Modelling lubricant related fuel economy in heavy duty diesel engines

Tribological Research and Design for Engineering Systems D. Dowson et al. (Editors) 9 2003 Published by Elsevier B.V.

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Modelling Lubricant Related Fuel Economy in Heavy Duty Diesel Engines R. Castle, S. Arrowsmith Infineum UK Ltd, Milton Hill Technology Centre, PO Box 1, Abingdon, OX13 6BB, UK

Current and future legislation controlling the exhaust emissions from Heavy Duty Diesel (HDD) vehicles requires an overall improvement in engine efficiency as well as the use of specialised after treatment systems to remove harmful species from the exhaust gas, and prevent these from entering the environment. Engine efficiency can be improved in a number of ways, and a combination of all approaches is required to ensure that the desired levels can be achieved. One approach is to improve the fuel efficiency of crankcase lubricants to optimise fuel consumption and minimise the amount of harmful combustion gas generated. In addition to the environmental benefits, reductions in overall fuel consumption in heavy duty diesel vehicles will represent significant cost savings to transport fleet operators. It is well known that vehicle fuel economy can be improved significantly by using oils of lower viscosity grades. More recently it has been shown that improvements in HDD fuel economy can also be achieved by optimisation of the lubricant additive package 1, and that the appropriate selection of lubricant viscometrics and additive package can generate improvements in fuel economy without compromising wear performance. In this study, the impact of both lubricant viscometrics and additive type and treat on HDD fuel economy has been evaluated in a series of laboratory bench tests which measure friction under different lubrication regimes, and also in a full driveline rig. It will be shown that fuel economy improvements of up to 4% can be achieved in a fired engine by moving from a 15W40 crankcase lubricant to an optimised 0W20. In addition, it will be shown that within a single viscosity grade, significant improvements in fuel economy can be achieved by tailoring the viscometrics and also b y optimising the additive package. Finally, it will be shown that by using laboratory bench test data to model quantitatively the response in the fired engine, the relative importance of different lubrication regimes in determining overall HDD vehicle fuel economy can be estimated. 1. INTRODUCTION Proposed legislation controlling exhaust emissions from heavy duty diesel vehicles will require significant reductions in the levels of harmful gases and particulates escaping to the environment. The allowed limits will decrease in some cases by an order of magnitude over the next five to eight years. A number of different approaches will be needed to ensure that these targets can be met, including the use of after treatment devices, design changes to improve engine efficiency and maximising the contribution of the crankcase lubricant to fuel economy. In addition to the legislative requirements for reduced emissions, there is considerable pressure from haulage operators to improve the fuel efficiency of heavy duty diesel vehicles. Fuel costs represent a significant part of the total running costs of any transport fleet. Estimates vary 2'3 between 30% and

50% and will depend upon taxation schedules in different ~countries, type of vehicle and service requirements. The ability of the engine lubricant to reduce fuel consumption is therefore an important route not only to the reduction of vehicular emissions, but also to enhanced operating profits for the consumer. The extent to which crankcase lubricants can reduce fuel consumption in heavy duty diesel vehicles is dependent on many factors, but previous estimates 24 suggest that 3.5% might be achievable relative to the consumption recorded when a conventionally formulated 15W40 HDD oil is used. In order to maximise lubricant related fuel efficiency it is necessary to minimise engine friction over a wide range of conditions of temperature, pressure, shear rate and speed. The combinations of these conditions will result in different lubrication regimes operating in the engine, including full film

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and mixed hydrodynamic lubrication, full film and mixed elastohydrodynamic lubrication and some boundary lubrication. In an HDD engine, all of these lubrication regimes will be in operation, although to differing extents. Previous estimates 5 have suggested that in an HDD engine, frictional losses in the bearings and piston assembly (mixed and full film HD lubrication) account for the up to 90% of engine friction, the rest being generated in the valvetrain (mixed/full film EHD and boundary lubrication). Thus the maximisation of fuel economy from the lubricant will require primarily minimisation of friction under mixed and full film HD lubrication conditions, and additionally (although the effects are expected to be less significant) minimisation of friction under mixed/full film EHD and boundary lubrication conditions. Under full film lubrication conditions, be it HD or EHD, frictional losses are dependent on the dynamic viscosity of the lubricant and upon the effective pressure viscosity coefficient, a. Engine friction can therefore be reduced by minimising the dynamic viscosity of the lubricant (and, under EHD conditions, using a basestock of lower a value). However, the benefits obtained from reductions in viscosity are not unlimited, since under HD and EHD conditions, both film thickness and friction depend equally on viscosity. Therefore reductions in viscosity lead to a decrease in film thickness until eventually the ratio of film thickness to composite surface roughness, 2, falls below a critical level (-3) and asperity contact occurs. This causes an increase in friction as the lubrication moves from full-film to the mixed regime. It has been demonstrated 6 previously that surface active additives can delay the transition from full film to mixed lubrication. This is achieved by the formation of discrete chemisorbed layers on the metal surfaces which effectively reduce the composite surface roughness and thereby delay the transition to mixed lubrication. Under boundary conditions, as might be experienced in the valvetrain (particularly with oils of lower viscosity grades), frictional losses can be reduced by appropriate additive selection, such that metal to metal contact is minimised, and boundary

friction controlled, through the action of the lubricant. In this study, the impact of three variables on lubricant related fuel economy has been studied. The effect of changing viscosity grade, varying VM type and treat and optimising the additive package on friction and fuel economy has been evaluated in laboratory bench tests and a driveline rig respectively. Finally, the complete data set has been analysed statistically in an attempt to understand the role of different lubrication regimes in the generation of total engine friction. 2. LUBRICANT S E L E C T I O N A matrix of oils with differing viscometric properties and formulations has been blended to evaluate the impact of viscosity grade, ratio of newtonian to non-newtonian viscosity and additive package on friction under different lubrication regimes and engine fuel efficiency. A description of the test oils is given in Table 1. Table 1 Visc Grade

Test oil matrix Newt. Visc

BS

I

Block Group II , copolymer .

15W40 0W20

Max

PAO

5W30

Max

PAO

5W30

Min

PAO

5W30 5W30 5W30 5W30 5W30

VM

Min PAO Min PAO ] Min 1 PAO ] Min' PAO Min PAO

5W30

Max

5W30

Max

PAO .

PAO

Block copolymer Block copolymer Block copolymer Star A Star B Star C OCPA OCP B Block ! copolymer Block . copolymer . L

Addpak A

A A A A A A A A B C

Table 1 shows how the viscometric and formulation parameters were varied in each of the 11 oils. Basestocks (BS) used were either a Group II or

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PAO and three different additive packages were studied. Six viscosity modifiers (VM) and three different viscosity grades have been evaluated. Within each viscosity grade it is possible to vary the Newtonian viscosity of the oil (that is the part which is independent of shear rate), and still achieve the same finished oil viscometrics (HTHS and kV100). This is achieved by using different amounts of viscosity modifier to boost the viscosity at low and medium shear rates. This is illustrated schematically in Figure 1.

3. LABORATORY BENCH TESTS The matrix of oils described in Table 1 has been evaluated using a number of laboratory bench tests, and a full driveline rig at Shell Global Solutions PAE laboratory, Hamburg. The bench tests aim to evaluate either friction or contact film thickness under a number of different lubrication regimes and the driveline rig measures the impact of the crankcase lubricant on vehicle fuel economy under controlled running conditions. The equipment used is described below.

3.1. Journal Bearing Friction Rig

Viscosity

-iiiii12 .

Higher Newtonian viscosity Lower Newtor~ian

viscosity

Shear Rate

Figure 1

Schematic showing how the same low shear rate viscometrics can be achieved using different ratios of Newtonian (or base viscosity) to non-Newtonian viscosity.

The Newtonian component of the viscosity is provided by the basestock and the additive package, whereas the non-newtonian contribution is provided by the viscosity modifier. At high shear rates, the viscosity modifiers undergo shear thinning and their contribution to viscosity is lost. At these shear rates, the viscosity of the oil will be determined by the Newtonian component of the viscosity which is independent of shear rate. Each of the 11 oils in Table 1 have been formulated either to maximise the Newtonian component of the viscosity, or to minimise it. These details are given in the second column of Table 1. Two oils formulated to the same low shear rate viscometrics, where the Newtonian viscosity has been maximised in once case and minimised in the other, might well be expected to behave differently in an engine or rig test where shear rates are often in excess of 106 s1.

The journal bearing friction rig, as shown in Figure 2, has been developed by Infineum and PCS Instruments. This rig uses automotive production crankshaft bearings and a modified production connecting rod. Load can either be held constant, or varied so as to replicate the load/crank angle profile of a real engine. Bearing friction is measured as a function of speed at a range of temperatures (up to 150 ~ Stribeck curves generated using this rig show the variation of friction with speed as the lubrication regime changes from full film hydrodynamic to mixed hydrodynamic.

Figure 2

Bearing Friction Rig.

3.2. EHD Optical Film Thickness Rig EHD film thickness between a steel ball and glass disk is measured as a function of entrainment speed using an optical EHD test rig 6, as shown

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schematically in Figure 3. In these tests, entrainment speed was varied between 10.3 and 102 ms 1, while test temperatures ranged between 40 ~ and 100 ~ Tests were carried out in pure rolling.

3.4. HFRR (Boundary friction rig) Boundary friction coefficients were measured at temperatures between 40 ~ and 140 ~ using a High Frequency Reciprocating Rig (HFRR), as shown schematically in Figure 5. The test comprises a steel ball reciprocating against a steel disc at 20 Hz with a stroke length of 1 mm. A 4 N load was used in all tests. Friction coefficient and temperature are measured as a function of time. Vibrator

Ba!!',

I

Fla t

[ Figure 3

I i

EHD Optical Film Thickness Rig.

i

3.3. Mini Traction Rig (MTM) Mixed and full film EHD friction (or traction) is measured using a ball on disc traction rig, as shown in Figure4. Traction measurements are made between a steel ball and a steel disc at a range of different slide/roll ratios using a fixed disc speed. Measurements were made at a series of temperatures ranging from 40 ~ to 135 ~ C. Load Beam /

Traction T r a n s d ~ e r / - - - -

Ball Motor

Disc Mot

Ball

Load Motor

Dis

Ball Gimbal Mount

Figure 4: Mini Traction Rig.

Figure 5

High Rig.

Frequency

Reciprocating

3.5. Driveline Rig A Volvo FH12 direct injected 6 cylinder engine, in line with EURO 2 emissions, was installed with its associated gearbox and axle system in the Shell Global Solutions PAE Hamburg test facility. A schematic of this rig is shown in Figure 6 below. Fuel consumption is monitored continuously and differences in performance are calculated relative to a 15W40 reference oil.

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Figure 6 Schematic of driveline rig used to measure fuel economy contribution of test lubricants.

Figure 7 Variation of bearing friction with journal speed at 100 ~ for oils with different viscometric properties.

The testing programme was conducted under both steady state and transient operations so as to reflect realistic driving conditions. The gearbox and axle were lubricated with a standard gear and axle oil. Table 2 shows the general loading conditions used for this study.

The graph in Figure 7 shows the impact of changing viscometrics on bearing friction. All oils were blended using the same additive package and viscosity modifier. In the 15W40, 0W20 and 5W30(max) oils, the Newtonian viscosity has been maximised. A comparison of these three oils shows clearly that the bearing friction is reduced over the majority of the speed range by moving to a lower viscosity grade. However, it should be noted that at low speeds, the lubrication transitions from full film to mixed hydrodynamic, and that this transition occurs at progressively higher speeds as the viscosity grade is reduced. The result is that at low speeds the 0W20 oil exhibits considerably higher friction than the 15W40 oil. In the 5W30(min) oil, the Newtonian component of the viscosity has been minimised so that a much greater contribution to viscosity is derived from the non-Newtonain viscosity modifier. A comparison of the two 5W30 oils demonstrates that within a given viscosity grade, significant reductions in friction can be achieved through tailoring of the viscometric properties. It should be noted that both these oils have an HTHS of 3.5 cP, and a kV100 of 12 cSt.

Table 2

General loading conditions. RPM

Power

Transient City Running

60O (min) 1400 (max)

Torque (Nm) 25O (min) 1100 (max)

15.7 (min)

Steady State Freeway

1200

750

94

Engine Condition

(kW) 160 (max)

4. RESULTS & DISCUSSION

4.1. Effect of Viscosity Grade and Newtonian Viscosity The effect of viscosity grade and Newtonian viscosity within a given viscosity grade on friction and fuel economy was evaluated using the rigs

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Figure 10 Measured improvements in fuel economy at 90 ~ relative to 15W40 under various running conditions in the driveline rig.

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The driveline rig data confirms the trends observed in the journal bearing rig, that is, the greatest fuel economy improvement is observed for the 0W20 oil, and the 5W30(min) oil shows a FE credit compared with the 5W30(max) oil. This validates the hypothesis that two oils formulated within the same viscosity grade, and with the same HTHS and kV100 viscosities can have markedly different performance in fuel economy tests. The final result on the graph shows the repeat measurement on the 15W40 oil. This repeat suggests that the error in the measurements is of the order of 0.3%. 4.2. Effect of Additive Package Three oils containing different additive systems were used to evaluate the impact of additive chemistry on HDD fuel economy. Figure 11 shows the measured bearing friction as a function of journal speed for these oils.

Figure 11 Variation of bearing friction with journal speed at 100~ for oils containing different additive systems. The graph in Figure 11 demonstrates that no reduction in friction is achieved under these conditions by optimising the additive package. However, the relative improvement in friction of the 5W30 oils compared with the 15W40 is clearly seen. Figure 12 shows the effect of optimising the additive package on boundary friction. In this case, the 0W20 oil is added for comparison.

Figure 13 Measured improvement in fuel economy at 90~ relative to 15W40 for oils containing different additive systems.

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Figure 14 Variation of bearing friction with journal speed at 100 ~ for oils formulated with different viscosity modifiers. The data from the journal bearing rig in Figure 14 shows a small advantage in friction reduction for the Star and block copolymer over the OCP. At very low speeds, the star polymer gives the lowest friction. All 5W30 oils show an advantage over the 15W40. Figure 15 shows the comparative data from the driveline fuel economy rig.

Improving fuel efficiency in any vehicle requires the minimisation of friction in all parts of the engine, and it is well known that different areas of the engine will operate under different lubrication regimes. Since friction reduction under each of these regimes requires a different formulation approach, the minimisation of total engine friction is facilitated by an understanding of the relative importance of each of the lubrication regimes in the engine. The bench rigs used in this study are designed to measure friction or film thickness under different lubrication regimes. It was anticipated that by

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analysing statistically the results from the bench test data and the driveline rig, it might be possible to estimate the relative importance in the engine test of friction generated under the different regimes. The data from the driveline rig (the "response") has been modelled with a series of predictors using a linear regression technique. The predictors included data from all the bench tests, chosen to represent the

behaviour of the lubricants under specific lubrication regimes. The analysis looked firstly at main effects and secondly at the interactions between those effects. The objective was to determine which of the predictors, if any, were most highly correlated with the response data, and thereby estimate the importance of various lubrication regimes in determining friction in the engine.

Statistical models from linear regression analysis using main effects only; JBM = journal bearing rig, FT = film thickness rig; MTM = traction rig R2 Associated Lubrication Regimes Model Response Table 3

i

Transient 90 Steady State 90 Combined 90 Table 4

0.97

Mixed h~dro@namic

ks + as[MTM 135]

0.35

Mixed EHD

ko + ao[FT1]

0.41

Mixed hydrodynamic

k t at-

at[JBM1] -t- bt[FZl]

Statistical models from linear regression analysis including interaction terms

Response Combined 90

Model

R2

Associated Lubrication Rel~imes

Kc + Ao[FT2*MTM80] + B~[FTI*FT3]

0.91

Mixed EHD/HD?

Table 3 shows some of the resulting models from this analysis, using main effects only. In each case, the R 2 value is also given to show the extent of the correlation. A strong correlation is observed between the transient data from the driveline rig at 90 ~ and a combination of two terms from the journal bearing rig and the film thickness rig respectively. This suggests that mixed hydrodynamic lubrication may be the dominant regime in the engine under these conditions. It was not possible to correlate the other responses at 90 ~ to the bench data, as shown by the low R 2 values for "Steady state 90" and "Combined 90". Table 4 shows that on adding interaction terms to the analysis, it was possible to improve the correlation between "Combined 90" and the bench data significantly. While this process allowed a large improvement in the R 2 value, the addition of interaction terms into the model makes any physical interpretation of the data almost impossible

The validity of this exercise is limited by the small data set (only 11 data points), and the results from this analysis must be understood to be indications at this stage as opposed to wellestablished predictions. 6. CONCLUSIONS Fuel economy benefits in heavy duty diesel engine testing have been observed for oils of lower viscosity grade. Optimising the balance of Newtonian to nonNewtonian viscosity within a grade has been shown to give measurable benefits in fuel economy. The addition of lubricant additives can provide improvements in fuel economy, particularly under transient running conditions. The chemical structure of viscosity modifiers has been shown to have a significant effect on both bearing friction and engine fuel economy, with an

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advantage observed for star polymers and block copolymers over OCPs. Statistical analysis of the data set suggests EHD/HD mixed lubrication are the dominant regimes, but engine test data also shows importance of boundary friction control in heavy duty diesel fuel economy. 7. R E F E R E N C E S

1. "The Role of Lubricant Additives in Improving Fuel Economy for Heavy Duty Diesel Applications", R. C. Castle, S. Arrowsmith, C. J. Locke, C. H. Bovington, Technische Arbeitstagung Hohenheim, Stuttgart, 2002 2. "European Requirements for a Super HighPerformance Diesel Oil", D.C. Colbourne, Truck Technology International, 1988, pp 108111 3. "Engine Friction Lubricant Sensitivities : A Comparison of Modem Diesel and Gasoline Engines", R. I. Taylor, Proceedings of the 11th International Colloquium, Esslingen, "Industrial and Automotive Lubrication", 1998 (published by the Technische Akademie Esslingen, 1998) 4. "Fuel Economy Improvement by Engine and Gear Oils", W.J. Bartz, Proceedings of the 24th Leeds-Lyon Symposium on Tribology (published in Tribology for Energy Conservation, Tribology Series, pp 13-24, 34, 1998, Elsevier, Editor: D.Dowson) 5. "Heavy Duty Diesel Engine Fuel Economy: Lubricant Sensitivities", R. I. Taylor, SAE Spring Fuels and Lubricants Conference, Paris, 2000 6. "The Importance of the Stribeck curve in the Minimisation of Engine Friction", C. Bovington, S. Korcek and J. Sorab; Leeds-Lyon Tribology Symposium, Lyon, Sept 1998