velocity squared for the cast insert material. The increase in recession observed is similar to that described by Zum-Gahr [2] for the increase in erosion rate with impact velocity of small particles. 99 Automotive Engine Valve Recession ORIGINAL MACHINING MARKS SURFACE CRACKING MATERIAL REMOVAL Fig. 6.20 Seating faces of seat inserts run against valves with closing velocities of: (a) 960 mm/s; (b) 1600 mm/s; (c) 2100 mm/s. Valve material V1 run against cast seat insert material S3 on motorized cylinder head. Major seat diameter is towards top of figure. See Table 6.1 for details of materials 6.2.2.8 Valve rotation When using rotation on the hydraulic loading apparatus, an even wear scar was achieved. Debris was observed at the valve/seat insert interface during testing (as shown in Fig. 6.22). The material was dark and powder-like in nature. It was being removed from the interface under the action of the rotation. It is possible, therefore, that valve rotation promotes debris removal. This would be useful in reducing abrasive wear caused by wear debris otherwise trapped in the interface and in reducing the build-up of deposits and the subsequent formation of hot spots. Examination of the valve seating face of a rotated valve revealed the presence of circumferential grooves. These were caused by rotation of the valve either on closing or while the valve was closed. 100 Experimental Studies on Valve Wear Fig. 6.21 Valve recession against closing velocity for a V1 valve run against an S2 sintered seat insert and a V1 valve run against an S3 cast seat insert on the motorized cylinder head (at 160 000 cycles). See Table 6.1 for details of materials Fig. 6.22 Debris at valve/seat insert interface during valve rotation Valve rotational speeds were measured during all tests run on the motorized cylinder head. This was achieved by marking the valve heads and then timing a set number of revolutions. Valve rotation was seen to vary between 8 r/min and 18 r/min, depending on test conditions and seat materials. Valve rotation was observed to decrease with an increase in lubricant supply to the cam/follower interface. This can be explained by looking at the mechanism by which valve rotation occurs. The cam is offset from the follower in order to prevent localized wear at the contact area. The offset cam also promotes valve rotation as a result of the follower rotation, when the valve and follower are in contact. An increase in the 101 Automotive Engine Valve Recession lubricant supply to the valve/follower contact will reduce the coefficient of friction. The frictional force will, therefore, also be reduced. This will decrease the rotational speed of the follower, and hence the valve will also rotate at a slower speed. As shown in Fig. 6.23, valve rotation was also observed to decrease with an increase in valve closing velocity. Fig. 6.23 Valve rotation against valve closing velocity for a V1 valve run against an S2 sintered seat insert and a V1 valve run against an S3 cast seat insert on the motorized cylinder head. See Table 6.1 for details of materials It is not possible to assess whether rotation influenced valve or seat insert wear in the motorized cylinder head as valve rotation could not be varied while keeping other test parameters constant. 6.2.2.9 Effect of temperature It has been shown that the general trend is that wear decreases as the temperature is increased from 150 to 600 °C [8]. This was thought to be because of oxide formation at high temperatures preventing metal-to-metal contact, and thus reducing adhesive wear. For this reason tests run on the hydraulic loading apparatus were not designed to study the effect of temperature on valve wear. The majority of the tests were run at room temperature. Running at an increased temperature of 130 °C using a hot air supply (see Fig. 5.2) was found to have little effect on the wear rate of the valves or seat inserts (as shown in Fig. 6.24). The valves and seat inserts exhibited similar wear features to those run at lower temperatures. It was not possible to run the hydraulic loading apparatus at temperatures higher than 130 °C. 102 Experimental Studies on Valve Wear Fig. 6.24 Valve recession for V1 valves run at two different temperatures against S2 sintered seat inserts using impact and sliding on the hydraulic loading apparatus. See Table 6.1 for details of materials At higher temperatures it is possible that the hardness of the valve or seat insert material may be reduced due to tempering, which could lead to increased valve recession or even fatigue failure of valves. Such temperatures are usually caused by a reduction in heat transfer from the valve head area as a result of deposit formation. 6.3 Seat insert materials Having analysed the root causes of the valve recession mechanism and considered the implications for valvetrain design and seat insert and valve materials, it was decided to investigate the potential for reducing valve recession by improving existing seat insert materials for use in inlet valve applications, and to identify potential new seat insert materials. Reduction of valve recession can also be achieved by introducing design changes, but this is a more time-consuming process and would have implications for other aspects of engine performance. The aim of this work was to test potential new seat insert materials and compare the results with those for existing materials. The wear mechanisms were investigated and the material performance was assessed. Testing was carried out using the bench test apparatus designed to simulate the loading environment and contact conditions to which a valve and seat insert are subjected (as described in Chapter 5). 103 Automotive Engine Valve Recession 6.3.1 Experimental details 6.3.1.1 Valve and seat insert materials Selection of the potential new seat insert materials was based upon knowledge of the prevalent wear mechanisms and problems such as valve misalignment due to thermal distortions and valve seat loosening. The work described in Section 6.2 showed that the two main causes of valve recession are impact wear and frictional sliding and that the effect of both are increased by misalignment of the valve due to thermal distortions in the seat. Before selecting materials it was, therefore, necessary to identify the properties required to reduce the effect of these mechanisms. One of the most important factors controlling impact wear has been shown to be fracture toughness [7]. The higher the fracture toughness of a material, the lower the impact wear. The effect of frictional sliding can be reduced by increasing the lubricity of the seat material, especially under dry running conditions. The problem of thermal distortion can be reduced by improving the conduction of heat through the seat area to the cooling channels in the cylinder head (a valve transfers approximately 75 per cent of the heat input to the top-of-head through its seat insert into the cylinder head [9]). The most important properties required of a seat insert material in order to decrease the possibility of valve recession occurring were, therefore, considered to be: ● high fracture toughness; ● good lubricity under dry running conditions; ● high thermal conductivity. The properties required of a seat insert material in order to reduce the likelihood of insert ‘drop-out’ occurring have been listed as [10]: ● high compressive yield strength; ● low modulus of elasticity; ● high thermal conductivity; ● normal thermal expansion. Three materials were selected using these criteria. Each material does not fulfil all the criteria listed, but as each valve recession problem has its own set of operating parameters and design features it was thought best to test materials with a range of properties. Maraging steel was selected for its high resistance to impact. EN 1A, a freecutting mild steel, was selected as it was thought the alloying elements incorporated to increase the machinability would help decrease the effect of frictional sliding while still giving good resistance to impact. Oilite, an oil-impregnated, porous, bronze material, was selected exclusively for its resistance to frictional sliding wear. Two types of cast iron were chosen to represent seats formed in-situ within the cylinder head: a grey cast iron and a ductile cast iron. 104 Experimental Studies on Valve Wear It was decided not to consider ceramic seat insert materials in this investigation as they are expensive to machine and the sponsor was more interested in low-cost options. It is possible that the wear properties of ceramic materials will be examined in the future. Results for the materials listed above were compared with those for two of the seat insert materials used in hydraulic loading apparatus work described in Section 6.2. The first was a sintered material (S2) consisting of a martensitic tool steel matrix with evenly distributed intra granular spheroidal alloy carbides, and the second a cast material (S3) consisting of a tempered martensitic tool steel matrix with a network of carbides evenly distributed. All selected seat materials and their properties (where available) are shown in Table 6.6. Other selection criteria, such as cost and machinability, will be discussed further on. Valves made from a martensitic, low-alloy steel (V1) were selected for use in the tests. 6.3.1.2 Specimen details The geometries of the valves and seat inserts used are shown in Fig. 5.4. The materials not available as seat inserts and the cylinder head materials were made up into specimens as shown in Fig. 5.5. These could be used in both test rigs. Table 6.6 Properties of the seat materials Material Fracture toughness (MNm3/2) Hardness (Hv) Tensile strength (MN/m2) Thermal conductivity (W/m°C) Coeff. of thermal expansion (µm/m°C) Modulus of elasticity (GN/m2) Cast seat insert (S3) 490 250–400 40 10.3–12.6 120 Sintered seat insert (S2) 490 Grey cast iron (250) 12 193 250 46 10.8 (20–400 °C) 110 Ductile cast iron (600) 20 319 600 32.5 (300 °C) 12.5 (20–400 °C) 174 205 360 43 (400 °C) 13.95 (20–400 °C) 185 358 1800 20.9 (100 °C) 10.1 (24–284 °C) 186 89.4 96.5 Free-cutting mild steel (EN 1A) Maraging steel (250) Oilite 120 17.64 (20–30 °C) 6.3.1.3 Test methodologies The impact and sliding test methodology (see Section 5.5.1.2) was utilized on the hydraulic loading apparatus in order to investigate the effect of the impact of the valve on the seat insert as the valve closes, in combination with the combustion loading. Test parameters used for selected tests are shown in Table 6.7. 105 Automotive Engine Valve Recession Table 6.7 Seat insert material impact and sliding test parameters Seat insert material Valve temp. (°C) Freq. (Hz) Valve lift (mm) Valve closing velocity (mm/s) Load Displacement Misalignment (kN) waveform (mm) No. of cycles Lubn. (Y/N) Grey cast iron (250) R.T. 10 0.6 18 13 Sinusoidal 0 100 006 N Ductile cast iron (600) R.T. 10 0.6 18 13 Sinusoidal 0 100 013 N Free cutting mild steel (EN 1A) R.T. 10 0.6 18 13 Sinusoidal 0 100 021 N Maraging steel (250) R.T. 10 0.6 18 13 Sinusoidal 0 100 074 N Oilite R.T. 10 0.6 18 13 Sinusoidal 0 100 212 N Grey cast iron (250) R.T. 10 0.6 18 18.5 Sinusoidal 0 100 008 N Ductile cast iron (600) R.T. 10 0.6 18 18.5 Sinusoidal 0 100 107 N In order to investigate the effect of impact of the valve on the seat insert materials as the valve closes, tests were run on the motorized cylinder head using seats manufactured from grey cast iron, ductile cast iron, free-cutting mild steel, maraging steel, and oilite. The same clearance was used for each valve, giving a constant closing velocity for each material. Details of the clearance used and the closing velocity, energy, and force are shown in Table 6.8. The tests were run for 100 000 cycles. Valve rotation was measured for each valve during the tests. Table 6.8 Motorized cylinder head test parameters Valve clearance (mm) 0.515 Closing velocity (mm/s) 1600 Closing energy (J) 0.234 6.3.2 Results Figure 6.25 shows valve recession (calculated from wear scar data) from tests run on the hydraulic loading apparatus using impact and sliding for the seat materials listed in Table 6.6. The materials selected mainly for their resistance to frictional sliding exhibited the largest amount of recession. Of these, free-cutting mild steel provided the best performance, but this material was expected to have a greater resistance to impact than grey cast iron and oilite. It is clear that resistance to impact wear is a key material property in reducing the likelihood of valve recession. 106 Experimental Studies on Valve Wear Fig. 6.25 Valve recession for V1 valves run against various seat materials using impact and sliding on the hydraulic loading apparatus The ductile cast iron recessed less than might have been expected. Given its low fracture toughness compared to maraging steel, a higher recession relative to this material would have been anticipated. The low wear may be explained by its higher hardness combined with the presence of graphite, which can act as a solid lubricant, providing improved resistance to sliding wear compared with maraging steel. It was not, however, unexpected that the ductile cast iron recessed less than the grey cast iron and the free-cutting mild steel. Additives, introduced into the molten iron just before casting of ductile cast iron, cause the graphite to grow as spheres rather than flakes, as in grey cast iron. Ductile cast iron is consequently stronger and more ductile than grey cast iron, giving it an increased resistance to impact. On examination of the wear scars from all the tests it was found that the material combinations exhibiting the greatest recession gave the least severe valve wear and the most severe seat wear. Wear features such as radial indentations were again observed on seat seating faces. On examination of the oilite seat before testing it was found that the seating face was covered in cracks and in places chunks of the material had been torn away on machining. However, after testing it was found that the cracks had disappeared. The only explanation was the occurrence of material flow at the seating face. Observation of the valve wear scar revealed the presence of scratches in the radial direction across the entire width, not observed with any other material combination. 107 Automotive Engine Valve Recession Figure 6.26 illustrates the increase in recession recorded when increasing the combustion loading while maintaining the valve closing velocity, using a ductile cast iron seat. Again radial indentations on the seat seating face increased in number at the higher load. Fig. 6.26 Valve recession for V1 valves run at two different combustion loads against ductile cast iron seats using impact and sliding on the hydraulic loading apparatus Figure 6.27 shows the valve recession (calculated from wear scar data) from tests run on the motorized cylinder head. Maraging steel and free-cutting mild steel have not been included as they exhibited exceptionally high recession in just the first 50 000 cycles (0.295 mm and 0.23 mm, respectively). They were both located in the seat position furthest from the driving belt and pulley system (see Fig. 5.12). It has been speculated that the unexpected level of impact wear was caused by torsional vibration of the camshaft, which only affected this seat position. In normal operation, in an engine, the camshaft would have a balancing flywheel at the opposite end to the driven end to eliminate such vibrations, but this was not present on this rig. The follower at this seat position eventually disintegrated and damage was also observed on the cam. This had not been a problem on previous tests as this valve position had not been utilized. Unfortunately, time constraints meant the tests could not be repeated to obtain more data for the two materials or to further study the effect. This could give cause, however, to believe that valve position influences the magnitude of wear. Instances where one valve has recessed far more than the others in the same cylinder head have been observed during engine testing (as described in Section 1.2). The cause, however, was not identified. Further work would be required to establish whether valve position affects recession and, if so, the mechanism which leads to its occurrence. 108 Experimental Studies on Valve Wear Fig. 6.27 Valve recession for V1 valves run against various seat materials with a valve closing velocity of 1600 mm/s on the motorized cylinder head It can be seen that, relatively, the wear follows the same pattern on the motorized cylinder head as it did on the hydraulic loading apparatus, with one exception. On the hydraulic loading apparatus, more recession was recorded for the grey cast iron seat than the sintered seat insert, whereas on the motorized cylinder head the sintered seat insert exhibited greater wear. Clearly the sintered material is less resistant to impact wear than grey cast iron, but more resistant to sliding wear. A similar relative position for the ductile cast iron verifies that as expected it has high resistance to impact. This explains the good performance shown by this material in the hydraulic loading apparatus. An accurate relationship between valve recession and seat hardness or seat fracture toughness could not be identified from the range of materials tested. With the little fracture toughness data for the materials available and ignoring the low recession exhibited when using a ductile cast iron seat, it could be said that wear decreases as fracture toughness increases. Ignoring the S2 sintered seat insert material, recession could be said to be inversely proportional to seat hardness. As well as looking at resistance to impact and sliding in the materials selection process, machinability and cost of materials must also be considered. Some of the properties which give low tool wear and hence good machinability, listed by Mills and Redford [11], are: 109 Automotive Engine Valve Recession 1. low yield strength and low work hardening rate; 2. high thermal conductivity; 3. low chemical reactivity with tool or atmosphere; 4. low fracture toughness. Clearly several of these properties are in direct contradiction to those required of the valve/seat materials in order to resist impact and sliding wear. It was noted that while, generally, for a group of similar materials machinability improves as the fracture toughness of the workpiece reduces, there are exceptions. Spheroidal (ductile) cast irons having higher fracture toughness than similar flake graphite cast irons actually give lower cutting-tool wear rates. The efficiency with which materials are machined can be measured effectively by assessing the power required to machine a unit volume of material in unit time (specific power consumption). Data in the Metals Handbook of Machining [12] gives power required for machining for a range of engineering materials. Table 6.9 gives approximate powers for the seat materials tested. Approximate costs per kilogram for the materials [13] are also shown in Table 6.9. Considering all the factors discussed (wear resistance, machinability, and cost) two materials stand out as potential seat/seat insert materials. These are maraging steel and ductile cast iron. Both gave far higher wear resistance than the two seat insert materials tested. While more power is consumed machining the ductile cast iron it is relatively cheap compared to the other materials. Maraging steel is, however, both expensive and difficult to machine. Ductile cast iron has the added advantage in that it can be alloyed with small amounts of nickel, molybdenum, or copper to improve its strength and hardenability. Larger amounts of silicon, chromium, nickel, or copper can be added for improved resistance to corrosion or for high temperature applications. This may mean it could also be suitable for exhaust valve applications, making it a viable option for use in manufacturing the entire cylinder head. Table 6.9 Approximate power consumption and costs for the seat materials Seat material Power required for turning (Watts per mm3 per min ×104) 1 Cost per kg (£) 2 Oilite Maraging steel Free-cutting mild steel Grey cast iron Ductile cast iron 135 420 210 180 540 1.1–1.4 1.2–1.8 0.25–0.35 0.2–0.35 0.2–0.35 1 ASM [12] 2 Granta Design Ltd [13] 110 Experimental Studies on Valve Wear 6.4 Conclusions Stating the root cause of valve recession is difficult. Each valve recession problem will have its own unique set of operating parameters, design features, and material combinations. The investigation, however, has clearly shown that: 1. The bench test apparatus provides a valid simulation of the wear of both inlet valves and seat inserts used in automotive diesel engines. 2. The valve and seat insert wear problem involves two distinct mechanisms: impact of the valve on the seat insert on valve closure and sliding of the valve on the seat insert under the action of the combustion pressure. 3. Wear increases with valve closing velocity, combustion load, and misalignment of the valve relative to the seat insert. When using S3 cast seat inserts, valve recession was proportional to the square of the closing velocity (see Fig. 5.21). 4. Lubrication of the valve/seat interface reduced valve recession, on the material combination tested, by a factor of 3.5. 5. Resistance to impact is a key material property in reducing the likelihood of valve recession. 6. Considering all the factors discussed (resistance to impact and sliding wear, machinability and cost), two materials stand out as potential seat/seat insert materials. These are maraging steel and ductile cast iron. Both gave far higher wear resistance than the two seat insert materials tested. While more power is consumed machining the ductile cast iron it is relatively cheap compared to the other materials. Maraging steel is, however, both expensive and difficult to machine. 7. Data is available for the development of a semi-empirical model for predicting valve recession. 6.5 References 1. Van Dissel, R., Barber, G.C., Larson, J.M., and Narasimhan, S.L. (1989) Engine valve seat and insert wear, SAE Paper 892146. 2. Zum-Gahr, K.H. (1987) Microstructure and wear of materials, Tribology Series No. 10, Elsevier, Amsterdam. 3. Fricke, R.W. and Allen, C. (1993) Repetitive impact-wear of steels, Wear, 163, 837-847. 4. Hutchings, I.M., Winter, R.E., and Field, J.E. (1976) Solid particle erosion of metals: the removal of surface material by spherical objects, Proc. R. Soc. Lond. Ser. A., 348, 379-392. 5. Lewis, R. (2000) Wear of diesel engine inlet valves and seats, PhD Thesis, University of Sheffield, UK. 6. Rabinowicz, E. (1995) Friction and wear of materials, Second Edition, John Wiley and Sons, New York. 111 Automotive Engine Valve Recession 7. Kawachi, R., Tujii, H., Kawamoto, M., and Okabayashi, K. (1983) On the impact wear of carbon steel and cast iron, J. Japan Inst. Metals, 47, 225–230. 8. Wang, Y.S., Narasimhan, S., Larson, J.M., Larson, J.E., and Barber, G.C. (1996) The effect of operating conditions on heavy duty engine valve seat wear, Wear, 201, 15–25. 9. Giles, W. (1971) Valve problems with lead free gasoline, SAE Paper 710368. 10. Lane, M.S. and Smith, P. (1982) Developments in sintered valve seat inserts, SAE Paper 820233. 11. Mills, B. and Redford, A.H. (1983) Machinability of engineering materials, Applied Science Publishers, London. 12. ASM (1967) Metals handbook: Vol. 3 – Machining, ASM International, Ohio. 13. Granta Design Ltd (1994) Cambridge materials selector software, Version 2.02. 112 Chapter 7 Design Tools for Prediction of Valve Recession and Solving Valve Failure Problems 7.1 Introduction An important aspect of research is the industrial implementation of the results. This chapter looks at the provision of tools that will enable the results of the review of literature, analysis of failed specimens, and bench test work to be applied in industry to assess the potential for valve recession and solve problems that occur more quickly. The development of a semi-empirical model for predicting valve recession is described. Flow charts outline steps to be used to reduce the likelihood of recession occurring during the design process, as well as offering solutions to problems that do occur. The model developed for predicting valve recession was based on the mechanisms of valve and seat wear determined during investigations carried out using purpose built valve wear test apparatus (as detailed in Chapter 6). Existing models for the mechanisms of wear observed were combined to produce the final model. Constants required for the model were taken from curves fitted to experimental data. An iterative software program called RECESS1 was developed to run the model. The model was then validated against both engine tests and tests run on the hydraulic loading apparatus. The flow charts were developed from data collected from the review of literature, analysis of failed specimens, and modelling. 7.2 Valve recession model 7.2.1 Review of extant valve wear models Existing models developed for assessing the likelihood of the occurrence of valve wear have been simplistic. They focussed on the frictional sliding between the valve and seat 1 RECESS is a commercially available software package and database. Contact the authors for further details. 113 Automotive Engine Valve Recession under the action of the combustion pressure and took no account of the impact of the valve on the seat on valve closure. The valve wear factor Fw derived by Pope [1] for medium-speed diesel engines, equation (7.1) varies directly with the coefficient of friction. Other variables are dependent on engine design. It was found that a wear factor of 200 gave a satisfactory service life and wear factors greater than 250 gave a poor service life. The length of a ‘satisfactory life’, however, was not defined. Fw = µ P 2 N D 5 tan θ 2 E B C t tm (7.1) where µ is the coefficient of friction between valve and seat, P is the maximum cylinder pressure (psi), N is the engine speed (r/min), D is the valve disc diameter (in), θ is the seat angle (degrees), E is the Young’s modulus for the valve material (psi), B is the Rockwell Hardness number, C is the height of seat (in), t is the distance from the valve disc face to the seat top (in), and tm is the height of the valve disc cone (in). Details of the valve dimensions required are shown in Fig. 7.1. tm t C θ P D Fig. 7.1 Valve details [1] For a N/A, 1.8 litre, IDI, diesel engine inlet valve and seat (valve material V1 and seat insert material S3) P = 1885.5 psi; θ = 45 degrees; C = 0.087 in; N = 4800 r/min; E = 29 007 367.9 psi; t = 0.122 in; D = 1.417 in; B = 79.5 RA; tm = 0.481 in; µ = 0.4 (steel on steel, dry) or 0.05 (steel on steel, greasy). 114 Design Tools for Prediction of Valve Recession and Solving Valve Failure Problems Therefore, for a ‘dry’ valve/seat insert contact Fw = 3154 and for a ‘greasy’ valve/seat insert contact Fw = 394, both of which lie in the range likely to give an ‘unsatisfactory’ life. Engine test data (shown in Fig. 4.4), however, clearly indicate that wear falls within acceptable boundaries for this material combination. Giles [2] used the following expression, based on the equation developed by Archard [3] for predicting volume of material lost through adhesive wear V, to investigate ways in which valve recession could be reduced V= KWL CP (7.2) where K is the wear coefficient, W is the load on sliding surfaces, L is the sliding distance, P is the penetration hardness of the softer of the two contacting surfaces, and C is a constant depending on units of measurement. By admission, however, this was only intended to provide a generalized approach to minimizing wear rather than perform a quantitative analysis. It was clear that a model was required that would provide a quantitative analysis of valve recession. In order to achieve this it needed to encompass all wear mechanisms occurring at the valve/seat interface. To allow a range of inlet valve and seats to be considered, valve parameters, design parameters, and material properties also needed to be included. 7.2.2 Development of the model Development of the model followed the stages detailed in the flow chart shown in Fig. 7.2. It was decided to consider the two fundamental wear mechanisms identified as causing valve recession separately (frictional sliding at the valve/seat insert interface under the action of the combustion pressure, and impact of the valve on the seat insert on valve closure) as they occur as two definable events in the valve operating cycle. Approaches for modelling the wear mechanisms identified for each were appraised and parameters were then derived either from test results or directly from the valve and seat design and engine operating conditions. An allowance for effects such as misalignment and variation in lubrication was also built in at this stage. The two parts were then combined to form the final valve recession model. 7.2.2.1 Frictional sliding Abrasive, adhesive, and fretting wear were observed to have occurred as a result of the frictional sliding between the valve and seat under the action of the combustion pressure. It was, therefore, decided to use the equation developed by Archard [3], for deriving wear volume V in sliding situations, to model the wear caused by the frictional sliding 115 Automotive Engine Valve Recession V= kPc x h (7.3) where k is the wear coefficient, Pc is the contact force (N), x is the sliding distance (m) and h is the penetration hardness of the softer of the two contacting surfaces (N/m2). Equation (7.3) was originally developed for modelling adhesive wear, but it has also been used successfully to model both abrasive wear [4] and fretting wear [5]. Calculation of the parameters to be used in equation (7.3) is outlined below. Load The peak load normal to the direction of sliding at the valve/seat insert interface Pc was calculated using the peak combustion pressure pp and the valve head geometry, as shown in Figure 7.3 Pc = ppπ Rv 2 (1 + µ )sin θ v (7.4) where θv is the valve seating face angle (degrees) and µ is the coefficient of friction at the valve/seat interface. 116 EQUATIONS MECHANISMS EFFECT OF EXTERNAL INFLUENCES Fracture Toughness Deformation IMPACT (on valve closure) PARAMETERS W = KNe n or W = KNv n Surface Cracking Sub-Surface Cracking Closing Velocity Increases with misalignment Decreases with increasing recession Valve Mass MODEL Sliding Distance Increases with misalignment and contact width Wear Coefficient Decreases with lubrication Adhesion SLIDING (under action of combustion load) Abrasion k Pc δ N V= h Hardness Fretting Load Fig. 7.2 Development of the valve recession model Design Tools for Prediction of Valve Recession and Solving Valve Failure Problems CAUSES OF WEAR 117 Automotive Engine Valve Recession The load on the valve seat is initially zero then rises to Pc and falls back down to zero. For the purposes of calculating the sliding wear volume an average load P was assumed equal to half Pc. In the absence of other data, µ was estimated to be 0.1 for the valve/seat interface, which is a typical value for boundary lubricated steel surfaces. pp θv µ Pc µ Pc Pc Pc Rv Fig. 7.3 Valve geometry and loading Sliding distance Slip at the interface can be found either by measurement of scratches or by using finite element analysis. Data generated by Mathis et al. [6] using the latter were utilized in this model. In the absence of other data it was assumed that slip at the interface δ is proportional to combustion load Pc. The total sliding distance is calculated by multiplying the slip δ by the number of loading cycles N. Wear coefficient Rabinowicz [7] carried out a series of studies to systematically generate wear coefficients for sliding metals. These were derived from wear volumes using equation (7.3). Different states of lubrication and material ‘compatibility’ (‘the degree of intrinsic attraction of the atoms of the contacting metals for each other, as demonstrated by the solid solubility or liquid miscibility’ [7]) were investigated. The published data were used to construct a chart of wear coefficients, as shown in Fig. 7.4. Adhesive wear data are shown as well as data applicable to abrasive, fretting, and corrosive wear. The lubrication states and material compatibilities for valve/seat insert material pairings used in both engine tests and rig tests are shown in Table 7.1. The wear coefficients, also listed, were taken from Fig. 7.4. It should be noted that the wear coefficients given in Fig. 7.4 were derived using the form of Archard’s equation in which wear is proportional to sliding distance multiplied by load divided by three times the hardness. In this work the three was taken out and the wear coefficients adjusted accordingly. 118 Design Tools for Prediction of Valve Recession and Solving Valve Failure Problems Identical Metals Compatible Metals ADHESIVE WEAR Poor Lube Unlubed Poor Lube Unlubed Partly Compatible Metals Good Lube Good Lube Good Lube Excellent Lube Unlubed Poor Lube Good Lube Unlubed Non-metal on Metal or Non-Metal ABRASIVE WEAR High Abr. 3 - Body Concentr. - Body 2 -Body CORROSIVE WEAR Rampant Excellent Lube Poor Lube Unlubed Incompatible Metals Excellent Lube Low Abr. Concentr. Benign - EP Action Unlubed FRETTING 10-1 10-2 Lubed 10-3 10-4 Lubed 10-5 10-6 WEAR COEFFICIENT Fig. 7.4 Wear coefficients to be anticipated in various sliding situations [7] Hardness The penetration hardness of the softer of the two materials (valve and seat insert) was used. ‘Hardness’ of the valve and seat materials used during engine and rig tests are shown in Tables 6.2 and 6.6. Table 7.1 Lubrication states, material compatibilities, and wear coefficients for valve/seat insert material pairings used in both engine tests and rig tests Test apparatus Valve material Seat material Lubrication Material compatibility Wear coefficient Engine V1 S1 Poor Compatible 5×10−5 Engine V1 S3 Poor Compatible 5×10−5 Hyd. loading app. V1 S2 Unlubed Compatible 1×10−3 Hyd. loading app. V1 S3 Unlubed Compatible 1×10−3 Hyd. loading app. V1 S3 Good Compatible 1×10−6 Hyd. loading app. V1 Grey cast iron Unlubed Compatible 1×10−3 Hyd. loading app. V1 Ductile cast iron Unlubed Compatible 1×10−3 Hyd. loading app. V1 Free-cutting mild steel Unlubed Compatible 1×10−3 Hyd. loading app. V1 Oilite Poor Intermediate 5×10−5 Hyd. loading app. V1 Maraging steel Unlubed Compatible 1×10−3 119 Automotive Engine Valve Recession 7.2.2.2 Impact The observed deformation on the valve seating faces and surface cracking on the seat inserts are characteristic of processes leading to wear by single or multiple impact of particles [8] (see Section 6.2.2.1). Fricke and Allen [9] used a relationship of the same form as that used in erosion studies to model impact wear of poppet valves operating in hydropowered stoping mining equipment W = KNe n (7.5) where W is the wear mass (Kg), e is the impact energy per cycle (J) and K and n are empirically determined wear constants and e= 1 2 mv 2 where m is the mass of the valve added to the mass of the follower and half the valve spring mass (kg), and v is the valve velocity at impact (m/s). Wellinger and Breckel [10], in their repetitive impact studies, also found that wear loss could be described by W = KNv n (7.6) Fricke and Allen [9] justified the use of such a relationship for impact wear of valves, citing work by Hutchings et al. [11] in which it was shown that erosion can be satisfactorily modelled by the impact of large particles. In their work they used hard steel balls up to 9.5 mm in diameter. It was thus assumed that a relationship exists between impacts on a macroscale (greater than 1 mm) and impacts on a microscale (less than 1 mm), such as those found typically in erosive wear. The similarity of the wear features observed during testing to those attributed to erosive wear further supports this approach (see Section 6.2.2.1) as well as the fact that wear was found to be approximately proportional to the square of the closing velocity (see Section 6.2.2.7). It was, therefore, decided that an equation of this form would be appropriate to model the wear caused by the impact of the valve on the seat insert on valve closure. Although the valve closing velocities achieved using the hydraulic loading apparatus were not representative of those found in an engine, equation (7.5) was thought to be more suitable as the energies on valve closure were quite close to those found in an engine (see Section 5.5.3.2). Calculation of the parameters used in equation (7.5) is outlined below. Velocity Valve closing velocity was determined from the initial valve clearance ci using the velocity curves illustrated in Section 5.5.3.1. This was achieved by first fitting a line to the valve lift curve (fourth-order polynomial) to obtain lift as a function of time and then differentiating this equation to give velocity as a function of time. Velocity was then plotted against lift and a further line fit gave velocity as a function of lift. 120 Design Tools for Prediction of Valve Recession and Solving Valve Failure Problems As the valve recesses, the valve clearance at closure will decrease, thereby decreasing the valve closing velocity. This needs to be considered in the application of the model. Wear constants K and n Values of K and n for the seat materials used in these studies were derived using an iterative process to fit equation (7.5) to experimental data from tests run on the motorized cylinder head. Equation (7.5) was used to calculate wear volumes rather than wear mass to fit in with Archard’s sliding wear equation, equation (7.3), when combining the two to create the final model. These were then used to calculate recession values using equations relating wear volume to recession [12]. At each data point the velocity was recalculated to take account of the change in clearance due to recession. An iterative process was used to fit the recession values to those derived for tests run on the motorized cylinder head. The results of this process for tests run with V1 valves and S3 cast seat inserts, V1 valves and S2 sintered seat inserts, and V1 valves and grey cast iron, ductile cast iron, and oilite seats are shown in Figs 7.5, 7.6, and 7.7 respectively. Fig. 7.5 Modelling of valve recession for V1 valves run with three different closing velocities against S3 cast seat inserts on the motorized cylinder head (solid line – experimental data; broken line – model prediction) 121 Automotive Engine Valve Recession Fig. 7.6 Modelling of valve recession for V1 valves run with two different closing velocities against S2 sintered seat inserts on the motorized cylinder head (solid line – experimental data: broken line – model prediction) Fig. 7.7 Modelling of valve recession for V1 valves run against various seat materials with a closing velocity of 1600 mm/s on the motorized cylinder head (solid line – experimental data; broken line – model prediction) 122 Design Tools for Prediction of Valve Recession and Solving Valve Failure Problems As can be seen for seat insert materials S1 and S2, the values of K and n derived give good correlation over a range of velocities. This proves that the impact model is relevant. The values of K and n used to produce the model predictions shown in Figs 7.5–7.7 are listed in Table 7.2. Table 7.2 Values of impact wear constants K and n for valve/seat insert material pairings Valve Material Seat Material K n V1 S3 5.3×10−14 1 V1 S2 3.5×10−14 0.3 V1 Grey cast iron 4.8×10−14 0.77 V1 Ductile cast iron 5×10−14 1.2 V1 Free-cutting mild steel * * V1 Oilite 5×10−14 0.2 V1 Maraging steel * * * Test incomplete (see Section 6.3.2) It was hoped that a relationship between the impact wear factor n and seat fracture toughness would emerge. However, the low recession exhibited by the ductile cast iron seats provided an anomaly. In general it could be said that n reduces with increasing seat hardness and seat fracture toughness. Without being more exact it would be impossible to build in such a relationship to the model. This means that testing on the motorized cylinder head is required in order to model the recession likely to occur with a particular material combination. 7.2.2.3 Final model Putting together the equations for sliding and impact wear [(7.3) and (7.5)] gives the final wear model as V= kPNδ + KNen h (7.7) In order to incorporate the change in pressure at the interface and any other effects likely to lead to a reduction in the wear rate with time, a term consisting of the ratio of the initial valve/seat contact area Ai to the contact area after N cycles, A, to the power of a constant j was included. The term j was determined empirically using bench and engine test data.  kPN δ  A  + KNe n   i  V=   h  A  j (7.8) 123 Automotive Engine Valve Recession Equation (7.8) gives a wear volume which is then converted to a recession value using equations derived previously using the valve and seat geometries [12]. Equation (7.9) gives the geometrical relationship between r and V for the case where valve and seat angles are equal   V 2 + − w w   sin θ s i i r =  π R cos θ sin θ i s s   (7.9) where Ri is the initial seat insert radius, θs is the seat insert seating face angle, and wi is the initial seat insert seating face width (as measured). Ri can be calculated using wi and the radius and seating face width as specified for the seat insert (Rd and wd). Valve closing velocity is related to valve recession, which needs to be considered in the application of the model. Closing velocity decreases as recession increases. 7.2.3 Implementation of the model A flow chart outlining the use of the model is shown in Fig. 7.8. The wear volume is determined incrementally. The initial valve closing velocity and contact area are used to calculate the volume of material removed over the first N cycles. This is then converted to recession, and new values for the clearance (and hence closing velocity) and contact area are determined. The calculation is repeated until the total number of iterated cycles equals the required run duration. Recalculate v = f(c) ci m N K v = f(c) e 1 = 2 mv 2 IMPACT WEAR VOLUME = KNe n n k P h δ Recalculate c = ci - r + Wear Volume For N Cycles x Ai A j Add to Total Volume Calculate Recession = f(V) SLIDING WEAR VOLUME k P δN = h N Fig. 7.8 Valve recession model application flow chart (Reprinted with permission from SAE paper 2001-01-1987 © 2001 Society of Automotive Engineers, Inc.) 124 Design Tools for Prediction of Valve Recession and Solving Valve Failure Problems In order to provide a tool for running the valve wear model, an iterative software program called RECESS was developed. Within the program the recession calculation is carried out in three stages. In the first and second, the sliding and impact wear volumes, Vs and Vi, are calculated for N cycles. In the final stage, Vs and Vi are added together to calculate the wear volume for the set of N cycles. This is then used to calculate the recession value r for the total number of cycles. The three stages are explained in more detail below. Sliding wear calculation At this stage, as shown in Fig. 7.9, the following are entered: ● valve and initial seat insert geometry; ● seat insert material properties; ● maximum combustion pressure, valve misalignment (if any), coefficient of friction at the valve/seat interface, wear coefficient for the material combination and the number of cycles. The sliding wear volume is calculated for N cycles. WEAR VOLUME DUE TO SLIDING Initial SI Seating Face Width ( w i ) = Initial SI Radius ( R i ) = Max. Combustion Pressure ( p p ) = Valve Head Radius ( R v ) = Seating Face Angle ( θ v )= Coeff. of Friction at Interface ( µ ) = Max. Contact Force at Interface ( P c ) = Avg. Contact Force at Interface ( P bar ) = Valve Misalignment Relative to SI = Slip at Interface ( δ ) = Number of Cycles ( N ) = Sliding Distance ( x ) = Wear Coefficient ( k ) = Hardness of SI Material ( h ) = Wear Volume Due to Sliding ( V s ) = 2.00E-03 0.01685 2.00E+07 1.80E-02 45 0.1 26172.61949 13086.30975 0 8.06E-06 1440000 11.6064 1.00E-05 4.90E+02 1.05324E-10 m m Pa m Degrees N N m m m 2 Hv (Kg/mm ) 3 m Fig. 7.9 Sliding wear calculation Impact wear calculation At this stage, as shown in Fig. 7.10, the following are entered. ● Valve clearance. Note that, for the 1.8 litre diesel inlet cam to be used in validating the model (see Section 7.2.4), valve clearances between 0 and 0.4 mm give a constant valve closing velocity and the first column should be used. For clearances above 0.4 mm the velocity varies with clearance. The clearance should then be entered in the second column where it is used to calculate a closing velocity using a curve fitted to clearance/closing velocity data for the cam. 125 Automotive Engine Valve Recession ● Mass of the valve, follower, and half the spring. ● Wear constants K and n (derived from experimental data). Note that r at the top of the sheet should be left as zero for the first set of N cycles. After the first set of N cycles a value for r is calculated (at the Recession calculation stage, see below) which should then be entered for the subsequent set of N cycles in order to recalculate valve clearance and the closing velocity (not necessary for initial valve clearances below 0.4 mm). The number of cycles N is taken directly from the Sliding Wear stage. The impact wear volume is calculated for N cycles. WEAR VOLUME DUE TO IMPACT Recession (r ) = Valve Clearance ( c ) = Valve Misalignment Relative to SI = Total Clearance = Velocity on Impact ( v ) = Mass (m ) = Energy on impact ( e ) = Wear Constant K = Wear Constant n = Number of Cycles ( N ) = Wear Volume Due to Impact ( V i ) = Clearance: 0 to 0.0004 0.00E+00 0.000235 0 0.000235 0.288274 0.18295 0.007601746 5.30E-14 1 1440000 5.80165E-10 Clearance: 0.0004 + 0 0.0004 0 0.0004 0.800492234 0.18295 0.058616065 5.30E-14 1 1440000 4.47358E-09 m m m m m/s Kg J m3 Fig. 7.10 Impact wear calculation Recession calculation At this stage, as shown in Fig. 7.11, the value of j to be used is entered as well as the initial value of (Ai/A). The sliding wear volume Vs for the set of N cycles, and the impact wear volume Vi for the set of N cycles, are then added together to calculate the wear volume for the set of N cycles. This is multiplied by the (Ai/A)j term. This number should then be entered into the table on the sheet. The values in this table are added together to calculate the total wear volume Vt which is used to calculate the valve recession r for the total number of cycles, using the equation relating Vt to r for an equal valve and seat angle of 45 degrees [12]. The recession r calculated should then be re-entered at the Impact Wear stage in order to recalculate the valve clearance, c, and the valve closing velocity, v, for the next set of N cycles (not necessary for initial clearances below 0.4 mm), and the new value of (Ai/A) should be entered at the Recession stage. 126 Design Tools for Prediction of Valve Recession and Solving Valve Failure Problems RECESSION V s (Wear Volume Due to Sliding) = 1.05324E-10 m V i (Wear Volume Due to Impact) = 5.80165E-10 m j (A i /A ) = 3 No of Cycles Total Wear Volume 3 0.762 3 Wear Volume for N Cycles = j Wear Volume for N Cycles * ( A i /A ) 6.8549E-10 m 5.22343E-10 Total Wear Volume = Recession ( r ) = Seat Insert Seating Face Width ( w ) = SI Radius ( R ) = 5.97E-09 3.93035E-05 0.002055584 0.016889304 Seating Face Area ( A ) = Initial Seating Face Area ( A i ) = 0.000218136 m 2 0.000211743 m A i /A = j = j (A i /A ) = 0.9706955 10 0.742728627 m m m m 3 2 (A i /A ) j N (1) 6.85E-10 1 N (2) 6.62E-10 0.966 N (3) N (4) 6.40E-10 6.20E-10 0.934 0.904 N (5) N (6) N (7) N (8) 6.00E-10 5.83E-10 5.66E-10 5.50E-10 0.876 0.851 0.826 0.803 N (9) N (10) 5.36E-10 5.22E-10 0.782 0.762 Fig. 7.11 Recession calculation 7.2.4 Model validation In order to validate the model valve, recession predictions were calculated for engine tests and tests run on the hydraulic loading apparatus. 7.2.4.1 Engine tests When calculating recession predictions for engine test results, the model was used as described above. Equal valve and seat angles of 45 degrees and the same initial seating face widths were assumed. It was also assumed that the initial clearances were set within the constant velocity region (0–0.4 mm, see Fig. 5.14). Initial conditions used for the test simulated are shown in Table 7.3. Table 7.3 Initial values used in calculating model predictions Valve material Seat material Test type w (mm) k P (N) h (Hv) v (mm/s) K n V1 S1 Engine 2 5×10−5 8053 490 288 3.5×10−14 0.3 V1 S3 Engine 2 5×10−5 8053 490 288 5.3×10−14 1 V1 S3 Hyd. app. 0.719 1×10−3 8053 490 18 5.3×10−14 1 8053 490 18 3.5×10−14 0.3 V1 S2 Hyd. app. 0.760 1×10−3 V1 Grey C.I. Hyd. app. 0.653 1×10v3 8053 193 18 4.8×10−14 0.77 V1 Ductile C.I. Hyd. app. 0.796 1×10−3 8053 319 18 5×10−14 1.2 V1 Oilite Hyd. app. 0.504 5×10−5 8053 89.4 18 5×10−14 0.2 0.726 1×10−6 18 5.3×10−14 1 V1 S3 (lubed) Hyd. app. 8053 490 127 Automotive Engine Valve Recession Figure 7.12 shows the model prediction for an engine test run using S3 cast and S1 sintered seat inserts. As can be seen, the model produces a good prediction of valve recession. In order to determine which parameters were having the largest influence on the wear, the percentage split of the total wear between impact and sliding for the engine test predictions was calculated (see Fig. 7.13). Clearly the contribution of impact wear to the total was higher than that of sliding. The parameters having the largest influence on the wear are, therefore, valve closing velocity, valve mass, and the impact wear constants K and n, which are related to the resistance to impact wear of a seat material. Fig. 7.12 Model prediction (broken line) versus engine test data. (Reprinted with permission from SAE paper 2001-01-1987 © 2001 Society of Automotive Engineers, Inc.) 128 Design Tools for Prediction of Valve Recession and Solving Valve Failure Problems Fig. 7.13 Percentage of total wear caused by impact and sliding for a cast seat insert (S3) engine test 7.2.4.2 Bench tests Figures 7.14 and 7.15 illustrate recession predictions using the model for tests run on the hydraulic loading apparatus. Initial conditions used for each test simulated are shown in Table 7.3. As can be seen, the model produces a good approximation of valve recession. Figure 7.16 shows the contributions of impact and sliding wear to the total wear for each case. In general, the effect of impact is reduced, compared to the engine test prediction, as a result of the lower valve closing velocities experienced in the hydraulic loading apparatus. The two exceptions are the lubricated S3 cast seat insert, where the lubricant reduced the sliding wear to about 0.1 per cent of the total, and the oilite seat, which has a very low resistance to impact. It has been shown that the valve recession model produces good predictions of valve recession for both engine tests and bench tests. It could clearly be used, therefore, to give a quick assessment of the valve recession to be expected with a particular material pair under a particular set of engine/test rig operating conditions. This will help speed up the process undertaken in selecting material combinations or in choosing engine operating parameters to give the least recession with a particular combination. A quick examination of the model immediately reveals how wear can be reduced by varying engine operating parameters and material properties. For example, reducing valve closing velocity, valve mass, and valve seating face angle, or increasing the valve head stiffness and seat material hardness will reduce valve recession. 129 Automotive Engine Valve Recession Fig. 7.14 Model prediction (broken lines) versus hydraulic loading apparatus data for a cast seat insert (S3), sintered seat insert (S2), and a lubricated cast seat insert (S3). (Reprinted with permission from SAE paper 2001-01-1987 © 2001 Society of Automotive Engineers, Inc.) Fig. 7.15 Model prediction (broken line) versus hydraulic loading apparatus data for grey cast iron, ductile cast iron, and oilite seats. (Reprinted with permission from SAE paper 2001-01-1987 © 2001 Society of Automotive Engineers, Inc.) 130 Design Tools for Prediction of Valve Recession and Solving Valve Failure Problems Fig. 7.16 Percentage of total wear caused by impact and sliding for each bench test modelled By studying the individual contributions of impact and sliding wear it is possible to focus on the particular parameters that need to be altered in order to produce the largest feasible reduction in the total wear for a particular material combination. In order to improve the model it would be beneficial to be able to relate the impact wear constants K and n to material properties. Currently they have to be determined from test data. It is likely that as the tests progress the hardness at the seating face of the seats increases. It would also be desirable to include such data in the model. When selecting wear coefficients using Fig. 7.4 there was a large range of possible values for each material combination/lubrication state being considered (at least one order of magnitude). In order to further improve the model it would be beneficial to obtain more accurate wear coefficient data referring to the actual material pair used. 131 Automotive Engine Valve Recession 7.3 Reducing valve recession by design The flow charts included as Figs 7.17 and 7.18 encapsulate the review of literature, analysis of failed specimens, bench test work, and modelling to create flow charts. The first of these was produced to help reduce the likelihood of valve and seat wear occurring during the design process of a new engine or in the modification of an existing engine (see Fig. 7.17). It was developed directly from the valve recession model and looks at ways in which wear can be reduced by varying the model parameters. The chart looks at sliding and impact wear separately, as they are treated as such in the model. Changing engine design variables may have an adverse effect on engine performance. For example, a reduction in the peak cylinder pressure could reduce the wear, but could also reduce the engine thermal efficiency if a decreased compression ratio is used. Likewise, stiffening the valve head to reduce the relative movement at the seating face by adding material or altering the valve seating face angle could adversely affect air flow characteristics, and if the valve weight increases significantly, it may affect valvetrain dynamics. The introduction of lubrication at the valve/seat interface is clearly unacceptable as this goes against the original stated intention to reduce oil in the air stream of diesel engines. A possible alternative could be the use of solid lubricants incorporated in sintered seat insert materials. However, the problems involved with the use of such materials have been highlighted and some work is needed to improve their performance, certainly with respect to resistance to impact. Given the problems outlined above and the constant demand for improved performance and efficiency and reduced emissions, reducing the likelihood of valve recession occurring by design may prove impossible. However, wear reduction through the use of new materials, materials selection, or the use of wear-resistant coatings provides an acceptable alternative. As already discussed, the two test rigs developed are suitable for use in the testing of new valve, seat, and valvetrain designs and material ranking. 7.4 Solving valve/seat failure problems The second flow chart (see Fig. 7.18) was produced to assist in solving valve/seat failure problems. It was developed using experience gained during the progression of the work on failure analysis, experimental work and modelling, and using information obtained in the review of literature. Three stages to solving a valve/seat failure problem were identified. i) Determination of valve/seat failure During engine testing the criterion for determining valve/seat failure is pressure loss in a cylinder. ii) Analysis of problem This involves establishing what has happened and why it has happened. Techniques that may be used in this process are: 132 Design Tools for Prediction of Valve Recession and Solving Valve Failure Problems ● profilometry; ● ovality (of seating faces); ● visual inspection; ● optical microscopy. The profilometry and ovality measurements will show how much wear has occurred and whether it is even. Visual inspection should indicate whether deposit build-up played any part in the failure, and optical microscopy will reveal the wear mechanisms that have occurred which should indicate what has caused the wear to occur (impact or sliding). iii) Solving the problem This involves establishing ways in which the causes of the failures identified in stage (ii) can be eliminated. The flow chart goes through the stages detailed above, giving likely reasons why a valve or seat may have failed and possible causes of these failures and then goes on to offer solutions to the problems. This information should help focus any engine testing required to solve a problem, thereby saving time, effort, and cost. This chart is intended to be a working document that should be added to as work is carried out or as different problems arise. The flow chart can be used to solve problems unique to one cylinder or common to several/all of the cylinders in a head. 133 134 Reduce Impact Energy on Valve Closure REDUCING EFFECT OF IMPACT WEAR Decrease Wear Constant, n Reduce Valve Closing Velocity Reduce Valve Mass Increase Valve/Seat Fracture Toughness REDUCTION OF VALVE RECESSION Reduce Sliding Distance Increase Valve Head Stiffness Increase Valve Disc Thickness Reduce Valve Reducing Valve Face Angle Angle Seating Face REDUCING EFFECT OF SLIDING WEAR Reduce Contact Load Decrease Wear Coefficient Increase Hardness of Valve/Seat Reduce Valve Head Diameter Increase Valve/ Seat Lubrication Decrease Valve/Seat Material Compatibility Use Wear Resistant Coatings Fig. 7.17 Reducing valve recession by design. (Reprinted with permission from SAE paper 2001-01-1987 © 2001 Society of Automotive Engineers, Inc.) Automotive Engine Valve Recession Reduce Peak Combustion Pressure ANALYSING THE PROBLEM FAILURE What Has Happened? What Caused it to Happen? Excessive Valve Wear Due to Impact Fracture Toughness too Low Select New Material Due to Guttering Poor Cam Design Redesign Cam Profile Clearance too High Reduce Clearance Dynamic Problem with Camshaft Investigate Camshaft Dynamics and Correct Any Problems Due to Sliding Excessive Seat Wear Valve Closing Velocity too High Due to Impact Due to Sliding Poor Valve Design IDENTIFY FAILURE Seating Angle too High Head Stiffness too Low Valve Fatigue Failure Use Wear Resistant Coating on Seating Face Flaking of Deposit/Varnish (formed from lubricant) Thermal Softening Due to Excessive Temperature Uneven Seat Wear Redesign Valve Head Reduce Lubricant Supply to Valve/Seat Interface Seat Area not Hardened Adequately Check/Improve Induction Hardening Process Inadequate/Non-Uniform Cooling Redesign Head Cooling Channels Deposit Build-Up Reducing Heat Transfer Valve Misalignment Relative to Seat Incorrect Fit on Seat Insert Causing High Hoop Stresses Poor Manufacturing Tolerances Improve Tolerances on Head Machining Reduce Interference Fit on Seat Insert 135 Fig. 7.18 Solving valve/seat failure problems. (Reprinted with permission from SAE paper number 2001-01-1987 © 2001 Society of Automotive Engineers, Inc) Design Tools for Prediction of Valve Recession and Solving Valve Failure Problems Hardness too Low Material not Suitable SOLVING THE PROBLEM Automotive Engine Valve Recession 7.5 References 1. Pope, J. (1967) Techniques used in achieving a high specific airflow for highoutput medium-speed diesel engines, Trans. ASME J. Engng Power, 89, 265–275. 2. Giles, W. (1971) Valve problems with lead free gasoline, SAE Paper 710368. 3. Archard, J.F. (1953) Contact and rubbing of flat surfaces, J. Appl. Physics, 24, 981–988. 4. Suh, N.P. and Sridharan, P. (1975) Relationship between the coefficient of friction and the wear rate of materials, Wear, 34, 291–299. 5. Stower, I.F. and Rabinowicz, E. (1973) The mechanism of fretting wear, Trans. ASME, J. Lubrication Technol., 95, 65–70. 6. Mathis, R.J., Burrahm, R.W., Ariga, S., and Brown, R.D. (1989) Gas engine durability improvement, Paper GRI-90/0049, Gas Research Institute. 7. Rabinowicz, E. (1981) The wear coefficient – magnitude, scatter, uses, ASME Paper 80-C2/LUB-4, Trans. ASME J. Lubrication Technol., 103, 188–194. 8. Zum-Gahr, K.H. (1987) Microstructure and wear of materials, Tribology Series No. 10, Elsevier, Amsterdam. 9. Fricke, R.W. and Allen, C. (1993) Repetitive impact-wear of steels, Wear, 163, 837-847. 10. Wellinger, K. and Breckel, H. (1969) Kenngrossen und Verscheiss Beim Stoss Metallischer Werkstoffe, Wear, 13, 257–281, in German. 11. Hutchings, I.M., Winter, R.E., and Field, J.E. (1976) Solid particle erosion of metals: the removal of surface material by spherical objects, Proc. R. Soc. Lond. Ser A., 348, 379–392. 12. Lewis, R. (2000) Wear of diesel engine inlet valves and seats, PhD Thesis, University of Sheffield, UK. 136 Index Abrasion 22 Acceleration 10 Adhesion 22 Cams 9 Cam followers 8 Camshaft, overhead 8 Causes of valve recession 22 et seq. Closing velocity 10, 70, 71, 96, 97, 120 effect on wear 96 Collet 8, 9 Combustion: load 55, 73, 116, 118 effect on wear 97 particles 47 Compression ignition (CI) engines 7 Corrosion 22 hot 3 Deposits 34, 35, 46, 50 Design Tools 113 et seq. fault tree 135 valve recession model 113 et seq. Development of recession model 115 et seq. Dynamics 9, 69 Engine cycle, four-stroke 7 Engine operating parameters 69 et seq. Engine recession data 92 Erosion 87, 120 Exhaust gas recirculation 45 Fault tree 135 Formation of wear scars 88 Four-stroke engine cycle 7 Fracture toughness 91, 104, 105, 109, 110 Fretting 22 Guttering 13, 21, 29 Hardness 81, 82, 105, 119 Heat transfer 13 Hot corrosion 3 Impact: load 55 wear 87, 108, 120, 125, 128, 129, 131 Implementation of recession model 124 et seq. Lacquer formation 45–47 Lead replacement petrol 2 Load: combustion 55, 73, 96, 116, 118 impact 55 Lubrication 34, 63, 93 effect on wear 93 Materials: composition 18 cost 109, 110 properties: fracture toughness 91, 104, 105, 109, 110 hardness 81, 82, 105, 119 machinability 109, 110 thermal conductivity 104, 105 seat insert 19, 103 selection 92, 104, 110 valve 17 Misalignment 25, 51, 55, 72, 94 effect on wear 25, 94 Operating stresses 12 Operating systems 8 Overhead camshaft 8 Overhead valves 8 Poppet valve 1, 15 Push rods 8 Recession: causes of 22 et seq. model 113 et seq. design software - RECESS 125 et seq. development of 115 et seq. implementation of 124 et seq. validation of 127 et seq. reduction of 28 by design 132 Rocker arms 8 Rotation 9, 34, 35, 61, 63, 100, 101 effect on wear 34, 100, 101 Seat insert materials 19, 103 composition 18 Shims 9 Spark ignition (SI) engines 7 137 Automotive Engine Valve Recession Temperature 13, 14, 31, 102 effect on wear 31, 102 Thermal conductivity 104, 105 Torching 13, 21, 29 Validation of recession model 127 et seq. Valve: acceleration 10 bounce 11–13 closing velocity 10, 70, 71, 97, 120 deposits 34, 35, 46, 50 design 15, 16 dynamics 9, 69 energy 72 guides 9 lift 10, 70, 71 lubrication 34, 63, 93 materials 17 misalignment 25, 51, 55, 72, 94 motion 9 operating stresses 12 operating systems 8 operation 7, 8 overhead 8 poppet 1, 15 recession: causes of 22 et seq. model 113 et seq. reduction of 28, 132 138 rotation 9, 34, 35, 61, 63, 100, 101 spring 8, 11 temperature 13, 14, 31, 102 Valve/seat failure problems 132 et seq. Velocity, closing 10, 70, 71, 96, 97, 120 Wear: abrasive 22 adhesive 22 characterization 26 coefficient 118, 119 debris, appearance 100 erosive 87, 120 fretting 22 impact 87, 108, 120, 125, 128, 129, 131 mechanisms 81 scars, formation of 88 sliding 115, 125, 128, 129, 130, 131 test methods 56 testing apparatus 55 et seq. Wear testing: block on ring 56 crossed cylinder 56 extant wear test rigs 56 thrust washer 56 University of Sheffield wear test rigs 59 et seq. hydraulic loading apparatus 60, 74 motorized cylinder head 67 worn surfaces 84

With years of expertise in powertrain aggregates, AVTEC is now setting benchmarks in the manufacture of high-precision engine and transmission components for the Automotive and Off-highway Industry.

AVTEC's state-of-the-art manufacturing units at Pithampur, Hosur & Chennai, are ISO 16949 certified and equipped with the best of international machines like Liebherr, Gleasen, Hyundai, Mori Sekie, Reishauer and Escofier. Its in-house design and development centre and capability to handle the entire supply chain management, enable it to provide the complete solution from art to part.

The range of Components manufactured by AVTEC includes:

- 5 'C's of Engine: Cylinder head, Cylinder blocks, Crank shaft, Connecting rod and Camshaft

- Complex Casting Parts: Exhaust & Inlet manifold Flywheel, Rocker cover

- Transmission components for Automotive & Off-highway

- High precision Gears & Shafts

- Transfer case

- Planetary gear drive

- Double synchro pack

Having already won the trust of leading global and domestic OEM's like Daimler, GM, Ford, Volvo Eicher, M&M, Tata, Jaguar, Allison, Caterpillar, Eaton, Telcon, L&T Komatsu and many more, is the testimony to AVTEC’s commitment towards quality.

Automatic pump control methods are described as a pump system that automatically delivers water when a tap is opened, and automatically stops the pump after all taps are closed.

A variety of pump control methods have been used in the past.

The Cycle Stop Valve replaces:

Variable Frequency Drives (VFD's) or "Constant Pressure Pumps"

Large pressure tanks and water towers

Flow switches and shuttle valves

Standard Pump Control Valves

Other so called Constant Pressure Valves

Cycle Stop Valves

The function of a Cycle Stop Valve (CSV) is to:

Provide variable flow and constant pressure control superior to VFD systems

Replace large pressure tanks and water towers

Provide minimum flow required to cool the pump and/or motor

Provide minimum flow to replenish the pressure tank when needed

Eliminate transient pressure waves and water hammer, stop line breaks

When selecting a Cycle Stop Valve, certain information must be known:

System pressure required

System flow required

Maximum output pressure of the pump

These devices are known by several names. Constant Pressure Valve (or CPV), Cycle Stop Valve (or CSV), are the names most commonly used to describe a valve that mechanically controls the output flow from a pump to match the usage. These valves have no electronics. The valve mechanically senses down stream pressure, and a pilot valve or spring mechanically controls the valves position. When pressure decreases below the set point, the valve moves toward the open position. When the pressure increases above the set point, the valve moves towards the closed position. By varying the position of the valve to maintain a constant pressure, such as 50 PSI, the output of the pump exactly matches the amount of water being used. In this way, there is no excess water produced, so large pressure tanks are no longer needed to minimize cycling.

A standard pressure switch and pressure tank is used for starting and stopping the pump. The CSV is installed before the pressure tank and switch, and is usually adjusted to the middle of the pressure switch setting. For example, a CSV set at 50 PSI is used with a 40/60 pressure switch, or a CSV set at 55 PSI is used with a 50/60 pressure switch. Small or large pressure tanks can be used with these devices, and the size of the tank determines the exact pressure settings.

When a tap is opened, the compressed air in the pressure tank, forces water from the tank to supply the usage. The pressure drops from 60 to 40 PSI as the tank is emptied. This utilizes all the water stored in the pressure tank, and keeps the pump from having to start for small, intermittent uses of water. At 40 PSI, the pressure switch starts the pump. With the CSV set at 50 PSI, the pressure quickly rises to 50 PSI. The pump will continue to run, and the pressure will remain constant, as long a small amount of water (usually 1 to 5 GPM) continues to be used. This keeps the pump from cycling on and off during long showers, small irrigation zones, and low heat pump discharge rates.

When all water outlets have been closed, the CSV also closes, and a small amount of flow (1 to 5 GPM) is bypassed from the inlet to the outlet of the CSV. This bypass rate maintains proper cooling for the pump/motor while slowly filling the remainder of the pressure tank, and the pressure switch shuts off the pump at 60 PSI. The larger the pressure tank, the less number of times the pump must start for intermittent uses of water such as ice makers, rinsing toothbrushes, or flushing a toilet. The smaller the pressure tank, or the more narrow the pressure switch differential, the more often the pump will need to start but, the sooner constant pressure is achieved.

The CSV creates a mechanical soft start and soft stop, which eliminates water hammer. An electronic soft starter can be used with a CSV, but is rarely needed as the CSV minimizes the number of on/off cycles.

When the points in the pressure switch open, no voltage is maintained on the system while the pump is off. The CSV does not use any power itself when the motor is running, or when the motor is off. The power consumption of the pump/motor naturally decreases as flow is decreased with the CSV. A CSV controlled pump uses the least amount of energy per gallon when the pump is delivering maximum flow.

The pump is sized to the maximum GPM or peak requirements of the system. When using flow less than maximum pump output, the CSV reduces the output of the pump accordingly. This keeps the pump running continuously, when flow rates required are less than maximum pump output. Very small flows or leaks, (less than 1 to 5 GPM) are fed by the pressure tank, as the pump slowly cycles on and off, depending on the size of the pressure tank, and the pressure bandwidth.

Maintaining 50 PSI constant for a shower or sprinklers, can be noticeably different than an average 50 PSI, as when a pump is cycling on and off between 40 and 60 PSI. A constant 50 PSI will deliver a consistent spray pattern for sprinklers, compared to when the pressure is continually changing between 40 and 60 PSI.

Large water systems that supply multiple houses, communities, and cities, can also use CSV controls. Varying the pump flow to match the usage eliminates the need for water towers, large hydro-pneumatic tanks, or multiple pressure tanks.

There are other manufacturers of Constant Pressure Valves or CPV's, and they use different types of controls and bypasses, and have different pressure tank size requirements than a CSV.

While the basic principle is the same, most brands of CPV's have external or drilled hole by-passes. These type bypasses have many disadvantages compared to the non-closing type by-pass of a Cycle Stop Valve. Small debris in the water can clog a hole at any time. Also water spewing through a small hole at 200 feet per second causes minerals to precipitate out of solution, and forms scale build up. This will clog a hole the same way holes in your showerhead clog up. Also some bacteria love areas of high velocity and will also clog small holes. Either way, what looks like barnacles on a boat hull, quickly build up and clog the small hole. This small hole is responsible for the flow needed to cool the pump and motor. When this hole clogs up, the pump is destroyed in only a few minutes.

To try and prevent the hole from clogging, a much larger hole is drilled. The size of this hole is also very important. It needs to be large enough to properly cool the pump and motor but, too large and the pump will still cycle at low flow. This means the pump can still be cycled to death and the pressure is not constant. Even a larger hole will still clog it just takes it a little longer.

The Cycle Stop Valve does not have a hole to clog. It has a non-closing seat, with two half moon notches, that come together to create a hole when the valve closes. This allows the use a very small 1 GPM bypass. When the valve opens, the two half moon notches split, and any debris, scale, or buildup breaks loose and flushes away. This prevents the valve from ever clogging, while maintaining a minimum of 1 to 5 GPM. Now the pump cannot cycle, even when flow as low as 1 GPM is being used. The pump also has the required minimum of 1 to 5 GPM to remain cool, without fear of clogging a small drilled bypass hole.

The non-closing bypass of the Cycle Stop Valve makes it the only control that reacts fast enough, to have wave canceling technology, that eliminates water hammer and line breaks.

BASIC PUMPING STANDARD OPERATING GUIDELINE DATE APPROVED: March 2008 DATE REVISED: June 2014 I. Scope This standard establishes a guideline for pumping a fire apparatus. II. Definitions 1. Appliance – A device, other than a hand held nozzle, where the direction of water flow is interrupted or changed. 2. Bleeder Valve - Valve on a gate that allows air from an incoming supply line to be bled off before allowing the water into the pump. 3. Compound or Vacuum Gauge – A gauge capable of measuring positive or negative pressures. This is the gauge that measures the intake pressure on a pump. 4. Cavitation – A condition that is created by water vapor bubbles (air) in the pump. 5. Centrifugal Pump – A non-positive displacement pump where water is introduced at the center of a revolving impeller, and moved outward. Can not pump air. 6. Discharge - Valve used to move water from the pump to the hose line. 7. Discharge Gauge - Shows the operator the pressure at each of the discharge valves being used. 8. Drain - Valve used to drain water from piping and pumps. 9. Engine Pressure - The actual pressure at the pump panel. 10. Friction Loss - The part of the total pressure lost due to turbulence of water moving against the interior surfaces of pipes, hose, and appliances. 11. G.P.M. – Gallons per minute. 12. Gutter Line – A hand line used to flow water so the pump does not overheat. 13. Intake - Valve used to allow water to enter the pump. 14. Master Gauge - Shows the highest pressure being pumped. 15. Master Stream - Any fire stream that is flowing over 350 gpm. 16. Nozzle Pressure – Pressure at which water is being discharged from the nozzle. 17. Pressure – A measure of the energy contained in water and is stated in pounds per square inch (psi). 18. Primer – A small positive displacement pump that allows for the air to be displaced from the pump and suction hose. This allows the pump to receive water from a static water source. 19. Pressure Governor-Pressure control device that controls engine speed. Designed to eliminate a hazardous condition resulting from excessive pressures. 20. Pressure Relief Valve- Automatic device designed to release excess pressure from a pump while multiple lines are flowing. 21. Pump Shift Override- Allows the operator to bypass the electric pump shift and still engage the pump manually. 22. Residual Pressure - Pressure left over in a water system after water is flowing. 23. RPM Gauge - Revolutions per minute of the motor. 24. Static Pressure - Water pressure available in a system prior to water flowing. 25. Tank to Pump Valve –Valve that allows water from the tank into the pump. 26. Tank Fill Valve – Valve that allows the operator to fill the booster tank from water coming in to the pump. Can also be used to recirculate water, to cool the pump. 27. Water Hammer – The concussion effect of a moving stream of water, when its flow is suddenly stopped. 28. Water Temperature Gauge - Allows the operator to monitor the water temperature of the motor. III. Standards and Measurements One gallon of fresh water weighs 8.33 pounds (use 8.3 in formulas.) Atmospheric pressure at sea level is 14.7 pounds.. 50 foot section of 1 3/4 inch hose contains 6.24 gallons. 50 foot section of 2 1/2 inch hose contains 12.75 gallons. 50 foot section of 3 inch hose contains 18.3 gallons. 100 foot section of 5 inch hose contains 102 gallons. ( Approx. 950 lbs. ) 100 foot section of 5 inch hose uncharged weighs approx. 103 lbs. IV. Placing Pump in Gear A. Automatic Transmission l. Bring apparatus to full stop. Come to idle speed. 2. Shift transmission to neutral. Set the parking brake. 3. Operate pump shift device. 4. Shift road transmission into proper gear. This is usually drive. 5. Check the indicator lights to see if pump is in gear, check speedometer, and listen as pump goes in gear. 6. Depress accelerator to ensure shift is complete. V. Operating From the Booster Tank and Pressurized Water Source l. Set Wheel chocks. 2. Check "OK to pump light". 3. Open tank to pump valve. 4. Set throttle to 100 to 1200 rpm’s 5. Engage primer (If needed) a. Approximately 30 seconds for 1250 gpm pumps or less b. Approximately 45 seconds for 1500 + gpm pumps c. Add 15 seconds for front or rear intakes d. Engage primer until steady stream of water is flowing from the primer discharge hose. e. Look for pressure reading on master gauge and vacuum on the compound gauge. 6. Be sure hose is clear of hose bed and hose crew is ready for water. 7. Slowly open appropriate discharge. 8. Increase the throttle control to desired pressure. 9. Set pressure control devices. 10. Connect supply line to intake valve. 11. Open bleeder (if available) to purge air and leave open until steady stream of water flows from the opening. The following procedures need to be done together to stop from losing pressure on the lines being supplied 12. Open intake valve slowly. Close tank to pump valve slowly (This needs to be done simultaneously when possible ) 13. Adjust throttle to maintain desired pressure. 14. Open "tank fill" valve to refill tank. ** Partially open tank fill valve to recirculate water when no water is flowing, or use a gutter line ** Check all gauge readings VI. Shutting Down Procedures 1. Reduce throttle control to idle. 2. Close discharge valves. 3. Make sure tank is full of water. 4. Close intake valves. 5. Place transmission in neutral. 6. Wait for engine speedometer to go to zero. 7. Operate pump shift device. VII. Friction Loss Formula Method: FL= CQ2 L FL= Friction loss in psi C= Coefficient – from a predetermined chart Q= Quantity – GPM divided by 100 L= Length – length of hose divided by 100 Pump Discharge Pressure= Nozzle Pressure+ Friction Loss Coefficients: 1 ¾”- 15.5 2 ½”- 2 3”- .8 5”- .08 Nozzle Pressures: All standard Fog: 100 psi Smooth bore ( handline): 50 psi Smooth bore (masterstream): 80 psi Standard Tip Sizes: Tip Size GPM 7/8 160 15/16” 185 1” 200 1 1/8” 250 1 ¼” (handline) 325 1 ¼” (master stream) 400 1 3/8” 500 1 ½” 600 1 ¾” 800 2“ 1000 Appliances: Rules of thumb to remember are: - 10 psi FL. for hose appliances, such as wyes and Siamese. - Insignificant for flows < 350gpm. Elevation: - add 5 psi of friction loss per story - add or subtract .5 lbs. of friction loss per foot of elevation VIII. Drafting Procedures 1. Select Draft Site a. Optimum usage is within 10 ft vertical lift b. Need minimum 18" of water on all sides of the strainer c. Keep strainer off the bottom to avoid picking up debris (Use ladder if needed) 2. Position pumper as near as possible to the water source. a. Set parking brake 3. Attach suction hose to pump. a. Suction hose should be even to or lower than the intake b. Ensure that all connections are tight c. Ensure all drains and valves on the intake side of the pump are closed d. Use the front or opposite side intake if possible (front intake piping reduces capacity) 4. Ensure you have a means for water circulation 5. Place pump in gear in accordance with transmission instructions: 6. Primer Operation a. Set throttle to 1000 to 1200 rpm's b. Engage primer 1. Approximately 30 seconds for 1250 gpm pumps or less 2. Approximately 45 seconds for 1500 + gpm pumps 3. Add 15 seconds for front or rear intakes c. Engage primer until steady stream of water is flowing from the primer discharge hose. d. Look for pressure reading on master gauge and vacuum on the compound gauge. 7. Open circulation valve. 8. Open discharge valves slowly while increasing rpm's to maintain or increase pressure. 9. If pump fails to prime, check for the following: a. Air leaks b. Debris on strainer c. Oil level low in priming tank d. Defective priming valve e. Drafting lift to high f. Not enough water above strainer- may cause whirl pooling g. Hard sleeve higher than intake h. Primer not activated long enough 10. Maintenance after drafting: a. Refill primer oil ( if applicable) b. Back flush pump with clean water IX. Pump and Tank Capacities Engine 71 Tanker 72 Engine 74 Brush 75 - 1500 GPM 1250 GPM 1250 GPM 300 GPM 1000 Gallons 2500 Gallons 1250 Gallons 300 Gallons 35 Foam 40 Foam 10 Foam X. Foam procedure for E-71 1. Place pump in gear by following procedure in section IV 2. Open tank to pump valve 3. Prime