Downloaded from SAE International by Univ of Central Lancashire, Thursday, July 02, 2020
SAE TECHNICAL
PAPER SERIES
2007-01-0264
Design and Development of the Valve Train
for a Racing Motorcycle Engine
Phil Carden and Ken Pendlebury
Ricardo UK
Naji Zuhdi
Petronas Malaysia
Andrew J G Whitehead
Del West USA
Reprinted From: New SI Engine and Component Design and Engine Lubrication
and Bearing Systems
(SP-2093)
2007 World Congress
Detroit, Michigan
April 16-19, 2007
400 Commonwealth Drive, Warrendale, PA 15096-0001 U.S.A. Tel: (724) 776-4841 Fax: (724) 776-0790 Web: www.sae.org
Downloaded from SAE International by Univ of Central Lancashire, Thursday, July 02, 2020
By mandate of the Engineering Meetings Board, this paper has been approved for SAE publication upon
completion of a peer review process by a minimum of three (3) industry experts under the supervision of
the session organizer.
All rights reserved. No part of this publication may be reproduced, stored in a retrieval system, or
transmitted, in any form or by any means, electronic, mechanical, photocopying, recording, or otherwise,
without the prior written permission of SAE.
For permission and licensing requests contact:
SAE Permissions
400 Commonwealth Drive
Warrendale, PA 15096-0001-USA
Email: permissions@sae.org
Fax:
724-776-3036
Tel:
724-772-4028
For multiple print copies contact:
SAE Customer Service
Tel:
877-606-7323 (inside USA and Canada)
Tel:
724-776-4970 (outside USA)
Fax:
724-776-0790
Email: CustomerService@sae.org
ISSN 0148-7191
Copyright © 2007 SAE International
Positions and opinions advanced in this paper are those of the author(s) and not necessarily those of SAE.
The author is solely responsible for the content of the paper. A process is available by which discussions
will be printed with the paper if it is published in SAE Transactions.
Persons wishing to submit papers to be considered for presentation or publication by SAE should send the
manuscript or a 300 word abstract of a proposed manuscript to: Secretary, Engineering Meetings Board, SAE.
Printed in USA
Downloaded from SAE International by Univ of Central Lancashire, Thursday, July 02, 2020
2007-01-0264
Design and Development of the Valve Train for a Racing
Motorcycle Engine
Phil Carden and Ken Pendlebury
Ricardo UK
Naji Zuhdi
Petronas Malaysia
Andrew J G Whitehead
Del West USA
Copyright © 2007 SAE International
ABSTRACT
This paper describes the design and development of a
direct acting valve train for high speed operation in a
racing motorcycle engine. At the outset of the project the
engine speed limiter was set to 14000 rpm and this was
eventually raised to 16000 rpm. The paper covers the
evolution of the design and includes descriptions of the
components including camshaft, tappet, shim, retainer,
valve and valve springs. Valve train dynamic analysis
software was used for the following tasks.
x
x
x
x
Assessment of the influence of the changed parts on
valve train dynamics and durability
Design of new cam profiles
Setting speed limit for each build level
Investigation of failures
These activities are covered in this paper.
INTRODUCTION
Petronas began racing the FP1 motorcycle in the World
Superbikes series in 2003. From 2004 until 2006
Petronas and Ricardo worked together to improve the 3cylinder 900cc engine of the FP1 despite the World
Superbikes rule change that permitted other teams to
use 1000cc 4-cylinder engines. Petronas decided
against changing the FP1 engine to 1000 cc due to
costs of re-homologation and so the bore (88.0 mm) and
stroke (49.3 mm) could not be changed during the
project.
During the project it became obvious that an increase in
engine speed was required to raise the power output
and to give a wider power band.
The development of high speed valve trains has a long
history and has many known problems, pitfalls and
conflicting objectives (see References 1, 2 and 3). Even
a simple direct acting valve train has a bewildering
number of design variables and constraints. Traditionally
valve train designs evolved following prolonged
prototype testing. The approach of the Engine
Development Team was to rely on advanced analysis
tools and to support decisions with minimal testing. This
paper forms a report on the valve train development.
COMPONENT DEVELOPMENT HISTORY
OVERVIEW
The rules of World Superbikes did not limit the engine
speed for a 900cc 3-cylinder engine so the Engine
Development Team hoped to increase engine power by
increasing the maximum operating speed of the engine.
The valve train was identified as the limiting factor for
engine speed and so the main objective of the valve
train development became the need for increased
engine speed. A target of 16000 rpm was set. This
corresponds to a challenging mean piston speed of
26.29 m/s for the FP1 with stroke of 49.3 mm.
The rules of World Superbikes prohibit changes to major
castings and so it was impossible to change the valve
train type from direct acting to the pad finger follower
(which has potential for lower effective mass and is used
on most Formula 1 engines). The use of pneumatic
springs was also prohibited. Thus it was necessary to
reduce the mass of all the valve train moving parts to the
minimum required for durability.
The following sections summarise the changes made to
each component.
Downloaded from SAE International by Univ of Central Lancashire, Thursday, July 02, 2020
CAMSHAFTS
The camshafts were machined from steel 16NiCr11. The
cam bearing journals and cam lobes were hardened to a
depth of 1.6 mm and then carbonitrided to give a final
surface hardness of 40-42 HRc. The bearing journals
were then lapped.
The baseline camshafts had outside diameter of 24.5
mm between lobes and inside diameter of 16 mm.
During the project the inside diameter was increased to
17 mm and the timing gear mounting flange was
scalloped as shown in Figure 1. These measures
resulted in an 8.4% reduction in mass and a 10.3%
reduction in rotating inertia of the camshafts.
However the reduction in second moment of area
resulting from this 2.25 mm diameter hole in the 5.0 mm
diameter valve stem was just 3.5% so the hollowing
procedure proved a reliable way to save 1.4 g.
The second design iteration on the intake valve was to
reduce the mass further by changing the form of the
back of the valve head from a 40 mm radius to an 11o
back angle. This gave a further mass reduction of 2.1 g.
This technique does reduce the stiffness and strength of
the valve and this can give problems with valve seat
wear or even valve failure. In addition the valve
geometry change can affect the flow characteristics of
the charge on entry to the cylinder. However, in this
case the valves proved durable and the extra engine
speed resulting from the significant mass reduction
outweighed any loss in flow.
The basic dimensions of the valves are shown in Table 1
and these were unchanged except for a slight reduction
in length to suit the change from lash cap to lash disc.
Figure 1 Camshaft end flange
Table 1 Valve dimensions
Parameter
Intake valve head diameter (mm)
Intake valve stem diameter (mm)
Intake valve length (mm)
Exhaust valve head diameter (mm)
Exhaust valve stem diameter (mm)
Exhaust valve length (mm)
VALVES
TAPPET
The baseline valves (intake and exhaust) were made
from titanium alloy Ti 6242+0.2Si. The intake valves
were solution treated and aged to give a tightly packed
+ (alpha+beta) grain structure, exhibiting high
hardness (42-45HRc) and excellent fatigue and creep
resistance at intake valve temperatures (up to 500qC).
The exhaust valves were heat treated and aged to give
a lamella alpha beta structure within prior beta grains,
exhibiting fatigue and creep resistance up to 800qC.
The tappets were made from a through hardened tool
steel (H11), heat treated to 51-54 HRc, hard turned,
ground and then polished to around 0.05 Ra prior to
application of a diamond-like carbon (DLC) coating
developed. The DLC coating was approximately 3 Pm
thick and was applied by a plasma-assisted chemical
vapor deposition process following surface preparation.
This coating was extremely hard (2500 Hv) and was
used to minimise friction and wear.
Both valves had a molybdenum alloy coating, plasma
sprayed on the stem to minimize friction and prevent
wear of the valve guide. The valve tip was protected by
a Cr2C3 NiCr cermet coating applied via a high velocity
oxy-fuel flame process. Further to these coatings a thin
film plasma vapor deposition CrN coating was tested on
the seat of the intake valve to address a valve seat
degradation problem. The coating performed well but it
was determined that the seat wear was due to intake
debris being ingested and so this coating was not
adopted for race production parts.
The baseline tappet was of traditional design and the
first design iteration involved the introduction of large
slots into the skirt of the tappet to reduce mass (see
Figure 2). This concept had been previously subject to
limited development but had not been used because
most very high speed engines had switched to finger
follower valve trains for lowest mass. However, the
durability of the slotted tappet was proven quickly and
resulted in a significant mass reduction of 4.9 g.
The valve supplier had developed a technique to reliably
produce hollow valves with stem diameter as small as
4.5 mm. In an effort to reduce the mass of the intake
valve this method was applied to the FP1 intake valve.
Assuming the valve train was under control the main risk
of failure from this modification was valve stem bending.
Baseline
36.0
5.0
93.7
30.0
5.0
95.0
Final
36.0
5.0
93.2
30.0
5.0
94.5
Downloaded from SAE International by Univ of Central Lancashire, Thursday, July 02, 2020
Table 3 Spring dimensions
Baseline
Parameter
inner
outer
Wire diameter (mm)
2.90
3.95
Mean coil
17.9
24.75
diameter (mm)
Fitted length (mm)
32.6
32.6
Fitted force (N)
153
298
Final
inner
outer
2.90
3.7
17.1
23.7
28.75
130
28.75
279
SPRING RETAINER, SHIM AND COLLETS
Figure 2 Slotted tappet skirt
A second iteration with reduced crown thickness and
reduced wall thickness was introduced (see Table 2)
and this resulted in a further mass reduction of 2.6 g.
These parts were tested and proved to have adequate
durability although they did occasionally exhibit some
signs of wear of the DLC probably due to flexing of the
tappet crown.
Table 2 Tappet dimensions
Parameter
Outside diameter (mm)
Overall length (mm)
Central crown thickness (mm)
Minimum wall thickness (mm)
Slots in skirt
Baseline
32.0
23.5
3.3
0.8
No
Final
32.0
23.5
3.0
0.7
Yes
The spring retainers were machined from maraging steel
(C300) and nitrided to achieve hardness of 61-64 HRc
on the mating surfaces for the springs. A through
hardened steel lash cap was located on the top of each
valve stem. This design resulted in several problems as
the speed of the engine was increased (see later
sections) and so the components were redesigned to
improve durability and to reduce mass.
The first design iteration involved a move away from the
lash cap concept to a lash disc held captive in a revised
spring retainer (see Figure 3). This allowed for a
reduction in mass, an improvement in durability of the
retainer and elimination of valve tip wear. C300 was
used for the revised retainer and the lash disc was made
from cold worked die steel, through hardened to HRc 5761, then ground and polished.
SPRINGS
The springs used throughout the project were made
from steel. The spring manufacturers were very
protective regarding materials, surface treatments and
fatigue strength data so no further data can be provided.
The baseline springs had been designed for a rated
speed of 13500 rpm and problems soon emerged as
engine speed was raised (see later section).
Springs from several different suppliers were procured
and tested but eventually the spring supplier was
changed to NHK who provided a spring pack of their
design to meet a target maximum speed of 15500 rpm
for a cam profile with 12 mm peak lift and a target life of
3 million cycles. The outer spring was given a right hand
coil and the inner spring had a left hand coil. The coils
were designed with a target interference of 0.1 mm and
the ends were ground and carefully chamfered. This
spring pack was used for most of the project but during
the final season a further optimized spring pack (see
Table 3) was designed to suit the cam profile with
reduced peak lift of 11 mm and to give a higher rated
speed of 16000 rpm.
Figure 3 Sections through baseline and final intake
valve trains
The revised spring retainer had different diameters to
suit the final spring pack design. The position of the
collet groove relative to the valve tip was changed to suit
the new retainer design as shown in Figure 3 but the
collets remained unchanged during the project.
Downloaded from SAE International by Univ of Central Lancashire, Thursday, July 02, 2020
COMPONENT MASSES
The changes made to
summarised in Table 4.
Table 4 Valve train mass
Mass (kg)
Intake valve
Exhaust valve
Tappet
Shim (cap or disc)
Collets
Retainer
Outer spring effective mass
Inner spring effective mass
Intake valve train
effective mass
Exhaust valve train
effective mass
component
masses
are
Baseline
0.0243
0.0222
0.0335
0.0010
0.0002
0.0080
0.0115
0.0062
0.0847
Final
0.0208
0.0214
0.0260
0.0008
0.0002
0.0072
0.0093
0.0051
0.0694
0.0826
0.0700
The final valve train is shown in Figure 4.
Figure 4 Section through final valve train
VALVE TRAIN DESIGN ANALYSIS
The valve train analysis software had a kinematics
solver (used to design cam profiles and calculate
pseudostatic forces and oil film thickness etc) and a
dynamics solver (used to determine dynamic valve
motion, dynamic forces and spring surge vibration).
The results of the analysis of the intake valve train,
comparing the baseline design with the final design, are
presented in this section. The larger valve and more
aggressive, higher lift cam profile conspired to make the
intake valve train the worst case in almost every respect
and so results are shown for intake only.
Figure 5 shows a comparison between the kinematic
acceleration of the intake valve for baseline and final
designs.
Figure 5 Kinematic valve acceleration against crank
angle
Table 5 shows some important kinematic parameters for
the baseline and final designs. The peak lift was reduced
by 1 mm and the period was increased slightly. The
choice of peak valve lift and period was obviously a
compromise and the lift was reduced to enable the high
speed operation. The final choice was guided and
supported by performance simulation and by extensive
engine performance testing on a series of cam profiles.
The reduction in peak lift entailed a corresponding drop
in intake valve L/D ratio but it was possible to slightly
increase the lift area integral (a non-dimensional
parameter defined as the area under the lift curve
divided by the theoretical maximum area under the lift
curve).
Table 5 Kinematic parameters
Parameter
Baseline
Peak kinematic valve lift L
12.0
(mm)
Inner seat diameter D (mm)
35.0
L/D
0.343
Lift area integral
0.555
Period – top of ramp (deg)
307.2
Ramp height (mm)
0.20
Ramp velocity (m/s)
0.432 @
14000 rpm
Valve acceleration on
29818 @
opening flank (m/s2)
14000 rpm
Valve acceleration on cam
11530 @
nose (m/s2)
14000 rpm
Valve acceleration on
36962 @
closing flank (m/s2)
14000 rpm
Opening side acceleration
2.51
ratio
Closing side acceleration
3.21
ratio
Final
11.0
35.0
0.314
0.557
310.0
0.20
0.500 @
16000 rpm
33404 @
16000 rpm
13305 @
16000 rpm
41554 @
16000 rpm
2.51
3.12
The ramp height was maintained at 0.2 mm and the
ramp kinematic velocity was also maintained at the
same value despite the increase in engine speed and
the corresponding increase in actual ramp velocity
indicated in the table. The peak kinematic valve
accelerations (nose and flanks) were decreased by 13-
Downloaded from SAE International by Univ of Central Lancashire, Thursday, July 02, 2020
16% as shown in Figure 5 but the dynamic peak
acceleration values were increased by 12-15% due to
the increase in engine speed. Acceleration ratios were
maintained at similar values.
The durability of the cam/tappet contact was assessed in
terms of contact stress, lubricant film thickness and
tappet edge clearance (the proximity of the cam/tappet
contact line to the edge of the tappet). The results are
shown in Table 6.
The highest values of contact stress occurred at low
engine speed at the nose of the cams but racing engines
spend very little of their lives at low speed. The final
design had a larger nose radius and so correspondingly
lower contact stress. At high engine speed this situation
was reversed (see Figure 6) and the final valve train
experienced higher contact stress than baseline at rated
speed.
characteristic shape associated with direct acting valve
trains for baseline and final valve trains. The film
thickness is high on the flanks and low over the nose.
The lubricant film thickness at the nose is often
assessed by using the Deschler and Wittman number
(see Reference 2). This is usually in the range 0.15 to
0.25. For the final design this upper limit was exceeded
slightly but no problems occurred.
At the flank/nose transition the film thickness passes
through a very low region as the lubricant entrainment
velocity passes through zero. The lubrication in this
region is assessed by consideration of the number of
consecutive crank degrees at which predicted oil film
thickness is less than 0.1 Pm and this is usually
expected to be less than 10 deg.
The cam/tappet contact stress over the nose was also
calculated using dynamic forces.
Table 6 Cam/tappet durability
Parameter
Baseline
Peak cam tappet contact
831 @
stress at idle (N/mm2)
3500 rpm
400 @
Peak cam tappet contact
14000 rpm
stress at rated speed
(N/mm2)
Lubricant film thickness at
0.295
peak cam lift (Pm)
Deschler and Wittman
0.207
number at peak lift
8.26
Maximum number of
consecutive crank degrees
at which oil film thickness is
less than 0.1 Pm at rated
speed
Minimum tappet edge
0.30
clearance (mm)
Final
764 @
3500 rpm
436 @
16000 rpm
0.278
0.272
7.86
Figure 7 Lubricant film thickness at cam/tappet
contact against crank angle
For the baseline cam profile with high lift the tappet edge
clearance was very low (0.3 mm) but as the lift was
reduced this increased to a value higher than necessary.
Reduction in tappet diameter was considered but in fact
this was limited by the need for clearance between the
tappet and the outer spring and so the potential benefit
was small.
1.90
Figure 6 Pseudostatic cam/tappet contact stress at
high engine speed
The lubricant film thickness is shown plotted against
crank angle in Figure 7. This graph shows the
The valve trains were also assessed in terms of their
dynamic performance. This software used began as a
dedicated valve train dynamics package and although it
has expanded to have extensive multi-body dynamics
capability it retains exceptional speed and ease-of-use
for valve train dynamic analysis. The model used for
assessment of valve train dynamics is shown in Figure
8.
Downloaded from SAE International by Univ of Central Lancashire, Thursday, July 02, 2020
Figure 9 Valve seating velocity against engine speed
Figure 8 Single valve line model
The cam node was suspended on a stiffness element
representing camshaft bending stiffness and camshaft
bearing support stiffness. The tappet top stiffness was
modeled as a function of eccentricity of the cam/tappet
contact. The valve stem was generally modeled as a
single axial stiffness although multiple stiffness and
mass models were tried. Valve seat and valve head
bending were also modeled as a single stiffness. The
valve, spring retainer, shim and collets were modeled as
a single lumped mass. Each valve spring was modeled
as a series of lumped masses (8 per coil) connected by
stiffnesses using a special macro element that accounts
for coil clash effects as the spring closes and loss of
contact between spring ends and mating parts.
The level of damping due to interference between valve
springs is intimately dependent on the fit between the
springs (which changes as the springs are compressed)
and so is very difficult to model explicitly. For this project
the approach taken was to make two analysis runs to
assess each design; one with high damping (assumed
20% of critical damping) and one with very low damping
typical of independent springs with no interference (0.5%
of critical damping). The sensitivity of all aspects of
system dynamics to damping was thus considered at
every stage.
At low speeds the valve seating velocity was controlled
by the closing ramp on the cam but as engine speed
increased this control was progressively lost. The
baseline design showed an increase in seating velocity
from ~13500 rpm and a sharp transition to very high
seating velocity at ~14800 rpm. During the project, as a
result of engine testing and cylinder head rig testing, it
was discovered that the titanium intake valves would fail
suddenly at the valve stem below the retainer if valve
seating velocity exceeded 4 m/s. This information, with a
suitable safety margin, was used to set the engine
speed limiter for each valve train build. For the final
design the valve seating velocity did not exceed 2 m/s
until the engine reached speeds in excess of 16500 rpm
and did not experience a sharp transition to very high
seating velocity at engine speeds below 17000 rpm. This
characteristic proved very significant as the valve train
was able to survive the inevitable over-rev incidents that
occur during racing. It was interesting to note that the
valve seating velocity was not very strongly dependent
on the assumed damping due to interference between
the springs.
Further insight into intake valve closing can be obtained
by observation of Figure 10 which shows the valve
closing event for the baseline design at 9000 rpm, 13000
rpm and 15000 rpm.
The software was used to calculate the dynamic
response of the valve train in the high speed range and
the results were mainly presented in terms of plots of
critical parameters against engine speed. Figure 9
shows a plot of valve seating velocity which was used to
assess the severity of the valve closing event.
Figure 10 Valve lift at closing against crank angle
Downloaded from SAE International by Univ of Central Lancashire, Thursday, July 02, 2020
At 15000 rpm the valve hit the seat before the top of the
ramp, seating velocity was high and initial contact was
followed by a large valve bounce. The height of this first
bounce can also be plotted against engine speed and
the results are shown in Figure 11.
Cam/tappet separation showed more sensitivity to spring
interference damping than valve seating phenomena but
the general conclusions were not changed. The baseline
design exhibited a significant loss of contact between
cam and tappet over the cam nose at speeds in excess
of ~14500 rpm while for the final design separation did
not exceed ~0.2 mm at speeds below 17000 rpm.
Spring surge is a well-known problem associated with
high-speed coil-sprung valve trains. The springs are
excited by the harmonic content of the cam profile and
may continue to vibrate following valve closure, possibly
affecting the motion during the next valve event. The
team quantified this phenomena by plotting the
amplitude of the residual vibration of the centre coil of
each spring just after valve closure as shown in Figures
13 and 14.
Figure 11 Valve bounce height against engine speed
The results show a very similar pattern to the valve
seating velocity results. Typically The team regarded
bounce height in excess of 0.1 mm as unacceptable and
this limit correlated reasonably well with the limit
established for valve seating velocity.
The next phenomena considered during dynamic
analysis was valve jump. This was characterized by loss
of contact between the cam and the tappet during the
valve event when inertia force (including vibration)
exceeds available spring force. Figure 12 shows a plot
of maximum distance between cam and tappet against
engine speed.
Figure 13 Outer spring surge against engine speed
Figure 14 Inner spring surge against engine speed
Figure 12 Cam/tappet separation against engine
speed
The spring surge amplitude for the baseline springs was
very high (~1 mm expected for production engines but
higher values can be tolerated for racing) and this
contributed to limited high speed dynamic performance.
The surge amplitude for the final design was greatly
Downloaded from SAE International by Univ of Central Lancashire, Thursday, July 02, 2020
reduced by improved matching between cam and spring
characteristics. The surge amplitude was moderately
sensitive to spring interference damping.
One of the problems associated with high spring surge
amplitude was spring seat hammering. The spring
vibration leads to loss of contact between spring and
seat followed by large impact forces as contact is reestablished (for example see Figure 15 at end of event).
Figure 17 Maximum force at bottom of outer spring
against engine speed
The final cam profile and spring design showed
significant improvement in valve spring seat force
characteristic.
Figure 15 Force at bottom of outer baseline spring
against crank angle
This problem manifested itself as breakage of spring end
tangs with some intermediate spring designs during this
project (see later section) and so plots of minimum and
maximum spring force as shown in Figures 16 and 17
were made routinely to assess the risk of this problem.
Figure 16 Minimum force at bottom of outer spring
against engine speed
The software was also used to calculate dynamic stress
in the springs. Stress at the worst case location on the
baseline outer spring is shown plotted against crank
angle in Figure 18 and stress range is plotted against
engine speed in Figure 19. Although the spring fatigue
strength was not known the effect of valve train
dynamics on spring stress was quantified and this
proved useful during the project as a basis for
comparison between designs.
Figure 18 Stress at worst case location in baseline
outer spring against crank angle
Downloaded from SAE International by Univ of Central Lancashire, Thursday, July 02, 2020
Figure 20 Whole engine rotational dynamics model
Figure 19 Stress range at worst case location in
baseline outer spring against crank angle
Some results from spring analysis are given in Table 7.
Table 7 Spring cover factor and stress
Parameter
Baseline
Pseudostatic spring cover
1.21 @
factor at rated speed
14000 rpm
298 fitted
Shear stress in outer
1065 max
spring -pseudostatic
(N/mm2)
356 fitted
Shear stress in inner
1066 max
spring -pseudostatic
(N/mm2)
Final
1.39 @
16000 rpm
410 fitted
1294 max
291 fitted
1199 max
Pseudostatic spring stress levels were increased for the
final design but spring quality was such that spring
fatigue failures were not a significant problem.
An analysis model of the whole engine was created (see
Figure 20). Measurements of torsional vibration
displacement were made at each gear in the timing drive
and the values of stiffness, damping and clearance in
the model were adjusted to give reasonable correlation
with measured data. The model was used to investigate
the following.
x
x
x
Effects of crankshaft dynamics and timing drive
dynamics on valve motions
Dynamic loads on timing gears
Dynamic torques at gear fasteners
A typical plot comparing intake valve seating velocity, as
calculated using the single valve line model of the final
valve train, with predicted values for each intake valve
train from the whole model is shown in Figure 21. It is
interesting to note that in this case (3 cylinder engine
with geared timing drive) although the valve motions
were obviously affected by the timing drive dynamics the
magnitude of the effect was not very significant. It was
thus possible to use a single valve train analysis model
to set speed limits for valve train builds.
Figure 21 Comparison of intake valve seating results
from single valve model and whole engine model
VALVE TRAIN TESTING
A cylinder head test rig was used for durability testing of
valve train variants during the later stages of the project.
In particular, testing of alternative valve spring packs
proved very valuable (see later section). The rig was
also used to measure valve train friction power loss.
The idler gear on the timing drive was driven by an
electric motor via a drive shaft with a flexible element
and the cylinder head was supplied with lubricating oil at
100qC and coolant at 80qC.
The rig was generally operated at steady state speeds
for predefined periods.
Downloaded from SAE International by Univ of Central Lancashire, Thursday, July 02, 2020
VALVE TRAIN FAILURE MODES
TAPPET BORE FAILURE
The first major problem to occur as the speed of the
engine was increased was the failure of tappet bores.
The cylinder head cracked and the tappet bore structure
broke off on several engines as shown in Figure 22.
Use of die penetrant on used cylinder heads revealed
that the cracks began at the fillet radius in the slot
machined to provide clearance for the cam lobe as
shown in Figure 23.
Figure 24 Stress concentration in tappet bore slot
VALVE TIP WEAR
The lash cap used to control valve clearance on the
baseline valve train resulted in sporadic examples of
high wear rate at the valve tip surface as shown in
Figure 25.
Figure 22 Tappet bore failure
Figure 25 Valve tip wear
Figure 23 Crack in tappet bore
The software was used to calculate the pseudostatic
contact force between the cam and the tappet and the
eccentricity of the contact point on the tappet top. These
values were then used to calculate the tipping moment
acting on the tappet which in turn was used to calculate
the reaction forces between the tappet side and the
tappet bore at the top and bottom of the tappet. These
forces were applied to a local finite element model of the
cylinder head in the tappet bore region (see Figure 24),
low safety factors were confirmed and various schemes
for improvement were investigated. Eventually the fillet
radius was increased and this failure was eliminated.
This was thought to be due to inadequate control of
tolerances on the bore of the lash cap leading to variable
fit between the cap and the valve. Large diametral
clearance between these components led to tipping of
the cap and edge loading of the valve tip. This problem
could probably have been solved by improving the fit
and surface finish in the lash cap but it was eventually
eliminated by the change of design to the lash disc.
VALVE STEM FAILURES
As engine speed was increased a series of failures of
the intake valve stems occurred. The stems typically
broke just below the bottom of the spring retainer, at the
end of the collet contact area, as shown in Figure 26.
Downloaded from SAE International by Univ of Central Lancashire, Thursday, July 02, 2020
The transition to damaging level of seating velocity
occurred at ~300 rpm lower speed and the resulting
impacts were strong enough to break the valve stems.
The problem was resolved by changing the spring
supplier and introducing an improved spring design for
high speed operation. The relaxation of the final springs
was greatly reduced even though they were subjected to
higher stresses as shown in Table 7. Fitted force was
typically reduced by less than 3% if the engine was run
at less than 16000 rpm and this rose by up to 10%
following prolonged operation at 16000 rpm.
VALVE SPRING END TANG FAILURES
Eventually the root cause was traced to the very large
relaxation of the baseline valve springs. Spring
relaxation (or loss of spring force) following use is known
to be dependent on stress level, time and temperature
and typically it results in a reduction in fitted force of
~5%. Measurement of spring force/displacement curves
before and after use revealed that the fitted force of the
baseline springs was reduced by up to 40% following
high speed operation. Figure 27 shows the effect of a
20% reduction in spring fitted force on the valve seating
velocity. Loss of control of the valve spring can also
result in application of bending loads on the valve stem
and this may also have contributed to the failure.
Springs were subjected to periods of constant speed
operation followed by measurement of spring pack fitted
force and further testing at increased speed as
illustrated in the typical test results graph shown in
Figure 28.
100
18000
90
16000
80
70
14000
period 1
period 2
12000
60
50
period 3
Fitted relaxation
40
10000
30
Spring relaxation force (N)
The failure surface indicated pure tensile fatigue so the
applied loads were probably far higher than expected.
Previous experience suggested that loss of valve train
control could be to blame despite the fact that analysis
suggested the dynamics should be acceptable at the
failure speed. This led to a general re-investigation of
the valve train dynamics.
Engine speed (rpm)
Figure 26 Valve stem failure
The most important valve train failure mode was valve
stem failure following spring relaxation (as described
previously) and with this in mind a procedure for
evaluation of alternative spring designs on the cylinder
head rig was developed.
20
8000
10
6000
0
0
100000
200000 300000 400000
Spring cycles
500000
600000
Figure 28 Spring relaxation chart
The final springs relaxed by just 7 N after 500000 cycles
(including 120000 cycles at 16000 rpm engine speed).
This increased to 33.5 N after a further 50000 cycles at
16500 rpm. On average this process revealed
interesting differences between the final spring pack
(that was eventually chosen) and the best competitor as
shown in Figure 29. The relaxation of the final springs
was the lowest of the springs tested.
Figure 27 Effect of spring relaxation on valve seating
velocity
Downloaded from SAE International by Univ of Central Lancashire, Thursday, July 02, 2020
x
140
Competitor
Averrage relaxation force (N)
120
NHK
x
100
Success was achieved by making extensive use of
dynamic simulation combined with minimal rig
testing.
The contribution of world class component suppliers
to the success of the project was invaluable.
80
ACKNOWLEDGMENTS
60
40
The authors would like to thank senior management at
Petronas for providing permission to publish this paper.
20
REFERENCES
0
Fitted
Compressed
Figure 29 Comparison of spring relaxation following
final testing with 50000 cycles at 16500 rpm
The best competitor spring pack also suffered from
failure of the end tangs (as shown in Figure 30) when
subjected to speeds in excess of 16000 rpm.
1. Valve train design for multivalve automotive gasoline
engines. Heath, A.R. SAE 885133, 1988
2. Optimal design of high speed valve train systems.
David, J.W. Kim, D. Covey, J.A. SAE 942502, 1994.
3. The application of advanced simulation methods in
the design of Formula 1 valve trains. Di Paola, G.
Smith, A. 19th International Vienna Engine
Symposium, 1998, Vol. 1.
4. The design of cams for flat faced followers with
regard to elastohydrodynamic lubrication. Deschler,
G. and Wittman, D. MTZ March 1978.
CONTACTS
Phil Carden
Ricardo UK
Shoreham Technical Centre
Shoreham-by-sea
BN43 5FG
UK
Tel +44 1273 794959
e-mail phil.carden@ricardo.com
Figure 30 Failure of spring end tang
This failure probably occurred due to loss of contact
between spring and seat followed by impact force when
contact was re-established (see Figures 16 and 17).
CONCLUSIONS
The design/development work on the valve train for the
Petronas FP1 had the following conclusions.
x
The final valve train had exceptional durability at
rated speed (speed limiter set to 16000 rpm) and
was able to survive overspeed events at up to 17000
rpm without failure.
Naji Zuhdi
Powertrain Technology,
Block E, PETRONAS Research & Scientific Services
Sdn. Bhd.,
LOT 3288 & 3289, Off Jalan Ayer Itam,
Kawasan Institusi Bangi, 43000 Kajang,
Selangor D.E., Malaysia
Phone : +603-89244500
Fax : +603-89244548
e-mail ahmadnaji@petronas.com.my
Andrew Whitehead
Del West USA
28128 W Livingston Av
Valencia, CA 91355
USA
Tel +1 661 295 5700
e-mail andrew.whitehead@delwestusa.com
Downloaded from SAE International by Univ of Central Lancashire, Thursday, July 02, 2020
DEFINITIONS, ACRONYMS, ABBREVIATIONS
Cam flank: Part of cam profile with positive acceleration
DLC: Diamond like carbon
Tappet edge clearance: The tappet radius minus the
maximum eccentricity of the cam/tappet contact point
L/D ratio: Peak valve lift divided by inner valve seat
diameter
Cam nose:
acceleration
Part
of
cam
profile
with
negative