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Predictivewearmodellingoflubricatedpistonringsinadieselengine

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Ž .Wear 231 1999 89–101

Predictive wear modelling of lubricated piston rings in a diesel engine

M. Priest ), D. Dowson, C.M. TaylorSchool of Mechanical Engineering, The UniÕersity of Leeds, Leeds, LS2 9JT, UK

Received 15 March 1999; accepted 18 March 1999

Abstract

The tribological performance of piston rings in reciprocating internal combustion engines can only be fully understood when bothlubrication and wear are considered in combination. To this end, a numerical model has been developed that predicts the dynamics,lubrication and wear of piston rings interactively for the first time. This paper reports the application of this new model to the piston ringpack of a diesel engine. With the overall aim of evaluating the correlation between theory and experiments, this analysis is divided intotwo discrete parts. First, the model is used to predict the lubrication performance of measured ring packs before and after periods ofrunning, at constant speed and load, in a Caterpillar 1Y73 single-cylinder diesel engine: the objective being to establish the change intribological behaviour with observed wear in the engine. Secondly, the model is used interactively to predict the lubrication and wear ofthe top compression ring from the same engine. This research advances the understanding of piston ring profile evolution with time andits dependence on complex interactions between lubrication and wear. q 1999 Elsevier Science S.A. All rights reserved.

Keywords: Piston rings; Lubrication; Wear; Diesel engine

1. Introduction

The tribological behaviour of piston rings has long beenrecognised as an important influence on the performanceof internal combustion engines in terms of power loss, fuelconsumption, oil consumption, blow-by and harmful ex-haust emissions.

The primary role of the piston ring pack is to maintainan effective gas seal between the combustion chamber andthe crankcase. The rings of the piston ring pack, whichtogether effectively form a labyrinth seal, achieve this byclosely conforming to their grooves in the piston and to thecylinder wall. The small quantity of gas that does find itsway into the crankcase, blow-by, is normally piped back tothe inlet valve and fed back into the cylinder.

In addition to causing a dramatic increase in pressure,the combustion event generates a large amount of heat.Much of this thermal energy is convected into the pistoncausing a marked increase in the temperature of the piston,which is dissipated by heat transfer to adjacent compo-nents and the engine coolant. The secondary role of thepiston ring pack is to transfer this heat from the piston intothe cylinder wall and thence into the coolant.

) Corresponding author. Tel.: q44-113-2332178; fax: q44-113-2332150; E-mail: m.priest@leeds.ac.uk

The final function of the piston ring pack is to limit theamount of oil that is transported from the crankcase to thecombustion chamber. This flow path is probably the largestcontributor to the oil consumption of an engine and leadsto an increase in harmful exhaust emissions as the oilmixes and reacts with the other contents of the combustion

w xchamber 1 . The desire to extend service intervals ofengines and minimise harmful exhaust emissions to meetever more stringent legislative requirements, means thatthe permissible oil-consumption levels of modern enginesare very low compared to their predecessors of 10 or 20

w xyears ago 2 .The piston ring pack must fulfil these three roles with a

minimum of frictional power loss, most notably at thesliding interface with the cylinder wall, and a minimum ofwear in order to maximise component life. Unfortunately,the piston ring pack is one of the largest sources of frictionin the internal combustion engine over the normal range of

w xengine speeds and loads encountered in service 3–5 .Exact figures vary from engine to engine, but typically thepiston assembly, comprising both the piston rings and thepiston skirt, accounts for 40–50% of total engine frictionw x1 . Piston ring pack friction losses are greater than thoseof the piston skirt at low to moderate engine speeds but thesituation may be reversed at high engine speeds due to thelarge wetted area of the piston skirt contributing to viscous

0043-1648r99r$ - see front matter q 1999 Elsevier Science S.A. All rights reserved.Ž .PII: S0043-1648 99 00125-8

( )M. Priest et al.rWear 231 1999 89–10190

Fig. 1. Caterpillar 1Y73 piston and ring pack.

w xfriction 6 . In terms of wear, there is insufficient under-standing of the interaction with the lubrication process. So,even though manufacturers can produce rings that have anexcellent life expectancy, these components may be farfrom optimum from a lubrication and friction standpoint.

As a consequence of their importance to engine perfor-mance, the theoretical and experimental study of piston-ringlubrication has received much attention in the literature,leading to major advances in our understanding of their

w xbehaviour 7 . The mathematical analysis of piston-ringlubrication is complex and by necessity requires simplify-ing assumptions. However, rapid development in numeri-cal methods over the last 25 years has resulted in sophisti-cated piston-ring lubrication models that are finding appli-

w xcation in the design process 8 .It is the proposition of the research presented in this

paper, however, that a complete understanding of thetribological performance of piston rings in reciprocatinginternal combustion engines can only be achieved whenboth lubrication and wear are considered in combination.The running profile of the piston ring that slides againstthe cylinder wall wears significantly in service, even withwear-resistant materials and coatings, such that the ringprofile after only a short period of running in the engine

w xdiffers greatly from that of the component as new 9–14 .Modification of the ring profile by wear has a large effecton lubrication, friction and oil transport at the interfacebetween the piston ring and cylinder wall which then inturn modifies the wear conditions. This interaction be-tween lubrication and wear has important implications forthe performance and life of the internal combustion engine.

The piston ring is perhaps the most complicated tribo-logical component in the internal combustion engine. It issubjected to large, rapid variations of load, speed, tempera-

ture and lubricant availability. In one single stroke of thepiston, the piston ring may experience boundary, mixed

w xand full fluid film lubrication 7 . Elastohydrodynamiclubrication of piston rings is also possible in both gasolineand diesel engines on the highly loaded expansion stroke

w xafter firing 15 .Incorporating a consideration of wear into the dynamics

and lubrication analysis of piston rings adds a further layerof complexity to the model. This is compounded by thefact that wear is the least understood of the three mainprocesses in tribology: friction, lubrication and wear.

A piston ring tribology model incorporating predictionof the change in ring face profile with wear in the engine

w xhas recently been developed 16,17 . It assumes that wearof the ring profile may be described by the Archard wear

w xequation, in the form proposed by Lancaster 18 :

VskWx 1Ž .s

The wear factor, k, is a function of the interactingmaterials, their surface topography, the lubricant and theoperating conditions. This can alternatively be expressedas a variation of wear factor with specific film thicknessrelative to the wear factor in the boundary lubricationregime, k . The wear factor in the boundary lubrication0

regime is determined from bench test rig experimentsusing actual components and lubricant at operating condi-tions of load, speed and temperature indicative of bound-ary lubrication. This empirical input to the model clearlyexposes our lack of fundamental understanding of the wearprocesses taking place in such tribological interfaces. Thisapproach, however, has been applied successfully in auto-

w xmotive valve train wear modelling 19,20 .With this relationship and the cyclic variation of mini-

mum film thickness predicted by the lubrication model, thewear factor can be determined at each crank angle in theengine cycle. Thus it is possible to predict, interactively,the changes in wear and lubrication of the piston ring that

Fig. 2. Piston ring measurement on the Form Talysurf profilometer.

( )M. Priest et al.rWear 231 1999 89–101 91

take place with running time in the engine. Full details ofw xthe analysis may be found in Refs. 16,17 .

This paper considers the application of this new interac-tive dynamics, lubrication and wear model to the pistonring pack of a Caterpillar 1Y73 single-cylinder dieselengine. After describing piston-ring and cylinder-wall wearexperiments undertaken on the engine, the theoretical anal-ysis is presented as two discrete parts. In the first, the

model is used to predict the lubrication performance ofmeasured ring packs before and after periods of running inthe engine, the objective being to establish the change intribological behaviour with observed wear in the engine.Secondly, the model is used interactively to predict thelubrication and wear of the top compression ring from thesame engine, with the objective of evaluating the correla-tion between the new model and experiments.

Fig. 3. Variation of measured ring profiles with time.

( )M. Priest et al.rWear 231 1999 89–10192

2. Experimental results

Fig. 1 shows the geometry of the piston ring pack andthe piston from the Caterpillar 1Y73, single-cylinder, dieselengine. The piston ring pack consists of three compressionrings and an oil-control ring mounted on a cast-iron pistonwith a steel insert forming the top ring groove. The topcompression ring, Ring 1, is a barrel-faced, chromium-plated, cast-iron ring and the other two compression rings,Rings 2 and 3, are of the same design and are taper-faced,plain cast-iron, scraper rings with an internal step. Thisstep gives them a slight upward twist when fitted, givingthem a dished appearance, such that they offer their loweredge to the cylinder wall. The oil-control ring, Ring 4, is atwin-land, single-piece design with interland drainage slotsand a coil spring expander, the lands of which arechromium plated. The cylinder is a wet liner design and ismanufactured from induction hardened cast iron.

Further physical data for the ring pack is given forreference in Appendix A. The engine was run at acrankshaft speed of 1200 rpm and a nominal brake load of

Ž .either 10 or 14 bar brake mean effective pressure BMEP .Basic engine data and operating parameters relevant to thepiston assembly for these two conditions are also given inAppendix A.

Extended duration tests, at constant speed and load,were undertaken to study the changes occurring in thepiston ring profiles and the surface roughness of the ringsand the cylinder wall. Piston ring profiles and surfaceroughness data were measured using a Rank Taylor Hob-son Form Talysurf profilometer with a laser-referencedstylus rather than the more traditional inductive system.

The arrangement used to measure the Caterpillar 1Y73piston rings on the Form Talysurf is shown in Fig. 2. A jigwas devised to locate the piston rings, which consisted oftwo vee block supports attached to tee slots in the machinebase with a solid steel cylinder suspended between them.The cylinder had a flat machined along its length to whicha small vice was fixed which held the piston ring and alaboratory slip gauge as shown in Fig. 2. The slip gaugewas used to establish a datum for the ring profiles toenable comparison to be made between different measure-ments. Before each ring profile measurement was taken,the relative heights of the vee block supports were adjusteduntil a 2.0-mm traverse across the slip gauge gave amaximum vertical displacement of the stylus of less than1.0 mm, which was equivalent to a maximum slip gaugesurface inclination of 0.0288. As the slip gauge wasclamped against the ring flank, this gave a reference for allthe measurements, assuming that the ring flanks did notwear greatly in service.

Ring profiles were measured at three circumferentiallocations remote from the ring gap using a stylus with a2-mm tip radius. It was decided at this stage in the researchnot to measure ring profiles near the ring gap, whereunknown dynamic effects in service may lead to a wear

pattern that could not confidently be reproduced usinganalytical wear models. A simple template was used tomark the ring flanks prior to measurement.

Fig. 3 shows the measured ring profiles at a circumfer-ential position 908 offset from the ring gap after 0, 120 and628 h running.

The rings at 0 at 120 h are the same rings measuredbefore and after running at a constant speed and load of1200 rpm and 14 bar BMEP. The data reported for 628 h,however, is for a completely different ring pack that wasrun for half the test at 1200 rpm and 10 bar BMEP andhalf at 1200 rpm and 14 bar BMEP. Although not a truecomparison with the results at 0 and 120 h, the measure-ments at 628 h show an entirely consistent trend and are auseful indication of the long-term wear of the rings in thisengine.

Wear of the piston rings at the other two circumferentialpositions was similar for the compression rings, which isnot surprising given that all three measurement locationsare remote from the ring gap and the rings are free torotate around the piston. The oil-control ring exhibitedsome circumferential variation but this was small com-pared to the resolution of the measurement technique. Full

w xdetails can be found in Ref. 16 .The variation of root mean square surface roughness of

the components, in the direction of sliding, throughout thetest is given in Table 1. The surface roughness wasdetermined using the software provided with the FormTalysurf which simulates the standard electronic filteringtechniques developed for inductive stylus profilometersw x21 . The data for the piston rings is for the full profile at 0h and for the worn regions of the profile at 120 and 628 h.The surface roughness values for the cylinder wall are themean of a series of measurements taken at top dead centre,midstroke and bottom dead centre. This approach wasdictated by the use of these parameters in the mathematicalmodel.

The root mean square surface roughness of the cylinder,measured using the Form Talysurf with standard electronicfiltering, is strongly affected by the deep honing marksintroduced into the surface during manufacture. This pa-rameter is used in the mixed lubrication model of the ringpack, in which the deep honing marks play no part, and it

Table 1Variation of axial surface roughness parameters with time

Ž .Component Time h

0 120 628

Ž .s mmRing 1 0.112 0.008 0.078Ring 2 0.227 0.029 0.023Ring 3 0.227 0.016 0.012

Ž .Ring 4 upper land 0.033 0.007 0.017Ž .Ring 4 lower land 0.039 0.039 0.037

Cylinder 0.476 0.139 0.180

( )M. Priest et al.rWear 231 1999 89–101 93

was therefore considered unrepresentative to input themeasured values including the honing marks into the model.A parametric study was undertaken of the surface rough-ness of small sections of the cylinder wall not containinghoning marks, manually chained together to give sufficientdata for statistically significant results. This suggested thatnew cylinder roughness values should be reduced by afactor of 0.50 and worn values by a factor of 0.32 before

input to the model. These scaling factors are included inthe data reported in Table 1.

3. Lubrication predictions for measured piston ringprofiles

Fig. 4 shows the variation of minimum lubricant filmthickness throughout the engine cycle for the new ring

Fig. 4. Predicted film thickness variation at 0 h.

( )M. Priest et al.rWear 231 1999 89–10194

pack in both fully flooded and starved lubrication condi-tions, 08 crank angle being top dead centre firing. The fullyflooded model of lubricant flow in a ring pack assumesthat there is an unlimited supply of lubricant available toeach ring at all stages in the engine cycle such that theinlet region of the ring profile is always full of lubricant.In reality, however, the quantity of lubricant available toeach ring is the thin film smeared on the cylinder wall by

the preceding ring and consequently the inlet region of thering profile may be starved of lubricant. This more sophis-

w xticated approach is termed starved lubrication 7 .Included on each graph in Fig. 4 is the upper limit of

surface contact, and hence wear, which is equal to fourtimes the composite root mean square roughness of thepiston ring and cylinder wall, 4s . This defines the transi-tion between the mixed and full fluid film lubrication

Fig. 5. Predicted film thickness variation at 120 h.

( )M. Priest et al.rWear 231 1999 89–101 95

regimes. Hence, if the film thickness is less than 4s ,surface contact and wear will occur. The transition be-tween the mixed and boundary lubrication regimes is equalto half the composite surface roughness, 0.5s . The lubri-cation model uses this latter value as the minimum permit-ted film thickness and it can be observed as a distinctminimum limit in Fig. 4. The upper limit of surface

contact for the oil-control ring was computed using thelargest of the two land roughness values as the minimumfilm thickness can occur on either land depending on theattitude of the ring to the cylinder wall.

In fully flooded conditions, all four rings experiencefull fluid film, mixed and boundary lubrication. The verythin film generated on the downstrokes by Ring 4, the

Fig. 6. Predicted film thickness variation at 628 h.

( )M. Priest et al.rWear 231 1999 89–10196

Table 2Predicted friction power loss

Ž . Ž .Time h Friction power loss W

Fully flooded lubrication Starved lubrication

Ring 1 Ring 2 Ring 3 Ring 4 Ring 1 Ring 2 Ring 3

0 162.7 138.1 132.0 135.4 356.8 147.1 143.8120 126.9 91.0 72.8 75.9 417.2 173.3 170.6628 119.5 87.6 73.3 46.0 270.2 160.0 158.6

oil-control ring, is of great concern as this ring controls thesupply of lubricant to the compression rings under starvedlubrication conditions. Extreme starvation of the compres-sion rings by the oil-control ring is indicated in the starvedresults of Fig. 4. All the compression rings have muchreduced film thicknesses compared to their fully floodedvalues, with no ring achieving a full fluid film at any stagein the engine cycle. Wear would occur at all stages in theengine cycle under these conditions.

A step change can be seen in film thickness in some ofthe curves of Fig. 4, e.g., at 4708 for Ring 1 and 5008 forRings 2 and 3. The occurs when the ring moves axiallyfrom the top to the bottom of the piston groove, or viceversa, resulting in a change in radial gas load, a phe-nomenon often referred to as ring lift.

Fig. 5 shows the film thickness parameters for the samering pack after 120 h running in the engine. First, it shouldbe noted that the lubrication regime transition limits havesmaller absolute values due to reductions in compositesurface roughness of the rings and the cylinder as reportedin Table 1. Looking first at the fully flooded results it canbe seen that all the compression rings, Ring 1 to 3, developa full fluid film for much more of the engine. In particular,Ring 1 no longer has the sustained period of damagingboundary lubrication from 7038 to 568 crank angle ob-served with the new ring in Fig. 4. Ring 4 exhibits verymodest hydrodynamic film action on all strokes but with afull fluid film only achieved on the downstrokes. Overall,the performance of this ring has apparently deterioratedfrom the new ring performance given in Fig. 4. Thisconsequently starves the compression rings of lubricanteven more severely than at 0 h with very low levels ofhydrodynamic film generation for all three compressionrings under starved lubrication conditions.

The results for 628 h running in the engine are given inFig. 6. Modest improvements in film thickness are observ-able for the compression rings in fully flooded conditions.Ring 4 generates full fluid films on the downstrokes with aboundary film on the upstrokes. This is the completeopposite to the behaviour of the new ring and very differ-ent to the performance after 120 h. The existence of a fullfluid film under the oil-control ring on the downstrokeslessens the severity of lubricant starvation of the compres-sion rings. The performance of all three compression ringsunder starved conditions has consequently improved

slightly. Ring 3 has improved the most such that on thedownstrokes it is predicted to have a full fluid film in themidstroke region.

Friction power loss predictions for the complete ringpack at 0, 120 and 628 h running in the engine aresummarised in Table 2. Looking first at the fully floodedresults, the compression rings show a pattern very muchakin to running in with significant reductions in friction inthe first 120 h and little change thereafter. The oil-controlring, Ring 4, however, has similar reductions at the end ofeach time interval. The starved lubrication results for thecompression rings give a much more confusing picture.The lubricant starvation caused by the oil-control ringapparently masks any friction changes due to profile modi-fication, reductions in surface roughness and interactionbetween the compression rings. The likelihood of thisextreme level of starvation occurring in practice is dis-cussed later in this paper.

4. Piston ring profile wear predictions

To investigate predicted changes in piston ring geome-try and performance due to wear, the piston ring lubrica-

w xtion and wear model 16,17 was applied to the top com-pression ring, Ring 1, of the Caterpillar 1Y73 engine overthe first 120 h of running. The ultimate objective was toachieve good correlation between predicted and measuredring profiles at 120 h.

The ring profile of Ring 1 at 0 and 120 h is shown inFig. 2. As each measurement of ring profile has a differentcoordinate origin, the ring profile measured after 120 hrunning cannot be simply plotted over the new profile toevaluate wear. Efforts were therefore made to overlay theprofiles and Fig. 7 includes the results of two alternativeapproaches. First there is what is referred to as a geometricoverlay where the worn ring profile was moved in x and zuntil the new profile and the apparently unworn regions ofthe worn profile appeared to correlate well. Second, there

Fig. 7. Measured profiles of the new and worn ring.

( )M. Priest et al.rWear 231 1999 89–101 97

is what has been labeled a mass overlay which wasdetermined from the weight loss of the ring over the 120-htest period, 19.9 mg, assuming that equal mass was lost atall locations around the circumference of the ring. Theworn profile was then moved in x and z until the areabetween the new and worn profiles correlated with theweight loss. The obvious difference between the geometricand mass overlays in Fig. 7 suggests that the wear of thering was not equal at all locations around the ring circum-ference. It should be noted that no rotation of the wornprofiles was introduced as part of the overlay procedures.

The first stage in the solution process was to input theinitial topography and operating conditions to the lubrica-tion model to determine the cyclic variation of lubricantfilm thickness, applied radial load and axial velocity whichare required by the wear model to determine wear rates. Asthis is a single-ring analysis and as such no lubricationpredictions are undertaken for the other rings in the pack,fully flooded lubrication conditions were assumed through-out. Also, the lubricant separation boundary conditions for

w xReynolds equation developed by Coyne and Elrod 22were adopted.

The lubrication model used in this analysis did notdetermine the cyclic variation of torsional twist angle ofthe piston ring, which is also needed by the wear model,and so an assumed ring twist pattern was required. Fig. 8shows the predicted cyclic variation of twist angle for thetop compression ring of a 203-mm bore, single-cylinder

w xdiesel engine from the published literature 23 .Based on this predicted behaviour, a sinusoidal cyclic

ring twist variation with an amplitude of 0.18, a period of2408 crank angle and zero offset, 08 ring twist at 08 crankangle, was assumed for this investigation. This was chosento give best agreement with the data in Fig. 8 during thefirst half of the power stroke when the wear rates wereexpected to be highest. A graphical representation of theassumed ring twist characteristic is given in Fig. 9.

w xFig. 8. Predicted ring twist angles from the literature 23 .

Fig. 9. Assumed ring twist variation.

Another important input parameter to the wear model isthe wear factor under boundary lubrication conditions, k .0

Reciprocating wear tests in boundary lubrication condi-tions were undertaken on samples of the piston ring andcylinder liner from this engine. The tests were run with agood quality diesel engine lubricant in both fresh anddegraded conditions and the results are summarised inTable 3.

After some initial trials, a value of k of 1.5=10y120

mm3 mmy1 Ny1 was adopted as most appropriate to thetop ring of this engine during the first 120 h of running.During this time, the lubricant will go from fresh to adegraded condition due, primarily, to exposure to heat. Thebulk temperature of the lubricant available to the pistonring at top dead centre, where much of the wear appearedto occur according to the initial trials, is assumed to beequal to that of the cylinder wall at this position, whichwas 1408C. It was also assumed that the surface roughnessof the piston ring and cylinder wall reduced linearly withsimulation time from the measured values at 0 h to themeasured values at 120 h as given in Table 1.

The wear model was then used to predict the develop-ment of the piston ring profile over a 2-h period assumingthe lubricant film thickness at any point in the engine cycleremained constant during this time. The worn piston ringtopography was then fed back to the lubrication model andthe results fed forward into the wear model for another 2-hperiod of simulated profile development. This interactiveprocess was continued until the total desired simulationtime of 120 h was achieved.

The minimum film thickness variation throughout theengine cycle, as predicted by the lubrication model for the

Table 3Piston ring wear factors from lubricated reciprocating wear tests

Bulk Lubricant k03 y1 y1Ž . Ž .temperature 8C condition mm mm N

y1 3200 Fresh 2=10y12300 Fresh 4=10y11300 Degraded 5=10

( )M. Priest et al.rWear 231 1999 89–10198

Fig. 10. Cyclic variation of specific film thickness at 0 h.

initial profile at 0 h, is plotted in Fig. 10 as specific filmthickness, l, where

hminls 2Ž .

s

Also marked on the graph are the transitions betweenboundary and mixed lubrication and mixed and full fluidfilm lubrication. The film-thickness predictions indicatethat at 0 h, wear occurs for the majority of the engine cycleas a full fluid film is only achieved in midstroke regionson the power, intake and compression strokes. Note, theshape of the curve is slightly different to that presented inFig. 4 due to the use of the Coyne and Elrod cavitationmodel rather than the more common Reynolds cavitation

w xand reformation model 22 .The wear-rate variation with crank angle predicted by

the wear model at 0 h is shown in Fig. 11 and indicatesthat, although wear takes place for a significant proportionof the engine cycle, the largest values occur around topdead centre firing where the applied load is greatest. Thewear rate tends to zero at the dead centres because thesliding velocity falls to zero as the piston changes direc-tion.

The predicted profile at 120 h is shown in Fig. 12 alongwith the measured ring profiles. Good agreement is appar-

Fig. 11. Cyclic variation of wear rate at 0 h.

Fig. 12. Predicted ring profile compared with the measured profiles.

ent in both the quantity of material worn away and theshape of the profile.

5. Discussion

5.1. Lubrication predictions with measured piston ringprofiles

The film thickness predictions from the fully floodedlubrication analysis indicate that the mathematical model iscapable of evaluating the expected improvement in perfor-mance of these rings with running time in the engine. In anattempt to quantify the general health of the interfacebetween the piston ring and cylinder wall, Table 4 presentsthe proportion of the engine cycle for which the filmthickness is less than the upper limit of surface contact andhence wear is likely to occur. The compression rings,Rings 1 to 3, all greatly improve their performance be-tween 0 and 120 h and then change little between 120 and628 h indicating the completion of running-in by 120 h assuggested by the friction results in Table 2. The samecannot be said of the oil-control ring, Ring 4, which showsno categorical improvement.

The oil-control ring operating conditions are crucial indetermining the predicted performance of the compressionrings through the starved lubrication analysis. The verylow film thicknesses predicted for the compression rings instarved conditions in Figs. 4–6 and the high levels of wear

Table 4Assessment of the fully flooded results

Ž .Time h Proportion of the engine cycle for which theŽ .minimum film thickness is less than 4s %

Ring 1 Ring 2 Ring 3 Ring 4

0 41 61 62 61120 7 11 9 78628 10 9 8 50

( )M. Priest et al.rWear 231 1999 89–101 99

Table 5Ž .Laser-induced fluorescence LIF film-thickness measurements

Ž . w xStroke Minimum film thickness mm degrees of crank angle after top dead centre

w x w x w x Ž . w x Ž . w xRing 1 1038 Ring 2 938 Ring 3 888 Ring 4 upper land 848 Ring 4 lower land 818

Power 0.55 3.2 5.2 9.3 7.6Exhaust 0.60 1.3 2.2 6.5 10.7Induction 1.2 2.5 5.1 9.1 4.5Compression 1.4 2.5 3.3 5.3 10.5

that would result seem rather unrealistic. It is proposed thatthe oil-control film thickness predictions are too severeand the oil-control ring is therefore excessively starvingthe compression rings of lubricant to a degree that seemsunlikely in practice. A possible cause of these severepredictions could be inaccuracies in the assumed oil-con-trol ring profile. In measuring the profile of the oil-controlring lands, in terms of inclination and relative heightdifference, we have attempted to resolve to within 1 mmover a Form Talysurf traverse of 6 mm. This correspondsto a slope of 0.00958, which is less than the tolerance of0.0288 allowed on setting the datum of the measurementsystem, as discussed previously.

To investigate this further, Table 5 shows laser-inducedŽ .fluorescence LIF film-thickness measurements under the

piston rings of this engine from a single transducer mountedin the cylinder liner near midstroke. Details of the applica-tion of this technique to the Caterpillar 1Y73 engine can

w xbe found in Ref. 24 .The measured film thicknesses for the oil-control ring

lands are large when compared to the theoretical predictionof oil-control ring minimum film thickness, which is out-put as the smallest of the two land film-thickness values atany instant in time, in Figs. 4–6. Only on the downstrokesat 628 h, Fig. 6, do the predictions come close to themeasurements with a predicted maximum film thickness ofapproximately 3.2 mm. This adds considerable weight tothe conclusion that the mathematical model is predictingfilm-thickness levels that are far too low for the oil-controlring.

As a direct consequence of this argument, it is likelythat the true running condition of the compression ringslies somewhere between the fully flooded and starvedpredictions of the analysis. Comparison of the measuredfilm thicknesses in Table 5 and film-thickness predictionsafter 120 h running in Fig. 5, probably the most compara-ble conditions to those of the measurements, supports thisview.

5.2. Piston-ring profile wear predictions

The prediction of the wear for the top compression ringfrom the Caterpillar 1Y73 diesel engine over the first 120h of running, reported in this paper, is the first applicationof the piston-ring dynamics, lubrication and wear model

w xdeveloped in Refs. 16,17 to fired engine test data. Assuch, the results are most encouraging with good agree-ment between measurements and theory in terms of theamount of wear taking place and the shape of the worn-ringprofile, as shown in Fig. 12. The most important aspects ofthe piston ring profile, in terms of lubrication performance,are its curvature and the axial position of the peak of theprofile. In both respects, the measured and predicted pro-files show a high degree of correlation.

As noted previously, incorporating a consideration ofwear into the dynamics and lubrication analysis of pistonrings adds a further layer of complexity to the model.Also, because of our basic lack of fundamental understand-ing of the wear process, it necessitates a degree of empiri-cism. Hence the need to input a wear factor for theboundary lubrication regime to the model based on benchtest rig results such as those reported in Table 3. Closerconsideration of the initial trials of the model, undertakento determine the actual value of wear factor to be inputfrom the range reported in Table 3, reveals a self-calibra-tion mechanism inherent in the model. More precisely,given the worn ring profile and a detailed knowledge ofthe operating history of the piston ring, it should bepossible to works backwards to determine the wear factorin the boundary lubrication regime. This process, naturally,depends upon the confidence that can be placed in theaccuracy of the model. It is envisaged, that further valida-tion of the model using data from fired engine tests andbench wear test rigs, will reduce the reliance of the modelon empirical input.

The assumption of fully flooded lubrication conditionsin these predictions is clearly an oversimplification butwas necessary as only one ring was analysed in this firstapplication of the model. However, the choice of wearfactor ameliorated the effect on the results.

Table 6Piston-ring data for the Caterpillar 1Y73

Parameter Ring number

1 2 3 4

Ž .Mass g 47.9 35.1 35.0 42.4Ž .Elastic modulus GPa 117 117 117 117Ž .Fitted ring gap mm 0.66 0.61 0.61 0.56

Ž .Tangential ring tension N 54 31 32 40

( )M. Priest et al.rWear 231 1999 89–101100

Table 7Basic engine data and operating conditions for the Caterpillar 1Y73

Parameter Operating conditions

1200 rpm, 10 bar 1200 rpm, 14 bar

Ž .Cylinder bore diameter mm 130.2Ž .Piston stroke mm 165.1

Ž . w xConnecting rod length mm between bearing centres 273.1Ž .Gudgeon pinrcrank axis offset mm 0

Ž .Lubricant dynamic viscosity at 408C mPa s 86.04Ž .Lubricant dynamic viscosity at 1308C mPa s 5.14

Ž .Engine speed rpm 1200Ž .Piston land temperature above ring 1 8C 244 270Ž .Piston land temperature below ring 1 8C 198 218Ž .Piston land temperature below ring 2 8C 184 200Ž .Piston land temperature below ring 3 8C 165 170

Ž .Cylinder wall temperature for ring 1 at top dead centre 8C 130 140Ž .Cylinder wall temperature at midstroke 8C 100 105

Ž .Cylinder wall temperature for ring 4 at bottom dead centre 8C 90 90

One criticism that can be fairly made of the results isthat rather bold assumptions were made about the torsionaltwisting behaviour of the piston ring based on very limited

w xstudies available in the literature 23 . This arose becausethe lubrication model used in this analysis did not deter-mine the cyclic variation of torsional twist angle of thepiston ring, which is needed by the wear model. Subse-quent development of the model has given it the capabilityto predict torsional twisting of the ring, albeit at an in-creased cost in terms of numerical complexity. Somewhatfortuitously, the predicted variation of torsional twist anglearound top dead centre firing, where much of the wear ispredicted to occur, as shown in Fig. 11, agrees well withthe assumed variation of twist angle given in Fig. 9.Hence, a greater level of confidence can be placed in thereported wear predictions than would have originally beenassumed. The application of the more advanced model,incorporating the calculation of torsional twist angle, willbe reported in a future paper.

6. Conclusions

Ž .i This paper reports the first application of a newinteractive dynamics, lubrication and wear model for pis-ton rings to data from fired engine tests.

Ž .ii The profile that the piston rings offer to the cylinderwall was observed to wear significantly during the enginetests, particularly for the top compression ring during thefirst 120 h. The surface roughness of the piston rings andthe cylinder wall also reduced dramatically over this pe-riod.

Ž .iii Lubrication predictions for the measured ring pro-files have highlighted the sensitivity of lubricant filmthickness and friction, between the piston ring and cylinderwall, to wear of the ring profiles.

Ž .iv The lubrication predictions also demonstrate thecapability of the model to evaluate the expected improve-

ment in piston ring performance during the early life of thecomponents, often referred to as the running-in period.

Ž .v Predicting the performance of the oil-control ringhas been shown to be very problematic, as the lubricationanalysis is very sensitive to measured profile parameterssuch as the relative heights of the two lands.

Ž .vi Interactive wear and lubrication predictions for thetop compression ring of the diesel engine studied showencouraging correlation between measured and predictedring profiles after 120 h running. This is particularlyencouraging in the light of assumptions made in the analy-sis in terms of wear behaviour in the boundary lubricationregime, availability of lubricant and torsional twisting ofthe piston ring.

Ž .vii Wear of the new top compression ring was pre-dicted to occur for much of the engine cycle with thelargest wear rates around top dead centre firing, where theapplied loads are high and film thicknesses low.

Ž .viii Strategies to overcome the limiting assumptions ofthe model have been identified.

Ž .ix This research advances the understanding of pistonring profile evolution with time and its dependence oncomplex interactions between lubrication and wear.

7. Nomenclature

h minimum film thickness between the pistonmin

ring and cylinder wallk wear factork wear factor in the boundary lubrication regime0

V worn volumeW normal loadx sliding distances

x axial coordinate of piston ring profilez radial coordinate of piston ring profilel specific film thicknesss composite root mean square surface rough-

ness of the piston ring and cylinder wall

( )M. Priest et al.rWear 231 1999 89–101 101

Acknowledgements

The authors would like to thank Shell Research, ShellResearch and Technology Centre, Thornton, UK and theIndustrial Unit of Tribology, University of Leeds, UK fortheir financial and technical support of this research. Par-ticular thanks go to Shell Research for use of the Caterpil-lar 1Y73 engine and for supplying the wear factor datareported in Table 3 and the LIF film thickness measure-ments summarised in Table 5. The advice of Dr. J.C. Bellof Shell Research in relation to surface topography andwear is greatly appreciated by the authors.

Appendix A

A.1. Caterpillar 1Y73 engine data

Further physical data for the ring pack is given forreference in Table 6. Basic engine data and operatingparameters relevant to the piston assembly are given inTable 7.

References

w x1 S.T. Gazzard, D.R. Eastham, R.J. Jakobs, R.L. Lunsford, Pistonsystem design for low emissions, Leading Through Innovation, T&NSymposium 1995, Paper 20, 1995.

w x2 R. Munro, Emissions impossible—the piston and ring support sys-tem, SAE Paper 900590, 1990.

w x3 M.L. Monaghan, Engine friction—a change in emphasis, 2nd BPTribology Lecture, Inst. Mech. Eng., 1987.

w x4 M.L. Monaghan, Putting friction in its place, 2nd Int. Conf.: Com-bustion Engines—Reduction of Friction and Wear, Inst. Mech. Eng.Conf. Pub. 1989-9, Paper C375rKN1, 1989, pp. 1–5.

w x5 D.A. Parker, D.R. Adams, Friction losses in the reciprocating inter-nal combustion engine, tribology—Key to the Efficient Engine, Inst.Mech. Eng. Conf. Pub. 1982-1, Paper C5r82, 1982, pp. 31–39.

w x6 R.J. Chittenden, M. Priest, The theoretical analysis of piston ringand piston skirt performance and its application in engine frictionmodelling, submitted to J. Tribol., Trans. ASME, 1998.

w x7 D. Dowson, Piston assemblies; background and lubrication analysis,Ž .in: C.M. Taylor Ed. , Engine Tribology, Tribology Series, 26,

Chap. 9, Elsevier, 1993, pp. 213–240.w x8 A. Mierbach, M.L. Hildyard, D.A. Parker, H. Xu, Piston ring

performance modelling, Leading Through Innovation, T and NSymposium 1995, Paper 15, 1995, pp. 15.1–15.13.

w x9 L.L. Ting, J.E. Mayer Jr., Piston ring lubrication and cylinder borewear analysis: Part I. Theory, J. Lubr. Technol., Trans. ASME,Paper 73-LUB-27, 1974, pp. 305–313.

w x10 L.L. Ting, J.E. Mayer Jr., Piston ring lubrication and cylinder borewear analysis: Part II. Theory verification, J. Lubr. Technol., Trans.ASME, Paper 73-LUB-25, 1974, pp. 258–266.

w x11 S. Furuhama, M. Hiruma, The relationship between piston ringscuffing and the formation of surface profile, Piston Ring Scuffing,Inst. Mech. Eng. Conf., Paper C67r75, 1976, pp. 35–43.

w x12 G.M. Hamilton, S.L. Moore, Measurement of piston ring profileduring running-in, Piston Ring Scuffing, Inst. Mech. Eng. Conf.,Paper C69r75, 1976, pp. 61–70.

w x13 A.V. Sreenath, N. Raman, Running-in wear of a compression igni-tion engine: factors influencing the conformance between cylinder

Ž .liner and piston rings, Wear 38 1976 271–289.w x14 A.V. Sreenath, N. Raman, Mechanism of smoothing of cylinder

Ž . Ž .liner surface during running-in, Tribol. Int. 9 2 1976 55–62.w x15 J.E. Rycroft, R.I. Taylor, L.E. Scales, Elastohydrodynamic effects in

piston ring lubrication in modern gasoline and diesel engines, in:Ž .Dowson et al. Ed. , Elastohydrodynamics—Fundamentals and Ap-

plications in Lubrication and Traction, Proc. 23rd Leeds-Lyon Sym-posium on Tribology, September 1996, Elsevier, 1997, pp. 49–54.

w x16 M. Priest, The wear and lubrication of piston rings, PhD thesis, TheUniversity of Leeds, 1996, 246 pp.

w x17 M. Priest, D. Dowson, C.M. Taylor, Prediction of the lubricationand wear of piston rings—theoretical model, World TribologyCongress, Abstracts of Papers, September, 1997, p. 82, full paper inpreparation for J. Eng. Tribol. Proc.—Inst. Mech. Eng., Part J, 1999.

w x18 J.K. Lancaster, Dry bearings: a survey of materials and factorsaffecting their performance, Tribology, December, 1973, pp. 219–251.

w x19 J.C. Bell, P.J. Davies, W.B. Fu, Prediction of automotive valve trainwear patterns with simple mathematical models, Mechanisms andSurface Distress, Proc. 12th Leeds-Lyon Symposium on Tribology,

Ž .Paper XI iii , 1985, pp. 323–333.w x20 J.C. Bell, T. Colgan, Pivoted follower valve train wear: criteria and

Ž .modelling, Lubr. Eng., J. STLE, February 1991 114–121.w x21 T.R. Thomas, Rough Surfaces, 1st edn., Longman Group, Harlow,

UK, 1982.w x22 M. Priest, D. Dowson, C.M. Taylor, Theoretical modelling of cavita-

tion in piston ring lubrication, accepted for J. Mech. Eng. Sci, Proc.—Inst. Mech. Eng., Part C, 1999.

w x23 B.L. Ruddy, D. Dowson, P.N. Economou, A.J.S. Baker, Piston ringlubrication: Part III. The influence of ring dynamics and ring twist,

Ž .in: S.M. Rohde, D.F. Wilcock, H.S. Cheng Eds. , Energy Conserva-tion Through Fluid Film Lubrication Technology: Frontiers in Re-search and Design, ASME pub., 1979, pp. 191–215.

w x24 M.A. Brown, H. McCann, D.M. Thompson, Characterization of theoil film behaviour between the liner and piston of a heavy-dutydiesel engine, SAE Paper 932784, 1993.