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
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3 © Ricardo plc 2010RD.10/174701.120th April 2010
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…
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4 © 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
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5 © Ricardo plc 2010RD.10/174701.120th April 2010
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
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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
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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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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
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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
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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
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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
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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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15 © 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
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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
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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
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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
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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
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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
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21 © Ricardo plc 2010RD.10/174701.120th April 2010
Presentation contents
Introduction
Some general comments and issues
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
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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
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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
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Presentation contents
Introduction
Some general comments and issues
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
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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
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Crankcase pumping loss - WAVE
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
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Crankcase pumping loss - WAVE
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
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Presentation contents
Introduction
Some general comments and issues
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
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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
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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
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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
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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
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33 © Ricardo plc 2010RD.10/174701.120th April 2010
Presentation contents
Introduction
Some general comments and issues
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
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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
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35 © Ricardo plc 2010RD.10/174701.120th April 2010
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
tappet/bore contacts
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
tappet/bore contacts
cam/tappet contacts
measured data minuspredicted PMEP
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36 © Ricardo plc 2010RD.10/174701.120th April 2010
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
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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
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38 © Ricardo plc 2010RD.10/174701.120th April 2010
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
tappet/bore contacts
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
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39 © Ricardo plc 2010RD.10/174701.120th April 2010
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
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40 © Ricardo plc 2010RD.10/174701.120th April 2010
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
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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
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45 © Ricardo plc 2010RD.10/174701.120th April 2010
Presentation contents
Introduction
Some general comments and issues
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
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46 © Ricardo plc 2010RD.10/174701.120th April 2010
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
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Thank you for your attention
Any Questions?
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