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