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lIBRARY
IVI Michigan State a
University

 

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thesis entitled

COMPUTATIONAL ANALYSIS OF PISTON RING WEAR AND
OIL CONSUMPTION FOR AN INTERNAL COMBUSTION ENGINE

presented by

BOON-KEAT CHUI

has been accepted towards fulfillment
of the requirements for

Master degree in Mechanical Engineering

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COMPUTATIONAL ANALYSIS OF PISTON RING WEAR AND OIL
CONSUMPTION FOR AN INTERNAL COMBUSTION ENGINE

By

Boon-Keat Chui

A THESIS

Submitted to
Michigan State University
in partial fulfillment of the requirements
for the degree of

MASTER OF SCIENCE

Department of Mechanical Engineering

2001

ABSTRACT

COMPUTATIONAL ANALYSIS OF PISTON RING WEAR AND OIL
CONSUMPTION FOR AN INTERNAL COMBUSTION ENGINE

By

Boon-Keat Chui

Improving the efficiency and reducing the exhaust gas emission have been the
goals in the research of internal combustion engines. This thesis presents the fundamental
theory in computational modeling of lubrication, piston ring wear and oil consumption in
an engine, which play significant roles in influencing the efficiency and emission of the
engine. This thesis also presents the predicted results of a study in which the relationship
between the piston ring wear and the oil consumption of a 4.6L, V8 engine was analyzed.
A computer program, the CASE system (Cylinder kit Analysis System for Engines),
together with WEAR and OILCONSUME models were used to simulate the results for

this thesis.

The predicted results show that the engine load and speed affected the piston ring
wear and the oil consumption significantly. The transient analysis of the piston ring wear
show that the piston ring surface experienced high volume loss due to wear during the first
20 hours simulation and reached steady state value after that time period. The transient
results of the oil consumption, which consists of oil vaporization and reverse flow through
the upper ring gap, did not show significant variation during the 500 hours simulation.
However, blow-by into the crankcase showed a significant increase during the 500 hours

simulation under the effect of wear.

To my Lord and Savior, Jesus Christ,
to my lovely wife, Gylin,
to my grandmother, my parents and family,
to all my friends.

ACKNOWLEDGEMENTS

I would like to express my appreciation for my advisor, Dr. Harold Schock for his
advice, guidance and encouragement toward my progress in completing this thesis. I am
also thankful to Jan Chappell, Bobbie Slider, Xiaobing Liu, Stephen Yen, Mahmood Rahi,

and all in the engine lab who helped me through the completion of my thesis.

I would like to thank my wife, Gylin for her support, encouragement and prayer.
Likewise, thanks to my close friends James and Karen Getz, Rex and Vangie Alocilja for
their support and prayer. Without their support and encouragement, it would not have been

possible for me to succeed in my Master degree program.

TABLE OF CONTENTS

LIST OF TABLES ........................................................................................................... viii
LIST OF FIGURES ........................................................................................................... ix
Chapter 1 INTRODUCTION .............................................................................................. 1
l . 1 Motivation .................................................................................................... 1

1.2 Literature Survey ......................................................................................... 3

1.3 Problem Statement ....................................................................................... 5
Chapter 2 FUNDAMENTALS OF PISTON RING PACK LUBRICATION ..................... 7
2. 1 Introduction .................................................................................................. 7

2.2 Formulating Oil Film Thickness ................................................................ 10

2.2.1 Oil Viscosity ................................................................................... 10

2.2.2 Oil Flow in Piston Ring Pack ........................................................ 13

2.2.3 Reynolds Equation for Ring Pack .................................................. 16

Chapter 3 FUNDAMENTALS OF PISTON RING WEAR ............................................. 20
3. 1 Introduction ................................................................................................ 20

3.2 Formulating Piston Ring Adhesive Wear .................................................. 22

3.2.] Surface Texture of Piston Ring ...................................................... 22

3.2.2 Wear Mechanism ............................................................................ 26

Chapter 4 FUNDAMENTALS OF OIL CONSUMPTION .............................................. 31
4.1 Introduction ................................................................................................ 3 l

4.2 Vaporization of Engine Oil from the Cylinder Liner ................................. 32

TABLE OF CONTENTS

4.3 Reverse Flow of Engine Oil through Top Compression Ring Gap ........... 37
4.3.1 Formulating Reverse Flow Equation ............................................. 40
4.4 Total Oil Consumption ............................................................................... 43

Chapter 5 SIMULATION ANALYSIS OF OIL CONSUMPTION AND PISTON RING

WEAR IN A 4.6L,V8 ENGINE ............................................................................ 45
5. 1 Introduction ................................................................................................ 45
5.2 Integration of Different Computer Models ................................................ 46
5.3 Engine Specification of A 4.6L-8 Cylinder IC Engine .............................. 47
5.4 Simulation Results ..................................................................................... 48
5.5 Results and Discussions ............................................................................. 49
5.5.1 Cylinder Pressures ......................................................................... 49

5.5.2 Oil Film Thickness ......................................................................... 51

5.5.3 Average Blow-by ........................................................................... 55

5.5.4 Piston Ring Wear ........................................................................... 57

5.5.5 Oil Consumption .................................................... . ........................ 75

5.5.6 Oil Consumption and Gas Back Flow ........................................... 80

5.5.7 Oil Consumption and Average Gas Blow-by ................................. 83

5.5.8 Oil Consumption and Piston Ring Wear ........................................ 86
Chapter 6 SUMMARY ..................................................................................................... 89
6. 1 Conclusions ................................................................................................ 89

vi

TABLE OF CONTENTS

6.2 Recommendation ....................................................................................... 91

BIBLIOGRAPHY .............................................................................................................. 93

vii

TABLE 2.]

TABLE 2.2

TABLE 2.3

TABLE 2.4

TABLE 5.1

TABLE 5.2

LIST OF TABLES

Regime of Piston Ring Lubrication ............................................................. 8
SAE Oil Viscosity ...................................................................................... 12
Vogel Constants for SAE Grades Oil ......................................................... 12
Boundary Conditions for Piston Ring Lubrication .................................... 18
Selected Simulation Cases ......................................................................... 46
Engine Data ............................................................................................... 47

viii

FIGURE 1.1

FIGURE 2.1

FIGURE 2.2

FIGURE 2.3

FIGURE 2.4

FIGURE 2.5

FIGURE 2.6

FIGURE 2.7

FIGURE 2.8

FIGURE 3.1

FIGURE 3.2

FIGURE 3.3

FIGURE 3.4

FIGURE 3.5

FIGURE 3.6

FIGURE 3.7

FIGURE 3.8

FIGURE 4.1

FIGURE 4.2

FIGURE 4.3

FIGURE 4.4

FIGURE 4.5

LIST OF FIGURES

Cylinder Kit Components of A Basic Engine .............................................. 6
Comparison of Major Categories of Friction Losses [ 9 ] ........................... 8
Compression Ring with Lubricating Oil Film ............................................ 9
Linear Velocity Distribution ...................................................................... 10
Nonlinear Velocity Distribution ................................................................. 11
Oil Flow Through a Rectangular Slot ........................................................ 13
Parabolic Velocity Distribution .................................................................. 15
Piston Ring Lubrication ............................................................................. 18
Forces Acting on a Piston Ring ................................................................. 19
Rubbing of Two Contact Surface Under Microscopic View ..................... 21
Surface Profile ........................................................................................... 23
Engineered Surface Profile ........................................................................ 24
Surface Representation Using Abbott Firestone Curve ............................. 25
Piston Ring Surface Profile ........................................................................ 25
Piston Ring Sliding on Cylinder Bore ....................................................... 26
Ring Section Contact with Cylinder Liner ................................................ 28
Discretized Piston Ring Surface ................................................................ 29
Oil Consumption Mechanism .................................................................... 31
Vaporization Process at Oil Film ............................................................... 35
Thermal Resistance Model for Oil Film .................................................... 36
Flow Through Piston Ring Gap ................................................................. 38
Different Flow in a Pipe System ................................................................ 39

FIGURE 4.6 Free-Body Diagram of A Cylinder Fluid ................................................... 41

FIGURE 5.1 Identification Numbering for Piston Ring ................................................. 48
FIGURE 5.2 Cylinder Pressure at Different Operating Conditions ................................ 49
FIGURE 5.3 Cylinder Pressure at 1600rpm .................................................................... 50
FIGURE 5.4 Cylinder Pressure at 3200rpm .................................................................... 50
FIGURE 5.5 Cylinder Pressure at 5000rpm .................................................................... 51
FIGURE 5.6 Predicted Oil Film Thickness (lst Cycle) .................................................. 52
FIGURE 5.7 Oil Film Development at 1600rpm 0.0 inch Manifold Vacuum ................ 53
FIGURE 5.8 Oil Film Development at 3200rpm 0.0 inch Manifold Vacuum ................ 54
FIGURE 5.9 Oil Film Development at 5000rpm 0.0 inch Manifold Vacuum ................ 54
FIGURE 5.10 Prediction of Average Blow-by (Hour 0.0) ................................................ 55
FIGURE 5.11 Prediction of Average Blow-by (Hour 500.0) ............................................ 56
FIGURE 5.12 Comparison of Average Blow-by Increase ................................................ 57

FIGURE 5.13 Wear at 1600rpm 0.0inch Manifold Vacuum(1st Simulation Cycle) ......... 58
FIGURE 5.14 Wear at 1600rpm 0.0inch Manifold Vacuum (After 500 hours) ................ 59
FIGURE 5.15 Wear at 1600rpm 12.0inch Manifold Vacuum(lst Simulation Cycle) ....... 60
FIGURE 5.16 Wear at 1600rpm 12.0inch Manifold Vacuum (After 500 hours) .............. 60
FIGURE 5.17 Wear at 1600rpm 20.0inch Manifold Vacuum(lst Simulation Cycle) ....... 61
FIGURE 5.18 Wear at 1600rpm 20.0inch Manifold Vacuum (After 500 hours) .............. 61
FIGURE 5.19 Wear at 3200rpm 0.0inch Manifold Vacuum( 1 st Simulation Cycle) ......... 62
FIGURE 5.20 Wear at 3200rpm 0.0inch Manifold Vacuum(Afier 500 hours) ................. 63
FIGURE 5.21 Wear at 3200rpm 12.0inch Manifold Vacuum(] st Simulation Cycle) ....... 63

FIGURE 5.22 Wear at 3200rpm 12.0inch Manifold Vacuum(Afier 500 hours) ............... 64

FIGURE 5.23 Wear at 3200rpm 20.0inch Manifold Vacuum(1st Simulation Cycle) ....... 64

FIGURE 5.24 Wear at 3200rpm 20.0inch Manifold Vacuum(After 500 hours) ............... 65
FIGURE 5.25 Wear at 5000rpm 0.0inch Manifold Vacuum(lst Simulation Cycle) ......... 65
FIGURE 5.26 Wear at 5000rpm 0.0inch Manifold Vacuum(After 500 hours) ................. 66
FIGURE 5.27 Wear at 5000rpm 12.0inch Manifold Vacuum(lst Simulation Cycle) ....... 66
FIGURE 5.28 Wear at 5000rpm 12.0inch Manifold Vacuum(After 500 hours) ............... 67
FIGURE 5.29 Wear Rate at 1600rpm ................................................................................ 70
FIGURE 5.30 Wear Rate at 3200rpm ................................................................................ 70
FIGURE 5.31 Wear Rate at 5000rpm ................................................................................ 71
FIGURE 5.32 Comparison of Steady State Wear Rate ..................................................... 71
FIGURE 5.33 Measured Wear Rate .................................................................................. 72
FIGURE 5.34 Ring Race Change at 1600rpm .................................................................. 73
FIGURE 5.35 Ring Face Change at 3200rpm ................................................................... 74
FIGURE 5.36 Ring Face Change at 5000rpm ................................................................... 75

FIGURE 5.37 Comparison of Time-average Oil Consumption over 500hrs Simulation .77

FIGURE 5.38 Estimated Oil Evaporation During One Cycle .......................................... 78
FIGURE 5.39 Total Oil Consumption Rate for Reverse Flow and Vaporization ............. 79
FIGURE 5.40 Oil Loss vs. Gas Backflow through lst Ring Gap (1) ................................ 81
FIGURE 5.41 Oil Loss vs. Gas Backflow through lst Ring Gap (II) ............................... 82
FIGURE 5.42 Oil Loss vs. Gas Blow-by (I) ..................................................................... 84
FIGURE 5.43 Oil Loss vs. Gas Blow-by (II) .................................................................... 85
FIGURE 5.44 Oil Loss vs. Wear Rate (1) .......................................................................... 87
FIGURE 5.45 Oil Loss vs. Wear Rate (11) ........................................................................ 88

xi

Chapter 1 INTRODUCTION

1.1 Motivation

The world energy shortage has motivated the automotive industry to produce
engines that will be more efficient in the future. Petroleum, the main source of energy for
today’s stereotypical engines will become very limited in a few decades; more energy-
saving engines can prolong its usage. Another factor that motivates the engine research is
the environmental pollution. Stricter emission regulations have been enforced on
automobiles in order to preserve the quality of the environment, because they are
considered to be a major contributor to pollution. The automotive industry has been
searching for possible solutions. One approach is to design engines that can use alternative
energy sources, such as bio-related fuel produced from crops. A second approach would
be to invent another type of engine to replace the internal combustion engine, such as a
fuel cell powered vehicle; while another approach is to increase the fuel efficiency of

engines.

With the two factors mentioned above, it is obvious that the criteria for designing
engine should include fulfilling the requirement of high energy efficiency and low
pollution level. In other words, engines should have low fuel consumption (high mileage
per gallon of fuel) and low emission of toxic gases. Generally, energy efficiency and
emission are determined by factors such as the friction level of an engine, the lubrication

system and the energy conservation system of an engine.

The main region that contributes to the engine friction and emission is the cylinder
kit, which consists of pistons, rings, cylinder bores, connecting rods, bearings as shown in
Figure 1.1. The cylinder kit of an engine is considered to be the power source of the
engine because it is the region where chemical energy from fuel is transformed into kinetic
energy that drives the whole automobile. When the fuel (chemical energy) is ignited in the
combustion chamber, the explosive force from the burning fuel pushes the piston into a
reciprocating motion (kinetic energy). The highly dynamic forces, during the
reciprocating motion in the cylinder kit, often result in the wearing of contact surfaces due
to friction. The contacting surfaces include the surfaces between the piston and piston
rings, piston and cylinder bore, and piston rings and cylinder bore. In normal operating
condition, friction and wear cut scratches in the metal surfaces mentioned above. This

causes compression and power loss, and increased oil consumption.

Understanding the physics within the cylinder kit system is crucial for automotive
research. It is an important step toward inventing high efficiency engines. However,
studying the physics of the engine cylinder kit system is a rather complicated process,
because it involves the understanding of heat transfer, fluid mechanics, thermodynamics,

and solid mechanics processes that occur during engine operation.

With the advanced computer systems available for complicated calculation, the
computational method is becoming a partial approach for engine simulation. Using this
method, engineering equations, formulas, and correlations derived from theory and
experiments can now be implemented into the computational models, which include

analytical and numerical models. The engine cylinder kit system can also be modeled

computationally. In this thesis, computer models of piston ring wear and oil consumption

are studied.

1.2 Literature Survey

Michigan State University Engine Research Laboratory(MSUERL) has been
putting much effort in the research related to the piston ring wear and the oil consumption.
Chung [1] in his Ph.D. dissertation, developed a piston ring wear model through the use of
ring dynamics analysis based on several results of experimental efforts in order to study
the ring loading effect on the piston ring wear. The simulation results, over the engine
speed range between 1200rpm to 2100rpm, showed that higher engine speed operations
had lower ring wear rate. Barkman [2] in his MS. thesis, performed experiment on the
measurement of piston ring wear rate of a one-cylinder engine. His results showed that the
piston ring wear rate decreased at a relatively high rate to a steady state value afier certain
hours of operation. Pint [3] developed a mathematical model for the prediction of the
tribological system in a simple internal combustion (IC) engine. He designed and
developed a software module for the analysis of a three dimensional piston ring wear. His
analysis related piston ring twist and piston tilt to the degree of wear in IC engines, under

different operating conditions.

For the oil consumption analysis, Ejakov [4] in his Ph.D dissertation, performed
ring pack behavior and oil consumption modeling of IC engines. His results showed the
importance of modeling ring twist in order to have a more accurate prediction of the piston
ring behavior and oil consumption. Qualitative analysis and discussion on oil consumption

were performed based on gas dynamic results.

Research in the piston ring wear and the oil consumption was also performed by
many researchers from other research institutes. Takashi Inoue et a1.[5] performed
experimental measurement of transient oil consumption of an automotive engine. Their
study showed that the reverse gas flow carries oil from engine crankcase to combustion
chamber because of the high vacuum under engine-brake conditions. Min et a1.[6]
experimentally studied the dynamic characteristics of an oil consumption relationship;
between the instantaneous oil consumption and the location of the piston ring gap in IC
engines. Their results showed a significant effect of piston rings rotation on oil
consumption, under different operation conditions. Tian et a1.[7] developed a complete
one-dimensional mixed lubrication model to predict oil film thickness and fiiction of the
piston ring-pack. Their study found that the surface roughness has an important impact on
oil transport across the piston ring pack into the region above TDC (Top Dead Center).
The preliminary comparison showed good agreement between model predictions and the
measurements for different engine oils. Wong et a1. [8] developed a model to estimate the
oil vaporization from the cylinder liner as a contributing mechanism to engine oil
consumption. They found that most (74%) of the oil vaporization occurred during intake

and compression strokes.

The researchers mentioned above have indeed contributed to a deeper
understanding of oil consumption and wear that happen in IC engines. But, discussion and
analysis on the relationship between oil consumption and piston ring wear were not
performed in depth. Therefore, the objective of this thesis is to extensively study the

relationship between the two mechanisms (piston ring wear and oil consumption).

1.3 Problem Statement

Piston ring wear and excessive oil consumption are common problems for an
internal combustion engine as the engine gets older. Piston ring wear occurs on the contact
surface of the piston rings and cylinder liner, after a certain amount of time of operation.
The contact surface experiences high pressure and forces that cause deformation of the
geometry and degrading of the quality of the surface material. The distorted surface
affects the functionality of the piston rings, and results in energy loss. Because of wear, the
piston rings not only lose their sealing power, which prevents gas flow through ring piston
ring pack; but also lose their functionality to help in the formation of an oil film layer for
lubrication purposes. The excessive oil flow rate through the ring pack and the formation
of oil film thickness may be the determining factors for oil consumption. Ejakov’s [4] in
research on oil consumption and ring pack behavior indicated a close relationship between
the ring pack gas flow and oil film thickness, and the oil consumption. There are several
mechanisms in oil consumption: oil loss through the vaporization process, and oil flow
through piston ring gap during the high pressure gradient between the combustion
chamber and piston second land. Other mechanisms such as oil consumption by oil mist

and throw-off of liquid droplets, but they are not included in this analysis.

The purpose of this thesis is to implement piston ring model and oil consumption
model in the design and analysis of a 4.6L, V8 engine in order to study the relationship
between piston ring wear and oil consumption. The design parameters of the 4.6L, V8
engine and its thermodynamic and heat transfer data are used in computer simulations of

piston and ring dynamics as well as the prediction of wear and oil consumptions. The

results are to be used to analyze the significance of one mechanism affecting another

during engine operation.

 

Top
Compression
R'ng

Second
Compression
R'ng

Oil Control
Ring Ra'l

Oil Control
Ring Expatder

 

 

 

 

 

 

 

 

 

 

FIGURE 1.] Cylinder Kit Components of A Basic Engine

Chapter 2 FUNDAMENTALS OF PISTON RING PACK
LUBRICATION

2.1 Introduction

The lubrication system in an IC engine plays a major role in reducing the friction
and enhancing the smoothness of operation. Without a good lubrication system, an IC
engine will encounter problems such as excessive energy loss and short engine life.
Friction losses directly affect the maximum brake torque and minimum brake specific fuel
consumption. Often the difference in the friction losses distinguishes a good engine design

from an average engine design [ 9 ].

Figure 2.1 shows the relative significance of different friction work in typical four-
cylinder automotive engines (Spark-Ignition, SI) and diesel engines (Compressed-
Ignition, CI) at different loads and speeds. The absolute value of the total friction work
changes with load, and increases as speed increases. It is observed that fiiction loss in the
cylinder kit (piston-crank) assembly is the major contributor to the total energy loss of

engines.

The major design factors that influence cylinder kit assembly friction are the
following: ring width, ring face profile, ring tension, ring gap, liner temperature, ring land
width and clearances, skirt geometry, and skirt-bore clearance. Usually, cylinder kit

assembly fiiction is dominated by the piston ring friction.

A typical compression ring is separated from the cylinder liner with a thin layer of
oil, which acts as the cushion that reduces friction, as shown in Figure 2.2. Depending on
the oil film thickness, there are three distinct regions of piston ring lubrication: boundary,

mixed, and hydrodynamic as shown in Table 2.1.

 

2 3600mm __
4.0baerep
175 1800 rpm .
' 2.0baerep
1.5
1800 rpm
125 .

6.0bar Bmep

II.

Mean Electve Pressue,bar
'3: a

P
m

0.25 r

 

 

 

     

FIGURE 2.1 Comparison of Major Categories of Friction Losses [ 9 ]

TABLE 2.1 Regions of Piston Ring Lubrication

 

 

 

 

Regime of Lubrication Condition
Lubrication
. Enough oil film between the surfaces, and
Hydrodynanuc
surface-to-surface contact does not occur
Surface-to-surface contact occurs because
Boundary
oil film is not thick enough
_ The intermediate region between
Mixed . _ .
hydrodynamic and boundary lubricatlon

 

 

 

 

Lubricant Ix

 

  

FIGURE 2.2 Compression Ring with Lubricating Oil Film

  

There are models for the ring and oil film behavior to predict the ring dynamic and

oil film formation. The Reynolds equation is used for oil film thickness model of

hydrodynamic lubrication. For the practical case where the oil film thickness, h is much

less than the ring width, the Navier-Stokes equations for the liquid film motion reduces to

a Reynolds equation of the form:

$03—10: —6 Upg—: + 121131:
where:
h oil film thickness
pressure
U velocity
)1 dynamic viscosity of oil

(2.1)

Oil film thickness is determined by several factors: oil availability to the ring pack,

lubricant properties, ring radial tension, gas pressure, ring funning face profiles, engine

speed, liner temperature and surface texture.

2.2 Formulating Oil Film Thickness

2.2.1 Oil Viscosity

Viscosity is a characteristic property of oil. Similar to modulus of an elastic solid
material, it is used for the shear analysis of lubricants. Oil viscosity affects the formation
of oil film thickness. Consider a lubricant film between a cylinder bore and a piston ring.

Assume that the piston ring has a flat surface profile, as show in Figure 2.3.

 

FIGURE 2.3 Linear Velocity Distribution

When the piston ring moves with a velocity U relative to the cylinder bore, a thin
layer of oil adheres to the bore such that the velocity of this layer is zero. Another layer of
oil simultaneously adheres to the moving ring face and it has the velocity U. The first
assumption to be made here is that the oil flows in parallel layers and there are no
secondary irregular motions and fluctuations superimposed on the flow; secondly, it is

assumed that the flow is laminar. According to the shear stress equation

(2.2)

isl‘i‘j

where
t shear stress

F force required to maintain the flow

A area of the ring face in contact with the oil

10

If it is assumed linear velocity distribution for the thin oil film, as shown in Figure

2.3, the rate of shearing strain of the oil, 7 will become

bIQ

r = (2.3)

By definition, the dynamic viscosity of oil is the ratio of shearing stress to the rate
of the shearing strain

shearing stress

 

ll = rate of shearing strain
_ F/A_ 1:
“ ‘ U/h— U/h

Thus, the shear stress equation becomes

=> 1 = u-L-I (2.4)
h
For nonlinear velocity distribution as shown in Figure 2.4, the shear stress is

t: 11% (2.5)

Piston Ring

 

FIGURE 2.4 Nonlinear Velocity Distribution

11

The SAE (Society of Automotive Engineers) standard engine oil is commonly
used by the majority of vehicles today. In the SAE standard, engine oil is separated into
different grades according to the viscosity. Viscosity changes with temperature as

indicated in Table 2.2.

TABLE 2.2 SAE Oil Viscosity

 

 

 

 

Oil Grade Temperature Viscosity(cP)
25 0C 41.45
SAE SW
100 0C 4.06
25 0C 229.47
SAE 20W
100 0C 10.50
25 0C 574.46
SAE 50
100 0C 17.99

 

 

 

 

 

The relationship between viscosity and temperature for the SAE grades of oil is

assumed to follow the Vogel viscosity equation:

 

( ” )
temperature + c

11 = ae (2.6)

where a, b, c are Vogel constants

TABLE 2.3 Vogel Constants for SAE Grades Oil

 

 

 

 

SAE Grade a b c
5W 0.05567 900.0 110.8
20W 0.02370 1361.0 123.3
50 0.01963 1518.0 122.6

 

 

 

 

 

 

12

2.2.2 Oil Flow in Piston Ring Pack

Oil flow in the piston ring pack needs to be determined for the calculation of oil
film thickness. Assume that oil flow in the piston and piston ring pack is similar to a

viscous flow through a wide, rectangular slot, as shown in Figure 2.5.

 

FIGURE 2.5 Oil Flow Through a Rectangular Slot

Consider the slot where the width b is much bigger than the gap h so that we can
neglect flow losses at the end; consider also that the length 6 is much bigger than the gap

so that we can neglect losses at the entrance and the exit of the slot. It is a fair assumption

13

because the clearance between piston and piston rings, and cylinder wall is very small (in

the order of 10'4 meter) in comparison to the perimeter of cylinder wall, which is about
100 times bigger in magnitude. The downward driving force for the flow is the pressure

gradient between point A and point B:
A p = P1 - P2

Using Newton’s Law, the force is
Fd=2ybAp (2.7)
If we assume a nonlinear velocity profile, the shear stress, I = u*(du/dy).

Thus, the resisting upward force will become

F” = 1: A = 2 f b “(du/dy) (2.8)

Applying static equilibrium for the forces F d and F u will results in

2ybAp = -(21bu:—:) (2.9)
2) AP —
—mydy — a'u

Equation 2.9 has negative sign on the right hand side because of the negative

velocity gradient where u decreases as y increases. Applying boundary conditions

le“

u=0aty=i

on the integration of Equation 2.9 results in

14

2
_ .Ai’. ’_’__ 2
u _ 2M[4 y] (2.10)

The velocity from Equation 2.10 has a parabolic profile in y direction as shown in

Figure 2.6. The maximum velocity at y = 0, is

_ Aph2
umax — 8‘16
The average velocity is
2 Aph2
uave = §umax = W (2-11)

With the velocity available, we can calculate the volume of flow, Q passes through

the slot. Multiplying the average velocity with cross-sectional area, we get

3
_ Apbh
_ 12M (2.12)

Q

 

 

FIGURE 2.6 Parabolic Velocity Distribution

15

2.2.3 Reynolds Equation for Ring Pack

With the flow rate being determined, we can further derive the oil film thickness

formula to obtain the Reynolds equation. The three-dimensional Reynolds equation has

the following general form:

£675?” + 3% (ZL 1”): — (rug—h — 6W3: + 12V (2.13)

where
oil pressure

[.1 oil dynamic viscosity

h oil film thickness

U ring axial velocity

W ring tangential velocity

V = £311: ring radial velocity

We can further simplify the Reynolds equation into two-dimensional form if we

assume an axisymmetric case of the piston ring lubrication, as shown in Figure 2.7:

“ax

8%[13—1’}: —6Ua—h+1 (2.14)
Equation 2.14 is valid for piston ring lubrication for conditions:
onegligible body forces
onegligible fluid inertia

oconstant lubricant pressure throughout the thickness of the oil film

onewtonian fluid for the lubricant

l6

- laminar flow

ono slip at the boundaries

The Reynolds equation will be in the form of Equation 2.1 when constant lubricant

viscosity, u is assumed.

a 33p _ _ ah ah

Assuming non-deformable surfaces for both the ring and the liner, integrating

Equation 2.1 yields
- 6 U] 12 3h] AI B 215
P-‘ll 1+ “372+ 3+ (-)
where:

A, B integration constant

12 = Jédx
1,... jg;

The integration constants A and B are found from the boundary conditions. Three

typical boundary conditions for piston ring lubrication are given in Table 2.4.

17

TABLE 2.4 Boundary Conditions for Piston Ring Lubrication

 

 

 

 

 

 

 

 

Position Sommerfeld Half-Sommerfeld Reynolds
x=0 p=PI p=P1 p=P1
x = b/2 -- p = 0 --
x=b p=P2 -- p=P2,‘—’L’=o
dx

 

With defined boundary conditions, and a given piston ring face profile, the
hydrodynamic pressure distribution can be determined from Equation 2.15. The
hydrodynamic pressure force must be equated by another source of forces. The forces are
fi'om the pressure force acting behind the ring and the tension of the ring. The

hydrodynamic pressure force acting on the ring face is determined by integrating the

 

 

FIGURE 2.7 Piston Ring Lubrication

pressure, p(x) over the ring face:

b
P = I0p(x)dx

l8

 

For equilibrium in the radial direction as show in Figure 2.8, the pressure force

obtained from Equation 2.16 must be equal to the radial ring force, F R

P = FR (2.17)

Then, the minimum oil film thickness (hmin) will be determined by satisfying

Equation 2.17.

 

p(x)

   
  

 

444?“? 4; ,4. 4,..,.,::.:,,
//////, 75/4,, 5,, .. .1 .g/ I
//.44;, /1. . _. .,

fix-2;. . ,7, I// . , v. /,-',". I" _
, /‘ .. ."- .. 'r .
fab/)1 : under I inergv/gxygw ,. )4 . .
’// . x2"

 

2727/

 

"X = b Piston Ring mm

.r'.“:<\ \"~.-.'-

.-.\\“~,‘)§- :

I .413:

. (Elk

\\\\\\

\ \:I;\H\‘QVV~:

X ~\\\\\\\:‘

FIGURE 2.8 Forces Acting on a Piston Ring

 

/

 

 

l9

Chapter 3 FUNDAMENTALS OF PISTON RING WEAR

3.1 Introduction

Cylinder kit components experience extreme mechanical stresses. These
mechanical stresses are caused by high kinetic energy and momentum from piston
dynamics, high temperature and pressure during combustion, and distortion of metal. A
common problem that arising in this environment is wear, as a result of contacts of surface
asperities between the rings and the cylinder wall. Therefore, the design of piston ring
pack must satisfy the criteria set by durability and efficiency, despite of the severe

conditions.

The reciprocating motion of the piston and piston rings against the cylinder liner
causes the wearing of the contact surfaces, as shown in Figure 3.1. Recalling that there are
three regions of lubrication: boundary, mixed and hydrodynamic, we assume that piston
ring wear only occurs in the regime of boundary and mixed lubricant where there is lack

of lubricant that avoids surface-to-surface contact.

Generally, wear is caused by the individual and combined effects of corrosion,
adhesion, and abrasion. Corrosion occurs when acidic substances within the cylinder kit
assembly, such as the products of combustion, attack metal surfaces. Those metal surfaces
include the major component of the cylinder kit assembly, such as piston and ring.
However, effective prevention has been discovered through the invention of chemical

detergent oils that neutralize the corrosive acids and with proper control of the engine

20

temperature. This type of wear is difficult to model because the issue is more about the

problem of chemical reaction. Hence, corrosive wear is assumed negligible in this thesis.

 

FIGURE 3.1 Rubbing of Two Contact Surface Under Microscopic View

Another type of wear is abrasion. Abrasive wear occurs when there are impurities
present in the lubricating oil at the contact surfaces. The impurities can come from
atmospheric dust, and from the metallic debris from corrosive and adhesive wear. When
the impurities stay between two contact surfaces, such as the piston ring and cylinder
liner, the ratio of asperity contact is increased. Abrasion can be eliminated by using air
filtration, periodic oil changes and filtration in the oil system. This type of wear will not be

included in this thesis.

21

The third type is adhesive wear. It mainly affects parts of the engine where there is
metal-to-metal contact. The cylinder kit assembly experiences this type of wear when
there is metal-to-metal contact between the piston, rings, and cylinder walls, and takes
place. Adhesive wear is especially significant when the engine is at the beginning of
operation and is cold because there is insufficient oil in the cylinder kit assembly. This

type of wear is of the focus for this thesis.

3.2 Formulating Piston Ring Adhesive Wear

3.2.1 Surface Texture of Piston Ring

Surface texture affects the wear rate of a surface. Surface texture can be vaguely
identified with the roughness and smoothness of a surface. Surface texture representation
can be very complex. A complete description of a surface involves many aspects: the
amplitude (roughness and smoothness), the waviness (the pattern of surface roughness)
and the density of surface asperities per unit surface area. Identifying a surface with only

one aspect while neglecting the other will result in inaccurate surface representation.

There are several methods for the representation of the surface texture. A common

parameter representing a surface texture is Ten-points-height, R2. It is the average distance

between the five highest peaks and the five deepest valleys of a surface within the

sampling length. The formula for R2 is

(Yp1+ Yp2+ Yp3+ Yp4+ Yp5)‘(Yv1 + Yv2+ Yv3+ er’r Yv3)

R2: 5

(3.1)

 

where:

lg’s the height of the five highest peaks

Y ’s the depth of the five lowest valleys

V

22

Another commonly used parameter is the Roughness Average, Ra. Roughness

average is the arithmetic average of the distance of the filtered or unfiltered roughness

profile fiom its mean line, as shown in Figure 3.2.

Mean L'ne

 

 

 

 

 

FIGURE 3.2 Surface Profile

The formula for roughness average is as following:

Ra = 1% j: lh(x)ldx (3.2)

Another commonly used parameter is the Root Mean Square Roughness Average,

Rq. It is the root mean square of the distance of the filtered or unfiltered Roughness profile
from its mean line, as shown in Figure 3.2. Rq is the standard deviation of the amplitude

density distribution. The amplitude of this parameter is more sensitive to peaks and

valleys compared to Ra. The formula for root mean square roughness average is as

following:

 

Rq = fl”; h2(x)dx (3,3)

0

23

Using Ra and Rq to represent a surface texture is sufficient, except when there are
differences in the density of the real contact area. To illustrate this point, we can use two
surfaces that have the same value of Ra and Rq, but have different surface texture. Figure

3.3 shows two different surfaces: surface 1 and surface 2.

Height

Average Reference Line

 

Surface 1

Height

      

Average Reference Line

Surface 2

FIGURE 3.3 Engineered Surface Profile

Applying Equation 3.2 and Equation 3.3 will give the same value of Ra and Rq,
even though these two surfaces are different. To distinguish surface 1 from surface 2,
another parameter is introduced: the Abbott Firestone Curve (AFC), which is also called
the Bearing Ratio Curve. It is a graphical representation of the bearing ratio parameter, tp
(length of bearing surface, expressed as a percentage of the assessment length of a surface
specimen, at a depth below the highest peak or a selected distance from the average
reference line) in relation to the profile level. The curve contains all the amplitude

information of a surface profile. The AFC for surfaces 1 and 2 is shown in Figure 3.4.

24

Abbott-F1 restone Curve

Proftifl1e
Wright de‘.

  

Average
Reference Line

 

 

in
Surface 1 A 100% JI V w
Profile
Height der th
Average
Reference Line
to

 

 

 

1oo%—>l' I

 

Surface 2

 

FIGURE 3.4 Surface Representation Using Abbott Firestone Curve
From the Abbott Firestone Curve, one can identify the difference between the two

surfaces, where surface 1 has lower amplitude density than surface 2.

In this thesis, we describe the piston ring surface in terms of R: and the Abbott
Firestone Curve. The piston ring surface profile is assumed to be in the shape of cones

with the peak angle, [3 of 45 degrees, as shown in Figure 3.5.

Abbott-Firestone Curve

‘ D th

      

 

 

,a“

r¢.__':i1—Oo%
FIGURE 3.5 Piston Ring Surface Profile

25

The surface of the cylinder liner needs to be modeled too. However, this step can
be avoided by assuming that the interaction of two surfaces can be simplified as the
interaction of one rough surface with a planar surface. The roughness value will become

the combination of the two surfaces roughness.

_ (3.4)
Rz _ R2] + R22

where
R2 the roughness of the combined surface profile
R2, the roughness of surface one
R22 the roughness of surface two

3.2.2 Wear Mechanism

Consider a small section of piston ring sliding on cylinder bore as shown in Figure

3.6.

 

 

 

 

FIGURE 3.6 Piston Ring Sliding on Cylinder Bore

26

The section of the piston ring is assumed to be very small such that the curvature
effect of the piston ring in circumferential direction is negligible. The amount of worn

volume from the piston ring surface is calculated using the following equation:

5 V = k g in (3.5)
where
5V incremental volume loss from the contact surfaces
5s sliding distance of one surface with respect to another surface
F the total load applied to the piston ring surface while it is sliding
H combined hardness of the piston ring and cylinder liner
k wear coefficient

Equation 3.5 is only valid when the lubrication is not hydrodynamic. To model
this, we can assume that the lubrication is not hydrodynamic when the oil film thickness
between the piston ring and the cylinder liner is less than the combined roughness. In the
regime of mixed or boundary lubrication, the entire load is assumed to be carried by the
metal surface contact between the piston ring and cylinder liner, and the support by the oil
film is neglected. To evaluate the total load that applies on the surface, consider that a
piston ring pressed against the cylinder bore by forces such as ring tension and the
pressure forces behind the piston ring. Assume that all the forces are transferred to the
cylinder liner as a normal force. Using a solid mechanic analysis, pressing the ring against
the cylinder bore will result in axial force distribution on the piston ring surface, as shown

in Figure 3.7.

27

if! 4S4;

‘ 4~ ,
igCMmder
{inner
\‘. I L‘::‘

 

. k4} " V) _4
47, p»
~ ~4 rrna

' f Force.N
_ 11%)") '1'.—
.. ..§;Pistoww
‘ ';.j;f.Ring Section; 5 .

 

 

 

To ‘ 4:;
\x» \ ‘~

FIGURE 3.7 Ring Section Contact with Cylinder Liner

Applying Newton Second Law, one can deduce that

N = F, (3.6)
where
Fr = If(x)dx (3.7)

In order to obtain f(x), the following assumptions are made:

olinear elastic behavior of the piston ring material

omodulus of elasticity of the cylinder liner is infinite

odeformation is very small

oshear stress on the piston ring is negligible compared to normal stress

With the assumptions above, the determination of force distribution begins with
the discretization of the piston ring surface into In slices with the same width as shown in
Figure 3.8. These slices are assumed to be completely separated from each other since the

normal stress is much higher than the shear stress.

28

 

Cylinder
~Liner

: Normal
‘ Force,N

FIGURE 3.8 Discretized Piston Ring Surface

Using Hook’s law:

where:

AI

cl = E 21. (3.8)
AEI.

E = T (3.9)
l

f: = 61. Al (3.10)

the normal stress

force at each slice

the modulus of elasticity of piston ring
the elastic deformation of the piston ring

the deformation of the length of a slice

29

If the original length of a slice

A the cross-sectional area of a slice

In order to obtain the force for each slice, an arbitrary reference point is assigned to
start the process of evaluating the total force. The process continues with the reference
point being varied until the criteria of Equation 3.6 is met. The final value for the force of

each slice will be used for the calculation of wear at each slice.

30

Chapter 4 FUNDAMENTALS OF OIL CONSUMPTION

4.1 Introduction

 
  

. . <:
Vaporization C
C

Reverse Oll Flow <72;

   
  

   

Oil Supply
a on Film

 
    
  
 
 
 
 
 

Cylinder
\ Liner

Major Oil Path
#-
Minor Oil Path

4.1 Oil Consumption Mechanism

Oil consumption is a phenomenon in which the engine oil escapes from the

engine crank case into the environment. The engine oil consumption can occur through

31

various ways: oil leakage through the engine gasket, vaporization of oil from the oil film
on the cylinder liner, flow of oil from the engine crank case to the combustion chamber
through the piston ring gap, throw-off and blow-by. Oil consumption can be reduced
primarily by a better design of the engine crank case sealing system to prevent oil leakage
through engine gasket. However, to further reduce oil consumption, one must solve the
problems of excessive oil loss through other methods. In this research, two contributors of
oil consumption are the main focus: oil vaporization and reverse flow of engine oil

through the piston ring endgap as shown in Figure 4.1.

Oil vaporization is associated with the magnitude of the film thickness left on the
cylinder bore. The engine oil, which vaporizes from the cylinder liner, diffuses into the
cylinder gases and leaves the combustion chamber as either partially burned or unburned
hydrocarbon. Hence, reducing the oil vaporization can reduce hydrocarbons in the exhaust

gases, and further reduce overall engine emission.

On the other hand, reverse flow of engine oil through the top piston ring endgap
is associated with the pressure drop across the piston ring pack. The design of the piston
and piston ring, as well as the configuration of piston ring endgaps, influence the oil flow

into the combustion chamber.

4.2 Vaporization of Engine Oil from the Cylinder Liner

A vaporization model is used here to predict the amount of oil loss due to heat
and mass transfer. It is appropriate to use this model by assuming that the oil film left on
the cylinder liner during the intake and exhaust strokes as a liquid surface, while the
combustion gas is a gas flow above the surface as shown in Figure 4.2. Because of the

32

reciprocating motion of the piston ring with respect to the cylinder liner, the area of the
cylinder liner exposed to the combustion gases is also changing. The volume of oil

vaporized per crank angle, Vemp is evaluated using the equation below:

A m”At
V j = —— (4.1)
€\ ap p0”
where:
A area of cylinder bore being exposed to combustion chamber
as a function of crank angle
m” evaporative mass flux

At time interval between two successive crank angle

poi, engine oil density

The following assumptions must be made in order to use Equation 4.1:

othe rate of oil vaporization is low to be considered a low mass transfer rate
convection process

othe change in the engine oil temperature is very small so that its density can be
treated as constant

othe temperature of the oil film does not exceed the engine oil boiling point so
that the burning of oil does not occur

othe engine oil is considered a single-species fluid for the simplification of the
problem

33

empirical equations are

where:

D

m evap
g m

mf

Sh

poi!

Dh

DC

P

vap

P

cylinder
MW’s

a,a',e

ll

m evap = gm mf

Shpoith
gm : ————D-——

C

m _PP[ vapor )[MWZH]
f cylinder MW

511- — a ReDd Sce

 

evaporative massflux
mass transfer coefficient
mass fraction

Sherwood number

engine oil density

binary diffussion coefficient

hydraulic diameter
instantaneous vapor pressure of oil
instantaneous pressure of combustion gas

molecular weight

constants

34

The evaporative mass flux is derived from the mass convection theory where the

(4.2)

(4.3)

(4.4)

(4.5)

(4.6)

ReD Reynolds Number
Sc Schmidt Number

vai, arr krnematrc vrscosrty

The vapor pressure of different grades of hydrocarbon can be obtained from the

Handbook of Vapor Pressures and Heats of Vaporization of Hydrocarbons and Related

Compounds [11]. The relationship between vapor pressure and surface temperature follow

the Antoine equation:

B
longap = A — —
C + Tsurface
where
A, B, C Antoine constants for engine oil
Oil film

 
  

Combustion Gasr-fLiquid ,
Chamber \ nte ace 1

I
l // I:95‘$\ Pvapoil
\ Gasflowll l: \

 

cylinder
~_ Bore

     

     

I
I
I
l
i
\ /
// l i
.\ I
\ W‘ :
\ y I
~' Piston: _ g, - :

  

(\4‘.\:

FIGURE 4.2 Vaporization Process at Oil Film

35

(4.7)

With the Antoine equation, the only unknown variable needed to calculate the
vapor pressure is the surface temperature of the oil film. Obtaining the surface temperature
of the oil film is possible by using the thermal resistance model. Consider Figure 4.3. With
the temperature of the combustion gas and liner known from measurement, we are able to

deduce the net heat transfer across the oil film using

T. , — T
q“ = inf—8‘" (43)
e Rtotal
H 1
Rm, = R1 + R2 = [a + a (4.9)

Since the net heat transfer across the oil film is the same as the heat transfer from

 

 
 

Tliner t0 Isurjface’ thus
Tliner _ Tsur ace
9. = q... = —R f (4.10)
l
:> Tsurface = Tliner _ R1 qnef (411)
Oil FUR
Combustion
Chamber "
p—H—>
I
1 Cylinder
: Bore
q < ', 0
L . fl _« .
hA I kA _
gas Tsurface T| ner
Piston

 
 

 

FIGURE 4.3 Thermal Resistance Model for Oil Film

36

We evaluate the amount of oil loss due to vaporization using Equation 4.1 to
Equation 4.11. The results will be added with the amount of oil loss due to another method

that will be discussed next.

4.3 Reverse Flow of Engine Oil through Top Compression Ring Gap

Engine oil consumption can be caused by the reverse flow of the oil through the
top compression piston ring, when pressure in the second land of the ring pack is higher
than the pressure in the combustion chamber. To model this oil loss mechanism, one can
apply the theory of laminar viscous flow in a channel or duct. As in Figure 4.4, we assume

that the ring gap is a non-circular duct with the length of the ring width, R”, and it is

completely filled with viscous engine oil.

The flow of the engine oil through the ring gap may be a laminar flow or it may
be a turbulent flow. Using the dimensionless parameter, Reynolds number, Re- the ratio of
the inertia to viscous effects in the flow, we can determine the type of flow through the

ring gap. The equation of Reynolds number is

ReD = ¥ . (4.12)

37

Ring Endgap

     
  
 

 

§I "I'm mpg commesgfii Ring Width
“rm "fl \\\\ Edge

   

FIGURE 4.4 Flow Through Piston Ring Gap
In the piston ring pack, one can assume the viscous oil flow to be very slow,
because the average cross sectional area for oil flow is relatively small (in the order of

10'7 meter square) and the viscous effects are dominant. Hence, a low Reynolds number
flow is expected and we can categorized the type of flow as laminar for the oil flow

through the piston ring end gap.

Another factor that we need to consider is the effect of the entrance region.

Including this factor into the calculation of oil flow will be complicated. Therefore, we

38

need to simplify the problem by assuming a fully developed flow. For this assumption to
be valid, the dimension of ring gap and the ring width is important. Consider a pipe system
as shown in Figure 4.5. The fluid typically enters the pipe with a nearly uniform velocity
profile at section 1. As the fluid moves through the pipe, viscous effects cause it to be
retarded by the pipe wall (the non-slip boundary condition.) Due to the boundary layer
effects, which are caused by the viscous effects, the velocity profiles change with distance
along the pipe. In the region where the velocity profiles are changing, it is called the
entrance region. After the entrance region, the velocity profile remains constant and the
flow is considered to be fully developed. In the fully developed region, viscous effects are
of considerable importance. As with many other properties of pipe flow, the dimensionless

entrance length, fe/D, correlates quite well with the Reynolds number. Typical entrance

lengths are given by

 

  
 

i
b9 = 0.06 Re for laminar (4.13)
E... R - LL F
gig: eglon r 4 ully DevelopedJV—_.
Boundary Layer D

 
   

FIGURE 4.5 Different Flow in a Pipe System

To simplify the modeling of the oil flow through the ring gap, we can assume that

the effect of the entrance length is negligible, provided that the ratio of the ring width to

39

the hydraulic diameter of the ring gap channel satisfies Equation 4.13. This assumption is
appropriate for a typical IC engine, since the ring gap is relatively small compared to the

ring width.

4.3.1 Formulating Reverse Flow Equation

Refer to Figure 4.4, the second land of the piston ring pack is labeled 2, while the
combustion chamber is 1. When the pressure at 2 is higher than the presure at l, the
pressure force will push the oil through the ring gap. For simplification in deriving the
equation for flow, hydraulic diameter, D will be used for the ring gap cross section and the
analysis is based on a circular pipe.

_ 2 (endgapxedge)
D — endgap + edge (4'14)

Drawing the control volume and applying a fluid statics analysis to it as shown in

Figure 4.6 yields

(p1)1Ir2 — (p1 — Ap)1tr2 — (t)21tr€ = o (4.15)

3912 = 23 (4.16)
f r

where the acceleration is zero during fully developed flow when the velocity is constant.

Equation 4.16 represents the basic balance in forces needed to drive each fluid

particle along the channel in the ring gap with constant velocity. Since neither Ap nor 6 are
. . . . 2 . .
filnctrons of the radial coordlnate, r, it follows that -r—I must be mdependent of r. Hence, 1t

can be concluded that

40

‘11-}

= Constant
Applying the boundary conditions
1: : 0 at r = 0 and 1: = tw, wall shear stress at r = D/2

21 r
=>t= w (4.17)
D

 

12nr€

 

r—q—o

p11tr'2 (p1-Ap)1rr2 x

 

 

 

 

 

FIGURE 4.6 Free-Body Diagram of A Cylinder Fluid

From Equation 4.16 and Equation 4.17, the pressure drop and wall shear stress

are related by

Ap 7);” (4.18)

For laminar flow of a Newtonian fluid, the shear stress is proportional to the

velocity gradient. In the notation associated with ring gap channel flow, the relationship is

t = 41% (4.19)

The negative sign is included to give 1: > 0 with a’u/a’r < 0 (the velocity decreases

from the centerline of channel to the wall.) Neglecting the gravitational effect, Equation

41

4.16 and Equation 4.19 are the two governing laws for a fully developed laminar flow of a

Newtonian fluid. Combining these two equations, we obtain

(it: - _ &)
dr — (21.16 r (4.20)

and we integrate it to give the velocity profile

jdu = —§%jrdr

OI'

__ All 2
— (41%} +Cl

where C1 is a constant. Applying boundary condition where u = 0 at r = D/2,

yields

C1 = (146%?)02
u<r>=(i:—:§)lI-eri

The maximum velocity is the centerline velocity, Vc = ApD2/16uf. The flowrate

can be obtained by integrating the velocity with the cross-sectional area

Q = IudA = Iu(r)21rrdr (4.22)

42

2
Q = “08 VC (4.23)

 

By definition, the average velocity is the flowrate divided by the cross-sectional

area, V = Q/A, so that for this flow

 

 

 

2 2
_1tDVc4 _Vc_ApD
Vave — 8 (RD‘?) '7 2 - 32“! (4.24)
and
1tD4Ap
Q = 1281.16 (4.25)

Equation 4.25 represents the formula for fillly developed laminar flow of oil

through the top compression ring gap where

Q volume flow rate of oil through ring gap

D hydraulic diameter of ring gap channel

Ap pressure difference between the second land of ring pack and combustion
chamber

11 viscosity of engine oil

i ring width, the distance travelled by the oil flow

4.4 Total Oil Consumption

Equation 4.1 and Equation 4.25 represent two formulas of the oil consumption

mechanism: oil vaporization and reverse flow through ring gap. Adding the oil loss from

43

these two mechanisms allows the prediction of the total oil consumption during a specific

time of engine operation.

Total Oil Consumption 2 Oil + Oilrev (4.26)

evap

In this thesis, it is assumed that vaporization and reverse oil flow are the only two
sources of oil loss. This assumption is made for the purpose that one can study the

behavior of oil consumption under the influence of vaporization and reverse flow through

ring gap.

44

Chapter 5 SIMULATION ANALYSIS OF OIL CONSUMPTION
AND PISTON RING WEAR IN A 4.6L,V8 ENGINE

5.1 Introduction

This chapter presents the results of a simulation analysis of the ring pack behavior,
oil film thickness, oil consumption and ring wear, and focus mainly on the relationships
among them. In this thesis, the developed computer programs, RING, TWIST, PISTON,
WEAR and OILCONSUME were used in combination with an engine thermodynamic
model, the Ford Motor Company General Engine Simulation (GESIM). RING, TWIST
and PISTON are parts of the Cylinder kit Analysis System for Engines (CASE) while

WEAR and OILCONSUME are programs developed to be implemented into the CASE.

RING is a model used to predict groove and inter-ring gas pressure in a piston ring
pack for reciprocating gasoline and diesel engines. It also predicts the oil film thickness
between piston and cylinder bore, and piston ring and cylinder bore. This program
requires the description of the engine intended to be studied. PISTON is a model used to
predict the piston tilt per engine cycle. TWIST is a model used to predict the piston ring
twist behavior. WEAR is a model used to compute the amount of worn volume from the
first compression ring surface for a complete cycle and it can be integrated for specified
time interval. OILCONSUME is a model used to compute the oil consumption caused by

vaporization and reverse flow.

The GESIM program is a computational tool used in the Ford ENGine SIMulation

system (ENGSIM) to model in-cylinder processes in an IC engine. The simulation results

45

include in-cylinder flow, heat transfer, and pressure vs. crank angle during engine cycle. It
accounts for heat transfer losses from the combustion chamber during intake,
compression, expansion, and exhaust stroke for both burned and unburned gases during

combustion.

5.2 Integration of Different Computer Models

For this analysis, GESIM was used to generate simulation data of a 4.6L V8 IC
engine. The data include in-cylinder pressure, in-cylinder temperature, bore temperature,
piston temperature and in-cylinder heat convection coefficient. The engine description and
the generated data were the input for RING, TWIST, PISTON, WEAR and
OILCONSUME. Data of 011 film thickness, pressure load of compression ring, inter-ring
gas pressure, ring twist and piston tilt were generated by RING, PISTON and TWIST.
These data were input into WEAR and OILCONSUME. The results of simulation include
blow-by, oil film thickness, oil consumption rate and wear rate. For this thesis, nine

simulated cases were selected as shown in Table 5.1.

TABLE 5.1 Selected Simulation Cases
Case Operating Condition
1 1600rpm WOT (0.0 inches manifold vacuum)

1600rpm 50.0% Load (12.0 inches manifold vacuum)
1600rpm 16.67% Load (20.0 inches manifold vacuum)
3200rpm WOT (0.0 inches manifold vacuum)
3200rpm 50.0% Load (12.0 inches manifold vacuum)
3200rpm 16.67% Load (20.0 inches manifold vacuum)
5000rpm WOT (0.0 inches manifold vacuum)
5000rpm 50.0% Load (12.0 inches manifold vacuum)
5000rpm 16.67% Load (20.0 inches manifold vacuum)

 

 

 

 

 

 

 

 

 

 

 

 

 

OOOVQMAUJN

 

46

5.3 Engine Specification of A 4.6L-8 Cylinder IC Engine

The basic specification of the engine model used in this work is shown in Table 5.2.

TABLE 5.2 Engine Data

 

 

 

 

 

 

 

Parameter Specified Value

Engine Speed 1600rpm, 3200rpm, 5000rpm

Load 16.67%(20 ian), 50.0%(12 ian), WOT(0 ian)
Bore diameter 90.4 mm

Stroke 90.0 mm

Connecting rod length 150.7 mm

Connecting rod mass 577 g

 

Number of rings Two compressron rings and one 3-p1ece 011 control

 

 

 

 

 

 

 

ring
Piston diameter 90.2 mm
Piston height 56.4 mm
Piston mass 356 g
agate; Compression Ring 1.47 mm
llzlifgrrlrlueigh'fop Compressron 3.83 mm
Upper Compression Ring 10.6 g
Mass
Upper Compression Ring 24.0 N
Tensron

 

 

 

 

The top compression ring has been divided into 9000 elements as shown in Figure
5.1. Each element is identified with Segment ID and Slice ID for the calculation and

analysis purpose.

47

  

U pper Compression Ring

Slice I

\J...
4.1m.

\\ I”

“—0

 

FIGURE 5.1 Identification Numbering for Piston Ring

5.4 Simulation Results

The results presented here is separated into two parts: comparison of results at
different operating conditions of the engine and comparison of the transient results. It is
important for one to notice the trend of each individual result affecting each other under
the same or different conditions as time increases. It needs to be emphasized here that only
the upper compression ring is considered in the wear analysis. The upper compression ring
is the focus because of the most severe environment it is exposed to compared to the

second ring and the oil ring.

48

5.5 Results and Discussions

5.5.1 Cylinder Pressures

The cylinder gas pressures versus engine crank angle for all simulated cases are
shown in Figure 5.2. Comparing the cylinder pressures at 1600rpm (Figure 5.3), 3200rpm
(Figure 5.4), and 5000rpm (Figure 5.5) indicated that the engine speed did not
significantly affect the magnitude of the cylinder peak pressure. However, the engine load
had significant effect on the peak pressure. Higher cylinder peak pressure was observed at
higher engine load. For all three engine speeds, 1600rpm, 3200rpm and 5000rpm, the
highest cylinder peak pressure was observed at 0.0 inch manifold vacuum (100% load),
and the lowest at 20.0 inch manifold vacuum (16.67% load). The cylinder pressure peaked

at about 365 degree crank angle, due to the fuel ignition during the power stroke.

 

 

 

 

moo Cylinder Pressure d4.6L V8 10 En®e
6000 .

,_, 5000 -

:1"

as 4000 ~

Q

g 3000 ~

61' 2000 -
1000

0 CA deg
0 180 360 W 720

--¢--16m_mlnch ~~°~1Bm.12lnch ~.~ 1600.20hch
- -x—32m_mhch -—-—3200_12lnch .. 3200mm
WSWJJ Inch SMJZ hch ---- 5000.20 Inch

 

FIGURE 5.2 Cylinder Pressure at Different Operating Conditions

*Images in this thesis are presented in color

49

 

Cylinder Pressue of 4.6L VB 10 Engine

7000-

Pressue(kPa)

 

 

 

 

 

 

 

 

 

- - --' -- -« CA deg
0 180 360 5$ 720
- -<> - 1600_0 Inch - 4 - 1600_12 Inch . - 1600_20 Inch
FIGURE 5.3 Cylinder Pressure at 1600rpm
Cylinder Pressue of 4.6L V8 10 Engine
7000 r
E
i
6:
CA deg
0 180 360 54) 720
- ~x — 3200_00 Inch - - - - 3200_12 hch ~ - 3200_20 Inch

FIGURE 5.4 Cylinder Pressure at 3200rpm

50

 

Cylinder Pressure 01 4.6L V8 10 Engine
7000 ’

6000 ~
5000 -
4000 -
3000 -
2000 -
1000 -

Pressure(kPa)

 

 

 

CA deg

 

0 180 360 540 720

W 5000_00 Inch -—- 5000_12 Inch --— 5000_20 Inch

FIGURE 5.5 Cylinder Pressure at 5000rpm

5.5.2 Oil Film Thickness

The cylinder pressures influence the piston ring pack dynamics including the oil
film formation. Using RING, the oil film thickness results were generated. The
comparison of the predicted oil film thickness for different simulated cases (1600rpm 0.0
inch, 1600rpm 12.0 inch, 1600rpm 20.0 inch, 3200rpm 0.0 inch, 3200rpm 12.0 inch,
3200rpm 20.0 inch, 5000rpm 0.0 inch, 5000rpm 12.0 inch, 5000rpm 20.0 inch manifold
vacuum) was shown in Figrre 5.6. The oil film reached its maximum value when the
piston was approaching TDC. At a same engine load, the highest maximum oil film
thickness was observed at 5000rpm, and the lowest at 1600rpm. On the other hand, at a
same engine speed, the highest maximum oil film thickness was observed at 20.0 inch
manifold vacuum, and the lowest maximum oil film was at 0.0 inch manifold vacuum. For

this engine, with only the speed and the load varied, the minimum oil film thickness

51

values for all operating conditions were the same at about 1.40um. This result concluded
that increasing the engine speed and reducing the engine load can increase the overall oil
film thickness. In general, sufficient thickness of the oil film is necessary at high engine

speed and high load to ensure efficient lubrication system.

 

Oil Film Thickness Per En ine 0 tie
8.000E-03 ~ g I

7.000e-03 _
6.000E-03 .
5.000503 -
4.000E-03
30005-03 ..
20005-03
1000503

000015400 1 . . CA deg
0 130 360 540 720

1800.00 inch - -- I - -1600_12 itch --— 10m.” itch

 

 

 

Oiltilmhicknessmm

 

 

--9-'50(Il.0.0'mch --¢--5000_121nch - 0- SOWJUMIWQ]

FIGURE 5.6 Predicted Oil Film Thickness (lst Cycle)

The transient analysis was performed for the oil film thickness. Comparing the
transient results at 1600rpm 0.0 inch (Figure 5.7), 3200rpm 0.0 inch (Figure 5.8) and
5000rpm 0.0 inch manifold vacuum (Figure 5.9) showed that, over the simulation period
of 500 hours, the oil film thickness varied at different degree in magnitude under different
engine speeds and loads. The greatest change of the oil film formation was observed at the
operating condition of 5000rpm 0.0 inch manifold vacuum in comparison with other
cases. This is because under this condition, the piston ring face underwent the most severe

deformation (discussed in the later section). However, the variation of oil film thickness

52

over the simulation period was small as shown in Figures 5.7, 5.8 and 5.9. This indicates
that the oil film formation may not be sensitive to the worn piston ring face profile. This
can be due to the limitation of the input data line in describing the piston ring face profile

for RING program simulation.

Note: The simulation using RING program predicted the ring collapse at the initial
operating condition of 5000rpm 20 inch-manifold. Since the ring collapse phenomenon
violates some assumptions in analysis, this operating condition was eliminated from

simulations discussed in the following sections.

 

450503 _ Transient Change ofOll Film Thickness at 1600rpm10096 Load

4mem-

3505-03 - //\
3.00 E03 l l It ,l \
2.501503 -\ '
2005-03 4 \ \ /
1.50503 4 IL____S I. ,1

1005-03 .
5005-04 4

0.00E+00 r l i
0 180 360 540 720

 

Oil film thickness (mm)

CA deg

 

 

 

—0hr -—300hr ——500hr

FIGURE 5.7 Oil Film Development at 1600rpm 0.0 inch Manifold Vacuum

53

Oil flm h’ckness (mm)

 

5.30E-03 4
4.80E-03 -l
4.30E-03 -
3.80E-03 4
3.30E-03 -
2.80E-03 -
2.30E-03 ‘
1.80E-03 -

Transient Chance ofOil Film Thickness at 3200rom100% Load

 

 

 

 

1 .30 E-03
0

T T

540

____l" , i

180 360

—0hr ——300hr —5(Ilhr

720

CAdeg

FIGURE 5.8 Oil Film Development at 3200rpm 0.0 inch Manifold Vacuum

OiFllm Thickness (mm)

 

7.30E-03 3

6.30 1503 ~

5.30E-03 .

4.30E-03 -

3305-03 -

2.30E-03 ~

 

Transient Change 010“ Film Thickness a15000rpm100% Load

-'.~

 

 

 

1 .30 E03

0

540
—0hr -—300hr -—5mhr

720

CAdeg

FIGURE 5.9 Oil Film Development at 5000rpm 0.0 inch Manifold Vacuum

54

5.5.3 Average Blow-by

It is important to examine the blow-by of the engine under different operating
conditions because the blow-by tells the effectiveness of the piston ring pack in
controlling mass flow within the cylinder kit. RING was used to generate the blow-by
results. Figure 5.10 shows the average blow-by obtained after the first cycle of simulation.
The highest blow-by was observed at 5000rpm 0.0 inch manifold vacuum, where the
engine experienced the highest speed and load. The lowest blow-by was observed at
1600rpm 20.0 inch manifold vacuum. This can conclude that increasing both the engine

speed and the engine load will increase the blow-by proportionally.

2.000E-03
1 .500E-03
I .000E-03

5.000E-04

   
    

Average Blow-by (kgls)

5000rpm
3200rpm
1 600 rpm

0.000E+00

0.0 inch 12.0 inch

Manifold Vacuum
FIGURE 5.10 Prediction of Average Blow-by (Hour 0.0)

20 .0 inch

In order to study the effect of the piston ring wear and the oil consumption on the
blow-by, the average blow-by after 500 hours of simulation was obtained for all simulated
cases, as shown in Figure 5.11. Comparing Figure 5.10 and Figure 5.11 showed a
significant increase of blow-by over the 500 hours of simulation. The increase of the

average blow-by ranged from 3.0% at 3200rpm 20.0 inch to 50.0% at 5000rpm 0.0 inch

55

manifold vacuum (Figure 5.12). The increase of the blow-by may depend on the degree of
deformation experienced by the ring surface, as shown in Figure 5.34, Figure 5.35 and
Figure 5.36. Generally, the functionality of the ring face relies on the piston ring face
profile and the ring dimension. The wear caused the deformation on the ring face as well
as the reduction of the ring width. This increased the inter-ring volume in which the blow-
by was evaluated on. The transient wear analysis in the later part of this chapter showed
that the piston ring underwent the most severe change under the operating condition of
5000rpm 0 inch manifold pressure, in which the maximum ring width was reduced by
14%. The 14% reduction of the maximum ring width had contributed to 50% increase in
the inter-ring volume (groove and land volume) of upper compression ring, which was in
proportion to the increase of blow-by. On the other hand, the reduction of the ring width
was 0.8% under the condition of 3200rpm 20 inch manifold pressure, and it contributed to
the 3.0% increase of the inter-ring volume of the upper compression ring, which was in

proportion to the increase of blow-by.

20005-03
15035-03
iniolsoa ’

5.003504

   

Average Blow-by(kgls)

5003mm
3200rpm
1600rpm

0.0IIIE+00
0.1] inch

 

12.0 inch

20.0 inch
Manifold Vacuum

FIGURE 5.11 Prediction of Average Blow-by (Hour 500.0)

56

 

 
 
    
 
 

 

 

 

 

 

 

 

1.; 2000503 Comparison of Blow-by at Hr. 0. 0 and Hr. 500.0
E” 10005—03

s

.Q

5 1.200503 IHr00

g .Hr 500.0

8) 0.000504

:4

4 4.000504

 

 

 

0
W65» (0 fOperating

° 41>
'lv "
(o 6‘ Q 6‘ Q
{Q \Q 9% $90 §3 Conditions
«g N6 " tbs tbs EFQ
FIGURE 5.12 Comparison of Average Blow-by Increase

5.5.4 Piston Ring Wear

WEAR was used to generate the distribution of wear on the piston ring surface, the
transient change of the ring profile affected by the wear, and the wear rate over the period
of simulation. Comparing the wear distribution at 1600rpm 0.0 inch manifold vacuum
after the first cycle (Figure 5.13) and after 500 hours of simulation (Figure 5.14) yield two
observations. First, the wear effect at the beginning of simulation (Figure 5.13) was
observed to be concentrated around the center region of the piston ring (Slice ID 40 to 60)
and at the thrust and anti-thrust side (Segment ID 23 and 68). After 500 hours of
simulation (Figure 5.14), the wear effect distributed more evenly over the piston ring

surface.

57

    

IDOE'OQ »~- I .300510
1600rpm 0 inch 130509
:3 8.00E-10 (Hour 0‘0) ' 135005.10.
E 8.00E-10
”E
.g 8.00510 (34.00510.
0’ 6.00510
E 4mm0
E
I2.00E-10-
é 2.00E-10 88 4.00540
0.00E+00 59
<- locomo-
v 1 2005.10
wmwv Sam
amp ”meg 9
FIGURE 5.13 Wear at 1600rpm 0.0inch Manifold Vacuum(1st Simulation Cycle)

Second, the magnitude of the maximum wear volume decreased (80%) over the
500 hours of simulation. When the piston ring surface underwent the deformation, greater
area on the piston ring surface were experiencing the wear effect. With greater surface
area against the reacting forces from the cylinder bore, less stress was experienced by each
divided surface element on the piston ring (since stress is force divided by area). As a

result, the magnitude of the maximum wear volume was observed to decrease.

58

,. -~—-—». _-._. l8.0(E-10-

1.00509 I 1006-09
1600rpm flinch

8.00E-10 (Hour 500) I 135.005.10—

l 8.005.111

1] 4.01510
6 006-10

Volume Loss (mmslcycle)

I 2.00E-10-
4.00E-10

I0.00E400-
SEQ ID 2.00510

 

. a 0
Slice ID 2

FIGURE 5.14 Wear at 1600rpm 0.0inch Manifold Vacuum (After 500 hours)

Similar analysis was performed for lower engine loads at 1600rpm. Figures 5.15
and 5.16 are the wear distribution plots for 1600rpm 12 inch manifold vacuum, after the
first cycle and after 500 hours of simulation respectively. Figures 5.17 and 5.18 are the
plots for 1600rpm 20 inch manifold vacuum. Comparing the wear distribution plots of
different engine loads at 1600rpm showed that the magnitude of the maximum wear
volume decreased as the engine load decreased. The highest magnitude of wear volume

was observed at 100% load and the lowest at 16.67% load.

59

I 7 .00E-10-
8 .00E-10
I 6 IKE-10-
7 012-10
I SHE-10-
6 005-10
I 4015-10-
5 .01E-10
El 3 .OE-10-
4.01E-10

D 2 .01E-10-
3 .0fE-10

I 1 .01E-10-
2.0(E-10

Volume Loss (mmatcycle)

I 0.0E+00-
1 .0IE-10

 

FIGURE 5.15 Wear at 1600rpm 12.01nch Manifold Vacuum(lst Simulation Cycle)

I 7006-10-
_ : ONE-10
{a 1500rpm1 . ' “NE-1°-
i‘ (Hour 500) ' ’ 71105-10
_ . ., Isms-10
" 6.01510
lune-10
Silt-10
D 3.11510
HIE-10

El HIE-10-
“IE-10

I 1 III-10-
2 IKE-10

I 0.00800
1 .IIE-10

Volume Loss (mmalcycle)

 

FIGURE 5.16 Wear at 1600rpm 12.0inch Manifold Vacuum (After 500 hours)

60

Volume Loss (mmalcycle)

FIGURE 5.17 Wear at 1600rpm 20.0inch Manifold Vacuum(1st Simulation Cycle)

Volume Loss (mm3loycle)

FIGURE 5.18 Wear at 1600rpm 20.0inch Manifold Vacuum (After 500 hours)

an
o
O
r.“
_L
o
a

J

 
    

1600rpm 20 inch -
(Hour 0.0)

  

[I
300E—10 ; ll;

l ly'hlllgllll‘.

l
I.
.V .‘u

,J-‘w ‘ — Wfi._

_1600rpm20lnch ’ I ' '
(Hour500) ' ‘9 ~-

 

. 5 0
Slice ID 9

61

I “IE-10-
SHE-10

El SEE-10-
HIE-10

El ZOE-10-
3 .lIE-10

I 1 .OE-10-
HIE-10

I 0.0%“!!-
HIE-10

I “IE-10-
SHE-10

D SHE-10-
4.00E-10

D 2.00E-10-
3.l'IE-10

I1.lI£-10-
HIE-10

I 011E441)-
1 .EIE-10

Figures 5.19 and 5.20 are the wear distribution plots for 3200rpm 0.0 inch at two
different time periods, after the first simulation cycle and afier 500 hours of simulation.
Comparing with the lower load conditions such as 12.0 inch manifold vacuum (Figures
5.21 and 5.22) and 20 inch manifold vacuum (Figures 5.23 and 5.24) at the same engine
speed of 3200rpm, similar trends as the wear results at 1600rpm were also observed. The
wear distribution was initially concentrated at the middle range of the piston ring surface,

and later distributed more evenly over the surface at lower magnitude.

       
   
     

1.20309 "”7”“ 11 l - . ,1 m? j “NE-0%
3200rpm flinch -; 1-205-09
if 1 00E-09 (Hour 00) loose-1o
3 1005419
U
"E 800510 06.005.10-
E sous-1o
F; 6.00E-10
8 04.005.10-
E 400510 sous-w
5' l2.ooE.1o
’9 200510 3 88 4005.10
0.00E+00 59 locomo-
‘— zoos—10

‘_ 1
Slice lD °° E 839 '0

FIGURE 5.19 Wear at 3200rpm 0.0inch Manifold Vacuum(lst Simulation Cycle)

62

I1.ll.'E-09-
1.2(E~09

I BEE-10-
1 IKE-09

El HIE-10-
HIE-10

D 4.00611}
HIE-10

Volume Loss (mmalcycle)

I ZINE-10-
“IE-10

I 0.[IE+00-
ZINE-10

 

8‘— ‘ 3 ID
SlicelD °°§ 39

FIGURE 5.20 Wear at 3200rpm 0.0inch Manifold Vacuum(After 500 hours)

I! 7 006-10-
8 HIE-10
I B DIE-10-
HIE-10
I {LINE-10o
SHE-10
I QUE-10-
5 .llE-10
Cl 311E40-
4.01510

El 2115-10-
3.01510

I 1 IKE-10-
2 III-10

I ODEeOO-
1 IKE-10

Volume Loss (mmalcycle)

 

FIGURE 5.21 Wear at 3200rpm 12.0inch Manifold Vacuum(lst Simulation Cycle)

63

I [OE-10-
8.0(E-10
I 6.01E-10-
? 005.10
I 5.011510-
6 JIM-10
I “IE-10-
5 IKE-10

. ;. ,y mooeao
w '~ . 4005.10
ozone—10.
3005.10

I 1 IKE-10
2 .0(E-10

loooem
389' '0 100510

I"
ind-Flu I. l|,-. .

Volume Loss (mm3lcycle)

 

Slice ID

FIGURE 5.22 Wear at 3200rpm 12.0inch Manifold Vacuum(After 500 hours)

    

. ‘ lune-10
1.40E 10 ..-. , » ;. mos-1o
g 1.20E-10 3200rpm 20 Inc , ' j l1.ooE-1o-

U .' . _
”In.” 1005-10 , , ”1:”
E g y x ' . .11.
g 8.00E 11 ;. ' . ‘ f . . ' woe-10
§ 6.00E-11 — ; . " 5’ 3 name-11.
_i - ‘ l -' ' 13' HIE-11

g 4.00E 11 .

2 ‘ .1 .. ‘1. D4005."—
; 2.00E-11 . ' 33 5005.11
0.00500 59 lame-11-

" 30 4.ooE-11
85m 1 Ioooewo.

“N ‘- Se ID '
Slice ID °° °° § 9 m‘"

FIGURE 5.23 Wear at 3200rpm 20.0inch Manifold Vacuum(lst Simulation Cycle)

_ ,-~ —-.__, I120E-10-
xx” ' “MW—1 Hoe-lo

. ‘* ' I1mE-10-

- 1. .
120E“) 3200rpm 20inch , 0:. 20510
100510‘ (HourSDU) _‘ ,1 [ ‘ I800E-11-
I ' * " '. * 1005.10
05,005.11.
8.00E-11

El 4.00E-11-
6.00E-11

Volume Loss (mm3lcycle)

I 2 HIE-11-
4.00E-11

loooem
N 1 2005.11
00 2,3 8 Say lD

 

Slice lD

FIGURE 5.24 Wear at 3200rpm 20.0inch Manifold Vacuum(After 500 hours)

The wear analysis for the engine speed of 5000rpm also gave similar observations
as discussed previously for 1600rpm and 3200 rpm. Figures 5.25 and 5.26 are the plots for

5000rpm 0.0 inch, and Figures 5.27 and 5.28 for 5000rpm 12 inch manifold pressure.

.1206-09-
1405.09
l1mE-09-
1.2(6-09
Imam.
" p __ 1.00509
110“} name-lo
lllrl .slll' . ~- lame-10

ll ll-

D 4006-10-
6 WG-10

Volume Loss (mmsfcycle)

I HIE-10-
4.00E-10

Ioooeioo.
8 399 ‘0 2005-10

 

Slice ID 5
FIGURE 5.25 Wear at 5000rpm 0.0inch Manifold Vacuum(lst Simulation Cycle)

65

I1.20E-09-

1.40E-08
5000rpm0 inch , . , ,7 ,1

(HourSUU) .. . ‘ . .4 _- I100E-09-

- - , .. 1205.09

l8.00E-10-

1005-09

[3 RUE-10-
BIKE-10
U 4.00510-
BIKE-10

Volume Loss (mm3lcycle)

I 2.wE-10-
“IE-10

 

N ‘_ loooem
co 0) 8 Seg lD 200510

Slice ID

FIGURE 5.26 Wear at 5000rpm 0.0inch Manifold Vacuum(After 500 hours)

    
       

fr, -- ~_‘_wm________5_ mime-lo
8.00E-10 * ' {2:111
5000rpm12inch ‘ '

E 7.00E-10 (Hour00) ,. 7' ‘ . 7006.10
:3" 6,00E-10 ' , , ' 'zfilg‘
E 5.00E-10 ’ I4.tlE-10-
g 400E10 .r m“
. l. - , ., . 03005.10
E 3.00E-10 ,1" ‘ l . 4005.10
% 2.00E-10 7V . . name-10
> 1 DOE-10 " H 88 3305-10
0.00E+00_ 59 I1.00E-10-
2005.10

“‘00
conF 1 Se ID IONS-00
Slice ID 00 m g g NOE-10

FIGURE 5.27 Wear at 5000rpm 12.0inch Manifold Vacuum(lst Simulation Cycle)

66

‘- ~ ~——-~.. .__ _ l:li.ooE-1o-
7 8.00610
5000rpm 12 inch ! “305.10.

7.00510 (Hour 5o0) l 1,005.10
I5.00E-10-
8.00E-10
“005.10
5.00E-10
l:l3.ooe.1o
4.0(E-10

nilli'ifiJJll-Jilln H‘ l:l2.00E-10-
acne-lo

Volume Loss (mm3lcycle)

I 1 00610-
2005-10

I 0.00800-
1 DOE-10

 

0
Slice ID 9
FIGURE 5.28 Wear at 5000rpm 12.0inch Manifold Vacuum(After 500 hours)

In order to study the transient characteristic of the piston ring wear, simulation was
performed over 500 hours of simulation period to obtain the piston ring wear rate. Since
different locations on the piston ring surface have different wear effect, the elements on
Segment ID of 45 (opposite the ring gap position) was selected for the wear rate to be
obtained. The elements on Segment ID of 45 were the elements that experienced less
effect from the piston tilt and ring twist so that the calculation of the wear rate would be
more stable. The average wear rate taken from all slices of Segment ID of 45 were plotted

in Figures 5.29 through 5.31 for different operating conditions.

Figure 5.29 shows the wear rate at 1600rpm under different engine loads. It was
observed that the wear rate of each operating condition decayed to their steady state values
at different time periods. The wear rate of 1600rpm 0.0 inch manifold vacuum reached its

steady state value afier about 100 hours, 1600rpm 12.0 inch manifold vacuum about 150

67

hours, and 1600rpm 20 inch manifold vacuum about 250 hours. In addition, it was
observed that increasing the engine load increased the wear rate during steady state.
During steady state, the highest wear rate (1.9e-4 mm/hr) was observed at 100% engine

load (0.0 inch manifold vacuum), and the lowest (8.0e-5 mm/hr) at 16.67% engine load.

At higher engine speeds, 3200rpm and 5000rpm, the wear rates under different
engine loads were plotted in Figures 5.30 and 5.31 respectively. Comparing the wear rates
of 3200rpm shown in Figure 5.30 showed that the wear rate of each engine load also
decayed to their steady state values at different time periods. The settling time periods for
3200rpm 0.0 inch, 12.0 inch and 20 inch manifold pressure were about 30 hours, 60 hours
and 100 hours respectively. F mm Figure 5.31, it was noticed that the settling time periods
of the wear rate were 20 hours and 40 hours for 5000rpm 100% load and 50% load

respectively.

It is interesting to note that the wear rates at the lower loads of each engine speed
stabilized at a level coincidentally lower than the other engine loads. The possible
explanation is that at lower engine load, the combined stress experienced by the piston

ring surface is low due to the low cylinder peak pressure (refer to Figure 5.2).

The decrease of the wear rate can also be explained using the Abbott Firestone
Curve in addition to the stress analysis discussed earlier in this section. Refer to the Abbott
Firestone Curve shown in Figure 3.5 in Chapter 3. Under the wear effect, the peaks of the
piston ring surface were reduced. This resulted in the increase of the bearing ratio. At the

microscopic scale, the increase of the bearing ratio means more real contact surface exists

68

between the piston ring face and the cylinder liner. The increase of the real contact surface
results in the reduction of stress (stress is force divided by area) because more surface
asperities are bearing the contact force at the same time. The decrease of the stress will

result in the decrease of the wear according to the wear formulation.

The steady-state wear rates in Figure 5.29(1600rpm), Figure 5.30(3200rpm) and
Figure 5.31(5000rpm) were compared (Figure 5.32) in order to draw an observation
associating the wear rate to the engine speed and load. It was observed that during the
steady state, increasing the engine load would result in the increase of the wear rate.
However, increasing the engine speed would not ensure the increase of the wear rate, as
shown in Figure 5.32. At the 16.67% load, rising the engine speed from 1600rpm to
3200rpm resulted in the reducing of the steady-state wear rate. This observation can be
explained using the oil film thickness analysis. It was discussed previously that increasing
the engine speed would increase the maximum oil film thickness. The increase of the
maximum oil film thickness ensured the occurrence of hydrodynamic lubrication, which
prevents wear. Based on this argument, it can be understood that the piston ring wear
could be reduced by increasing the engine speed(low range of engine speed), for the same
engine power output. This observation agrees with observation from Chung’s results[1],

which showed the increase of wear rate when the engine speed increased from 1200rpm to

2100mm.

69

9.00000 E-04
8.00000 E°04
7.00000 E-04
6.00000 E-04
5.00000 E-04
4.00000 E-04
3.00000 E-04

Wear Rate (mmfhr)

1 .00000 E-04 ~

0.00000E+00

1 .40000E-03
1 .20000E-03
1 .00000E-03
8.00000 E-04
6.00000 E-04

Wear Rate (mmfhr)

4.00000 E-04
2.00000E-04
0000605000

 

200000504

 

 

Wear Rate vs. Time

 

*1600100 "- 1600l12

 

 

' 1600120

 

 

Time (tr)

 

 

50 100 150 200 250 300 350 400 450

FIGURE 5.29 Wear Rate at 1600rpm

 

 

Wear Rate vs. Time

 

 

 

 

.‘"H——_g.--_..--—+---—..-_—..--——..--—u—.-——

.... 4..-—

«—.....4.._-

 

 

0

50100150200250300350400450500

FIGURE 5.30 Wear Rate at 3200rpm

70

500

Tme (hr)

4.00000E-03 -.,_. if7....V._.a._,._.fi_v_r_7 ”\___,_MM__;_#*
350000503 . Wear Rate vs. Time

3.00000E-03 .
2.50000E-03
2.00000E-03
1.500005-03 i
1.00000E-03 i.
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o 50100150200250300350400450500

FIGURE 5.31 Wear Rate at 5000rpm

 

[MN 5000100 N...” 500011 2

 

Wear Rate (mmfhr)

 

 

 

 

 

0.0012
0 001 ‘ ' "Comparison of The Steady State Wear Rate. """"

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Operating Conditions
FIGURE 5.32 Comparison of Steady State Wear Rate

Experimental results from Barkman[2] showed a similar trend in the measured
wear rate as displayed in Figure 5.33. The measurement of the piston ring wear rate was
performed on a single cylinder engine using radiotracer-based wear diagnostic method.
The wear rate decreased to a steady state value within 20 hours of simulation. A step input
at the 25th hour of operation increased the wear rate quickly, but the wear rate returned to

a steady state value within 10 hours.

71

5.00E-01 ~-
450501 at Measured Wear Rate vs. Run Time

4.00E-01 i
350601 ' Compression Ratio
300501 ’ Increase
2.50E-01 \L
2.00E-01
1.50E-01
1.00E-01
5.00E-02 -
0.00E+CIJ . ‘ . . Hours Run
0.00 10.00 20.00 30.00 40.00 50.00

FIGURE 5.33 Measured Wear Rate

 

Wear Rate (mm/hr)

    
   

 

 

   

 

Besides the wear distribution and the wear rate analyses, the effect of the wear on
the piston ring face profile is also presented in this section. The following figures display
the change of the cross-sectional ring face profile in 3-D(left) and 2-D(right) views for
different simulated cases. The 45th segment of the piston ring was selected for this
analysis. The 2-D plot shows the cross section of the piston ring profile at the 45th
segment where the x-axis is the ring width and the y-axis the ring height. The legend of the
2-D plot contains the curves for the ring face profile at different time period during the
simulation. The 3-D plot has the same description as the 2-D except with the time

reference on the z-axis.

The 2-D and 3-D views of the transient change of upper compression ring face
profiles at 45th Segment ID are displayed in Figures 5.34 through 5.36. Figure 5.34 show
how the ring face profile evolve over the period of simulation at 1600rpm. It was observed
that higher engine load caused greater deformation on the piston ring profile. Indeed, this

observation was already expected from the wear rate analysis presented in previous

72

section. It was also observed that the ring face profiles for all simulated cases changed

from the original parabolic-shaped to linear taper-shaped.

 

 

        
  

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FIGURE 5.34 Ring Race Change at 1600rpm

Figure 5.35 shows evolvement of the piston ring face profile at 3200rpm. Similar
to the simulation at 1600rpm, as expected, the degree of deformation was directly affected

by the engine load. The most severe deformation occurred at the 100% load, and the least

73

at 16.67% load. As discussed in the wear rate analysis, the least deformation was observed

at 3200rpm 16.67% compared to all other simulated cases at different loads and speeds.

    
   

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FIGURE 5.35 Ring Face Change at 3200rpm
Figure 5.36 shows the plots of the change of the piston ring profile at 5000rpm. As
expected by the wear rate analysis, the most severe deformation occurred under the
operating condition of 5000rpm 100% load in comparison with the simulated cases at

1600rpm (Figure 5.34) and 3200rpm (Figure 5.35).

74

Manifold Vaccum W

 

 

 

 

   

 

 

 

 

 

 

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FIGURE 5.36 Ring Face Change at 5000rpm

5.5.5 Oil Consumption

The oil consumption mechanisms evaluated in this thesis are the oil vaporization
and the reverse oil flow. Figure 5.37 showed the time-average oil consumption
components of different operating conditions. It was observed that increasing the engine
speed and reducing the engine load could increase the total oil consumption. The highest
total oil consumption rate was observed at the operating condition of 5000rpm 12 inch-

manifold pressure, while the lowest value was at 1600rpm 0.0 inch-manifold pressure.

75

For oil vaporization, it was observed that its rate increased when the engine speed
increased and the engine load decreased. This observation can be explained by using the
oil film thickness analysis. Comparing the oil film thickness result (Figure 5.6) with the
oil vaporization result (Figure 5.37), it was observed that the effect of the speed and load
on both the oil vaporization and the oil film thickness was the same. Since the oil
vaporization was evaluated using the oil film thickness, the increase of the maximum oil
film thickness (resulted from the increase the speed and the decrease the load) resulted in

the increase of the oil vaporization, as shown in Figure 5.37.

For the reverse oil flow, it was observed that its rate decreased when the engine
speed and the engine load increased. Comparing this results with the blow-by results gave
an interesting observation: the effect of the engine speed and the engine load on the
reverse oil flow is directly opposite from the effect on the blow-by (Figures 5.10, 5.11 and
5.12.) In contrast to the reverse oil flow, the blow-by increased when the engine speed and

the engine load increased.

The blow-by is the flow of combustion gas from the combustion chamber into the
engine crankcase, and its direction is opposite from the reverse oil flow. At high blow-by
state, there is high flow rate of combustion gas across the piston ring pack in the direction
from the combustion chamber to the engine crank. This means that the net pressure
gradient per cycle across the piston ring pack is negative for the calculation of the reverse

oil flow. As a result, the reverse oil flow per cycle was reduced.

Of the two oil consumption mechanisms studied, it was also observed that the oil

vaporization was dominant at high speed (5000rpm) while the reverse oil flow was the

76

dominant oil loss component at low engine speed (1600rpm). The temperature data fi'om
GESIM showed that the average operational temperature per cycle is directly proportional
to both engine speed and load. The higher the speed and load, the higher the operational
temperature. The oil vaporization mechanism dominated at high speed means that the heat
and mass transfer process with a concurrent high oil film thickness was playing a more
important role during high engine speed operation (higher operational temperature),
compared to lower engine speed operation (lower operational temperature.) The reverse

oil flow dominates as the oil vaporization rate is much lower for the low speed conditions.

 

 

 

 

 

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E 0-02 ' lReverse Flow
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Operating Conditions

FIGURE 5.37Comparison of Time-average Oil Consumption over 500hrs Simulation

Wong et al. [8] performed the oil vaporization analysis using a computer
simulation. Comparing the oil vaporization result obtained in this thesis, shown in Figure
5.38, with the result from Wong et al. [8] showed a good agreement. As mentioned in
Chapter 1, Wong et al. [8] found that most (74%) of the oil vaporization occurred during

the intake and compression strokes. The oil vaporization result in Figure 5.38 shows that

77

there was high percentage of oil vaporization during the intake and compression strokes of
an engine cycle. Figure 5.38 shows that the lowest percentage (58%) of the oil
vaporization occurred during the intake and compression strokes under the operating
condition of 1600rpm 0.0 inch manifold vacuum, and the highest percentage (85%) under
the operating condition of 3200rpm 20 inch manifold vacuum. The high percentage of
vaporization during the intake and compression strokes was caused by the low oil vapor
pressure and high convective coefficients that encourage heat and mass transfer process

during these two strokes compared to the power and exhaust strokes.

 

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FIGURE 5.38 Estimated Oil Vaporization During One Cycle

Combining the reverse oil flow and oil vaporization gives the total oil
consumption, as shown in Figure 5.39. The total oil consumption for all operating
conditions underwent very small change over the period of simulation time. This

observation was obtained since the transient results on the oil film thickness also

78

underwent small variation over the simulation period (refer to Figures 5.7 through 5.9).
Moreover, the small effect of the worn ring surface on the overall ring pack pressure could
also contribute to the small variation in the transient oil consumption results. Without
including the wear effect on the piston ring tension, piston groove and cylinder bore, the
ring pack pressure would not be significantly affected. As a result, the oil consumption
rate, which is also evaluated using the ring pack pressure, did not have significant change

over the simulation period.

 

 

 

 

 

 

 

 

 

 

 

 

 

 

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FIGURE 5.39 Total Oil Consumption Rate for Reverse Flow and Vaporization

79

5.5.6 Oil Consumption and Gas Back Flow

The relationship between the gas back flow and the oil consumption was also
studied. The RING program was used to generate the gas back flow through the upper ring
gap over 500 hours of simulation. The results of the oil consumption rate and the gas back
flow for each simulated case were plotted in Figures 5.40 and 5.41. Figure 5.40 contains
the plots of the oil consumption rate (using the lefi y-axis) and the back flow (using the
right y-axis) for the first four cases: 1600rpm 0.0 inch, 1600rpm 12.0 inch, 1600rpm 20.0
inch and 3200rpm 0.0 inch manifold pressure. Figure 5.41 contains the plots for the other
four cases: 3200rpm 12 inch, 3200rpm 20 inch, 5000rpm 0.0 inch and 5000rpm 12.0 inch
manifold pressure. It was observed that a correlation existed between the gas back flow

through the upper ring gap and the total oil consumption.

The transient change of the oil consumption rate and of the gas back flow was
similar. When there was an increase in the gas back flow, there was also an increase in the
oil consumption rate over the same period of simulation and vice versa. This correlation
can be understood since the pressure difference of the piston ring pack is a driving force
for both mechanisms. In addition, the flow of both mechanism are along the same path in
the same direction, from the engine crank case to the combustion chamber. Thus, both

mechanism showed a similar trend over the simulation period.

80

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

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FIGURE 5.40 Oil Loss vs. Gas Backflow through lst Ring Gap (I)

 

 

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FIGURE 5.41 Oil Loss vs. Gas Backflow through lst Ring Gap (II)

82

5.5.7 Oil Consumption and Average Gas Blow-by

The research on the relationship between the blow-by and the oil consumption was
previously carried out at MSUERL by Ejakov [4]. This thesis carried out further
investigation on how the blow-by relates to the oil consumption. The RING program was
used to generate the blow-by over 500 hours of simulation. The plotting of the blow-by
results with the oil consumption rate are shown in Figure 5.42 and Figure 5.43. Figure
5.42 contains the plots of the oil consumption rate (using the lefi y-axis) and the blow-by
(using the right y-axis) for the first four cases: 1600rpm 0.0 inch, 1600rpm 12.0 inch,
1600rpm 20.0 inch and 3200rpm 0.0 inch manifold pressure. Figure 5.43 contains the
plots for the other four cases: 3200rpm 12 inch, 3200rpm 20 inch, 5000rpm 0.0 inch and

5000rpm 12.0 inch manifold pressure.

It was observed that the oil consumption and the blow-by results underwent
different trend of change over the simulation period depending on the engine speed. At
1600rpm, the oil consumption rate underwent relatively high increase over the simulation
period, but the blow-by did not. At 5000rpm, the blow-by underwent relatively high
increase but the oil consumption rate did not. However, at 3200rpm, the transient results of
both mechanism were found to be correlated, in which the change of the blow-by was

directly proportional to the change of the oil consumption.

With the limited number of cases performed in this work, no obvious conclusion
could be drawn regarding the transient relationship between the blow-by and the oil

consumption rate. This concept requires further study.

83

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

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FIGURE 5.43 Oil Loss vs. Gas Blow-by (II)
85

5.5.8 Oil Consumption and Piston Ring Wear

The transient results of the wear rate and the oil consumption rate for each
simulated cases were plotted in Figures 5.44 and 5.45. The number of simulation cases for
this analysis is the same as the previous two analyses. The purpose for this analysis was to

study how the oil consumption changed with respect to the wear rate.

First, it was observed that the oil consumption rate started to increase linearly
when the wear rate was at steady state. Prior to steady state wear rate, the oil consumption

rate increased non-linearly.

Second, the transition point decreased as the engine speed increased. At 1600rpm
conditions, the change of the oil consumption trend from nonlinear to linear was at about
the 100th hour of simulation. At 5000rpm conditions, the transition point was at the 10th
hour of operation. This primary investigation on the relationship between the piston ring
wear and the oil consumption showed that the piston ring wear is indeed an important
factor in the oil consumption, which was modeled using the oil vaporization and reverse

oil flow.

86

 

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FIGURE 5.45 Oil Loss vs. Wear Rate (1])

88

Chapter 6 SUMMARY

6.1 Conclusions

In this work, models of piston ring wear and oil consumption were formulated and
implemented in a computer model called the CASE system. With the CASE system, the
computational design and analysis for a specific engine cylinder kit geometry could be
performed. During the analysis presented here, source data from GESIM was developed
for a 4.6L, V8 Ford engine and used as input to the CASE system; WEAR and
OILCONSUME are the two subroutines, which were developed to be implemented in the

CASE system.

In addition to the wear and the oil consumption analyses, the oil film formation,
the blow-by and the gas back flow through upper ring gap were also studied in this thesis

since they are also important factors in the analysis of an engine performance.

The engine oil consumption model, which included oil vaporization and reverse oil
flow mechanisms, did not have significant change over 500 hours of simulation.
Practically, the oil consumption would be expected to increase as the age of an engine
increases. A significant increase in oil consumption was not observed in this study. This
may be due to the fact that not all oil loss mechanisms were included in the simulation.
Those oil loss mechanisms not included in this study are oil throw-off, oil carrying
through blow-by, valve stem leakage and so on. These mechanisms are more difficult

models in a realistic manner, but it will be necessary to include them into the oil

89

consumption model. Including these mechanisms into the oil consumption model should
give a more complete and practical analysis on the oil consumption. For example, the
RING generated blow-by results showed significant increase of blow-by over the
simulation period. Generally, it is known that the blow-by influences the oil consumption.
Therefore, implementing the mechanism of the oil carrying through blow-by would result

in the increase of the oil consumption rate over a long time period of engine simulation.

The piston ring wear results of this thesis showed consistency with Chung’s
mathematical results[1], Barkman’s experimental results [2] and Pint’s simulation results
[3]. High wear rate was observed at the beginning of the simulation and declined
exponentially to a steady state value over the simulation period. In addition, the wear was
observed to affect the region on the piston ring surface where there were high
concentrating stress and influence of the ring twist and piston tilt. The piston ring region
on the thrust and anti-thrust sides were observed having higher effect than other locations

on the piston ring.

This work also predicted the direct impact of the wear on the blow-by. It was
observed that the blow-by increased accordingly to the reduction rate of the ring width.
The reduction of the piston ring width due to wear contributed to the increase of the inter-
ring volume. Because the inter-ring pressure is a function of the inter-ring volume, and the
blow-by is evaluated from the inter-ring pressure, the increase of the inter-ring volume

over the simulation period had resulted in the increase of the blow-by.

The engine speed and the engine load were significant factors in affecting an

engine performance related to the oil consumption, the piston ring wear, the oil film

90

formation, the gas back flow and the gas blow-by. It was found in this thesis that both
speed and load affect the wear and oil consumption mechanism. This means that low
speed did not guarantee low rates of wear and oil consumption, and vice versa. A proper
combination of the engine speed and the engine load provides the minimum oil
consumption and wear rates. For example, as discussed in Chapter 5, the operating
condition of moderate speed(3200rpm) and low engine load (20 inch manifold vacuum)
resulted in better oil film formation that kept the piston ring surface at lower wear rate
than the condition at 1600rpm. High engine speed (5000rpm) had a high oil film

thickness, but did not help in keeping low wear rate.

6.2 Recommendation

Several recommendations were suggested for the future improvement in this

research.

1. Reexamination of oil film thickness sensitivity to ring face profile. It was observed in
the results that the oil film did not have obvious change even though the piston ring face
profile had gone through a drastic change. The input parameter of the piston ring face
profile should be increased so that a more complete and detailed description of the piston

ring can be evaluated by the RING program.

2. Improvement of the wear model by accommodating more dimension of contacting
surfaces. In this thesis, other than the piston ring face and cylinder wall interaction, no
other surface interaction are included. The wear related surface-interaction between the
piston groove and ring upper surface and bottom surfaces should be studied. Wear on

cylinder bore also needs to be included because the hardness and toughness on this surface

91

is lower (sofier) than ring face. In addition, sensitivity studies of surface characteristic on

the piston ring with respect to wear should be studied.

3. Improvement of the oil consumption model by accommodating other oil loss
mechanisms such as blow-by, oil throw-off, oil burning on the cylinder wall and piston,

and valve stem leakage.

4. Obtaining more accurate values for chemical data of engine oil species. Chemical
data such as binary diffusion coefficient of two phases of oil vapor and combustion gas

need to be obtained so that the oil evaporation model can be modeled more accurately.

92

BIBLIOGRAPHY

1 Chung, Y. “Development and Analysis of A Fire Ring Wear Model for A Piston
Engine”. Ph.D. dissertation, Michigan State University, 1992

2 Barkman, D.A. “An Evaluation of The Parameters Involved in Developing A
Radiotracer Based Wear Diagnostics System”, M.S. thesis, Michigan State University,
1992

3 Pint, S., Schock, H.J., “Design and Development of a Soflware Module for Analysis of
3-D Piston Ring Wear.” SAE Paper 2000-01-0920

4 Ejakov, Mikhail A. “Ring Pack Behavior and Oil Consumption Modeling in IC
Engines”. PHD dissertation, Michigan State University, 1998

5 Inoue, T., Maeda Y., Takeda M., Nakada M., “Study of Transient Oil Consumption of
Automotive Engine”. Toyota Motor Corp. 1989

6 Min, B.S., Kim, J.S., Oh, D.Y., Choi, J.K, Jin, J.H., “Dynamic Characteristics of Oil
Consumption Relationship Between the Instantaneous Oil Consumption and the Location
of Piston Ring Gap”. SAE Paper 982442, 1998.

7 Tian, T., Wong, V., Heywood, J.B., “A Piston Ring-Pack Film Thickness and Friction
Model for Multigrade Oils and Rough Surfaces.” SAE Paper 962032, 1996

8 Wong, V., Audette III W.A., “A Model for Estimating Oil Vaporization from the
Cylinder Liner as a Contributing Mechanism to Engine Oil Consumption.” SAE Paper
1999-01-1520, .999

9 Heywood, J.B., “Internal Combustion Engine Fundamentals,”McGraw-Hill, Inc., 1988

10 Nakashima K., Ishihara S., “Influence of Piston Ring Gaps on Lubricating Oil Flow
into the Combustion Chamber.” SAE Paper 92546, 1992

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11 Wilhoit, R.C., Zwolinski, B.J. “Handbook of Vapor Pressures and Heats of
Vaporization of Hydrocarbons and Related Compounds.” Texas A&M University, 1971.

12 White, Frank M., “Fluid Mechanics”.McGraw-Hill, New York, 4th ed., 1999

13 Wong, V., Audette III W.A., “Oil Transport Along the Engine Cylinder Liner
Coupling Vaporization and Film-Thickness Analysis.” Massachusetts Institute of
Technology, 2000

14 Munson, Young, Okiishi, “Fundamentals of Fluid Mechanics.” 3rd Ed, John Wiley
Sons, Inc. New York, 1998

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