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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.
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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
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Postboks
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Nils Hansens Vei 2
0604 Oslo 6
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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
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4149-012 Porto
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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
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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
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Automotive Engine Valve Recession
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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
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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
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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
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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
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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).
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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
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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