Calculation of crank train friction in a heavy duty truck engine and comparison with measured data

Special Issue Article

Calculation of crank train friction in a

heavy duty truck engine and comparison

with measured data

Phil Carden1,Carl Pisani1,Emmanuel Laine′2,Ian Field2,

Maryann Devine3,Andreas Schoeni4and Peter Beyer4

Abstract

This article describes measurements and calculations of friction losses in the crank train of an IVECO heavy duty truck engine.The main objectives were to quantify the effect on friction of variation in lubricant viscosity and to demonstrate the ability of CAE tools to predict friction at contacts in the crank train.Results of various measurements of friction are described.Cycle-averaged whole engine friction was derived from measurements of cylinder pressure and brake torque. In addition a motored friction teardown test was performed to separate the losses from different engine systems. Commercially available analysis tools were used to calculate friction at contacts in the crank train under the same conditions as the tests and this article describes key assumptions and input data and presents calculated results compared with measured data.

Keywords

Friction,engine,piston,piston rings,bearings,crankshaft

Date received:22March2012;accepted:12June2012

Introduction

Manufacturers of all kinds of automotive engines are currently trying to reduce friction to improve fuel con-sumption and so reduce CO2emissions.One possibil-ity for friction reduction on heavy duty diesel engines is the use of low viscosity oil1and one of the aims of this project was to investigate the e?ect of low viscos-ity oil on the engine friction.

The test engine was an IVECO Cursor13EU4 engine which is a typical heavy duty truck engine of I6con?guration having12.9L displacement with bore of135mm and stroke of150mm.This engine delivers peak torque of2500Nm between1000and1600r/min and peak power of412kW at1900r/min.

Engine friction is di?cult to measure with con?-dence(see Van Basshuysen and Schafer2or Xin3for an introduction to the subject).The various measure-ment techniques for piston friction are reviewed in Noorman et al.4and Nagar and Miers.5Friction between the piston assembly and liner has previously been measured by two methods.The?oating liner method4–9has the advantage that friction force is measured directly by force sensors under the move-able liner but the method has the major disadvantage that it cannot be applied to a production engine and is usually applied to a single cylinder test rig.

The alternative method for measuring piston fric-tion is sometimes referred to as the instantaneous IMEP method10and involves accurate measurement of cylinder pressure,strain in the connecting rod and crankshaft position.The gas force acting on the piston is calculated from the cylinder pressure,the rod force is calculated from the rod strain and the iner-tia force is calculated from the motion of the piston. The friction force is then calculated by taking a bal-ance of forces acting on the piston.This technique has the advantage that it could be applied to a production engine(if telemetry could be used to get the rod strain signal out)but the major disadvantage is the level of 1Ricardo Consulting Engineers Ltd,Bridge Works,Shoreham-by-Sea, UK

2Infineum UK Limited,Milton Hill,Abingdon,UK

3Infineum US,Linden,NJ,USA

4FPT Industrial,Arbon,Switzerland

Corresponding author:

Phil Carden,Ricardo Consulting Engineers Ltd,Bridge Works, Shoreham-by-Sea,BN435FG,UK.

Email:phil.carden@https://www.wendangku.net/doc/f45379691.html,

Proc IMechE Part J:

J Engineering Tribology

227(2)168–184

!Ricardo UK Ltd2012

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DOI:10.1177/1350650112453679

https://www.wendangku.net/doc/f45379691.html,

accuracy that can be expected because the friction is calculated by subtracting one large number from another.11

Both methods have proved useful for software val-idation but there must remain doubts about the dif-ferences in operating conditions between rigs and real engines.Also both methods appear to have increasing problems in making high?delity measurements at high engine speeds and loads when vibration inter-feres with the desired signals.

For the work described here it was essential to use a production heavy duty engine and was also neces-sary to gain some understanding of the overall friction response of the whole engine rather than just the piston.Two di?erent measurement methods were adopted.In the?rst the whole engine friction was deduced from measurements of cylinder pressure and brake torque under a range of di?erent operating speeds and loads.This technique is always subject to large potential errors in the absolute level of friction as discussed in the next section but was expected to have enough resolution to reveal the in?uences of engine load and oil viscosity on friction.

In addition a?ve-stage motored friction teardown test was performed giving information about the variation of friction at sub-system level across the engine speed range with no load under controlled temperature conditions.All tests were performed using two di?erent lubricants.On the?rst stage of the motored test the crank train friction was mea-sured alone(losses due to friction at pistons and crankshaft)with all other sources of loss removed. The errors should be small in this test but,of course the engine is not?red so the in?uence of gas pres-sures is not felt by the contacts in the crank train and the component temperature pro?les,particularly for pistons and liners,are not representative of a ?red engine.Despite these drawbacks this test method is widely used in industry and is very valu-able for ranking engines against one another in terms of friction.The philosophy here was to validate the software by simulating the motored operating conditions.

Crank train friction is also hard to calculate,par-ticularly for the piston skirt and rings on a?red engine,where loads and speeds vary widely and rap-idly during each half stroke and the local oil tempera-ture must be assumed.Ricardo Software codes ENGDYN,PISDYN and RINGPAK were used to create detailed models to solve Reynolds equation and calculate oil?lm thickness at the crankshaft bear-ings,piston skirt/liner contact and piston ring/liner contacts.There is a vast literature on this subject and this has recently been reviewed thoroughly.3 The focus of this article is on the key assumptions and boundary conditions and the subsequent results rather than the details of the mathematical models used in the software,which have been published before.12,16–18Estimation of whole engine friction from measurements of cylinder pressure and brake torque

The friction of the whole engine(FMEP)was deter-mined from measurements of cylinder pressure(used to calculate IMEP)and brake torque(used to calcu-late BMEP)under?red conditions with oil and cool-ant temperature controlled to90?1 C.The cylinder pressure was measured in3of the6cylinders using water cooled Kistler6061B pressure sensors.The brake torque was measured using an HBM in-line torque meter.

Data was recorded at the speed and load points indicated in Table1with the two lubricants with properties described in Table 2.The engine was ?ushed3times and the oil?lter was changed before making measurements with the second lubricant.

It is relatively easy to calculate the friction mean e?ective pressure(FMEP)using equation(1) FMEP?IMEPàBMEPe1THowever,the technique is subject to large potential errors because IMEP and BMEP are both much larger than FMEP.For example if IMEP is21bar and BMEP is20bar and there is?1%error in IMEP and?0.1%error in BMEP then the error in FMEP could be as high as?30%.This means that the results can only be used as a rough guide to the abso-lute level of friction but the resolution and repeatabil-ity should enable sensitivity to variations of load and oil viscosity to be evaluated with better con?dence.

T able1.Load/speed points for fired whole engine testing. Engine speed

(r/min)Part load Full load

800Min load,5,10,15bar18bar 1000Min load,5,10,15,20bar24bar 1200Min load,5,10,15,20bar25bar 1400Min load,5,10,15,20bar25bar 1600Min load,5,10,15,20bar24bar 1800Min load,5,10,15,20bar22bar 2000Min load,5,10,15bar18bar 2200Min load,5,10bar13bar

T able2.Lubricant properties.

Name Baseline

Low

viscosity Kinematic viscosity at40 C(cSt)75.7733.64 Kinematic viscosity at100 C(cSt)12.28 6.53 Density(kg/m3)862.8850.1

Carden et al.169

Figure 1shows the FMEP plotted against speed for various engine load (BMEP)levels.The graph shows the expected trends (friction increases with engine speed and engine load)but the non-smooth curves and the outlying points are representative of the under-lying potential error.Figure 2shows a comparison between the friction levels with the baseline oil and the low viscosity oil at full load and minimum load.This also shows the expected trends (friction was reduced by the lower oil viscosity and a greater reduc-tion was observed at low load where lubrication regime at contacts was more likely to be hydrodynamic).

Measurement of engine friction by motored teardown test

A ?ve-stage motored teardown test was performed.At each stage the oil and coolant temperatures were con-trolled to 90?1 C.The details of the builds were as follows.

Build 1–Crank train only.Cylinder head removed and replaced by a steel plate designed to block oil and coolant passages to the cylinder head and enable cylinder head bolt loads to distort the bores in a real-istic way but with large holes above the cylinders so pumping losses are avoided.Valve train,timing drive,oil pump,coolant pump and FEAD belt were not ?tted.

Build 2–As Build 1but with pistons and connect-ing rods removed and replaced by dummy weights bolted to the crank pins.

Build 3–As Build 2but with cylinder head,timing drive gears,camshaft and valve trains ?tted.Injector rockers were not ?tted.

Build 4–As Build 3but with oil pump ?tted.External oil circuit still in place to condition the oil to 90?1 C.

Build 5–As Build 4but with coolant pump and other FEAD belt and gear-driven auxiliaries ?tted.The auxiliaries were alternator,compressor,air con-ditioning pump,power-assisted steering pump and low pressure fuel pump and none of these were loaded during the test.

For this motored test an-line torque meter with a range of 500Nm and accuracy of 0.1%was chosen.The oil and coolant temperatures were controlled to within ?1 C using Ricardo test bed heating/cooling control units and at each build stage the engine was motored for 30min at 1200r/min to stabilise before taking measurements.The oil gallery pressure was manually set at each speed to achieve the same pres-sure (?0.05bar)as that measured on the ?red engine.Each measured torque value was the average of 3points taken while going up and down the speed range at intervals of 200r/min.Each build was mea-sured with both lubricants as shown in Table 2and ?ushed 3times

beforehand.

Figure 1.Measured whole engine friction with baseline

lubricant.

Figure 2.Effect of oil viscosity on friction at full load and minimum load.

170Proc IMechE Part J:J Engineering Tribology 227(2)

The results were processed to give losses for the following groups by subtraction as necessary:.Crankshaft group loss from Build 2directly .Reciprocating group loss from Build 1àBuild 2.Valvetrain group loss from Build 3àBuild 2.Oil pump group loss from Build 4àBuild 3.

Auxiliaries group loss from Build 5àBuild 4

Figure 3shows the distribution of measured losses under motored conditions from the various groups across the speed range using the baseline lubricant.The losses due to the crank train group,which were the focus of this study,were given by directly from Build 1and include losses of the crankshaft group and the reciprocating group.It can be seen on Figure 3that the crank train losses account for 55%of the total at 1000r/min and 57%of the total at 2000r/min.Figure 4shows the measured losses for the crank train group with baseline and low viscosity lubricants.The low viscosity lubricant reduced crank train losses under motored conditions by 19%at 1000r/min and 15%at 2000r/min.

The crank train group includes the friction losses at top piston ring/liner contact,second piston ring/liner contact,oil control ring/liner contact,piston skirt/liner contact,small end bearings,big end bearings,main bearings and crankshaft seals.In addition wind-age losses as the crankshaft and connecting rod move through crankcase gas were present together with losses due to pumping of the crankcase gas from one bay to the next as pistons move up and down.Further losses due to shearing of the lubricant as it escaped through the thrust gaps at each big end and main bearing may also have been present.

It is not possible to separate these detailed losses by any known measurement technique but calculations can potentially o?er some insight into their relative magnitudes.

Calculation of piston ring friction

Piston ring friction was calculated using Ricardo RINGPAK software.This program has fully coupled and integrated models for ring axial and twist dynam-ics,inter-ring gas dynamics,ring radial dynamics,lubrication at the ring/liner interface,oil consumption and wear.There is not enough space in this article to describe the model fully but this has been done previously.12

RINGPAK was used to calculate the oil ?lm thick-ness and friction power loss at each piston ring through one engine cycle of 720 .The following key assumptions were made:

.Secondary motion of the piston was ignored in the ring analysis.

.Measured mean cylinder pressure data were used to excite the model for ?red engine cases and for the motored case without cylinder head the pres-sure above the piston was assumed to be atmos-pheric.Figure 5shows the peak values of cylinder pressure across the speed range.

.Inertial forces due to the piston and connecting rod were included and for the motored cases these were the only applied

forces.

Figure 4.Measured crank train loss from motored teardown test with two

lubricants.

Figure 3.Measured losses from motored teardown test with baseline lubricant.

Carden et al.171

.Shear thinning of the lubricant was ignored because it was a small e?ect for the oils used.

.Thickness of the oil on the liner before the oil ring on the downstroke was assumed to be5m m and thickness of the oil?lm presented to the other rings was calculated internally in RINGPAK.

.Composite surface roughness was calculated assuming liner roughness of0.2Rpk and ring face roughness of0.5Ra..The calculated oil?lm thickness at each piston ring was divided by the composite surface roughness at each time step.If the result was less than4then the Greenwood–Tripp algorithm13was used to calcu-late the asperity contact load.The asperity contact friction force was then calculated by multiplying the asperity contact force by an assumed friction coe?cient of0.12(typical value for boundary lubricated contact from Xin3)and integrating over the contact area.

.For motored cases the local oil temperature in the contact was assumed to be90 C.Since in this test the oil temperature in the main gallery was con-trolled to90 C the implication here was that the drop in oil temperature between main gallery and liner was matched by the increase in oil tempera-ture due to shear heating between piston and the liner.

.For full load and minimum load cases the local oil temperature in the contact region was assumed to be the average of the liner temperature and the ring temperature.Graphs showing typical assumed variation in these temperatures are shown in Figures6and7.These assumed temperatures were based on measured data from a similar engine.In practice the shear heating e?ect will raise the local oil temperature to a slightly higher value than that of the surfaces but this was not accounted for in this analysis.

.The ring face pro?les were taken from the drawings and so represent new pro?les rather than worn

pro?les.

Figure7.Assumed liner

temperature.

Figure6.Assumed ring

temperatures.

Figure5.Peak cylinder pressure.

172Proc IMechE Part J:J Engineering Tribology227(2)

There is not enough space in this article to include all the input data.

Some typical RINGPAK results are given in Figures 8,9and 10.The results presented in these graphs are at 1600r/min with baseline lubricant but similar calculations were made at other speeds and using the low viscosity oil.

Figure 8shows the calculated minimum oil ?lm thickness,friction force and friction power loss

against crank angle for the top piston ring under motored conditions,minimum load conditions and full load conditions.Under motored conditions the low oil temperature results in large oil ?lm thickness at each mid stroke and each stroke had similar level of power loss.Under full load conditions the oil tem-perature was much higher at the top ring and this led to a lower level of oil ?lm thickness in each stroke.At the start of the expansion stroke the cylin-der pressure was very high and this led to

high

Figure 8.Minimum oil film thickness,friction force and friction power loss at top ring/liner contact at 1600r/min under various load conditions.

Carden et al.173

pressure behind the top ring.The combination of high pressure and high oil temperature led to predicted period of boundary lubrication and thus the calcu-lated friction power loss at the top ring was signi?-cantly a?ected by the engine load.

The analysis predicts that the asperity contact fric-tion power loss at the top ring during the ?rst half of the expansion stroke is signi?cant.The authors expect that this predicted friction power loss would be lower than that indicated if the worn pro?le of the top ring was used instead of the new pro?le as indicated in Priest et al.14but there was no opportunity to measure worn ring pro?les in this study.

Figure 9shows the calculated minimum oil ?lm thickness,friction force and friction power loss against crank angle for the second piston ring

under

Figure 9.Minimum oil film thickness,friction force and friction power loss at second ring/liner contact at 1600r/min under various load conditions.

174Proc IMechE Part J:J Engineering Tribology 227(2)

motored conditions,minimum load conditions and full load conditions.This showed a similar trend to the top ring but the e?ect of the gas pressure was much reduced.The level of calculated minimum oil ?lm thickness at the two top rings was similar to results presented in Tian.15

Figure 10shows the calculated minimum oil ?lm thickness,friction force and friction power loss against crank angle for the upper rail of the oil control ring under motored conditions,minimum load conditions and full load conditions.The lower rail had a very similar response.In this case the friction power loss was a?ected by tem-perature assumptions as load was increased but not by the cylinder pressure.The thin rail face and the higher ring tension combined to result in lower

?lm

Figure 10.Minimum oil film thickness,friction force and friction power loss at oil control ring top rail/liner contact at 1600r/min under various load conditions.

Carden et al.175

thickness and so some asperity contact friction was predicted throughout each stroke.

Figure 11shows the e?ect of engine speed and load on the combined cyclic averaged power loss at the piston rings for one cylinder.Results indicate that under motored conditions the piston ring friction rises strongly with speed and the friction loss with low viscosity oil was signi?cantly lower than that with the baseline oil.In this case the lubricant tem-perature was assumed to be 90 C at all speeds and for both oils.Under minimum load conditions at low engine speed the calculated piston friction loss was similar to the motored condition but at high speed the predicted friction loss was lower.At full load con-dition the predicted piston ring friction was signi-?cantly increased.Under ?red (full load and minimum load)conditions the predicted piston ring friction with low viscosity oil was increased compared with the baseline.

Calculation of piston skirt friction

Piston skirt/liner friction was calculated using Ricardo PISDYN software.There is not enough space in this article to describe the model fully but this has been done previously.16,17

PISDYN was used to calculate the oil ?lm thick-ness and friction power loss on both sides of the piston skirt at intervals of 1 through one engine

cycle of 720 .The following key assumptions were made:

.Skirt/liner gap was assumed to be fully ?ooded with oil.

.Measured mean cylinder pressure data were used to excite the model for ?red engine cases and for the motored case without cylinder head the pres-sure above the piston was assumed to be atmos-pheric.Figure 5shows the peak values of cylinder pressure across the speed range.

.Inertial forces due to the piston and connecting rod were included and for the motored cases these were the only applied forces.

.Shear thinning of the lubricant was ignored because it was a small e?ect for the oils https://www.wendangku.net/doc/f45379691.html,posite surface roughness was calculated assuming a worn skirt roughness of 3.9Ra and liner roughness of 0.2Rpk.

.The calculated oil ?lm thickness at each location on the piston skirt was divided by the compos-ite surface roughness at each time step.If the result was less than 4then the Greenwood–Tripp algorithm 13was used to calculate the asperity contact load.The asperity contact friction force was then calculated by multiplying the asperity contact force by an assumed friction coe?cient of 0.12(typ-ical value for boundary lubricated contact from Xin 3)and integrating over the contact area.

.For motored cases the local oil temperature in the contact was assumed to be 90 C.Since in this test the oil temperature in the main gallery was controlled to 90 C the implication here was that the drop in oil temperature between main gallery and liner was matched by the increase in oil temperature due to shear heating between piston skirt and the liner..For full load and minimum load cases the local oil temperature in each calculation cell in the skirt/liner contact zone was assumed to be the average of the local liner temperature (Figure 7)and the skirt tem-perature (see Table 3for assumed skirt temperatures based on measured data from a similar

engine).

Figure 11.Effect of load,speed and oil viscosity on piston ring friction power loss.

T able 3.assumed piston skirt temperature data.

Piston skirt temperature ( C)

Speed (r/min)Motored Minimum load Full load 80090901011000909210212009092102140090941021600909510218009098103200090991042200

90

98

103

176Proc IMechE Part J:J Engineering Tribology 227(2)

.The cold piston skirt pro?le was taken from the drawing and adapted to the hot shape using FE models of the piston with di?erent imposed piston temperature distribution for each load/speed case.

.Circumferential bore distortion was ignored (partly because FE models of the complete engine struc-ture were not available and partly because this engine has wet liners so circumferential bore dis-tortion should be minimal).

.Axial bore distortion was modeled assuming a simple conical shape due to thermal expansion (higher bore temperature at the top of the liner than the bottom).A straight line was ?tted through the data shown in Figure 7and the high temperature region at the top of the bore

was

Figure 12.Side force,minimum oil film thickness and friction power loss the piston skirt at 1600r/min under various load conditions.

Carden et al.177

ignored as skirt/liner contact does not occur in this region.

There is not enough space in this article to include all the input data.

Some typical PISDYN results are given in Figure 12.The results presented in these graphs are at 1600r/min with baseline lubricant but similar cal-culations were made at other speeds and using the low viscosity oil.

Under motored conditions the piston side force was similar in each stroke and governed by inertial e?ects.The graph showing minimum oil ?lm thick-ness requires careful interpretation because it shows the minimum value found by PISDYN after consider-ing all locations on the piston at each crank angle.Thus the location of the minimum ?lm thickness changes from one crank angle to the next.The graph indicates that low ?lm thickness occurs shortly

after the start of each upstroke (compression and exhaust)and this is related to predicted piston tilting at these locations.In addition low ?lm thickness occurred before and after TDC ?ring under full load conditions due to the increased gas forces.Low ?lm thickness resulted in periods of predicted boundary lubrication and this led to predicted periods of asper-ity contact power loss but the overall mean piston skirt losses were dominated by hydrodynamic losses with peaks in power loss at the centre of each stroke as shown in the graph.Short periods of high min-imum ?lm thickness are also indicated (mainly close to TDC)and under these conditions a full oil ?lm may not be maintained in the gap between skirt and liner.Figure 13shows a typical map of hydrodynamic pressure on the thrust side of the piston at 90 after TDC on the expansion stroke.The ?gure shows half of the skirt and the model assumes the skirt is sym-metric.The grid area indicates the area of the lubri-cation mesh on the skirt.The contours indicate

areas

Figure 13.Map of hydrodynamic pressure on thrust side of piston skirt at 90 after TDC at full load and 1600

r/min.

Figure 14.Map of oil film thickness on thrust side of piston skirt at 90 after TDC at full load and 1600r/min.

178Proc IMechE Part J:J Engineering Tribology 227(2)

Figure 17.Effect of load,speed and oil viscosity on small end bearing friction power

loss.

Figure 16.Effect of load,speed and oil viscosity on crank bearing friction power

loss.

Figure 15.Effect of load,speed and oil viscosity on piston skirt friction power

loss.

Figure https://www.wendangku.net/doc/f45379691.html,parison of measured and calculated losses for crank train group under motored conditions with baseline oil.

Carden et al.179

of equal pressure.Figure 14shows the corresponding map of oil ?lm thickness at the same instant.

Figure 15shows the e?ect of engine speed and load on the cycle averaged power loss at the piston skirt for one cylinder.Results indicate that under motored conditions the piston skirt friction rises strongly with speed and the friction loss with low viscosity oil was signi?cantly lower than that with the baseline oil.In this case the lubricant temperature was assumed to be 90 C at all speeds and for both oils.Under minimum load conditions at low engine speed the calculated piston skirt friction loss was similar to the motored condition but at high speed the predicted loss was lower mainly due to the e?ect of the hotter oil at mid stroke.At full load condition the predicted piston skirt friction was signi?cantly increased and there was not much di?erence between baseline and low viscosity oil.The implication here was that the predicted reduction in hydrodynamic loss was balanced by an increase in asperity contact loss.

Calculation of bearing friction

Friction at main bearings,big end bearings and small end bearings was calculated using Ricardo ENGDYN software.This program has many options for alterna-tive models for bearing calculations.There is not enough space in this article to describe the models fully but this has been done previously.18

For the results presented in this article the follow-ing key assumptions were made:

.Main bearing loads were calculated using a dynamic model based on sti?ness-quality ?nite element models of the crankshaft and cylinder block.In this way the load sharing at the main bearings and the 3D vibration e?ects were accounted for.The ?ywheel and TV damper were also included in the model.

.Measured mean cylinder pressure data were used to excite the model for ?red engine cases.

.Inertial forces due to the piston,connecting rod and crankshaft were included and for the motored cases these were the only applied forces.

.Each main and big end bearing was modelled using Booker mobility method with Okvirk short bearing theory 19so calculated results include hydro-dynamic shear losses only and not asperity contact losses.Previous work using more sophisticated and time-consuming models suggested that this assumption was justi?ed in this case because the calculated oil ?lm thickness under the conditions of the test was always high enough to avoid signi?-cant asperity contact with oil supply temperature of 90 C.

.Although the piston pin was in fact a ?oating type it was assumed to be ?xed in the piston so the rela-tive motion at the small end occurred only between the pin and the connecting rod.

.The small end bearing was modelled using a ?nite di?erence solver for Reynold’s equation so when oil ?lm thickness was less than 4times the compos-ite surface roughness the asperity contact load was calculated using the Greenwood–Tripp algo-rithm.13The asperity contact friction force was then calculated by multiplying the asperity contact force by an assumed friction coe?cient of 0.12(typical value for boundary lubricated contact from Xin 3)and integrating over the contact area..Each oil was modelled using the relevant tempera-ture/viscosity curve calculated using the Walther equation and shear thinning was ignored because for the oils used it was a small e?ect.

.Perfect alignment of bearings and perfectly round bearings and journals were assumed.

.A thermal balance was included (so the e?ective oil temperature was calculated from oil supply tem-perature,calculated ?ow rate and calculated power loss due to oil shear).

.Mean bearing clearance was used at each bearing after measurements were made to validate this assumption.

.The measured oil supply pressure was entered at each speed Figure 16shows the e?ect of engine speed and load on the cycle averaged power loss for all main bearings and big end bearings.Results indicate that

under

Figure https://www.wendangku.net/doc/f45379691.html,parison of measured and calculated losses for crank train group under motored conditions with low viscosity oil.

180Proc IMechE Part J:J Engineering Tribology 227(2)

motored conditions the crankshaft bearing friction rises strongly with speed and the friction loss with low viscosity oil was signi?cantly lower than that with the baseline oil.In this case the lubricant supply temperature was assumed to be90 C at all speeds and for both oils.Under minimum load con-ditions at low engine speed the calculated crankshaft bearing friction loss was similar to the motored con-dition but at high speed the predicted loss was slightly lower partly due to the hotter oil and partly due to cancellation of gas and inertia forces during parts of the cycle.At full load condition the predicted crank-shaft bearing friction was increased due to higher loads during the expansion stroke.

Figure17shows the e?ect of engine speed and load on the cycle averaged power loss for one small end bearing.Results indicate that under motored condi-tions and minimum load conditions the power loss increased with speed and the low viscosity oil had a small e?ect on the friction level.However under full load conditions higher losses were predicted at the small end bearing and these were dominated by asper-ity contact loss rather than hydrodynamic e?ects.This was expected.20However despite this increase in pre-dicted friction the contribution of the small end bear-ing loss to the total friction was still very small (Figures18and19).The calculated friction at the small end bearing with low viscosity oil was increased. The small end bearing is one location that may bene?t from the use of advanced coatings to prevent wear if low viscosity oil is used.

Comparison of measurements with calculated results

The calculated power losses for the contacts under motored conditions were adjusted so that they accounted for the losses of the whole crank trains measured on the motored test(6sets of pistons, rings,small end bearings,big end bearings and

https://www.wendangku.net/doc/f45379691.html,parison between measured and calculated losses for the whole engine under fired conditions at minimum load with baseline oil.

Carden et al.181

7main bearings)and the power losses were converted into FMEP values in bar.The results are shown com-pared with measured data from the motored tests in Figures18and19for baseline and low viscosity oils, respectively.In each case the sum of the calculated losses was lower than the measured loss.This is likely to be partly because other power losses,such as crankshaft seal friction,crank shaft windage, crankcase pumping,oil churning,thrust bearing losses,etc.,were present on the actual engine.The response of the sum of the calculated values to engine speed is reasonably good(slope of calculated line is similar to measured slope).The comparison with baseline oil shows better agreement between the measured and calculated results than that with low viscosity oil.This indicates that the analytical methods have over-predicted the in?uence of the low viscosity oil as the gap between measured and calculated data was larger in Figure19.

Direct comparison between the calculated crank train friction under?red conditions(minimum load and full load)and measured data was not possible because the?red measurement technique used could only give measured friction for the whole engine.To circumvent this problem and provide a partial val-idation under?red condition it was assumed that the losses at the valve train group,oil pump,coolant pump and front end accessory drive(FEAD)were not sensitive to engine load and so the same as the values measured on the motored teardown test.

Figure20shows the measured losses for the whole engine under minimum load conditions with baseline oil compared with the sum of calculated crank train losses under minimum load conditions,measured losses at the valve train group,oil pump,coolant pump and FEAD and measured losses due to the fuel injection equipment(provided by IVECO). Figure21shows similar data for the baseline oil under full load conditions.

It is important to bear in mind the signi?cant potential error in the absolute measured data for the whole engine friction.However,Figures20and21 demonstrate a reasonable level of agreement between calculations and measurements under?red engine conditions.

Figure22shows calculated total crank train fric-tion presented as FMEP in bar under?red and motored conditions for the two di?erent oils.Thus

https://www.wendangku.net/doc/f45379691.html,parison between measured and calculated losses for the whole engine under fired conditions at full load with baseline oil.

182Proc IMechE Part J:J Engineering Tribology227(2)

showing the sensitivity of the models to load,speed and oil viscosity.

Conclusions

The following conclusions can be drawn:

.Measured whole engine friction data obtained under?red conditions indicated the expected trends in terms of variation with load,speed and oil viscosity but the inherent inaccuracy of the method dictates caution when interpreting the results.

.Measured data from the motored teardown test indicated the contribution of the crank train to the total friction under motored conditions and also quanti?ed the potential reduction in crank train friction due to the low viscosity lubricant.

.Analysis was used to quantify friction at each piston ring,piston skirt,each main bearing,big end bearing and small end bearing and their vari-ation with load,speed and oil viscosity.

.Total calculated crank train friction was lower than the measured data obtained from the motored tear-down test but some known sources of loss were not included in the calculations so this was expected. .However the total calculated crank train friction showed a greater response to the oil viscosity than the measurement(i.e.the gap between meas-urements and calculations was greater with the low viscosity oil)..CAE models were also used to calculate total crank train friction loss under?red engine conditions (minimum load and full load)and the results were compared with measured data for whole engine friction(measured motored data for those systems not expected to be sensitive to engine load were used in the comparison).This work showed reasonable level of agreement.

Funding

This research received no speci?c grant from any funding agency in the public,commercial,or not-for-pro?t sectors. Acknowledgements

The authors thank the senior management at Ricardo, In?neum and FPT Industrial for permission to publish this work.Thanks also to David Bell,Jon Plail and Jan Hronza of Ricardo Software for assistance with CAE model development.

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体中所占的重要程度，具有较强的说服力，被广泛应用于确定改革的主要目标和效果、调查产生缺陷及故障的原因。因此，企业管理人员必须掌握排列图的绘制，并将其应用到质量过程中去。 一般说来，绘制排列图的步骤如图7-1所示，即：确定调查事项，收集数据，按内容或原因对数据分类，然后进行合计、整理数据，计算累积数，计算累积占有率，作出柱形图，画出累积曲线，填写有关事项。 图7-1 排列图的绘制步骤 排列图的应用实例 某化工机械厂为从事尿素合成的公司生产尿素合成塔，尿素合成塔在生产过程中需要承受一定的压力，上面共有成千上万个焊缝和焊点。由于该厂所生产的十五台尿素合成塔均不同程度地出现了焊缝缺陷，由此对返修所需工时的数据统计如表7-1所示。 表7-1 焊缝缺陷返修工时统计表 序号项目返修工时fi 频率 pi/% 累计频率 fi/% 类别 1焊缝气孔14860.460.4A 2夹渣5120.881.2A 3焊缝成型差208.289.4B 4焊道凹陷15 6.195.5B 5其他11 4.5100C 合计245100 缝成型差、焊道凹陷及其他缺陷，前三个要素累加起来达到了89.4%。根据这些统计数据绘制出如图7-2所示的排列图：横坐标是所列举问题的分类，纵坐标是各类缺陷百分率的频数。

班主任激励学生的14种方法

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MATLAB实验：运算基础,并且附有答案

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572 633 694 755 816 877 >> b=a(ans) b = 122 183 244 305 366 427 488 549 610 671 732 793 854 915 976 >> length(b) ans = 15 2）建立一个字符串向量，删除其中的大写字母（提示：利用find函数和空矩阵。）a=’I am maying’; a( find(a>’A’&a<’Z’))=[] 3、已知A=[23 10 -78 0；41 -45 65 5；32 5 0 32； 6 -54 92 14]，取出其前3行构成矩阵B，其前两列构成 矩阵C，其左下角3x2子矩阵构成矩阵D，B与C的乘积构成矩阵E，分别求E>B=A([1 2 3 ],:)>> C=A(:,[1 2])>> D=A([2 3 4 ],[1 2 ])>> E=B*C >> E

激励教育的方法

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激励中学生的四个有效方法

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全面质量管理的基本方法

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