Friction-Reducing Surface-Texturing in ReciprocatingAutomotive Components

AVIRAM RONEN and IZHAK ETSION (Fellow, STLE)Faculty of Mechanical Engineering

TechnionHaifa, 32000, Israel

andYURI KLIGERMAN

Surface Technologies Ltd.Nesher, 36601, Israel

A model is presented to study the potential use of micro-sur-face structure in the form of micro pores to improve tribologicalproperties of reciprocating automotive components. The Reynoldsequation and the equation of motion are solved simultaneously fora simplified piston/cylinder system with surface texturing. Thesolution provides the time behavior of both the clearance and thefriction force between the piston ring and cylinder liner sur-faces. It is shown that surface texturing can efficiently be used to

maintain hydrodynamic effects even with nominally parallel sur-faces. It is also shown that optimum surface texturing may sub-stantially reduce the friction losses in reciprocating automotivecomponents.

KEY WORDSAutomotive; Friction; Hydrodynamic Lubrication; Piston

Rings; Surface Texturing

INTRODUCTIONAbout 40 percent of the total energy developed by a typical

automotive engine is consumed by engine friction (1). Reducingthis friction loss is a key factor in improving fuel consumption andenvironment protection. Of all the various engine components that

NOMENCLATURE

A = contact areac = nominal clearance between mating surfacesC = dimensionless clearance, c/hpF

e= external force acting on the stationary specimen

Ff(t) = instantaneous friction force between the sliding table and the specimen

Ff = average friction forceFh = hydrodynamic opening force acting between the sliding

table and the specimen= dimensionless force, F/(paA)

h = instantaneous local film thicknesshp = pore depthH = dimensionless instantaneous local film thickness, h/hpL

c= connecting rod length

L = axial length of specimenm = specimen massNp = number of poresp = pressurep

a= ambient pressure

pt = pressure above ringpb = pressure below ring

F

P = dimensionless pressure, p/parp = base radius of the porer1 = half side of imaginary square cellr

c= crankshaft radius

Sp = area density of the porest = timeU = sliding velocityW = specimen widthL = specimen lengthx = Cartesian coordinateX = dimensionless Cartesian coordinate, x/rpz = Cartesian coordinateZ = dimensionless Cartesian coordinate, z/rpP = dimensionless pressure differential, Pb-Pt = crankshaft radius over connecting rod length ratio, rc /Lc = pore diameter over crankshaft radius ratio, 2rp/rc1 = bearing number, 3 /pa2 = dimensionless inertia parameter, 2mrppa/(

2 A) = dynamic viscosity = pore depth over diameter ratio, hp/2rp = dimensionless time, tpa/ = crank angular velocity

Presented at the 56th Annual MeetingOrlando, FloridaMay 20-24, 2001

Final manuscript approved April 11, 2001Review led by Luis San Andrs

359

TRIBOLOGY TRANSACTIONSVol. 44 (2001), 3, 359-366

360 A. RONEN, I. ETSION AND Y. KLIGERMAN

Fig. 1Simple model of piston rings and cylinder liner simulation.

contribute to the friction losses the piston and piston ring systemaccounts for about 50-60 percent. It is not surprising therefore thatengine friction in general and the friction of the piston/cylindersystem in particular was the focus of many research work, e.g. (2)-(4).

Proper lubrication and surface roughness are key issues inreducing friction in the piston/cylinder system and, hence,received great deal of attention in the relevant literature. Some ofthe previous studies (5) and (6), which predicted the oil film thick-ness between a piston ring and cylinder liner, assumed that theliner and the ring surfaces were smooth. The first to develop a pis-ton ring lubrication model, which included surface roughnesseffect, was Rohde (7), in 1980. Sanda and Someya (8) examinedtheoretically and experimentally the effect of surface roughnesson lubrication between a piston ring and a cylinder liner regardingfour different magnitudes and directions of roughness. It wasfound that the effect of the roughness is significant only near thetop and bottom dead centers when the oil film is thinner. Higherfriction peaks were observed in the case of larger (high RMS) andlongitudinal roughness (asperities ridges parallel to the slidingdirection) than in the case of smaller (low RMS) and transverseroughness.

Michail and Barber (9), (10) developed a theoretical pistonring/cylinder wall model, based on the average Reynolds equationdeveloped by Patir and Cheng (11), to study the effects of cylin-der wall roughness, crosshatched angle and plateau honing on theoil film thickness between a piston ring and a cylinder wall. It wasfound that for a given set of operating parameters, the oil filmthickness is largest for transversely oriented surfaces. This corre-sponds to crosshatch angles of less than 45 degrees for the honedsurface considered. It was also found that for equal roughness, theoil film thickness is largest for unskewed surfaces. From the

above it is clear that microstructure of the sliding surface plays animportant role in friction control. This is probably due to themicrostructure effect on the build up of hydrodynamic fluid filmbetween the mating surfaces.

Anno, et al. (12), (13) show that the use of micro asperities isan effective and controllable technique for obtaining hydrody-namic operation in face seals. Hamilton, et al. (14) demonstratehow micro irregularities and cavities can be intimately involved inthe hydrodynamic process between two parallel surfaces andEtsion, et al. (15) show similar effects while using spherical pores.Burstein and Ingman (16) applied the concept of pore ensemble topiston ring lubrication. They solved the case of reciprocatingmotion of a compression piston ring with cylindrical microporeswhile assuming that the distance between neighboring pores islarge enough to neglect the interaction between them.

The present paper presents a novel idea with a potential toreduce the friction force between the piston rings and cylinderliner. For simplicity the piston/cylinder system is modeled by twonominally flat surfaces simulating segments of two piston ringsand reciprocating relative to a flat (see Fig. 1) that simulates thecylinder liner. The piston ring faces are laser textured to containmicro-pores that act as micro hydrodynamic bearings to enhancehydrodynamic lubrication (Etsion, et al. (15)). This model,although very simplistic, was selected to permit simple presenta-tion of the potential of the novel idea and to allow a comparisonof the theoretical results with experimental ones that will beobtained on a test rig very similar to the schematic description ofFig. 1.

ANALYTICAL MODELThe geometrical model of the two piston ring segments sur-

faces is displayed in Figs. 2(a) and 3. The pores are uniformly dis-tributed, with an area density Sp. Each pore is modeled by anaxisymmetric spherical segment with a base radius rp and depth hpand is located in the center of an imaginary square cell of sides 2r1x 2r1 (see Fig. 2(a)) where:

[1]

Assuming a slider-crank mechanism, as shown schematicallyin Fig. 4, the sliding velocity U corresponding to the crankshaftangle may be expressed as (Mabie and Ocvirk, (17)):

[2]

where () is a trigonometric function given by:

[3]

( ) =

+

sin

sin

sinp

p

pa

a

a

2

2 1 2 2

U rc= ( )

rr

Sp

p1 2

=

Friction-Reducing Surface-Texturing in Reciprocating Automotive Components 361

where = pat/ is a dimensionless form of the time. The two-dimensional, time dependent form of the Reynolds equation for anincompressible Newtonian fluid in laminar flow is given by:

[4]

where z and x are lateral and longitudinal direction Cartesiancoordinates, respectively, and h (see Fig. 3) is the local film thick-ness at a specific point of the textured surface:

Here c(t) is the time dependent nominal clearance between themating surfaces (see Fig. 3); and (see Fig. 2(a)) are the localCartesian coordinates for a single pore cell. In order to reduce Eq.[4] to nondimensional form the nondimensional Cartesian coordi-nates X and Z, nondimensional local film thickness H, and nondi-mensional pressure P are defined as:

h c t r

h c th r

hr h

hr

p

p p

p

p p

pp

= ( ) +

= ( ) + + +( )

+ 12 it is possible to save computation time by considering Np = 12only. This phenomenon was found also valid for the other valuesof 1 in Table 1 and can be explained by the vanishing influenceof the ends of the column as its length, L, increases. Above a cer-tain number of pores the effect of the boundary pressures at the

two ends of the specimen is confined to a few pores close to theseends. The rest of the pores further away from the ends generate thesame pressure distribution as their neighbors. Hence, the averagepressure over these pores as well as the resulting clearance andfriction force become independent of Np.

An interesting finding in Fig. 9 is the existence of an optimum value that minimizes the average friction force. This demon-strates the potential of friction reduction with the surface texturingas compared to non-textured surfaces where = 0. For example,for Np=24 there is a reduction of 28% in the dimensionless aver-age friction force as is increased from = 0.02 (which is closerto non-textured surfaces) to its optimum value of = 0.14. Itshould be noted that the present simple model with the flat andparallel piston rings and cylinder liner of Fig. 1 cannot beused with = 0 since no hydrodynamic load capacity can beobtained in this case. However, as can be seen from Fig. 9 the fric-tion force gradient increases rapidly at very small values of as deceases. Hence, the actual friction reduction with the surface tex-turing as compared to totally non-textured surfaces is expected tobe much higher than 28%.

Figure 11 presents the effect of the bearing number, 1, on thedimensionless average friction force for a wide range of values.As can be seen the average friction force at a given increaseswith 1. Since 1 depends linearly on the crank angular velocity,, and the lubricant viscosity, , (see Eq. [7]) the increase in thefriction force with 1 is expected. Again the existence of an opti-mum value that minimizes the average friction force can be seen.This optimum value depends on the bearing number 1 and

TABLE 1VALUES OF DIMENSIONLESS PARAMETERS OF THE PROBLEM CONSIDERED FOR THE ANALYTICALINVESTIGATION

DIMENSIONLESS RANGE OF VARIATIONPARAMETER LOWER LIMIT REFERENCE CASE UPPER LIMIT

Np 6 12 24Sp 5% 10% 20%

2.5 5 101 5.0E-4 1.0E-3 2.0E-32 5.0E+5 1.0E+6 2.0E+6 0.001 0.002 0.004

Fe

Fig. 9Dimensionless average friction force, vs. depth over diameterratio, , for various pores number, Np.

Fig. 10Dimensionless average friction force, , vs. number ofpores, Np.

Ff ( )

364 A. RONEN, I. ETSION AND Y. KLIGERMAN

varies from about 0.1 at 1 = 0.5E-3 to about 0.18 at 1 = 2.0E-3. Figure 12 presents the effect of the pore diameter over crank-

shaft radius ratio, , on the dimensionless average friction forcefor the range of . As can be seen the average friction force at agiven increases with a decrease of . Since depends linearly on1/rc, Eq. [7], and because the sliding velocity, U, depends linearlyon rc (see Eq. [2]) the increase in the friction force with thedecrease of is expected. The optimum value again is clearlyshown.

Figure 13 presents the effect of the external force, , on thedimensionless average friction force. As expected the averagefriction force at a given increases with since this force is act-ing to reduce the average clearance between the mating surfacesand thus increases the friction force (see Eq. [14]). As in Figs. 9,11 and 12 the optimal pore depth over diameter ratio is clearlydemonstrated.

CONCLUSIONA model was developed to demonstrate the potential of reduc-

ing friction losses in reciprocating systems by surface texturing inthe form of multiple spherical micro-pores. A simplified pis-

Fe

Fe

ton/cylinder system with a constant external normal force wasanalyzed demonstrating that hydrodynamic effects can be gener-ated by surface texturing even for nominally parallel mating sur-faces. The time variation of the clearance and the friction force atany given operating conditions were obtained by simultaneouslysolving the Reynolds equation and a dynamic equation.

The main parameters of the problem were identified and anintensive parametric investigation was performed. The followingconclusions summarize the outcome of the investigation:

1. The interaction between adjacent pores is significant andhence, its effect on the hydrodynamic pressure distributioncannot be neglected.

2. The maximum variation of the fluid film clearance over onerevolution is small being less than 30 percent of the poresdepth. Hence, the variation in the friction force with thecrank angle is mainly due to the variation in the slidingspeed.

3. A change of the area density of the pores, Sp, in the rangebetween 5% to 20% changes the friction force by less than7%.

4. The effect of the inertia of the piston rings is negligible.Changing the inertia parameter by 3 orders of magnitudecauses only 2% change in the average friction force.

5. Increasing the number of pores over the axial length of thepiston ring reduces the friction sharply up to Np = 12.Thereafter the reduction in friction becomes much moremoderate.

6. An optimum value of the pore depth over diameter ratio was found, which yields a minimum friction force. Thisoptimum ratio varies from about = 0.1 to about = 0.18for all the relevant parameters of the problem.

7. Although the model does not allow direct comparison with asystem that does not include the pores (the case of = 0) itis clear that a friction reduction of 30% and even more isfeasible with textured surfaces.

ACKNOWLEDGMENTThe research reported here was supported in parts by Surface

Technologies Ltd., by the GM-UMI Foundation and by the

Fig. 11Dimensionless average friction force, , vs. depth overdiameter ratio, , for various bearing number, 1.

Ff ( )

Fig. 12 Dimensionless average friction force, Ff, vs. depth over diameter ratio, , for various pore diameter over crankshaftradius ratio number, .

Fig. 13 Dimensionless average friction force, Ff, vs. depth over diameter ratio, , for various external force, F

e .

Friction-Reducing Surface-Texturing in Reciprocating Automotive Components 365

German-Israeli Project Cooperation (DIP). This support as wellas the permission by Surface Technologies to publish the work isgratefully acknowledged by the authors.

REFERENCES

(1) Nakada, M., Trends in Engine Technology and Tribology, Trib. Intl., 27, pp3-8, (1994).

(2) Knoll, G. D and Peeken, H. J., Hydrodynamic Lubrication of Piston Skirts,Trans. ASME Jour. of Lubr. Tech., 104, pp 504-509, (1982).

(3) Mitsuru, H. and Yasukazu, B., A Study of Piston Friction Force in an InternalCombustion Engine, ASLE Trans., 30, pp 444-451, (1987).

(4) Takiguchi, M., Machida, K. and Furuhama, S., Piston Friction Force of A SmallHigh Speed Gasoline Engine, Trans. ASME Jour. of Trib., 110, pp 112-118,(1988).

(5) Ting, L. L. and Mayer, J. E., Piston Ring Lubrication and Cylinder Bore WearAnalysis, Part I Theory, Trans. ASME Jour. of Lubr. Tech., 86, pp 305-314,(1974).

(6) Wakuri, Y., Hamatake, T., Soejima, M. and Kitahara, T., Piston Ring Frictionin Internal-Combustion Engines, Trib. Intl., 25, pp 299-308, (1992).

(7) Rohde, S. M., A Mixed Friction Model for Dynamically Loaded Contacts withApplication to Piston Ring Lubrication, AGARD Lecture Series, SurfaceRoughness Effects in Hydrodynamic and Mixed Lubrication, ASME, New York,pp 19-50, (1980).

(8) Sanda, S. and Someya, T., The Effect of Surface Roughness on LubricationBetween a Piston Ring and a Cylinder Liner, in Proc. of the Institution ofMechanical Engineers, Intl. Conf., Tribology-Friction, Lubrication and Wear,IMechE, 1, pp 135-143, (1987).

(9) Michail, S. K. and Barber, G. C., Effects of Roughness on Piston RingLubrication Part I: Model Development, Trib. Trans., 38, pp 19-26, (1995).

(10) Michail, S. K. and Barber, G. C., The Effects of Roughness on Piston RingLubrication Part II: The Relationship between Cylinder Wall Surface

Topography and Oil Film Thickness, Trib. Trans., 38, pp 173-177, (1995).(11) Patir, N. and Cheng, H. S., An Average Flow Model of Determining Effects of

Three-Dimensional Roughness on Partial Hydrodynamic Lubrication, Trans.ASME Jour. of Lubr. Tech., 100, pp 12-16, (1978).

(12) Anno, J. N., Walowit, J. A. and Allen, C. M., Microasperity Lubrication,Trans. ASME Jour. of Lubr. Tech., 91, pp 351-355, (1968).

(13) Anno, J. N., Walowit, J. A. and Allen, C. M., Load Support and Leakage fromMicroasperity-Lubricated Face Seals, Trans. ASME Jour. of Lubr. Tech., 90, pp726-731, (1969).

(14) Hamilton, D. B., Walowit, J. A. and Allen, C. M., A Theory of Lubrication byMicroirregularities, Trans. ASME Jour. of Basic Eng., 88, pp 177-185, (1966).

(15) Etsion, I., Kligerman Y. and Halperin, G., Analytical and ExperimentalInvestigation of Laser-Textured Mechanical Seal Faces, Trib. Trans., 42, pp511-516, (1999).

(16) Burstein, L. and Ingman, D., Pore Ensemble Statistics in Application toLubrication Under Reciprocating Motion, Trib. Trans., 43, pp 205-212,(2000).

(17) Mabie, H. H. and Ocvirk, F. W., Mechanisms and Dynamics of Machinery, 3rdEd., John Wiley and Sons Inc., New York, pp 19-20, (1975).

(18) Dowson, D. and Taylor, C. M., Cavitation in Bearings, Ann. Rev. Fluid Mech.,11, pp 35-66, (1979).

(19) Gerald, C. F. and Wheately P. O., Applied Numerical Analysis, 5th ed., Addison-Wesley Publishing Co., New York, pp 157-159, (1994).

(20) Cho, S. -W., Choi, S. -M. and Bae, C. -S., Frictional Modes of Barrel ShapedPiston Rings Under Flooded Lubrication, Trib. Intl., 33, pp 545-551, (2000).

366 A. RONEN, I. ETSION AND Y. KLIGERMAN