copyright by myoungjin kim 2005

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.


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


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


hot motoring---------------------------------------------------------------------134



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 5.14 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 5.19 Effects of engine speed on the predicted total piston assembly friction for


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


motoring conditions at 800rpm------------------------------------------------169


motoring condition at 1200 rpm-----------------------------------------------169


motoring conditions at 1600 rpm----------------------------------------------170


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



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



rpm--------------------------------------------------------------------------------175



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)


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)


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)


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


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


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


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


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


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



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


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


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


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


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

)


-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

)


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

)


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

)


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)


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)


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



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



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


Fric

tion

forc

e (N

)

C rank angle (deg)

-180 0 180 360 540-20

-10

0

10


Fric

tion

forc

e (N

)

C rank angle (deg)


-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


Fric

tion

forc

e (N

)

C rank angle (deg)


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


Fric

tion

forc

e (N

)

C rank angle (deg)


-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


Fric

tion

forc

e (N

)

C rank angle (deg)


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


Fric

tion

forc

e (N

)

C rank angle (deg)

-180 0 180 360 540-20

-10

0

10


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


Fric

tion

forc

e (N

)

C rank angle (deg)

172


-180 0 180 360 540-30

-20

-10

0

10

20

30

40


Fric

tion

forc

e (N

)

C rank angle (deg)

-180 0 180 360 540-30

-20

-10

0

10

20

30

40


Fric

tion

forc

e (N

)

C rank angle (deg)


-180 0 180 360 540-30

-20

-10

0

10

20

30

40


Fric

tion

forc

e (N

)

C rank angle (deg)

-180 0 180 360 540-30

-20

-10

0

10

20

30

40


Fric

tion

forc

e (N

)

C rank angle (deg)


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


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


Fric

tion

forc

e (N

)

C rank angle (deg)

-180 0 180 360 540-20

-10

0

10


Fric

tion

forc

e (N

)

C rank angle (deg)



176

-180 0 180 360 540-40

-30

-20

-10

0

10

20

30


Fric

tion

forc

e (N

)

C rank angle (deg)

-180 0 180 360 540-30

-20

-10

0

10

20


Fric

tion

forc

e (N

)

C rank angle (deg)


motoring condition at 1200 rpm.

-180 0 180 360 540-40

-30

-20

-10

0

10

20

30


Fric

tion

forc

e (N

)

C rank angle (deg)

-180 0 180 360 540-30

-20

-10

0

10

20


Fric

tion

forc

e (N

)

C rank angle (deg)



177

-180 0 180 360 540-40

-30

-20

-10

0

10

20

30


Fric

tion

forc

e (N

)

C rank angle (deg)

-180 0 180 360 540-30

-20

-10

0

10

20


Fric

tion

forc

e (N

)

C rank angle (deg)



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


Fric

tion

forc

e (N

)

C rank angle (deg)

-180 0 180 360 540-20

-10

0

10


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


Fric

tion

forc

e (N

)

C rank angle (deg)



181

-180 0 180 360 540-20

-10

0

10


Fric

tion

forc

e (N

)

C rank angle (deg)

-180 0 180 360 540-20

-10

0

10


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


Fric

tion

forc

e (N

)

C rank angle (deg)

-180 0 180 360 540-20

-10

0

10


Fric

tion

forc

e (N

)

C rank angle (deg)

182



-180 0 180 360 540-20

-10

0

10


Fric

tion

forc

e (N

)

C rank angle (deg)

-180 0 180 360 540-20

-10

0

10


Fric

tion

forc

e (N

)

C rank angle (deg)



-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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196

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

copyright by myoungjin kim 2005

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