calculation of friction in high performance engines

8/20/2019 Calculation of Friction in High Performance Engines

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www.ricardo.com

© Ricardo plc 2010RD.10/174701.1

Calculation of Friction in High Performance EnginesRicardo Software European User Conference 2010

Phil Carden

Ricardo UK


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2 © Ricardo plc 2010RD.10/174701.120th April 2010

Presentation contents

Introduction

Some general comments and issues

Semi-empirical models FAST

Piston assembly friction PISDYN and RINGPAK

Fluid film bearing friction ENGDYN

Crankcase pumping loss WAVE

Valve train and timing drive friction VALDYN

Case study

Conclusions


3/47


Introduction

Minimising friction in motorsport engines is vital because

– power is required elsewhere…

– high friction often means high wear

– high friction leads to unnecessary heat generation and greater need for cooling

Knowledge of friction level is required

– to evaluate potential of alternative low friction designs, coatings etc

– as input to performance simulation models Measurement of engine friction is

– highly problematic, particularly at high engine speed

– the results are often obtained too late to make key design changes

– but measurements are useful for validation of calculation methods

Analytical tools are beginning to reach required level of fidelity – but there are still some problems and mysteries…


4/47



Introduction







Case study

Conclusions


5/47


Which losses are included in “engine friction” ?

Pumping of gas above pistonLosses due to pumping ofcrankcase gas below pistons

Windage losses in engine

Oil churning losses in engine

Power required to drive

auxiliaries (oil pump, water

pump, alternator)

Transmission losses (gears,

bearings, seals, windage, oil

churning etc)

Piston ring/liner friction

Piston skirt/liner friction

Small end bearing friction

Big end bearing friction

Main bearing friction

Thrust bearing friction

Valve train friction

Camshaft bearing friction

Timing drive friction

Balancer bearing friction

Seal friction Auxiliaries friction (oil pump,

water pump, alternator,

scavenge pumps etc)

Definitely not includedOptionalDefinitely included


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0

0.02

0.04

0.06

0.08

0.1

0.12

0.14

0.16

0.18

0.2

0.22

0 2 4 6 8 10 12 14 16 18 20

Duty parameter (viscosity x speed / load)

or specific film thickness (film thickness / surface roughness)

f r i c t i o n c o e f f i c i e n t

Stribeck curve

Friction coefficient at any lubricated contact depends on relative speed, normal load, oil

viscosity, shape of contacting parts, roughness of contacting parts, temperature,

materials of contacting parts etc

Boundary

lubricationMixed

lubricationElastohydrodynamic

lubricationHydrodynamic

lubrication

Cam/follower

Piston rings / liner

Crankshaft bearings


7/47


How to measure friction ?

FMEP is small compared to IMEP and

BMEP so any measurement error is

magnified

Use of more intrusive instrumentation

usually involved changes to componentsand this can affect durability

When the whole engine is motored then

pumping loss cannot be separated from

friction by measurement

No combustion so gas load due to

compression only

No way to account for increased local

temperatures due to combustion

Issues

Operate engine normally

Measure cylinder pressure (with very

good knowledge of TDC) and brake torque

as accurately as possible

Calculate FMEP from IMEP – BMEP

Can extend test to make more detailed

measurements

Simply connect engine to motor

Drive the engine without firing and

measure the torque

Need to control temperature of oil and

water carefully

Test can be extended by progressively

stripping engine to give approximatebreakdown in friction by subsystem

Description

Fired testMotored test


8/47


What kinds of calculation are performed ?

Potentially long set up times

Potentially long run times

Need different tools to analyse all

contacts

Some losses are very difficult to model

Easy to lose sight of “big picture”

Need lots of test data to develop tool

Accuracy depends of how similar

engine is to those in database

Cannot accurately predict effect of

detail design changes

Limitations

Potentially good accuracy for anycontact if suitable model is available

Can account for differences in load,

detailed component geometry, oil grade,

temperature etc

Very quick to set up and run

Can potentially account for most

sources of loss

Advantages

Advanced CAE approachSemi-empirical approach


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0

0.2

0.4

0.60.8

1

1.2

1.4

1.6

1.8

2

2.2

0 5 10 15 20 25 30 35 40 45 50 55 60

Time (hours)

F M E P ( b a r )

Some special problems of motorsport engines

Friction testing usually calls for prolonged operation

under stabilised conditions – Many motorsport engines cannot be operated under

stabilised conditions at high speed/load

– During a race the piston temperature varies

significantly and can approach the limit at end of

straights after ~15 seconds

– If the engine is held at full load the piston temperaturewill rise significantly affecting clearances, local oil

temperature and so piston friction before failure

Motorsport engines are often not run-in properly before

use

– In typical passenger car engine friction reduces from

initial value for at least 20-30 hours

– Motorsport engines are sometimes used straight from

new build so friction level is reducing all the time

0

50

100

150

200

250

300

350

400

450

0 10 20 30 40 50 60 70 80 90 100 110

time (sec)

a v e r a g e p i s t o n t e m p e r a t u r e ( d e g C )

Simulation of 1 lap at Monza

Measurement on passengercar engine at 4000 rpm

5% reduction in 40 hours


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Introduction







Case study

Conclusions


11/47


Semi-empirical models - FAST

FAST (Friction Assessment Simulation Tool) is a semi-empirical engine friction

prediction program used internally by Ricardo

FAST predicts friction loss across the engine speed range for individual subsystems

and for the whole engine

FAST uses semi-empirical equations

– Coefficients and exponents were developed by Ricardo to give improved agreement

with the latest measured data obtained from motored teardown tests on modern

passenger car engines and with other published measured data

– New equations were developed for systems not covered by published sources

– Latest version extended to enable prediction of friction under fired conditions

FAST currently requires ~50 input values such as bore, stroke, bearing diameter, valve

train type etc. to make a prediction of motored friction loss

– All input values can be obtained easily from drawings or components – More detailed information such as cylinder pressure, piston mass, rod mass, ring

tension etc. is required to calculate the increase in piston friction under fired engine

conditions


12/47


Typical results for reciprocating group

Predictionsunder motored

conditions

compared with

measured data

from motored

teardown test

Prediction

under full load

conditions

shows

significant

increase in

friction at top

piston ring and

skirt0

0.1

0.2

0.3

0.4

0.5

0.6

0.7

0.8

0.9

1

1 0 0 0

1 5 0 0

2 0 0 0

2 5 0 0

3 0 0 0

3 5 0 0

4 0 0 0

4 5 0 0

5 0 0 0

5 5 0 0

6 0 0 0

6 5 0 0

7 0 0 0

Engine speed (rpm)

G a s o l i n e r e c i p r o c a t i n g g r o u p f r i c t i o n - m o t o r e d ( b a r )

CPMEP

FIRST RING

SECOND RING

OIL CONTROL RING

PISTON SKIRT

BIG END BEARING

MEASURED

0

0.1

0.2

0.3

0.4

0.5

0.6

0.7

0.8

0.9

1

1 0 0 0

1 5 0 0

2 0 0 0

2 5 0 0

3 0 0 0

3 5 0 0

4 0 0 0

4 5 0 0

5 0 0 0

5 5 0 0

6 0 0 0

6 5 0 0

7 0 0 0

En ine s eed r m

G a s o l i n e r e c i p r o c a t i n g g r o u p f r i c t i o n - f u l l l o a d ( b a r )

CPMEP

FIRST RING

SECOND RING

OIL CONTROL RING

PISTON SKIRT

BIG END BEARINGS


13/47


Typical results for other systems

Resultsaccurate to

+/-20% for

each

subsystem

and often

much better

than this0.00

0.05

0.10

0.15

0.20

0.25

0.30

0.35

1000 1500 2000 2500 3000 3500 4000 4500 5000 5500 6000 6500

Engine speed (rpm)

V 6 g a s o l i n e v a l v e t r a i n g r o u p f r i c t i o n ( b a r ) TIMING BELT

DIRECT ACTING HYDRAULIC VALVETRAINSCAMSHAFT SEALSCAMSHAFT BEARINGSMEASURED DATA

0

0.02

0.04

0.06

0.08

0.1

0.12

0.14

0.16

0.18

0.2

1000 1500 2000 2500 3000 3500 4000 4500 5000 5500 6000

Engine speed (rpm)

I 4 g a s o l i n e a l

t e r n a t o r f r i c t i o n ( b a r )

FAST V2

Measured data

0

0.02

0.04

0.06

0.08

0.1

0.12

0.14

0.16

0.18

0.2

1000 1500 2000 2500 3000 3500 4000 4500 5000 5500 6000

Engine speed (rpm)

I 4 g a s o l i n e o

i l p u m p f r i c t i o n ( b a r )

FAST V2

Measured data

0

0.05

0.1

0.15

0.2

0.25

0.3

0.35

1000 1500 2000 2500 3000 3500 4000 4500 5000 5500 6000 6500

Engine speed (rpm)

V 6 g a s o l i n e c r a n k s h a f t g r o u p f r i c t i o n

( b a r )

WINDAGE

SEALS

MAINS BEARINGS

MEASURED DATA


14/47


Typical results for high performance sports car engine

For wholeengine

friction the

results are

generally

accurate

to within

+/-10% FAST has

not been

validated

for

engines

designed

to rev

higher

than 8000

rpm

0

0.2

0.4

0.6

0.8

1

1.2

1.4

1.6

1.8

2

2.2

2.4

1 0 0 0

1 5 0 0

2 0 0 0

2 5 0 0

3 0 0 0

3 5 0 0

4 0 0 0

4 5 0 0

5 0 0 0

5 5 0 0

6 0 0 0

6 5 0 0

7 0 0 0

7 5 0 0

8 0 0 0

Engine speed (rpm)

W h o l e e n g i n e - n o l o a d ( b a r )

Auxiliaries group

Crankshaft group

Valvetrain group

Reciprocating group

0

0.2

0.4

0.6

0.8

1

1.2

1.4

1.6

1.8

2

2.2

2.4

1 0 0 0

1 5 0 0

2 0 0 0

2 5 0 0

3 0 0 0

3 5 0 0

4 0 0 0

4 5 0 0

5 0 0 0

5 5 0 0

6 0 0 0

6 5 0 0

7 0 0 0

7 5 0 0

8 0 0 0

Engine speed (rpm)

W h o l e e n g i n e - f u

l l l o a d ( b a r )

Auxiliaries group

Crankshaft group

Valvetrain group

Reciprocating group


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Introduction







Case study

Conclusions


16/47


Piston assembly friction prediction

Piston assembly friction

– involves losses at many sliding interfaces – is very difficult to measure even if cost is not an issue

Analysis can offer insights not possible by measurement

Need for validated software but

– Losses and rings and skirt are inter-related

– Some key inputs are hard to determine

• Temperature of surfaces and oil

• Film thickness on liner

• Worn surface shapes of skirt and rings

– The lubrication regime, and so the effective friction

coefficient, , can change dramatically during each

half stroke• boundary at end of stroke ( = ~0.1)

• hydrodynamic at mid stroke ( = ~0.005)

rings

skirt


17/47


Ricardo Software - PISDYN and RINGPAK

PISDYN can calculate

– friction loss due to oil shearingeffects and asperity contact

• at skirt/liner contact

• at small end bearing

– piston secondary motion

– wear loads

RINGPAK can calculate

– friction loss due to oil shearing

effects and asperity contact

• at each ring/liner contact

– wear loads – oil consumption

– blow-by and blow-back

Cylinder oil

evaporation

Oil supplyPiston tilt

Dispersed

oil flow

Ring

dynamics

Pin bearing

eccentricity

Ring lubrication,

friction

Inter-ring

gas flows

Ring - bore

conformability

Oil supplyPiston tilt

Dispersed

oil flow

Ring

dynamics

Pin bearing

eccentricity

Ring lubrication,

friction

Inter-ring

gas flows

Ring - bore

conformability

Asperity

Contact

Skirt Oil Film

Hydrodynamics

(Pressure)

Secondary Motions

Cut OutsElasticDeformation

Bore Distortion

Crown Land

ProfilesCrown Contact

and Friction

Thermal / Pressure

Inertial Skirt

Deformation

Bore Temps.

(Viscosity effect)

Lubrication

CylinderDynamics

Top Ring Groove

Friction Force

Asperity

Contact

Skirt Oil Film

Hydrodynamics

(Pressure)

Secondary Motions

Cut OutsElasticDeformation

Bore Distortion

Crown Land

ProfilesCrown Contact

and Friction

Thermal / Pressure

Inertial Skirt

Deformation

Bore Temps.

(Viscosity effect)

Lubrication

CylinderDynamics

Top Ring Groove

Friction Force


18/47


Model validation at low engine speed

Predictions validated at low engine speed (up to 2500

rpm) using force difference method Measured cylinder pressure, connecting rod strain,

crankshaft angular position, liner temperature, piston

temperature etc.

Piston friction force calculated from gas force, inertia

force and rod force

Future plans for validation at higher speed as part ofENCYCLOPAEDIC project

Results at2000 rpmfull load


19/47


Experience at high speed

Ricardo have used PISDYN and

RINGPAK to calculate piston friction at

very high engine speeds

– Credible ring/liner friction results

obtained from RINGPAK

– But at high speed losses are

dominated by piston/skirt asperity

contact which is predicted to occur in

each stroke even for a well-developedpiston skirt profile

Skirt/liner friction power loss shows strong

sensitivity to

– component temperature

• governs clearance, component

distortion, and local oil temperature

– oil supply assumptions

– component surface texture

rings

rings

skirt


20/47


Predicted result shows asperity contact pressure at 15000 rev/min

High Speed Case Study

Wear region above rings

Each skirt has patches of wear,

either side of the central region

AndCentrally under the rings


21/47



Introduction






Case study

Conclusions


22/47


Fluid film bearing friction

Journal bearings normally operate in the

hydrodynamic lubrication regime

Losses are dominated by oil shearing effects

– Relatively easy to quantify using simple

bearing analysis methods

– Can be used to study the effects of bearing

dimensions, clearance, oil viscosity etc

Surface contact (and so boundary or mixed

lubrication) is possible during running-in

process or if bearing is overloaded

– Friction prediction is theoretically possible

using elasto-hydrodynamic analysis with

asperity contact model but

• Measured worn axial bearing profilesand distorted shapes are needed for

sensible results

• so not very useful for designBearing axial position (mm)rear front

Main 2 B e a r i n g w e a r d e p t h ( µ )

Bearing axial position (mm)rear front

Main 2 B e a r i n g w e a r d e p t h ( µ )

1

2

3

4

5

6

1

2

3

4

5

6

0.00

0.05

0.10

0.15

0.20

0.25

0.30

4000 6000 8000 10000 12000 14000 16000 18000

Engine Speed (rpm)

T o t a l M a i n B e a r i n g F M E P ( b a r )

Oil A

Oil B

Axial profileNo axial profile


23/47


Camshaft bearing friction – effect of ENGDYN model level

Front camshaft bearings often exhibit edge wear due to

loads from the timing drive

– This wear usually occurs during run-in period and

then stabilises

ENGDYN used to predict friction at camshaft bearings

– Simple rigid model gives power loss due to oil shear

effects

– With flexible camshaft model edge contact leads tohigh friction loss

– When worn profile is introduced friction loss is

similar to results from simple model0

10

20

30

40

50

60

70

80

90

100

600 800 1000 1200 1400 1600 1800 2000

camshaft speed (rpm)

p o w e r l o s s ( W

statically determinate loading

dynamic loading

dynamic loading - with axial profile


24/47



Introduction






Case study

Conclusions


25/47


Crankcase pumping loss - WAVE

Crank case pumping incurs a power loss

– Power required to pump crankcase gas from onebay to the next as the pistons move up and down

– This can be significant loss on high speed

engines

• so “dry sump” designs with scavenge pumps

are often used to minimise this

Ricardo and others have used 1D gas dynamicsmodelling software such as WAVE to predict crank

case pumping loss with some success

– Complex geometry of crank case from CAD and

then automatically reduced to 1D network


26/47



BREATHER AREA

C P M E P > Critical Breather Area

Losses reduce with

increase in area

Reduced resistance to

inter-bay flow exchange

< Critical Breather Area

Losses increase with increase in area

Increasing inter-bay flow exchange

Reduced recovery of compression work

Max CPMEP at

a Critical

Breather Area

-0.18

-0.16

-0.14

-0.12

-0.1

-0.08

-0.06

-0.04

-0.02

0

0 300 600 900 1200 1500

EFFECTIVE BREATHER AREA (mm2)

C P M E P ( b a r )

7000

8500

11000

13500

16000


27/47



Inter-bay breathingholes removed

-0.50

-0.45

-0.40

-0.35

-0.30

-0.25

-0.20

-0.15

-0.10

-0.05

0.00

6000 8000 10000 12000 14000 16000 18000 20000 22000

ENGINE SPEED (rpm)

C P M E

P ( b a r )

-5.0

-4.5

-4.0

-3.5

-3.0

-2.5

-2.0

-1.5

-1.0

-0.5

0.0

P O W E R ( k W )0.38

bar 4.7

kW

Single external breather

exit through balancer

shaft center

4 scavenge points

Wet

Sump

Dry

Sump


28/47



Introduction






Case study

Conclusions


29/47


Valve train friction prediction

At system level valve train friction is reasonably well-understood

– The mean friction level of valve trains can be measured by driving the camshaft usingan electric motor and measuring mean drive torque

However

– These system level measurements mask the fact that the detailed sources of loss are

not very well understood

– Especially since the measurements often include the timing drive as well as the valve

train

Motorsport engines nearly always have

sliding contacts between cam and

tappet to enable use of a lightweight

follower and optimise high speed

dynamics

– Friction loss is dominated by losses

at the these highly-loaded sliding

contacts 0

0.1

0.2

0.3

0.4

0.5

0.6

9000 10000 11000 12000 13000 14000 15000 16000

Engine speed (rpm)

F M E P ( b a r )

total gear losses

camshaft bearings

tappet/bore contacts

cam/tappet contacts


30/47


Cam/follower friction – sliding contact

VALDYN can calculate friction power loss at sliding

contacts between cam/tappet or cam/follower – Friction due to oil shear and asperity contact are

calculated separately

– Friction is therefore dependent on oil viscosity as

well as surface texture

– The model also considers

• friction between tappet and cylinder • tappet rotation

0.000

0.010

0.020

0.030

0.040

0.050

0.060

4000 6000 8000 10000 12000 14000 16000 18000

Engine Speed (rpm)

M e a n E f f e c t i v e F

r i c t i o n C o e f f i c i e n t

Oil A

Oil B


31/47


Timing drive friction - VALDYN

VALDYN has the capability to calculate

the losses in chain timing drives – Models have been validated validated

against measured data at passenger

car engine speeds

– Results can be very dependent on

tensioner design

0

200

400

600

800

1000

1200

1400

1600

1800

1000 1500 2000 2500 3000 3500 4000 4500 5000 5500

engine speed (rpm)

p o

w e r l o s s ( W )

link/guide

roller/sprocket

pin/bush

cam bearings

idler bearings

measured


32/47


Timing drive friction - VALDYN

VALDYN has gear models for

calculation of contact forces andsliding speeds

– Friction coefficients can be entered

to calculate losses at gear meshes

– But actual losses in high speed

gears are mostly caused by

windage and oil churning effects

which cannot be modelled withconfidence

VALDYN can model belt drive

dynamics but detailed sources of loss

are not understood so cannot be

predicted

– Belt material hysteresis?

– Rubbing losses between teeth and

pulleys?

– Pumping air out during meshing?

Tensioner

Crankshaft

Camshafts

Tensioner

Crankshaft

Camshafts

Tensioner

Crankshaft

Camshafts

Tensioner

Crankshaft

Camshafts


33/47



Introduction






Case study

Conclusions


34/47


Case study

The following slides illustrate the use of CAE tools and the results of motored teardowntest to understand the distribution of friction in a racing engine

The engine was motored and a 4-stage teardown test was made

– 1 Whole engine with intake and exhaust systems

– 2 Intake and exhaust systems removed

– 3 Cylinder head removed

• including upper timing gears

• Cylinder head removed by plate to maintain bolt loads on structure

• Plate was non restrictive so no pumping losses

– 4 Pistons and rods removed

• Replaced by dummy weights

All stages include driven oil pump, water pump and transmission


35/47


Case study - Valve train group friction

Measured data obtained by

subtraction

– Test 2 – Test 3

Pumping loss under motored

conditions calculated using

WAVE

Camshaft bearing friction

calculated using ENGDYN

Cam/tappet losses and

tappet/bore losses calculated

using VALDYN

Gear contact losses estimated

using loads from VALDYN and

friction coefficients – Windage losses estimated

0

0.2

0.4

0.6

0.8

1

1.2

1.4

1.6

1.8

2

9000 10000 11000 12000 13000 14000 15000 16000

Engine speed (rpm)

F M E P ( b a r )

pumping losstotal gear losses

camshaft bearings


cam/tappet contacts

measured data

0

0.2

0.4

0.60.8

1

1.2

1.4

1.6

1.8

2

9000 10000 11000 12000 13000 14000 15000 16000

Engine speed (rpm)

F M

E P ( b a r )

total gear losses

camshaft bearings


cam/tappet contacts

measured data minuspredicted PMEP


36/47


Case study - Reciprocating component group friction

Measured data obtained by

subtraction – Test 3 – Test 4

Piston skirt friction loss and

small end bearing loss

calculated using PISDYN

Piston rings friction loss

calculated using RINGPAK

Crankcase pumping loss

calculated using WAVE

Big end bearing losscalculated using ENGDYN 0

0.2

0.4

0.6

0.8

1

1.2

1.4

1.6

1.8

2

9000 10000 11000 12000 13000 14000 15000 16000

Engine speed (rpm)

F M

E P ( b a r )

piston skirt

oil control rings

top piston ring

crankcase pumping

small end bearings

big end bearings

measured data


37/47


0

0.2

0.4

0.6

0.8

1

1.2

1.4

1.6

1.8

2

9000 10000 11000 12000 13000 14000 15000 16000

Engine speed (rpm)

F M E P ( b a r )

water pump

gear windage

oil pump

seals

balancer windage

rolling element bearings

gear meshes

crankshaft windage

main bearings

measured data

Crankshaft/transmission group friction

Measured data

obtained directly Test 4

Main bearing loss

calculated using

ENGDYN

Windage losses

estimated using simpleapproach

Gear contact losses

estimated using simple

approach

Oil pump losses and

water pump lossesestimated

Rolling element bearing losses calculated usingcalculated loads and friction coefficients

Seals losses estimated using suppliers database


38/47


Whole engine motored friction loss (1)

With motored pumping

losses included

Additional loads at main

bearings due to inertia of

pistons and rods

accounted

Allowance in calculations

for additional loads atpistons and bearings due

to motored gas force

with cylinder head on

0

0.5

1

1.5

2

2.5

3

3.5

4

4.5

5

9000 10000 11000 12000 13000 14000 15000 16000

Engine speed (rpm)

F M E P

( b a r )

pumping loss

timing gears

camshaft bearings


cam/tappet contacts

piston skirt

oil control rings

top piston rings

crankcase pumping

small end bearings

big end bearingswater pump

gear windage

oil pump

seals

balancer windage

rolling element bearings

gear meshes

crankshaft windage

main bearings

measured data


39/47


Whole engine motored friction loss (2)

Same data but with results grouped

by subsystem

Pumping losses excluded

At 15000 rpm

– Crankshaft 25.1%

– Transmission 7.9%

– Pumps 11.4%

– Crankcase pumping 8.5%

– Pistons 33.9%

– Valve trains 13.2%

0

0.5

1

1.5

2

2.5

3

3.5

4

4.5

5

9000 10000 11000 12000 13000 14000 15000 16000

Engine speed (rpm)

F M E P ( b a r )

valve trains and drive

piston assemblies

crankcase pumping

oil and water pumps

transmission (inc balancer)

crankshaft

measured minus pumping loss


40/47


Full load fired friction loss by analysis

Next step was to use CAE tools to model the

following under fired engine conditions – Piston rings

– Piston skirt

– Small end bearing

– Big end bearings

– Main bearings

The following were accounted for

– Increased cylinder pressure

– Increased component temperature and

effect on clearances

– Increased oil temperature and effect on

oil viscosity


41/47


Full load fired friction loss at piston group (1)

Fired conditions lead to

– slightly higher loss at small end

bearings

– no significant change at top piston

rings

– reduced loss at oil control rings

– significantly increased loss at piston

skirt (particularly at lower engine

speeds)

small end bearings

00.020.040.060.080.1

0.120.140.160.180.2

9000 10000 11000 12000 13000 14000 15000 16000

Engine speed (rpm)

F M E P (

b a r )

motored

full load fired

top piston rings

00.020.040.060.080.10.12

0.140.160.180.2

9000 10000 11000 12000 13000 14000 15000 16000

Engine speed (rpm)

F M E P ( b

a r )

motored

full load fired

oil control rings

00.020.040.060.080.1

0.120.140.160.180.2

9000 10000 11000 12000 13000 14000 15000 16000

Engine speed (rpm)

F M E P ( b a

r )

motored

full load fired

piston skirts

0.00

0.25

0.50

0.75

1.00

1.25

1.50

1.75

2.00

2.25

2.50

9000 10000 11000 12000 13000 14000 15000 16000

Engine speed (rpm)

F M E P ( b a r )

motored

full load fired


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Full load fired friction loss at piston group (2)

Predicted losses at piston

skirt are highly sensitive toskirt/liner clearance which is

sensitive to assumed

component temperatures

– Results shown with +/-5

micron clearance

– No change in componentor oil temperatures

Model could be refined with

better knowledge of

component temperatures

0.00

0.25

0.50

0.75

1.00

1.25

1.50

1.75

2.002.25

2.50

9000 10000 11000 12000 13000 14000 15000 16000

Engine speed (rpm)

F M E P ( b a r )

motored

full load fired (nominal)

full load (5um reduced clearance)

full load (5um increased clearance)


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Full load fired friction loss at crank train bearings

Fired conditions lead to

– slightly higher loss atlower speeds

– but similar losses at

high speed as inertia

forces dominate

0

0.05

0.1

0.15

0.2

0.25

0.3

9000 10000 11000 12000 13000 14000 15000 16000Engine speed (rpm)

F M E P ( b a r )

motored

full load fired

0

0.05

0.1

0.15

0.2

0.25

0.3

9000 10000 11000 12000 13000 14000 15000 16000

Engine speed (rpm)

F M E P ( b a r )

motored

full load fired

Mainbearings

Big endbearings


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Whole engine full load fired friction loss

Results grouped by subsystem

Pumping losses excluded

Measured line is whole engine

motored data for comparison

At 15000 rpm

– Crankshaft 22.2%

– Transmission 6.9% – Pumps 10.0%

– Crankcase pumping 7.5%

– Pistons 41.9%

– Valve trains 11.5%

0

0.5

1

1.5

2

2.5

3

3.5

4

4.5

5

9000 10000 11000 12000 13000 14000 15000 16000

Engine speed (rpm)

F M E P

( b a r )

valve trains and drive

piston assemblies

crankcase pumpingoil and water pumps

transmission (inc balancer)

crankshaft

measured minus pumping loss


45/47



Introduction






Case study

Conclusions


46/47


Conclusions

Overall friction level can be estimated easily based on historical data

If more complex calculations are performed there are many problems

– All major contacts involve asperity contact losses and oil shear losses

– Asperity contact losses dominate if dimensions from drawings are used

• but as components wear during run-in period these losses reduce

– This makes true prediction difficult

• but if worn component geometry and surface roughness data are used as modelinputs then better results are possible

Some interesting insights can be obtained by

– use of relatively inexpensive motored strip test

– combined with use of advanced CAE tools

But to make further progress we need

– measured data from fired engines at high speed to validate software

– improved models of mixed lubrication including 3D surface texture effects


47/47

Thank you for your attention

Any Questions?

calculation of friction in high performance engines

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