400 Commonwealth Drive, Warrendale, PA 15096-0001 U.S.A. Tel: (724) 776-4841 Fax: (724) 776-5760

SAE TECHNICALPAPER SERIES

and

2000-01-1780

Overview of Techniques for Measuring FrictionUsing Bench Tests and Fired Engines

Mike T. NoormanExxonMobil Research and Engineering Company

Dennis N. Assanis and Donald J. PattersonUniversity of Michigan

Simon C. Tung and Spyros I. TseregounisGeneral Motors Research and Development Center

Reprinted From: Advances in Powertrain Tribology 2000(SP–1548)

International Spring Fuels & LubricantsMeeting & Exposition

Paris, FranceJune 19-22, 2000

The appearance of this ISSN code at the bottom of this page indicates SAE’s consent that copies of thepaper may be made for personal or internal use of specific clients. This consent is given on the condition,however, that the copier pay a $7.00 per article copy fee through the Copyright Clearance Center, Inc.Operations Center, 222 Rosewood Drive, Danvers, MA 01923 for copying beyond that permitted by Sec-tions 107 or 108 of the U.S. Copyright Law. This consent does not extend to other kinds of copying such ascopying for general distribution, for advertising or promotional purposes, for creating new collective works,or for resale.

SAE routinely stocks printed papers for a period of three years following date of publication. Direct yourorders to SAE Customer Sales and Satisfaction Department.

Quantity reprint rates can be obtained from the Customer Sales and Satisfaction Department.

To request permission to reprint a technical paper or permission to use copyrighted SAE publications inother works, contact the SAE Publications Group.

No part of this publication may be reproduced in any form, in an electronic retrieval system or otherwise, without the prior writtenpermission of the publisher.

ISSN 0148-7191Copyright © 2000 CEC and SAE International.

Positions and opinions advanced in this paper are those of the author(s) and not necessarily those of SAE. The author is solelyresponsible for the content of the paper. A process is available by which discussions will be printed with the paper if it is published inSAE Transactions. For permission to publish this paper in full or in part, contact the SAE Publications Group.

Persons wishing to submit papers to be considered for presentation or publication through SAE should send the manuscript or a 300word abstract of a proposed manuscript to: Secretary, Engineering Meetings Board, SAE.

Printed in USA

All SAE papers, standards, and selectedbooks are abstracted and indexed in theGlobal Mobility Database

1

2000-01-1780

Overview of Techniques for Measuring FrictionUsing Bench Tests and Fired Engines

Mike T. NoormanExxonMobil Research and Engineering Company

Dennis N. Assanis and Donald J. PattersonUniversity of Michigan

Simon C. Tung and Spyros I. TseregounisGeneral Motors Research and Development Center

Copyright © 2000 CEC and SAE International.

ABSTRACT

This paper presents an overview of techniques formeasuring friction using bench tests and fired engines.The test methods discussed have been developed toprovide efficient, yet realistic, assessments of newcomponent designs, materials, and lubricants for in-cylinder and overall engine applications.

A Cameron-Plint Friction and Wear Tester was modifiedto permit ring-in-piston-groove movement by the testspecimen, and used to evaluate a number of cylinderbore coatings for friction and wear performance. In asecond study, it was used to evaluate the energyconserving characteristics of several engine lubricantformulations. Results were consistent with engine andvehicle testing, and were correlated with measured fueleconomy performance.

The Instantaneous IMEP Method for measuring in-cylinder frictional forces was extended to higher enginespeeds and to modern, low-friction engine designs. Acomparison of historical cylinder friction measurementsshows reductions of 85% for late model piston /cylinderbore designs. A technique for accurately measuringoverall engine friction was developed and used to assessthe benefits of friction modifiers with an ability to measurechanges in friction less than 1%.

INTRODUCTION

Engine design and tribology engineers are constantlychallenged to innovate advanced products to meet moredemanding emissions and fuel economy targets. Currentresearch and development on engine cylindercomponents include new designs, materials, coatingsand surface treatments, with the goals of reduced weight,

longer life, higher operating temperatures, and reducedfriction.

Tribological bench testing in the laboratory can providerapid and cost effective information, and is often used forscreening or ranking purposes in the developmentprocess of new engine materials and lubricants. Theauthors will review the current development of selectedbench tests employing either rotary or reciprocatingmotion for evaluating friction and energy-conservingcharacteristics of lubricants. The main advantage of thedeveloped bench tests is that parts of real components(cylinder liners and piston rings) are tested, preservingthe geometry and metallurgy of the engine, therebypermitting a representative evaluation of surface finishesand oils. Results from materials and lubricants studieswill be presented and correlated with vehicle testing.

Since the engine environment cannot be completelysimulated in bench tests, engine tests are often neededto verify and validate findings. The authors havedeveloped several techniques to measure overall andcycle-resolved, reciprocating component friction in firedengines. These techniques, and current research anddevelopment efforts (including investigation of newengine designs and energy-conserving lubricants) will bereviewed.

TRIBOLOGICAL BENCH TESTS

This section reviews bench testers employing eitherrotary or reciprocating motion for evaluating the friction,wear, and energy-conserving characteristics of interfacematerials and lubricants. Several configurations, asshown in Table 1, are in common use for friction and wearevaluation [1]. General information is presented onexperience and practice with one rotary (Falex Type) andone reciprocating tester (Modified Cameron-PlintMachine) developed in the authors' laboratory.

2

Table 1. Common Bench Test Classification

Figure 1. Schematic of Falex Type Block-On-Ring Friction and Wear Tester

ROTARY BENCH TEST METHOD – Standard rotarybench devices include pin-on-disk (ASTM G99), block-on-ring (ASTM G77), cross cylinders, and four ball tests[1,2,3]. While these tests can be designed for uni-directional, rotary or reciprocating motion, they all involvea non-conformal contact geometry. In the last 25 years,the rotary block-on-ring tester [4] (designed as LFW-1,also referred to as Falex friction and wear tester), shownin Figure 1, has been used in automotive laboratories forevaluation of piston rings, engine blocks, liners, valveguides, as well as surface coatings. Scuff results havebeen obtained by incrementally increasing load untilfailure is detected by friction measurement. Severalprocedures have been used for wear testing with thisdevice. These have involved variations in load, speed,lubrication, temperature, and duration. The coefficient offriction is determined at the beginning and end of eachtest from the weight applied and the force transduceroutput. Block wear is most often evaluated by calculatingthe volume removed from the block wear scar area. Thering wear is evaluated by weight loss or a subjectiveevaluation of visual appearance. Test procedures for thisrotary block-on-ring tester, used extensively for pistonring and cylinder liner evaluations, have been describedin a reference paper published by Patterson, Hill, andTung [5]. The non-realistic, non-conformal contact ofstandard specimens used in the block-on-ring test

(versus the conformal contact of the piston ring againstthe cylinder bore) limits the relevance of the test. Ingeneral, the main disadvantage of rotary bench testmethods is that real components cannot be tested;hence, the geometry of engine components is notpreserved, and representative surface finishes cannot beevaluated.

Figure 2. Schematic of Cameron-Plint High Frequency Friction Tester

RECIPROCATING TEST METHOD - MODIFIEDCAMERON-PLINT FRICTION AND WEAR MACHINE –

A Cameron-Plint High-Frequency Friction and WearTester has been modified (Figure 2), to serve as a tool forevaluating advanced lubricants or advanced enginematerials for possible use on engine cylinder bores. Themodified Cameron-Plint High Frequency Friction andWear Tester provides a reciprocating motion whichmakes it suitable for simulating piston/cylinder linerdynamics [6]. The main advantage of this bench testmethod is that real components can be tested, and sincethe geometry of the engine is preserved, representativesurface finishes can be meaningfully evaluated. Aspecimen holder is moved back and forth across acylinder segment as would occur in an engine. Inaddition, the holder has been modified to allow the pistonring to move in a slot, simulating the movement thatoccurs in a piston ring groove. Friction and electricalresistance are continuously measured during the test,and the wear of the cylinder liner and piston ring sectionsare measured with the use of a surface analyzer, usingstandard techniques [7]. In order to evaluate the stablefriction characteristics of engine bores, friction coefficientmeasurements are performed after a sufficiently longsliding test time to ensure stabilization, typically 5 hours.Friction force is reported as the root mean square (RMS)of the instantaneous friction force over one cycle. Thereciprocating bench test procedures are described indetail in [6,7]. Selected results from representativestudies of advanced engine materials, coatings andenergy-conserving lubricants are presented below toillustrate the potential and use of this method.

3

Figure 3. Friction Coefficients of Tested Engine Materials

Figure 4. Wear Resistance of Tested Engine Materials

RECIPROCATING BENCH TEST RESULTS ONSEVERAL ADVANCED ENGINE MATERIALS – Newthermal spray processes are used to coat aluminumbores with either aluminum bronze or iron oxide coatingsto provide a hard surface with good lubricationcharacteristics. Thermal spray coatings are consideredadvantageous replacements for cast iron since they havereduced weight, better heat transfer properties and lowercosts. In order to assess possible friction and wearmerits of such coatings, engine bore and piston ringcomponents have been used in bench tests underconditions typical of the operating engine. In particular,temperatures and loads were chosen to represent theactual operating conditions at top-dead center,immediately after combustion. Advanced bore materialsand coatings that were considered as gray cast ironalternatives for a particular 3.8L engine design includedtwo thermal sprayed alloys produced by plasma transferwire arc processes (PTWA-bronze and PTWA-1040steel) [7]; two thermal sprayed alloys produced bydiamond jet processes (DJ -bronze and DJ-1025 steel)[8]; and 390 Al-Si alloy and Nikasil composites.

Results show that the coefficient of friction did not varysignificantly between the different test cylinder bores, asshown in Figure 3. After 5 hours of sliding, all materialsreached stable friction characteristics. Friction was alsomeasured after 20 hours and 40 hours to determine thesurface impact over extended periods of sliding. Theaverage coefficient of friction of the thermal sprayed steelcoatings (either PTWA 1040 steel or DJ 1025 steelcoatings) was 10% lower than the thermal sprayedbronze coatings, but slightly higher than the cast ironbore, as shown in Figure 3; the latter finding is attributedto the rougher surface finish of the thermal sprayedcoatings.

The reciprocating bench test is especially useful instudying the wear patterns of surface coatings, which candiffer from those of iron. Thermal spray microstructuresconsist of layered structures, called “splats,” from theflattening of the molten metal as it hits the surface. Thisresults in layers of the parent material interlaced withoxides. Oxide layers between the splats are the weaklink in the coating, as cracking occurs along the oxideboundaries, eventually releasing the splats [7,8]. Thiswear mechanism is usually referred to as “splatdelamination.” Figure 4 shows bore wear depth as afunction of test time for all specimens. Clearly, thediamond jet (DJ) aluminum bronze coating experiencesthe greatest wear. The PTWA aluminum bronzeperforms similarly to that of the aluminum alloy 390. Theferrous PTWA 1040 coating performed best, and was asgood as or better that the Nikasil coating.

These findings correlate very well with previouslyreported engine friction and wear tests [6,7]. Frictionbenefits followed the same trends, and consistent wearmechanisms were observed in both bench and enginetests.

RECIPROCATING BENCH TEST RESULTS ONENERGY-CONSERVING OILS – Bench tests can alsobe used to assess the tribological properties of engineconserving lubricants, including studies of base stockformulation, viscosity index modifier, and friction modifier.For instance, Table 2 lists a set of oils, which is exploredin detail using bench tests in a companion paper [9].These oils have been also described in reference [10],where related fuel economy characteristics in vehicletests were reported. The selection of test oils spans arange of viscosity grades. The influence of frictionmodifiers, including organo-molybdenum additives(MoDTC), on friction and wear was also investigated. Byincluding oils REO8 and REO9, the effects of using aformulation with a high viscosity index basestock wascompared to one with a conventional basestock. Notethat the ASTM FM-8 formulation is a friction-modifiedversion of HR-4, the Sequence VI reference oil.

Figure 3 Friction C oeffic ients o f Tested E ng ine M ateria ls

0.095

0.100

0.105

0.110

0.115

0.120

0.125

Cas t IronBore

Alum inum390

Nikas il PTW A1040 Steel

D J 1025 Steel

PTW ABronze

DJ Bronze

Fri

cti

on

Co

eff

icie

n5th hour

20th hour

40th hour

Figure 4 W ear Resistance of Tested E ng ine Materia ls

0

2

4

6

8

10

12

14

16

18

20

0 20 40 60 80

Tim e [h ours]

We

ar

De

pth

[m

icro

ns DJ - Bronz e

Ca s t Iro n

A lumin um 39 0

PTW A - B ronz e

Nikas il

PTW A - 10 40 Stee l

DJ - 1025 S tee l

4

Table 2. Engine Oils Used in Cameron-Plint Bench Tests and in Fuel Economy Tests

Figure 5. Friction Coefficient vs. Mo Concentration. Solid line represents a power fit of all data points. Dashed line is the average value of the friction coefficients of the four oils with the lowest friction coefficient values (REO 8, REO 9, 529, M-2)

A detailed surface mechanism study on the impact ofenergy-conserving lubricants has been published in acompanion study [9]. Very significant differences incylinder bore wear depths were observed after 5 hours ofsliding using the different lubricants. Surface analysistechniques were used to reveal the presence of elementsfound in oil formulations, and to correlate them withfriction reduction mechanisms of the energy conservinglubricants. A plot of measured friction coefficient versusthe molybdenum concentration on the cylinder surface isshown in Figure 5. These results indicate a relationshipbetween friction reduction and molybdenumconcentration. A critical concentration of molybdenum(about 1.0 – 1.5 micrograms/cm2) appears to benecessary for the friction reducing film formation tobecome evident. Beyond the critical concentration,friction stabilizes and approaches an asymptotic level.These observed effects of molybdenum surfaceconcentration on bench-measured friction coefficientcorrelates well with ultimate fuel economy improvements

on a vehicle tested over the FTP driving cycle, as shownin Figure 6.

Figure 6. Fuel Economy vs. Molybdenum Concentration

Table 3. Vehicle Fuel Economy Data, Viscometric Properties of the Oils, and Friction Coefficient Data

Table 4. Fuel Economy Regression Results

The bench-measured friction coefficient data and theviscometric properties of the oils (Table 3) weresystematically correlated to vehicle dynamometer fueleconomy data, as measured in two GM vehicles: a 1993Buick LeSabre with a 3.8L V6 engine and a 1993 PontiacGrand Am with a 2.3L Quad4 engine [10]. The engineswere chosen to represent the two major types of GMvalve train configurations currently produced, a rollerfollower design with low tension rings and two valves percylinder, and a bucket tappet, dual overhead cam withfour valves per cylinder design. In the correlation, thevehicle fuel efficiency data were assumed to depend on

F ig u r e 5 . F r i c t i o n C o e f f i c i e n t v s . M o C o n c e n t r a t io n .S o l id L i n e R e p r e s e n t s a P o w e r F i t o f a l l D a t a P o i n t s .

T h e D a s h e d L i n e i s t h e A v e r a g e V a lu e o f t h e F r ic t io nC o e f f ic ie n t o f t h e F o u r O i ls w i t h t h e L o w e s t F r i c t i o nC o e f f ic ie n t V a l u e s ( R E O 8 , R E O 9 , 5 2 9 , M - 2 )

M o C o n c e n t r a t i o n , m ic r o g r a m s /c m

1 0 3 0 H R -4

O i l B

B C

2 0 5 0

F M -8

R E O 8

5 2 9

R E O 9 0 . 0 3 4

M -2

0 . 0 9 0

0 . 0 8 0

0 . 0 7 0

0 . 0 6 0

0 . 0 5 0

0 . 0 4 0

0 . 0 3 0

0 . 0 2 0

0 . 0 1 0

0 . 0 0 00 . 0 0 . 5 1 . 0 1 . 5 2 . 0 2 . 5 3 . 0

2

Fr

icti

on

C

oe

ffi

ci

en

t

2

5

one of the viscometric variables (to describe thehydrodynamic contribution to fuel economy) and on oneof the friction variables (taken as a measure of theboundary friction contribution of the oil to the vehicle'sfuel economy). Analysis of variance was used to identifywhich one of the viscometric and friction variablescorrelated best with the fuel economy data. The resultantcorrelation equations are shown in Table 4 for the twovehicles, as well as for the average fuel economy data ofthe two vehicles. A schematic plot of the predicted fueleconomy for each vehicle and their average (ascalculated from the equations) versus the measured fueleconomy is shown in Figure 7.

Figure 7. Predicted vs. Measured Fuel Economy in Vehicle Tests

Figure 8. Force Balance Around the Piston Assembly

The friction coefficient measured during the (5th) hour inthe friction test (FC5th hr) correlated best with the fueleconomy data. This is not surprising, since the 5th hourfriction coefficient represents friction values measuredwith broken-in surfaces of the friction pair, a conditionthat is expected in the engine cylinder-ring area that thebench friction test was designed to simulate.

By examining the relative contribution of the variousterms in the equations in Table 4, we conclude that the2.3L engine has a higher sensitivity to boundary frictionthan the 3.8L engine. This in agreement with the well-known notion that in small-displacement engines, agreater portion of the mechanical friction is attributed toboundary losses than in large displacement engines. Inaddition, the valvetrain of the 2.3L engine (buckettappets) is more sensitive to boundary lubrication thanthe valvetrain of the 3.8L engine (roller followers).

FIRED ENGINE FRICTION METHODS

CYLINDER FRICTION – Motoring of an engine by adynamometer is the easiest method of quantifyingfriction. By progressively disassembling the engine, thefriction of various components can be inferred. Literaturesuggests that the piston ring assembly may make up asmuch as 60-75% of total engine friction [11]. The primarydrawbacks of the motoring technique are the loss of thecontribution of firing compression and the in-cylinder heateffects. More advanced techniques seek to study thecylinder friction under firing conditions using moving orfloating cylinder liners [12,13,14,15]. These techniquesrequire considerable modifications to the test engine,which raise questions about the representative nature ofthe resulting modified engine.

The Instantaneous IMEP Method was developed in theearly 1980’s as a way to measure cylinder friction in afired engine with minimal modifications [16]. The methodis based on an axial force balance around the pistonassembly, which includes gas force, crankshaft force,inertial force and friction force (Figure 8). Friction can bedetermined by simultaneously measuring cylinderpressure and connecting rod strain, and calculatinginertia. Instrumentation of the test engine is not trivial,but once in place, is minimally intrusive and therefore hasminimal impact on the in-cylinder dynamics of the originalsystem design. Instrumentation includes:

1. A water cooled, piezo-electric pressure transducer inthe cylinder head,

2. Four strain gages affixed to the connecting rod, twoon each side, positioned parallel and perpendicularto the axis,

3. A shaft encoder on the crankshaft, capable ofindicating 1 degree increments as well as engineTDC, and

4. Construction of a light-weight, yet durable,“grasshopper” linkage connecting the instrumentedconnecting rod cap to the engine block. This linkageis used to transport the strain gage wires out of theengine.

A high-speed data acquisition unit is used to collectpressure and strain signals at each crank angle. Signalsare conditioned and converted to forces. Cylinderpressure is first decreased by barometric pressure (whichacts on the underside of the piston) and is multiplied by

-2.0

-1.5

-1.0

-0.5

0.0

0.5

1.0

1.5

-2.5 -1.5 -0.5 0.5 1.5 2.5

Measured FE [mpg], relative to BC oil

Pre

dic

ted

FE

[m

pg

], r

ela

tiv

e t

o o

il B

2.3L, (R^2=0.66)

3.8L, (R^2=0.67)

AVG, (R^2=0.75)

Weight

friction = ƒ(gas pressure, inertia, weight, wrist pin force)

6

bore area to determine the gas force acting on thetopside of the piston. Strain data are converted directlyto force via the conditioning step, but must bedecomposed into forces acting parallel and perpendicularto the cylinder axis. This is done using the crank angleinformation and slider-crank theory. Examples of axialgas and conrod forces over a complete combustion cycleare presented in Figure 9.

Figure 9. Gas & ConRod Forces Acting on Piston Assembly 1500 rpm, 1/6 WOT Load

Total inertial force is composed of the inertia of thepiston, and the inertia of the length of the connecting rodfrom the midpoint of the strain gages to the wrist pinopening. Inertia is calculated directly from mass andacceleration (via slider-crank theory). Since the piston,rings and wrist pin travel exclusively in an axiallydirection, piston inertial force can be determined byassuming a point mass, while the complex motion of theconrod through the cylinder bore requires that its inertiabe rigorously determined using a distributed massapproach. Friction force of the piston assembly (Figure10) is calculated by subtracting the resulting gas force,inertial force and component weights from the axialconrod force.

One drawback to the Instantaneous IMEP Method is thatthe resulting friction calculation is the difference of tworelatively large force measurements determined usingtwo differing techniques; therefore, the accuracy of themethod is very dependent on the accuracy of the cylinderpressure and connecting rod force measurements.Calibration is therefore critical. Also, thermal shock of thepressure transducer following the combustion event is acommon problem and care must be taken to select amodel and an individual unit that is stable. Thermalshock causes a short-term drift in the pressuretransducer, and since the strain gages are unaffected bythe combustion event, the drift will show up in thecalculated friction. Strain gages should be installed onthe connecting rod to minimize sensitivity to bending, andat the same time be as close to the center of gravity ofthe rod as possible. Wiring the gages in a WheatstoneBridge format increases sensitivity and compensates fortemperature variations.

Figure 10. Cylinder Friction Acting on Piston Assembly 1500 rpm, 1/6 WOT Load

Complete details of the Instantaneous IMEP method canbe found in earlier SAE publications by Uras andPatterson [16,17]. The method has been applied toseveral engines, and used to evaluate the effects of oilviscosity, friction modifiers, ring tension and piston designon cylinder friction [16,17,18,19]. Generalized findingsinclude:

1. The ability to measure cylinder friction, on a cycleresolved basis, with enough accuracy to identifyboundary, mixed and hydrodynamic lubricationregimes. Boundary lubrication exists at TDC andBDC, where piston speeds diminish to zero, causingthe collapse of the oil film. Mixed lubrication, whereincreasing piston velocity increases oil film thicknesscausing a decrease in observed friction, is commonlyfound at the beginning and end of the stroke (first andlast 30 degrees). Both mixed lubrication andhydrodynamic lubrication (where increasing speedscause increased friction) have been observed in themid stroke region, depending on engine speed andload.

2. Extremes in oil viscosity (high and low) showedincreased cylinder friction. At high viscosity levels (29cSt @ 100 °C), mixed lubrication regimes gave way tohydrodynamic, increasing friction. The effects ofreducing viscosity to very low levels (2 cSt @ 100 °C)were relatively small at mid stroke, but very pronouncedapproaching TDC and BDC, as mixed lubrication gaveway to boundary conditions. The use of friction modifiershas been shown to decrease peak friction force by 30-40% at TDC and BDC, depending on oil viscosity.Overall cylinder friction, however, was not significantlyreduced.

3. Cylinder friction losses under fired conditions werehigher than those under motoring conditions, at thesame coolant and oil temperatures, engine speedand intake pressure, due to ring loading and heattransfer. At low engine speeds, overall differences ofup to 30% were observed, with the biggest changesobserved in the power and exhaust strokes. Thesedifferences decreased at higher speeds and reducedloads where lubrication was more hydrodynamic.

Figure 9 Gas & ConRod Forc es Ac t ing on P is ton As sem bly1500 rpm , 1/6 W OT Load

-8000

-6000

-4000

-2000

0

2000

4000

6000

8000

0 90 180 270 360 450 540 630 720

Crank Angle [degrees]

Fo

rce

[N

Gas Force

Ax ial ConRod Force

Figure 10 Cy linder F ric tion A c ting on P is ton A s sem bly1500 rpm , 1/6 W OT Load

-80

-60

-40

-20

0

20

40

60

80

0 90 180 270 360 450 540 630 720

Crank Ang le [degrees]

Fri

cti

on

Fo

rce

[

7

4. While cylinder friction varies with speed and load,results show that it represents only 20-30% of overallengine friction. This is contrary to literature findings[11] which suggest cylinder friction is responsible for60-75% of overall engine friction.

Initial work by Uras and Patterson was conducted at 500rpm, later increasing to 1600 rpm. The InstantaneousIMEP Method was extended by Clampitt [20] in the late1980’s to measure cylinder friction at engine speeds upto 4200 rpm, based on advances in high-speed dataacquisition systems which made it possible to collect dataat higher speeds and for a longer duration. Datacovering 100 cycles could be captured, up from amaximum of 9 just a few years before. Increased cycleaverages helped smooth data considerably, which provedvaluable at higher engine speeds.

A program [20] was conducted to evaluate theperformance benefits of a new, light-weight pistonassembly. A 1989 fuel injected 2.5L four cylinder enginewas instrumented with pressure transducer, strain gagesand grasshopper linkage on cylinder 1. Operated from2000 to 4200 rpm at full load conditions, increased brakepower measurements indicated improved performancefor the light weight piston over the stock assembly; whileno substantial differences were observed in cylinderfriction. Subsequent work by Clampitt [20] showed thatbenefits were related to improved combustion with thenew design rather than weight savings. These resultsare consistent with work done by Uras and Patterson [16]which showed that reduced weight pistons can actuallyincrease friction slightly (due to deformation), dependingon speed. Additional work by Uras and Pattersontargeted at investigating the effect of piston weight atconstant design confirmed that decreasing weight alonehas negligible effects on cylinder friction.

Figure 11. Normalized Cylinder Friction Stock and Lightweight Piston Assemblies

Cylinder friction results from the 2.5L4 engine,normalized to overall engine friction, are presented inFigure 11. The light-weight piston design exhibitedoverall similar friction characteristics to the stock units,with slightly higher levels observed in the 2600 to 3000speed range. Results ranged from 30-40% over the 2000to 3400 rpm range, which are in closer agreement to theUras and Patterson results of 20-30% (over the 500 to

1600 rpm speed range) than the 60-75% previouslysuggested by literature.

Clampitt compared fired and motored frictionmeasurements, with mixed findings. Virtually nodifference between fired or motored conditions wereobserved for the overall and cylinder frictionmeasurements using stock pistons (Figure 12), while anaverage of 12 – 15% reductions were found for the lightweight piston design (Figure 13). These findings show aconsiderably smaller difference between fired andmotored friction than the 30% reported earlier by Urasand Patterson.

Figure 12. Fired vs. Motored Friction – Stock Piston

Figure 13. Fired vs. Motored Friction – Lightweight Piston

A new program is underway designed to extend theInstantaneous IMEP Method to late model, low-frictionengine designs. Tests have shown limited success inapplying the method to a 1995 2.5L V6 engine. The all-aluminum, split-block design uses low-friction, rollered,multivalve heads and low-friction, molybdenum coatedshort skirt pistons. Limited testing has been conductedat 1500 rpm and 1/6 full load, which approximates asteady road cruise operation of 45-50 mph. Cylinderfriction was measured at 20% of overall engine friction,consistent with the low end of the 20-30% range reportedby Uras and Patterson.

The most challenging aspect to the success of thisprogram has been the low absolute friction levels of theengine design. In terms of friction mean effectivepressure (FMEP), nominal cylinder friction valuesaveraged 110 mbar. This compares to nominal values of

F ig u re 11 Norm aliz ed Cy linder F ric tionS toc k and Ligh twe ight P is ton A s s em blies

0 %

1 0 %

2 0 %

3 0 %

4 0 %

5 0 %

6 0 %

7 0 %

8 0 %

9 0 %

1 0 0 %

2 0 0 0 2 4 0 0 2 8 0 0 3 2 0 0 3 6 0 0

Engine Spee d [rpm]

Cy

lin

de

r F

ric

tio

n,

as

% o

f

Fri

cti

on

S toc k P is ton A s s em bly

Ligh twe ight P is tonA s s em bly

Figu re 12 Fired vs . M otored Fric tion - S toc k P is ton

0

5 0 0

1 0 0 0

1 5 0 0

2 0 0 0

2 5 0 0

2 0 0 0 2 4 0 0 2 8 0 0 3 2 0 0 3 6 0 0 4 0 0 0 4 4 0 0

Engine Speed [rpm]

Fri

cti

on

Me

an

Eff

ec

tiv

e P

res

s[m

ba

r]

Fired O verall FM E P

M otored O verall FM E P

Fired Cy linder FM E P

M otored Cy linder FM E P

Figure 13 Fired vs . M otored Fric tion - Lightwe ight P is ton

0

5 0 0

1 0 0 0

1 5 0 0

2 0 0 0

2 5 0 0

2 0 0 0 2 4 0 0 2 8 0 0 3 2 0 0 3 6 0 0 4 0 0 0 4 4 0 0

Engine Speed [rpm]

Fri

cti

on

Me

an

Eff

ec

tiv

e P

res

[mb

ar] Fired O verall FM E P

M otored O verall FM E P

Fired Cy linder FM E P

M otored Cy linder FM E P

8

480 mbar and 730 mbar for engine work done by Assanisin the late 1980’s and Uras and Patterson in the early1980’s, respectively [16,20]. On a force basis, averagecylinder friction forces of 60 N were sought from adifference between gas and conrod forces that peakedover 6000 N.

Other design issues also raised new challenges. Limitedspace in the four-valve cylinder head necessitated theuse of the smallest available pressure transducers. Non-water-cooled, Kistler 6125A units equipped with thermalshock resistant covers were installed through the waterjacket of the cylinder head. Thermal shock wasminimized, but still apparent for 30 degrees following thestart of combustion. Powdered metal technologyconnecting rods made strain gage selection difficult, ascompensation for temperature related expansion of theconnecting rod is done by selecting strain gages with like-metal backings. This problem was circumvented bycalibrating the instrumented connecting rod at in-usetemperatures.

OVERALL ENGINE FRICTION – The InstantaneousIMEP Method has shown that cylinder friction makes upmuch less of the overall engine friction than wasoriginally expected. While the Method remains asensitive tool for understanding in-cylinder events, it maybe inappropriate to generalize findings to overall enginefriction performance. This would be particularly truewhen assessing lube oil properties such as viscosity andfriction modifier content.

While numerous bench tests and motored engines areavailable for measuring frictional changes and benefits,there is often a strong desire to augment this informationwith data generated from “real world” engine operation.A research program was begun in early 1996 with theprimary objective of developing a testing protocol inwhich the benefits from engine oil friction modifier couldbe quantified in a fired-engine stand. This involveddeveloping a method for measuring engine friction, andthen applying the technique to measure changes infriction.

Work lost to engine friction is the difference between thework generated within the cylinders and the sum of thework available at the flywheel and the work used to runthe engine accessories (generator, water pump, powersteering pump, air conditioning compressor etc.). Whileinternal work can be determined through the use ofcylinder pressure indicators (commonly expressed interms of mean effective pressure as "indicated meaneffective pressure" or IMEP) and flywheel work by thepower absorbed by a dynamometer (brake meaneffective pressure or BMEP), the work given up to driveaccessories is harder to measure. By eliminating orminimizing the number of accessories driven by theengine, friction can be approximated as the differencebetween the internal work and the work at the flywheel:

Friction Mean Effective Pressure (FMEP) = IMEP – BMEP

The general developmental strategy was to operate anengine at a given steady-state operating condition,accurately measure cylinder pressure and powerabsorption at the dynamometer, introduce the frictionmodifier additive, and then re-measure cylinder pressureand brake power. The largest obstacles were achievingtrue steady-state operation in a modern, microprocessor-controlled engine, and data processing.

A 1995 2.5V6 engine was used in this program, mountedon an electric dynamometer designed to maintain a setengine speed by absorbing flywheel power. Engine loadwas controlled by throttle position. All accessories wereremoved from the engine, leaving a single drive belt thatoperated the stock water pump. Engine coolant and oiltemperatures were controlled to 95 and 105 °C,respectively.

The engine was equipped with complete factoryemissions equipment and was controlled by an electroniccontrol module (ECM) through an OEM wiring harness.Provisions were made to monitor and control such inputparameters as spark advance and barometric pressure.In addition, the ECM was removed from adaptive learningmode. While these changes do not provide absolutecontrol over the engine operation, the ECM was morelikely to operate using a limited number of operationalmaps. This had a marked effect on maintaining steady-state operation.

Cylinder-to-cylinder variations in IMEP can exceed 10%.Since FMEP is a relatively small value determined as thedifference between much larger IMEP and BMEPmeasurements, extrapolating engine IMEP data from asingle cylinder can introduce substantial error.Therefore, six Kistler pressure transducers (Model6125A) were machined one in each combustionchamber, adjacent to the spark plug. The four-valveconfiguration of this engine necessitated the use ofextremely small, non-water-cooled, pressuretransducers. However, the transducers were mountedthrough the cylinder head’s internal coolingpassageways, essentially providing constant temperatureoperation to the bulk of the unit.

A Tektronix high-speed data acquisition unit collecteddata once every crank angle, triggered by a Kistler shaftencoder. The unit converted analog signals to digital andheld the data for the duration of the measuring event,which typically lasted 87 cycles. Once completed, theentire data set was uploaded to a Pentium-basedpersonal computer for processing.

The upload data set included the followingmeasurements for each crank angle of each cycle: sixcylinder pressures, six cylinder volumes, manifoldabsolute pressure (MAP), barometric pressure, anddynamometer load. At 720 crank angles per cycle for 87cycles, the entire data set consists of almost one milliondata pieces, too large to save as is. The results wereprocessed by first correcting for drift in the pressuretransducers by pegging the cylinder pressure

9

measurements at bottom dead center of the intakestroke, for each cylinder, to the MAP reading. The data ateach crank angle was then used to make individualcylinder IMEP and PMEP calculations for each cycle.Using the average BMEP measured over the cycle, 87individual estimates of FMEP were made.

At the beginning of each test, the engine was flushed andfresh lube oil installed. Speed and load conditions wereset, and the engine was allowed to stabilize. Dataacquisition was started at engine start up andautomatically repeated measurements at roughly two-minute intervals for the duration of the test. Frictiondropped rapidly as the engine warmed up, and stabilizedin as little as one hour, depending on speed and load.Stabilization was signified by 6 to 10 data sets with FMEPvalues within 10 mbar (preferably 5 mbar) of each other.Figure 14 is given as an example. Once stabilized, theFMEP of the base engine oil was measured, then frictionmodifier was weighed into a syringe and injected into therunning engine. Friction reduction resulting from theadditive was evident within two minutes. Stabilization atthe new friction level typically occurs within fiveacquisition cycles. The test was either terminated at thispoint, or a second additive dose was injected.

Figure 14. Engine FMEP 1500 rpm, 1/6 WOT Load, Reference Oil

Effects of the friction modifier were calculated bycomparing the average values for each stabilized FMEPregion following an injection to the values generated forthe base region prior to the initial injection (Figure 14).Since additive effects occurred almost instantly, therewas no need to account for drift in engine or lubeperformance. The method has since been extended toaccount for time-dependent events, such as viscosityshear, and will be the subject of a future publication.

A correction factor for BMEP was calculated to determinethe power increases resulting exclusively from the frictionmodifier. Ideally, IMEP remains constant under “steadystate” conditions once the engine warms up. In realitythough, slight changes sometime occur from onestabilized region to another (Figure 15). IMEP changesdirectly affect changes in BMEP, so if more or less poweris generated in the engine, a similar increase or decreaseis seen at the flywheel. Friction losses are essentiallyconstant over the small variations observed in IMEP at agiven steady-state condition. These changes in IMEP, if

any, were accounted for by normalizing BMEP to IMEP.Power benefits generated by friction modifier use weredetermined by comparing these normalized powerfigures between the stabilized regions.

Figure 15. Engine IMEP & BMEP 1500 rpm, 1/6 WOT Load, Reference Oil

Overall Engine Friction Test Results – Friction modifiereffects were investigated at engine speeds of 1500, 2000and 3000 rpm at selected loads varying from 1/12 to 3/4WOT load. The test results are presented in Figure 16.Initial inspection shows that additive-related frictionbenefits increase with decreasing engine speed, whilepower benefits increase with decreasing load. Maximumfriction changes were observed at 1500 rpm and 1/6WOT load. Interestingly this speed/load condition isconsistent with steady-state driving at suburban speeds,and is similar to the operating conditions of the SequenceVI test (1500 rpm, approximately 6 hp), an enginelubricant testing standard used for fuel economycertification.

Closer inspection of Figure 16 yields an intuitivelyobvious finding: the magnitude of engine friction (FMEP)changes are equal to the increases in corrected power(BMEPc). This firmly establishes the relationshipbetween friction and power. Less intuitive is the unusualfinding that the magnitude of the friction drop (and hencepower increase) is nearly constant over the speed/loadranges investigated. This explains why friction benefits(change on a percentage basis) are maximized at lowspeeds and power benefits are maximized at low loads:absolute friction levels are lowest at low speeds andabsolute power levels are lowest at low loads.

A series of subsequent tests were conducted at thissame 1500 rpm & 1/6 WOT load condition to evaluatetest repeatability, the effects of various lube formulations,and ring tension. Using a commercially available 5W-30mineral oil meeting API Energy Conserving II standards,repeat testing over several days yielded an averageFMEP of 562 mbar, +/- 23 mbar at the 95% confidenceinterval. When used to determine instantaneouschanges, sensitivities less than 1% were achieved at the95% confidence interval.

Examples of friction measurements with a variety ofdifferent commercial lubes are presented in Figure 17.The FMEP of four commercial engine oils varied from

450

470

490

510

530

550

570

590

610

630

650

0:00 1:00 2:00

Tim e [h r]

FM

EP

[m

ba

r]

Friction Modif ie r A dded

Friction Modif ie r A dded

Figure 15 E ng ine IM E P & B M E P1 5 0 0 R M P , 1 /6 W O T L o ad , R e fe re nc e O il

2 60 0

2 62 0

2 64 0

2 66 0

2 68 0

2 70 0

2 72 0

2 74 0

2 76 0

2 78 0

2 80 0

0 :00 1 :00 2 :00

Time [hr]

IME

P [

mb

a

1 40 0

1 42 0

1 44 0

1 46 0

1 48 0

1 50 0

1 52 0

1 54 0

1 56 0

1 58 0

1 60 0

BM

EP

[b

a

< - IM E P-> B M E P

Fric tion Modifie r Added

F ric tion Modifie r Added

10

488 – 604 mbar, and the effectiveness of friction modifier,at constant dosage, varied between 6-11%. Powerincreases ranged from 2.5-4%, and followed theobserved friction reductions.

To get a qualitative estimate for the benefits associatedwith low-friction piston design, the stock pistons werereplaced with a duplicates equipped with flat-faced,gapless compression rings, and oil control rings with

three times greater tension. The molybdenum disulfidecoatings were also removed from the piston skirts. Afterbreak-in, tests were conducted using the reference oil at1500 rpm & 1/6 WOT load. FMEP increased 29% to 724mbar. Addition of friction modifier resulted in FMEPreductions of 16% (compared to 11% for stock pistons),suggesting that advances in piston design have reducedrubbing friction.

Figure 16. Friction and Power Measurements at Various Speeds & Loads

Figure 17. Effect of Lube Oil Formulation on FMEP 1500 RPM & 1/6 WOT Load

RPM 1500 Power, hp 25.7 RPM 3000 Power, hp 51.4Torque, ft*lbs 90 Rated Load 3/4 Torque, ft*lbs 90 Rated Load 3/4

BMEPc FMEP BMEPc FMEPBase Condition, mbar 5865 589 Base Condition, mbar 6031 860FM Added, mbar 5930 522 FM Added, mbar 6100 792Change, mbar 65 -67 Change, mbar 69 -68

Change, % 1.1% -11.4% Change, % 1.1% -7.9%

RPM 2000 Power, hp 22.8Torque, ft*lbs 60 Rated Load 1/2

BMEPc FMEPBase Condition, mbar 3875 713FM Added, mbar 3939 645Change, mbar 64 -68

Change, % 1.7% -9.5%

RPM 1500 Power, hp 5.7 RPM 3000 Power, hp 5.7Torque, ft*lbs 20 Rated Load 1/6 Torque, ft*lbs 10 Rated Load 1/12

BMEPc FMEP BMEPc FMEPBase Condition, mbar 1460 523 Base Condition, mbar 853 752FM Added, mbar 1525 458 FM Added, mbar 937 671Change, mbar 65 -65 Change, mbar 84 -81

Change, % 4.5% -12.4% Change, % 9.8% -10.8%

Test Conditions

Test Results

Test Conditions Test Conditions

Test Conditions Test Conditions

Test Results Test Results

Test Results Test Results

400

450

500

550

600

650

700

750

800

Lube A (5W -30) Lube B (20W -50) Lube C (5W -30) Lube D (5W -30)

FM

EP

[m

ba

r]

Neat Lube O il

w/ added Friction Modif ie r

-6%

-9%

-8%

-11%

-14%6%8%Change from

Lube A ->

Effect of Friction Modifier

11

SUMMARY AND CONCLUSIONS

• A modified Cameron-Plint High-Frequency Frictionand Wear Tester was used to evaluate the effect ofengine materials and energy-conserving engine oilson friction and wear of the bore/piston ring interface.

• Friction coefficient data from bench tests correlatedwell with vehicle-dynamometer fuel economy data asmeasured in GM vehicles.

• The Instantaneous IMEP Method has beensuccessfully applied to higher engine speeds andmodern, low-friction engine designs. Comparison ofthe measured cylinder and engine friction figuresgenerated since its development shows greatadvances on the part of engine builders to reducefriction, on the order of 80-85%.

• Based on numerous tests using the InstantaneousIMEP Method, cylinder friction makes up about 20–40% of overall engine friction, depending on speedand load.

• A method has been developed to determine overallengine friction capable of measuring differences inengine friction as small as 0.4%. The method hasbeen successfully used to evaluate the effects offriction modifiers and lube oil formulations.

• Consistent with expectations, measurements confirmthat decreases in friction result in increases of similarabsolute magnitude in power at constant throttleconditions.

• For the friction modifier evaluated, absolute changesin friction levels, resulting from friction modifier use,remained constant over a wide range of enginespeed/load conditions. On a percentage basis,friction reduction was maximized at low speedconditions, while power was maximized at low loadconditions.

REFERENCES

1. Benzing, R., M. Peterson, et.al, “Friction and WearDevices,” ASLE Publication, Park Ridge, Illinois,1976.

2. ASTM G99 Standard: Test Method for Ultra Testingwith a Pin-On-Disk Apparatus, ASTM, PA, 1990.

3. ASTM G77 Standard: Test Method for RankingMaterials to Sliding Wear Using Block-On-Ring WearTester, ASTM, PA,1983.

4. United States Steel (USS) Lubrication EngineersManual, Edited by Charles Bailey and JosephAarons, 1st Edition, 1971.

5. Patterson, D., Hill, S., and S. Tung, ”Bench WearTesting of Engine Power Cylinder Components,”Presented at the ASME Fall Technical Conference,Muskegon, Michigan, 1991.

6. Hartfield-Wunsch, S., Tung, S. and C. Rivard,“Development of a Bench Test for the Evaluation ofEngine Cylinder Components and the Correlationwith Engine Test Results,” SAE 1993 Transactions,Section 3, P. 1131-1138, Paper No. 932693, October15, 1993.

7. Hartfield-Wunsch, S. and S. Tung, “ The Effect ofMicrostructures on the Wear Behavior of ThermalSprayed Coatings,” Reprint from the 1994 7thThermal Spray Conference Proceedings, Boston,Massachusetts, June 20-24, 1994.

8. Fessenden, K. S., Zurecki, Z., and T. P. Slavin, “Thermal Sprayed Coatings: Properties, Processes,and Applications”, ASM Materials Park, Ohio, 1991.

9. Tung, S., and S. Tseregounis, “An Investigation ofTribological Characteristics of Energy-ConservingEngine Oils Using A Reciprocating Bench Test,” Tobe presented at the SAE Spring Fuels andLubricants Conference, June 20, 2000.

10. Tseregounis, S. and M. McMillan, “Engine Oil Effectson Fuel Economy in GM Vehicles - Comparison withthe ASTM Sequence VI-A Engine DynamometerTest,” SAE Paper No. 952347, 1995 SAE Fuels andLubricants Meeting, October 16-19, 1995.

11. McGeehan, J. A., “A Literature Review of the Effectsof Piston and Ring Friction and Lubricating OilProperties,” SAE Paper No. 780673, 1978.

12. Furuhama, S. and M. Takiguchi, “Measurement ofPiston Frictional Force in Actual Operating DieselEngine,” SAE Paper No. 790855, 1979.

13. Furuhama, S. et al., “Effect of Piston and Piston RingDesigns on the Piston Friction Force in DieselEngines,” SAE Paper No. 810977, 1981.

14. Sherrington, I and E. H. Smith, “The Measurement ofPiston-Ring Friction by the ‘Floating-Liner’ Method,”SAE Paper No. 884707, 1988.

15. Kitahara, T. et al., “Studies on the Characteristics ofPiston Ring Friction,” SAE Paper No. 928434, 1992.

16. Uras, H. M. and D. J. Patterson, and “Measurementof Piston and Ring Assembly Friction –Instantaneous IMEP Method,” SAE Paper No.830416, 1983.

17. Uras, H. M. and D. J. Patterson, “Effect of SomeLubricant and Engine Variables on InstantaneousPiston and Ring Assembly Friction”,” SAE Paper No.840178, 1984.

18. Uras, H. M. and D. J. Patterson, “Oil and Ring Effectson Piston Ring Assembly Friction by theInstantaneous IMEP Method,” SAE Paper No.850440, 1985.

19. Uras, H. M. and D. J. Patterson, “Effect of SomePiston Variables on Piston and Ring AssemblyFriction,” SAE Paper No. 870088, 1987.

20. Clampitt, G. D., “The Effects of LightweightReciprocating Components on Engine Friction, HeatTransfer, and Performance,” Master's Thesis, D. N.Assanis (Advisor), University of Illinois, May 1991.