piston ring friction loss behaviour for motored and fired reciprocating engines

L.L. Ting and T.S. Shih Research Laboratory, Ford Motor Company, Dearborn, MI, USA

A previously developed piston ring 1 ubrication model has been further extended so that the piston ring frictional losses can be predicted in both hydrodynamic lubrication and metal-to-metal contact regions for various engine operating conditions. Ring friction results for two engine types are presented for both hot motoring and engine firing conditions. The hot motoringpredic- tions were found to be in good agreement with tests. Results show that when the engine is motored, piston ring friction losses in the hydrodynamic lubrication region predominate. If the engine is fired, the losses in the metal-to-metal contact region become dominant due to high gas pressure and temperature effects. Ring friction loss can be significantly reduced by using low tension rings with a correct ring sliding face profile.

Abstract

Piston Ring Friction Loss Behaviour for Motored and Fired Reciprocating Engines

piston ring, friction loss, reciprocating engine, fired engine, lubrication model

Keywords

It has long been recognised that, among the many engine moving components, the piston ring is considered to be the one contributing most significantly to the total engine mechanical friction 1 0 s s e s . ~ ~ ~ ~ ~ An engine friction reduction study by a hot engine motoring method has also shown that the major motoring friction components ranked in the order of importance are: pistodrings, valve train, crank shaft, and oil pump.4 Methods developed which can be used to reduce the piston and ring friction losses, especially the piston rings, therefore, would have a far greater impact on the improvement of engine fuel economy than those of the other friction components quoted.

However, successful piston ring friction reduction methods could not be efficiently and effectively developed without a thorough understanding of the fundamental nature of ring loss

INTRODUCTION

Lubrication Science 8-1, October 1995. (8) 37 0954-0075 $1 0.00 + $4.00

38

L.L. Ting and T.S. Shih: Piston Ring Friction Loss Behaviour for Motored and Fired Reciprocating Engines

behaviour. To aim at this, a previously developed piston ring lubrication and cylinder bore wear model, RINBOW5s6 was further extended and refined, so that the piston ring and the ring pack frictional losses in the metal-to-metal contact region, as well as in the hydrodynamic lubrication region, could be both calculated and evaluated.

In this paper, some of the piston ring and piston assembly frictional loss calculation for two types of gasoline engines, considering both hot motoring and engine firing conditions, will be presented and discussed. The hot motoring calculations have been compared with the engine experiments, and the agreement was found to be good. This indicates that the refined RINBOW computer model is relevant and suitable for predicting the piston ring and piston assembly friction loss behaviour for various engine operating conditions. By using this model, the necessary parameter changes, such as ring face profile, ring tension, piston and ring designs, crank radius to connecting rod length ratio, weight, . . . etc., may be determined to minimise the piston ring and piston assembly frictional losses.

Computer modeldprograms of a similar nature, predicting engine component friction losses, but with different degrees of sophistication and complexity, can be seen in references 7 to 10.

PISTON ASSEMBLY FRICTIONAL Loss COMPUTATIONAL METHOD

The total piston assembly frictional losses are calculated considering the friction contributions due to the piston rings, piston skirt, wrist pin bearings, and the connecting rod bearings.

The RINBOW computer program begins by requesting in- put of the design and weight information of all the piston assembly moving components. Since the ring friction losses will be computed first, detailed information on ring face geometry, ring tension, ring material Young's modulus, combustion chamber pressure vs. crank angle relation, cylinder wall and piston temperature distributions, type of oil and its viscosity-temperature relation, and ring-bore coefficient of friction, is also needed. As the ring face geometry is an extremely important parameter, measured ring profiles must be available.

Figure 1 shows the ring profiles of the top ring, second ring, and oil ring segment of a 2.01 engine, designated as En- gine B, measured the use of a surface analyser. I t is interesting to note that the top ring of this ring set has an off-centred parabolic face profile.

Lubrication Science 8-1, October 1995. (8) 38 0954-0075 $10.00 + $4.00

39

~~~~~~~ ~~~


Figure 1 Measured piston ring face profile of Engine B (2.01)

(a) Top ring face profile: off-centred parabolic

r = $ Top surface

(b) Second ring face profile: tapered

200 pm

(1 p in = 0.0254 pm)

(c) Oil ring segment face profile

Ring friction computational procedures start with the calculation of inter-ring gas pressures and blowby rate over a com- plete firing cycle. The ring-bore oil film thickness loci, considering the effects of behind-ring gas pressure, ring tension, and cylinder wall temperature (viscosity) distribution, are computed next. By examining whether the minimum oil film thickness is less or greater than a critical oil film thickness value, known as the lower limit for film lubrication, and varying from 1.0 to 1.3 pm depending on surface finish characteris- t i c ~ , ~ , ~ the ring metal-to-metal contact operation region can then be separated from the hydrodynamic lubrication region.

In the metal-to-metal contact region, the frictional force of each ring is the product of the ring contact pressure load and


40


Figure 2 Instantaneous

distribution friction coefficient 0.10

Symbol LP Li 2. r= .g 0.08 0 .- o 4448N 445N

(1000 Ib) (100 Ib) - 0 6672N 890 N 0

(1500 Ib) (200 Ib) .- F = Friction force E

L - 0.06 al

0

al Contact load 0.04

3 0 a, C tQ

m In C

E 0.02 c

- 0

Figure 3 Looping Stribeck relation -friction coefficient vs. Summerfeld number

f

-100 -80 -60 -40 -20 TDC 20 40 60 Crank angle, deg

80 100

1 I I I

Top ring face profile: barrel shaped Oil type: Mobil5W30 Downstroke

3 Upstroke

-0.2 c Dl -1 00 -50 0 50 100


41

L.L. l ing and T.S. Shih: Piston Ring Friction Loss Behaviour for Motored and Fired Reciprocating Engines

the friction coefficient. The value of the friction coefficient can be a constant in average, or a function of either the crank angle degree, or, more generally, the Sommerfeld number, as shown in Figure 2,l and Figure 3,3 respectively. The constant value of 0.04 has been used in all the current computations.

In the hydrodynamic lubrication region, since the ring sliding velocity, oil viscosity, and oil film thickness are all known at any given crank angle degree, the corresponding frictional force due to viscous shear can then be calculated by the use of the following relation:"

Frictional force = [Oil viscosity] x [Sliding velocity1 x [Ring face area]/[Oil film thickness]

Finally, the integration of the product of the ring velocity and the frictional force over a cycle gives the frictional power loss per ring. The total engine frictional power loss due to all the rings can then be determined with the known number of the piston rings per cylinder, and the number of cylinders per engine. The piston skirt friction force is the product of piston skirt side thrust and a friction coefficient value, which is a function of engine rotational speed, and the piston side thrust pressure. l2 Piston pin and connecting rod bearing frictional forces are calculated based on the McKee's friction coefficient relation," modified with the consideration of boundary lubrication effect.

Taking into consideration the engine hot motoring test conditions with the ambient combustion chamber pressure simula- tion, the friction losses of the piston ring pack, piston skirt, and the connecting rod and wrist pin bearings, were computed for four different engine types (1.61,2.01,2.31,3.31) with the engine speed varying from 1,000 to 4,000 r/min. They were all found to be in good agreement with the tests. Results of Engine A (1.61) and Engine B (2.01) are given in Figures 4 and 5 respectively (see overleaf).

As can be seen, generally, the total motoring friction, Fric- tion Mean Effective Pressure (FMEP), increases as the engine speed increases, and it is the piston ring friction loss which makes up the largest portion of the total. Figures 4 and 5 also show that the motored piston ring friction losses in the hydrodynamic lubrication region predominate when the engine

RESULTS AND DISCUSSION



10 -

h * 8- 4 2 II 6-

m

.- rn Q Y

4 - rn a a ; 2-

0

Figure 4 Piston assembly friction losses, hot motoring, Engine A (1.61)

connecting rod RINBOW - - -

- Experiment

. - - - - - 0

0 /

0 0

0 c - 0

I /

4 c

Piston skirt friction I- / -

I I I I I _- - - -

Figure 5 Piston assembly friction losses, hot motoring, Engine B (2.01)

.- rn - 4 u .- 2 a W

h

RINBOW Experiment Pistoi pin &

connecting rod

’/

I 1000 2000 3000 4000

Engine speed, r/min


43


speed becomes greater than about 1,500 r/min. Below that, the losses in the metal-to-metal contact region are either some- what larger than or approximately the same as those in the hydrodynamic lubrication region, depending on engine design.

This is understandable because, when the engine is motored with ambient combustion chamber gas pressure conditions, the piston ring oil film thickness is mainly a function of engine speed. A low engine speed run means low oil film thickness loci generation, so the rings will operate primarily in the metal-to-metal region, resulting in large friction loss in that region. As the engine speed begins to increase, oil film thickness increases accordingly. The ring operation region then changes from more of the metal-to-metal contact to more of the hydrodynamic. Further speed increases, obviously, lead to predom- ination of the hydrodynamic operations and, consequently, the large hydrodynamic friction losses, as Figures 4 and 6 show.

Now, if the motored ring friction losses of these two engines in the metal-to-metal contact regions are compared, it can be seen that Engine B is the one with the much lower FMEP values than those of Engine A, at all the engine speeds. Some of the important factors contributing to Engine B’s low friction loss behaviour have been identified. They are: off-centred parabolic face profile of the top ring, thin ring width, low oil ring tension, lightweight piston, and smaller crank radius to connecting rod length ratio. It is also worth noting that it is mainly the small metal-to-metal contact region ring friction losses of Engine B that are responsible for its low friction loss behaviour.

Excluding the FMEP losses due to connecting rod and piston pin bearings, Figure 6 (overleaf) shows that, by the use of low tension rings, the FMEP can be reduced by an amount of 0.2 to 0.6 psi (1.38 to 4.14 kPa) depending on engine speed and simulated constant combustion chamber pressure conditions. Engine fuel economy, therefore, can be improved by using the low tension rings. However, it should be pointed out that the ring tension reduction should not be done too aggressively, in order to avoid the adverse effects of high engine oil consump- tion and degradation of the ring sealing capability.

Since the RINBOW engine hot motoring friction predic- tions were found to be in good agreement with the experiments, it is reasonable to expect that good friction results can be ob- tained if RINBOW is applied to fired engine friction loss

Lubrication Science 8-1, October 1995. (8) 43 0954-0075 $1 0.00 + $4.00

44

Figure 6 Effects of ring tension and combustion chamber pressure on piston ring friction losses, hot


Ring tension [Top/second oil] Ib (1 Ib = 4448 N) Production rings [6.1/4.9/10.8]

- Low tension rings [2.9/3.4/7.6] motoring, Engine A (1.61)

h m 4 Q,

2 II m a .-

7 v .- m a a W

h

- _ _ - - - I I

/ I - '1 0 ' I I I

1000 2000 3000 4000 Engine speed, r/min

computations. Using the same Wide Open Throttle (WOT) engine firing conditions, the piston ring and the piston skirt friction losses for Engine A and Engine B were calculated, and the results are given in Figures 7 and 8, respectively. As can be seen, the ring FMEP losses in the metal-to-metal contact region are now becoming larger than those in the hydrodynamic lubrication region, which is completely opposite to the hot motoring engine results previously shown.

This friction loss behaviour change is not surprising since, for the fired engines, the high behind-ring gas pressure and the low oil viscosity due to the cylinder wall temperature rise, will cause a substantial reduction of the ring oil film thickness. Pis- ton rings, therefore, are operating mostly in the metal-to-metal contact region over the cycle, resulting in large FMEP losses in that region. On the other hand, in addition to the loss of the hydrodynamic lubrication region operation, the low oil viscosity value also contributes to the decrease of the ring friction loss in that region. This is because the hydrodynamic friction loss is proportional to oil viscosity, and the oil viscosity decreases sharply as the oil film temperature becomes high. The fired


45


Figure 7 Piston ringskirtfriction losses, WOT, Engine A (1.61)

7

/ I

/ /

_ / - - A -

/ I E-.------ +

-: I I I I I

1000 2000 3000 4000 Engine speed, r/min

Figure 8 Piston rindskirt friction losses, WOT, Engine A (1.61), and Engine B (2.01)

Engine A (1.61)

Engine B (2.01) - - - _ _ _ _ - - - - - - ~etA- to - Piston ring metal friction

2.01- - I

/ I

/ I I - - / - - - --- 4

_ / - -

0 1000 2000 3000 4000

Engine speed, rlmin

~ ~~~~~~ ~ ~~ ~~~


46

~~ ~


engines, of course, always run at higher temperatures, and sometimes much higher than those of the motored engines.

Comparing Figures 4 and 5 with Figures 7 and 8, it can be seen that, for the motored engines, the ring losses in the hydrodynamic lubrication region predominate. For the fired engines, metal-to-metal contact region ring friction losses become dominant. The ‘well-recognised’ claim that ‘there are measured engine friction differences between firing and motoring condi- tionj4 thus can now be considered as in theory correct, since piston rings are the key friction components. This also means that one should be extremely careful, if it is attempted to determine the fired engine friction loss characteristics simply based on motored engine test results. Further experimental investiga- tions of fired engine friction loss behaviour, therefore, are still needed in order to verify fully and confirm the RINBOW predic- tions. Such work should begin by developing some in situ piston ring oil film thickness and friction loss measuring techniques, which can be proven as accurate and reliable.

Figure 8 also shows that, under similar engine WOT firing conditions, the piston assembly FMEP loss of Engine B is about 1.8 psi (12.4 kPa) lower than that of Engine A in the speed range of 1,000-4,000 r/min. Again, it is mainly the small piston ring metal-to-metal contact region friction losses of En- gine B that are responsible for such FMEP differences.

In order to examine how the top ring profile would affect the piston assembly friction losses, and to determine the optimum shape minimising the friction losses, additional FMEP computations have been made by changing the ring face crown height, 6, systematically, considering again the engine WOT firing conditions. Figure 9 shows the FMEP vs. 6 relations at engine speeds of 1,500 and 3,000 r/min for the top rings of Engine A and Engine B, having the symmetrical and off-centred face profiles, respectively, as shown.

As can be seen, for Engine A and at both engine speed levels, the 6 value of minimum FMEP loss is 0.000064 in. (1.63 pm). Obviously, if this 6 is a value somewhere between 0.00004 in. (1.02 pm) and 0.00015 in. (3.81 pm), it still can be considered as an optimum, because the curves are rather flat within those two limits. The measured 6 value of the top ring of Engine A, however, is 0.000923 in. (23.4 pm), and the corresponding FMEP values, represented by triangles ‘A’ , are about 0.6 psi (4.13 kPa) higher than the optimum. This means that, if the

~~ ~ ~


47

6.0 -

5.0 -

4.0 -


Figure 9 Optimum top ring profiles for Engine A and Engine B, WOT

Top ring face profile symmetrical

h RI highly curved R Y m E

I I .- IJY Q

.- v) P a W

h

3000 r/min Optimum 6 Engine A

1500 r/min I I

Optimum 6 I I t

3000 rlmin 1500 r/min

! I ' Top ring face profile

I I off -centred I I parabolic

3.0 ' I I I I I I I I I I I I I I I I I I I I I O!OOOOl 0100003 0.0001 010003 0.001

6, in. (1 in. = 0.0254 m)

highly curved top ring of Engine A can be redesigned by changing the crown height value from the present of 0.000923 in. (23.4 pm) to the optimum of 0.000064 in. (1.63 pm), the ring frictional loss can be reduced by an amount of about 0.6 psi (4.13 kPa).

The top ring optimum crown height of Engine B has also been determined, and is 0.000048 in. (1.22 pm), as shown. It is interesting to note that the measured top ring crown height of this engine, represented by the solid triangles 'A', is already very close to the optimum, because the corresponding FMEP losses, compared at the same engine speed levels, are only slightly higher than the minimum values. Figure 9 also shows that, in general, the friction loss of Engine B is about 1 psi (6.89 kPa) lower than that of Engine A. This clearly demonstrates the engine friction reduction advantage of using top compres- sion rings with off-centred face profiles having the optimum crown height value. Top rings with highly curved and symmetrical face profile are generally not desirable.


48


Piston assembly friction losses predicted by RINBOW were found to be in good agreement with engine hot motoring test results, indicating that RINBOW is relevant and suitable for an- alysing and predicting piston assembly friction loss behaviour. Results show that piston ring friction loss behaves very differ- ently depending on engine operating conditions. When the engine is motored, ring friction losses in the hydrodynamic lubrication region predominate. The metal-to-metal contact region friction losses, however, become dominant if the engine is fired. This is due to the high behind-ring gas pressure, and the high oil film temperature (low oil viscosity), resulting in the decreases of ring oil film thicknesses. Piston ring friction losses can be significantly reduced by using a low tension top ring with an off-centred face profile having the correct crown height value.

CONCLUSION

Rererences 1. Ting, L.L., ‘Lubricated piston rings and cylinder bore wear, Wear Control Handbook, ASME, p. 609,1980. 2. Ting, L.L., ‘A review of present information on piston ring tribology’, SAE Paper 852355,1985. 3. Ting, L.L., ‘Development of a reciprocating test rig for tribological studies of piston engine moving components part I: Rig design and piston ring friction coefficients measuring methods’, SAE Paper 930685, 1993. 4. Kovach, J.T., Tsakiris, E.A., and Wong, L.T., ‘Engine friction reduction for improved fuel economy’, SAE Paper 820085,1982. 5. Ting, L.L., and Mayer, J.E., Jr., ‘Piston ring lubrication and cylinder bore wear analyses, part I: Theory’, J. of Lubr. Tech., 96F, 305 (1974). 6. Ting, L.L., and Mayer, J.E., Jr., ‘Piston ring lubrication and cylinder bore wear analyses, part II: Theory verification’, J. of Lubr. Tech., 96F, 258 (1974). 7. Hoshi, M., ‘Reducing friction losses in automobile engines’, Trib. lnt., 17, 185

8. Hamai, K., et a/., ‘Development of a friction prediction model for high performance engines’, Lubr. Eng., 47,567 (1991). 9. Goenka, P.K., et a/., ‘FLARE: An integrated software package for friction and lubrication analysis of automotive engines - part I: Overview and application’, SAE Paper 920487,1992. lO.Jeng, Y.R., ‘Friction and lubrication analysis of a piston-ring pack‘, SAE Paper 920492,1992. 11. Fuller, D.D., Theory and practice of lubrication for engineers, John Wiley & Sons, New York, 1956. 12. Awano, S., Thermodynamical performances of four-cycle gasoline engines, Report of the Research Institute of Technology, Nihon University, Japan, 1962.

(1 984).


piston ring friction loss behaviour for motored and fired reciprocating engines

Documents