copyright by myoungjin kim 2005
TRANSCRIPT
Copyright
by
Myoungjin Kim
2005
The Dissertation Committee for Myoungjin Kimcertifies that this is the approved version of the following dissertation:
Friction Force Measurement and Analysis of the Rotating Liner Engine
Committee:
Ronald D. Matthews, Supervisor
Thomas M. Kiehne
Matthew J. Hall
Ofodike A. Ezekoye
Charles E. Roberts, Jr.
Friction Force Measurement and Analysis of the Rotating Liner Engine
by
Myoungjin Kim, B.S., M.S.
Dissertation
Presented to the Faculty of the Graduate School of
The University of Texas at Austin
in Partial Fulfillment
of the Requirements
for the Degree of
Doctor of Philosophy
The University of Texas at Austin
August, 2005
Dedication
To my wife, Jaesun Lee, and my two sons, Teayoung Kim and Joonyoung
Kim, who love and encourage me.
To my parents, Jaehwan Kim and Youngsun Joo, and my parents-in-law,
Kyuhyung Lee and Mooin Jung, who support and pray for me.
Their love and belief in me made this dissertation possible.
v
Acknowledgements
I would like to express my thanks to the following persons whose help and advice
were essential in guiding me to this dissertation. Firstly I appreciate my supervisor, Dr.
Matthews. I was very impressed and tried to learn from your passion and attitude toward
research and teaching. My committee members (Dr. Kiehne, Dr. Hall, Dr. Ezekoye, and
Dr. Roberts) advised and guided me in the right direction in my dissertation. Dr.
Dardalis, who is the inventor of the Rotating Liner Engine, gave me an opportunity to
participate in this project and helped me in doing the experiments. Dr.Liechti, in the UT
Aerospace Engineering Department, taught me how to install, measure, and calibrate the
strain gages. The machine shop personell, including Curtis Johnson, Danny Jares, John
Pedracine, Don Artieschoufsky, and Tho Huynh, helped me whenever I needed their aid.
Undergraduate students, Ian, Ryan, and Sanggyu Lee, did a great job in modeling and
installing the hardware. Byungsoon Min, Dr.Myungjun Lee, and Dr. Joo were good
friends and helped me during my Ph.D studies at the University of Texas at Austin.
Many Korean colleagues, including Seyoon Kim, Deajong Kim, Jihoon Choi, Seokyoung
Ahn, Seunghan Lee, Youngkeun Park, Younghoon Han, Jaebum Hur, Dohyung Kim, and
others, shared many good memories with me. My parents and parents-in-law always
prayed and endured for my four years of dissertation research. Finally, the love from my
wife, Jaesun Lee, and two sons, Taeyoung Kim and Junyoung Kim, made me do my best
and finish this dissertation successfully.
vi
Friction Force Measurement and Analysis of the Rotating Liner Engine
Publication No.________________
Myoungjin Kim, Ph.D.The University of Texas at Austin, 2005
Supervisor: Ronald D. Matthews
As emissions regulations become more stringent and fuel prices increase at a
significant rate, the fuel efficiency of piston engines becomes more important than ever.
Since most of the engine’s friction losses are from the piston/ring assembly, it is
indispensable to reduce the piston/ring assembly friction for better fuel economy. The
Rotating Liner Engine (RLE) was developed to remove the boundary lubrication of the
piston/ring assembly friction through cylinder liner rotation.
Even though the RLE was initially developed mainly by Dr. Dardalis several
years ago, the friction reduction effect of the RLE was not confirmed except via
preliminary motoring tests using a crude dynamometer. The main purpose of this
dissertation is to confirm the RLE effect on piston assembly friction reduction using
sophisticated measurement methods. Three different friction measurement methods were
applied in measuring the friction force difference between a baseline engine and a
prototype RLE. Through the use of three different friction measurement methods, the
friction reduction of the RLE has been confirmed via this dissertation research and each
of the friction measurement methods is also compared based on its measurement results.
The analysis of the friction mechanism of the baseline engine was performed using the
instantaneous IMEP method and a commercial simulation program called RINGPAK.
Through the use of experimental methods and the simulation, the friction mechanism of
vii
the piston/ring assembly is analyzed. The limitation of the experimental and the
calculation methods is also discussed.
viii
Table of Contents
List of Tables------------------------------------------------------------------------------------xii
List of Figures----------------------------------------------------------------------------------xiii
Chapter 1 Introduction-----------------------------------------------------------------------1
1.1 Overview of piston assembly friction------------------------------------------------1
1.2 Motivation------------------------------------------------------------------------------6
1.3 Dissertation overview and scope of work-------------------------------------------7
Chapter 2 Test Methods----------------------------------------------------------------------9
2.1 Overview of friction measurement methods----------------------------------------9
2.1.1 Direct motoring test with tear-down--------------------------------------9
2.1.2 Morse test-------------------------------------------------------------------10
2.1.3 Willans line method-------------------------------------------------------10
2.1.4 Measurement of FMEP from IMEP and BMEP-----------------------11
2.1.5 Floating liner method------------------------------------------------------14
2.1.6 Instantaneous IMEP method----------------------------------------------15
2.1.7 P-w method-----------------------------------------------------------------16
2.2 Friction measurement methods in this research-----------------------------------19
2.2.1 Instantaneous IMEP method----------------------------------------------20
1) Strain gage-------------------------------------------------------------20
2) Piston dynamics-------------------------------------------------------22
2.2.2 P-w method-----------------------------------------------------------------26
1) Crankshaft dynamics-------------------------------------------------27
2) Transfer matrix method----------------------------------------------33
Chapter 3 Test Setup-------------------------------------------------------------------------38
3.1 Test engine-----------------------------------------------------------------------------38
3.1.1 Baseline engine-------------------------------------------------------------38
ix
3.1.2 Rotating Liner Engine-----------------------------------------------------39
3.2 Torque sensor and coupling----------------------------------------------------------45
3.3 Strain gage measurements------------------------------------------------------------46
3.3.1 Strain gage specifications-------------------------------------------------47
3.3.2 Strain gage installation and measurement------------------------------50
3.3.3 Strain gage calibration-----------------------------------------------------52
3.3.4 Bending and temperature compensation--------------------------------54
3.4 Cylinder pressure measurement and data acquisition----------------------------58
Chapter 4 Test Results----------------------------------------------------------------------61
4.1 Hot motoring tests--------------------------------------------------------------------61
4.1.1 Cycle-averaged friction torque and tear-down tests------------------62
1) Baseline engine------------------------------------------------------62
2) Rotating Liner Engine-----------------------------------------------66
4.1.2 Friction force measurement using the instantaneous IMEP method
---------------------------------------------------------------------------------------70
1) Baseline engine-------------------------------------------------------70
1-1) Motoring friction during cold motoring tests (oil
temperature: 20°C)---------------------------------------------70
1-2) Motoring friction during hot motoring tests (oil
temperature: 90°C)---------------------------------------------84
2) Rotating Liner Engine-----------------------------------------------89
4.1.3 Friction torque measurement using the P-w method------------------92
4.1.4 Dynamic characteristics of the crankshaft system (using transfer
matrix method)------------------------------------------------------------------100
4.2 Firing tests----------------------------------------------------------------------------111
4.2.1 Baseline engine-----------------------------------------------------------111
4.2.2 Cyclic variations----------------------------------------------------------118
4.3 Piston assembly friction force analysis using the IMEP and P-w methods--128
4.3.1 Piston assembly friction modeling-------------------------------------129
4.3.2 Experimental piston ring assembly friction torque values----------132
x
Chapter 5 Error Analysis in Friction Force Measurement-----------------125
5.1 Introduction---------------------------------------------------------------------------125
5.2 Measured friction errors and analysis in the p-w method----------------------125
5.3 Measured friction erros and analysis in the instantaneous IMEP method---136
5.3.1 Sensitivity analysis-------------------------------------------------------136
5.3.2 Measurement errors in the strain gage---------------------------------138
5.3.3 Possible error sources in the instantaneous IMEP method----------139
Chapter 6 Friction Force Calculation Using RINGPAK--------------------141
6.1 Introduction--------------------------------------------------------------------------141
6.2 Details of the RINGPAK models-------------------------------------------------142
6.2.1 Ring dynamics------------------------------------------------------------143
1) Axial ring motions--------------------------------------------------143
2) Ring twist------------------------------------------------------------145
6.2.2 Inter-ring dynamics-------------------------------------------------------146
1) Governing equations and flow models---------------------------146
2) Blow-by and blow-back of gas flow------------------------------149
6.2.3 Ring-liner lubrication and radial ring dynamics----------------------150
1) Radial ring motion--------------------------------------------------150
2) Ring-liner hydrodynamic lubrication-----------------------------151
3) Ring-liner boundary lubrication-----------------------------------153
4) Ring-liner friction and power losses------------------------------153
6.2.4 Liner oil transport--------------------------------------------------------154
6.2.5 Oil consumption mechanisms-------------------------------------------155
1) Oil evaporation------------------------------------------------------155
2) Oil throw-off from inertia------------------------------------------160
3) Oil entrainment in blow-back gases------------------------------160
6.3 Input data for RINGPAK simulations--------------------------------------------162
6.3.1 Input parameters----------------------------------------------------------162
6.3.2 Engine operating condition----------------------------------------------165
6.4 Simulation results--------------------------------------------------------------------165
6.4.1 Motoring friction results (hot motoring)------------------------------165
xi
6.4.2 Firing friction results----------------------------------------------------169
6.4.3 Parameter study----------------------------------------------------------173
1) Effect of ring tension----------------------------------------------173
2) Effect of surface roughness---------------------------------------177
Chapter 7 Summary and Conclusions-----------------------------------------------185
Chapter 8 Recommendations for Future Work---------------------------------188
References---------------------------------------------------------------------------------------190
VITA----------------------------------------------------------------------------------------------196
xii
List of Tables
Table 2.1 Modeling of each friction component------------------------------------------------32
Table 3.1 Baseline engine specifications--------------------------------------------------------53
Table 3.2 Torque sensor specifications----------------------------------------------------------59
Table 3.3 Thermal expansion coefficients of engineering materials-------------------------63
Table 3.4 Strain gage specifications--------------------------------------------------------------64
Table 4.1 Component values in the equivalent dynamic model of the crankshaft--------121
Table 4.2 IMEP and COV of the IMEP at firing test-----------------------------------------137
Table 4.3 Mean friction work and friction work COV---------------------------------------139
Table 4.4 Engine basic parameters used by friction model----------------------------------144
Table 5.1 Piston and piston ring terminology-------------------------------------------------171
Table 5.2 Base RINGPAK input data for baseline engine-----------------------------------172
xiii
List of Figures
Figure 1.1 California emissions regulations---------------------------------------------------1
Figure 1.2 Typical fuel energy distribution in an internal combustion engine-----------2
Figure 1.3 Representative mechanical loss distribution ------------------------------------3
Figure 1.4 Stribeck diagram showing the various regimes of lubrication----------------5
Figure 2.1 Example of measured cylinder pressure ----------------------------------------11
Figure 2.2 Schematic view of the floating liner method-----------------------------------14
Figure 2.3 Free body diagram of a piston ---------------------------------------------------15
Figure 2.4 Wheatstone bridge circuit---------------------------------------------------------20
Figure 2.5 Acceleration of the piston and connecting rod---------------------------------23
Figure 2.6 Lumped mass model of a connecting rod---------------------------------------29
Figure 2.7 Free body diagram for the wrist pin forces-------------------------------------30
Figure 2.8 Mass-less shaft with disks--------------------------------------------------------34
Figure 2.9 Free body diagrams of a shaft and disk-----------------------------------------34
Figure 2.10 Mass-less shaft with six disks--------------------------------------------------36
Figure 3.1 Prototype Rotating Liner Engine assemblies----------------------------------40
Figure 3.2 Rotating Liner Engine components---------------------------------------------41
Figure 3.3 Final seal design and installation------------------------------------------------42
Figure 3.4 Cross-section showing the face seal---------------------------------------------43
Figure 3.5 Driving mechanisms for rotating the liner--------------------------------------44
Figure 3.6 Photograph of driving mechanism-----------------------------------------------44
Figure 3.7 Torque sensor calibration curve--------------------------------------------------46
Figure 3.8 Typical thermal output variations with temperature for self-temperature-
compensated constantan (A-alloy) and modified Karma (K-alloy) strain
gages--------------------------------------------------------------------------------47
Figure 3.9 Schematic of strain gage installation--------------------------------------------51
Figure 3.10 2100 series signal conditioner and amplifier----------------------------------51
Figure 3.11 Tension test system for the connecting rod tests----------------------------- 53
Figure 3.12 Connecting rod installed in the servo-hydraulic test machine--------------53
Figure 3.13 Strain gage calibration test results----------------------------------------------54
xiv
Figure 3.14 Wheatstone bridge circuit used in tests----------------------------------------55
Figure 3.15 Gage factor variation with temperature for constantan (A-alloy) and
isoelastic (D-alloy) strain gages------------------------------------------------56
Figure 3.16 Measurement system configurations-------------------------------------------59
Figure 4.1 Baseline engine hot motoring torque--------------------------------------------63
Figure 4.2 Teardown test results for the baseline engine----------------------------------65
Figure 4.3 Rotating Liner engine hot motoring torque-------------------------------------66
Figure 4.4 Total hot motoring friction reduction through liner rotation-----------------67
Figure 4.5 Teardown test results for the Rotating Liner engine--------------------------68
Figure 4.6 Piston assembly friction torque of the baseline engine and the RLE--------69
Figure 4.7a Measured pressure force and connecting rod force for an oil temperature
of 20 °C at 500 rpm and 800 rpm-----------------------------------------------71
Figure 4.7b Measured pressure force and connecting rod force for an oil temperature
of 20 °C at 1200 rpm, 1600 rpm, and 2000 rpm------------------------------72
Figure 4.8 Effects of engine speed on pressure and connecting rod force variations
throughout the cycle at an oil temperature of 20 °C-------------------------74
Figure 4.9a Measured angular speed and the calculated angular acceleration, linear
speed, and linear acceleration at 500 rpm and 800 rpm---------------------74
Figure 4.9b Measured angular speed and the calculated angular acceleration, linear
speed, and linear acceleration at 1200 rpm and 1600 rpm------------------75
Figure 4.9c Measured angular speed and the calculated angular acceleration, linear
speed, and linear acceleration at 2000 rpm------------------------------------76
Figure 4.10 Strain gage location, nomenclature used in the equation set, and
accelerations-----------------------------------------------------------------------77
Figure 4.11 Modeled connecting rod using SOLIDWORKS-----------------------------78
Figure 4.12a Inertial forces of the piston and the connecting rod at 500 rpm-----------78
Figure 4.12b Inertial forces of the piston and the connecting rod at 800 rpm and 1200
rpm----------------------------------------------------------------------------------79
Figure 4.12c Inertial forces of the piston and the connecting rod at 1600 rpm and 2000
rpm----------------------------------------------------------------------------------80
xv
Figure 4.13 Effects of engine speed on the variation of the inertia force throughout the
cycle for motoring conditions---------------------------------------------------80
Figure 4.14 Effect of oil temperature on oil viscosity--------------------------------------82
Figure 4.15 Friction force of the piston assembly at an oil temperature of 20 °C------83
Figure 4.16 Effects of engine speed on the variation of the friction force throughout
the cycle at an oil temperature of 20 °C---------------------------------------84
Figure 4.17 Measured pressure force and connecting rod force (90 °C oil
temperature)-----------------------------------------------------------------------85
Figure 4.18 Effects of engine speed on the variations of the pressure and connecting
rod forces throughout the cycle at an oil temperature of 90 °C -----------------86
Figure 4.19 Friction force of the piston assembly at an oil temperature 90°C---------87
Figure 4.20 Effect of engine speed on the friction force throughout the cycle for an oil
temperature of 90 °C-------------------------------------------------------------88
Figure 4.21 Friction force comparison between the baseline engine and the RLE at
1200 rpm--------------------------------------------------------------------------89
Figure 4.22 Sensitivity analysis for the friction force obtained using the instantaneous
IMEP method---------------------------------------------------------------------91
Figure 4.23 Measured instantaneous motoring torque of the baseline engine and the
RLE as obtained using the p-w method---------------------------------------92
Figure 4.24 Pressure torque at 1200 rpm----------------------------------------------------94
Figure 4.25 Crankshaft assembly 3-dimensional modeling-------------------------------95
Figure 4.26 Rotational and translational speed and acceleration of the baseline engine
and the Rotating Liner Engine at 1200 rpm----------------------------------96
Figure 4.27 Inertia torques developed by translational and rotational motion at 1200
rpm---------------------------------------------------------------------------------97
Figure 4.28 Measured output torque, pressure torque, inertia torque, and friction
torque of the baseline engine and the RLE at 1200 rpm--------------------98
Figure 4.29 Equivalent dynamic model of the crankshaft system-----------------------100
Figure 4.30 Derivation of the field matrix--------------------------------------------------102
Figure 4.31 Derivation of the point matrix-------------------------------------------------104
xvi
Figure 4.32 Measured and calculated motoring torque using harmonic components-----
-------------------------------------------------------------------------------------106
Figure 4.33 Mesh generated for ANSYS analysis-----------------------------------------107
Figure 4.34 Comparison between the measured and the calculated instantaneous speed
at 1200 rpm-----------------------------------------------------------------------109
Figure 4.35a Measured cylinder pressure and connecting rod forces at WOT firing
condition (800 and 1200 rpm)-------------------------------------------------111
Figure 4.35b Measured cylinder pressure and connecting rod forces at WOT firing
condition (1600 and 2000 rpm)-----------------------------------------------112
Figure 4.36 Effect of engine speed on the pressure and connecting rod forces
throughout the cycle for WOT firing conditions---------------------------112
Figure 4.37a Measured inertial forces of the piston assembly and the connecting rod
under WOT firing conditions (800 and 1200 rpm)------------------------114
Figure 4.37b Measured inertial forces of the piston assembly and the connecting rod
under WOT firing conditions (1600 and 2000 rpm)----------------------115
Figure 4.38 Effects of engine speed on the inertia force throughout the cycle under
WOT firing conditions----------------------------------------------------------116
Figure 4.39 Friction force of the piston assembly under WOT firing conditions-----117
Figure 4.40 Cyclic cylinder pressure variations for the baseline engine under WOT
firing conditions-----------------------------------------------------------------119
Figure 4.41 Pressure force and friction force variations during the cycle at 800 rpm----
-------------------------------------------------------------------------------------119
Figure 4.42 Crank angles at peak pressure and friction forces at 800 rpm-------------120
Figure 4.43 Pressure force and friction force variations during the cycle at 1200 rpm---
-------------------------------------------------------------------------------------121
Figure 4.44 Pressure force and friction force variations during the cycle at 1600 rpm---
-------------------------------------------------------------------------------------122
Figure 4.45 Crank angles at peak pressure and friction forces at 1200 rpm-----------122
Figure 4.46 Crank angles at peak pressure and friction forces at 1600 rpm-----------123
Figure 4.47 Position of the piston top relative to the head and distance swept by the
piston at 800 rpm----------------------------------------------------------------125
xvii
Figure 4.48 Piston assembly instantaneous friction torque at 1200 rpm---------------132
Figure 4.49 Friction torque obtained for the baseline engine using the p-w method for
cold motoring--------------------------------------------------------------------133
Figure 4.50 Friction torque obtained for the baseline engine using the p-w method for
hot motoring---------------------------------------------------------------------134
Figure 4.51 Friction torque obtained for the baseline engine using the p-w method for
WOT firing conditions----------------------------------------------------------135
Figure 5.1 Primary phenomena associated with a piston ring pack---------------------136
Figure 5.2 Schematic of ring motion and associated force and moment components
---------------------------------------------------------------------------------------138
Figure 5.3 Schematic of the various flow passages around a ring----------------------141
Figure 5.4 Schematic of blowby and blowback gas flows-------------------------------144
Figure 5.5 Schematic of radial ring motion with the associated force components--145
Figure 5.6 Cross-section of the gas-oil film-liner-coolant system at an arbitrary axial
location---------------------------------------------------------------------------151
Figure 5.7 Piston configuration--------------------------------------------------------------157
Figure 5.8 Ring configuration----------------------------------------------------------------158
Figure 5.9 Predicted piston ring friction at 500 rpm for hot motoring conditions----161
Figure 5.10 Predicted piston ring friction at 800 rpm for hot motoring conditions---161
Figure 5.11 Predicted piston ring friction at 1200 rpm for hot motoring conditions-162
Figure 5.12 Predicted piston ring friction at 1600 rpm for hot motoring conditions-162
Figure 5.13 Predicted piston ring friction at 2000 rpm for hot motoring conditions-163
Figure 5.14 Effects of engine speed on the total piston assembly friction for hot
motoring conditions-------------------------------------------------------------163
Figure 5.15 Predicted piston ring friction at 800 rpm for WOT firing conditions----164
Figure 5.16 Predicted piston ring friction at 1200 rpm for WOT firing conditions---165
Figure 5.17 Predicted piston ring friction at 1600 rpm for WOT firing conditions---165
Figure 5.18 Predicted piston ring friction at 2000 rpm for WOT firing conditions---166
Figure 5.19 Effects of engine speed on the predicted total piston assembly friction for
WOT firing conditions----------------------------------------------------------166
Figure 5.20 Predicted crank angles at peak pressure and friction forces---------------167
xviii
Figure 5.21 Predicted effects of high ring tension on piston ring friction under hot
motoring conditions at 500 rpm-----------------------------------------------168
Figure 5.22 Predicted effects of high ring tension on piston ring friction under hot
motoring conditions at 800rpm------------------------------------------------169
Figure 5.23 Predicted effects of high ring tension on piston ring friction under hot
motoring condition at 1200 rpm-----------------------------------------------169
Figure 5.24 Predicted effects of high ring tension on piston ring friction under hot
motoring conditions at 1600 rpm----------------------------------------------170
Figure 5.25 Predicted effects of high ring tension on piston ring friction under hot
motoring conditions at 2000 rpm----------------------------------------------170
Figure 5.26 Comparison of the predictions of piston ring friction between the baseline
and the high ring tension over a range of engine speeds-------------------170
Figure 5.27 Predicted effects of decreasing the asperity radius of curvature by a factor
of 5 (Case 1) on piston ring friction under hot motoring conditions at 500
rpm--------------------------------------------------------------------------------173
Figure 5.28 Predicted effects of decreasing the asperity radius of curvature by a factor
of 5 (Case 1) on piston ring friction under hot motoring conditions at 800
rpm--------------------------------------------------------------------------------174
Figure 5.29 Predictions of the effects of decreasing the asperity radius of curvature by
a factor of 5 (Case 1) on piston ring friction under hot motoring conditions
at 1200 rpm----------------------------------------------------------------------174
Figure 5.30 Predicted effects of decreasing the asperity radius of curvature by a factor
of 5 (Case 1) on piston ring friction under hot motoring conditions at 1600
rpm--------------------------------------------------------------------------------175
Figure 5.31 Predicted effects of decreasing the asperity radius of curvature by a factor
of 5 (Case 1) on piston ring friction under hot motoring conditions at 2000
rpm--------------------------------------------------------------------------------175
Figure 5.32 Friction force comparison between Case 1 (asperity radius decreased by a
factor of 5) and the baseline asperity radius of curvature------------------176
Figure 5.33 Friction force comparison between Case 2 (asperity radius of curvature
decreased by a factor of 10) and the baseline radius of curvature--------176
1
Chapter 1. Introduction
1.1 Overview of piston assembly friction
As the fuel economy and emissions regulations become more stringent,
automakers have to strive to meet the imposed emissions standards. For example, In
California’s emissions regulations, the Tier1/LEV standards were applied for the 2003
model year, and more stringent LEVII standards were effective from 2004. The demand
for fuel economy has also become a more urgent problem. Figure 1.1 shows an example
of emissions regulations enforced in California since 1992.
0
0.1
0.2
0.3
0.4
0.5
NMHC NMOG NOx
g/m
i
Tier1 TLEV LEV ULEV
Figure 1.1. California emissions regulations. .
The emissions standards in Figure 1.1 are simplified examples of gasoline vehicle
emissions regulations. The current emissions regulations enforced in the United States
and Europe are more complicated and stringent. Automakers have invested a tremendous
amount of money for improving fuel economy and developing emissions reduction
techniques in order to meet these strict regulations about fuel economy and exhaust
2
emissions. Fuel economy has become one of the most important factors when consumers
choose a new car as gas prices increase. For better fuel economy, the vehicle and the
powertrain efficiencies should be evaluated and improved. Better vehicle efficiencies can
be achieved through the reduction of vehicle weight, rolling resistance and aerodynamic
drag. The powertrain efficiency, especially internal combustion engine fuel economy,
can be improved through the analysis and understanding of how the fuel energy is
consumed during the engine cycle.
Brake Power25%
Cooling Losses
30%
Exhaust30%
Mechanical Losses
15%
Figure 1.2. Typical fuel energy distribution in an internal combustion engine.
The distribution of available fuel energy in an internal combustion engine is
shown in Figure 1.2. Even though the values in Figure 1.2 depend on the engine type and
operating conditions, the overall proportion of each component are meaningful. In Figure
1.2, it can be known that only about 25% of the fuel energy is available as brake power.
The rest of the fuel energy is eventually dissipated in the form of heat. Approximately
15% of the input fuel energy is consumed as mechanical losses at full load. Although the
mechanical friction losses consume only 15% of the input fuel energy at moderate and
high loads, most of the indicated power in the idle or low load condition is used for
engine mechanical and pumping losses. In addition to effects of mechanical losses on
fuel economy, the power demand for starting the engine is diminished and so the electric
3
parts such as start motor, battery and alternator could be smaller if the mechanical losses
are minimized. Therefore, the reduction of mechanical friction losses in an internal
combustion engine is important for better fuel economy and compact electrical
components. Via both theoretical and experimental studies, many researchers have
studied the frictional contributions and tribological characteristics of each engine
component.
The engine frictional losses can be classified into four main components:
piston/ring assembly, valve train system, bearing system, and auxiliaries (water pump, oil
pump, alternator, etc). Figure 1.3 shows the general proportions of the frictional loss of
each engine component, although these proportions can be changed according to engine
speed, load, and the type of engine and also vary greatly within the literature.
Piston Assembly45%
Valve Train10%
Pumping Losses20%
Bearing 25%
Figure 1.3. Representative mechanical loss distribution.
From Figure 1.3, it can be said that most of the engine friction losses are from the
piston assembly. Therefore, it is necessary to attack the piston assembly friction to
achieve low engine friction since the piston assembly (piston skirt, piston rings and piston
4
pin) accounts for about half of total engine friction for motoring conditions, and an even
higher fraction of total engine friction for firing operation. Many researchers have made
a great effort to understand the tribological phenomena and reduce the frictional losses of
the piston assembly. However, the friction reduction of the piston assembly still remains
a challenging area due to the complexities of the tribological phenomena and the
interrelations among friction, emissions, durability, noise and vibrations, oil consumption,
blow-by, etc.
In general the characteristics of the lubrication phenomena in the piston
assemblies can be explained using the Stribeck curve. Figure 1.4 shows a general
Stribeck diagram representing the various lubrication regimes. In the Stribeck diagram,
the lubrication regime can be classified into three regions: boundary lubrication,
hydrodynamic lubrication, and mixed lubrication. In the boundary lubrication region, the
asperities between two rubbing surfaces come into contact and become dry friction. The
surface properties and lubricant additives influence the friction losses in this region while
the friction coefficient is independent of lubricant viscosity, surface speed, and load. In
the hydrodynamic region there is no direct contact between the two surfaces. The
lubricant film separates the two surfaces completely. Therefore, the friction losses in the
hydrodynamic region mainly come from shear forces of the fluids moving at different
velocities between two surfaces. The mixed lubrication region lies between these two
extremes.
5
Valve train
Coe
ffic
ient
of
fric
tion
(Viscosity) x (Speed) / Load
Engine bearings
Piston rings
Mixed LubricationBoundary Lubrication Hydrodynamic Lubrication
Piston skirt
Figure 1.4. Stribeck diagram showing the various regimes of lubrication.
Basically the main functions of the piston ring assembly are to seal 1) the high
pressure combustion gas from the combustion chamber to the crankcase and 2) the oil
from the crankcase to the combustion chamber. Optimized piston/ring pack designs
should fulfill their functions with minimum friction losses and wear. That is, the
tribological characteristics such as friction, lubrication and wear of the piston/ring
assembly are equally important. However in this dissertation the main concern will be
concentrated to minimize the friction losses of the piston/ring assembly.
The lubrication phenomena of piston rings are extremely complicated due to the
variation of the piston speed, piston ring dynamics, and interactions of the cylinder gas
and lubricant film among the ring, ring groove, and the cylinder liner. Many researchers
have made progress in analyzing the lubrication phenomena of the piston assembly. The
results of their research have proven that the basic frictional mechanism of the piston/ring
assembly is the combination of boundary, mixed and hydrodynamic lubrication. As the
oil viscosity, piston speed, and load are changed, the lubricant regime of the piston
assembly also changes. As the piston approaches TDC (Top Dead Center) and BDC
(Bottom Dead Center), the piston speed becomes zero momentarily and the duty
parameter (the x-axis) in the Stribeck curve approaches zero. Therefore, the lubrication
regime near the TDC and BDC positions becomes boundary lubrication. In the mid
piston stroke, the piston speed attains a maximum value and hydrodynamic lubrication
6
becomes dominant. Mixed lubrication occurs during the transitions between
hydrodynamic and boundary lubrication. On the up-stroke, the squeeze film effect of the
oil can delay the transition to mixed and boundary lubrication. In addition to the
complicated friction mechanism of the piston/ring assembly, the friction reduction of the
piston assembly is also interrelated with oil consumption, blow-by, wear and other engine
durability problems. Therefore, it is still a challenging problem to reduce the piston/ring
assembly friction losses in spite of its importance.
1.2 Motivation
Although the reduction of piston assembly friction is a difficult and challenging
problem due to its complexities in lubrication phenomena and its interrelation with other
problems such as oil consumption and blow-by, it is the most effective way to minimize
the engine mechanical friction losses. It is generally known that at BDC and TDC the
lubrication regime becomes boundary lubrication because the piston speed is
momentarily zero in the vicinity of each dead center. However, because of the squeeze
film effect of the oil film, the direct surface contact between the ring and the liner is
minimized. The lubrication regime between the piston ring and the liner is hydrodynamic
during the middle region of the piston stroke, where the piston speed is high and the gas
pressure is relatively low. Therefore, the hydrodynamic, the boundary, and the mixed
lubrication regimes are all encountered while the piston is moving from BDC to TDC.
Basically, the Rotating Liner Engine (RLE) concept is to force the portions of the
stroke for which the piston assembly normally operates in the boundary and mixed
lubrication regimes to operate instead in hydrodynamic lubrication. The RLE rotates the
cylinder liner to ensure relative motion between the piston skirt and the liner and also
between the piston rings and the liner. That is, rotating the liner can produce a large
relative velocity between the piston and the liner in the very low piston speed regions
near BDC and TDC. Thus, even at the BDC and TDC positions, the lubrication regime
of the piston can remain hydrodynamic in the RLE. Thus, the RLE concept can reduce
piston assembly friction, especially in low piston speed applications such as heavy-duty
diesels and large bore stationary natural gas engines for which, due to the high gas
7
pressures, a large portion of the friction loss is due to mixed-boundary lubrication. The
RLE hardware, especially the seal between the rotating liner and the stationary cylinder
head, was first developed and designed by the friction research team at the University of
Texas at Austin. However, the friction-reducing effect of the RLE has only been
tentatively confirmed via initial experimental measurements. The goal of the present
research was to quantify the friction reduction of the piston assembly in the RLE using
more sophisticated experimental methods than were used for the initial comparisons
(Dardalis, 2003).
1-3 Dissertation overview and scope of work
Even though the RLE was initially developed mainly by Dr. Dimitrois Dardalis
several years ago (Dardalis, 2003), the friction reduction effect of the RLE was not
confirmed except via preliminary motoring tests using a crude dynamometer. The main
purpose of this dissertation is to confirm the RLE effect on piston assembly friction
reduction using sophisticated measurement methods. Three different friction
measurement methods were applied in measuring the friction force difference between
the baseline engine and the RLE. Through the use of three different friction
measurement methods, the friction reduction of the RLE is confirmed via the present
research and each friction measurement method is also compared based on its
measurement results. The analysis of the friction mechanism of the baseline engine was
performed using the instantaneous IMEP method and a commercial simulation program
called RINGPAK. Through the use of experimental methods and the simulation, the
friction mechanism of the piston/ring assembly is analyzed. The limitations of the
experimental and the calculation methods are also discussed.
This dissertation is composed of seven chapters. Chapter 1 describes the
background knowledge about mechanical friction losses of internal combustion engines
and basic concepts of the RLE. Chapter 2 summarizes the different friction measurement
techniques and their basic ideas. Chapter 3 explains the overall test setup including the
baseline engine and the RLE component descriptions, sensors such as the in-line torque
sensor, cylinder pressure sensor, strain gage, dynamometer, and couplings. Chapter 4
8
includes the friction measurement test results of the baseline engine and the RLE. In
Chapter 4 the measured friction results using the different measurement techniques are
analyzed and discussed. In Chapter 5 the simulation results using RINGPAK software
are presented and compared with the experimental test data. Chapter 6 summarizes the
results of this research and presents the conclusions that can be drawn from this study.
Recommendations for future work in this area are provided in Chapter 7.
9
Chapter 2. Test Methods
2.1 Overview of friction measurement methods
Due to the importance of friction losses in an internal combustion engine, many
methods have been developed to measure the engine friction loss with high accuracy. In
spite of its importance and endeavors to measure the engine friction, accurate
measurements of the engine friction losses are not easy since the amount of friction is
relatively small compared with other powers such as brake power, cylinder pressure
power, and so on. Accurate measurements of the piston/ring assembly friction are an
especially difficult and challenging problem in spite of its importance. Several different
friction measurement methods are introduced and explained in the following subsections.
2.1.1 Direct motoring test with tear-down
Direct motoring of the engine, under conditions as close as possible to firing, is
widely used to estimate mechanical frictional losses of internal combustion engines due
to its simplicity and capability to measure the friction loss of each engine component via
tear-down tests [1, 2, 3]. The engine temperature should be maintained as close as
possible to normal operating temperature by heating the cooling water and the engine oil.
The power required to motor the engine without firing includes the engine mechanical
friction loss and pumping loss. Thus, in order to get the pure mechanical friction losses
from measured motoring power the pumping loss should be measured. To minimize the
pumping loss, the throttle plate should be held wide open during these tests, but this may
stil yield a small pumping loss. In general the engine pumping loss can be calculated
using cylinder pressure measurement. The engine motoring torque with tear-down
techniques can be used to identify the contribution of each major engine component to
total friction losses. Although the motoring condition tries to simulate the firing
condition as closely as possible, the motoring frictional losses are different from those for
firing conditions due to the following reasons:
10
The cylinder pressure acting on the piston, piston rings and bearings is
lower for motoring than for firing conditions.
Piston and cylinder bore temperatures are lower under the motoring
condition than the firing condition. Thus, the viscosity of the lubricant in
motoring is higher than that of a firing engine.
The clearance between the piston and the cylinder is not the same for
motoring and firing conditions.
The pumping loop is different due to the exhaust effect of firing engine
conditions.
2.1.2 Morse test
The Morse test involves firing a multi-cylinder engine and recording its power
output. Then, power is cut from a single cylinder, the engine is adjusted to the previous
speed, and then power is again measured [4]. The power difference is the piston
assembly friction for that particular cylinder. This is done for all cylinders, and the
values are added to determine the total engine friction. This method has the same
accuracy problems as the direct motoring method, even though the error is smaller
because the continuous firing of the other cylinders will keep the non-firing cylinder
closer to normal operating temperatures. The major difficulty for this test is that a multi-
cylinder engine is required.
2.1.3 Willans line method
The Willans line is a method that can be used when sophisticated instrumentation
is not available [3, 4]. Engine load is measured and plotted at constant speed as a
function of engine fueling rate. The curve is then linearly extrapolated to zero fuel flow
which results in a negative load. This negative load indicates the losses of the engine.
This method gives not only the mechanical friction but includes the pumping losses also.
Generally the plot has a slight curve, which makes accurate extrapolation difficult.
11
2.1.4 Measurement of FMEP from IMEP and BMEP
FMEP (Friction Mean Effective Pressure) can be calculated after measuring
IMEP (Indicated Mean Effective Pressure) and BMEP (Brake Mean Effective Pressure).
The MEP (Mean Effective Pressure) is computed by
Nd
VR
PnMEP = (2.1)
where
P: power
nR: number of crank revolutions for each power stroke per cylinder
Vd: engine displaced volume
N: engine speed
IMEP is calculated from the cylinder pressure which is usually measured using a
piezoelectric pressure transducer. Figure 2.1 shows an example of measured cylinder
pressure.
Cyl
inde
r Pr
essu
re
Cylinder Volume
CB
A
Figure 2.1. Example of measured cylinder pressure.
12
IMEP is defined as
NV
nIPIMEP
d
R⋅= (2.2)
R
i
n
NWIP
⋅= (2.3)
∫= PdVi
W (2.4)
where
IP: indicated power
Wi: gross indicated work per cycle
The area between the exhaust and intake strokes (the pumping loop) in Figure 2.1
is negative work (pumping work). There is a lack of universal agreement regarding
whether this loss should be accounted for in the determination of mechanical efficiency
(i.e., treated as a component of the mechanical losses) or in the determination of the
indicated thermal efficiency. In turn, this controversy leads to two definitions of the
indicated mean effective pressure and of the indicated power. Gross IMEP (GIMEP) is
the constant pressure acting over the stroke that would yield the same power delivered to
the top of the piston as the variable pressure over the compression and expansion strokes
only. Thus, from Figure 2.1, GIMEP is defined as (area A + area B) divided by the swept
volume per cylinder. That is, the GIMEP and corresponding gross indicated power are
used by those who prefer to treat the pumping work (and corresponding pumping mean
effective pressure, PMEP) as a component of the mechanical losses. However, it can be
logically argued that the pumping work is the thermodynamic penalty paid for operating
a spark ignition engine, which requires use of a throttle plate in order to operate at part
load. In this case, the appropriate indicated power involves the pumping power which is
required to purge the cylinder of exhaust gases and fill it with fresh fuel and air. Thus, in
order to calculate the net power transferred to the piston during the engine cycle, the
pumping power should be accounted for in the measured indicated power. Pumping
work during the engine cycle is shown in Figure 2.1 as (area B + area C). Thus, the net
IMEP is computed by subtracting the PMEP from the gross IMEP and identified as (area
A) in Figure 2.1. Brake power is defined as the available power output from the engine.
13
BMEP can be measured using a load cell or torque sensor in the engine dynamometer
test. Through the use of the measured net IMEP and BMEP, the FMEP (Friction Mean
Effective Pressure) can be calculated in a manner that includes only mechanical losses
but not pumping losses. FMEP is an integral value for the total engine friction during
fired engine operation. In summary
NMEP=GIMEP-PMEP (2.5)
GIMEP=BMEP+FMEP+PMEP (2.6)
FMEP=GIMEP-BMEP-PMEP (2.7)
where
PMEP: Pumping Mean Effective Pressure
NMEP: Net Indicated Mean Pressure
FMEP results include all mechanical frictional losses and so do not provide
detailed information about piston/ring friction forces.
14
2.1.5 Floating liner method
The floating liner method, which was developed at Furuhama’s Internal
Combustion Engine Research Laboratory (Furuhama et al., 1981; Hoshi et al., 1989), is
widely used for fired engines by many researchers because this method can directly
measure the friction force of the piston assembly without any assumptions. The basic
idea of this method is to measure the friction force using a load cell installed in the
cylinder liner, which is allowed to float in the axial direction. Figure 2.2 shows the
typical friction measurement system using the floating liner method.
Cylinder liner
Lateral Stopper
Load cell
Seal o-ring
O-ring holder
Figure 2.2. Schematic view of the floating liner method.
The liner is separately fabricated from the cylinder block, which supports it with
specially designed liner supporting devices. Therefore, the piston system friction causes
a small displacement of the floating liner in the axial direction of the liner, and the load
cell installed between the lower part of the liner and the block senses the force, which is
due to piston assembly friction. The advantage of this method is that the piston friction
force can be measured directly without any assumptions or calculations. Thus, many
researchers have used this method. The load cell is installed to sense a small
displacement of the cylinder liner due to the piston friction force between the block and
15
the liner. In the floating liner method, the most important thing is to separate the cylinder
pressure force and piston thrust force from the piston axial force. This is because the
cylinder pressure force and the piston thrust force are much greater than the piston axial
friction force, so the small magnitude of these forces can affect the real piston friction
force measurement. Therefore, many researchers suggested their own hardware
modifications to prevent the effect of the cylinder pressure and piston thrust force on the
friction force measurement [2, 5, 9]. The greatest shortcoming of this method is the need
for extensive engine hardware modifications. That is, the liner should float in the
cylinder block to move in the axial direction only and also the combustion chamber has
to be sealed to prevent combustion from being unstable.
2.1.6 Instantaneous IMEP method
Uras and Patterson (1984) at the University of Michigan developed the
instantaneous IMEP method to measure the instantaneous friction force of an actual
production engine without extensive engine hardware modifications. The main idea of
the instantaneous IMEP method is that, since the direction of the forces at the piston are
known at each crank angle, if any three of the total forces (gas force, friction force, thrust
force, inertial force) in the axial direction are known, the fourth force can be determined
from a force balance. Figure 2.3 shows the simplified forces and moment acting on the
piston.
Fgas
FthrustFfriction
Fcon_rod
M
Figure 2.3. Free body diagram of a piston.
16
This method requires measurement of the instantaneous connecting rod force,
instantaneous cylinder pressure, engine speed and crank position, and calculation of the
inertial force. The gas force is calculated from the piston area and the cylinder pressure
measured using a piezoelectric pressure transducer. The connecting rod force is
measured by a strain gage installed in the connecting rod. A grasshopper linkage is
normally used to transmit the connecting rod force signal measured by a strain gauge
bridge to the instrumentation. The strain gage must be carefully located to compensate
for bending and temperature effects.
The advantage of this method is that there is no need for the modification of the
block and cylinder head except the need for modification of the oil pan to install the
grasshopper mechanism. However this method also has several drawbacks. One of the
disadvantages of this method is that the resulting calculated friction force is dependent
upon the two relatively large force measurements using two different measuring
techniques. That is, because the friction force is calculated from the force balance
between the large cylinder gas force and the connecting rod force, accurate measurement
of these two large forces is indispensable for the accuracy of the friction force calculation.
The factors affecting the accuracy of the cylinder pressure measurement are the accuracy
of the TDC position and the sensitivity and the thermal drift of the piezoelectric pressure
transducer. The accuracy of the connecting rod force measurement is dependent upon
minimizing the sensitivities of the bending moment and installing the strain gage as close
to the center of gravity of the rod as possible. Also, the sensitivity of the strain gage
should be calibrated according to the temperature. Therefore, elaborate effort is needed
to install and calibrate the strain gage. Another shortcoming is the requirement of the
linkage mechanism for transmitting the connecting rod force signal to the instrumentation.
Usually, a grasshopper mechanism is used for this purpose.
2.1.7 P-w method
Rezeka and Henein (1983) at Wayne State University developed the P-w method
to determine the instantaneous friction in internal combustion engines. The main idea of
17
this method is based on the fact that the instantaneous cylinder pressure, the
instantaneous frictional, inertial, and load forces cause an instantaneous variation in the
flywheel angular velocity. Thus, the P-w method requires the measurement of the
cylinder pressure and the instantaneous angular velocity of the flywheel to calculate the
instantaneous components of the frictional losses. The total friction torque (Tf) can be
expressed in terms of the torque generated by the gas pressure (TG), the inertia of the
reciprocating parts (TIN(rec)), the inertia of the rotating parts (TIN(rot)), and the load (TL).
LrotINretINGf TTTTT −−−= )()( (2.8)
The instantaneous friction torque can be expressed as a function of gas pressure,
instantaneous angular velocity, engine design parameters (such as connecting rod length
and crankshaft radius, etc.) and operating parameters.
LINgasf Tw
Il
rm
wwrKFmgFrKT −∆
∆−∆∆−+= θθαθ 2/
,...,,2/
,]cos[)(22
(2.9)
where
r: crank shaft radius
l: connecting rod length
K: geometrical transformation factor
I: rotational inertia calculated from the geometry and masses of rotating parts which
consist of the flywheel, crankshaft, and part of the connecting rod
Fgas: gas pressure force calculated from the measured gas pressure and piston area.
m: mass of the reciprocating parts which consist of the piston assembly and the
reciprocating part of the connecting rod.
g: gravitational acceleration
Fin: inertia force of the reciprocating parts calculated from )( 2
θd
dKwwKmrFIN += •
(2.10)
TL: load torque (brake torque).
18
The friction torque Tf(θ) can be calculated from the measured cylinder pressure, the
load torque and the calculated inertia torque. Basically, the friction torque Tf(θ) includes
all engine friction torques such as crankshaft bearing friction, piston assembly friction,
camshaft friction, etc. Thus, the calculated friction torque Tf(θ) using measured data
should be compared with the theoretically modeled friction torque in order to determine
the contributions of each engine component to total friction torque. The theoretical
friction torque is derived from modeling of the individual frictional losses. Table 2.1
summarizes each frictional loss model.
Friction components Modeling
Ring viscous lubrication friction 1105.0
11 ||)4.0(])([ waKrnnDwPPvaT cgasef ≡+•+= µ
Ring mixed lubrication friction 2222 |||sin|1)( waKrPPwDnaT gasecf ≡−+= θπ
Piston skirt friction 3333 )()( warKDMh
vaTf ≡••= µ
Valve train friction 4444 /|| waKrGLaT sf ≡= ω
Auxiliaries and unloaded
bearing friction5555 waaTf ≡= µω
Loaded bearing friction 662
66 /|cos|4
waPrDaT gascf ≡= ωθπ
Table 2.1. Modeling of each friction component.
The theoretical total friction torque is modeled as the sum of each friction
component’s friction loss. That is, the modeled total friction torque is assumed as the
linear combination of its individual components in the P-w method as follows.
19
∑=
=++=6
16611)(
jjjth wawawaT Lθ (2.11)
On the other hand, the measured friction torque Tf(θ) can be related to the modeled
friction torque Tth(θ) by
θεθθ += )()( thf TT (2.12)
That is
θθ εθ +=∑=
6
1
)(j
jjf waT (2.13)
where εθ is an error from the uncertainty of the measured Tf(θ) and the non-linearity
connected with the modeling. Mathematical techniques are used to find the coefficients
aj which minimize the error εθ. The optimum coefficient of the linear combination of the
individual torque components is determined by the linear regression technique to satisfy
the experimental torque results. The main advantage of the P-w method is that the
minimum measurement effort is required to obtain the frictional loss of each engine
component. However, the oversimplified modeling of the frictional loss of the engine
components makes it difficult to be used to analyze the engine friction loss in detail.
2.2 Friction measurement methods in this research
In this study, the hot motoring test with tear-down is used to examine the
frictional difference between the baseline engine and the RLE. Although the motoring
friction is not exactly the same as firing mechanical friction, the motoring method with
tear-down tests can give information about the engine component’s frictional losses.
Using the motoring method, the piston assembly frictional losses between the baseline
engine and the RLE can be determined. Since the motoring method does not provide any
information about differences in the lubrication mechanism between the baseline engine
and the RLE, the lubrication mechanism difference has been evaluated using both the
instantaneous IMEP method and the P-w method. Basically, the instantaneous IMEP
20
method and the P-w method are indirect methods in that both methods calculate the
friction from the measured cylinder pressure, crankshaft speed and the connecting rod
force. Therefore, the crankshaft and piston dynamics are required to compute the friction
forces when applying the instantaneous IMEP method and the P-w method.
2.2.1 Instantaneous IMEP method
The instantaneous IMEP method can determine the piston assembly friction force
from the measured cylinder pressure, the connecting rod force, and the piston assembly
inertia force. The cylinder pressure is measured using a piezoelectric pressure transducer.
For the measurement of the connecting rod force and the piston assembly inertia force,
information about the strain gage measurements and the piston dynamics are needed, as
discussed in the following subsections.
1) Strain gage measurements
Strain gages were installed in the connecting rod to measure the connecting rod
force. The force measurements using strain gages usually utilize the constant voltage
Wheatstone bridge circuit. Figure 2.4 shows the basic Wheatstone bridge circuit.
R1 R2
R3 R4
Ei
Eo
Figure 2.4. Wheatstone bridge circuit.
21
In Figure 2.4, the Wheatstone bridge circuit is composed of three parts, which are a
constant voltage source Ei, four resistors R1, R2, R3, R4, and the measured output
voltage Eo. For a Wheatstone bridge, the output voltage Eo is given by
io ERRRR
RRRRE
))(( 4321
4231
++−= (2.14)
In Equation 2.14, the output voltage Eo will be zero when the numerator goes to zero.
That is,
4231 RRRR = (2.15)
When Equation 2.15 is satisfied, it is said that the Wheatstone bridge circuit is balanced.
Thus, from the balanced bridge condition the small unbalanced voltage caused by a
change in resistance can be measured. An output voltage oE∆ is developed if the
resistances of R1, R2, R3, and R4 are changed by 4321 ∆Rand,∆R,∆R,∆R . An output
voltage oE∆ can be expressed as in Equation 2.16:
iER
R
R
R
R
R
R
R
r
rE )1)((
)1( 4
4
3
3
2
2
1
120 η−∆−∆+∆−∆
+=∆ (2.16)
In Equation 2.16, r represents R2/R1=R3/R4. In this project r is set to 1. Also, η is defined as:
1
3
3
2
2
4
4
1
1
11
−
∆+∆+∆+∆++=
R
R
R
Rr
R
R
R
R
rη (2.17)
22
In Equation 2.17, η can be neglected if the strain is less than 5%. In most cases, η is
neglected since the strain is less than 5%. If r is set to one and η is neglected, then
Equation 2.16 can be simplified as
iER
R
R
R
R
R
R
RE )(
4
1
4
4
3
3
2
2
1
10
∆−∆+∆−∆=∆ (2.18)
Therefore, the resistance change of the strain gages can be observed by the measurement
of 0E∆ .
2) Piston dynamics
The piston assembly’s inertia forces must be calculated to get the friction forces
from the measured cylinder gas force and the connecting rod force. Figure 2.5 illustrates
the acceleration of the piston assembly where
R: crank radius
L: connecting rod length
φ ( tω ): angle of the crankshaft
θ: angle that the connecting rod makes with X axis
ω: crank angular velocity
x: instantaneous piston position
23
TDC
x
xr
it
in
2
r
...
..r
L
R
A
B
Figure 2.5. Acceleration of the piston and connecting rod.
The strain gage is installed at point A in Figure 2.5. The strain gage installed at
point A measures not only the transmitting force between the piston and crankshaft but
also the inertial forces of the connecting rod. Thus the inertial force at point A should be
removed from the measured strain gage value. Acceleration at point A can be expressed
as in Equation 2.19.
ABBA aaa += (2.19)
where
aA: absolute acceleration of point A
aB: absolute acceleration of point B
aAB: acceleration of point A with respect to the piston pin (point B)
The acceleration at A on the connecting rod is the sum of the absolute
acceleration of point B and the relative acceleration of point A with respect to the piston
24
pin. The absolute acceleration of point B is equal to the linear piston acceleration, x&& .
The absolute acceleration of point A can be rewritten using the unit vectors in and it. The
vector component in indicates the normal direction of the acceleration and it is transverse
directional vector component. Then the acceleration of A is
tnA irxirxa )sin()2cos( θθθθ &&&&&&& +−+−= (2.20)
In Equation 2.20, θ& and θ&& represent the first and second time derivatives of angle θ .
The inertial force in the transverse direction tends to bend the connecting rod. However,
since the Wheatstone bridge circuit is designed to compensate for the bending component
of the inertial force, the transverse inertial force can be neglected. Thus, the acceleration
of point A can be described as the normal component only. That is
` 2cos θθ &&& rxaA −= (2.21)
Equation 2.21 is expressed as a function of x&& , θ , r, and θ& . Basically, since the engine
crank angle and speed are measured using an optical encoder installed in the engine front
end, such as on the damper pulley, the parameters in Equation 2.21 have to be expressed
as a function of crank angle φ , crank rotation speed φ& , and crank rotational acceleration
φ&& . The parameter r represents the position of a strain gage on the connecting rod and is
constant. The linear acceleration of the piston translational motion, x&& , can be expressed
as a function of crank angle variables such as φ , φ& , and φ&& .
In Equation 2.21, θcos can be converted to a function of φ . From trigonometry
θφ sinsin LR = (2.22)
Therefore,
φθ sinsinL
R= (2.23)
25
22 )sin(1sin1cos φθθL
R−=−= (2.24)
In order to express x&& as a function of φ , φ& , and φ&& , x can be expressed as
θφ coscos LRx += (2.25)
From Equations 2.24 and 2.25,
2)sin(1cos φφL
RLRx −+= (2.26)
Equation 2.26 is an exact expression for the piston position x as a function of R, L, and φ.
This exact expression for the piston position can be differentiated versus time to obtain
the velocity and acceleration of the piston. However, it is difficult to determine the
effects of design parameters, such as R and L, on velocity and acceleration. Thus, a
simpler expression for the piston position is needed to examine the effects of the engine
design parameters on the piston velocity and acceleration. To do this, the binomial
theorem is used. The binomial theorem is as follows.
L+−−+−++=+ −−− 33221
!3
)2)(1(
!2
)1()( ba
nnnba
nnbnaaba nnnnn (2.27)
Using the binomial theorem
)2cos4
(cos4
]sin)2
(1[cos2
22
2
φφφφL
RR
L
RL
L
RLRx ++−=−+≅ (2.28)
Thus, the linear velocity x& and the linear acceleration x&& can be expressed as
)2sin2
(sin φφφL
RRx +−= && (2.29)
26
)2sin2
(sin)2cos(cos2 φφφφφφL
RR
L
RRx +−+−= &&&&& (2.30)
In Equation 2.21, θ& can be expressed if Equation 2.23 is differentiated; then
φφθθ coscos &&L
R= (2.31)
2)sin(1
cos
cos
cos
φ
φφθφφ
θL
RL
R
L
R
−==
&&& (2.32)
Using Equations 2.30 and 2.32, Equation 2.21 can be expressed in terms of φ , φ& , and φ&&That is
2
)sin(1
cos)2)sin(1))(2sin
2(sin)2cos(cos2(
2
−−−+−+−=
φ
φφφφφφφφφ
L
RL
R
rL
R
L
RR
L
RR
Aa
&&&&
(2.33)
Basically the friction force measurement using the instantaneous IMEP is
calculated from the measured pressure force, connecting rod force and the inertia force.
Therefore, a small error in each force measurement can cause an error in the calculated
friction forces.
2.2.2 P-w method
The main advantage of the P-w method is that there is no need to change the
engine hardware. However, the P-w method necessitates modeling the friction of each
engine component and the inertia torque of the crankshaft system including the piston
assembly must be calculated. For the inertia torque calculation, the piston and the
27
connecting rod dynamics should be investigated. Previous researchers also reported that
the calculated friction torque showed a negative value at some crank angles [14, 15].
However, negative friction torque is not physically possible since this would mean that
friction torque drives the engine at some crank angles. It was reported [16, 17] that one
of the possible reasons for the negative friction torque is due to the torsional vibration of
the crankshaft system. Therefore, the crankshaft dynamic characteristics must also be
investigated to understand the vibration characteristics of the crankshaft system.
1) Crankshaft dynamics
Crankshaft dynamics includes the piston, the connecting rod and the crankshaft
movement during the engine operation. Basically the piston and the connecting rod
movement can be described as a slider crank kinematics, as shown in Figure 2.5. In case
of the slider crank mechanism, the crankshaft is in pure rotation and the piston is in pure
translation. Therefore, their kinematical motions can easily be determined by the
assumption of geometries and materials. However, the connecting rod has a more
complex motion. For an exact dynamic analysis of the connecting rod, it is needed to
determine the linear acceleration of the CG of the connecting rod for all crank angles.
However, the connecting rod can be modeled as two lumped masses with negligible
errors in this application. That is, the connecting rod is modeled as two lumped masses,
concentrated one at the crankpin, and one at the wrist pin. The lumped mass
concentrated at the wrist pin has pure translation motion and the lumped mass at the
crankpin has pure rotation. From simple dynamic arguments, the two mass model of the
connecting rod should have dynamically equivalent characteristics with the original rod.
The following conditions are required to satisfy the dynamic equivalence between the
two lumped masses model and the original rod.
1. The mass of the model must equal that of the original body
2. The center of gravity must be in the same location as that of the original body
3. The mass moment of inertia must equal that of the original body
28
Figure 2.6 shows the lumped mass model of a connecting rod. The requirements for
dynamic equivalence can be expressed mathematically.
b
Gp
bp
pb
bp
bp
ML
IL
LL
LMM
LL
LMM
=+=+=
(2.34)
In Equation 2.34, Lp represents the center of percussion which is a point on a body at
which there is a zero reaction force when struck with a force. Therefore Lp is the location
of the center of percussion corresponding to a center of rotation at Lb. As shown in
Figure 2.6, the second mass Mp is placed at the link’s center of percussion P to obtain
exact dynamic equivalence. However, as an approximate model, it can be assumed that
one lumped mass is placed at the crankpin end A and the other mass belongs at the wrist
pin end, as shown in Figure 2.6 with a relatively small error. From the approximate
model as shown in Figure 2.6:
ba
ab
ba
ba
LL
LMM
LL
LMM
+=+=
(2.35)
Therefore, the dynamic behavior of the single cylinder engine can be described as
the rotational motion (such as crankshaft and connecting rod portion lumped at the
crankshaft pin end) and the translational motion of the piston assembly and connecting
rod part lumped at thewrist-pin end.
29
L
LaLb
CG
A
B
IGM ,
(a) Original connecting rod
CG
Mp
Mb
LpLb
(b) Exact dynamic model
CG
Ma
Mb
LaLb
(c) Approximate model
Figure 2.6. Lumped mass model of a connecting rod.
30
The instantaneous rotational speed of the crankshaft is determined from the torque
balance of the several torque sources. Engine output torque, often called the load torque,
is the resultant torque after the pressure torque generated from the combustion of a fuel in
the combustion chamber is balanced with the inertia torque and the friction torque of the
engine’s moving components, such as the piston assembly, connecting rod, bearings,
camshaft, crankshaft, and so on. That is
FIPL TTTT −−= (2.36)
where
TL: load torque
TP: pressure torque
TI: inertia torque
TF: friction torque
Therefore, each torque component in Equation 2.36 has to be calculated and
measured at each crank angle in order to apply the P-w method. Firstly, the pressure
torque generated from the combustion energy of the fuel should be calculated. The gas
pressure torque is the torque generated by the cylinder pressure due to the combustion of
the fuel and air mixture. This gas force can be divided into two components, as shown in
Figure 2.7.
F
2Fg
g
Fg1
Figure 2.7. Free body diagram for the wrist pin forces.
31
θcos1 gg FF = j (2.37)
θtanˆ2 ggg FiFF −= j (2.38)
where
Fg : cylinder pressure force acting on the piston
Fg1 : cylinder pressure force component acting on the cylinder wall
Fg2 : cylinder pressure force component acting on the connecting rod
The gas torque is the multiplication of the force component perpendicular to the slider
motion and the distance x, which is the length from the center of the crankshaft to that of
the wrist pin. Distance x can be expressed as in Equation 2.28 in terms of the
geometrical parameters. Therefore, the pressure torque generated by cylinder pressure is
+⋅+−⋅=•= )2cos4
(cos4
tan2
1 φφθL
RR
L
RLFxFT ggg (2.39)
Here the connecting rod angle θ can be expressed using crank angle φ.
2)sin(1
sintan
φφθ
L
RL
R
−= (2.40)
The denominator of the equation can be approximated using the binomial theorem.
φφ
22
2
2
sin2
1
)sin(1
1
L
R
L
R+≅
−(2.41)
Using Equations 2.40 and 2.41, the cylinder gas pressure torque in Equation 2.39 can be
expressed as
32
++−⋅
+⋅⋅≅ )2cos
4(cos
4sin
21sin
22
2
2
φφφφL
RR
L
RL
L
R
L
RFT gg (2.42)
The next step is to calculate the inertia torque due to the acceleration of the
masses in the system. The simplified lumped mass model is used to calculate the inertia
torque. The piston assembly experiences pure translation motion during engine
operation. Thus, the inertia torque of the piston assembly is caused by the linear
acceleration of the piston. The crankshaft motion can be simplified as a lumped mass,
which has pure rotational movement. Thus, the inertia torque of the crankshaft is from its
rotational acceleration. However, the kinematical characteristics of the connecting rod
are rather complicated. Here, the two-mass model is used in order to simplify the
calculation of the inertia torque of the connecting rod. Two lumped masses are
concentrated at the crankpin and wrist pin. The inertia torque results from the action of
two lumped masses on the crankshaft axis. The inertia force of the concentrated mass at
the crank pin can be divided into two components: radial and tangential. The radial
component has no effect on the moment. The total inertia torque caused by the rotational
inertia is from the crankshaft mass and the lumped mass of the connecting rod centered at
the crank pin end. Therefore, the rotational inertia torque is expressed in Equation 2.43.
dt
dMMT cccir
ω•+= )( (2.43)
where
Tir : inertia torque generated by rotational acceleration
Mc : lumped mass of the crankshaft
Mcc : lumped mass of the portion of the connecting rod centered at the crank pin
The inertia torque of the mass concentrated at the wrist pin can be expressed as follows.
33
( )( )
++−⋅
+⋅
+−⋅+=
+⋅+−⋅⋅⋅+−−≅•−=
)2cos4
(cos4
)sin2
1(sin)2cos(cos
)2cos4
(cos4
tan)(
22
2
22
2
tL
RtR
L
RLt
L
Rt
L
Rt
L
RtRMM
L
RR
L
RLxMMxFT
cwm
cwmiit
ωωωωωωω
φφθ&&
(2.44)
where
Tit: inertia torque generated by linear acceleration
Mm: lumped mass of the piston assembly (piston, piston rings, and wrist pin)
Mcw: lumped mass of the portion of the connecting rod centered at the wrist pin
Consequently, the friction torque can be calculated from the calculated gas pressure
torque, the inertia torque, and the measured load torque using Equations 2.42 and 2.44.
2) Transfer matrix method
The crankshaft system can be modeled using lumped masses, springs, and
dampers. The modeled dynamic system can be described as the combination of the linear
differential equations. In order to establish the correlation among the cylinder pressure,
the angular speed fluctuations, and the engine load, the differential equations of the
crankshaft system should be solved. Usually, there are three different ways to solve these
linear differential equations. Those are the transfer matrix method, direct numerical
integration, and modal analysis. Each of these three methods has its own advantages.
Direct integration and modal analysis can simulate transient operation of the system but
the transfer matrix method can be used only for steady state operation. However, the
transfer matrix method can be used for reverse calculation. That is, it is possible to
calculate the exciting torque using the measured crankshaft speed using the transfer
matrix method. In this dissertation the transfer matrix method will be used to analyze and
simulate the crankshaft system dynamics.
The state vector at a point of an elastic system can be expressed as combinations
of displacement and internal forces of the point i. The simple torsional system is
composed of an elastic massless shaft with concentrated masses along its length. Figure
34
2.8 shows a torsional system which has a massless shaft with distributed disks of
concentrated mass.
I I I I
i-2 i-1 i i+1
z zL R
i-1 i i+1i-2
i i
Figure 2.8. Massless shaft with disks.
In Figure 2.8, the shaft is assumed to be elastic and have no rotational inertia. The disks
are also considered to be rigid and have a rotational moment of inertia Ii. Figure 2.9
represents a free body diagram of a shaft and disk.
M M i
i-1
LRØ Ø
i-1
i
LR I w2i Øi
iML
M iR
Figure 2.9. Free body diagrams of a shaft and disk.
In Figure 2.9, the equilibrium of the shaft can be used to derive two equations. The first
is from the equilibrium of the shaft torque.
35
Ri
MLi
M1−= (2.45)
The second is from simple strength of materials.
iT
iRiR
iL
i GJ
lM
)(1
1−
− =−φφ (2.46)
In Equation 2.46, li is the length of the shaft, JT is the polar second moment of area of the
shaft, and G is the shear modulus of the material. Equations 2.45 and 2.46 can be
expressed using matrix notation:
R
ii
T
L
iM
GJM
110
11
−
•
=
φφ
(2.47)
Rii
Li zFZ 1−= (2.48)
iZ is a state vector on either size of the disk i. The matrix F is called the field
transfer matrix. For the disk i, the twist angle of the left and right side of the disk is the
same.
Ri
Li φφ = (2.49)
From the torque balance of the free body diagram in Figure 2.9:
02 =+− iiLi
Ri IMM φω (2.50)
Therefore, as a matrix expression
36
L
i
R
iMIM
•
−=
φ
ωφ
1
012 (2.51)
Lii
Ri zPz = (2.52)
The matrix Pi is known as the point transfer matrix. Therefore, a more complex dynamic
system can be depicted using the combination of a field transfer matrix and a point
transfer matrix. For example, a shaft with six disks is shown in Figure 2.10..
0 1 2 3 4 5 6 7
Figure 2.10. Mass-less shaft with six disks.
In Figure 2.10, the matrix relation between state vectors can be expressed as follows.
011 ZFZ L = LR ZPZ 111 = RL ZFZ 122 = … RL ZFZ 566 = LR ZPZ 666 = RZFZ 677 = (2.53)
From Equation 2.53, it is possible to obtain the relation between the state vectors at the
two ends of the shaft:
0112233445566 ZFPFPFPFPFPFZ = (2.54)
37
In this manner all the intermediate state vectors have been eliminated and the states at the
two ends of the beam can be expressed as
06 UZZ = (2.56)
11223344556 FPFPFPFPFPFU = (2.57)
U is transfer matrix between the state vectors. The transfer matrix is a function of the
stiffness, damping and inertia of the system. The transfer matrix can be determined if the
physical properties of the system is calculated. Thus, the state vector of the shaft can be
determined using the transfer matrix and the boundary conditions.
38
Chapter 3. Test Setup
This chapter explains the experimental equipment and instrumentation that were
used for the measurements that are discussed in this dissertation. The test engines are
discussed in the next section. The in-line torque sensor and couplings are discussed in
Section 3.2. Section 3.3 provides details regarding the strain gage measurements. The
cylinder pressure measurement and data acquisition systems are discussed in Section 3.4.
Section 3.5 is a summary of this chapter.
3.1 Test engines
Two nominally identical engines were used for the present research. The
conventional (“baseline”) engine is discussed in Subsection 3.1.1. The prototype
Rotating Liner Engine is discussed in Subsection 3.1.2.
3.1.1 Baseline engine
As the baseline engine, a General Motors 2.3-L dual overhead cam, 16-valve, 4-
cylinder Quad 4 engine was used. This engine was converted to a single cylinder version
by removing 3 pistons and connecting rods, replaced with bob weights with mass =
piston assembly plus half connecting rod. The engine specifications are provided in
Table 3.1.
39
Type Inline 4 cylinder
Displacement (L) 2.26
Stroke (mm) 85
Bore (mm) 92
Compression ratio 10:1
Connecting rod length (mm) 147
Intake valve opening 22° before TDC
Intake valve closing 45° after BDC
Exhaust valve opening 120° after TDC
Exhaust valve closing 20° after TDC
Table 3.1. Baseline Engine Specifications
3.1.2 Rotating Liner Engine
The RLE engine hardware was first developed by Dr. D. Dardalis (2003) of the
friction research team at the University of Texas at Austin. The prototype RLE is a
single cylinder version converted from an in-line 4-cylinder baseline engine. The
components of the RLE are comprised of three main parts: the face seal, the rotating liner,
and the liner-driving mechanism. Figures 3.1 and 3.2 are pictures of the RLE assembly
and its components.
40
Figure 3.1. Prototype Rotating Liner Engine assemblies.
41
Face seal Rotating liner
Thrust bearing Head insert
Driving gear Stationary liner
Figure 3.2. Rotating Liner Engine components.
42
The active cylinder is the number 2 cylinder from the 4-cylinder baseline engine.
In Figure 3.1, the face seal for cylinder 2 can be seen as an insert in the cylinder head.
The face seal achieves the sealing between the rotating liner and stationary cylinder head.
Figure 3.2 shows each engine component of the RLE in more detail. Figure 3.3 shows an
assembly drawing of the seal in the head insert and the secondary seals (O-rings).
Figure 3.3. Final seal design and installation.
The side view cross-section in Figure 3.3 shows the ring along with the O-rings
and O-ring glands. Oil is pumped between the ring insert (the part that surrounds the
sealing ring) and the ring itself. Then the oil flows through the holes through the ring
(not shown), and into the ring face annular groove (barely visible in the bottom of the
picture). The seal is placed into the head insert, which is installed inside a groove
machined on the cylinder head such that the lower face of the seal is on about the same
plane as the face of the cylinder head. The functions of the head insert part are multiple.
First, it seals against the head water jacket that was opened as the groove in the head was
machined to house the insert. Second, it houses the inboard and outboard O-rings. The
inboard O-ring achieves the secondary gas sealing. The inboard O-ring contains the
pressurized oil that is pumped by the oil pump on the upper region of the seal, between
the insert and the seal. The outboard O-ring also helps contain the pressurized oil in the
upper seal area. The pressurized oil has two functions. First, it provides about 2/3rds of
the pre-load of the seal (about 1/3rd of the pre-load is provided by coil springs
compressed between the ring and the head insert). Second, it provides the lubrication for
the face of the seal through the holes.
43
Figure 3.4 Cross-section showing the face seal
In Figure 3.4, the combustion chamber is on the right and the seal ring installation
is illustrated. The head insert and O-rings are missing in Figure 3.4 to illustrate the
machining in the cylinder head necessary to install the insert. Also, in this picture, the
rotating liner flange is in two parts. Cylinder liner rotation is actuated by an electric
motor that rotates a shaft that drives a gear in the upper part of cylinder 1 which, in turn,
rotates a gear at the top of the rotating liner of cylinder 2. This shaft extends through the
spark plug hole of cylinder 1. Figure 3.5 indicates the general driving mechanism for
rotating the liner. Figure 3.6 is a real picture of the RLE driving mechanism.
44
Electric MotorTorque Sensor
Coupling
Figure 3.5. Driving mechanisms for rotating the liner.
Figure 3.6. Real view of driving mechanism.
45
An electric motor rotates the liner and the torque generated by an electric motor is
transferred to the liner drive shaft using a belt-pulley mechanism. The functions of the
belt-pulley system include both speed reduction and power transmission. A torque sensor
is installed to measure the torque required to rotate the liner. The gear installed at the end
of the driving shaft transmits the torque to the mating gear at the top of the rotating liner.
3.2 Torque sensor and coupling
Two torque sensors were used to measure the motoring friction torques. One is
installed between the engine and the dynamometer to measure the overall engine friction
torque and the other is for measuring the torque required for RLE liner rotation. The
detailed specifications of the two torque sensors are provided in Table 3.2.
ModelCooper.
LXT963
Cooper.
LXT962
Capacity 2000 in-lbs 100 in-lbs
Linearity ± 0.1% F.S. ± 0.1% F.S.
Histeresis ± 0.1% F.S. ± 0.1% F.S.
Repeatability ± 0.1% F.S. ± 0.1% F.S.
Output 4.1418 mV/V 2.0861 mV/V
Table 3.2. Torque Sensor Specifications
Figure 3.7 shows the calibration data of the two torque sensors. The coupling
used to connect the torque sensor to the AC motor (KTR Corp.) is a torsionally stiff,
flexible steel laminar coupling (RADEX 60-NN, KTR Corp.) which is able to
compensate for shaft misalignment caused by, for example, thermal expansion.
46
0 10 20 30 40 50 60 70 80 90 1000.0
0.5
1.0
1.5
2.0
2.5
3.0
Out
puts
(m
V/V
)
Load (in lbf)
Torque Sensor Calibration
LXT962
0 200 400 600 800 1000 1200 1400 1600 1800 20000.0
0.5
1.0
1.5
2.0
2.5
3.0
Out
puts
(m
V/V
)
Load (in lbf)
LXT 963
Figure 3.7. Torque sensor calibration curves.
3.3 Strain gage measurements
The connecting rod force, required for use in the instantaneous IMEP method,
was measured using strain gages. Of the many experimental methods available for
measuring strain, the electrical resistance strain gage is used for the instantaneous IMEP
method. The three basic types of electrical strain gages are metal foil, semiconductor,
and liquid metal. All are widely used for measuring a strain in mechanical systems. In
this research, metal foil strain gages were used for strain measurement. Metal foil strain
gages were first developed by Sanders and Roe in England in 1952. The photographic
production of a master image of the grid configuration followed by a foil etching process
used to produce this type of electrical resistance strain gage permits miniaturization of the
grids, versatility of the grid configuration, economics of production, and enhanced
control of quality of the gage.
47
3.3.1 Strain gage specifications
The metal foil type electrical strain gage produced by Measurement Group Inc.
was installed and used for connecting rod strain measurements. The part number of the
strain gage is CEA-06-062UT-350. “CE” means the flexible gages with a cast polyimide
backing and encapsulation featuring large, rugged, copper coated solder tabs. This
construction provides optimum capability for direct lead-wire attachment. “A” represents
constantan alloy in self-temperature-compensated form. Thus, “CEA” is a universal
general purpose strain gage which has a constantan grid completely encapsulated in
polymide, with large, rugged copper-coated tabs. CEA is primarily used for general
purpose static and dynamic stress analysis. The normal operating temperature range is
from -100ºF to 350ºF. The strain range is ±3% for gage lengths under 1/8 inch and ±5%
for 1/8 inch and over. “06” represent the self temperature compensation (STC) number.
Figure 3.8. Typical thermal output variations with temperature for self-temperature-compensated constantan (A-alloy) and modified Karma (K-alloy) strain gages.
48
Figure 3.8 illustrates the thermal output characteristics of typical A- and K- alloy
self-temperature-compensated strain gages. As demonstrated by the data in Figure 3.8,
the gages are designed to minimize the thermal output over the temperature range from
about 0° to +400° F (-20° to +205° C). When the self-temperature-compensated strain
gage is bonded to a material having the thermal expansion coefficient for which the gage
is intended, and when operated within the temperature range of effective compensation,
strain measurements can often be made without the necessity of correcting for thermal
output. The two-digit S-T-C number identifies the nominal thermal expansion coefficient
(in ppm/°F) of the material on which the gage will exhibit optimum thermal output
characteristics, as shown in Figure 3.8. Table 3.3 lists a number of engineering materials,
and gives nominal values of the Fahrenheit and Celsius expansion coefficients for each,
along with the S-T-C number which would normally be selected for strain measurements
on that material. “062” indicates the active gage length in millimeters. “UT” represents
the grid and tab geometry. Therefore, 062UT is a small general purpose two-element 90º
tee rosette. The exposed solder tab area is 0.07 x 0.04 inch. The last number, 350, means
the resistance in ohms.
49
Material Description Expansion Coefficients ** Recommended S-T-C
Alumina , fired 3.0 ppm/° F (5.4 ppm/° C) 03
Aluminum Alloy , 2024-T4*, 7075-T6 12.9 ppm/° F (23.2 ppm/° C) 13*
Beryllium , pure 6.4 ppm/° F (11.5 ppm/° C) 06
Beryllium Copper , Cu 75, Be 25 9.3 ppm/° F (16.7 ppm/° C) 09
Brass, Cartridge , Cu 70, Zn 30 11.1 ppm/° F (20.0 ppm/° C) 13
Bronze, Phosphor , Cu 90, Sn 10 10.2 ppm/° F (18.4 ppm/° C) 09
Cast Iron, Gray 6.0 ppm/° F (10.8 ppm/° C) 06
Copper , pure 9.3 ppm/° F (16.7 ppm/° C) 09
Glass , Soda, Lime, Silica 5.1 ppm/° F (9.2 ppm/° C) 05
Inconel ,Ni-Cr- Fe alloy 7.0 ppm/° F (12.6 ppm/° C) 06
Inconel X Ni-Cr- Fe alloy 6.7 ppm/° F (12.1 ppm/° C) 06
Invar Fe-Ni alloy 0.8 ppm/° F (1.4 ppm/° C) 00
Magnesium Alloy *, AZ-31B 14.5 ppm/° F (26.1 ppm/° C) 15*
Molybdenum *, pure 2.2 ppm/° F (4.0 ppm/° C) 03*
Monel , Ni-Cu alloy 7.5 ppm/° F (13.5 ppm/° C) 06
Nickel-A ,Cu-Zn-Ni alloy 6.6 ppm/° F (11.9 ppm/° C) 06
Quartz , fused 0.3 ppm/° F (0.5 ppm/° C) 00
Steel, Alloy ,4340 6.3 ppm/° F (11.3 ppm/° C) 06
Steel, Carbon , 1008, 1018* 6.7 ppm/° F (12.1 ppm/° C) 06*
Steel, Stainless , Age Hardenable (17-4PH) 6.0 ppm/° F (10.8 ppm/° C) 06
Steel, Stainless , Age Hardenable (17-7PH) 5.7 ppm/° F (10.3 ppm/° C) 06
Steel, Stainless , Age Hardenable (PH15-7Mo) 5.0 ppm/° F (9.0 ppm/° C) 05
Steel, Stainless , Austenitic (304*) 9.6 ppm/° F (17.3 ppm/° C) 09*
Steel, Stainless , Austenitic (310) 8.0 ppm/° F (14.4 ppm/° C) 09
Steel, Stainless , Austenitic (316) 8.9 ppm/° F (16.0 ppm/° C) 09
Steel, Stainless , Ferritic (410) 5.5 ppm/° F (9.9 ppm/° C) 05
Tin , pure 13.0 ppm/° F(23.4 ppm/° C) 13
Titanium , pure 4.8 ppm/° F (8.6 ppm/° C) 05
Titanium Alloy , 6Al-4V* 4.9 ppm/° F (8.8 ppm/° C) 05*
Titanium Silicate *, polycrystalline 0.0 ppm/° F (0.0 ppm/° C) 00*
Tungsten , pure 2.4 ppm/° F (4.3 ppm/° C) 03
Zirconium ,pure 3.1 ppm/° F (5.6 ppm/° C) 03
50
Table 3.3. Thermal expansion coefficients of engineering materials
Gage Type CEA-06-062UT-350Resistance at 24°C 350.0±0.4%Gage factor at 24°C 2.105±0.5%Transverse sensitivity at 24°C 1.2±0.2%
Table 3.4. Strain gage specifications
The detailed characteristics of the strain gage are summarized in Table 3.4. The
gage factor is the measure of the sensitivity produced by a resistance strain gage. The
gage factor is determined through the calibration of the specific gage type and is the ratio
between 0R
R∆ and
L
L∆(strain), where R0 is the initial unstrained resistance of the gage.
StrainR
R
L
LR
R ∆=∆
∆=Factor Gage (3.1)
3.3.2 Strain gage installation and measurements
The metal foil type strain gage (CEA-06-062UT-350) was bonded to the
specimen (connecting rod) with an adhesive. The bonded resistance strain gage consists
of a strain-sensing element, a thin film that serves as an insulator and a carrier for the
strain-sensing element, and tabs for lead wire connections. M-Bond GA-61 was used for
the strain gage adhesive. M-Bond GA-61 has two components, partially filled, 100%
solids epoxy adhesive for general purpose stress analysis. This adhesive forms a hard,
chemically-resistant material when fully cured. Two strain gages were installed on the
connecting rod using M-Bond GA-61and cured about 3 hours an electric oven at 300 ºF.
After bonding and curing the strain gage in the connecting rod, the lead wire was
soldered to the exposed solder tab. In order to transfer the measured strain gage signal to
the data acquisition system, a flexible flat wire (ribbon cable) was used. Figure 3.9
shows the detailed wiring harness diagram.
51
Strain gage
Connector1
Connector2
Connector3
Cable
Figure 3.9. Schematic of strain gage installation.
The output side of connector 3 in Figure 3.9 is connected to a strain gage signal
conditioner and amplifier (Measurement Group Inc. 2100 system). The signal
conditioner and amplifier is shown in Figure 3.10 in detail.
Figure 3.10. 2100 series signal conditioner and amplifier.
52
The 2100 system consists of a 2120B two channel plug-in amplifier module that
includes bridge completions, bridge balance, amplifier balance, bridge excitation
regulator and shunt calibration, and a 2110B plug-in module capable of powering up to
ten channels at maximum rated voltage and current. The Wheatstone bridge excitation
voltage was set to 5 volts using the 2110B power supply. The strain gage amplification
factor was set to 500.
3.3.3 Strain gage calibration
The strain gage installed on the connecting rod should be calibrated to determine
the relation between the excited rod force and the strain gage output voltage. If we know
the exact material properties of the connecting rod, we can calculate the connecting rod
force using the measured strain gage output signal without a calibration process.
However, in most cases it is indispensable to perform the calibration process to find out
the elastic properties of the sample which, in the present case, is a connecting rod. A
computer controlled Instron 8500 servo-hydraulic test machine was used to perform the
tension test on the connecting rod. Figure 3.11 represents the overall test facility used for
strain gage calibration.
53
Figure 3.11. Tension test system for the connecting rod tests.
Figure 3.12 shows how the connecting rod was installed in the servo-hydraulic test
machine for the connecting rod tension tests. The tension test results are shown in Figure
3.13.
Figure 3.12. Connecting rod installed in the servo-hydraulic test machine.
54
0.0 0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.80
1000
2000
3000
4000
5000
6000
7000
8000
For
ce (
N)
Output Voltage (V)
Measured data point Linear fit
Strain Gage Calibration
Figure 3.13. Strain gage calibration test results.
3.3.4 Bending and temperature compensation
The bending and temperature variation of the connecting rod during the engine
experiments could affect the true connecting rod force measurements. In order to
calculate the true connecting rod force from the measured strain gage signal, the bending
and temperature effects of the strain gage should be eliminated or compensated for. The
bending effect of the connecting rod can be compensated for by using a full Wheatstone
bridge circuit such as that shown in Figure 3.14.
55
R3
R1
Ei
R4
R2
Eo
Figure 3.14. Wheatstone bridge circuit used in tests.
In Figure 3.14, R1 and R4 are the fixed resistors adopted for the full Wheatstone
bridge circuit and R2 and R3 are active strain gage resistances for the two strain gages
installed on the connecting rod. The output signal E0 of this strain gage circuit can be
expressed as in Equation 3.2.
iER
R
R
R
R
R
R
R
r
rE )1)((
)1( 4
4
3
3
2
2
1
120 η−∆−∆+∆−∆
+=∆ (3.2)
In Equation 3.2, 1R∆ and 3R∆ are zero because R1 and R3 are fixed resistors.
Then, Equation 3.2 becomes
iER
R
R
R
r
rE )1)((
)1( 4
4
2
220 η−∆+∆
+−=∆ (3.3)
The resistances of two fixed resistors, R1 and R3, are selected to be equal to the
unstrained resistances of the two strain gages, R2 and R4. The measured strain in the
connecting rod was less than 5%. Therefore, in Equation 3.3, r becomes 1 and η is zero.
Then, Equation 3.3 is reduced to
iER
R
R
RE )(
4
1
4
4
2
20
∆+∆−=∆ (3.4)
56
In Equation 3.4 RR /∆ can be expressed in terms of the gage factor and the strain via
εGRR =∆ / (3.5)
where
G: gage factor for the strain gages installed on the connecting rod
ε: strain in the connecting rod
The gage factor is the dimensionless proportionality factor between the relative
change of the resistance and the strain. The gage factor of the strain gage is determined
by sample measurements and is given on each package as the nominal value with its
tolerance. Using the gage factor, Equation 3.4 can be expressed as
ii EG
EGGE )(4
)(4
142420 εεεε +−=+−=∆ (3.6)
Usually, the strain in the connecting rod can be expressed as the sum of strains due to
longitudinal and bending stresses.
bl εεε += (3.7)
Thus,
222 bl εεε += and 444 bl εεε += (3.8)
In the case of the connecting rod,
lll εεε == 42 and 42 bb εε −= (3.9)
Using Equations 3.7, 3.8, and 3.9, Equation 3.6 becomes
El
GE ε
20
−=∆ (3.10)
57
Therefore it can be known from Equation 3.10 that the bending effect can be
compensated using the full Wheatstone bridge circuit by placing two strain gages on
opposite sides of the connecting rod such that they bend in opposite directions.
The measured strain gage signal could also have errors due to temperature effects
in addition to the bending effects. That is, the electrical resistance of the strain gage can
vary not only with the strain variation, but with a temperature variation as well.
Therefore, the errors caused by the temperature effect should be compensated for or
eliminated in order to obtain the true longitudinal stress in the connecting rod. Basically,
the thermal output of the strain gage is caused by two different mechanisms. One is from
the dependence of the electric resistances of the gage grid conductors on temperature
variation. The other is from differential thermal expansion between the grid conductor
and the test part. Two approaches were used to minimize errors due to temperature
effects. First, thermal errors can be eliminated using the temperature coefficient of the
gage factor and the balance of the Wheatstone bridge circuit. Since all experiments were
performed under steady state conditions, it is assumed that the connecting rod
temperature does not change much during the strain measurement. Thus, the gage factor
variation with temperature change can be compensated for using the strain gage sensor
data sheet. Figure 3.15 shows the gage factor variation with temperature for A-alloy and
D-alloy strain gages. For A-alloy, the gage factor is linearly dependent on the
temperature variation. Thus, if the gage temperature does not change much during the
measurement, the gage factor variation with temperature can be compensated for using
the measured gage temperature. In the present experiments, the Wheatstone bridge
circuit was balanced at experiment temperatures. In this case, the strain gage has no
voltage output due to the temperature effects. The second technique that was used to
minimize errors due to temperature effects was the choice of self-temperature
compensated strain gages. Through the use of self temperature compensated strain gages,
even if a small temperature variation occurs during the test, the errors from the
differential thermal expansion between the grid conductor and the test part can be
compensated for. However, if the connecting rod operating temperature changes
considerably during the strain measurement, the actual strain gage factors could be
different from the steady state temperature values. Thus, in order to avoid the possible
58
thermal errors in the strain measurements, the engine temperature should be maintained
constant during engine operation.
Figure 3.15. Gage factor variation with temperature for constantan (A-alloy) and isoelastic (D-alloy) strain gages.
3.4 Cylinder pressure measurement and data acquisition
A spark plug-mounted piezoelectric type pressure transducer (Kistler 6052) was
used to measure the cylinder pressure. The pressure transducer signal was input to a
charge amplifier (Kistler 5120). The charge amplifier output of the cylinder pressure
signal, strain gage signal, torque sensor output, and other engine output signals were
analyzed using a DSP Technology, Inc., combustion analyzer based on crank-angle-space.
An optical encoder (BEI model H25) was used to provide the crank angle (1440
signals/cycle) and TDC signals (two signals/cycle) to the combustion analyzer. Figure
3.16 illustrates the measurement system configuration for the cylinder pressure, torque
sensor, and strain gage signals. Cooling water and oil temperatures were maintained at
90°C during both the hot motoring tests and the firing tests.
59
2
13
1 14
3
4
11
865
7
12
Torque signal
Pressure signal
TDC signal
Crank angle signal
9
10
1) Encoder 8) AC motor2) Coupling 9) Combustion analyzer 3) Pressure transducer 10) Personal computer4) Charge amplifier 11) Water heater 5) Drive shaft 12) Oil heater6) Coupling 13) Strain gage conditioner7) Torque sensor 14) Strain gage
Figure 3.16. Measurement system configurations.
3.5. Summary of Measurement Systems Used for the Present Research
This chapter provided detailed information about the experimental systems that
were used to acquire the data that is discussed in Chapter 4. Early in this chapter, the
baseline engine and the prototype Rotating Liner Engine were discussed. This was
followed by a discussion of the two torque sensors that were used to measure the torque
at the crankshaft (engine input torque for motoring and engine output torque when firing)
and the torque input required to rotate the liner of the RLE. Most of this chapter was
devoted to a discussion of the system used to measure the force transmitted through the
connecting rod. Two strain gages were used for this purpose. They were located near the
60
CG of the connecting rod, but on opposite sides to eliminate potential erroneous signals
due to bending of the connecting rod. Erroneous signals can also be caused by
temperature effects via 1) differential thermal expansion between the gage grid
conductors and the connecting rod, and 2) the dependence of the electrical resistances of
the gage grid conductors on temperature. Two techniques were used to minimize or
eliminate these potential sources of error. First, self temperature-compensated strain
gages were used. This should be sufficient unless the connecting rod temperature varies
significantly during the experimental run. In this case, the gage factor may differ from its
nominal value. The second method used to minimize thermal errors in the strain gage
signal was to balance the Wheatstone bride circuit after the oil and coolant were warmed
to 90 oC and circulated through the engine. The engine was motored after the wheatstone
bridge circuit was balanced. In spite of the methods used to minimize thermal errors, this
remains an area of concern. The data acquisition system and the instrumentation used to
measure cylinder pressure and crank angle were discussed in the final part of this chapter.
61
Chapter 4. Test Results
4.1 Hot motoring tests
The hot motoring test is one of the methods for estimating engine frictional losses.
Although the motoring friction losses are not exactly the same as those for firing
operation, the hot motoring method is widely utilized to assess engine friction and is a
common bench-marking test due to its simplicity and the possibility of improving the
understanding of the sources of engine friction via tear-down tests. In the following
subsections, the engine motoring friction, especially piston ring assembly friction, of the
baseline engine and the RLE are measured using three different measurement techniques:
the direct motoring, the instantaneous IMEP method, and the P-w method. First, the
motoring friction is measured by direct motoring tests and, through the tear-down tests,
the friction of each engine component is measured. In the tear-down test, the engine
components are disassembled one by one and the motoring friction of each component is
measured. Through the direct motoring and the tear-down tests, the friction reduction
effects of the RLE are confirmed in the present experiments. Second, the instantaneous
IMEP method is applied to the baseline engine and the RLE under both motoring and
firing conditions. The difference of the lubrication mechanism between the baseline
engine and the RLE can be confirmed using the instantaneous IMEP method. Finally, the
P-w method is applied to both engines during the motoring tests in an attempt to
determine the friction differences of the engine components between the baseline engine
and the RLE.
62
4.1.1 Cycle-averaged friction torque and tear-down tests
The direct motoring and the tear-down tests are used to compare the friction
torque between the baseline engine and the RLE. In-line torque sensor signals for
measuring the motoring torque were logged and averaged during twenty engine cycles.
4.1.1(1) Baseline engine
The engine was motored from 1200 rpm to 2000 rpm in 200 rpm increments
during the hot motoring tests. Since the face seal of the RLE is designed based on 60
psig (~500 kPa absolute) of oil pressure, the oil pressure of the baseline engine should be
maintained at 60 psi to be the same as the rotating liner engine during the present hot
motoring tests. The 60 psi requirements set the lower end of the rpm test range since the
engine oil pressure falls below 60 psi at speeds lower than 1200 rpm. During the hot
motoring tests, the oil and water temperatures were maintained at 90 °C. The engine
coolant and oil were heated and circulated before starting the hot motoring tests. During
the hot motoring tests, the cylinder pressure and torque signals were measured for 20
consecutive engine cycles. Figure 4.1 shows the measured cycle-averaged motoring
torque of the baseline engine during the hot motoring test. The measured friction torque
changed from 6.31 lb-ft (8.55 N-m) at 1200 rpm to 7.09 lb-ft (9.61 N-m) at 2000 rpm.
The measured friction torque can be converted to friction mean effective pressure using
the Equation 4.1.
dV
TR
nkPaFMEP
28.6)( = (4.1)
where:
nR: number of crank revolutions for each power stroke per cylinder
T: measured friction torque (N-m)
Vd: engine displaced volume (dm3)
63
1 2 0 0 1 4 0 0 1 6 0 0 1 8 0 0 2 0 0 04
5
6
7
8
Fric
tion
torq
ue (
lbft)
E n g in e S p e e d ( rp m )
1 2 0 0 1 4 0 0 1 6 0 0 1 8 0 0 2 0 0 01 0 0
1 5 0
2 0 0
2 5 0
3 0 0
FM
EP
(kP
a)
E n g in e S p e e d ( rp m )
Figure 4.1. Baseline engine hot motoring torque.
Since the baseline engine is a single cylinder version converted from four
cylinders, the FMEP is much higher than that of a normal mass produced engine. The
reason is because, although the engine swept volume used in Equation 4.1 is for a single
cylinder, the mechanical friction losses (such as main bearings, water pump, oil pump,
etc.) except for the piston assembly are the same as the original four-cylinder engine.
For the tear-down tests, after the oil and coolant are heated to normal operating
temperature (90 oC) and circulated through the engine, the engine is motored at a selected
speed and the friction torque is measured with the throttle wide open. This, of course, is
simply the normal hot motoring method. The following steps differentiate the tear-down
test from normal hot motoring, and allow more information to be extracted from
relatively simple tests. For the tear-down tests, the engine is progressively disassembled
and hot motoring tests are performed after each stage of disassembly. First, the piston
and connecting rod are removed so that the difference in friction torque between this test
and the prior test with all components must be due to the friction in the piston assembly.
Next, the camshaft timing belt is removed. The difference in friction torque between this
test and the prior test with only the piston assembly removed must be due to the friction
in the valvetrain (chain, camshaft journal bearings, and components in the valve motion
mechanism) and water pump, which is driven by this same chain on the engine used for
64
these experiments. Furthermore, the friction measured during this test is only that from
the crankshaft’s journal bearings and oil pump, which is gear driven off the crankshaft.
Figure 4.2 presents the tear-down test results for the baseline engine. In the
teardown tests the friction losses of the water pump, oil pump, and other mechanical parts
are accounted for. The friction of the water pump and the timing chain are included in
the camshaft assembly loss. The oil pump friction is included in the crankshaft assembly
loss. In the upper graph in Figure 4.2, the piston assembly friction is responsible for
about 35% of the total motoring friction losses. The portion of total friction that is due to
piston assembly friction losses in this engine is relatively small compared with normal
gasoline engines because only one piston is installed in this 4 cylinder engine. In Figure
4.2, the red column (horizontal pattern) represents the piston assembly friction losses, the
yellow column (cross pattern) is for crankshaft friction losses, and the green column
(hatch pattern) is for camshaft friction losses. As shown in Figure 4.2, the crankshaft
assembly friction loss increases with increasing engine speed. This means that the
dominant friction mechanism of the crankshaft assembly is in the hydrodynamic region.
In the case of the camshaft assembly, friction losses also increase, as the engine speed
increases, but not as strongly as was the case for the crankshaft assembly. This can be
explained as the main lubrication region of the camshaft assembly is boundary or mixed
lubrication. The friction losses of the piston assembly decrease with increasing speed in
the low speed regions but increase in the high speed regions,. That is, as expected, at low
engine speeds the piston assembly friction shows that the boundary and mixed lubrication
characteristics are dominant. However, as the engine speed becomes higher, the
dominant lubrication mechanism is converted to hydrodynamic.
65
1000 1200 1400 1600 1800 2000 22000
1
2
3
4
5
6
7
8
9
10
Tor
que
(lbft)
Engine speed (rpm)
Baseline assembly Crankshaft+Camshaft Crankshaft Piston assembly
1000 1200 1400 1600 1800 2000 22000
1
2
3
4
Tor
que
(lbft)
Engine Speed (rpm)
Piston Assembly Camshaft assembly Crankshaft assembly
Figure 4.2. Teardown test results for the baseline engine.
66
4.1.1(2) Rotating Liner Engine
Figure 4.3 shows the measured motoring torque of the RLE. Figure 4.4 indicates
the hot motoring friction torque and friction reduction rate of the Rotating Liner Engine
compared with the baseline engine. From Figure 4.4 it can be said that through the use of
liner rotation the total mechanical friction is diminished by 23% ~ 31% compared with
the baseline engine. The hot motoring torque of the RLE is lower than that for the
baseline engine by 31% at 1200 rpm and somewhat less at the higher speeds. At lower
speeds, the RLE has more pronounced friction reduction effects. This friction reduction
of the RLE during the hot motoring test is believed to be due to the reduction of piston
assembly friction through liner rotation. However, since the motoring torque in Figures
4.3 and 4.4 includes the total engine friction, such as piston assembly, camshaft, and
crankshaft friction, additional tests were performed to identify the exact source of the
decreased friction.
1 2 0 0 1 4 0 0 1 6 0 0 1 8 0 0 2 0 0 03
4
5
6
7
Fric
tion
torq
ue (
lbft)
E n g in e S p e e d ( rp m )
1 2 0 0 1 4 0 0 1 6 0 0 1 8 0 0 2 0 0 05 0
1 0 0
1 5 0
2 0 0
2 5 0
FM
EP
(kP
a)
E n g in e S p e e d ( rp m )
Figure 4.3. Rotating Liner engine hot motoring torque.
67
1000 1200 1400 1600 1800 2000 22000
1
2
3
4
5
6
7
8
9
10 Baseline engine RLE
Tor
que
(lbft)
Engine Speed (rpm)
1200 1400 1600 1800 20000
10
20
30
Fric
tion
redu
ctio
n (%
)
Engine speed (rpm)
Figure 4.4. Total hot motoring torque and friction reduction through liner rotation.
68
Figure 4.5 presents the tear-down test results of the Rotating Liner Engine. In Figure 4.5,
the friction losses of the crankshaft and camshaft assemblies are not much different from
those of the baseline engine. However, piston assembly friction is totally different from
that of the baseline engine. That is, as the engine speed increases, the piston assembly
friction is also monotonically increasing. This means that the dominating friction
mechanism of the piston assembly of the Rotating Liner Engine is converted from
boundary and the mixed lubrication regimes for the baseline engine to the hydrodynamic
lubrication regime for the RLE.
1000 1200 1400 1600 1800 2000 22000
1
2
3
4
5
6
7
8
9
10
Tor
que
(lbft)
Engine Speed (rpm)
RLE assembly Crankshaft+camshaft Crankshaft assembly Piston assembly
1000 1200 1400 1600 1800 2000 22000
1
2
3
4
Tor
que
(lbft)
Engine Speed (rpm)
Piston Assembly Camshaft Assembly Crankshaft Assembly
Figure 4.5. Teardown test results for the Rotating Liner Engine.
69
Figure 4.6 clearly shows the piston assembly friction reduction effects of the
Rotating Liner Engine.
0
20
40
60
80
100
1200 1400 1600 1800 20000
1
2
3
4
5
Pis
ton
Ass
embl
y F
rictio
n T
orqu
e (f
t-lb
)
Engine Speed (rpm)
Baseline Engine RLE
RLE
Pis
ton
Ass
embl
y F
rictio
n re
duct
ion
(%
)
Figure 4.6. Piston assembly friction torque of the baseline engine and the RLE.
Piston assembly friction of the RLE is reduced by 70% to 90% compared with
that of the baseline engine over the entire test speed range. Figure 4.6 highlights the
piston assembly friction reduction effect more clearly. At 1200 rpm, almost 90% of the
piston assembly friction is reduced by liner rotation. As the engine speed increases, the
friction reduction decreases, eventually producing a 70% benefit at 2000 rpm. This
occurs because more of the stroke experiences hydrodynamic lubrication as the engine
speed increases.
70
4.1.2 Friction force measurement using the instantaneous IMEP method
Through the direct hot motoring torque tests of the baseline engine and the RLE,
it can be concluded that the liner rotation is effective in reducing the piston assembly
friction. Especially from the tear-down test, it was confirmed that the dominant friction
mechanism of the piston assembly was changed from boundary and mixed lubrication to
hydrodynamic lubrication. In this subsection, the piston assembly friction of the baseline
engine and the RLE are measured using the instantaneous IMEP method. That is, under
motoring conditions, the friction reduction of the RLE will be explored using the
instantaneous IMEP method. In the instantaneous IMEP method, the piston assembly
friction force is computed from the measured cylinder pressure, connecting rod force and
crankshaft rotational speed.
4.1.2(1) Baseline engine
4.1.2(1-1) Cold motoring friction (oil temperature: 20°°°°C)
Figure 4.7 shows the measured cylinder pressures and the connecting rod forces
of the baseline engine at an oil temperature of 20 °C. The pressure forces are calculated
based on the measured cylinder pressure and cylinder bore area. The connecting rod
force is measured from the strain gages installed on the connecting rod and indicates the
force transferred through the connecting rod during engine operation. The difference
between the cylinder pressure force and the connecting rod force is the sum of the friction
force and the inertial force of the piston and the connecting rod assembly. That is,
basically the connecting rod force has information about the piston assembly friction and
the inertia forces. In Figure 4.7, the connecting rod force shows almost the same shape as
the pressure force at low engine speeds since the inertia force is not large at low rpm.
However, as the engine speed increases, the connecting rod forces looks different from
the cylinder pressure forces as the inertia forces become bigger and bigger. Figure 4.8
shows the characteristics of the pressure force and the connecting rod force at different
engine speeds more clearly. In Figure 4.8, the pressure forces are increasing as the
71
engine speed increases. This is mainly because of the valve timing and the intake system
tuning. Since the intake valve closing timing is set to maximize the use of the intake flow
inertia at a specific engine speed, some of the intake air is flowing back to the intake
manifold at low engine speed. The volumetric efficiency also increases as the engine
speed increases due to the intake flow inertia. The variation of connecting rod force can
clearly be confirmed in Figure (4.8). As the engine speed goes up, the inertia forces of
piston and connecting rod have more influences on connecting rod forces.
-180 0 180 360 540-2000
0
2000
4000
6000
8000
10000
12000
14000
Pre
ssur
e fo
rce
(N)
Crank angle (deg)
500 rpm
-180 0 180 360 540-2000
0
2000
4000
6000
8000
10000
12000
14000
Con
_rod
forc
e (N
)
Crank angle (deg)
500 rpm
-180 0 180 360 540-2000
0
2000
4000
6000
8000
10000
12000
14000
Pre
ssur
e fo
rce
(N)
Crank angle (deg)
800 rpm
-180 0 180 360 540-2000
0
2000
4000
6000
8000
10000
12000
14000
Con
_rod
forc
e (N
)
Crank angle (deg)
800 rpm
Figure 4.7a. Measured pressure force and connecting rod force for an oil temperature of20 °C at 500 rpm and 800 rpm.
72
-180 0 180 360 540-2000
0
2000
4000
6000
8000
10000
12000
14000P
ress
ure
forc
e (N
)
Crank angle (deg)
1200 rpm
-180 0 180 360 540-2000
0
2000
4000
6000
8000
10000
12000
14000
Con
_rod
forc
e (N
)
Crank angle (deg)
1200 rpm
-180 0 180 360 540-2000
0
2000
4000
6000
8000
10000
12000
14000
Pre
ssur
e fo
rce
(N)
Crank angle (deg)
1600 rpm
-180 0 180 360 540-2000
0
2000
4000
6000
8000
10000
12000
14000
Con
_rod
forc
e (N
)
Crank angle (deg)
1600 rpm
-1 8 0 0 1 8 0 3 6 0 5 4 0-2 0 0 0
0
2 0 0 0
4 0 0 0
6 0 0 0
8 0 0 0
1 0 0 0 0
1 2 0 0 0
1 4 0 0 0
Pre
ssur
e fo
rce
(N)
C ra n k a n g le (d e g )
2 0 0 0 rp m
-1 8 0 0 1 8 0 3 6 0 5 4 0-2 0 0 0
0
2 0 0 0
4 0 0 0
6 0 0 0
8 0 0 0
1 0 0 0 0
1 2 0 0 0
1 4 0 0 0
Con
_rod
forc
e (N
)
C ra n k a n g le (d e g )
2 0 0 0 rp m
Figure 4.7b. Measured pressure force and connecting rod force for an oil temperature of20 °C at 1200 rpm, 1600 rpm and 2000 rpm.
73
-180 0 180 360 540-2000
0
2000
4000
6000
8000
10000
12000
14000
Con
n_ro
d fo
rce
(N)
Crank angle (deg)
-180 0 180 360 540-2000
0
2000
4000
6000
8000
10000
12000
14000
500 rpm 800 rpm1200 rpm1600 rpm2000 rpm
500 rpm 800 rpm1200 rpm1600 rpm2000 rpm
Pre
ssur
e fo
rce
(N)
Crank angle (deg)
Figure 4.8. Effects of engine speed on pressure and connecting rod force variationsthroughout the cycle at an oil temperature of 20°C.
The inertial force of the piston assembly should be calculated from the measured
crankshaft rotational speed in order to determine the friction force of the piston assembly.
Figure 4.9 shows the engine rotational speed, the linear speed, and the acceleration of the
crankshaft and the piston assembly at different engine speeds. In Figure 4.9, the
rotational speed of the crankshaft is measured at the front end of the crankshaft using an
optical encoder. The angular acceleration, linear speed and linear acceleration of the
piston and connecting rod are computed based on the measured rotational speed using
Equations 2.29 and 2.30. The angular acceleration in Figure 4.9 is computed using an
equation such as
φωω
φφωωωω
φωωφ
φωωα ∆
−=
−−
+≈===
22
21
22
12
1212
d
d
dt
d
d
d
dt
d(4.2)
2
2
dt
d φα =
dt
dφω =
74
-180 0 180 360 54052
54
56
58
60
62
Rot
atio
nal s
peed
(ra
d/s)
Crank angle (deg)
-180 0 180 360 540-1000
-500
0
500
1000
Rot
atio
nal a
ccel
erat
ion
(rad
/s^2
)
Crank angle (deg)
-180 0 180 360 540-10
-5
0
5
10
Line
ar s
peed
(m
/s)
Crank angle (deg)
-180 0 180 360 540-1000
-500
0
500
1000
Line
ar a
ccel
erat
ion
(m/s
^2)
Crank angle (deg)
500 rpm
-180 0 180 360 54080
84
88
92
96
100
Rot
atio
nal s
peed
(ra
d/s)
Crank angle (deg)
-180 0 180 360 540-2000
-1500
-1000
-500
0
500
1000
1500
2000
Rot
atio
nal a
ccel
erat
ion
(rad
/s^2
)
Crank angle (deg)
-180 0 180 360 540-10
-5
0
5
10
Line
ar s
peed
(m
/s)
Crank angle (deg)
-180 0 180 360 540-1000
-500
0
500
1000
Line
ar a
ccel
erat
ion
(m/s
^2)
Crank angle (deg)
800 rpm
Figure 4.9a. Measured angular speed and the calculated angular acceleration, linear speed, and linear acceleration at 500 rpm and 800 rpm
75
-180 0 180 360 540125
130
135
140
145
150
Rot
atio
nal s
peed
(ra
d/s)
Crank angle (deg)
-180 0 180 360 540-3000
-2000
-1000
0
1000
2000
3000
Rot
atio
nal a
ccel
erat
ion
(rad
/s^2
)
Crank angle (deg)
-180 0 180 360 540-20
-15
-10
-5
0
5
10
15
20
Line
ar s
peed
(m
/s)
Crank angle (deg)
-180 0 180 360 540-2000
-1500
-1000
-500
0
500
1000
1500
2000
Line
ar a
ccel
erat
ion
(m/s
^2)
Crank angle (deg)
1200 rpm
-180 0 180 360 540150
155
160
165
170
175
180
Rot
atio
nal s
peed
(ra
d/s)
Crank angle (deg)
-180 0 180 360 540-3000
-2000
-1000
0
1000
2000
3000
Rot
atio
nal a
ccel
erat
ion
(rad
/s^2
)
Crank angle (deg)
-180 0 180 360 540-20
-15
-10
-5
0
5
10
15
20
Line
ar s
peed
(m
/s)
Crank angle (deg)
-180 0 180 360 540-2000
-1500
-1000
-500
0
500
1000
1500
2000
Line
ar a
ccel
erat
ion
(m/s
^2)
Crank angle (deg)
1600 rpm
Figure 4.9b. Measured angular speed and the calculated angular acceleration, linear speed, and linear acceleration at 1200 rpm and 1600 rpm
76
-180 0 180 360 540180
190
200
210
220
230
Rot
atio
nal s
peed
(ra
d/s)
Crank angle (deg)
-180 0 180 360 540-6000
-4000
-2000
0
2000
4000
6000
Rot
atio
nal a
ccel
erat
ion
(rad
/s^2
)
Crank angle (deg)
-180 0 180 360 540-20
-15
-10
-5
0
5
10
15
20
Line
ar s
peed
(m
/s)
Crank angle (deg)
-180 0 180 360 540-4000
-3000
-2000
-1000
0
1000
2000
3000
4000
Line
ar a
ccel
erat
ion
(m/s
^2)
Crank angle (deg)
2000 rpm
Figure 4.9c. Measured angular speed and the calculated angular acceleration, linear speed, and linear acceleration at 2000 rpm
The inertial forces of the piston assembly and the connecting rod are computed
using the angular and linear speed and acceleration. In the case of the piston assembly,
the motion is basically linear. Thus, the inertia forces of the piston can be expressed by
the linear acceleration. The connecting rod experiences both linear acceleration and
rotational acceleration simultaneously. Equation 2.33 presents the acceleration of the
connecting rod during engine operation. The inertial force of the connecting rod in the
direction of the connecting rod axis can be expressed by the following integration:
∫−= rr
adrrAnF n2
)(ρ (4.3)
In Equation 4.3, ρ and Ar are the density and the cross-sectional area of the connecting
rod.
77
The detailed notation for calculating the connecting rod inertia force is shown in
Figure 4.10.
Strain gage
t
an
a
r3 r
dr
r1
x
r2
Figure 4.10. Strain gage location, nomenclature used in the equation set, and accelerations.
In Figure 4.10, the unit mass dr has two components of acceleration: ta and na .
Therefore, using Equation 4.3, the inertial force of the connecting can be calculated. The
integration in Equation 4.3 requires a numerical approach to obtain an accurate solution.
However, with a negligible error the integration can be simplified under the assumption
of uniform cross-sectional area and density of the connecting rod. The connecting rod
mass from r1 to r2 can be included in the piston mass due to its translational motion. In
this research, the connecting rod was 3-dimensionally modeled using SOLIDWORKS.
Figure 4.11 shows the modeled connecting rod. Figures 4.12 and 4.13 present the
calculated inertial forces of the connecting rod and the piston assembly for a range of
engine speeds.
78
Figure 4.11. Modeled connecting rod using SOLIDWORKS.
- 1 8 0 0 1 8 0 3 6 0 5 4 0- 2 0 0- 1 5 0- 1 0 0
- 5 00
5 01 0 01 5 02 0 0
C r a n k a n g le ( d e g )
Tot
al in
ertia
forc
e (N
)
- 3 0
- 2 0
- 1 0
0
1 0
2 0
3 0
Con
_rod
iner
tia fo
rce
(N)
- 2 0 0- 1 5 0- 1 0 0
- 5 00
5 01 0 01 5 02 0 0
Pis
ton
iner
tia fo
rce
(N) 5 0 0 r p m
Figure 4.12a. Inertial forces of the piston and the connecting rod at 500 rpm
79
- 1 8 0 0 1 8 0 3 6 0 5 4 0- 5 0 0- 4 0 0- 3 0 0- 2 0 0- 1 0 0
01 0 02 0 03 0 04 0 05 0 0
C r a n k a n g l e ( d e g )
Tot
al in
ertia
forc
e (N
)
- 1 0 0- 8 0- 6 0- 4 0- 2 0
02 04 06 08 0
1 0 0C
on_r
od in
ertia
forc
e (N
)- 5 0 0- 4 0 0- 3 0 0- 2 0 0- 1 0 0
01 0 02 0 03 0 04 0 05 0 0
Pis
ton
iner
tia fo
rce
(N) 8 0 0 r p m
- 1 8 0 0 1 8 0 3 6 0 5 4 0- 1 0 0 0
- 5 0 0
0
5 0 0
1 0 0 0
C r a n k a n g l e ( d e g )
Tot
al in
ertia
forc
e (N
)
- 2 0 0- 1 5 0- 1 0 0
- 5 00
5 01 0 01 5 0
2 0 0
Con
_rod
iner
tia fo
rce
(N)
- 1 0 0 0
- 5 0 0
0
5 0 0
1 0 0 0
Pis
ton
iner
tia fo
rce
(N) 1 2 0 0 r p m
Figure 4.12b. Inertial forces of the piston and the connecting rod at 800 and 1200 rpm
80
- 1 8 0 0 1 8 0 3 6 0 5 4 0- 2 0 0 0- 1 5 0 0- 1 0 0 0
- 5 0 00
5 0 01 0 0 01 5 0 02 0 0 0
C r a n k a n g le ( d e g )
Tot
al in
ertia
forc
e (N
)
- 2 0 0- 1 5 0- 1 0 0
- 5 00
5 01 0 01 5 02 0 0
Con
_rod
iner
tia fo
rce
(N)
- 2 0 0 0- 1 5 0 0- 1 0 0 0
- 5 0 00
5 0 01 0 0 01 5 0 02 0 0 0
Pis
ton
iner
tia fo
rce
(N) 1 6 0 0 r p m
- 1 8 0 0 1 8 0 3 6 0 5 4 0- 2 5 0 0- 2 0 0 0- 1 5 0 0- 1 0 0 0
- 5 0 00
5 0 01 0 0 01 5 0 02 0 0 0
C r a n k a n g le ( d e g )
Tot
al in
ertia
forc
e (N
)
- 3 0 0
- 2 0 0
- 1 0 0
0
1 0 0
2 0 0
3 0 0
Con
_rod
iner
tia fo
rce
(N)
- 2 0 0 0- 1 5 0 0- 1 0 0 0
- 5 0 00
5 0 01 0 0 01 5 0 02 0 0 0
Pis
ton
iner
tia fo
rce
(N) 2 0 0 0 r p m
Figure 4.12c. Inertial forces of the piston and the connecting rod at 1600 and 2000 rpm
81
-180 0 180 360 540-2000
-1750
-1500
-1250
-1000
-750
-500
-250
0
250
500
750
1000
1250
1500
500 rpm 800 rpm1200 rpm1600 rpm2000 rpm
Iner
tia fo
rce
(N)
Crank angle (deg)
Figure 4.13. Effects of engine speed on the variation of the total inertia force throughout the cycle for motoring conditions.
Figure 4.13 presents the total inertia force throughout the cycle for several engine
speeds. As expected, the inertia force is has a larger magnitude at higher engine speeds.
The resulting friction force of the piston assembly can be computed using the information
of the pressure force, the connecting rod force, and the inertial forces of the connecting
rod and piston assembly. For cold motoring, the oil viscosity is much higher than for hot
motoring. Figure 4.14 indicates the effect of temperature on the oil viscosity for 5W-30
engine oil. As shown in Figure 4.14, at 20°C oil temperature, the viscosity is ~50%
higher than for fully warmed up conditions (90 °C oil temperature). Thus, with cold oil
the duty parameter (viscosity*speed/load) in the Stribeck curve becomes high and there is
more possibility for hydrodynamic lubrication than for boundary or mixed lubrication in
the piston assembly friction.
The experimental results for the piston assembly friction for cold motoring are
shown in Figures 4.15 and 4.16. As expected, the friction forces near TDC and BDC are
smaller than near mid piston stroke. That is, due to the high viscosity of the cold engine
oil, the boundary friction region has become minimized and the friction from the
hydrodynamic lubrication is maximized at the mid stroke.
82
200 300 400 500 600 700-0.1
0.0
0.1
0.2
0.3
Oil
visc
osity
(P
a se
c)
Oil temperature (K)
Figure 4.14. Effect of oil temperature on oil viscosity.
As the engine speed increases, the hydrodynamic friction region of the stroke has
become wider than at lower speeds. Figure 4.16 clearly shows the characteristics of
piston ring assembly friction at low oil temperature as the engine speed increases. In
Figure 4.15, the friction force at low engine speeds, such as 500 and 800 rpm, shows a
little friction peak near the compression TDC position. However, as the piston speed
increases over 800 rpm, the friction peak near compression TDC completely disappears
and the hydrodynamic lubrication is more dominant. In the hydrodynamic lubrication
region, since the friction coefficient is increasing dependent on the piston speed, at the
mid stroke the piston friction force is maximized, as shown in Figure 4.16.
83
-180 0 180 360 540-1000-750-500-250
0250500750
1000
Crank angle (deg)
Fric
tion
forc
e 2000 rpm
-1000-750-500-250
0250500750
1000
Fric
tion
forc
e (N
) 1600 rpm
-1000-750-500-250
0250500750
1000
Fric
tion
forc
e (N
)
1200 rpm
-1000-750-500-250
0250500750
1000
Fric
tion
forc
e (N
)
800 rpm
-1000-750-500-250
0250500750
1000
Fric
tion
forc
e (N
) 500 rpm
Figure 4.15. Friction force of the piston assembly at an oil temperature of 20 °C.
84
-180 0 180 360 540-1000
-750
-500
-250
0
250
500
750
1000
Pis
ton
Ass
embl
y F
rictio
n F
orce
(N
)
Crank angle (deg)
500 rpm 800 rpm 1200 rpm 1600 rpm 2000 rpm
Figure 4.16. Effect of engine speed on the variation of the piston assembly friction force throughout the cycle at an oil temperature of 20 °C.
4.1.2(1-2) Hot motoring friction (oil temperature: 90°°°°C)
The oil viscosity of 5W-30 at 90 °C is just 5% of that at 20°C. The small
viscosity yields an increased possibility for boundary and mixed lubrication between the
piston rings and the liner. At elevated oil temperature, the oil film thickness between the
piston assembly and the liner could be less than that of metal asperities at some piston
positions, and thus there could be more possibilities of metal-to-metal contact between
them during the engine operation. That is, the lubrication regime could be boundary and
mixed lubrication near the TDC and BDC positions in which the cylinder pressure is high
and the piston speed is low. Figures 4.17 and 4.18 show the measured cylinder pressure
and connecting rod forces at an oil temperature of 90 °C at various engine speeds. The
trend of pressure force variation as the engine speed increases is the same as that for low
oil temperature. However, the absolute value of cylinder pressure at high oil temperature
is less than that at low oil temperature due to the difference in volumetric efficiency.
Since the air densities trapped in the cylinder become lower and the volumetric
efficiencies are lower at high oil temperature, the cylinder pressure at high oil
temperature is lower than that for low oil temperature at the same engine speed.
85
-180 0 180 360 540-2000
0
2000
4000
6000
8000
10000
12000
14000
800 rpm
500 rpm
800 rpm
Pre
ssur
e fo
rce
(N)
Crank angle (deg)
500 rpm
-180 0 180 360 540-2000
0
2000
4000
6000
8000
10000
12000
14000
Con
_rod
forc
e (N
)
Crank angle (deg)
-180 0 180 360 540-2000
0
2000
4000
6000
8000
10000
12000
14000
Pre
ssur
e fo
rce
(N)
Crank angle (deg)
-180 0 180 360 540-2000
0
2000
4000
6000
8000
10000
12000
14000
Con
_rod
forc
e (N
)
Crank angle (deg)
-180 0 180 360 540-2000
0
2000
4000
6000
8000
10000
12000
14000
1600 rpm
1200 rpm
1600 rpm
Pre
ssur
e fo
rce
(N)
Crank angle (deg)
1200 rpm
-180 0 180 360 540-2000
0
2000
4000
6000
8000
10000
12000
14000
Con
_rod
forc
e (N
)
Crank angle (deg)
-180 0 180 360 540-2000
0
2000
4000
6000
8000
10000
12000
14000
Pre
ssur
e fo
rce
(N)
Crank angle (deg)
-180 0 180 360 540-2000
0
2000
4000
6000
8000
10000
12000
14000
Con
_rod
forc
e (N
)
Crank angle (deg)
Figure 4.17. Measured pressure force and connecting rod force (90 oC oil temperature).
86
-180 0 180 360 540-2000
0
2000
4000
6000
8000
10000
12000
14000
Con
n_ro
d fo
rce
(N)
Crank angle (deg)
-180 0 180 360 540-2000
0
2000
4000
6000
8000
10000
12000
14000
500 rpm 800 rpm1200 rpm1600 rpm2000 rpm
500 rpm 800 rpm1200 rpm1600 rpm2000 rpm
Pre
ssur
e fo
rce
(N)
Crank angle (deg)
Figure 4.18. Effects of engine speed on the variations of the pressure and connecting rod forces throughout the cycle at an oil temperature of 90 °C.
The friction forces are calculated from the measured pressure force, connecting
rod force and the inertial forces. The calculated friction forces for the hot motoring
condition are shown in Figures 4.19 and 4.20. Figure 4.19 indicates the overall trend of
the piston ring friction forces according to the change of engine speed. As the engine
speed increases, the friction force right after compression TDC decreases. However, in
contrast to cold motoring, the friction force near the mid-stroke does not change a lot.
This is basically due to the low oil viscosity at high oil temperature. At high speed, the
measured friction forces are more influenced by the increased signal noise that occurs at
higher speeds. Figure 4.20 shows the effect of engine speed on the hot motoring friction
force variation more directly. As the engine speed increases, the friction forces decrease
slightly near compression TDC and increase near mid piston stroke even if the
differences are small. However, in contrast to expectations, the friction force peak near
compression TDC at 800 rpm is greater than that for 500 rpm.
87
-180 0 180 360 540-1000-750-500-250
0250500750
1000
Crank angle (deg)
Fric
tion
forc
e (N
)
2000 rpm
-1000-750-500-250
0250500750
1000
Fric
tion
forc
e (N
)
1600 rpm
-1000-750-500-250
0250500750
1000
Fric
tion
forc
e (N
)
1200 rpm
-1000-750-500-250
0250500750
1000
Fric
tion
forc
e (N
)
800 rpm
-1000-750-500-250
0250500750
1000
Fric
tion
forc
e (N
)500 rpm
Figure 4.19. Friction force of the piston assembly at an oil temperature of 90 °C.
88
-180 0 180 360 540-500
-400
-300
-200
-100
0
100
200
300
400
500 500 rpm 800 rpm1200 rpm
Fric
tion
forc
e (N
)
Crank angle (deg)
Figure 4.20. Effect of engine speed on the friction force throughout the cycle with an oil temperature of 90 °C.
4.1.2(2) Rotating Liner Engine
The instantaneous IMEP method was applied to the Rotating Liner Engine in
order to find out the difference of the friction mechanism between the baseline engine
and the Rotating Liner Engine. The oil temperature was set to 90°C for the instantaneous
IMEP method.
-180 0 180 360 540-500
-400
-300
-200
-100
0
100
200
300
400
500
Fric
tion
forc
e (N
)
Crank angle (deg)
Baseline engine RLE
Figure 4.21. Friction force comparison between the baseline engine and the RLE at 1200 rpm.
89
Figure 4.21 shows the friction force differences between the baseline engine and
the RLE at 1200 rpm. As expected, the piston assembly friction force of the RLE before
and after TDC is much less than that of the baseline engine. In addition to the friction
reduction at TDC, the RLE friction force near the mid-stroke of expansion is also smaller
than that for the baseline engine. This means that the effect of liner rotation is also
effective to reduce the friction in the hydrodynamic lubrication region. From Figure
(4.20) and (4.21), the signs of the friction forces after compression TDC are not
reasonable. During the expansion at motoring tests, the measured friction forces show the
positive values. This sign changes during the expansion process is not physically possible.
In chapter 5, the errors of the measured friction force using the instantaneous IMEP
method will be discussed.
90
4.1.3 Friction torque measurement using the P-w method
In order to analyze the friction reduction effect of the RLE, it is necessary to
examine the instantaneous measured motoring torque signal. Figure 4.23 presents the
measured torque signal at 1200 rpm, 1600 rpm and 2000 rpm. The dashed line is for the
baseline engine and the thick line is for the RLE.
-180 0 180 360 540-150
-100
-50
0
50
100
150
Crank Angle (deg)
Tor
que
(Nm
)
-150
-100
-50
0
50
100
150
Tor
que
(Nm
)
-150
-100
-50
0
50
100
150
1600 rpm
2000 rpm
Tor
que
(Nm
)
1200 rpm
Figure 4.23. Measured instantaneous motoring torque of the baseline engine and the RLE as obtained using the p-w method.
91
The instantaneous motoring of Figure 4.23 was measured using an in-line torque
sensor (Cooper. LXT963). In Figure 4.23 crank angle 0° indicates the compression TDC
position during the engine cycle. Although the averaged motoring torque for the RLE is
less than that of the baseline engine, the instantaneous motoring torque of the RLE is
higher in some crank angle regions than for the baseline engine and lower in other
regions. Therefore, it is not easy to analyze the friction reduction mechanism of the RLE
just using measured instantaneous motoring torque. Although the cycle-averaged values
of inertia and pressure torque are zero during motoring, the motoring torque at each crank
angle is the resultant torque generated by inertia, cylinder pressure, and friction. That is,
the inertia torque, the pressure torque, and the friction torque influence the instantaneous
motoring torque at each crank angle.
fpiL TTTT ++= (4.4)
where
orquefriction t:T
torquepressure:T
torqueinertia:T
torquemotoringmeasured:T
f
p
i
L
Therefore, the inertia and pressure torque must be calculated to allow the friction torque
information to be extracted from the measured motoring torque at each crank angle. At
1200 rpm, the cylinder pressure torque is shown in Figure 4.24. From Figure 4.24 it can
be said that the cylinder pressures of the baseline engine and the RLE are almost the same
during hot motoring tests. The next task is to calculate the inertia torque. The inertia
torque is generated by rotating motion of the crank assembly and by reciprocating motion
of the piston assembly. If the instantaneous angular speed of the crankshaft is maintained
as constant during the engine cycle, then the rotational inertia force of the crankshaft
assembly is zero. However, although the engine speed attains a steady state operating
condition and becomes a constant average speed, the instantaneous rotational speed is not
constant. As the piston compresses trapped air, the instantaneous engine speed decreases,
and after compression the instantaneous engine speed increases during the expansion
92
stroke. These effects can be minimized via a large rotational mass of the flywheel and
dyno.
93
-180 0 180 360 540-200-150-100-50
050
100150200
Crank Angle (deg)
Tor
que
(N-m
)
-20000
2000400060008000
100001200014000
For
ce (
N)
-200-150-100-50
050
100150200
Tor
que
(N-m
)
-20000
2000400060008000
100001200014000
RLE
RLE
Baseline
For
ce (
N)
Baseline
Figure 4.24. Pressure torque at 1200 rpm.
94
That is, as the engine speed accelerates and decelerates, the inertia torque is constantly
applied to the crankshaft and so it is necessary to calculate the inertia torque of the
piston-connecting rod and crankshaft assembly. Firstly, the piston assembly (piston +
wrist pin + piston rings) can be considered as a point mass experiencing translational
motion during engine operation. The connecting rod motion is more complex than the
piston assembly because both rotation and translation are included. Using a dynamically
equivalent model of the connecting rod shown in Figure 2.6, the inertia torques
developed by the motion of the piston assembly and connecting rod were calculated.
Figure 4.25. Crankshaft assembly 3-dimensional modeling.
In order to calculate the rotational inertia torque of the crankshaft assembly, the
crankshaft systems including the flywheel and drive shaft were modeled using
SOLIDWORKS which is a 3-dimensional computer modeling tools. Figure 4.25 shows
the modeled crankshaft system. Thus, the rotational moment of inertia of the crankshaft
system was calculated using 3-dimensional computer modeling. Figure 4.26 indicates the
measured instantaneous speed and calculated translational and rotational acceleration of
the baseline engine and the RLE.
95
-180 0 180 360 540120
122
124
126
128
130
Ang
ular
spe
ed (
rad/
s)
Crank angle (deg)
-180 0 180 360 540
-8
-4
0
4
8
Line
ar S
peed
(m
/s)
Crank angle (deg)
-180 0 180 360 540-1000
-800
-600
-400
-200
0
200
400
600
800
1000
Line
ar A
ccel
eart
aion
(m
/s^2
)
Crank angle (deg)
-180 0 180 360 540-1000
-800
-600
-400
-200
0
200
400
600
800
1000
Ang
ular
acc
eler
atio
n (r
as/s
^2)
Crank angle (deg)
Baseline engine
-180 0 180 360 540120
122
124
126
128
130
Ang
ular
spe
ed (
rad/
s)
Crank angle (deg)
-180 0 180 360 540
-8
-4
0
4
8
Line
ar S
peed
(m
/s)
Crank angle (deg)
-180 0 180 360 540-1000
-800
-600
-400
-200
0
200
400
600
800
1000
Line
ar A
ccel
eart
aion
(m
/s^2
)
Crank angle (deg)
-180 0 180 360 540-1000
-800
-600
-400
-200
0
200
400
600
800
1000
Ang
ular
acc
eler
atio
n (r
as/s
^2)
Crank angle (deg)
RLE
Figure 4.26. Rotational and translational speed and acceleration of the baseline engine and the Rotating Liner Engine at 1200 rpm.
96
Using the information of linear speed, linear acceleration, angular speed, and
angular acceleration, the linear acceleration torque and the angular acceleration torque
were calculated and some of the results are shown in Figure 4.27. The total inertia torque
is just the sum of linear inertia torque and angular inertia torque.
-180 0 180 360 540-20
-16
-12
-8
-4
0
4
8
12
16
20
Baseline Engine
Line
ar a
ccel
erat
ion
torq
ue (
Nm
)
Crank angle (deg)
Baseline Engine
-180 0 180 360 540-200
-160
-120
-80
-40
0
40
80
120
160
200
Ang
ular
acc
eler
atio
n to
rque
(N
m)
Crank angle (deg)
-180 0 180 360 540-20
-16
-12
-8
-4
0
4
8
12
16
20
RLE
Line
ar a
ccel
erat
ion
torq
ue (
Nm
)
Crank angle (deg)
RLE
-180 0 180 360 540-200
-160
-120
-80
-40
0
40
80
120
160
200
Ang
ular
acc
eler
atio
n to
rque
(N
m)
Crank angle (deg)
Figure 4.27. Inertia torques developed by translational and rotational motion at 1200
rpm.
From Figure (4.27) it can be concluded that the inertia torque generated by piston
translational motion is about 1/10 of the crankshaft rotational inertia torque. Now, using
the calculated pressure and inertia torques, the frictional torque of the baseline engineand
the RLE can be calculated. Figure 4.28 illustrates the measured motoring torques, the
pressure torques, the inertia torques, and the friction torques from the baseline engine and
the RLE. The blue (thick line) is for the baseline engine and the red (dashed line) is for
the RLE.
97
-180 0 180 360 540
-100
-50
0
50
100
Crank angle (deg)
Fric
tion
torq
ue (
N-m
)
-200-150-100-50
050
100150200
Iner
tia to
rque
(N
-m)
-200-150-100-50
050
100150200
Pre
ssur
e to
rque
(N
-m)
-200-150-100-50
050
100150200
Baseline RLE
Out
put t
orqu
e (N
-m)
Figure 4.28. Measured output torque, pressure torque, inertia torque, and friction torque of the baseline engine and the RLE at 1200 rpm.
98
In Figure 4.28 the friction torque shows negative values (especially near 15
degrees BTDC of compression) during the compression stroke and the sign of the friction
torque also crosses zero during the other three strokes, which is physically impossible.
However, other researchers reported that negative friction torques were observed in some
crank angles. Since the crankshaft rotational speed is measured by an optical encoder
installed in the engine front case, there could be a phase difference between the optical
encoder and each crankshaft position if there is a torsional deformation of the engine
drive system. In fact if the crankshaft and the driveshaft system of the dynamometer do
not rotate as a rigid body, the basic equation (such as Equation 2.9) cannot describe the
real physical system. Therefore, the calculated frictional torques could become negative
due to the angular deflection between the angle encoder measurement position and the
active cylinder. In Figure (4.28) the noticeable negative friction torque can be observed
right before the compression TDC. It can be imagined that the torsional stress on the
crankshaft will be maximized just before the compression TDC since the pressure torque
from the cylinder is maximized and reverse to the rotational direction in this region. After
compression TDC, the friction torque shows positive values for most of crank angles.
Therefore, in order to understand and explain the negative friction torque, the dynamic
characteristics of the crankshaft and the drive system were studied and analyzed, as
discussed in chapter 5.
99
4.2 Firing tests
4.2.1 Baseline engine
Even though hot motoring tests are widely used to estimate the friction of engine
components, the friction loss during motoring is different from that of firing. Thus, it is
best to measure the engine friction under firing conditions. However, the friction
measurement of the piston assembly under firing condition is still a challenging problem.
In this research the piston assembly friction measurement under firing condition was
done by using the Instantaneous IMEP method. For the IMEP method, the cylinder
pressure and connecting rod forces were measured during fifty engine cycles at 800 rpm,
1200 rpm, and 2000 rpm engine speed. Figures 4.29 and 4.30 show the ensemble
averaged cylinder pressure forces and connecting rod forces measured at the WOT
condition.
-180 0 180 360 540-2000
0
2000
4000
6000
8000
10000
12000
14000
16000
18000
1200 rpm
800 rpm
1200 rpm
Pre
ssur
e fo
rce
(N)
Crank angle (deg)
800 rpm
-180 0 180 360 540-2000
0
2000
4000
6000
8000
10000
12000
14000
16000
18000
Con
_rod
forc
e (N
)
Crank angle (deg)
-180 0 180 360 540-4000
0
4000
8000
12000
16000
20000
24000
Pre
ssur
e fo
rce
(N)
Crank angle (deg)
-180 0 180 360 540-4000
0
4000
8000
12000
16000
20000
24000
Con
_rod
forc
e (N
)
Crank angle (deg)
Figure 4.29a. Measured cylinder pressure and connecting rod forces at the WOT firing condition (800 and 1200 rpm).
100
-180 0 180 360 540-4000
0
4000
8000
12000
16000
20000
24000
28000
2000 rpm
1600 rpm
2000 rpm
Pre
ssur
e fo
rce
(N)
Crank angle (deg)
1600 rpm
-180 0 180 360 540-4000
0
4000
8000
12000
16000
20000
24000
28000
Con
_rod
forc
e (N
)
Crank angle (deg)
-180 0 180 360 540-4000
0
4000
8000
12000
16000
20000
24000
28000
Pre
ssur
e fo
rce
(N)
Crank angle (deg)
-180 0 180 360 540-4000
0
4000
8000
12000
16000
20000
24000
28000
Con
_rod
forc
e (N
)
Crank angle (deg)
Figure 4.29b. Measured cylinder pressure and connecting rod forces at WOT firing condition (1600 and 2000 rpm).
-180 0 180 360 540-5000
0
5000
10000
15000
20000
25000 800 rpm1200 rpm1600 rpm2000 rpm
Con
n_ro
d fo
rce
(N)
Crank angle (deg)
-180 0 180 360 540
0
5000
10000
15000
20000
25000 800 rpm1200 rpm1600 rpm2000 rpm
Pre
ssur
e fo
rce
(N)
Crank angle (deg)
Figure 4.30. Effect of engine speed on the pressure and connecting rod forces throughout the cycle for WOT firing conditions.
101
As shown in Figure 4.30, the cylinder pressure under firing conditions is quite
different at different engine speeds. Thus, the effect of engine speed on the friction force
of the piston assembly is complicated by the effect of engine speed on cylinder pressure
in addition to the direct dependence on engine speed. Basically, it is known that the gas
pressure behind a ring provides the major contribution to the sealing force. Thus, in the
case of the top compression ring the gas in the combustion chamber passes down the
clearance space between the piston crown land and the cylinder liner and then into the top
ring groove to load the rear face of the ring. Thus, this top groove pressure affects the
piston ring assembly friction since the high groove pressure exerts piston ring side force
and increases the normal load of the ring against the cylinder liner. It is usual in
lubrication analyses of piston ring packs to assume that the pressure in the top ring
groove is at all times equal to the combustion chamber pressure. Figure 4.31 shows the
inertia forces of the connecting rod and piston assembly at different speeds. Figure 4.32
clearly shows the total inertia force variation with engine speed variation. The inertia
forces under firing conditions are a little different compared with the motoring case
mainly due to the different instantaneous engine speed of firing from that of motoring.
The friction forces of the piston ring assembly under firing conditions were calculated for
each engine cycle using the measured cylinder pressure and calculated inertia forces for
each cycle. The friction force at each cycle was then ensemble averaged for 50 engine
cycles. The friction forces shown in Figure 4.32 are the ensemble-averaged friction forces
at different engine speeds.
102
- 1 8 0 0 1 8 0 3 6 0 5 4 0- 4 0 0
- 3 0 0- 2 0 0- 1 0 0
01 0 02 0 03 0 0
4 0 0
C r a n k a n g l e ( d e g )
Tot
al in
ertia
forc
e (N
)
- 6 0
- 4 0
- 2 0
0
2 0
4 0
6 0
Con
_rod
iner
tia fo
rce
(N)
- 3 0 0
- 2 0 0
- 1 0 0
0
1 0 0
2 0 0
3 0 0
Pis
ton
iner
tia fo
rce
(N) 8 0 0 r p m
- 1 8 0 0 1 8 0 3 6 0 5 4 0- 1 0 0 0
- 5 0 0
0
5 0 0
1 0 0 0
C r a n k a n g l e ( d e g )
Tot
al in
ertia
forc
e (N
)
- 1 5 0
- 1 0 0
- 5 0
0
5 0
1 0 0
1 5 0
Con
_rod
iner
tia fo
rce
(N)
- 1 0 0 0
- 5 0 0
0
5 0 0
1 0 0 0
Pis
ton
iner
tia fo
rce
(N) 1 2 0 0 r p m
103
Figure 4.31a. Measured inertial forces of the piston assembly and the connecting rod under WOT firing conditions (800 and 1200 rpm).
- 1 8 0 0 1 8 0 3 6 0 5 4 0- 1 5 0 0
- 1 0 0 0
- 5 0 0
0
5 0 0
1 0 0 0
1 5 0 0
C r a n k a n g l e ( d e g )
Tot
al in
ertia
forc
e (N
)
- 2 0 0- 1 5 0- 1 0 0
- 5 00
5 01 0 01 5 02 0 0
Con
_rod
iner
tia fo
rce
(N)
- 1 5 0 0
- 1 0 0 0
- 5 0 0
0
5 0 0
1 0 0 0
1 5 0 0
Pis
ton
iner
tia fo
rce
(N) 1 6 0 0 r p m
- 1 8 0 0 1 8 0 3 6 0 5 4 0- 2 5 0 0- 2 0 0 0- 1 5 0 0- 1 0 0 0
- 5 0 00
5 0 01 0 0 01 5 0 02 0 0 0
C r a n k a n g l e ( d e g )
Tot
al in
ertia
forc
e (N
)
- 3 0 0
- 2 0 0
- 1 0 0
0
1 0 0
2 0 0
3 0 0
Con
_rod
iner
tia fo
rce
(N)
- 2 0 0 0- 1 5 0 0- 1 0 0 0
- 5 0 00
5 0 01 0 0 01 5 0 02 0 0 0
Pis
ton
iner
tia fo
rce
(N) 2 0 0 0 r p m
Figure 4.31b. Measured inertial forces of the piston assembly and the connecting rodunder WOT firing conditions (1600 and 2000 rpm).
104
-180 0 180 360 540-2000
-1500
-1000
-500
0
500
1000
1500
800 rpm1200 rpm1600 rpm2000 rpm
Iner
tia fo
rce
(N)
Crank angle (deg)
Figure 4.32. Effects of engine speed on the total inertia force throughout the cycle under
WOT firing conditions for the baseline engine.
105
-180 0 180 360 540-2000-1500-1000-500
0500
10002000 rpm
Crank angle (deg)
Fric
tion
forc
e (N
)
-2000-1500-1000-500
0500
10001600 rpm
Fric
tion
forc
e (N
)
-2000-1500-1000-500
0500
10001200 rpm
Fric
tion
forc
e (N
)
-1000
-500
0
500
1000F
rictio
n fo
rce
(N)
800 rpm
Figure 4.33. Friction force of the piston assembly under WOT firing conditions.
106
In Figure 4.33, the friction forces have two peaks after the compression TDC
position, especially at low engine speed. The first peak in the friction force occurs just
after compression TDC at which time the main piston ring friction is in the boundary
lubrication region due to its slow piston speed. The second peak can be seen near the
peak cylinder pressure position. Under firing conditions, the peak cylinder pressure is
observed around 10° ~ 20° crank angles after compression TDC. The lubrication regime
between the piston ring and the cylinder liner can be changed to boundary or mixed
lubrication near the peak cylinder pressure location since the normal load of the piston
ring on the liner increases rapidly even though the piston speed is not zero at this crank
angle (but is still relatively slow). The pressure force, the inertia force, and the friction
force in Figures 4.30, 4.32, and 4.33 are the ensemble-averaged values. However, in the
case of firing tests, the cylinder pressure and the instantaneous engine speed are
continuously variant during the measurements. Therefore, it should be questioned
whether the ensemble-averaged friction forces are truly representative. More specifically,
it is common for modelers to use the ensemble-averaged cylinder pressure to predict
piston assembly friction. However, ring/liner friction is a highly nonlinear function of
cylinder pressure via, for example, the Stribeck diagram. Since the cyclic variation of the
cylinder pressure can be expressed using statistical parameters such as the CoV of the
IMEP, the cycle-by-cycle variations of the friction forces should be also defined and
calculated for the individually measured engine cycles.
4.2.2 Cyclic variation s
Figure 4.34 indicates the measured cylinder pressure variation during firing
engine operation at WOT conditions. The cylinder pressures were measured during
consecutive 50 engine cycles. The pressures in Figures 4.29 and 4.30 are the ensemble-
averaged values calculated from total measured pressures. As the cylinder pressure is
fluctuating from the minimum to maximum values during 50 cycles, the piston assembly
frictions is also influenced by the cylinder pressure variation each engine cycle.
Therefore, in addition to the ensemble-averaged friction force, it is valuable to find out
the maximum and the minimum range of piston ring assembly friction force during the
107
measured engine cycles. Through the analysis of each engine cycle, the new parameter
which can define the cyclic variation of the friction forces is defined and the friction
force variations during the measured engine cycles are explained based on this parameter.
-180 0 180 360 540-2000
0
2000
4000
6000
8000
10000
12000
14000
16000
18000
20000
1200 rpm
2000 rpm
800 rpm
1600 rpm
Pre
ssur
e fo
rce
(N)
Crank angle (deg)
-180 0 180 360 540-2500
0
2500
5000
7500
10000
12500
15000
17500
20000
22500
25000
Pre
ssur
e fo
rce
(N)
Crank angle (deg)
-180 0 180 360 540-2500
0
2500
5000
7500
10000
12500
15000
17500
20000
22500
25000
27500
30000
Pre
ssur
e fo
rce
(N)
Crank angle (deg)-180 0 180 360 540
-2500
0
2500
5000
7500
10000
12500
15000
17500
20000
22500
25000
27500
30000
Pre
ssur
e fo
rce
(N)
Crank angle (deg)
Figure 4.34. Cyclic cylinder pressure variations for the baseline engine under WOT
firing conditions.
108
-180 0 180 360 540-5000
-2500
0
2500
5000
7500
10000
12500
15000
17500
20000
-1000
-800
-600
-400
-200
0
200
400
600
800
1000
Pre
ssur
e fo
rce
(N)
Crank angle (deg)
Ensemble averaged value
Maximum pressure cycle
Minimum pressure cycle
Fric
tion
forc
e (N
)
Figure 4.35. Pressure force and friction force variations during the cycle at 800 rpm.
Figure 4.35 presents the ensemble-averaged, maximum and minimum of the
cylinder pressure and friction forces during 50 engine cycles at 800 rpm. As expected, the
piston ring assembly friction is dependent on the cylinder pressure variation. The
variation of the second peak of friction forces is very dependent on that of the peak
cylinder pressure. The crank angles at which the friction forces become maximum or
minimum are coincident with that of maximum pressure or minimum pressure. Figure
4.36 indicates the relation of the cylinder pressure and friction force variation from 0°
crank angle (compression TDC) to 100° in more detail. Figure 4.36 illustrates more
clearly the dependence of the friction force on cylinder pressure. However, the crank
angle at which the peak friction force is observed is not exactly coincident with that of
peak cylinder pressure. That is, the peak friction force occurs around 6° crank angle after
the peak cylinder pressure is observed. This peak friction force delay compared with peak
cylinder pressure can be observed at different engine speeds.
109
0 20 40 60 80 100-5000
-2500
0
2500
5000
7500
10000
12500
15000
17500
20000
-1000
-800
-600
-400
-200
0
200
400
600
800
1000
Pre
ssur
e fo
rce
(N)
Crank angle (deg)
Fric
tion
forc
e (N
)
Figure 4.36. Crank angles at peak pressure and friction forces at 800 rpm.
Figures 4.37 and 4.38 show the measured ensemble-averaged, maximum, and
minimum cylinder pressure and friction force at 1200 rpm and 1600 rpm. The friction
forces at different engine speeds show the same characteristics with that of 800 rpm. That
is, the friction forces show two peaks after the compression TDC and the second peak in
the friction forces is related with the peak cylinder pressure. Figures 4.39 and 4.40
indicate the relation between the cylinder pressure and the friction force in detail at 1200
rpm and 1600 rpm. Even though the engine speed increases, the crank angle differences
between the position of the peak pressure and that of the peak friction force are constant
independent of the engine speed. That is, there is about a 6° crank angle difference
between two peaks. It is considered that the phase lag of the friction forces is from the
squeeze film effect of oil between the piston ring and the liner. It is believed that the peak
friction forces are experienced at which the oil thickness is minimized. As the cylinder
pressure reaches its maximum value after TDC, the groove pressures are also maximized
and push the top compression ring toward the liner and then the film thickness comes to
its local minimum value. However, due to the squeeze film effect there is a time lag
between the crank angle of peak cylinder pressure and that of minimum film thickness.
110
-180 0 180 360 540-5000
-2500
0
2500
5000
7500
10000
12500
15000
17500
20000
22500
25000
-1400
-1200
-1000
-800
-600
-400
-200
0
200
400
600
800
1000P
ress
ure
forc
e (N
)
Crank angle (deg)
Ensemble average Maximum pressure cycle Minimum pressure cycle
Fric
tion
forc
e (N
)
Figure 4.37. Pressure force and friction force variations during the cycle at 1200 rpm.
-180 0 180 360 540-5000
0
5000
10000
15000
20000
25000
30000
-2500
-2000
-1500
-1000
-500
0
500
1000
Pre
ssur
e fo
rce
(N)
Crank angle (deg)
Ensemble average Maximum pressure cycle Minimum pressure cycle
Fric
tion
forc
e (N
)
Figure 4.38. Pressure force and friction force variations during the cycle at 1600 rpm.
111
0 20 40 60 80 100-5000
-2500
0
2500
5000
7500
10000
12500
15000
17500
20000
22500
25000
-1400
-1200
-1000
-800
-600
-400
-200
0
200
400
600
800
1000
Pre
ssur
e fo
rce
(N)
Crank angle (deg)
Fric
tion
forc
e (N
)
Figure 4.39. Crank angles at peak pressure and friction forces at 1200 rpm.
0 20 40 60 80 100-5000
0
5000
10000
15000
20000
25000
30000
-2500
-2000
-1500
-1000
-500
0
500
1000
Pre
ssur
e fo
rce
(N)
Crank angle (deg)
Fric
tion
forc
e (N
)
Figure 4.40. Crank angles at peak pressure and friction forces at 1600 rpm.
112
During the firing tests, the combustion stability or variation can usually be
expressed as the coefficient of variation in indicated mean effective pressure, CoV of
IMEP as defined in Equation 4.5.
100×=IMEP
COV IMEPIMEP
σ (4.5)
where:
IMEPσ : standard deviation in IMEP
IMEP: mean IMEP
Basically IMEPCOV defines the cyclic variability in indicated work per cycle and is used
to represent the combustion stability. Table 4.2 indicates the measured IMEP and CoV of
IMEP in this research.
800 rpm 1200 rpm 1600 rpm 2000 rpm
IMEP (kPa) 713.8 870.3 933.2 1005.6
Indicated work (Nm) 448.2 546.4 585.9 631.4
CoV of IMEP (%) 3.195 2.079 1.139 2.012
Table 4.1 IMEP and CoV of IMEP for the firing tests of the baseline engine
Therefore, from Table 4.1 it can be said that the mean IMEP or ensemble
averaged pressure cycle can have a variation that is the same as the CoV of IMEP in the
logged cycles. For example, at 800 rpm, the measured pressures show a 3.195 %
variation around the ensemble averaged pressure cycle. Since the piston ring assembly
friction is quite nonlinearly dependent on the cylinder pressure, the friction force will
have a variation, which is probably much greater than the combustion variation. It is
known that the oil film thickness between the piston ring and the liner has cycle-by-cycle
variations. Therefore, due to the combustion variation, the oil film thickness variation, the
113
dynamic variation, and so on, the measured friction forces have cyclic variability during
the measured cycles. In order to quantify the cyclic variation of the measured friction
forces, the friction work was calculated. Friction work can be defined via Equation 4.6
which is the product of the friction force and the distance over which the friction forces
load on the piston.
pfricitonfriction xFW ⋅= (4.6)
where:
frictionW : friction work
frictionF : piston assembly friction forces
px : distances swept by the piston during each crank angle under this friction force
In Equation 4.6 px is the distance swept by the piston during each crank angle.
Figure (4.41) represents x and px , which are the position from the TDC position and the
distance swept by the piston during each crank angle.
114
-180 0 180 360 5400.0000
0.0001
0.0002
0.0003
0.0004
0.0005
Sw
ept d
ista
nce
Xp
(m)
Crank angle (deg)
-180 0 180 360 5400.08
0.10
0.12
0.14
0.16
0.18
0.20
Pis
ton
posi
tion
x (m
)
Crank angle (deg)
Figure 4.41. Position of the piston top relative to the head and distance swept by the
piston at 800 rpm.
As shown in Figure 4.41, the swept distance of the piston during each crank angle
can be used to calculate the friction work during the firing operation. As with the CoV of
IMEP, the friction work COV can be defined via Equation 4.7.
100×=friction
frictionfriction W
COVσ
(4.7)
where:
frictionσ : standard deviation of friction work
frictionW : mean friction work
Table 4.2 shows the calculated the mean friction work and the COV of the piston
friction work. In Table 4.2 the friction work at 2000 rpm was not included since the noise
affected the friction work for this firing condition. It is reasonable that the measured
piston friction work occupies 7.7% ~ 9.2% of the indicated work. Thus, it can be said that
the piston friction measurement using the instantaneous IMEP method is useful to
115
analyze the piston friction quantitatively in addition to qualitative manner. As expected,
the mean friction work COV is much greater than the CoV of IMEP due to the cyclic
variation of oil film thickness, dynamic instability, and so on in addition to combustion
variation. Roughly the friction work COV is three times greater than the CoV of IMEP.
Therefore, it is more reasonable to assume that the average piston assembly friction
cannot be calculated using the ensemble-averaged pressure than to assume that it can.
800 rpm 1200 rpm 1600 rpm
Mean friction work (Nm) 34.73 42.07 53.85
Mean friction work COV (%) 10.697 7.216 2.824
Friction work/Indicated work (%) 7.75 7.7 9.19
Table 4.2 Mean friction work and friction work COV
4.2.3 Conclusions on the friction forces of the firing condition
Through the application of the IMEP method on firing condition, the friction
forces of the piston assembly were measured. The measured friction forces show the
several characteristics. During the firing the friction forces show the two peaks during the
engine cycle. The first peak is observed near at compression TDC and the other one is
near at peak cylinder pressure. The first peak is connected with the boundary friction
force, which is happening at TDC due to the zero piston speed. The second friction peak
is related with the peak cylinder pressure, which applies to the compression ring and
squeezes the oil between the ring and the liner. Due to the high cylinder pressure, the
compression ring becomes contacts with the liner and the boundary lubrication is
dominant during the high cylinder pressure. The maximum friction forces have shown
from 500N to 1500N at measured engine speeds. These maximum friction forces are
much greater than those of other researchers’ measurements. In fact the maximum
friction forces of the piston assembly are changed according to the measurement method
and the researchers [2, 8, 9, 21, 25, 37, 38]. From the literature survey the measured peak
friction forces range from 20N to 1000N. The measured friction forces are different
116
according to test engine, test method. In this research the measured friction forces are
confirmed using the friction force energy method. That is, the measured friction forces
are converted to the friction energy and the calculated friction energy was changed from
7% to 10% of the IMEP energy according to the measured engine speed. Thus, it can be
said that the measured friction forces are acceptable from the stand point of the energy
balance. In addition to the magnitude of the friction force the phase difference between
the peak cylinder pressure and the peak friction force is another point to be discussed.
The measured pressure and friction forces show the phase difference around 6° crank
angle. From the theoretical analysis it is natural that these forces have a phase difference.
However, this phase difference was independent of the engine speed. Even if the engine
speeds are changed, the phase difference is not much changed. The experimental results
are different from the expectation. However, it was hard to find the literature about the
experimental analysis of the phase difference between the cylinder pressure and the
piston friction force. Therefore, the theoretical analysis of the cylinder pressure and the
piston friction force should be more studied for future work.
117
4.3 Piston assembly friction force analysis using the instantaneous IMEP and P-w
methods
Although the instantaneous IMEP method can be used to measure the piston
assembly friction force, the measured piston friction force cannot provide any
information about the lubrication mechanism. That is, the measured piston assembly
friction using the IMEP method includes lots of information, such as piston viscous
lubrication friction components, mixed lubrication friction components, and piston skirt
friction. Thus, the piston assembly friction modeling is required to determine each
friction component’s contribution to the total piston assembly friction. In fact, a number
of engine friction models have been developed for better understanding and design about
engine mechanical friction losses. Historically, the first studies about engine mechanical
losses were published in the late fifties and mid-sixties. Even though many researchers
have measured the engine mechanical friction and developed the friction models to
explain the physical phenomena in engine friction losses since the late fifties, the
measurement and analysis of frictional losses incurred during engine running have never
been satisfactorily resolved. One of the problems of engine friction modeling is that the
friction models usually include all components of the engine. Thus, it is a challenging
problem to correlate the friction model with the measured engine friction. Another
problem of the engine friction modeling is that lots of mechanical friction models predict
the engine frictional losses on a cycle base rather than a crank angle base. Of course, the
engine frictional models based on the engine cycle are effective to develop a new engine
designs and estimate the frictional losses of the developed engine. However, since the
cycle based frictional models cannot analyze the engine frictional loss based on a crank
angle basis, it is difficult to understand and explain the frictional phenomena happening
during the engine cycle. Therefore, the crank angle based frictional model is also needed
for better understanding of the friction mechanism of the mechanical components and
reducing the frictional losses. In this subsection the measured piston assembly friction
will be analyzed using the P-w method to determine the contribution of each friction
mechanism to the total measured piston assembly frictional losses.
118
4.3.1. Piston assembly friction modeling
In the P-w method, the piston assembly friction can be divided into three main
parts such as ring viscous lubrication friction, ring mixed lubrication friction, and piston
skirt friction. These three friction components can be modeled mathematically. In ring
viscous friction modeling the friction force can be expressed as
Friction force due to ring viscous lubrication = oil film coefficient of friction * normal
load (4.8)
In Equation 4.8 the normal load is computed as the product of pressure on the ring
(including ring tension) and the projected ring area. The pressure and the projected area
of the ring can be expressed as:
Pressure on the ring = Pgas + Pe (4.9)
Projected area = ring)ofnumber equivalent(⋅⋅ owD (4.10)
where:
Pe: elastic pressure of the ring (ring tension)
D: cylinder diameter
wo: width of oil ring
The oil film coefficient of friction increases with the increase in the dynamic oil film
viscosity and with increasing piston speed. It decreases with the increase of the pressure
on the ring and the ring width because of the possibility of a favorable hydrodynamic
wedge. From the bench test [41]
Oil film friction coefficient
5.0
)(
⋅+⋅∝
oegas
p
wPP
vµ (4.11)
where:
µ: oil dynamic viscosity
vp: piston speed
119
Therefore, the friction torque generated from ring viscous lubrication is expressed.
Ring viscous lubrication friction:
lcoeoprvl RrDnnppwvcT ⋅⋅⋅⋅+⋅+⋅⋅⋅⋅= )4.0()(1 µ (4.12)
where:
no: number of oil rings
nc: number of compression rings
r: crank radius
wr
cRl ⋅= (4.13)
lRwr
L
r
L
r
wrdt
dxc ⋅⋅=
⋅
−
⋅
+⋅⋅==
2/1
22
sin1
cos1sin
φ
φφ : Instantaneous piston velocity
(4.14)
The mixed lubrication friction torque can be described according to Winer and Cheng
[40]
lrml RrT ⋅−⋅⋅= )1(loadnormalfrictionoft coefficien λ (4.15)
In Equation 4.15 λ is a parameter to determine the mode of lubrication. λ can be
simplified and assumed to be equal to sinφ. Therefore, the ring mixed lubrication friction
torque has the form like:
Ring mixed lubrication friction:
leccrml RrppwnDcT ⋅⋅−⋅+⋅⋅⋅⋅⋅= )sin1()(2 φπ (4.16)
120
The piston skirt friction is formulated by applying Newton’s law for viscous friction and
calculating the corresponding torque,
lp
lps
RrMDh
v
RrT
⋅⋅⋅⋅=⋅⋅⋅=
)()(
area)skirt (projectedstress)shear (Oil
µ (4.17)
Therefore, the piston skirt friction torque is finally expressed as
Piston skirt friction:
lpsl
ps RrLDh
RwrcT ⋅⋅⋅⋅⋅⋅⋅⋅= )(
3 µ (4.18)
where:
h: oil film thickness
Lps: length of piston skirt
From Equations 4.12, 4.16, and 4.18 the piston assembly friction torque can be expressed
as the linear combination of its individual components. That is,
∑=
=++=3
1332211
jjjf wcwcwcwcT
th (4.19)
Baseline engine parameters employed for frictional modeling are given in Table 4.3.
L Connecting rod length 0.147 m
r Crank radius 0.0425 m
µ Oil dynamic viscosity 0.015 kg/ms
wo Width of oil ring 0.00292 m
wc Width of compression ring 0.00119 m
Pe Elastic ring pressure force 26050 N/m2
No Number of oil rings 1
Nc Number of compression rings 2
Lps Length of piston skirt 0.03 m
h Oil film thickness 0.000005 m
D Cylinder diameter 0.092 m
Table 4.3 Engine basic parameters used for the friction model
121
The experimental friction torque and theoretical friction torque can be compared
at each crank angle. That is,
kff kTkTth
ε+= )()(exp
(4.20)
In Equation 4.20 kε indicates the error between the measurement and the prediction.
Thus, the problem is to determine the optimum coefficients c1, c2, and c3 which minimize
the error kε and fit the experimental results. In order to find out the optimum coefficients
which minimize the error, the linear regression technique was applied in this research.
4.3.2 Comparison of predicted and experimental piston ring assembly friction
torque values
The piston assembly friction torque can be calculated using the measured friction
force. Figure 4.42 represents the calculated friction torque using the measured piston
friction forces at 1200 rpm. Since the friction torque of the piston assembly includes all
kinds of friction sources, the P-w method was used to determine the contribution of each
friction components.
-180 -120 -60 0 60 120 1800
5
10
15
20
25
30
35
40
Fric
tion
torq
ue (
Nm
)
Crank angle (deg)
Cold motoring Hot motoring Firing
Figure 4.42. Measured piston assembly instantaneous friction torque at 1200 rpm.
122
Figures 4.43 and 4.44 show the measured and simulated friction torque using the
P-w method for cold and hot motoring conditions. During motoring, the cylinder pressure
is symmetrical around compression TDC. That is, the compression and expansion
pressure are almost the same during the motoring for a specific number of crank angles
before TDC as for the same crank angle increment after TDC. From the friction torque
models such as Equations 4.12, 4.16, and 4.18, it can be known that the cylinder pressure
affects the ring viscous lubrication and mixed lubrication friction. Thus, since the
motoring cylinder pressure has symmetrical characteristics between the compression and
the expansion, the modeled friction torque should show a symmetrical pattern during
motoring. However, the experimental friction torque does not indicate the symmetrical
characteristics during motoring. That is, the friction torque of the compression stroke is
higher than that of the expansion stroke. Therefore, as can be seen in Figures 4.43 and
4.44 the modeled friction torque during motoring does not simulate the experimental
friction torque successfully. This means that the present friction torque modeling cannot
represent the nonlinearity of the motoring friction torque, or that there is an error in the
measurements. However, although the modeled friction torque could not simulate the
measured friction torque perfectly, the change of friction mechanism between the cold
and the hot motoring can be proved using the friction torque modeling.
-180 -120 -60 0 60 120 180
0
5
10
15
20
25
Fric
tion
torq
ue (
Nm
)
Crank angle (deg)
Ring viscous lubrication Ring mixed lubrication Skirt lubrication Total calculated friction torque Experimental friction torque
Figure 4.43. Friction torque obtained for the baseline engine using the p-w method for cold motoring.
123
-180 -120 -60 0 60 120 180
0
5
10
15
Fric
tion
torq
ue (
Nm
)
Crank angle (deg)
Ring viscous lubrication Ring mixed lubrication Skirt lubrication Total calculated friction torque Experimental friction torque
Figure 4.44. Friction torque obtained for the baseline engine using the P-w method for hot motoring.
That is, as can be seen in Figure 4.43 during cold motoring the ring viscous
lubrication and the skirt lubrication mechanism are dominant and the ring mixed
lubrication is negligibly small in the theoretical friction torque. This result can be
expected because the oil film viscosity during cold motoring is much greater than that for
hot motoring and so the dominant lubrication could be hydrodynamic rather than mixed
lubrication. The predicted friction torque shows the same trend what we expect during
cold motoring friction tests. Figure 4.44 indicates the measured motoring torque and the
break down results of the hot motoring torque. It can be concluded from Figure 4.44 that
the main friction torque is from the ring viscous lubrication and the ring mixed
lubrication, not from skirt friction torque. The skirt lubrication friction is negligible
during hot motoring. The difference of the lubrication mechanism between the cold and
the hot motoring is mainly caused by the difference of oil viscosity. During the low
temperature motoring test the oil viscosity is so high that the main lubrication mechanism
is the ring viscous and the skirt lubrication. However, during high temperature motoring,
the oil viscosity is low enough to neglect the skirt friction torque.
124
-180 -120 -60 0 60 120 1800
5
10
15
20
25
30
35
40
Fric
tion
torq
ue (
Nm
)
Crank angle (deg)
Ring viscous lubrication Ring mixed lubrication Skirt lubrication Total calculated friction torque Experimental friction torque
Figure 4.45. Friction torque obtained for the baseline engine using the p-w method for
WOT firing conditions.
Figure 4.45 shows the friction torque comparison between the measured and the
theoretical torques under firing condition. The theoretical friction torque using the P-w
method follows the measured friction torque much closer than was true for the motoring
case. That is, the theoretical friction modeling shows better simulated results, which are
similar to the measured friction torque under firing conditions, than the motoring
simulation. During the firing tests, the ring viscous and mixed lubrication have influential
effects on friction torque. The skirt friction torques during the firing can be negligible
compared with the ring viscous and mixed lubrication friction torque.
A more comprehensive engine friction model, Ricardo’s RINGPAK software, is
discussed in Chapter y, after the problems with the experimental measurements of crank-
angle-resolved friction are explored in detail.
125
Chapter 5. Error Analysis in Friction Force Measurement
5.1 Introduction
In chapter 4, the friction forces were measured using three different measurement
methods: direct motoring, p-w method, and instantaneous IMEP method. The friction
forces using the p-w method and the instantaneous IMEP show the negative friction
during the engine cycle, which is physically impossible. Therefore, in this chapter, the
measured errors connected with the p-w method and the instantaneous IMEP method are
discussed and analyzed.
5.2 Measured friction errors and analysis in the p-w method
The friction torque generated from the p-w method showed the negative values
around the compression TDC positions. Since the negative friction torques are physically
impossible, the negative friction torque should be corrected. From the literature [15, 16,
17], one of the important reasons of the negative friction torque is from the torsional
vibrations of the engine crankshaft system. In those researches [15, 16, 17], the engine
operating conditions were no load. The friction torque was calculated from the measured
inertia torque and the pressure torque. Therefore, the analysis of the torsional vibration
for resolving the negative friction torque was only concentrated in the crankshaft system.
However, in this research, the engine output load was measured using the in-line torque
sensor and the friction torque was calculated from the measured inertia torque, the
pressure torque and the output load torque. Thus, in order to analyze the negative friction
torque, the torsional vibration of the crankshaft and the driveshaft should be analyzed. In
order to analyze the torsional vibration, the dynamic characteristics and the torsional
vibration modes of the engine and the driveshaft system are analyzed. In this research the
transfer matrix method was used to simulate the dynamic system of the engine and the
driveshaft. In order to analyze the dynamic characteristics of the crankshaft it is needed
to model the engine crank shaft system. Basically the crankshaft and piston assembly can
be modeled as a combination of equivalent masses, mass-less springs and dampers. The
126
dynamic characteristics of a crankshaft system can be described as a series of linear
differential equations. The single cylinder engine used in this experiment can be modeled
as in Figure (5.1).
D3J1
D1J2
K1
D2
J3
K2 K4
D4
J4
K3
J5
D5
K5
R1
J6D6
K6
J7
Figure 5.1. Equivalent dynamic model of the crankshaft system.
In Figure 5.1 the crankshaft system is described as the combinations of the mass
moment of inertia J, a mass-less elastic shaft with torsional stiffness K, absolute damping
R and relative damping D. Absolute damping R mainly indicates the piston damping
simulating the friction between the piston assembly and the cylinder wall. In Figure (5.1)
one absolute damper is used since the test engine is a single cylinder engine converted
from a four cylinder engine. Relative damping D is used to simulate the main journal
bearings in the crankshaft. The seven inertias represent the inertia of the crankshaft and
connecting rod, flywheel, damper pulley, and driveshaft. The equivalent dynamic model
of the crankshaft system can be expressed as seven linear differential equations.
127
776766767
66766766565565
55655655454454
44415445443433434
33433433332232
22322322121121
11211211
)()(
)()()()(
)()()()(
)()()()(
)()()()(
)()()()(
)()(
θθθθθθθθθθθθθθθθθθθθθθθ
θθθθθθθθθθθθθθθθθθθθθθθθθθθθ
θθθθθ
&&&&&&&&&&&&&&&&
&&&&&&&&&&&&&&&&&&&
&&&&
JDKM
JDKDK
JDKDK
JRDKDKM
JDKDK
JDKDK
JDK
=−−−−=−−−−−−−−=−−−−−−−−
=−−−−−−−−−=−−−−−−−−=−−−−−−−−
=−−−−
(5.1)
MKDJ =++ θθθ &&& (5.2)
=
7000000
06
00000
005
0000
0004
000
00003
00
000002
0
0000001
J
J
J
J
J
J
J
J (5.3)
−−+−
−+−−++−
−+−−+−
−
=
6600000
66550000
05544
000
0041433
00
0003322
0
00002211
0000011
DD
DDDD
DDDD
DRDDD
DDDD
DDDD
DD
D (5.4)
−−+−
−+−−+−
−+−−+−
−
=
66
6655
5544
4433
3322
2211
11
00000
0000
0000
0000
0000
0000
00000
KK
KKKK
KKKK
KKKK
KKKK
KKKK
KK
K (5.5)
128
Equation 5.1 can be expressed as a matrix like Equation 5.2. In Equation 5.2 the
components in the J, D, and K matrices can be expressed in Equations 5.3, 5.4, and 5.5.
In Equation 5.1 M4 indicates the driving torque from the combustion. M7 represents the
engine load torque, and θ ,θ& , and θ&& indicate the angular displacement, the angular speed
and the angular acceleration of the disks, respectively. In this research the transfer matrix
method was used to analyze the crankshaft dynamic characteristics.
The equivalent dynamic model of the crankshaft system can be divided into two
main components. One is the parallel connection of mass-less spring and relative damper
between the disks and the other is the serial connection of the absolute damper with the
disk.
Ki
Ci
(i-1)N
(i-1)c
(i-1)kN
N Nic
ikN
iN
Figure 5.2. Derivation of the field matrix.
In Figure 5.2 the damper forces at point i and i-1 are the same. By inspection of
Figure 5.2
)( 1)1( −− −== iiiciic xxcNN && with ))(Re()Re( 1pt
iiipt
icic exxpceNN −−== (5.6)
The complex damper force is
)( 1)1( −− −== iiiicci xxpcNN (5.7)
129
The complex spring force is:
)( 1)1( −− −== iiiikki xxkNN (5.8)
The total complex force of the parallel spring-damper assembly is:
))(( 11 −− −+=+== iiiiikicii xxpckNNNN (5.9)
Hence
pck
Nxx
ii
iii ++= −−
11 (5.10)
From Equation 5.10, the matrix relation between the complex state vector 1−iz and iz is:
1−= iii zFz (5.11)
110
11
−
•
+=
ii
iN
xcpk
N
x (5.12)
In Equation 5.11, Fi means the field matrix in Figure 5.2. In order to calculate the point
matrix in Figure 5.1 the serial connection between an absolute damping and inertia can be
simplified as in Figure 5.3.
130
NRNL
mpxi
i
i2
i
irpx
Figure 5.3. Derivation of the point matrix.
The point matrix in Figure 5.3 can be derived as:
Lii
Ri zPz = (5.13)
L
i
R
iN
x
rpmpN
x
•
+=
1
012 (5.14)
If the vibration system is steady state forced vibration, p in Equation 5.12 and 5.14 can be
expressed as Ωj , and then the matrices in Equation 5.12 and 5.14 become
11100
010
01
1
1 −
•
Ω+
=
i
L
i
N
xjckN
x
(5.15)
L
i
R
i
N
x
PmN
x
•
−Ω−=
1100
1
001
1
2 (5.16)
The extended matrices in Equation 5.15 and 5.16 can be converted into complex form.
131
1
222222
222222
110000
01000
010
00010
001
1−
•
Ω+Ω+Ω−
Ω+Ω
Ω+=
i
i
i
r
rL
i
i
i
r
r
N
x
N
x
ck
k
ck
c
ck
c
ck
k
N
x
N
x
(5.17)
L
i
i
i
r
r
i
r
R
i
i
i
r
r
N
x
N
x
Pkmr
Prkm
N
x
N
x
•
−+Ω−Ω
−Ω−+Ω−=
110000
10
00100
01
00001
1
2
2
(5.18)
In Equation 5.18, P can be the exciting torque produced by the active cylinder or the load
torque. The torque generated from the active cylinder can be expressed as a sum of
harmonic components.
∑=
−+=
K
kiiii t
kTt
kTTT
kk1
ImRe0 )
2sin)(
2cos)( ωω (5.19)
Basically the measured engine output torque can be decomposed as a series of
sine and cosine waves using DFT (Discrete Fourier Transformation). In Equation 5.19
kiT )( Re and
kiT )( Im are the real and imaginary components of the exciting engine torque
or load torque. K is the number of the harmonic components and ω the mean angular
velocity of the crankshaft. In this research the number of the harmonic components, K,
was set as 24. Figure 5.4 shows the measured torque values and the calculated torques as
a sum of harmonic components
132
-180 0 180 360 540-150
-100
-50
0
50
100
150
Mot
orin
g to
rque
(N
-m)
Crank angle (deg)
Experimental torque Calculated torque using harmonics
Figure 5.4. Measured and calculated motoring torque using harmonic components.
As can be seen in Figure 5.4 the calculated motoring torque using 24 harmonic
components can fully describe the measured motoring torque. Twenty four harmonic
components are sufficient to express the measured values. Basically, the governing
equations of the systems are linear differential equations. Thus, the harmonic components
of the exciting torque make an angular motion having their corresponding frequencies.
Therefore, the resulting angular displacement of the disk can be expressed as a sum of the
contributions of all harmonic components of the exciting torques.
∑=
−+=
K
kiii t
kt
kkk
1
ImRe0 2
sin)(2
cos)( ωθωθθθ (5.20)
Using the Equation 5.19 and 5.20 the transfer matrices can be calculated using the same
relation as:
Lk
LNNN
RN kkkkkkkk
zUzPFPFPz )()()( 1111)1()1( ⋅=⋅⋅⋅⋅⋅⋅⋅= −− (5.21)
133
In Equation 5.20, Lik
z )( and Rik
z )( indicate the state vector on the left-hand and right-
hand side of the disk i. The ki
z is defined as:
=
1
)(
)(
)(
)(
Im
Im
Re
Re
k
k
k
k
k
i
i
i
i
i
T
T
z θ
θ
(5.22)
In applying the transfer matrix method to the crankshaft system, the first step is to
find out the component values such as rotational inertia, torsional stiffness, torsional
damping and absolute damping. The rotational inertia of the crankshaft is calculated
using the 3-D modeled crankshaft in Figure 4.25. The torsional stiffness was computed
using FEM software (ANSYS). Figure 5.5 illustrates the generated mesh for ANSYS
analysis. Table 5.1 shows the calculated component values of equivalent dynamic model
of the crankshaft.
Figure 5.5. Mesh generated for ANSYS analysis.
134
J1 J2 J3 J4 J5 J6 J7Rotational inertia
(kgm2) 0.0056 0.014 0.0138 0.0145 0.014 0.08 0.04
K1 K2 K3 K4 K5 K6Torsional stiffness
(Nm/rad) 3.5*E5 7.85*E5 9.36*E5 7.85*E5 12.4*E5 40*E5
D1 D2 D3 D4 D5 D6Relative damping
(Nms/rad) 15 15 15 15 15 15
R1Absolute damping
(Nm/s/rad) 2
Table 5.1 Component values in the equivalent dynamic model of the crankshaft
Applying the boundary conditions to Equation 5.21, the components Re)(ki
θ and
Re)(ki
θ are computed for all considered harmonics. The boundary conditions are
expressed by the absence of exciting torques on the first disk of the dynamic system. The
relation between the first disk and the last disk can be expressed as follows:
kkzUz k 17 ⋅= (5.23)
⋅
=
1
)(
)(
)(
)(
1
)(
)(
)(
)(
Im1
Im1
Re1
Re1
5554535251
4544434241
3534333231
2524232221
1514131211
Im7
Im7
Re7
Re7
k
k
k
k
kkkkk
kkkkk
kkkkk
kkkkk
kkkkk
k
k
k
k
T
T
UUUUU
UUUUU
UUUUU
UUUUU
UUUUU
T
T
θ
θ
θ
θ
(5.24)
At the first disk, Re1 )(
kT = Im
1 )(k
T = 0.
Then, Equation 5.24 can be expressed as:
135
⋅
=
1
0
)(
0
)(
1
)(
)(
)(
)(
Im1
Re1
5554535251
4544434241
3534333231
2524232221
1514131211
Im7
Im7
Re7
Re7
k
k
kkkkk
kkkkk
kkkkk
kkkkk
kkkkk
k
k
k
k
UUUUU
UUUUU
UUUUU
UUUUU
UUUUU
T
T
θ
θ
θ
θ
(5.25)
Then, Equation 5.25 can be simplified as:
+⋅+⋅+⋅+⋅+⋅+⋅+⋅+⋅+⋅+⋅
=
kkkkk
kkkkk
kkkkk
kkkkk
kkkkk
k
k
k
k
UUU
UUU
UUU
UUU
UUU
T
T
55Im
153Re
151
45Im
143Re
141
35Im
133Re
131
25Im
123Re
121
15Im
113Re
111
Im7
Im7
Re7
Re7
)()(
)()(
)()(
)()(
)()(
1
)(
)(
)(
)(
θθθθθθθθθθ
θ
θ
(5.26)
In Equation 5.26, the unknowns are Re1 )(
kθ , Im
1 )(k
θ , Re7 )(
kθ , and Im
7 )(k
θ . Thus, there are
four unknowns and four equations. Therefore, Equation 5.26 can be solved and the
angular displacements of all other disks can be computed. Figure 5.6 shows the measured
and the calculated speed of disk 1 of the baseline engine.
-180 0 180 360 540120
122
124
126
128
130
Inst
anta
neou
s ve
loci
ty (
rad/
s)
Crank angle (deg)
CalculationExperiments
Figure 5.6. Comparison between the measured and the calculated instantaneous speed at
1200 rpm.
136
In Figure 5.26 the calculated instantaneous speed signal represents the overall
characteristics of the measured signal even though both signals do not coincide perfectly.
In order to analyze the torsional vibration of the crankshaft and the driveshaft system, the
dynamic modeling should be accurate and the calculated instantaneous speed should be
the same with the measured engine speed. However, as shown in Figure 5.6, the
calculated engine speed is not exactly same with the measured one even if the calculated
speeds resemble the measured one. Therefore, the torsional vibration using this dynamic
modeling of the crankshaft and the drive shaft system cannot be correct. The more
sophisticated modeling of the crankshaft and the driveshaft is needed for the torsional
vibration analysis and this modeling would be a good topic for future research.
5.3 Measured friction errors and analysis in the instantaneous IMEP method
The friction forces using the instantaneous IMEP method in chapter 4 show the
unreasonable signs during the expansion stroke. These friction forces are physically
impossible. Therefore, the errors connected with the friction measurement using the
instantaneous IMEP method should be analyzed and corrected. Since the friction force in
the instantaneous IMEP method is calculated from the measured inertia force, the
pressure force, and the connecting rod force, the possible errors in each force should be
estimated. Before the error estimation at each measured force, the sensitivity analysis of
each force is made to determine the dominant factors of the friction force errors.
5.3.1 Sensitivity analysis
A sensitivity analysis was used to examine the main parameters which can affect
the friction force calculations. The main parameters which can affect the friction force
calculations are the sensitivities of the strain gage and the piezoelectric pressure
transducer, and the inertia force calculations. One of the possible errors in the strain gage
measurements is due to temperature variation during the measurement, even if the
measurement was done during steady state engine operation. From Figure 3.15, the gage
factor can vary by approximately 0.6 % during 100 °F temperature variations. If the
137
strain gage temperature is changed about 30 °C during the steady state operation, the
gage factor could change by about 0.3%. Figure 5.7 shows the piston assembly friction
force variation with (1) a 30 oC change in the strain gage temperature, (2) a 0.3% change
in the output from the piezoelectric pressure sensor, and (3) a 10% change in the inertia
forces. None of these yield the correct sign for the friction force around the mid-stroke of
expansion. From Figure (5.7), it can be said that the inertia force has a negligible effect
on the peaks in the friction forces that occur before and after TDC compared to the strain
gage and the pressure forces.
-180 0 180 360 540-500-400-300-200-100
0100200300400500
0.3% high pressure force 0.3% low pressure force
Crank angle (deg)
Fric
tion
forc
e (N
)
-500-400-300-200-100
0100200300400500
10% high inertia force 10% low inertia force
Fric
tion
forc
e (N
)
-500-400-300-200-100
0100200300400500
+30deg high temp. - 30 deg low temp
Fric
tion
forc
e (N
)
Figure 5.7. Sensitivity analysis for the friction force obtained using the instantaneous IMEP method.
138
Even if the inertia forces could have some errors in their calculations, the effect of the
inertia force errors on the friction forces can be neglected. Therefore, the possible error
sources of the friction forces are from the measured pressure force and the connecting rod
force. The cylinder pressure and the connecting rod forces are measured by the
piezoelectric pressure transducer and the strain gage. The next two subsections deal with
the measurement errors connected with the strain gage and the piezoelectric pressure
transducer.
5.3.2 Measurement errors of the strain gage
In this subsection the measurement errors of the strain gages are discussed. The
tensile test was performed in order to find out the strain gage output variation at constant
load. Through this test the confidence level of the measured strain gage values can be
determined at the specific load. Table 5.2 represents the variation of the strain gage
values at each constant tensile load.
Tensile force (N)Average strain gage
output (V)Deviation (V) Deviation/Average (%)
177.9289 0.003205 0.002747 85.72
333.6167 0.025786 0.003971 15.4
489.3044 0.046692 0.001831 3.92
600.51 0.060883 0.002747 4.51
867.4033 0.106049 0.002747 2.59
1467.913 0.204773 0.003662 1.79
2313.075 0.344696 0.004578 1.33
4425.981 0.693187 0.003927 0.67
17603.84 2.782593 0.001831 0.06
Table 5.2 Strain gage output variations at constant load
From Table 5.2 the strain gage variations at each constant load range from 0.002 V to
0.005 V. This strain gage output variations do not depend on the magnitude of the tensile
force. That is, the strain gage force can have variation around 0.003V at any load
conditions. At low load the measured strain gage values can have more errors than that of
139
high load. From Table 5.2 the strain gage variation at each load can be changed from
0.06% to 85.72% of the measured load. Therefore, the strain gage variation can affect the
accuracy of the connecting rod force at low load condition. The measured strain gage
variations can be converted to force and show the force variation around from 10N to
40N at tensile test. That is, the measured connecting rod force can have variations about
10N ~ 40N at each crank angle.
5.3.3 Possible error sources in the instantaneous IMEP method
Even if the strain gage itself can have a variation about 30N at every crank angle,
the sign problem of the friction force at expansion stroke cannot be explained only using
the strain gage resolution problem. The measured friction forces show the different errors
according to the measurement condition. For cold motoring (Figure 4.16), the signs are
pretty good but the cross-overs are a bit off, and the results seem to improve with
increasing engine speed. For hot motoring of the baseline engine (Figure 4.20), the cross-
overs are pretty good and the signs are good except for mid-expansion. That is, the
friction force at hot motoring shows a problem only at the expansion stroke. This cannot
be explained by the resolutions of the strain gage and the pressure transducer. The main
difference between the cold motoring and the hot motoring is the oil viscosity. The oil
viscosity at cold state is much higher than that of hot state. Thus, it can be deduced that
the hydrodynamic friction force between two states would be different and can affect the
piston dynamics such as the piston slap and the piston secondary motion. These piston
dynamics can affect the strain gage measurement values and mislead the calculation of
the piston friction force. Another possible error in friction forces during the motoring is
from the piston pin friction. In the instantaneous IMEP method, the piston pin friction
force was neglected for calculating the piston friction force from the measured pressure
force, connecting rod force, and the inertia force. The piston pin exerts the friction force
to the piston assembly and affects the piston friction force and dynamics. However, it is
very difficult to measure the piston pin friction force. For hot motoring of the RLE at
1200 rpm (Figure 4.21), the cross-overs are pretty good except end of intake/beginning of
compression, signs are good for the baseline engine except for mid-expansion, late
140
intake/early compression, and has a bad sign for most of the exhaust stroke. For the RLE
motoring, the piston dynamics are totally different from that of the baseline engine.
Therefore, the measured friction force of the RLE has the different characteristics. For
firing of the baseline engine (Figure 4.39), the cross-overs are close but not exact, and
there are sign problems: 1) last half (or so) of expansion, 2) first one-third to one-half of
intake, and 3) early exhaust. In firing condition the piezoelectric pressure transducer can
cause the errors in calculating the friction forces in addition to the strain gage
measurement. In this research the operating condition of the firing test was set to the full
load. In general the piezoelectric pressure transducer without the cooling passage can
experience the thermal drift during the measurement at high load condition. Thus, the
thermal drift of the pressure transducer is another error source in friction force
measurement during the firing.
The friction force measurement using the instantaneous IMEP method shows the
possibility for measuring the piston friction force easily compared with the floating liner
method. The measured friction forces show the reasonable values except the sign
problems at several crank positions. The sign problems of the measured friction force
cannot be fully explained only using the resolution of the strain gage and the pressure
transducer. The sign problems are mainly occurring near at expansion stroke during the
motoring and the firing. This means that the strain gage measurement is connected with
the piston dynamics. Therefore, the piston dynamics including piston slaps and the piston
secondary motions should be analyzed for better piston friction force measurement using
the instantaneous IMEP method.
141
Chapter 6. Friction Force Calculations Using RINGPAK
6.1 Introduction
RINGPAK is an advanced CAE (Computer Aided Engineering) tool for the
design and analysis of piston ring packs in internal combustion engines, developed by
Ricardo Co. RINGPAK is widely used commercial software to investigate the various
physical phenomena associated with piston ring operations as shown in Figure 6.1.
Inter-ring gas dynamics
Radial ring motion
Axial ring motion
Ring toroidal twist
Oil consumption from throw-off
Oil consumption from evaporation
ring-liner interfacesMixed lubricationat
oil mixed in blow-backConsumption due to
gas
transportLiner oil
Effect of distorted
conformancebore on ring
Figure 6.1. Primary phenomena associated with a piston ring pack.
RINGPAK has been developed utilizing a completely integrated approach using various
sub-models. The sub-models include:
1) Ring axial and twist dynamics
2) Inter-ring gas dynamics
3) Ring radial dynamics and mixed lubrication at the ring-liner interface
4) Liner oil transport
142
5) Oil consumption
6) Ring-face/liner and ring/groove wear
7) Ring conformance to the distorted bore
In this chapter, the measured frictional loss of the piston assembly of the baseline
engine will be compared to the simulated results using RINGPAK. RINGPAK provides
the cycle-averaged results related to ring pack performance, such as:
a) Friction and power loss
b) Gas blow-by and blow-back
c) Approximate oil consumption and
d) Approximate wear rates for the ring faces, groove-ring side faces and liner.
Therefore, through the use of the RINGPAK software, we can understand the effects of
speed, load, and other operating conditions on the friction mechanism in some detail.
6.2 Details on RINGPAK models
Piston ring performance controls friction, power loss, blow-by, oil consumption,
wear, and so on in the internal combustion engine. Thus, their parameters are of interest
and importance due to their impact on engine performance, efficiency, emissions, and
durability. The ring pack system is not fully understood due to its complexity in spite of
its importance. Basically the ring pack system involves the interactions of various
phenomena such as ring axial and radial motions, ring twist, gas flow through end gaps
and ring-groove side-clearances, ring-bore conformability, hydrodynamic and boundary
lubrication, oil transport, wear and oil consumption. The detailed sub-models used in
RINGPAK program are introduced and explained in the following sub-sections.
143
6.2.1 Ring dynamics
In this subsection, axial ring motions and ring twist are modeled.
1) Axial ring motions
Axial ring motions within the grooves are important characteristics in the ring
pack operations because they determine the piston ring sealing capabilities. The ring
sealing capabilities influences the blow-by gas which leaks from the combustion chamber
to the crankcase and the oil consumption. Additionally axial ring motions affected the
wear of contacting surfaces between the ring side faces and grooves. Figure 6.2 is a
schematic of ring motion and associated force and moment components.
Fa,ine and Ra,ine
Center of mass
x=x2
Pback
x
Fa,asp and Ra,asp
x=x1
hc
h(x)
x=x1
Ra,rad
Fa,frc and Ra,frc
Pdown
tdown
Fa,gas and Ra,gas
Pup
tup
Figure 6.2. Schematic of ring motion and associated force and moment components.
From Figure 6.2 the axial force balance can be expressed by
0,,,,, =−+++ ineaaspamixafrcagasa FFFFF (6.1)
144
gasaF , , acting on the ring due to a differential in land pressures above and below the ring
is expressed as
[ ] BtPtPF upupdowndowngasa π)()(, −= (6.2)
where:
B: bore diameter
P: land pressure
t: land-liner clearance
frcaF , , the friction force component at the ring face-liner interface is calculated by the
ring-liner lubrication model. mixaF , , the axial force component due to a mixture of oil and
gas is calculated based on solution of the Reynolds Equation in the upper and lower ring-
groove clearance regions. That is:
t
xh
x
Pxh
x mix ∂∂=
∂∂
∂∂ )(
12
)( 3
µ (6.3)
Boundary conditions are as follows in the upper and lower ring groove regions:
backPP = at 1xx = , upPP = at 2xx = : Upper ring-groove region (6.4)
backPP = at 1xx = , downPP = at 2xx = : Lower ring-groove region
The resultant axial force due to the mixtures is
BPdxPdxF up
x
xdown
x
xmixa π
−= ∫∫ )()(2
1
2
1, (6-5)
aspaF , is applied due to the increase of surface roughness when the ring groove side
clearance, h(x), becomes extremely small. The contact pressure aspP is calculated using
the Greenwood-Tripp model [42].
145
))(
()(15
216 2
σβσσβηπ xh
FEPasp = where ∫∞ −−=0
25.2 )
2exp()(
2
1)( ds
sxsxF π (6.6)
Thus,
BdxPdxPF up
x
x aspdown
x
x aspaspa π
−= ∫∫ )()(2
1
2
1, (6.7)
where:
σ: mean asperity height
β: asperity radius of curvature
η: asperity density
E: composite elastic modulus of the contacting materials
ineaF , , the axial force component due to inertia associated with the ring and piston motion
is given by
)(2
2
, pistoncg
ringinea At
hmF +∂
∂= (6-8)
where:
ringm : mass of the ring
cgh : instantaneous location of the ring within the groove
pistonA : acceleration of the piston
2) Ring twist
Ring twist motion has also its effect on the sealing and the scraping action at the
ring face-liner conjunction. The ring face scraping action influences the liner oil transport
and lubricating oil consumption. Ring twist can be calculated based on the moment
balance applied on the center of gravity of the ring.
146
0,,,,,, =−++++ ineaaspamixaradafrcagasa RRRRRR (6-9)
The moments in Equation 6.9 can be calculated using the axial force components in
Equation 6.1. The Rotational inertia ineaR , is expressed by
ααringringinea K
tIR +∂
∂=2
2
, (6-10)
where:
ringI : moment of inertia of the ring
ringK : ring cross-sectional torsional stiffness
α : ring twist
6.2.2 Inter-ring dynamics
In this subsection the blow-by and blow-back is calculated using the gas dynamics
model which has all the land and groove sub-volumes and performs instantaneous gas
mass balance for each sub-volume.
1) Governing equations and flow models
The blow-by and blow-back over an engine cycle can be calculated using a gas
dynamics model which assembles all the land and groove sub-volumes and performs
instantaneous gas mass balances for each sub-volume.
147
mdrho
m face+gap m m+ conf
belowm
abovem
Figure 6.3. Schematic of the various flow passages around a ring.
Figure 6.3 shows the various passages for gas flow around a ring. The passages between
adjacent lands are the end gap ( gapm& ), the ring face-liner clearance ( facem& ) when the ring
radially lifts out of the oil film on the liner and the flow through areas ( confm& ) generated
by the non-conformance of the ring to a distorted bore. Additionally the gas flow between
the lands and grooves through the ring-groove clearance regions are mass flow rates
abovem& and belowm& . The gas dynamics model for the ring pack assembly is as follows:
iaboveiconfifaceigapibelowiconfifaceigapiland mmmmmmmm
dt
dM,,,,1,1,1,1,
, &&&&&&&& −−−−+++= −−−− (6-11)
idrhoibelowiaboveigroove mmm
dt
dM,,,
, &&& −−= (6-12)
where ilandM , and igrooveM , are the mass of gas of the ith land and groove, respectively.
The instantaneous pressures in the land and groove regions are calculated by the use of
the ideal gas equation of state:
148
iland
ilandilandiland V
RTMP
,
,,, = (6-13)
igroove
igrooveigrooveigroove V
RTMP
,
,,, = (6-14)
where:
ilandP , , igrooveP , : pressure associated with the ith land and groove
R : gas constant
ilandT , , igrooveT , : area weighted average temperature at each volume
The sub-volumes such as ilandV , and igrooveV , are calculated from the land-liner
clearance profiles dependent on the land diameters and bore profile and ring-groove
clearances from the inner groove diameters. The gas mass flow rates ( gapm& , facem& , confm& )
between adjacent lands are calculated from the orifice flow equation such that
2/1)1(/12/1
11
2
−
−=
− γγγ
γγ
u
d
u
d
u
ud P
P
P
P
RT
PACm& (6-15)
and
)1(2)1(
2/1
1
2 −+
+=
γγ
γγu
ud
RT
PACm& when
1
1
2 −
+≤
γγ
γu
d
P
P(6-16)
where:
dC : discharge coefficient
A: flow area of orifice
uT : upstream gas temperature
uP , dP : upstream and downstream pressures
γ : polytropic exponent
The mass flow rates such as abovem& and belowm& are calculated an equation for isothermal
compressible flow through a narrow channel, such that:
149
Bxh
dx
RT
PPm
mix
du πµ1
3
22
)(24
)(−
−= ∫& (6-17)
where:
B: bore diameter
h(x): ring-groove side clearance distribution
T: average temperature of the sub-volume
mixµ : oil-gas mixture viscosity
2) Blow-by and blow-back of gas flow
Figure (6.4) illustrates the blowby and blowback that happen during the engine
operation.
Blowby to crankcase
mm mdrho + below+ mgapm + face+ conf
m
m
Last Ring/Groove
mbelow+ mgapm + face+
Top Ring/Groove
mabove+ mgapm + face+
Blowback past top ring
conf
Blowback to topland crevice
conf
Figure 6.4. Schematic of blowby and blowback gas flows.
150
High pressure in-cylinder gas tends to flow from the combustion chamber to the
crankcase through the ring pack and is called blow-by gas. As in-cylinder gas pressure
decreases during the late expansion stroke, the pressurized gas in the land and groove
volumes may flow back into the cylinder. This gas mass is call gas blowback. The gas
blowby and blowback flow rates can be calculated using:
∫ ++++cycleT
belowconffacegapdrho dtmmmmm0
)( &&&&& (kg/cycle): blowby to crankcase (6-18)
∫ +++cycleT
conffacegapabove dtmmmm0
)( &&&& (kg/cycle): blowback to topland crevice (6-19)
∫ +++cycleT
belowconffacegap dtmmmm0
)( &&&& (kg/cycle): blowback past the top ring (6-20)
6.2.3 Ring-liner lubrication and radial ring dynamics
In this section the ring-liner lubrication mechanism including the viscous
lubrication and the boundary lubrication is modeled. In order to determine the lubrication
regime the radial ring motion is calculated via a radial force balance.
1) Radial ring motion
The lubrication condition between the ring and the liner is dependent on the
minimum film thickness, which is based on radial ring dynamics. Thus, it is necessary to
understand and calculate the radial ring dynamics. Figure 6.5 shows the forces associated
with the radial ring motion.
151
Ffrc
groF
Ften inF
Fgas
Fasp
oilF
Figure 6.5. Schematic of radial ring motion with the associated force components.
From Figure 6.5, the radial force balance in the ring pack can be expressed as
0=−−−−++ inefrctengrogasaspoil FFFFFFF (6-21)
where:
oilF : Radial force due to oil film pressure
aspF : Radial force due to contact pressure acting on the face due to ring-liner asperity
interaction
gasF : Radial force due to gases acting on the non-lubricated portion of the ring face
groF : Radial force due to groove pressure acting behind the ring
tenF : Radial force due to ring tension
frcF : Ring groove friction force
ineF : Radial inertial force
2) Ring-liner hydrodynamic lubrication
152
The Reynolds equation governing the ring-liner hydrodynamic lubrication
condition is solved using the mass conserving (cavitation) scheme.
t
zh
z
zhV
z
Pzh
z oil ∂∂+∂
∂=
∂∂
∂∂ )()(
212
)( 3
µ (6-22)
where:
upPP = at 1zz = and downPP = at 2zz =)(zh : Clearance profile at the ring face-liner conjunction
V : Piston velocity
oilµ : Oil viscosity
upP , downP : Land pressures above and below the ring
1z , 2z : Lubricated extent of the ring-face
The characteristics of the mass conserving algorithm to solve the Reynolds equation are
as follows
Implementation of the Reynolds boundary condition ( 0=dz
dP) at the point of film
detachment and the JFO (Jakobsson-Floberg-Olsson) boundary condition at the
point of possible oil film re-attachment
Inclusion of the effect of slight compressibility of the lubrication via the bulk
modulus β
ρρβ ∂∂= P
where ρ : Oil density and P : Oil film pressure (6-23)
Incorporation of a variable α
cρρα = , when 1≥α (Oil filled zone) (6-24)
α = Fraction of clearance occupied by oil when 1<α (Cavitation zone)
cρ : Density of cavitation
Expressing the oil film pressure using α and β)1( −+= αβcPP 1≥α (6-25)
153
cPP = 1<α (6-26)
Introduction of a cavitation switch function g
g=1 1≥α (6-27)
g=0 1<α (6-28)
Expressing the Reynolds equation in terms of α
)( nncpc h
tz
m
z
m αρ∂∂=∂
∂+∂∂ &&
(6-29)
-n and n mean the upstream and current nodes, respectively
cm& is the Couette mass flow rate per unit circumferential length
−++−= −−
−−−−− )(2
)1(2 nn
nnnnnnncc hh
gghggh
Vm αρ& (6-30)
pm& is the Poisseuille mass flow rate per unit circumferential length
∆
−−−
= −−
z
gghm nnnn
coil
p
)1()1(
12
3 ααβρµ& (6-31)
From the solution of the oil film pressures, the radial oil force is calculated by
BdxzPFz
zoil π
= ∫ 2
1
)( (6-32)
3) Ring-liner boundary lubrication
When the ring-liner clearances are small, asperities on the opposing surfaces
begin to interact with each other and the lubrication becomes the boundary or mixed
lubrication condition. In order to calculate the contact pressures under boundary
lubrication, the Greenwood-Tripp model is used.
= σβσσβηπ )(
)(15
216 2 zhFEPasp (6-33)
154
dss
xsxF ∫∞
−−=
0
25.2
2exp)(
2
1)( π (6-34)
where:
σ : Mean asperity height
β : Radius of curvature of the asperity
η : Asperity density per unit surface area
E : Composite elastic modulus of the contacting materials.
Therefore the radial force on the ring face under the boundary lubrication condition is
computed by
BdzzPFt
aspasp π
= ∫0 )( (6-35)
where t is the axial thickness of the ring face
4) Ring-liner friction and power losses
Using the instantaneous oil film and contact pressure distribution the ring-liner
friction force can be calculated using
∫ ∫
∂∂−=
A A
oilhyd dA
z
PzhVdA
zhFR
2
)(
)(
µ(6-36)
∫=A
aspbdy dACPFR (6-37)
Thus, the total friction force is
bdyhydfrca FRFRF +=, (6-38)
where C is the friction coefficient for the ring-liner interface. Power losses due to ring-
liner friction are given by
155
∫∫
∂∂+=
AA
hyd dAz
PzhdAV
zhL
232
12
)(
)( µµ
(6-39)
∫=A
aspbdy dAVCPL (6-40)
6.2.4 Liner oil transport
The liner oil transport model has four functions
Calculating the instantaneous amount of oil available ( enh ) for lubricating each
ring based on the liner oil film profile
Calculating the volume of oil accumulating at the leading edge of the ring which
contributes to throw-off consumption
Computing the instantaneous oil film thickness ( exh ) trailing each ring
Generating the liner oil film profile at the end of each time step using
instantaneous values of ( exh )
The oil transport model determines the lubrication regimes such as fully flooded, partially
flooded and fully starved lubrication. The features of each lubrication regime are as
follows:
a. Fully flooded ring: The entire ring face is lubricated by oil and the loads are borne
by the oil film and asperity forces.
b. Partially flooded ring: A fraction of the ring face is lubricated by oil and the
remaining portion is under gas lubrication. The loads are borne by oil/asperity
forces and gas forces.
c. Completely starved ring: Due to low lubricant availability on the liner or a high
pressure gradient across the ring face, the oil film may detach from the ring face.
The loads are supported by the gases or gas/asperity forces.
156
6.2.5 Oil consumption mechanisms
It is generally accepted that three main mechanisms of oil consumption
(evaporation, throw-off at the top ring, and flow back to the combustion chamber) can be
attributed to the ring pack system. Each is discussed in the following subsections.
1) Oil evaporation
The oil consumption due to evaporation is based on the oil film thickness distribution
on the portion of the liner which is exposed to the high temperature combustion chamber
gas. The average rate of evaporation of oil from the cylinder surface is computed via
integration/summation over a) time, b) space and c) oil constituents.
∫∫∑= dxdttxRFtxmT
E ie ),(2),(1
, π& (6-41)
where:
E : Average oil evaporation rate
),(, txm ie& : Local instantaneous mass flux of evaporation of an oil constituent (i) at axial
location (x) on the liner at time (t)
R : Cylinder radius
F(x,t) : Weighting factor (0 to 1) indicating if location (x) is covered by the piston or
exposed to the cylinder gases at time t
T : Engine cycle period
The liner/cylinder surface is divided into a number of axial zones. It is assumed that there
is no heat or mass transport between the zones and that the thickness of the liner is small
compared to the bore diameter. Under these assumptions, the heat and mass transport
between the cylinder gas and liner surface can be regarded as one dimensional. In order
to account for the presence of compounds of varying volatility within the oil, the oil is
represented by a number of component species each of which has a different normal
157
boiling point. It is assumed that the evaporative mass flux of the oil is the sum of the
evaporative mass fluxes of each of the fractions.
hg(Tg-To)
(Pg,Tg,hg)Gas side Coolant side
Oil film thickness
-Qf
efgh m
To Tl
Liner thickness
(Tc,hc)
Figure 6.6. Cross-section of the gas-oil film-liner-coolant system at an arbitrary axial
location.
In Figure 6-6 the gas side boundary condition comprises the instantaneous values
of the gas temperature and pressure ),( gg PT and convective heat transfer coefficient
( )gh . On the other side of the liner/cylinder a prescribed fixed coolant temperature ( )cT
and coolant side heat transfer coefficient ( )ch are used. The net heat flux penetration into
the oil film can be obtained by a heat balance at the film interface,
0)( =−−− fefgogg QmhTTh & (6-42)
where:
oT : Gas/oil interface temperature
fgh : Enthalpy of evaporation
em& : Mass flux of evaporation
158
fQ : Net heat flux which penetrates the oil film
A one-dimensional transient conduction equation is solved utilizing )( fQ and the oil side
boundary condition to calculate the temperature distribution within the oil film and liner
and the interface temperature ( )oT .
2
2
x
T
t
T
∂∂=∂
∂ α (6-43)
where:
t : Time variable
α : Thermal diffusivity
x : Spatial variable
If the interface temperature 0T is less than the boiling temperature or the vapor pressure
is less than the ambient pressure, the evaporation flux ( )em& is diffusion limited and can be
calculated by solving a diffusion equation. For binary diffusion from a planar surface into
a gas stream a closed form solution of the diffusion equation exists.
d
ge
Dm δ
βρ )1log( +=& (6-44)
s
s
Y
YY
−−= ∞
1β (6-45)
where:
D : Diffusion coefficient
gρ : Gas density
dδ : Diffusion boundary layer thickness
Y : Mass fraction of the diffusing component (discussed below)
159
The subscripts ( )∞ and (s) pertain to the ambient and gas/oil interface. The diffusion
boundary layer thickness is estimated using the heat and momentum transfer analogy
(Colburn Analogy).
pc
gf CU
hC ρ
667.0Pr2= (6-46)
where:
fC : Skin friction coefficient
Pr : Prandtl number of the in-cylinder gas
cU : Characteristic gas velocity
pC : Constant pressure specific heat of the cylinder gas
The characteristic gas velocity is modeled as
221 psc VUKU += (6-47)
where:
1K : Constant
sU : Swirl component of the in-cylinder velocity
pV : Instantaneous piston velocity
To relate the skin friction coefficient to the Reynolds number, a correlation valid for flow
over a flat plate or for a fully developed pipe flow is used. This correlation has the form
42 )(Re fCK= (6-48)
where:
2K : Constant
Re : Reynolds number based on the momentum boundary layer thickness.
The diffusion boundary layer thickness can be calculated using
160
3
=
d
mSc δδ
(6-49)
where:
Sc : Schmidt number
mδ : Momentum boundary layer thickness
In Equation 6.45, the ambient mass fraction of the diffusing gas is assumed to be
negligibly small, whereas the saturation mass fraction (at the interface) is calculated by
−+
=
l
g
s
s
W
W
P
PY
11
1(6-50)
where:
sP : Saturation pressure at the interface temperature ( )oT
gW : Molecular weight of the cylinder gas
lW : Molecular weight of the diffusing phase (oil vapor)
The heat of vaporization in Equation 6.42 is computed from the following relation.
−×=
l
ofg
Th ρ
95.93910093.2 5 (6-51)
where lρ is the density of the oil film
2) Oil throw-off from inertia
During the up-stroke of the piston the scraping effect of the top ring is responsible
for oil accumulation at the leading edge and a fraction of this volume of oil is discharged
towards the combustion chamber due to inertia effects at TDC reversal positions. This
consumption mode indicates coupling between ring lubrication conditions (fully flooded,
161
partially flooded or starved ring face), liner oil transport and operating conditions
(cylinder pressure and piston velocity).
3) Oil entrainment in blow-back gases
During engine operation, oil present in the ring belt regions gets entrained into the
flowing gas. The oil for this mode of consumption comes from several sources such as oil
film on the liner, oil at the end gap and leading and trailing edges of the rings and oil
trapped between the asperities of the groove surfaces. For annular flow, the entrainment
rate is given by a correlation
316.0
22
,5 )(1075.5
−×= −
σρρg
fcritffg
DGGGEn (6-52)
where:
D: Bore diameter
fρ : Oil film density
gρ : Gas density
σ : Surface tension
gG : Gas mass velocity
fG : Liquid film surface velocity
critfG , : Critical oil film flow rate for the onset of entrainment
critfG , is calculated from a critical Reynolds number given by
+=
5.0
4249.08504.5expRef
g
f
gcrit ρ
ρµµ
(6-53)
where:
fµ : Oil film viscosity
gµ : Gas viscosity
162
6.3 Input data for RINGPAK simulations
In this section the input parameters required for running the RINGPAK program
are defined and specified of the baseline engine.
6.3.1 Input parameters
Piston ring assembly performance of the baseline engine was evaluated using
RINGPAK. The input data for the RINGPAK simulations were inserted via RAPID.
RAPID is a graphical user interface that allows the user to build, edit, import, and
exchange data and execute simulations of the cylinder kit. The cylinder kit may include a
cylinder bore, a piston, a wrist pin and a connecting rod. The piston may include a crown
and a skirt. The terminology connected with the piston ring assembly should be clarified
for RINGPAK input data. Figures 6.7 and 6.8 show a typical piston and piston ring
geometry.
163
Figure 6.7. Piston configuration.
Figure 6.8. Ring configuration.
164
Piston Piston ringA Ring land Free gapB Heat dam Compressed gapC Compression height Radial wall thicknessD Ring belt Ring diameterE Piston head Inside diameterF Piston pin Ring sidesG Skirt Ring faceH Major thrust face Side clearanceI Minor thrust face Ring widthJ Piston pin bushing Torsional twistK Back clearanceLMN Scuff bandO Groove depthP Groove root diameterQ Land diameterR Land clearanceS Skirt clearanceT Skirt grooveU Pin bore offsetV Groove Spacer
Table 6.1 Piston and piston ring terminology
Table 6.1 represents the terminology for the piston and the piston ring. Table (6.2) is the
real piston and piston ring input data for the baseline engine for the RINGPAK
simulations.
165
Cylinder liner Bore (mm) 92
Length (mm) 137
Thickness (mm) 4.93
Piston Top land width/clearance (mm) 5.969/0.6752
Second land width/clearance (mm) 4.064/0.6752
Third land width/clearance (mm) 2.921/0.612
1st compression ring End gap (mm) 0.3
Side clearance (mm) 0.06
Back clearance (mm) 0.76
Width (mm) 1.19
Rail thickness (mm) 3.55
Tension (N) 9.5
Mass (g) 9.85
Twist angle (deg) 0
2nd compression ring End gap (mm) 0.4
Side clearance (mm) 0.05
Back clearance (mm) 0.6
Width (mm) 1.49
Rail thickness (mm) 3.82
Tension (N) 10
Mass (g) 10
Twist angle (deg) 0
Oil ring End gap (mm) 0.76
Side clearance (mm) 0.06
Back clearance (mm) 1.35
Width (mm) 3
Rail thickness (mm) 3
Tension (N) 15
Mass (g) 8.41
Table 6.2 Base RINGPAK input data for baseline engine
166
6.3.2 Engine operating condition
For this dissertation research, piston/ring assembly friction was simulated under
motoring and firing conditions using RINGPAK. The detailed simulation conditions are
shown in Table 6.3.
Hot motoring Speed (rpm) 500, 800, 1200, 1600, 2000
Load (%) 100
Oil temperature (°C) 90
Firing Speed (rpm) 800, 1200, 1600, 2000
Load (%) 100
Oil temperature (°C) 90
Table 6.3 Engine operating conditions for the present RINGPAK simulations
6.4 Simulation results
This section is composed of three subsections. The first two sections represent the
simulation results of the baseline engine at the specified operation condition (hot
motoring and firing condition). All test conditions are based on WOT operating
condition. The last subsection deals with the parametric study of the input data such as
ring tension and surface roughness. Through the parametric study of those input data, the
effects of the ring tension and the surface roughness on the piston friction are shown in
the last subsection.
6.4.1 Hot motoring friction results
Figures 6.9 ~ 6.13 show the simulation results from RINGPAK for the hot motoring
condition. Each figure presents the results for the friction associated with each ring (top
ring, second ring and oil ring) and the total friction. From these graphs, it can be shown
that in the case of hot motoring, most of the friction loss at compression TDC is due to
the top ring. Additionally, the top ring friction is not strongly affected by engine speed.
167
As the engine speed increases from 500 to 2000 rpm, the hydrodynamic friction losses
are responsible for a greater portion of the total friction loss than at lower engine speed.
At 500 rpm and 800 rpm, most of the friction in the mid-stroke between TDC and BDC is
from the boundary and mixed lubrication. However, as the engine speed increases, the
boundary and mixed friction region are reduced and the portion of the stroke that enjoys
hydrodynamic lubrication is increasing. The boundary and mixed lubrication friction loss
at low speed is mainly caused by oil ring friction. The oil ring friction is decreased as the
engine speed increases.
-180 0 180 360 540-30
-20
-10
0
10
20
30 Tota l friction H ydrodynam ic friction
Fric
tion
forc
e (N
)
C rank angle (deg)
-180 0 180 360 540-20
-10
0
10
20 Top ring Second ring O il ring
Fric
tion
forc
e (N
)
C rank angle (deg)
Figure 6.9. Predicted piston ring friction at 500 rpm for hot motoring conditions.
168
-180 0 180 360 540-30
-20
-10
0
10
20
30 Tota l friction H ydrodynam ic friction
Fric
tion
forc
e (N
)
C rank angle (deg)
-180 0 180 360 540-20
-10
0
10
20 Top ring Second ring O il ring
Fric
tion
forc
e (N
)
C rank angle (deg)
Figure 6.10. Predicted piston ring friction at 800 rpm for hot motoring conditions.
-180 0 180 360 540-30
-20
-10
0
10
20
30 Total friction H ydrodynam ic friciton
Fric
tion
forc
e (N
)
C rank angle (deg)
-180 0 180 360 540-20
-10
0
10
20 Top ring Second ring O il ring
Fric
tion
forc
e (N
)
C rank angle (deg)
Figure 6.11. Predicted piston ring friction at 1200 rpm for hot motoring conditions.
169
-180 0 180 360 540-30
-20
-10
0
10
20
30 Tota l fric tion H ydrodynam ic fric tion
Fric
tion
forc
e (N
)
C rank angle (deg)
-180 0 180 360 540-30
-20
-10
0
10
20
30 Top ring Second ring O il ring
Fric
tion
forc
e (N
)
C rank angle (deg)
Figure 6.12. Predicted piston ring friction at 1600 rpm for hot motoring conditions.
-180 0 180 360 540-30
-20
-10
0
10
20
30 Total friction H ydrodynam ic friction
Fric
tion
forc
e (N
)
C rank angle (deg)
-180 0 180 360 540-30
-20
-10
0
10
20
30 Top ring Second ring O il ring
Fric
tion
forc
e (N
)
C rank angle (deg)
Figure 6.13. Predicted piston ring friction at 2000 rpm for hot motoring conditions.
170
-180 0 180 360 540-30
-20
-10
0
10
20
30
Fric
tion
forc
e (N
)
Crank angle (deg)
500 rpm 800 rpm 1200 rpm 1600 rpm 2000 rpm
Figure 6.14. Effects of engine speed on the total piston assembly friction for hot
motoring conditions.
Figure 6.14 represents the total friction loss variation throughout the cycle for a range
of engine speed. At low speed, most of the friction force is from the boundary and mixed
lubrication of the top ring and the oil ring. As the engine speed increases, boundary
lubrication friction force due to the top ring is not much changed, but the boundary
friction force from the oil ring is decreasing.
6.4.2 Firing friction results
The piston ring assembly friction for the firing condition is shown in Figures 6.15
~ 6.19. In the firing condition, the peak friction force is observed near the peak cylinder
pressure position. As the cylinder pressure is increased due to combustion, compared to
the hot motoring case, the friction force of the top ring is increased. The high cylinder
pressure pushes the top compressing ring toward the liner and so the friction force
between the top ring and the cylinder liner increases. In the mid stroke, the total friction
force shows the same trend with that for hot motoring as the engine speed increases. That
171
is, at low engine speed the friction force in the mid-stroke is mainly due to the boundary
and mixed lubrication of the oil ring. As the engine speed goes up, the mixed and
boundary lubrication of the oil ring decreases and the hydrodynamic lubrication friction
increases.
-180 0 180 360 540-30
-20
-10
0
10
20
30 Tota l fric tion H ydrodynam ic fric tion
Fric
tion
forc
e (N
)
C rank angle (deg)
-180 0 180 360 540-20
-10
0
10
20 Top ring Second ring O il ring
Fric
tion
forc
e (N
)
C rank angle (deg)
Figure 6.15. Predicted piston ring friction at 800 rpm for WOT firing conditions.
-180 0 180 360 540-30
-20
-10
0
10
20
30
40 Tota l fric iton H ydrodynam ic fric tion
Fric
tion
forc
e (N
)
C rank angle (deg)
-180 0 180 360 540-30
-20
-10
0
10
20
30
40 Top ring Second ring O il ring
Fric
tion
forc
e (N
)
C rank angle (deg)
172
Figure 6.16. Predicted piston ring friction at 1200 rpm for WOT firing conditions.
-180 0 180 360 540-30
-20
-10
0
10
20
30
40
50 Tota l fric tion H ydrodynam ic fric tion
Fric
tion
forc
e (N
)
C rank angle (deg)
-180 0 180 360 540-30
-20
-10
0
10
20
30
40
50 Top ring Second ring O il ring
Fric
tion
forc
e (N
)
C rank angle (deg)
Figure 6.17. Predicted piston ring friction at 1600 rpm for WOT firing conditions.
-180 0 180 360 540-30
-20
-10
0
10
20
30
40
50 Tota l fric tion H ydrodynam ic fric tion
Fric
tion
forc
e (N
)
C rank angle (deg)
-180 0 180 360 540-30
-20
-10
0
10
20
30
40
50 Top ring Second ring O il ring
Fric
tion
forc
e (N
)
C rank angle (deg)
Figure 6.18. Predicted piston ring friction at 2000 rpm for WOT firing conditions.
173
-180 0 180 360 540-30
-20
-10
0
10
20
30
40
50
Fric
tion
forc
e (N
)
Crank angle (deg)
800 rpm 1200 rpm 1600 rpm 2000 rpm
Figure 6.19. Effects of engine speed on the predicted total piston assembly friction for
WOT firing conditions.
The cylinder pressure measured during the baseline engine firing experiments was
used for the RINGPAK simulations. The real cylinder pressure data (ensemble-averaged
pressure) is shown in Figure 4.35. From Figure 6.19 the peak friction force at 1600 rpm is
greater than that at 2000 rpm. This is because the peak cylinder pressure at 1600 rpm is
higher than that at 2000 rpm. Figure 6.20 represents the relation between the peak friction
force and the peak cylinder pressure force. The crank angle at which the peak friction
force is observed is coincident with that of the peak cylinder pressure force. This
coincidence of crank angles between the peak cylinder pressure and the peak friction
force is different from the experimental results. As discussed in Chapter 4, I believe that
the crank angle offset observed in the experimental data is mainly due to the oil squeeze
film effect. Therefore, either the RINGPAK simulation does not reflect the physical
phenomena of the squeeze film effect or there is a problem with the experimental data.
174
0 20 40 60 80 100-30
-20
-10
0
10
20
30
40
50
0
5000
10000
15000
20000
25000
30000
35000
40000
Fric
tion
forc
e (N
)
Crank angle (deg)
800 rpm 1200 rpm 1600 rpm 2000 rpm
Pre
ssur
e fo
rce
(N)
Figure 6.20. Predicted crank angles at peak pressure and friction forces.
6.4.3 Parametric study
The piston ring design factors which can affect the friction force include piston
ring tension, ring width, end gap size, surface roughness and stiffness, ring cross section,
and so on. In this research the effects of ring tension and surface roughness were
investigated using RINGPAK simulation.
1) Effect of ring tension
Piston ring tension is one of the most important factors which can influence piston
ring friction, oil consumption, blow-by, etc. In this research, the main concern is about
the ring tension effect on piston assembly friction. Figures 6.21 ~ 6.25 show the
simulation results for 100 % higher ring tension compared with that of the baseline
engine. As the ring tension increases, the boundary and mixed lubrication friction are
responsible for a larger portion of the total friction forces than that of the baseline engine.
175
-180 0 180 360 540-40
-30
-20
-10
0
10
20
30
40 Total fric tion Hydrodynam ic friction
Fric
dtio
n fo
rce
(N)
C rank angle (deg)
-180 0 180 360 540-20
-10
0
10
20 Top ring Second ring O il ring
Fric
iton
forc
e (N
)
C rank angle (deg)
Figure 6.21. Predicted effects of high ring tension on piston ring friction under hot
motoring conditions at 500 rpm.
-180 0 180 360 540-40
-30
-20
-10
0
10
20
30
40 Tota l fric tion H ydrodynam ic fric tion
Fric
tion
forc
e (N
)
C rank angle (deg)
-180 0 180 360 540-20
-10
0
10
20 Top ring Second ring O il ring
Fric
tion
forc
e (N
)
C rank angle (deg)
Figure 6.22. Predicted effects of high ring tension on piston ring friction under hot
motoring conditions at 800 rpm.
176
-180 0 180 360 540-40
-30
-20
-10
0
10
20
30
40 Tota l fric tion H ydrodynam ic fric tion
Fric
tion
forc
e (N
)
C rank angle (deg)
-180 0 180 360 540-30
-20
-10
0
10
20
30 Top ring Second ring O il ring
Fric
tion
forc
e (N
)
C rank angle (deg)
Figure 6.23. Predicted effects of high ring tension on piston ring friction under hot
motoring condition at 1200 rpm.
-180 0 180 360 540-40
-30
-20
-10
0
10
20
30
40 Tota l fric tion H ydrodynam ic fric tion
Fric
tion
forc
e (N
)
C rank angle (deg)
-180 0 180 360 540-30
-20
-10
0
10
20
30 Top ring Second ring O il ring
Fric
tion
forc
e (N
)
C rank angle (deg)
Figure 6.24. Predicted effects of high ring tension on piston ring friction under hot
motoring conditions at 1600 rpm.
177
-180 0 180 360 540-40
-30
-20
-10
0
10
20
30
40 Tota l fric tion H ydrodynam ic fric tion
Fric
tion
forc
e (N
)
C rank angle (deg)
-180 0 180 360 540-30
-20
-10
0
10
20
30 Top ring Second ring O il ring
Fric
tion
forc
e (N
)
C rank angle (deg)
Figure 6.25. Predicted effects of high ring tension on piston ring friction under hot
motoring conditions at 2000 rpm.
The friction force variation at higher ring tension for the various engine speeds is
also shown in Figure 6.26. In Figure 6.26 the black and the red line (thick line) indicate
the friction force with high ring tension. The green and the blue line (thin line) are for the
friction force of the baseline condition. The friction force differences between the base
condition and the high tension ring in the mid-stoke are huge at 500 rpm. At low engine
speed the ring tension has a greater effect on the friction force since the boundary and
mixed lubrication are dominant at low engine speed. As the engine speed is increasing,
the ring tension effect on the friction force becomes smaller than that for low engine
speeds.
178
-180 0 180 360 540-40
-30
-20
-10
0
10
20
30
40
Fric
tion
forc
e (N
)
Crank angle (deg)
500 rpm, High tension 2000 rpm, High tension 500 rpm, Base 2000 rpm, Base
Figure 6.26. Comparison of the predictions of piston ring friction between the baseline
and the high ring tension over a range of engine speeds.
2) Effect of surface roughness
The boundary lubrication model in RINGPAK requires detailed input of surface
roughness characteristics such as asperity height, asperity radius of curvature and asperity
density of contacting surfaces. In the Greenwood-Tripp model the boundary lubrication
parameters used for boundary lubrication are σβη and βσ
, where σ is the composite
asperity height of the contact surfaces, β is the asperity radius of curvature and η is the
asperity density. Based on a data-base for general engineering surfaces the following
relationships are suggested for general cases.
01.0 < σβη < 0.05
179
0.01 < βσ
< 0.1
For the present RINGPAK simulations, the following boundary lubrication parameters
were used for the base condition.
σ = 2.5E-7 m, β = 0.005 m, η = 6E7 / m2
σβη = 0.075, βσ
=.00707
In order to examine the boundary lubrication parameter effects on piston ring friction
forces, two different parameters were simulated.
Case 1 : σβη = 0.015, βσ
=.016
Case 2 : σβη = 0.0075, βσ
=.022
Specifically, for Case 1 and Case 2 the asperity radius of curvature was changed from
0.005 m to 0.001 m (decreased by a factor of 5) and 0.0005 m (decreased by a factor of
10).
180
-180 0 180 360 540-30
-20
-10
0
10
20
30 Tota l friction H ydrodynam ic friction
Fric
tion
forc
e (N
)
C rank angle (deg)
-180 0 180 360 540-20
-10
0
10
20 Top ring Second ring O il ring
Fric
tion
forc
e (N
)
C rank angle (deg)
Figure 6.27. Predicted effects of decreasing the asperity radius of curvature by a factor of
5 (Case 1) on piston ring friction under hot motoring conditions at 500 rpm.
-180 0 180 360 540-30
-20
-10
0
10
20
30 Tota l fric iton H ydrodynam ic fric tion
Fric
tion
forc
e (N
)
C rank angle (deg)
-180 0 180 360 540-20
-10
0
10
20 Top ring Second ring O il ring
Fric
tion
forc
e (N
)
C rank angle (deg)
Figure 6.28. Predicted effects of decreasing the asperity radius of curvature by a factor of
5 (Case 1) on piston ring friction under hot motoring conditions at 800 rpm.
181
-180 0 180 360 540-20
-10
0
10
20 Tota l fric tion H ydrodynam ic fric tion
Fric
tion
forc
e (N
)
C rank angle (deg)
-180 0 180 360 540-20
-10
0
10
20 Top ring Second ring O il ring
Fric
tion
forc
e (N
)
C rank angle (deg)
Figure 6.29. Predictions of the effects of decreasing the asperity radius of curvature by a
factor of 5 (Case 1) on piston ring friction under hot motoring conditions at 1200 rpm.
-180 0 180 360 540-20
-10
0
10
20 Tota l fric tion H ydrodynam ic fric tion
Fric
tion
forc
e (N
)
C rank angle (deg)
-180 0 180 360 540-20
-10
0
10
20 Top ring Second ring O il ring
Fric
tion
forc
e (N
)
C rank angle (deg)
182
Figure 6.30. Predicted effects of decreasing the asperity radius of curvature by a factor of
5 (Case 1) on piston ring friction under hot motoring conditions at 1600 rpm.
-180 0 180 360 540-20
-10
0
10
20 Tota l fric tion H ydrodynam ic fric tion
Fric
tion
forc
e (N
)
C rank angle (deg)
-180 0 180 360 540-20
-10
0
10
20 Top ring Second ring O il ring
Fric
tion
forc
e (N
)
C rank angle (deg)
Figure 6.31. Predicted effects of decreasing the asperity radius of curvature by a factor of
5 (Case 1) on piston ring friction under hot motoring conditions at 2000 rpm.
-180 0 180 360 540-30
-20
-10
0
10
20
30
Fric
tion
forc
e (N
)
Crank angle (deg)
500 rpm, Case1 2000 rpm, Case1 500 rpm, Base 2000 rpm, Base
183
Figure 6.32. Friction force comparison between Case 1 (asperity radius decreased by a
factor of 5) and the baseline asperity radius of curvature.
-180 0 180 360 540-30
-20
-10
0
10
20
30
Fric
tion
forc
e (N
)
Crank angle (deg)
500 rpm, Case2 2000 rpm, Case2 500 tpm, Base 2000 rpm, Base
Figure 6.33. Friction force comparison between Case 2 (asperity radius of curvature
decreased by a factor of 10) and the baseline radius of curvature.
Case 1 simulation results at different engine speeds are shown in Figures 6.29 ~
6.31. Figures 6.32 and 6.33 show the effects of surface roughness on the friction force. In
Figure 6.32 the black and the red line (thick line) represent the base condition and the
green and the blue (thin line) is for Case 1. At low speed the effect of smoother surfaces
is effective in overall crank angle. That is, near TDC and BDC the smoother surface is
effective to reduce the boundary lubrication friction. Since the boundary lubrication
friction is dominant near mid stroke at low engine speed, the effect of a smoother surface
is also effective to reduce the friction. However, as the engine speed increases, the
smoother surface shows the same friction force as the base condition although it is
effective near BDC and TDC. Figure 6.33 compares the Case 2 with the base condition.
In Case 2 the surface roughness is even smoother than for Case1. The friction reduction
effects of Case 2 are profound at low engine speeds. At high engine speed, the smoother
surface does not show any benefits during the cycle except near TDC and BDC. Since the
184
main lubrication regime becomes hydrodynamic at high engine speed, the smooth surface
effect is limited to near the TDC and BDC regions at which still the lubrication region
remain as boundary and mixed lubrication due to its low piston speed.
185
Chapter 7. Summary and Conclusions
The Rotating Liner Engine was developed to eliminate the boundary and mixed
lubrication friction between the piston assembly and the cylinder liner. In this study, the
friction forces of the piston assembly in the baseline engine and the RLE have been
measured using three different measurement techniques: 1) direct motoring with tear-
down, 2) the instantaneous IMEP method, and 3) the P-w method. For better
understanding of the friction mechanism and comparison between the experimental
results and theory, the commercial software RINGPAK (Ricardo Software Co. Ltd)
which can simulate the friction loss of the piston ring assembly was used. Through the
application of different friction measurement techniques, the pros and the cons of each
measurement technique were examined and the friction mechanisms of the baseline
engine and the RLE can be better understood. Through the simulation using RINGPAK,
the friction mechanism of the baseline engine can be explained well and the effects of the
engine speed and other piston assembly design factors on the piston ring assembly
friction were found and understood from the perspective of better piston assembly design.
Applying the friction measurement techniques and simulations for the piston assembly
friction on the baseline engine and the RLE, several conclusions can be drawn.
The cycle-averaged motoring torque of the RLE represents a friction reduction of
23~31% compared to the baseline engine (single cylinder version of a 4-cylinder
engine) for hot motoring tests. It is estimated that the friction reduction due to
liner rotation would be 37.5-42.5% with all four pistons for hot motoring
experiments.
Using tear-down tests, it was found that the piston assembly friction of the
baseline engine is reduced by 90% at 1200 rpm and 71% at 2000 rpm through
liner rotation. This reduction corresponds to 64 and 59 kPa of FMEP reduction
respectively. Also, it can be concluded that, through liner rotation, the main
lubrication regime of the piston assembly is changed from mixed and boundary
for the baseline engine to predominantly hydrodynamic for the RLE.
186
The instantaneous IMEP method was successfully applied to both engines and
showed that most of the boundary lubrication friction near TDC in the baseline
engine was eliminated through the liner rotation.
The friction force measurement under firing condition of the baseline engine
showed two friction peaks. One friction peak occurs at compression TDC and the
other peak is observed after TDC and connected with the peak cylinder pressure.
The crank angle of the second peak in the friction force is not exactly coincident
with that of the peak cylinder pressure and has a phase lag around 6° crank angles.
The phase lag between the peak friction force and the peak cylinder pressure is
believed to be attributable to the squeeze effect of the oil film between the piston
ring and the cylinder liner.
The firing tests of the baseline engine showed reasonable values for piston friction
work. The cyclic variation of the piston friction work represents three times
higher than that of the cylinder pressure during firing. More cyclic variations of
the piston friction forces are attributable to the variation of oil film thickness,
dynamic instabilities, and so on. Perhaps most importantly, cycle variations in the
cylinder pressure are strongly nonlinearly related to ring-liner friction via, for
example, the Stribeck diagram.
Through the application of the instantaneous IMEP method and the P-w method,
the dominant friction mechanism during cold motoring is the ring viscous friction
and the skirt friction. In the hot motoring and the firing tests the dominant
mechanisms are the ring viscous and mixed lubrication friction. The RINGPAK
simulations revealed that the skirt friction mechanism in the hot motoring and the
firing tests is negligible.
The simulation results using RINGPAK also showed the friction mechanism
variation accord to the changes of engine operating conditions and design factors
187
of the piston ring. However, the phase lag between the peak friction force and the
peak cylinder pressure were not observed in the RINGPAK simulations.
In this research the P-w method and the instantaneous IMEP method were applied
to compare the piston assembly friction between the baseline engine and the RLE.
It was very difficult to determine the piston friction force via the P-w method due
to its simple assumption about the friction components. Although the piston
friction force using the instantaneous IMEP method showed sign problems around
expansion stroke, the instantaneous IMEP method gave much information about
the friction force and the lubrication mechanism of the piston assembly.
188
Chapter 8. Recommendations for Future Work
Through this research it was demonstrated the potential of the RLE to reduce the
piston/ring assembly friction. The first prototype RLE used in this research has showed
some hardware problems such as oil leak, water leak, and so on during its operation.
However, the second generation of the RLE can be better designed and solve most of
these hardware problems. Therefore, future work in this area should be concentrated on
the fundamental friction research which can include the development of friction
measurement techniques and the theoretical analysis of the RLE. Several topics for
further research are listed below.
1) Development of the IMEP method using telemetry
Friction measurement results using the instantaneous IMEP method in this
research showed the good potential for measuring the piston/ring assembly friction.
However, this method needs an improvement for better measurement accuracy and
endurance. The temperature compensation of the strain gage should be reinforced. During
the operation, the temperature in connecting rod might be changed and affect the
sensitivities of the strain gage even in steady state operation. Thus, the temperature in
connecting rod should be monitored and compensated during engine operation. The
calculation of the inertia force of the connecting rod was based on lumped mass analysis
in this research. The inertial force error based on lumped mass analysis could be
negligible at low engine speed. However, as the engine speed increases, the inertial force
based on lumped mass might be different from the real inertia values. Thus, it is needed
to calculate the inertia force of the connecting rod using 3-D dynamic analysis tool and
compare with that of the lumped mass. In this research the flexible flat cable was used to
transmit the strain gage signal to strain gage signal conditioner. This method was good
enough for low speed and short time measurements. However, the flexible flat cable
couldn’t last long at high speed. Therefore, for longer measurement at high engine speed,
it is indispensable to use the telemetry techniques to transmit the strain gage signals.
Since the connecting rod has a harsh condition for strain measurement (high temperature,
189
high rotating speed, etc.), it is still a challenging problem to use telemetry techniques in
measuring the strain of the connecting rod. Even if it would be a quite difficult to use
telemetry, it is attractive to measure the strain gage signals without the wiring.
2) Development of the friction rig for analysis of the RLE
It was difficult to find out the RLE effect on piston/assembly friction using engine
dynamometer tests since it took long time to install and uninstall the RLE. If the test rig
can be designed to simulate the RLE operation, it would be easier to examine and analyze
the RLE effects.
3) Development of micro-pattern of the cylinder liner for better friction
characteristics
Through the RLE engine operation, it was found that the specific micro-pattern
was made in the cylinder liner. Since the cylinder is rotating while the piston is
reciprocating, the specific hatch pattern is generated in the cylinder liner. There are
possibilities in reducing the friction loss due to its specific scratch pattern in the cylinder
liner. Therefore, if this specific pattern could be reproduced in the baseline engine using
3-D lithography techniques, it is possible to assess the friction reduction effect of the
specific scratch pattern. Recently, many researchers have tried to produce a special
surface pattern to reduce the boundary lubrication. Thus, the specific surface pattern in
the RLE could be a clue to remove the boundary lubrication without the liner rotation.
190
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VITA
Myoungjin Kim was born in Seoul, Korea on September 9th, 1966. He lived in
Seoul during his schooldays. He graduated Osan Middle School and Baemoon High
School with honors and entered Seoul National University in 1985. His major was
mechanical engineering and got his Bachelor of Science in 1989. After finishing
undergraduate studies in Seoul National University, he entered Korea Advance Institute
of Science and Technology (KAIST) in 1989 for graduate studies. In KAIST, he studied
the two phase flow and got his Master of Science in 1991. During his studies in KAIST,
LG electronics supported him financially. He started his first career in LG electronics in
1991. In LG electronics, he involved in developing the mechanic machine such as ATM
(Auto-Teller Machine), CD (Cash Dispenser). He worked for LG electronics for three
years and quit in 1994. He started his new career at Hyundai Motor Company in 1994. He
worked at Hyundai Motor Company for six and a half years. He involved in developing
engine hardware, calibrating an ECU, sample engine test, and combustion analysis. He
developed the swirl and tumble measurement rig and compared with 2-Dimensional PIV
and 3-Dimensional PTV water rig. He also compared the characteristics of the intake
flow field with that of the combustion. In 2001, he quit Hyundai Motor Company
temporarily and moved to Unites States of America for pursuing Ph.D. He studied the
internal combustion engine in the University of Texas at Austin and graduated in
summer, 2005. He will teach and research in University of Texas at El Paso as an
assistant professor from fall, 2005.
Address : 1155 Upper Canyon
El Paso , TX 79912
This dissertation was typed by the author