Automotive Engine Tribology

Wear 221 Ž1998. 1–8

Automobile engine tribology—design considerations for efficiency and durability C.M. Taylor

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School of Mechanical Engineering, The UniÕersity of Leeds, Woodhouse Lane, Leeds LS2 9JT, UK Received 19 June 1998; accepted 19 June 1998

Abstract In the United Kingdom, the Technology Foresight Programme wHMSO, Progress Through Partnership 1 Ž1995. 126 pp.x, through its Transport Panel, revealed the requirement for key generic technologies and scientific research in relation to automobiles. One of the ‘three major development opportunities which will help accommodate increased demand in a sustainable way’ was identified as ‘ vehicles with greater efficiency and reduced environmental impact’. In addition, further studies in ‘fuel efficiency’ and ‘simulation and modelling’ were recommended. The total scope of tribological considerations with regard to the above prospective research themes is immense and the present paper will focus upon the major frictional components of the automobile engine, that is, the bearings, the valve train and the piston assembly. In particular, the current position surrounding the modelling of these components will be reviewed and future possibilities identified. Prediction of overall engine friction will be addressed and the specific issues of modelling of lubricant behaviour and the role of surface topography touched upon. q 1998 Elsevier Science S.A. All rights reserved. Keywords: Tribology; Efficiency; Durability

1. Introduction Recent UK Government statistics w1x revealed that in 1994, there were 25 million vehicles registered for use, 85% petrol-fuelled and 15% diesel, with a very small gas-, liquefied petroleum gas- and electricity-powered group. The reciprocating internal combustion engine accounts for the vast majority of power units used in petrol and diesel engines, and whilst it is remarkably reliable and versatile, it does have drawbacks. These drawbacks can have enormous influences on national and international economies. The UK road vehicle figures mentioned above are reflected in even more impressive global statistics with the USA, Europe and Asia having some 500 million motor vehicles in 1993. The drawbacks of the internal combustion engine include relatively low thermal and mechanical efficiencies with a very significant proportion of the energy of the fuel

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being dissipated as heat. In addition, the IC engine contributes to environmental pollution through particulate, nitrous oxide and hydrocarbon emissions and to the greenhouse effect via carbon dioxide emissions. Fig. 1 shows in the broadest sense the manner in which the energy of the fuel is distributed. About 60% of the energy is dissipated in the form of heat, either from the engine surfaces or down the exhaust pipe. Mechanical actions may account for a further frictional loss of the order of 15%, leaving only a quarter of the original energy in terms of brake power. In a similar spirit, Fig. 2 describes the breakdown of the mechanical losses in the engine with the piston assembly being responsible for the lions share, approaching one-half. It will be noted that other losses associated with accessories and pumping can be 20% or more of the mechanical losses. It is clear that the energy distribution and mechanical loss distribution will vary with engine type and operating conditions. Indeed, an oil pump in say, a 1600 cm3 modern day petrol engine can absorb some 2–3 kW at the limiting engine speed, whilst the corresponding figure for a formula 1 racing car can be 20 kW.

0043-1648r98r$ - see front matter q 1998 Elsevier Science S.A. All rights reserved. PII: S 0 0 4 3 - 1 6 4 8 Ž 9 8 . 0 0 2 5 3 - 1

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C.M. Taylorr Wear 221 (1998) 1–8

Fig. 3 represents data provided by Anderson Ž1991. w2x identifying the distribution of fuel energy in a passenger car during an urban driving cycle. These data reveal that only 12% of the available power finds its way to the driving wheels, whilst 73% is lost to coolingrexhaust and pumping and 15% to mechanical losses. It is possible to do some amusing and alarming sums on the basis of this as indicated below. The automobile

Fig. 1. Typical fuel energy distribution in an internal combustion engine.

The laws of thermodynamics indicate that with the Carnot cycle, an ideal engine could achieve a maximum efficiency of about 85% for typical heat generation and rejection temperatures. The low heat rejection and the high thermal efficiency cannot be achieved in practice for a number of reasons. Quite naturally, the substantial thermal losses which are encountered have led to engineers and researchers devoting enormous energies to improving the efficiency of the IC engine, particularly since the energy crisis of the mid 1970s. In particular there, have been intense studies of the combustion process and engine designrfuelling development improvements linked to this. Further, low heat loss rejection engines have received attention and maximum brake efficiencies of over 45% have been reported. However, it has been recognised that worthwhile efficiency gains might be effectively had by paying attention to the more modest mechanical losses.

Fig. 2. Mechanical losses distribution in an internal combustion engine.

Power to wheels during city driving Cost of wasted powerrl Ž$1rl petrol, UK. Total cost to UKryear Ž25 million vehicles, 10,000 mileryear, 5.5 mile rl. Total worldwide costryear Ž625 million vehicles.

12% $0.88

$40 billion

$1 trillion

From the information in Fig. 3, it will be noted that a reduction in the mechanical losses of, say, 10% would lead to a reduction in the fuel consumption of 1.5%. For an average vehicle lifetime, driving a distance of 125,000 miles, using the data in the box above, this would amount to a fuel saving of 340 l of petrol during the lifetime of the car, equating to about $340, quite apart from environmental gains due to reduced emissions. On a worldwide scale,

Fig. 3. Power distribution in an automobile during city driving.


the potential economies, in the round, are clearly substantial. In order to address the durability and frictional losses in the major tribological components of an automobile engine, it is necessary to have a clear understanding of the engineering science underpinning their operation. Currently, the ability to achieve improved modelling of components and deliver reliable predictions of behaviour, before cutting metal in a prototype, is a major objective of all industrial sectors including the automotive field. This is being made possibly by the spectacular advance in computational power now available on the desktop of the engineering designer, coupled with substantive developments in the powerful solvers needed to secure solutions to the complex equations arising in the modelling. The remainder of this paper will examine the major tribological engine components from this standpoint, investigating the current position and prospective developments, and will identify some important developments and possible contributions. In interpreting this, it is important to have an understanding of the basic aspects of lubrication as reflected in the sketch version of the Stribeck diagram shown in Fig. 4. In summary the modes of lubrication shown in Fig. 4 are the following. Ža. Hydrodynamic—in which the surfaces are completely separated by a film of lubricant, and the generation of pressures in the film to carry the load is achieved by classical hydrodynamic action, in which the dynamic viscosity of the lubricant is the prime lubricant characteristic.

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Žb. Elastohydrodynamic—in which the surfaces are also in theory separated, but the contact is much more concentrated, the films are thinner and other physical phenomena Želastic distortion of the surfaces and the effect of pressure on dynamic viscosity. are influential. Žc. Mixed—in which a lubricated contact experiences some degree of asperity contact between surfaces and the overall behavioural characteristics and load capacity are defined by both Želasto. hydrodynamic and boundary Žsee below. influences. Žd. Boundary—in which it is the physical, and in particular chemical, actions of thin films attached to the surfaces which define performance—lubricant dynamic viscosity is not important but the additive package will have a significant role. Fig. 4 reflects the enormous challenge facing the lubricant technologist with regard to the use of a single product in an engine. The tribological components rely upon different lubrication mechanisms at different times, and the oil specification will be a compromise in terms of durability, energy efficiency and emission requirements. 2. Automobile engine tribological components 2.1. Engine bearings The modern day analysis of engine bearings dates back to the mid 1960s and is associated primarily with the method developed by Booker w3,4x. In this approach, now computerized, the applied bearing loading which is taken to be known a priori, is balanced by the load capacity generated by hydrodynamic action in the thin fluid film. The method yields the important design parameter, the minimum film thickness, together with an estimate of maximum lubricant pressure and other parameters as desired. The Mobility Method developed by Booker has become a remarkably robust and reliable design tool for engine bearing design. As with the other engine friction components, its predictions are, however, only benchmark because of the assumptions made in the analysis. The list below is helpful in scoping the conditions typically encountered in a modern day main or big end bearing w5x. Conditions in an automobile engine bearing Minimum lubricant film thickness Maximum shear rate Žup to. Maximum bearing temperature Specific loading Žup to. Maximum film pressure Lubricant dynamic viscosity Power loss per bearing Žhigh speed. Flow rate per bearing Žtypically.

Fig. 4. The Stribeck diagram identifying the regimes of lubrication as conventionally associated with specific lubricated engine components.

1 mm 10 8rs 120–1508C 60 MPa 250 MPa 2.5 = 10y3 Pa s 0.25 kW 0.015 lrs

The assumptions made in the mobility analysis are quite restrictive and these are listed below, with a corresponding

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indication of some instances where these have been relaxed by researchers. Mobility method Relaxation of assumption— of analysis of engine design challenges bearings—assumptions Short-bearing analysis

Circumferential symmetry Rigid surfaces

Perfect alignment Lubricant freely available Newtonian lubricant No vibrational instability

Isothermal

Full width-bearing analysis straightforwardly possible, but at substantial cost in analysis complexity and time. Analysis of groovingrhole effects emerging w6,7x but complex. Elastic and thermal effects being considered w8–12x, but complex and difficult to derive generalised design data. Misalignment rarely considered. Oil film history effects possible now, with computational effort becoming reasonable w13,6,7x. Consideration of shear thinning and pressure viscosity effects beginning to emerge w14,15x. Neglect of component inertia reasonable but little consideration of bearing stiffness and damping on overall vibrational characteristic of engine bearings. Few studies of thermal effects in bearings w16,17x or total engine thermal environment.

Studies of the performance of engine bearings, which were so active in the 1960s and 1970s, have taken second place to development of the tribological understanding of piston ringsrpiston assemblies and the valve train. However, evidence is beginning to emerge that with the continued development of smaller engines, with higher power outputs and speeds, that bearing problems are re-emerging. Increasing attention to engine bearing design in all its aspects is a priority for the future. 2.2. ValÕe train Valve train lubrication problems have presented difficulties to engine designers throughout the century. However, the introduction of the overhead camshaft in the early 1970s led to increased difficulties. OHC arrangements allowed improved production arrangements since the cylinder head could be assembled separately from the block. In addition, they operated better at high speeds and proved to be less expensive w18x. It is unfortunate that OHC designs were found to be inherently poor from a tribological point

of view. For the last 20 years, almost all the major automobile manufacturers have experienced cam and follower failures and problems persist. There are, of course, a wide variety of valve train mechanisms Že.g., push rod, pivoted follower, direct acting, roller follower, desmodromic, etc.. and general considerations of cam and follower lubrication may be found in Taylor w19,20x. A cam and follower contact is a form of concentrated contact Žnotionally elastohydrodynamic. which is traditionally reflected as operating in the boundary lubrication regime Žsee Fig. 4.. That is, the viscosity of the lubricant is seen to be unimportant in the lubrication process, but the establishment of thin surface films of molecular proportions due to physical and chemical actions linked to the additive package, are seen to be crucial. This is indeed the case, though more recent studies Žsee Ref. w19x. have already evidenced the contribution of elastohydrodynamic lubrication and the importance of effective mechanical design to establish the best lubrication conditions possible. Forms of zinc dialkyldithiophosphate ŽZDDP. are the most common extreme pressure additive used in oils to establish the thin films required; however, the prospect of the emission of zinc and phosphorous being seen as unacceptable for an environmental standpoint is a real one. The additive to replace ZDDP is not apparent at the current time. In the broadest sense, the valve train presents a wide spectrum of contacts to interest the tribological designer— the camsrfollowers, camshaft bearings, valve guides, lash adjusters, bucket guides, pivots and belt drives. Even if one restricts interest to the most important couple, the cam and follower, the range of tribological interests are extensive— metallurgyrfailurerwear, geometryrleftrkinematics, dynamicsrvibrationsrshockrnoise, additive packagerboundary lubricationrsurface layers, friction, variable valve timing, fluid film lubrication and design philosophy. The design of a cam and follower from a tribological context has improved significantly in the last 20 years, and has lead to the development of straightforward and sound procedures and software Že.g., Ref. w18x.. The steps are: Ø Kinematic considerations using standard methods of differential geometry; Ø Load calculation, normally considering only inertia and spring effects; Ø Film thickness evaluation from standard elastohydrodynamic formulae; Ø Stress calculations from Hertzian theory; Ø Friction and power calculations with a limiting traction assumption. Although such design procedures do not address directly boundary lubrication effects and the failure mechanisms which are encountered—pitting, scuffing and polishing wear—they do enable sound decisions as to mechanical design which contribute extensively to the production of effective mechanisms. The figures below will help to enable a feel for the operating conditions of the cam and


follower in a modern, four-cylinder, 1.8 l, four valvesrcylinder IC engine with double overhead camshafts w21x. Conditions in the cam and follower of a modern IC petrol engine with a direct acting cam and follower Camshaft speed 750 rpm 1500 rpm 2250 rpm 0.08 mm 0.13 mm 0.18 mm

Film thickness Žat the nose. Hertzian stress 635 MPa 609 MPa Žat the nose. Power loss 160 W 320 W Ž8 camsrfollowers. — intake camshaft Power loss 170 W 350 W Žintake valve train total.

570 MPa 430 W

500 W

The trends in the usage of types of valve trains on a global scale are interesting and an effective way of inferring design problems and potential solutions. In Europe, 20 years ago, the use of direct acting cams and followers in medium class automobile petrol engines was not extensive. With increasing specific power outputs, smaller and faster engines and the tribological problems associated with pivoted follower valve trains, there has been a strong move to the use of direct acting followers. It is also clear that roller follower valve trains are now receiving much closer attention in Europe. In the USA and Japan, the use of roller follower valve trains has been much more extensive, though there is a discernible trend towards an increasing use of direct acting followers in North America. The satisfactory lubrication of the cam and follower contact in IC engines has proved to be the most difficult of all the tribological components. To this day, extensive problems persist for many manufacturers, both in the mass and higher quality markets. Attention is drawn below to a range of issues which will challenge the cam and follower designer in the coming years and lead to improved durability and efficacy. Tribological design challenges for the automotive cam and follower v Improved surface profile, surface roughness and mixed lubrication considerations. v Development of a linkage between lubrication mechanics and chemical mechanics, with a better understanding of the role of additives in reaction films. v Consolidation of the developments in the understanding of lubricant rheology to make more effective design prognostications. v Wear modelling Že.g., Ref. w22x. linking to failure, materials, lubrication and thermal considerations. v Satisfactory provision of lubricant.

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2.3. Piston assembly For over 25 years, tribological studies of the piston assembly, particularly the piston ringrcylinder liner contact, have dominated the interest of researchers in the fields of lubrication, friction and wear as linked to IC engine components. The learned society literature abounds with publications and the reader is referred to two Proceedings of the Leeds-Lyon Symposium on Tribology for initial background reading w24,25x. Studies of the hydrodynamic lubrication of a single, flooded ring, critically dependent on the ring face profile, have been expanded to embrace both compression and oil control rings in a pack, with consideration of lubricant continuity and ring starvation. In the vicinity of the dead centres, where the entraining velocity of the lubricant into the ringrliner interface is zero, consideration of elastohydrodynamic lubrication and squeeze film effects have been undertaken and the importance of boundary lubrication in preserving satisfactory performance has become evident. Considerations of piston ring friction were originally based upon viscous shear calculations but have now been extended to incorporate mixed lubrication conditions. The use of statistical models incorporating surface topography effects has rendered encouraging agreement with regard to power loss predictions and markedly influenced the views on lubricant flow effects. There are increasing studies of the variety of lubricant transport mechanisms with the aim of understanding oil consumption and emission effects. The range of factors which can influence the tribological behaviour of the piston assembly is alarmingly extensive and whilst some have received research attention much remains to be done. The following list will serve to indicate the scope of the problem: Ø Ring dynamics including torsional effects Ø Gas pressure variations Ø Ringrgroove clearance Ø Groove flank profile Ø Ring flankrgroove interaction Ø Piston motion Ø Elastic and thermal deformation effects Ø Non-axisymmetric considerations including non-circular bores Ø Ring and cylinder liner wear Ø Surface topography influences Ø Ring and piston mechanical design Ø Integration of piston and ring dynamics Ø Materials consideration—piston and ring materials, coatings and treatments Ø Lubricant properties and lubricant degradation Ø Lubricant fuel dilution Ø Crevice volume Ø Cavitation considerations. The extent of issues outlined above is not complete but it does indicate that there is still a long way to go towards achieving optimum designs. Of course, the operating tribo-

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logical conditions of a piston ring can vary widely with a host of influential parameters. However, the details below will once again give an engineering feel for the scale of important factors. On this occasion, the data are representative of modern, 2-l, four-cylinder petrol engine with fuel injection, four valvesrcylinder and a double overhead camshaft w26x. Conditions associated with the piston ring and cylinder liner contact of a modern IC petrol engine Engine speed 2500 rpm; maximum cylinder pressure 4 MPa; ISO VG 100 lubricant; oil inlet and water outlet temperature 808C; BMEP 0.5 MPa Piston land temperature above ring 1 2008C Piston land temperature below ring 3 1908C Žoil control ring. TDC Žring 1. mean cylinder liner 1258C temperature Minimum oil film thickness ŽTDC ring 1. 0.9 mm Minimum oil film thickness Žring 3. 0.5 mm Mid-stroke oil film thickness Žring 1. 1.6 mm Ring 1 friction Žrun in. 0.15 kW Ring pack friction Žrun in. 0.49 kW Some particularly interesting developments surrounding piston assembly analysis include consideration of hydrocarbon emission effects Že.g., Ref. w27x. and the linkage for the first time of piston ring lubrication and wear models w26x. The latter aspect will be touched upon later in this paper but the following will serve to identify a range of important potential design and analytical considerations for piston assemblies in the future. Future piston assembly research and design challenges v Enhanced ring and liner wear modelling linked to lubrication model and engine life history predictions v Integrated piston and piston ring analyses v Improved materialsrcoating considerations linked to durability and failure v Improved understanding of lubricant transport mechanisms v Influence of surface topography and mixed lubrication v Three dimensional analysis improvements, including non-circular borerpiston considerations v Lubricant characteristicsrdegradation and their influences v Lubricant chemistryrreaction film studies

3.1. Lubricant modelling A recent paper w28x has described practical applications of lubricant modelling to IC engines. The adoption of wear models linked to lubrication conditions has already been touched upon in the present paper and Coy referred to such studies at the Shell Thornton Research Centre, UK in relation to camrfollower and piston ringrcylinder liner contacts. He also commented upon engine bearing wear, considerations of which will be an important area for development in the not too distant future. In particular, though, his paper drew attention to the importance of improved modelling of the lubricant. Typically, lubrication analyses have taken dynamic viscosity Žas a function of operating temperature. to be the significant lubricant characteristic. For elastohydrodynamic conditions, the effect of pressure on viscosity has also been considered. Coy identified that we are now in a position to be more extensive in incorporating a range of lubricant characteristic behaviours in analyses. Such considerations have already begun in relation to the prediction of the performance of the individual engine tribological components, and upon engine friction and durability in particular. The improved incorporation of lubricant behaviour is an important area for automotive engine tribology and the box below identifies current and prospective rheological considerations. Lubricant rheological analytical considerations in relation to automotive engine tribology v Temperature on dynamic viscosity v Pressure on dynamic viscosity v Shear rate on dynamic viscosity a v Viscoelastic effects for engine bearingsa a With relaxation time effects as influenced by molecular structure, temperature- and pressure-dependent viscosity and density

3.2. Surface topography In its broadest context, the ability to influence the behaviour of a lubricated contact by careful consideration of surface topography influences, and thus achieve a desired optimum performance, is a vibrant area of research at the current time. By surface topography influences, the potential effect of both profile and surface finish are implied. The area is a vast one upon which to comment and it is only possible to draw out a few examples with regard to the frictional elements of the internal combustion engine.

3. Other considerations 3.3. Engine bearings It is proposed to touch briefly upon two specific aspects of engine tribology which merit particular focus at the current time—lubricant modelling and surface topography.

The remarkably thin films which are predicted to occur in automobile engine bearings are suggesting that mixed


lubrication conditions may be encountered. This is, however, still a matter for contention and overall behaviour is clearly affected by thermal and elastic distortions. There is better evidence that the machining process for a big endbearing crankpin, creating a lobed shaft, can be significant in leading to premature failure. Mehenny et al. w29x have considered this issue. The analytical problem is one of circumferential lobing on a shaft with different amplitudes and frequencies possible. Prospectively, the reduction in lubricant film thickness may accelerate wear, whilst increases in the amplitude and frequency of the fluctuations in pressure can promote fatigue of the bearing surface. In Fig. 5, the effect of lobing on the pressure distribution with a nine lobed shaft is shown. In this sketch presentation, the load has rotated through 48 and the unusual form of the pressure distributions will give rise to minimum film thickness and pressure changes which can be significant in wear and failure. Some modest correlation with experimental data has been achieved for shafts with different numbers of lobes and different amplitudes. 3.4. ValÕe train Wear modelling of cams and followers has been mentioned previously w22x. This has led to some impressive agreements between measured and predicted profile wear. However, little has been reported in the literature relating to the influence of surface finish of cam and follower components on wear. Roylance et al. w23x have presented details of the use of surface profilometry to measure wear during running in on pivoted followers, with surface finish and hardness varied. Harrison w30x and Zhu w31x used a laboratory test apparatus to examine a direct acting mechanism. The cams were steel, induction hardened to 2.5 mm depth, stress relieved, ground and phosphated. In the first series of tests, cams with different initial surface roughness were run against nominally similar followers under the same conditions. Average Ž R a . surface roughness changes are recorded in Fig. 6. Each test was of 100 h at a fixed camshaft speed of 1200 rpm and apparatus temperature 508C.

Fig. 5. Effect of shaft lobing Žhere nine lobes. on mean hydrodynamic pressure in an automotive bearing.

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Fig. 6. Average surface roughness values on the cams of a direct acting mechanism.

It can be seen that on the cam flanks, there was little change in roughness, reflecting healthy oil film thickness at these sites. The cam and follower were mechanically separated over the base circle and hence, there was no change in this region. Over the base region, there was evidence on the rougher cams of an improvement in surface finish reflecting removal of asperity tips and surface flattening. It will be noted that the smoother cam roughened up, and there was a tendency for the final roughness at the nose to move towards the same value irrespective of initial roughness. It is worthy of note that in another series of tests, the temperature of the bulk apparatus was varied up to 1208C, but this did not appear to be an influential factor. 3.5. Piston assembly The influence of the surface topography of piston rings and cylinder liners upon their behaviour has been known for many decades. Running-in of the engine, with an appreciable reduction in friction due to ring profile changes, has been recognised for some time. The importance of the

Fig. 7. Predicted ring profile after 120 h as compared with the measured profiles.

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honing process in finishing cylinder liners in order to promote satisfactory lubrication and avoid scuffing has been recognised, if not understood. Past research on piston ringrcylinder liner tribological performance has concentrated entirely on the lubrication of the contact. Surface topography effects have been incorporated into lubrication analysis on the basis of statistical influences, using flow and friction factors associated with the lubrication equations. However, there is currently a proposition that a sound appreciation of the tribological performance of piston rings in an IC engine can only be achieved by combining lubrication and wear analyses. The sensitivity of ring performance to its profile has been shown in previous studies and the ability to predict wear of the rings and integrate this into the lubrication analysis has been recently achieved by Priest w26x. He used a simple wear model, analogous to that adopted for camsrfollowers by Colgan ad Bell w22x. Fig. 7 shows a comparison of measured and predicted results for the top compression ring of a Caterpillar IY73 single cylinder diesel engine. The predicted profile after 120 h of running compares well with the measure profile, both in shape and quantity of material worn away. The latter has been determined in two ways—the ‘geometric’ profile has been achieved by overlaying measured traces such that unworn portions correlated well, whilst the ‘mass’ profile was determined by measuring ring weight loss and assuming this was achieved by equal wear around the ring circumference. The more detailed research data presented by Priest is an encouraging step forward.

4. Conclusions Ža. The importance of the tribological design of the major frictional components of the automotive internal combustion engine has been emphasised from the point of view of efficiency, durability and emissions. Žb. The current operating conditions of engine bearings, amsrfollowers and piston ringsrcylinder liners have been reviewed and the state of the art of analytical design procedures explored. Challenges for the future in the design procedures for each machine element have been identified. Žc. Particular attention has been drawn to the importance of improved modelling of lubricant behaviour and to the enormous prospect for optimised design based on surface profiles and surface finish considerations. Žd. The field of automotive engine tribology represents a rich and varied environment for researcher and practi-

tioner, with potentially significant influence upon wealth creation and quality of life.

References w1x HMSO Transport Statistics Great Britain 2 Ž1995. 57-58. w2x B.S. Andersson, Vehicle tribology, 17th Leeds-Lyon Symposium on Tribology, Elsevier Tribology Series 18 Ž1991. 503–506. w3x J.F. Booker, Trans. ASME J. Basic Eng. 187 ŽD. Ž1965. 537–546. w4x J.F. Booker, Trans. ASME J. Lub. Tech. 91 ŽF. Ž1969. 534–543. w5x C.M. Taylor, ŽEd.., Engine tribology, Elsevier Series 26 Ž1993. 301 pp. w6x C.K.K. Lai, PhD Thesis, Univ Leeds UK, 1993, 285 pp. w7x D.S. Mehenny, PhD Thesis, Univ. Leeds UK, 1997, 269 pp. w8x B. Fantino, M. Godet, J. Frene, Studies of Engine Bearings and Lubrication SAE 539 Ž1983. 23–32. w9x B. Fantino, J. Frene, Trans. ASME J. Lub. Tech., 505–515. w10x H.H. Priebsch, J. Krasser, Elastohydrodynamics ’96, 23rd LeedsLyon Symposium on Tribology, Elsevier Tribology Series 32 Ž1996. 651–659. w11x H. Xu, Elstohydrodynamics ’96, 23rd Leeds-Lyon Symposium on Tribology, Elsevier Tribology Series 32 Ž1996. 641–650. w12x H. Xu, E.H. Smith, Proc. IMechE 204 Ž1990. 187–196, Part C. w13x G.J. Jones, Tribology of reciprocating engines, 9th Leeds-Lyon Symposium on Tribology, 1983, pp. 88–98. w14x D. Han, PhD Thesis, Univ. Leeds, UK, 1993, 240 pp. w15x D. Han, C.M. Taylor, Abstracts World Tribology Conference, London, 1997, p. 91. w16x J.M. Conway-Jones, R. Gojon, Vehicle tribology, 17th Leeds-Lyon Symposium on Tribology, Elsevier Tribology Series 18 Ž1991. 33–42. w17x A.O. Mian, G.J. Jones, The third body concept, 22nd Leeds-Lyon Symposium on Tribology, Elsevier Tribology Series 31 Ž1996. 291–298. w18x A.D. Ball et al., Tribological design of machine elements, 15th Leeds-Lyon Symposium on Tribology, Elsevier Tribology Series 14 Ž1989. 111–130. w19x C.M. Taylor, Vehicle tribology, 17th Leeds-Lyon Symposium on Tribology, Elsevier Tribology Series 18 Ž1991. 119–131. w20x C.M. Taylor, Proc. IMechE 208 Ž1994. 221–234, Part J. w21x L.S. Yang, PhD Thesis, Univ. Leeds, UK, 1992, 230 pp. w22x T. Colgan, J.C. Bell, Int. Trib. Conf., Nagoya, Japan 2 Ž1990. 1231–1235. w23x B.J. Roylance et al., Vehicle tribology, 17th Leeds-Lyon Symposium on Tribology, Elsevier Tribology Series 18 Ž1991. 143–143. w24x Tribology of reciprocating engines, 9th Leeds-Lyon Symposium on Tribology, 1983, 348 pp. w25x Vehicle tribology, 17th Leeds-Lyon Symposium on Tribology, Elsevier Tribology Series 18 Ž1991. 526 pp. w26x M. Priest, PhD Thesis, Univ. Leeds, 1996, 246 pp. w27x A. Thomson, C.D. Radcliffe, C.M. Tayor, Proc. IMechE 211 Ž1997. 223–233, Part J. w28x R.C. Coy, New directions in tribology, World Tribology Conference, London, MEP, 1997, pp. 197–209. w29x D.S. Mehenny, C.M. Taylor, G.J. Jones, H. Xu, SAE Technical Series Paper 970216 Ž1997. 9 pp. w30x P. Harrison, PhD Thesis, Univ. Leeds, UK, 1988, 281 pp. w31x G. Zhu, PhD Thesis, Univ. Leeds, 1988, 236 pp.

Automotive Engine Tribology

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