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Factors Influencing Boundary Friction and Wear of Piston Rings
Thinning Films and Tribological Interfaces / D. Dowson et al. (Editors) © 2000 Elsevier Science B.V. All rights reserved.
409
Factors Influencing Boundary Friction and Wear of Piston Rings M. Priest School of Mechanical Engineering, University of Leeds, Leeds, LS2 9JT, UK
Models have been developed for asperity friction and the rate of wear of piston tings in the boundary and mixed lubrication regimes. However, the surface interactions occurring between the piston rings, piston and cylinder bore in the presence of the engine lubricant with its complex chemistry, contamination and degradation behaviour are extremely difficult to interpret in terms of basic boundary friction and wear mechanisms. This paper aims to assist the analyst by reporting a series of friction and wear experiments on a high frequency reciprocating tribometer using specimens manufactured from real components in the presence of engine lubricant at speeds, loads and temperatures characteristic of modern automotive gasoline engines. Data is presented for the critical tribological interfaces between the piston ring face and the cylinder wall and the piston ring flank and the piston groove. The results yield a useful matrix of asperity friction coefficients and wear factors and clearly highlight the sensitivity of the wear factor to the selection of the materials of both surfaces. Measured wear factors at the piston ring flank / piston groove interface were much higher than at the piston ring face / cylinder liner interface. Furthermore, wear factors for the critical top piston ring / piston groove interface were found to be the highest and increased markedly with rising temperature.
NOTATION
AW
hlllill k ko V W X Xs
k ~t t3
cross-sectional area of the wear scar of the cylinder liner sample in the direction of sliding at the approximate circumferential centre of the contact with the ring mininmm lubricant film thickness wear factor wear factor in the boundary lubrication regime worn volume normal load stroke sliding distance specific film thickness root mean square coefficient of friction composite root mean square roughness of two interacting surfaces
1. INTRODUCTION Piston rings are critical components in the modem internal combustion engine, their optimum tribological performance having a controlling influence in minimising frictional power loss, fuel consumption, oil consumption, blow-by and h a m ~ l
exhaust emissions. The basic roles of the piston ring pack, the collective name for the three or more rings mounted on the piston, are to form an effective gas seal between the combustion chamber and the crankcase, to limit the upward transport of oil from the crankcase to the combustion chamber and to provide a flow path for heat transfer from the piston to the engine coolant. All this must be achieved with minimum frictional power loss, most notably at the sliding interface with the cylinder wall, and the least possible wear in order to maximise component life. Unfortunately, the piston ring pack is one of the largest contributors to friction in the internal combustion engine over the normal range of engine speeds and loads encountered in service [1-3]. The relative importance varies from engine to engine, but typically the piston assembly, comprising both the piston rings and the piston skirt, accounts for 4050% of total engine friction [4]. In terms of wear, there is insufficient understanding of the interaction with the lubrication process. So, even though manufacturers can produce rings that have an excellent life expectancy, these components may be far from optimum from a lubrication and friction standpoint. As a consequence of their importance to engine
410
performance, the theoretical and experimental study of piston ring lubrication has received much attention in the literature, leading to major advances in our understanding of their behaviour [5]. The mathematical analysis of piston ring lubrication is complex and by necessity requires simplifying assumptions. However, rapid development in numerical methods over the last thirty years has resulted in sophisticated piston ring lubrication models that are finding application in the design process [6]. Figure 1 shows a typical prediction of specific film thickness, defined as the ratio of the minimum lubricant film thickness to the composite root mean square surface roughness ~, = hmi" (3-
(1)
throughout the engine cycle for the top compression ring of a diesel engine from Priest et al [7], where zero degrees crank angle is top dead centre firing. ,10~
boundary/mixedtransition
]
relatively well understood but the surface interactions that take place present much more of a challenge. Models have been developed for asperity friction [8] and the rate of wear [7] of piston tings in the boundary and mixed lubrication regimes. However, the surface interactions occurring between the piston rings, piston and cylinder bore in the presence of the engine lubricant with its complex chemistry, contamination and degradation behaviour are extremely difficult to interpret in terms of basic boundary friction and wear mechanisms [9]. Such models, therefore, are empirical and assume a variation of boundary friction coefficient or wear factor with specific film thickness. Both are zero in the fluid film lubrication regime and increase through the mixed lubrication regime, as the degree of surface contact increases, to a maximum in the boundary lubrication regime. Figure 2 shows an example for the dimensional wear factor, k, after Priest [ 10]. LUBRICATION REGIMES mixed
boun [ary
fluid film
ko 6
{
4 .................................................
0 0
90
180 270 360 450 540 Crank Angle (degrees)
630
0 0.5
1
Figure 1. Cyclic variation of specific film thickness between a top compression ring and the cylinder wall [7] It can be seen from Figure 1 that the piston ring experiences the full range of lubrication regimes in one cycle, from full fluid film through mixed lubrication to boundary, lubrication. This is characteristic of the interface between the piston ring and the cylinder wall. The fluid film aspects of this behaviour are
2
3
4
~(-)
720
Figure 2. Variation of wear factor with specific film thickness [ 10] The wear factor, k, is related to the volume of material removed by wear, V, the applied normal load, W and the sliding distance, xs, by v = kWx.
(2)
as proposed by Lancaster [11 ]. Such models are clearly very sensitive to the values chosen for the boundary lubrication regime,
411
ko in Figure 2, and the situation is further complicated by the fact that the specific film thickness tends to be influenced by changes in surface roughness as the components run-in. The discussion thus far has considered only the conditions between the piston ring face and cylinder wall but the interface between the piston ring flank and the piston groove is also important in terms of engine reliability and durability [12]. This interface has received relatively little attention in the literature but is likely to experience much more mixed and boundary lubrication than the ring face. This paper aims to assist the analyst in the selection of appropriate values for asperity friction coefficients and boundary wear factors by reporting a series of friction and wear experiments on a high frequency reciprocating tribometer. Tests were undertaken on piston ring, piston and cylinder specimens manufactured from real components in the presence of engine lubricant at speeds, loads and temperatures characteristic of a modem automotive gasoline engine. Results are presented for the critical tribological interfaces between the piston ring face and the cylinder wall and the piston ring flank and the piston groove. The experiments were undertaken on a complete ring pack from a single engine with the aim of establishing the relative importance of the various interfaces in the ring pack. The results provide a useful matrix of asperity friction coefficients and wear factors for the designer/analyst and highlight the sensitivi .ty of the wear factor to the selection of the materials of both interacting surfaces. In addition, wear factors at the piston ring flank / piston interface are shown to be much higher than at the piston ring face / cylinder interface.
components. The design and dimensions of the piston rings are shown in Figure 3. The top compression ring, ring 1, is a barrel faced, spheroidal graphite cast iron ring (150GPa minimum modulus of elasticity) with a full-face, flame sprayed molybdenum wear resistant coating (650-950 HV hardness) on the ring face. The surface of this coating has an open porous structure. Ring 2, the second compression ring, is a plain grey cast iron, Napier scraper ring (85-115 GPa modulus of elasticity, 96-106 HRB, 210-290 lib hardness) with a fine turned finish on the nmning face but no wear resistant coating. The oil-control ring, ring 3, is a steel rail, three-piece, oil-control ring (207GPa modulus of elasticity) with a chromium plating wear resistant coating on the faces of the rails (700-900 HV hardness).
3.56mm ring 1
1.50mm
3.54 mm
2.68mm 0.52mm ~
I3.00mm*
0.52mm m ( 2.68mm Figure 3. Ring pack geometry
2. METHOD 2.1. Test Specimens
The experiments were undertaken on the piston assembly components from a single-cylinder Ricardo Hydra gasoline engine. This is a 0.5 litre capacity engine of 86mm bore and 86mm stroke. Although a research engine, it is based on a General Motors two litre, four-cylinder production engine and the piston and ring pack are standard production
The piston is manufactured from a hyper-eutectic aluminium silicon alloy (56-78 GPa modulus of elastici .ty, 114 BHN hardness). The Hydra engine has a specially manufactured wet liner as the General Motors four-cylinder equivalent engine has parent metal bores. The liner is machined from grey cast iron and the surface treatment is a 450-65 ° crosshatch plateau hone. In the current experiments, however, a different,
412
similarly honed, cylinder liner of 90mm bore diameter was used. This was selected in order that small sections of the piston rings cut from near the ring gaps had good circumferential conformity with the liner. This allowed small ring samples to be used in the tests, giving more accurate assessment of wear by weight loss. The chemical composition and physical properties of the two liners is very similar as shown in Table 1.
specimen and a stationary specimen from a scotch yoke mechanism with a maximum stroke of 15mm and a maximum frequency of 50Hz. Constant loads of up to 500N can be applied and the temperature of the specimens can be controlled from ambient up to 600°C. Friction is measured by a load cell, which restrains the stationary specimen in the direction of sliding. 2.3. Test Conditions
Table 1. Comparison of cylinder liner materials Hydra Engine Total Carbon (%) Silicon (%) Sulphur (%) Phosphorus (%) Manganese (%) Chronfium (%) Tensile Strength (tons per square inch) Hardness (BHN)
current experiments
3.0-3.4 1.8-3.0 0.12 max 0.5-1.0 0.7-1.2 0.35-1.0 16
3.0-3.5 1.8-2.5 0.10 max 0.4-0.6 0.6-1.0 0.3-0.6 15
230-300
230-275
Boundary lubrication is most likely to occur in the first half of the power stroke when gas pressures and temperatures are high and, in the case of the piston ring and cylinder liner, sliding and lubricant entrainment velocities are low. This can be clearly seen in the predicted film thickness behaviour for the top compression ring of Figure 1, the first half of the power stroke being from 0 ° to 90 °. The operating parameters for the reciprocating experiments were therefore based on this region using data for the running conditions of the piston ring lubrication and wear experiments under fired conditions undertaken on the Ricardo Hydra gasoline engine [ 10]. 2.3.1
For the experiments, 30mm long sections of piston ring were cut from all three ring types adjacent to the ring gaps. Rectangular sections of cylinder liner, 58mm (axial) and 38mm (circumferential), were prepared. The piston samples were 8mm diameter pins with a 50mm crown radius on one end, machined from the crown of the piston. The lubricant was a fully formulated SAE 30 monograde engine oil with a performance level of API SF/CD. This was selected for consistency with piston ring lubrication and wear experiments under fired conditions undertaken on the Ricardo Hydra gasoline engine [ 10].
Piston ring and cylinder liner
The cylinder liner section was held stationary in a shallow bath which contained suffzcient lubricant to submerge the interface between the ring and liner. The piston ring section was held perpendicular to the liner and reciprocated as illustrated in Figure 4. The stroke, x, was 5mm and frequency 42Hz, giving a mean sliding speed of 0.42m/s.
;
Z
2.2. Test Machine
The experiments were undertaken on a standard Cameron Plint TE77 high frequency reciprocating friction and wear machine. This apparatus is highly suited to piston ring wear studies, providing a simple harmonic reciprocating motion between a moving
Figure 4. Piston ring and cylinder liner test configuration
413
Tests of two hours duration were undertaken at normal load, W, of 100N and 200N and bulk oil temperatures of 100°C and 200°C.
2.3.2
Piston ring and piston
Figure 5 shows the test configuration with a stationary ring section and reciprocating piston sample in the form of a pin with a crown end radius. Again there was sufficient lubricant in the bath to submerge the interface. The stroke, x, was 2mm and the frequency 42Hz, yielding a mean sliding speed of 0.17m/s. The reduced sliding speed when compared to the ring and liner tests reflects the much more restricted motion of the ring within the groove. One practical consequence of this change in the test procedure was that an increased test duration of three hours was required to achieve significant wear.
10 mm
W
Wear of the piston ring and piston samples was determined by weighing the samples before and after testing using an mass balance with an accuracy of one microgram. Specimens were ultrasonically cleaned in acetone before weighing. The liner specimens could not be treated in the same way because an oil lacquer tended to develop on the surface outside the region of sliding during the test. Wear could therefore not be correlated with mass loss due to the mass increase caused by the lacquer. An alternative approach is to determine wear by volume change measured by profilometry. Three-dimensional profilometry to determine volume change is complex and time-consuming for curved surfaces such as the liner. It was therefore not considered for these relatively simple wear tests, especially as piston ring wear is more significant for the tribological performance of the interface. Wear of the liner specimens was alternatively determined on a relative basis by two-dimensional profilometry. The cross-sectional area of the wear scar, Aw, in the direction of sliding at the approximate circumferential centre of the contact with the ring was selected as an indication of the severity of wear. Measurements were taken using a Rank Taylor Hobson Fornl Talysurf stylus profilometer with a vertical resolution of 0.01 ~tm. The depth of the wear scars developed during the tests was generally in the range 1~un to 8~tm.
Figure 5. Piston ring and piston test configuration Tests were undertaken with a normal load, W, of 100N for rings 1 and 2 and 70N for ring 3, reflecting the fact that it is more lightly loaded against the piston groove in the engine. Bulk oil temperatures of 100°C, 150°C, 200°C and 300°C were investigated. It should be noted that the loads, speeds and temperatures at the piston ring flank and piston groove interface are far less well understood than the ring and liner [12]. The operating conditions used in the experiments therefore represent a best estimate of the conditions experienced in the engine.
2.4. Measurement Techniques Speed, load, bulk oil temperature and friction force were monitored throughout the test period using a computer based data acquisition system.
3. RESULTS
3.1. Piston Ring and Cylinder Liner The influence of bulk oil temperature on wear for all three ring .types sliding against the cylinder liner is shown in Table 2. The chromium plated steel oil-control ring, ring 3, exhibited the lowest wear factors closely followed by the flame sprayed molybdenum coated, spheroidal graphite cast iron top ring, ring 1. By far the highest wear factors are associated with the plain grey cast iron scraper ring, ring 2. In contrast to tings 1 and 3, the wear factor of ring 2 was also sensitive to the imposed temperature change. The data reported for cylinder liner wear needs to be considered with a degree of caution given the simplistic approach. Nevertheless, it can be observed from Table 2 that the most severe wear
414
occurred when ring 3 was the counterface. The 200N applied load was, however, optimised for the compression tings and is much higher than this ring would see in service. More interesting is the reduction in worn area at 200°C for ring 1. Table 2. Effect of bulk oil temperature on piston ring and cylinder liner wear at 200N normal load piston ring, ko liner, Aw (10 -11 mm3mm-lN -1) (!03 ~tm2)
ring 1 ring 2 ring 3
100°C
200°C
100°C
200°C
2.75 4.95 1.63
2.43 8.32 1.40
5.9 6.9 20.4
1.7 6.4 21.6
Table 3 shows the results at 200°C for all three ring .types at two different applied loads. Very. similar trends to Table 2 are evident for each ring .type and the cylinder liner. Table 3. Effect of load on piston ring and cylinder liner wear at 200°C bulk oil temperature Piston ring, ko liner, Aw (10 -11 mm3mm-lN -1) (103 ~tm2)
ring 1 ring 2 ring 3
100 N
200 N
100 N
200 N
2.27 4.50 0.93
2.43 8.32 1.40
2.1 3.1 16.8
1.7 6.4 21.6
Table 4 summarises the frictional behaviour in terms of root mean square (r.m.s.) coefficient of friction throughout the test period. Very. little variation is observed but more interestingly the coefficients are rather low for boundary, lubrication with the surface materials and lubricant under study, which did not contain a friction modifier additive. Normally, a coefficient of friction of the order of 0.12 would be expected for such contacts. This suggests the experiments were operating, for at least
part of the stroke, with severe mixed lubrication rather than boundary lubrication. The omissions in Table 4 were due to measurement equipment faults experienced during these experiments. Given the lack of variation in coefficient of friction, these experiments were not repeated. Table 4. Variation of r.m.s, coefficient of friction between piston ring and cylinder liner with load and bulk oil temperature
~t (-) 200 N
ring 1 ring 2 ring 3
200°C
100°C
200°C
100 N
200 N
0.09 0.10
0.08 0.07
0.09 0.09 0.09
0.08 0.07
3.2. Piston Ring and Piston The emphasis in this sequence of experiments was placed on the top ring, ring 1, as it experiences the highest loads and temperatures at the piston ring flank / piston groove interface. The influence of increasing temperature on wear factor and r.m.s, coefficient of friction for ring 1 is given in Table 5. As a matter of information, it was found necessary, to replenish the lubricant in the bath during the test at 300°C. Table 5. Effect of bulk oil temperature on piston and top piston ring wear and friction ko (10 -9 mm3mmlN 1) piston 100°C 150°C 200°C 300°C
0.56 2.92 7.27 13.19
ring 1 0.24 0.48 1.22 3.69
~t(-) 0.07 0.11 0.13 0.13
415
First of all, it should be noted that the wear factors are generally two orders of magnitude larger than observed in the piston ring / cylinder liner interface. There is also a rapid increase in wear factors for both the piston and piston ring with increasing temperature. The data is shown graphically in Figure 6 with linear least squares fit lines through the experimental data points. 15-
•
piston
,.-,, 10"7
,
o
0 50
100
150
200
250
300
350
bulk oil temperature (°C) Figure 6. Variation of wear factor with temperature for the piston sliding against the top ring flank
The wear factors for both the piston and ring 1 basically increase linearly with temperature. This has potentially serious implications for current and future engine designs, where the trend is for higher temperatures in the top ring groove of the piston. The r.m.s, coefficient of friction between the piston and top ring flank reported in Table 5 also increases with temperature up to 200°C where it appears to a reach a plateau at a value of 0.13. This is shown graphically in Figure 7. The friction behaviour for the other two tings at 200°C is reported in Table 6 along with the wear factors. The values for ring 1 are included to aid comparison. The coefficients of friction are very similar for the three ring types and of a magnitude consistent with boundary lubrication. The wear factors for the different ring types are, however, interesting in terms of their ranking. The top ring flank / piston groove, which is the most critical in the engine in terms of operation, exhibits the largest wear factors. There appears to be a compatibility problem between the spheroidal graphite cast iron ring, the piston alloy and the lubricant. The plain cast iron second ring, ring 2, produced lower wear factors, followed by the steel oil-control ring, ring 3. Table 6. Piston and piston ring wear and friction at 200°C bulk oil temperature
0.15
ko (10 .9 mm 3mm-1N1)
,/'
piston ring 1 ring 2 ring 3
0.10::t.
7.27 3.15 2.35
ring
~t (-) 1.22 0.65 0.30
0.13 0.11 0.12
4. DISCUSSION 0.05
50
160
1;0
260
2;0
360
3;0
bulk oil temperature (°C) Figure 7. Variation of r.m.s, coefficient of friction with temperature for the piston sliding against the top ring flank
As stated at the outset, the intention of these experiments was not to investigate the complex lubrication, friction, wear and tribochemical processes that occur in piston assemblies. Rather, the aim was to observe trends in basic boundary friction coefficients and wear factors. The experiments were undertaken on a complete ring pack from a single gasoline engine to evaluate the
416
relative importance of the various interfaces in the ring pack. The experiments on the piston ring face / cylinder liner interface show a clear sensitivity to the materials in terms of wear but little sensitivity in terms of the r.m.s, coefficient of friction. Measured wear factors at the piston ring flank / piston interface were found to be much higher than at the piston ring face / cylinder interface. This highlights the sensitivi.ty of the wear factor to the selection of the materials of both surfaces. However, the sliding distances at piston ring flank / piston groove interface are much less than for the typically boundary, or mixed lubricated proportion of the piston ring face / cylinder liner stroke. Hence the volumetric wear, V in equation (2), at the piston ring flank / piston groove interface will not be as large as expected from a simple examination of the wear factors. Wear at the piston ring flank / piston groove interface was also found to be sensitive to material type. Most worryingly, wear for the critical top piston ring and piston was found to be the highest and increased markedly with rising temperature. 5. CONCLUSIONS •
Simple bench tests for all the critical interfaces in the piston ring pack for a single engine have yielded a useful matrix of asperity friction coefficients and wear factors for the designer / analyst.
•
The sensitivity of the wear factor to the selection of the materials of both surfaces is clear.
•
Wear factors at the piston ring flank / piston interface are much higher than at the piston ring face / liner interface.
•
Wear for the critical top piston ring / piston groove interface was found to be the highest and increased markedly with rising temperature.
REFERENCES
1. Monaghan, M.L., "Engine friction- a change in emphasis", Instn. Mech. Engrs., 2nd BP Tribology Lecture, 1987.
2. Monaghan, M.L., "Putting friction in its place", 2nd Int. Conf.: Combustion Engines- Reduction of Friction and Wear, IMechE conf. pub. 1989-9, Paper C375/KN1, 1989, pp. 1-5. 3. Parker, D.A. and Adams, D.R., "Friction losses in the reciprocating internal combustion engine", Tribology - Key to the Efficient Engine, Instn. Mech. Engrs. Conf. Pub. 1982-1, Paper C5/82, 1982, pp. 31-39. 4. Gazzard, S.T., et al., "Piston system design for low emissions", Leading Through Innovation, T&N Symposium 1995, Paper 20, 1995. 5. Dowson, D., "Piston assemblies; background and lubrication analysis", Engine Tribology, C.M. Taylor (ed.), Tribology Series, 26, Elsevier, Ch. 9, 1993, pp.213-240. 6. Mierbach, A., et al., "Piston ring performance modelling", Leading Through Innovation, T&N Symposium 1995, Paper 15, 1995, pp. 15.115.13. 7. Priest, M., Dowson, D. and Taylor, C.M., "Predictive wear modelling of lubricated piston tings in a diesel engine", Wear, 231, 1, 1999, pp.89-101. 8. Rohde, S.M., "A mixed friction model for dynamically loaded contacts with application to piston ring lubrication", Surface Roughness Effects in Hydrodynamic and Mixed Lubrication, ed. S.M. Rohde and H.S. Cheng, ASME pub., 1980, pp. 19-50. 9. Coy, R.C., "Practical applications of lubrication models in engines", New Directions in Tribology, MEP, 1997, pp. 197-209. 10. Priest, M., "The wear and lubrication of piston tings", PhD thesis, University of Leeds, October. 1996, pgs. 246. l l. Lancaster, J.K., "Dry bearings : a survey of materials and factors affecting their performance", Tribology, December, 1973, pp.219-251. 12. Barrell, D.J.W., Priest, M. and Taylor, C.M., "Bench test study of piston ring flank and piston groove interaction", Lubrication at the Frontier, Proc. 25th Leeds-Lyon Symposium on Tribology, Lyon, 1998, Elsevier, Amsterdam, 1999, pp.343-351.
409
Factors Influencing Boundary Friction and Wear of Piston Rings M. Priest School of Mechanical Engineering, University of Leeds, Leeds, LS2 9JT, UK
Models have been developed for asperity friction and the rate of wear of piston tings in the boundary and mixed lubrication regimes. However, the surface interactions occurring between the piston rings, piston and cylinder bore in the presence of the engine lubricant with its complex chemistry, contamination and degradation behaviour are extremely difficult to interpret in terms of basic boundary friction and wear mechanisms. This paper aims to assist the analyst by reporting a series of friction and wear experiments on a high frequency reciprocating tribometer using specimens manufactured from real components in the presence of engine lubricant at speeds, loads and temperatures characteristic of modern automotive gasoline engines. Data is presented for the critical tribological interfaces between the piston ring face and the cylinder wall and the piston ring flank and the piston groove. The results yield a useful matrix of asperity friction coefficients and wear factors and clearly highlight the sensitivity of the wear factor to the selection of the materials of both surfaces. Measured wear factors at the piston ring flank / piston groove interface were much higher than at the piston ring face / cylinder liner interface. Furthermore, wear factors for the critical top piston ring / piston groove interface were found to be the highest and increased markedly with rising temperature.
NOTATION
AW
hlllill k ko V W X Xs
k ~t t3
cross-sectional area of the wear scar of the cylinder liner sample in the direction of sliding at the approximate circumferential centre of the contact with the ring mininmm lubricant film thickness wear factor wear factor in the boundary lubrication regime worn volume normal load stroke sliding distance specific film thickness root mean square coefficient of friction composite root mean square roughness of two interacting surfaces
1. INTRODUCTION Piston rings are critical components in the modem internal combustion engine, their optimum tribological performance having a controlling influence in minimising frictional power loss, fuel consumption, oil consumption, blow-by and h a m ~ l
exhaust emissions. The basic roles of the piston ring pack, the collective name for the three or more rings mounted on the piston, are to form an effective gas seal between the combustion chamber and the crankcase, to limit the upward transport of oil from the crankcase to the combustion chamber and to provide a flow path for heat transfer from the piston to the engine coolant. All this must be achieved with minimum frictional power loss, most notably at the sliding interface with the cylinder wall, and the least possible wear in order to maximise component life. Unfortunately, the piston ring pack is one of the largest contributors to friction in the internal combustion engine over the normal range of engine speeds and loads encountered in service [1-3]. The relative importance varies from engine to engine, but typically the piston assembly, comprising both the piston rings and the piston skirt, accounts for 4050% of total engine friction [4]. In terms of wear, there is insufficient understanding of the interaction with the lubrication process. So, even though manufacturers can produce rings that have an excellent life expectancy, these components may be far from optimum from a lubrication and friction standpoint. As a consequence of their importance to engine
410
performance, the theoretical and experimental study of piston ring lubrication has received much attention in the literature, leading to major advances in our understanding of their behaviour [5]. The mathematical analysis of piston ring lubrication is complex and by necessity requires simplifying assumptions. However, rapid development in numerical methods over the last thirty years has resulted in sophisticated piston ring lubrication models that are finding application in the design process [6]. Figure 1 shows a typical prediction of specific film thickness, defined as the ratio of the minimum lubricant film thickness to the composite root mean square surface roughness ~, = hmi" (3-
(1)
throughout the engine cycle for the top compression ring of a diesel engine from Priest et al [7], where zero degrees crank angle is top dead centre firing. ,10~
boundary/mixedtransition
]
relatively well understood but the surface interactions that take place present much more of a challenge. Models have been developed for asperity friction [8] and the rate of wear [7] of piston tings in the boundary and mixed lubrication regimes. However, the surface interactions occurring between the piston rings, piston and cylinder bore in the presence of the engine lubricant with its complex chemistry, contamination and degradation behaviour are extremely difficult to interpret in terms of basic boundary friction and wear mechanisms [9]. Such models, therefore, are empirical and assume a variation of boundary friction coefficient or wear factor with specific film thickness. Both are zero in the fluid film lubrication regime and increase through the mixed lubrication regime, as the degree of surface contact increases, to a maximum in the boundary lubrication regime. Figure 2 shows an example for the dimensional wear factor, k, after Priest [ 10]. LUBRICATION REGIMES mixed
boun [ary
fluid film
ko 6
{
4 .................................................
0 0
90
180 270 360 450 540 Crank Angle (degrees)
630
0 0.5
1
Figure 1. Cyclic variation of specific film thickness between a top compression ring and the cylinder wall [7] It can be seen from Figure 1 that the piston ring experiences the full range of lubrication regimes in one cycle, from full fluid film through mixed lubrication to boundary, lubrication. This is characteristic of the interface between the piston ring and the cylinder wall. The fluid film aspects of this behaviour are
2
3
4
~(-)
720
Figure 2. Variation of wear factor with specific film thickness [ 10] The wear factor, k, is related to the volume of material removed by wear, V, the applied normal load, W and the sliding distance, xs, by v = kWx.
(2)
as proposed by Lancaster [11 ]. Such models are clearly very sensitive to the values chosen for the boundary lubrication regime,
411
ko in Figure 2, and the situation is further complicated by the fact that the specific film thickness tends to be influenced by changes in surface roughness as the components run-in. The discussion thus far has considered only the conditions between the piston ring face and cylinder wall but the interface between the piston ring flank and the piston groove is also important in terms of engine reliability and durability [12]. This interface has received relatively little attention in the literature but is likely to experience much more mixed and boundary lubrication than the ring face. This paper aims to assist the analyst in the selection of appropriate values for asperity friction coefficients and boundary wear factors by reporting a series of friction and wear experiments on a high frequency reciprocating tribometer. Tests were undertaken on piston ring, piston and cylinder specimens manufactured from real components in the presence of engine lubricant at speeds, loads and temperatures characteristic of a modem automotive gasoline engine. Results are presented for the critical tribological interfaces between the piston ring face and the cylinder wall and the piston ring flank and the piston groove. The experiments were undertaken on a complete ring pack from a single engine with the aim of establishing the relative importance of the various interfaces in the ring pack. The results provide a useful matrix of asperity friction coefficients and wear factors for the designer/analyst and highlight the sensitivi .ty of the wear factor to the selection of the materials of both interacting surfaces. In addition, wear factors at the piston ring flank / piston interface are shown to be much higher than at the piston ring face / cylinder interface.
components. The design and dimensions of the piston rings are shown in Figure 3. The top compression ring, ring 1, is a barrel faced, spheroidal graphite cast iron ring (150GPa minimum modulus of elasticity) with a full-face, flame sprayed molybdenum wear resistant coating (650-950 HV hardness) on the ring face. The surface of this coating has an open porous structure. Ring 2, the second compression ring, is a plain grey cast iron, Napier scraper ring (85-115 GPa modulus of elasticity, 96-106 HRB, 210-290 lib hardness) with a fine turned finish on the nmning face but no wear resistant coating. The oil-control ring, ring 3, is a steel rail, three-piece, oil-control ring (207GPa modulus of elasticity) with a chromium plating wear resistant coating on the faces of the rails (700-900 HV hardness).
3.56mm ring 1
1.50mm
3.54 mm
2.68mm 0.52mm ~
I3.00mm*
0.52mm m ( 2.68mm Figure 3. Ring pack geometry
2. METHOD 2.1. Test Specimens
The experiments were undertaken on the piston assembly components from a single-cylinder Ricardo Hydra gasoline engine. This is a 0.5 litre capacity engine of 86mm bore and 86mm stroke. Although a research engine, it is based on a General Motors two litre, four-cylinder production engine and the piston and ring pack are standard production
The piston is manufactured from a hyper-eutectic aluminium silicon alloy (56-78 GPa modulus of elastici .ty, 114 BHN hardness). The Hydra engine has a specially manufactured wet liner as the General Motors four-cylinder equivalent engine has parent metal bores. The liner is machined from grey cast iron and the surface treatment is a 450-65 ° crosshatch plateau hone. In the current experiments, however, a different,
412
similarly honed, cylinder liner of 90mm bore diameter was used. This was selected in order that small sections of the piston rings cut from near the ring gaps had good circumferential conformity with the liner. This allowed small ring samples to be used in the tests, giving more accurate assessment of wear by weight loss. The chemical composition and physical properties of the two liners is very similar as shown in Table 1.
specimen and a stationary specimen from a scotch yoke mechanism with a maximum stroke of 15mm and a maximum frequency of 50Hz. Constant loads of up to 500N can be applied and the temperature of the specimens can be controlled from ambient up to 600°C. Friction is measured by a load cell, which restrains the stationary specimen in the direction of sliding. 2.3. Test Conditions
Table 1. Comparison of cylinder liner materials Hydra Engine Total Carbon (%) Silicon (%) Sulphur (%) Phosphorus (%) Manganese (%) Chronfium (%) Tensile Strength (tons per square inch) Hardness (BHN)
current experiments
3.0-3.4 1.8-3.0 0.12 max 0.5-1.0 0.7-1.2 0.35-1.0 16
3.0-3.5 1.8-2.5 0.10 max 0.4-0.6 0.6-1.0 0.3-0.6 15
230-300
230-275
Boundary lubrication is most likely to occur in the first half of the power stroke when gas pressures and temperatures are high and, in the case of the piston ring and cylinder liner, sliding and lubricant entrainment velocities are low. This can be clearly seen in the predicted film thickness behaviour for the top compression ring of Figure 1, the first half of the power stroke being from 0 ° to 90 °. The operating parameters for the reciprocating experiments were therefore based on this region using data for the running conditions of the piston ring lubrication and wear experiments under fired conditions undertaken on the Ricardo Hydra gasoline engine [ 10]. 2.3.1
For the experiments, 30mm long sections of piston ring were cut from all three ring types adjacent to the ring gaps. Rectangular sections of cylinder liner, 58mm (axial) and 38mm (circumferential), were prepared. The piston samples were 8mm diameter pins with a 50mm crown radius on one end, machined from the crown of the piston. The lubricant was a fully formulated SAE 30 monograde engine oil with a performance level of API SF/CD. This was selected for consistency with piston ring lubrication and wear experiments under fired conditions undertaken on the Ricardo Hydra gasoline engine [ 10].
Piston ring and cylinder liner
The cylinder liner section was held stationary in a shallow bath which contained suffzcient lubricant to submerge the interface between the ring and liner. The piston ring section was held perpendicular to the liner and reciprocated as illustrated in Figure 4. The stroke, x, was 5mm and frequency 42Hz, giving a mean sliding speed of 0.42m/s.
;
Z
2.2. Test Machine
The experiments were undertaken on a standard Cameron Plint TE77 high frequency reciprocating friction and wear machine. This apparatus is highly suited to piston ring wear studies, providing a simple harmonic reciprocating motion between a moving
Figure 4. Piston ring and cylinder liner test configuration
413
Tests of two hours duration were undertaken at normal load, W, of 100N and 200N and bulk oil temperatures of 100°C and 200°C.
2.3.2
Piston ring and piston
Figure 5 shows the test configuration with a stationary ring section and reciprocating piston sample in the form of a pin with a crown end radius. Again there was sufficient lubricant in the bath to submerge the interface. The stroke, x, was 2mm and the frequency 42Hz, yielding a mean sliding speed of 0.17m/s. The reduced sliding speed when compared to the ring and liner tests reflects the much more restricted motion of the ring within the groove. One practical consequence of this change in the test procedure was that an increased test duration of three hours was required to achieve significant wear.
10 mm
W
Wear of the piston ring and piston samples was determined by weighing the samples before and after testing using an mass balance with an accuracy of one microgram. Specimens were ultrasonically cleaned in acetone before weighing. The liner specimens could not be treated in the same way because an oil lacquer tended to develop on the surface outside the region of sliding during the test. Wear could therefore not be correlated with mass loss due to the mass increase caused by the lacquer. An alternative approach is to determine wear by volume change measured by profilometry. Three-dimensional profilometry to determine volume change is complex and time-consuming for curved surfaces such as the liner. It was therefore not considered for these relatively simple wear tests, especially as piston ring wear is more significant for the tribological performance of the interface. Wear of the liner specimens was alternatively determined on a relative basis by two-dimensional profilometry. The cross-sectional area of the wear scar, Aw, in the direction of sliding at the approximate circumferential centre of the contact with the ring was selected as an indication of the severity of wear. Measurements were taken using a Rank Taylor Hobson Fornl Talysurf stylus profilometer with a vertical resolution of 0.01 ~tm. The depth of the wear scars developed during the tests was generally in the range 1~un to 8~tm.
Figure 5. Piston ring and piston test configuration Tests were undertaken with a normal load, W, of 100N for rings 1 and 2 and 70N for ring 3, reflecting the fact that it is more lightly loaded against the piston groove in the engine. Bulk oil temperatures of 100°C, 150°C, 200°C and 300°C were investigated. It should be noted that the loads, speeds and temperatures at the piston ring flank and piston groove interface are far less well understood than the ring and liner [12]. The operating conditions used in the experiments therefore represent a best estimate of the conditions experienced in the engine.
2.4. Measurement Techniques Speed, load, bulk oil temperature and friction force were monitored throughout the test period using a computer based data acquisition system.
3. RESULTS
3.1. Piston Ring and Cylinder Liner The influence of bulk oil temperature on wear for all three ring .types sliding against the cylinder liner is shown in Table 2. The chromium plated steel oil-control ring, ring 3, exhibited the lowest wear factors closely followed by the flame sprayed molybdenum coated, spheroidal graphite cast iron top ring, ring 1. By far the highest wear factors are associated with the plain grey cast iron scraper ring, ring 2. In contrast to tings 1 and 3, the wear factor of ring 2 was also sensitive to the imposed temperature change. The data reported for cylinder liner wear needs to be considered with a degree of caution given the simplistic approach. Nevertheless, it can be observed from Table 2 that the most severe wear
414
occurred when ring 3 was the counterface. The 200N applied load was, however, optimised for the compression tings and is much higher than this ring would see in service. More interesting is the reduction in worn area at 200°C for ring 1. Table 2. Effect of bulk oil temperature on piston ring and cylinder liner wear at 200N normal load piston ring, ko liner, Aw (10 -11 mm3mm-lN -1) (!03 ~tm2)
ring 1 ring 2 ring 3
100°C
200°C
100°C
200°C
2.75 4.95 1.63
2.43 8.32 1.40
5.9 6.9 20.4
1.7 6.4 21.6
Table 3 shows the results at 200°C for all three ring .types at two different applied loads. Very. similar trends to Table 2 are evident for each ring .type and the cylinder liner. Table 3. Effect of load on piston ring and cylinder liner wear at 200°C bulk oil temperature Piston ring, ko liner, Aw (10 -11 mm3mm-lN -1) (103 ~tm2)
ring 1 ring 2 ring 3
100 N
200 N
100 N
200 N
2.27 4.50 0.93
2.43 8.32 1.40
2.1 3.1 16.8
1.7 6.4 21.6
Table 4 summarises the frictional behaviour in terms of root mean square (r.m.s.) coefficient of friction throughout the test period. Very. little variation is observed but more interestingly the coefficients are rather low for boundary, lubrication with the surface materials and lubricant under study, which did not contain a friction modifier additive. Normally, a coefficient of friction of the order of 0.12 would be expected for such contacts. This suggests the experiments were operating, for at least
part of the stroke, with severe mixed lubrication rather than boundary lubrication. The omissions in Table 4 were due to measurement equipment faults experienced during these experiments. Given the lack of variation in coefficient of friction, these experiments were not repeated. Table 4. Variation of r.m.s, coefficient of friction between piston ring and cylinder liner with load and bulk oil temperature
~t (-) 200 N
ring 1 ring 2 ring 3
200°C
100°C
200°C
100 N
200 N
0.09 0.10
0.08 0.07
0.09 0.09 0.09
0.08 0.07
3.2. Piston Ring and Piston The emphasis in this sequence of experiments was placed on the top ring, ring 1, as it experiences the highest loads and temperatures at the piston ring flank / piston groove interface. The influence of increasing temperature on wear factor and r.m.s, coefficient of friction for ring 1 is given in Table 5. As a matter of information, it was found necessary, to replenish the lubricant in the bath during the test at 300°C. Table 5. Effect of bulk oil temperature on piston and top piston ring wear and friction ko (10 -9 mm3mmlN 1) piston 100°C 150°C 200°C 300°C
0.56 2.92 7.27 13.19
ring 1 0.24 0.48 1.22 3.69
~t(-) 0.07 0.11 0.13 0.13
415
First of all, it should be noted that the wear factors are generally two orders of magnitude larger than observed in the piston ring / cylinder liner interface. There is also a rapid increase in wear factors for both the piston and piston ring with increasing temperature. The data is shown graphically in Figure 6 with linear least squares fit lines through the experimental data points. 15-
•
piston
,.-,, 10"7
,
o
0 50
100
150
200
250
300
350
bulk oil temperature (°C) Figure 6. Variation of wear factor with temperature for the piston sliding against the top ring flank
The wear factors for both the piston and ring 1 basically increase linearly with temperature. This has potentially serious implications for current and future engine designs, where the trend is for higher temperatures in the top ring groove of the piston. The r.m.s, coefficient of friction between the piston and top ring flank reported in Table 5 also increases with temperature up to 200°C where it appears to a reach a plateau at a value of 0.13. This is shown graphically in Figure 7. The friction behaviour for the other two tings at 200°C is reported in Table 6 along with the wear factors. The values for ring 1 are included to aid comparison. The coefficients of friction are very similar for the three ring types and of a magnitude consistent with boundary lubrication. The wear factors for the different ring types are, however, interesting in terms of their ranking. The top ring flank / piston groove, which is the most critical in the engine in terms of operation, exhibits the largest wear factors. There appears to be a compatibility problem between the spheroidal graphite cast iron ring, the piston alloy and the lubricant. The plain cast iron second ring, ring 2, produced lower wear factors, followed by the steel oil-control ring, ring 3. Table 6. Piston and piston ring wear and friction at 200°C bulk oil temperature
0.15
ko (10 .9 mm 3mm-1N1)
,/'
piston ring 1 ring 2 ring 3
0.10::t.
7.27 3.15 2.35
ring
~t (-) 1.22 0.65 0.30
0.13 0.11 0.12
4. DISCUSSION 0.05
50
160
1;0
260
2;0
360
3;0
bulk oil temperature (°C) Figure 7. Variation of r.m.s, coefficient of friction with temperature for the piston sliding against the top ring flank
As stated at the outset, the intention of these experiments was not to investigate the complex lubrication, friction, wear and tribochemical processes that occur in piston assemblies. Rather, the aim was to observe trends in basic boundary friction coefficients and wear factors. The experiments were undertaken on a complete ring pack from a single gasoline engine to evaluate the
416
relative importance of the various interfaces in the ring pack. The experiments on the piston ring face / cylinder liner interface show a clear sensitivity to the materials in terms of wear but little sensitivity in terms of the r.m.s, coefficient of friction. Measured wear factors at the piston ring flank / piston interface were found to be much higher than at the piston ring face / cylinder interface. This highlights the sensitivi.ty of the wear factor to the selection of the materials of both surfaces. However, the sliding distances at piston ring flank / piston groove interface are much less than for the typically boundary, or mixed lubricated proportion of the piston ring face / cylinder liner stroke. Hence the volumetric wear, V in equation (2), at the piston ring flank / piston groove interface will not be as large as expected from a simple examination of the wear factors. Wear at the piston ring flank / piston groove interface was also found to be sensitive to material type. Most worryingly, wear for the critical top piston ring and piston was found to be the highest and increased markedly with rising temperature. 5. CONCLUSIONS •
Simple bench tests for all the critical interfaces in the piston ring pack for a single engine have yielded a useful matrix of asperity friction coefficients and wear factors for the designer / analyst.
•
The sensitivity of the wear factor to the selection of the materials of both surfaces is clear.
•
Wear factors at the piston ring flank / piston interface are much higher than at the piston ring face / liner interface.
•
Wear for the critical top piston ring / piston groove interface was found to be the highest and increased markedly with rising temperature.
REFERENCES
1. Monaghan, M.L., "Engine friction- a change in emphasis", Instn. Mech. Engrs., 2nd BP Tribology Lecture, 1987.
2. Monaghan, M.L., "Putting friction in its place", 2nd Int. Conf.: Combustion Engines- Reduction of Friction and Wear, IMechE conf. pub. 1989-9, Paper C375/KN1, 1989, pp. 1-5. 3. Parker, D.A. and Adams, D.R., "Friction losses in the reciprocating internal combustion engine", Tribology - Key to the Efficient Engine, Instn. Mech. Engrs. Conf. Pub. 1982-1, Paper C5/82, 1982, pp. 31-39. 4. Gazzard, S.T., et al., "Piston system design for low emissions", Leading Through Innovation, T&N Symposium 1995, Paper 20, 1995. 5. Dowson, D., "Piston assemblies; background and lubrication analysis", Engine Tribology, C.M. Taylor (ed.), Tribology Series, 26, Elsevier, Ch. 9, 1993, pp.213-240. 6. Mierbach, A., et al., "Piston ring performance modelling", Leading Through Innovation, T&N Symposium 1995, Paper 15, 1995, pp. 15.115.13. 7. Priest, M., Dowson, D. and Taylor, C.M., "Predictive wear modelling of lubricated piston tings in a diesel engine", Wear, 231, 1, 1999, pp.89-101. 8. Rohde, S.M., "A mixed friction model for dynamically loaded contacts with application to piston ring lubrication", Surface Roughness Effects in Hydrodynamic and Mixed Lubrication, ed. S.M. Rohde and H.S. Cheng, ASME pub., 1980, pp. 19-50. 9. Coy, R.C., "Practical applications of lubrication models in engines", New Directions in Tribology, MEP, 1997, pp. 197-209. 10. Priest, M., "The wear and lubrication of piston tings", PhD thesis, University of Leeds, October. 1996, pgs. 246. l l. Lancaster, J.K., "Dry bearings : a survey of materials and factors affecting their performance", Tribology, December, 1973, pp.219-251. 12. Barrell, D.J.W., Priest, M. and Taylor, C.M., "Bench test study of piston ring flank and piston groove interaction", Lubrication at the Frontier, Proc. 25th Leeds-Lyon Symposium on Tribology, Lyon, 1998, Elsevier, Amsterdam, 1999, pp.343-351.