PUBLICATJONS NATL INST OF STANDARDS & TECH R.I.C. IMERGE

Nation Is and Technology

(formerly National Bureau of Standards)

NIST Special Publication 755

The Design and Construction of a

State-of-the-Art High Temperature

Tribometer

Jeffrey P. Yellets, Stephen M. Hsu, and E. Erwin Klaus i

DOE ECUT Tribology Program
I

!

j

i
NIST Special Publication 755

The Design and Construction of a

State-of-the-Art High Temperature

Tribometer

Jeffrey P. Yellets and Stephen M. Hsu

Ceramics Division

Institute for Materials Science and Engineering

National Institute of Standards and Technology

(formerly National Bureau of Standards)

Gaithersburg, MD 20899

E. Erwin Klaus

Department of Chemical Engineering

The Pennsylvania State University

University Park, PA 16802

NOTE: As of 23 August 1988, the National Bureau of Standards (NBS) became the National Institute of Standaids and Technology (NIST) when President Reagan signed into law the Omnibus Trade and

Competitiveness Act

September 1988

U.S. Department of Commerce

C. William Verity, Secretary

National Institute of Standards and Technology

(formerly National Bureau of Standards)

Ernest Ambler, Director
Library of Congress

Catalog Card Number: 88-600579

National Institute of Standards

and Technology

Special Publication 755

212 pages (Sept. 1988)

CODEN: XNBSAV

U.S. Government Printing Office

Washington: 1988

For sale by the Superintendent

of Documents,

U.S. Government Printing Office,

Washington, DC 20402

For sale by the Superintendent of Documents, U.S. Government Printing Office Washington, D.C. 20402
FOREWORD

This report is born out of a cooperative effort between the Chemical

Engineering Department, at Pennsylvania State University (PSU) , the

Tribology Group of the National Bureau of Standards, and the partial

support of the DOE ECUT Tribology Program. For years, the NBS has

encouraged and sponsored graduate cooperative programs in which

promising young scientists (some of them NBS staff members) come to

NBS to conduct research, part of which often becomes the graduate's

thesis. While it is not the objective of NBS to teach and train

graduate students , such a program often furthers NBS programmatic

goals by attracting high caliber scientists working on areas where NBS

mission lies. Many excellent research papers have results and many

students, upon graduation, have chosen to stay at NBS to continue

their research careers. Such programs also draw many first rate

university professors to NBS through the participation of these

students in NBS research programs and projects, thus fostering NBSuniversity interactions and enhancing the scientific caliber of the

work and reputation at NBS and the participating university. At the same time, through the frequent contacts that NBS has with industries,

a natural university/government lab/industry relationship evolves,

bringing a team focus on many research projects of significant

economic and/technological impacts.

In 1984, primarily through the interactions of Dr. Stephen M. Hsu of

NBS and Professors Elmer Klaus and Larry Duda of PSU, a cooperative

program in Tribology was started. Three graduate students: Mr.

Richard Gates (a NBS staff member), Mr. Jeffrey Yellets, and Mr.

Douglas Deckman were enrolled at the Chemical Engineering Department

at PSU. Three reports have been prepared to describe the fruit of

their relentless efforts in the last three years. These studies were conducted at NBS under the guidance of Dr. Stephen Hsu with the close

participation of Prof. Elmer Klaus who visited NBS frequently.

Dr. James Eberhardt, Mrs. Terry Levinson, and Mr. David Mello of the

DOE ECUT Tribology Program have at the same time sponsored a major

Tribology program at NBS. That program also benefited from the

studies conducted by Mr. Gates, Mr. Yellets, and Mr. Deckman. These

students, while not working on sponsored projects directly, enabled

NBS to explore some high risk, high pay off projects parallel to ECUT

projects. When appropriate, the students were supported by ECUT for some time. To this, we gratefully acknowledge the generous support of

DOE ECUT, without whose support many ideas would not be explored.

Stephen M. Hsu

Chief, Ceramics Division

ABSTRACT

High temperature ceramic tribology is one of the fastest growing

and least understood areas in tribology. The need to understand the

mechanisms of friction and wear for ceramic materials is critical.

Ceramic materials are being utilized in tool development, bearing

design, materials development for low-heat-rejection diesels,

automotive gas turbines, Stirling engines, and aerospace applications.

The development of a ceramics industry capable of manufacturing high

technology wear resistant ceramics is essential in the United States,

for if we do not develop such an industry, we will lose a substantial

market to an already rapidly growing foreign technology.

This effort describes the design, development, and construction

of a unique, state-of-the-art, high temperature tribometer for the

mechanistic study of novel ceramic materials. An overview of ceramic

tribology is included in this work as well as a review of many common

wear test configurations. The final designs and the construction of

the High Temperature Wear Facility is presented in this work.

Recommendations for future work on high temperature wear mechanisms of

ceramic materials and the tribological testing of such materials are

presented at the conclusion of this work.

V

TABLE OF CONTENTS

Page

FOREWORD iii

ABSTRACT v

LIST OF TABLES xi

LIST OF FIGURES xiii

Chapter

1 INTRODUCTION 1

2 A LITERATURE REVIEW OF CERAMIC TRIBOLOGY 4

2.1 Introduction 4

2.2 An Overview of Ceramic Tribological

Literature 5

2.2.1 Bearing Studies 5 2.2.2 Engine Studies 9 2.2.3 Tool Studies 15

2.2.4 Grinding, Polishing, and Other Studies . . 17

2.2.5 Ceramic Lubrication 18

2.2.5.1 Solid Lubrication 19

2.2.5.2 Other Techniques 21

2.3 Summary/Conclusions 23

3 DETERMINATION OF DESIGN PARAMETERS AND

SPECIFICATIONS 28

3.1 Wear Test Configuration 30

3.2 Temperature Limit 30

3.3 Environment and Lubrication 31

3.4 Speed Ranges 32

3.5 Load Range 33

vii
TABLE OF CONTENTS (continued)

Page

3.6 Summary of Specifications 34

3.7 Introduction to Design 36

4 HEAT TRANSFER 37

4.1 Heat Transfer Calculations 39

5 MATERIALS SELECTION FOR THE HIGH TEMPERATURE

WEAR FACILITY 54

5.1 Introduction 54

5.2 Partially Stabilized Zirconia 55

5.3 Silicon Nitride 57

5.4 Silicon Carbide 59

6 FLUID DYNAMICS 64

7 MECHANICAL DESIGN OF COMPONENTS 77

7.1 Design of Ceramic/Metal Components to

Accommodate Differential Thermal Expansion .... 77

7.2 Design and Selection of the High

Temperature Wear Facility Drive Mechanism .... 84

7.2.1 Selection of a Drive Motor 84

7.2.2 Selection of an Alignment System 85

7.3 Design of the High Temperature Wear

Facility Machine Framework 86

8 DESCRIPTION OF THE HIGH TEMPERATURE WEAR FACILITY ... 87

9 INSTALLATION AND INITIAL TESTING 105

9.1 Installation of Equipment 105

9.2 Testing of Equipment 109

9.2.1 Preliminary Testing 109

9.2.2 Friction Testing - Integrated Mechanical

Wear Facility Testing 110

viii
TABLE OF CONTENTS (continued)

Page

10 CONCLUSIONS AND RECOMMENDATIONS FOR FUTURE

WORK 118

REFERENCES 121

APPENDIX A. A REVIEW OF WEAR TEST CONFIGURATIONS 131

APPENDIX B. DEVELOPMENT OF FINITE DIFFERENCE EQUATIONS

FOR HIGH TEMPERATURE WEAR FACILITY

FURNACE DESIGN 159

APPENDIX C. GAS BEARING DESIGN CALCULATION FOR THE HIGH

TEMPERATURE WEAR FACILITY 175

ix

LIST OF TABLES

Page

Table 1: Bearing Studies: A Summary of Conditions 7

Table 2: A Summary of Heat Engine Studies 12

Table 3: Tool Studies: A Summary of Parameters 16

Table 4: A List of Materials Examined from H. E. Sliney's

"Solid Lubricant Materials for High Temperatures - A Review" 20

Table 5: Ceramic Mechanical Characterization Parameters .... 25

Table 6: High Temperature Wear Facility Design

Specifications 35

Table 7: Properties of Selected Ceramic Materials 56

Table 8: A Typical Properties Table of a-SiC 62

Table 9: T^/F^ Ratios for "V" Shaped Bearing Design 72

xi

LIST OF FIGURES

Page

Figure 1: General Schematic of Space Heater System 40

Figure 2: Heat Staging Scheme 43

Figure 3: Porous Insulators as Radiation Shields 45

Figure 4: Top View of High Temperature Furnace Design 47

Figure 5: Side View (3-D) of Insulation 48

Figure 6: Smallest Symmetric Element for High Temperature

Furnace Design 49

Figure 7: Summary of Boundary Conditions for High Temperature

Wear Facility Furnace Design 51

Figure 8: Results of Finite Element Analysis 53

Figure 9: Linear Thermal Expansion Curve of PSZ from

0-1500°C 58

Figure 10: Tensile Strength of Materials to 1500°C 60

Figure 11: Original Preliminary Hydrostatic Gas Bearing Layout

(Side View) 68

Figure 12: "V"- shaped Hydrostatic Gas Bearing Layout 70

Figure 13a: Side View of Hydrostatic Bearing 73

Figure 13b: Top View of Hydrostatic Bearing 73

Figure 14: Inherent and Orifice Compensated Supply Holes

(Sectional View) 74

Figure 15a: Tensile Hoop Stress on a Ceramic Tube Interference

Fit (Top View, Sectional) 79

Figure 15b: Compressive Hoop Stress on a Ceramic Tube

Interference Fit (Side View, Sectional) 80

Figure 16: General Schematic of Drive System 85

xiii
LIST OF FIGURES (continued)

Page

Figure 17: Overall Schematic Diagram of the Wear Tester 88

Figure 18: Unidirectional Test Configuration 89

Figure 19: Unidirectional Wear Test Configuration 90

Figure 20: Pin-on-Disk Configuration 91

Figure 21: Lower Ceramic Specimen Holder Assembly 92

Figure 22: Upper Sample Holder Diagram 94

Figure 23: View of Complete Upper Assembly 95

Figure 24: Furnace Chamber and Heating Elements 97

Figure 25: Design Layout - Upper and Lower Assemblies

and Furnace 98

Figure 26: Air Bearing Spindle Assembly 100

Figure 27: General Schematic of Drive and Spindle System .... 101

Figure 28: Front View - Motor Spindle System 102

Figure 29: Side View - Motor Spindle System 103

Figure 30: Addition of Motor Assembly - Front View 104

Figure 31: Frictional Testing Plan 110

Figure 32: Initial Furnace Testing Ramp Up 112

Figure 33: Furnace Temperature Ramping Program 113

Figure 34: Unidirectional Speed Calibration 114

Figure 35: Force Transducer Calibration 115

Figure 36: Initial Room Temperature Friction Data 116

Figure 37: Room Temperature Vibrational Friction Transients . . . 117

Figure 38: Two-Ball Contact 133

xiv
LIST OF FIGURES (continued)

Page

Figure 39: Ball on Flat/Ball on Disk Configuration 134

Figure 40: Ball on Ring Configuration 135

Figure 41: Ball on Cylinder Configuration 136

Figure 42: Fin-on-Disk - Rim Loaded 137

Figure 43: Fin-on-Disk - Face Loaded 138

Figure 44a: Pin-on-Disk - Fretting Motion 139

Figure 44b: Pin-on-Disk - Reciprocating Motion 140

Figure 45: Cylinder on Disk Rim Configuration 141

Figure 46a: Crossed Cylinder Device - Reciprocating Cylinder . . . 142

Figure 46b: Crossed Cylinder Device - Rotating Cylinder 143

Figure 47a: Four-Ball Wear Test Configuration 144

Figure 47b: Ball on 3-Flat Configuration 145

Figure 48: Five Ball Wear Test Configuration 146

Figure 49: Multiple Pin-on-Disk Configuration 147

Figure 50a: Disk Rim on Disk Rim 148

Figure 50b: Implementation #2 Disk Rim on Disk Rim 149

Figure 51: Rotating Cylinder on Plate Configuration 150

Figure 52: Ring-on-Block Configuration 151

Figure 53: Three Pellet on Disk Configuration 152

Figure 54: Ring on 2 Block Configuration 153

Figure 55a: Pellet on Reciprocating Plate 154

Figure 55b: Slider on Reciprocating Plate 155

Figure 56: Pellet on Disk Configuration 156

XV
LIST OF FIGURES (continued)

Page

Figure 57: Hydrostatic Bearing Design Surface 181

Figure 58: Hydrostatic Supply Surface Approximation 182

Figure 59: Fig. 5.3.13-MTI Gas Bearing Design Course 184

Figure 60: Fig. 5.3.5-MTI Gas Bearing Design Course 186

Figure 61: Fig. 5.3.9-MTI Gas Bearing Design Course 192

Figure 62: Fig. 5.3.19-MTI Gas Bearing Design Course 196

XVi
Chapter 1

INTRODUCTION

Ceramic materials are some of the oldest and most widely used

man-made materials known. Evidence suggests that the use of ceramic

pottery wheels began over 5000 years ago in the Near East.^ The first

tribological problems were found in the use of bitumen to lubricate

potters wheels around 3250 ± 250 B.C. at Ur of the Chaldees.^ Perhaps

the earliest recorded tribologist, an Egyptian slave pouring water in

front of a stone sledge that was transporting a statue of Ti (C. 2400

B.C.), was obviously intimately concerned with the friction of stone

on stone.^

Recently, man is entering a new stage of materials development,

the new stone age -- the age of advanced ceramics. These materials

are currently being explored for use in a number of tribological

applications

It has been estimated that in the U.S. alone, advanced ceramic

projects could add 6 billion dollars per year to the economy in 1990,

11 billion dollars in 1995, and 20 billion dollars in the year 2000.

If we assume linear industry growth in the periods 1990-1995 and 1995-

2000 consecutively, over 135 billion dollars may be added to our

economy in the coming decade. In addition to this boost to our

economy, it is estimated that, in the year 2000, if we have developed

wear and heat resistant ceramic materials we will save per year 1.5

billion dollars due to decrease in fuel consumption, 300 million
annually due to a reduction in maintenance costs, 530 million dollars

due to tool life increase and improved machining when ceramic tools

are developed, and greater than one billion dollars due to the use of

ceramic related materials in integrated optics.^

On the other hand, the impact of not having a United States

advanced ceramics program is great. In just the United States

transportation sector in 1985, more than 8.1 million cars and 3.5

million trucks were built. Between now and 1995, 11,000 to 12,000

aircraft engines will be built. ^ If a foreign competitor was to

dominate the development of a more economical and fuel efficient

ceramic engine, the U.S. based transportation industry might be at a

great competitive and monetary disadvantage.

Since the advanced ceramic industry seems to be an industry that

will be developed, tribologists , scientists, or researchers should

ascertain the barriers that need to be overcome in the development of

the U.S. ceramics program. Many of these properties relate to

tribology. According to the 1984 assessment report by the Charles

River Associates, some of these barriers are standardized testing

techniques, understanding and modeling of fracture behavior and time

dependent properties, standard reference materials, surface and bulk

chemistry, defect chemistry, and thermodynamic analysis.''

This report describes the systematic design and development of a

state-of-the-art high temperature wear tester, the High Temperature

Wear Facility. This apparatus is designed to measure friction and

wear of ceramic materials at high temperatures under an
environmentally controlled condition to understand the mechanisms of

wear of ceramics

The need for such a high temperature wear tester is supported by

an overview of the ceramic tribology literature. This gives guidance

to the initial wear test configuration, temperature limit, speed

limit, contact pressure, and environmental conditions. A review of

wear test configurations also has been conducted (see Appendix A for

more details) to help the initial choices.

Because of the very high temperature limits involved with

ceramics, key technical challenges exist: 1) Heat Transfer, 2)

Fluid Dynamics, 3) Materials Science, and 4) Mechanical and ThermoMechanical Design of Components. Sections on each of these areas and

the final design and construction of the High Temperature Wear

Facility are presented.
4

Chapter 2

A LITERATURE REVIEW OF CERAMIC TRIBOLOGY

2.1 Introduction

In recent years, many articles on ceramic tribology have

appeared. Unfortunately, a complete tribological database does not

exist. A literature survey is conducted here to: 1) assess the

current status of ceramic tribology information, 2) give a broad-based

information database on pre-test and post-test analysis of ceramic

tribological testing, 3) establish an information base for future

work, and 4) pin-point areas where improvement in methods and data

analysis may improve the useability and comparability of test methods

in ceramic tribology.

Since there are many potential applications for ceramics in heat

engines, turbines, bearings, machine tools, etc., the review is

divided into the following areas:

1) Bearing Studies

2) Engine Studies

3) Tool Studies

4) Grinding, Polishing, and Other Studies

5) Ceramic Lubrication

6) Conclusions
5

2.2 An Overview of Ceramic Tribological Literature

2.2.1 Bearing Studies

In ceramic tribology, one of the most critical applications is

the use of ceramics as bearing materials particularly rolling element

bearings. In one of the earlier studies by Sibley and Allen,® in

1962, a number of different ceramic materials were studied in order to

find a suitable material for high temperature bearing and seal

applications. To do this, sliding tests were conducted and several

interesting results were found. First, it was noted that thermal

streaking and hot spots occurred within the wear contact and that, in

general, the abundance of "hot spot" behavior increased the wear of

the test specimens.^ Second, it was noted that there appeared to be a

correlation between high friction and high wear of ceramic materials

and that the formation of surface films may indeed affect the

frictional behavior of the material. ^° In a study published the

following year, Taylor, Sibley, and Lawrence^ •' studied a ceramic

rolling bearing contact for three materials -- alumina, silicon

carbide, and titanium carbide. In this study it was shown that for a

standard bearing configuration, the elimination of vibration (even at

the expense of greatly increasing the load) could increase the life of

the bearing. It was also noted that superficial pitting and grain

removal occurred on the outer race of the bearing possibly indicating

that high surface stresses caused a structural breakdown of the

surface material. Baughman and Bamberger-'^ ^ also studied ceramic
boarinj',.': '''"^1 'hey found Lhat: the \or.r. of geometrical integrity due to

wear, plastic deformation, or catastrophic cracking couLd cause

failure of high temperature ceramic bearings. Other more recent

studies have been conduc ted . ^ ' ^ ^ Among these, Dalai- Chiu, and

Ral) i nowiczl'* found that for a silicon nitride ceramic, surface cracks

seem to be responsible for spall formation at high stress levels, and

Lucek, Sibley, and Rosenleib^^ found that for silicon nitride balls,

if the surface defects could be blunted, a much better tribological

material could be obtained.

For quick reference, a list of experimental conditions and

conclusions of references [8] to [23] is given in Table 1.

From the studies done in ceramic tribological bearing testing,

several fundamental principles for the testing of ceramic materials

emerge. First, in the design of ceramic components, extreme care

should be taken in bearing configuration and design. A well designed

metal bearing may perhaps become a poorly designed ceramic bearing if

the same design is followed. Differential thermal expansion between

the ceramic bearing and metal surrounding components may cause

vibration of the bearing system. A sharp edge on a ceramic ball

bearing cage may cause extremely high surface stresses in ceramic

components. This may cause failure of the bearing system. In eight

of the thirteen studies presented, major problems were attributed to

inadequate bearing design parameters . ^ ^ • ^ • ^ ^ • ^ ^ - ^ ° " ^ These problems

ranged from vibrations occurring due to inadequate design^ ^

, to poor

mounting of the inner raceway. ^° Therefore, to design a more

effective ceramic bearing tribological test, special care should be
7

Table 1

Bearing Studies: A Summary of Conditions

Ref

.

Speed Load/Pressure Temp. (°C) Slidine/RollinE

8 3000-6100 3.4X10'' -3.4X105 500-1000 Sliding

(cm/sec) Pa

11 27 cm/sec to 2.4N to 800 Sliding

5300 rpm 2.5X10^ Pa 500 Rolling

13 25,000 stress to 2.5X10^ Pa to 700 Rolling

14 to 3300 rpm to 4.7X10^ Pa no control Rolling

16 1500 rpm lOON 500 Rolling

17 24,000 rpm ION no control Rolling

18 to 1X10^ cm/sec to 780N to 500 Rolling

19 10,000 rpm 2.7X10^ Pa 538 Rolling

20 to 41,200 rpm not reported to 370 Rolling

21 10,000 rpm 2.7X10^ Pa 538 Rolling

22 10,000-55,000 rpm 110 and 445N to 540 Rolling

23 2000 rpm 1.9X10^ Pa to 800 Rolling
Table 1 (continued)

Ref

.

Major Conclusions Drawn

8 1

.

2.

Thermal spots occurred in wear contact affecting wear

properties

.

High friction may correlate with high wear in ceramic

materials

.

i i i .

2.

Elimination of bearing vibration is critical to

bearing life.

High surface stresses may cause ceramic surface

structural breakdown.

13 1.

2

.

Failure is due to loss of geometrical integrity.

Wear may occur by plastic deformation or catastrophic

cracking.

1 /. 1

.

The structural strength of ceramic was inadequate for

the bearing design and imposed loads.

16 1 . Z . Lubricants are ineffective at test conditions used.

"Ceramic" and HP-Si3N^ rated poorly under large loads.

17 1. Damage on bearing races is due to TiC particles

spalling off ball.

18 1. Properties that are important for making SijN^ a viable bearing component are low elastic modulus and

low density (compared to steel)

.

19 1. If oxidation film can be formed and ground off, a better Si3N^ ball bearing with "blunted" surface

defects would result.

20 1. Primary cause of failure was due to inner race mounting inadequacy.

21 1.

oz

.

Optical and SEM microscopy detected microf issuring of

Si3N4 surface.

Extreme care must be taken in bearing and race design.

22 1. The following factors contributed to bearing failure:

a) inadequate diametrical clearance,

b) thermal gradients

,

c) poor housing design,

d) solid lubricant removal, and

e) inadequate cage dampening & piloting.

23 1. The ball retainer needs further attention in

strengthening high temperature materials and reducing

friction at ring lands.
9

taken to assure that the bearing fails due to a limitation in the

ceramic bearing material and not due to poor engineering design of the

ceramic components. This requires that special attention be given to

seemingly trivial design parameters, but in the end, more meaningful

data will result.

Second, in the testing of ceramic materials, particular attention

should be given to the surfaces of the ceramic tribological

components, because the surfaces are vital parts of the tribo-system.

Questions like "Did I perform a process in the preparation of the

component that might give rise to unusually high surface stresses?"

and "Has the surface in some way altered itself such that a layer of

material at the surface may have different properties from that of the

bulk?" need to be asked. In the studies noted above, many of the

wearing processes were thought to be initiated from high surface

stresses acting at or just below the tribo- surface . In fact, the

maximum shear stress in the region of a tribo-contact is often below

the surface. Therefore, concern with surface and subsurface defects,

inclusions, machining defects, bulk and surface porosity, and residual

stresses caused by machining or processing techniques is necessary.

2.2.2 Engine Studies

Several articles dealing with the development of ceramic

materials for high temperature, low-heat-rejection engine usage have

been written in recent years. Unfortunately, much of the information
10

relating to heat engine studies is not readily accessible nor easily

understood. This is true for several reasons.

First, in studies dealing with high temperature engine materials,

much of the information that is discovered is proprietary and

therefore not available to the general scientific community. This

produces many localized databases of quantifiable information which is

not accessible to the research community.

Second, it is extremely difficult to identify and measure

tribologically important parameters in a ceramic engine test. For

instance, what are the contact pressures of a ring on a cylinder over

the stroke length? What are the velocities and accelerations along

the stroke length? What are the bulk temperatures in the ceramic

components along the stroke length? What are the surface temperatures

along the stroke length? Is "hot-spot" behavior important in the

system? These questions are not easily addressed and answers are not

readily available. Therefore, interpretation of heat engine test data

needs to be done carefully.

In the Highway Vehicle Systems Contractor's Coordination

Meeting,^'' R. N. Katz of the Army Materials Research Center points out

that the processing of ceramic materials can greatly affect their

strengths. He found that for silicon nitride rotors, the sintering of

the ceramic part had to be carefully compacted and sintered and that a

full-density ceramic component was essential. This meant that 1) hot

isostatic pressing (HIPing) should be done, 2) nitrogen over-pressures

were to be used in the sintering, 3) additives which promote diffusion

were to be used, or 4) the starting powder size distribution had to be
carefully controlled . ^

^ This problem of the careful fabrication and

design of ceramic components needs to be implemented if ceramic

tribological testing is to be effective. In this case, special

processing techniques had to be implemented to insure that a fully

dense ceramic component was produced. Katz also found that thermal

cycling can greatly affect the mechanical [and perhaps the tribomechanical] properties of silicon nitride rotors, and under these

conditions, strengths may decrease by as much as 40%.^^

Recently, two reviews of ceramic tribology in engine development

were written. George M. Thur gives a review of the Department of

Energy's Heat Engine Program. "^^ This review gives special insight

into the wide range of national impacts of a successful low-heatrejection engine program. Second, D. J. Godfrey, in the July 1983

issue of Materials and Design , assesses the use of silicon carbide,

silicon nitride, partially stabilized zirconia, SiAlON, and SiON as

candidate materials in low-heat- rej ection engine design.

Several studies have been conducted recently. Details and major

conclusions of these studies are given in Table 2.^^""^^ From these

studies, three major materials emerge as potential candidates for hea

engine materials. These ceramics are partially stabilized zirconia,

silicon carbide, and silicon nitride. In 1982, Woods and Oda^^ looke

at partially-stabilized zirconia (PSZ) for engine components. They

discovered that PSZ could be useful because it has a low thermal

conductivity (good insulating properties) , a high coefficient of

thermal expansion (which matches to metals better than some ceramics)
Table 2

A Sunimary of Heat Engine Studies

Ref

.

Author Organization Material(s) Used

29 Woods Cummins Engine Co. Partially stabilized

zirconia

30 Hannink CSIRO Materials Mg-partially stabilized

zirconia

31 Storm Carborundum Co

.

Silicon carbide

32 Marmach Nilsen Sintered Products Mg-partially stabilized

zirconia

33 Bryzik U.S. Army TACOM Partially stabilized

zirconia

34 Timoney Univ. College Dublin Silicon carbide

35 Edwards Sir W.G. Armstrong

Whitworth and Co.

PSZ/ceramic

36 Radanovic Cummins Engine Co

.

Silicon nitride, chromium

carbide
Table 2 (continued)

Ref

.

Ma] or Conclusions Drawn

29 1. PSZ is a good material because of its low thermal

conductivity

.

30 1.

2.

Mg-PSZ's show a wide degree of variability in many

applications

.

Mg-PSZ's properties can be tailored to many specific

applications

.

31 1

J. •

2.

T^V»£i X5 1" f" ;i r* ViTn*=»'n 1" (tF t~V»f» r* ^ T" a m i r* l^o ("Vip^ Tnpit~;il n^ivt" i q important

.

Design stress analysis and interaction with manu- facturers is important.

32 1. Mg-PSZ posses unique properties that make it suitable

as wear resistant components for adiabatic engine

d p U X J_C- ct L. X Llo .

33 1

.

The work showed dramatic results including the world's

best fuel economy.

34 1. On high temperature wearing of SiC, friction is low

due to either a graphite film or gas film lubrication,

attachment

.

35 1. A swing beam 2 -Stroke diesel engine may be feasible

for adiabatic diesel engine

applications

.

36 1. A minimum cooled engine could run as long as 1500

hours with minimum oil consumption.
and yet retains good thermal shock properties. In 1983, Timoney and

Flynn explored the use of silicon carbide as "adiabatic" diesel engine

components . ^

^ In that same year, Storm, McBeth, and Flynn found that

silicon carbide components could be operated in an adiabatic diesel

engine.'^ ^ In that study, a silicon carbide rotor was 1) subjected to

800°C and 110,000 rpm, 2) mounted in a test stand for 40 hours, and

then 3) put into a diesel engine tested in road service for over 8000

miles. Radovanovic , Kamo , and Dufrane explored silicon nitride as a

material for high temperature engine components . ^ Also, silicon

nitride already performs well as a high temperature bearing

material^^ • • and therefore, it would be a primary candidate for

high temperature bearings in an adiabatic diesel engine.

To conclude this brief overview of high temperature engine

testing, a missing common denominator in high temperature engine

testing should be examined. No laboratory simulation that effectively

measures friction and wear under simulated engine conditions exists.

Each of the programs referenced ran tests in different engines under

different speeds, loads, environments, and temperatures, thus making

results difficult to compare. Furthermore, such a wide variety of

tests would not easily lend itself to a general understanding of the

basic mechanisms of wear of ceramic tribological components.

Therefore, it is necessary to have a high temperature tribometer that

can test ceramic materials under 1) high bulk temperatures, 2)

controlled atmospheres, 3) controlled loads, and 4) have the

capability inherent in the design to perform mechanistic studies as

well as engine simulations. Such a laboratory simulation would
15

further the understanding of the tribological properties of ceramic

materials as candidates for low-heat-rejection engines.

2.2.3 Tool Studies

Several studies have been conducted in which ceramic materials

have been used as machining and tooling materials . ^ ^ Table 3 gives

a summary of test conditions for these references. Bayoumi , Bailey,

and Stewart"'^ studied cemented carbide ceramics for use in wood

machining. They found that, in the machining of materials under wet

conditions, tool wear occurs by mechano- chemical removal of the binder

material. Enomoto and Fujise^^ studied alumina cutting tools using

cathode- luminescence and observed that the tool tip changed both

composition and structure when machining iron. Young, Becker, and

Rhee*^ in 1983, studied drills coated with TiN and TiC/TiN and found

that wear in their system greatly depended on how the coating was

applied (i.e.. Physical Vapor Deposition versus Chemical Vapor

Deposition, etc.).

All of the above cited references reinforce the issue that wear

of ceramics is not merely a mechanical effect, but may also include

both chemical effects and tribo-chemical effects. Thus, in the design

of a ceramic tribological test series, special attention to the

possibility of complex chemical and tribo-chemical interactions which

take place at surfaces, material-material interfaces, and solidliquid interfaces in a tribological system need to be considered.

Although tribologists know this, models which incorporate chemical,
16

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17

environmental, and mechanical effects are difficult to develop.

However, when we begin to understand these effects more fully, such

modelling will lead to a greater understanding of friction and wear in

ceramic tribological systems.

2.2.4 Grinding, Polishing, and Other Studies

Many studies have been conducted in the study of the friction and

wear of materials which find use in the grinding and polishing

industry . ^ ^ ® Karaki and Watanabe^'^ looked at zirconia as an

abrasive for the mechano- chemical polishing of crystals used in the

electronic industry, and they found that mechano-chemical polishing

can increase material removal rates several fold over chemical etching

or mechanical polishing. Komanduri, Laverty, and Shaw'' ^ studied

ceramic abrasives for a simulated hot grinding technique. In this

technique, either alumina or silicon carbide abrasives were used in a

grinding process at temperatures to 1000°C. Suzuki and Sugita^^

studied polishing of MgO crystals while lubricating with water and

using a vibratory motion. They found that during the test, the pH of

the water rose because Mg(0H)2 was being generated.

Other interesting studies done^^"^'^ include a study done by Uhrig

and Waubke''^ in which an investigation was conducted to explore the

wearing properties of a ceramic flooring material. Wallbridge and

Dowson^° studied an alumina ceramic as a potential replacement for a

prosthetic hip joint. They found that only a marginal increase in

life was gained by using this material over other materials.
18

2.2.5 Ceramic Lubrication

No review of ceramic tribology would be complete without an

inspection of the methods of ceramic lubrication. Traditionally,

lubrication of materials for low temperature purposes has been

achieved by the use of a formulated liquid lubricant package and a

precisely machined bearing and lubrication system. At intermediate

temperatures, more sophisticated synthetic liquid lubricants are used

and sophisticated bearing lubrication systems are employed such as 1)

lubricant atomization, 2) lubricant cooling and filtering, and 3)

lubricant pumping.

However, at high temperatures, the above methods fail because:

1) the oxidation stability of the lubricant used is poor in high

temperature oxidative environments, 2) liquid lubricant vaporization

tends to cause many problems in bearing systems, 3) thermal

degradation of lubricants is rapid at high temperatures, and 4) sludge

formation due to increased oxidation and thermal degradation causes an

increase in system wear and shortens system life.

If, however, a lubricant can be continuously replenished at a

steady state level, a lubricating film can be maintained. Vapor phase

deposition of lubricants is a technique which seeks to replenish a

lubricant film at a steady state level. This technique shows promise

and will be discussed in the following sections. However, important

lubricant-ceramic chemical interactions are unknown for high

temperature ceramic systems. Many of the ceramic materials may or may
not be reactive to conventional lubricants at elevated temperatures.

Because of all of the above problems for ceramic lubrication at high

temperatures, different approaches are being taken to solve this

problem. Current efforts are divided into two major categories: 1)

Solid and "self" lubrication and 2) Other techniques for ceramic

lubrication

.

2.2.5.1 Solid Lubrication

Solid lubrication of metals is an area that has been studied

extensively for high temperature aerospace bearing studies. Recently,

solid lubricants are being studied for use in high temperature

engines. An excellent review of solid lubrication was written in

October 1982 by H. E. Sliney.^'' Sliney separates solid lubricants

into four major classes: a) layer lattice solid lubricants, b)

polyimides, c) unconventional solid lubricants, and d) hard

coatings. Table 4 gives a summary of the lubricants examined.

Many other articles have been written that relate to solid

lubrication and self lubricating compos ites . ^ ^ " ^ For solid

lubrication, many technical challenges still exist. These challenges

are to describe, formulate, and verify the mechanisms for solid

lubrication at high temperatures for 1) a variety of lubricants, 2) a

variety of materials, 3) a variety of environmental conditions, and 4)

a variety of test conditions. Such mechanisms are difficult to

understand. Thus, after 30 years of intensive solid lubrication
Table 4

A List of Materials Examined from H. E. Sliney's

Solid Lubricant Materials for High Temperatures - A Review"

Layer Lattice Solid Lubricants

1 \ i

)

My^ C rlOb2

->) Graphite

(jrapnite riuoriae

Polyimides

1) Polyimide Coatings

roiyimioe - lionueQ vjrapnite r±uoriae uoacings

3) Polyimide Composites

- Unconventional Solid Lubricants

1) Bonded PbO-Si02

2) LiF

3) CaF2

4) BaF2

5) Fluoride Metal Composites plus Coatings

6) Fluoride -Oxide Composite Coatings

Hard Coatings

1) B,C

2) Tie

3) SiC

4)

5) WC

6)

7) TiN

8) Cr2 03
21

research, the fundamental mechanisms of solid lubrication are still

not well understood and characterized at extreme temperatures.

Thus, in high temperature lubrication studies, there is a

need for mechanistic studies. Perhaps these tests could be most

effectively conducted at high bulk temperatures while minimizing the

effects of frictional heating in the tribological contact junction.

2.2.5.2 Other Techniques

Several lubrication techniques other than solid lubrication

are being explored for ceramic lubrication. A method that deserves

note is the use of thin gas films to lubricate tribological contact

junctions. Gas supported bearings have been used for many years.

Moore, Brandon, and Smyly,^* patented the design of an alumina

ceramic, gas lubricated bearing. In 1980, Zacherl patented a gas

bearing design for a ceramic rotor in a gas turbine engine or turbo

machine . ^

^

Several advantages and disadvantages exist for a gas bearing

design. First, a great advantage of lubricating a bearing using a

thin gas film is that, if the bearing operates in an ideal fashion,

tribological surfaces do not contact but a thin gas film maintains

separation of the surfaces. However, problems could exist with gas

bearing designs if 1) a stable and pure constant pressure gas supply

is not maintained for a pressurized hydrostatic gas bearing, or 2) a

certain speed and load regime is not maintained for a hydrodynamic gas

bearing. Thus, designs of this type represent challenges in design
22

but oppor: tun i tJ it these designs can be perfected. For instance,

hydrostatic 'assist' on a hydrodynainic bearing couid improve greatly

the porformnnce of such a bearing.

A second, technique for high temperature lubrication of

ceramic materials is the Chemical Vapor Phase Deposition of

Lubricants. This technique, pioneered by Dr. E. E. Klaus of the

Pennsylvania State University, involves 1) vaporizing a lubricant at

relatively low temperatures in the presence of a carrier gas (which

may or may not contain oxygen) , and 2) contacting this carrier gas

with a hot substrate. When the vaporized lubricant contacts the hot

substrate, the lubricant oxidizes or thermally degrades forming a thin

lubricant coating on the surface of a component. This coating then

provides lubrication for a limited period of time and, if replenished

at a steady state level, can provide continuous steady state

lubr icat ion

.

This process is encouraging for high temperature tribology

because it relies on the thermal degradation of a material. In 1974,

Lai studied such a vapor phase lubrication system for stainless steel

die casting. Min, in 1980, conducted a limited study of vapor

deposited films^'' while Pinto, in November 1982, studied high

temperature vapor phase lubrication for organic liquid lubricants.^®

These studies show promising results for Chemical Vapor Phase

Deposition of Lubricants as a method for high temperature lubrication

of metals in low-heat - re j ec t ion engines. Feasibility of this concept

on ceramic surfaces has not been demonstrated. Research in this area

is continuing.
2.3 Summary/Conclusions

Tribology of ceramic materials is an interdisciplinary science

involving materials science, physics, chemistry, and the engineering

sciences. To fully understand ceramic tribology, engineers, materials

scientists, ceramists, and tribologists need to work together to

understand tribological processes in ceramic wear contacts. A first

step towards that goal was to review the work done in the area of

ceramic tribology, for in reviewing the work several conclusions can

be made

:

1. In many cases, materials parameters are not sufficiently

described or reported. Therefore, effective comparisons between data

sets simply cannot be made. For example, let's look at the aluminum

oxide ceramic system. If a researcher reports that he is working with

an "alumina" powder, question such as, "Is this alumina an alpha, chi

,

delta, eta, gamma, kappa, iota, theta, or xi alumina?"^^ might be

asked. Furthermore, if it is reported that gamma alumina is used, is

the gamma alumina really gamma alumina, eta alumina, delta alumina, or

a mixture of chi and gamma alumina? Which conventions were used when

samples were received? Was the conversion Alcoa's or was it Haber's?

Was the convention British or French?^°°

Apart from general nomenclature challenges, commercial materials

should be specified carefullv. If a ceramic is reported as "company

X's alumina" and a researcher wants to follow up on another's

research, then the questions, "How was the processing done? What was
24

the purity of the alumina? What was done in finishing the components

after fabrication? What batch number was used? Is that material

still available?" can be asked. These questions require answers

before a researcher can duplicate another's experiment.

In addition to the key parameters mentioned above , it is

necessary to report materials parameters when working with ceramic

materials. Charles F. Bersch, in "Properties of Ceramics for

Structural and/or High Temperature Use: Need for Control,

Measurement, and Compilation" , ^ °

^ states that several parameters are

needed to specify the mechanical properties of a ceramic ° ^ These

are given in Table 5. Furthermore, to describe the tribological

properties of the ceramic material, the processes involved in

machining of the ceramic surface are important. For instance, was

grinding used, and was the surface subsequently polished or lapped?

Are the surfaces stressed? Are there surface defects? If the above

properties were reported, a better understanding and comparison of

ceramic tribological tests could be realized.

2. In many cases throughout the tribological testing community,

different loads, speeds, temperatures, and materials are tested. This

variation in load, speed, temperature, and materials is to be

expected. Every application is not the same. Every wear tester is

not the same. However, within a data set, batch to batch variations

in materials may cause problems in internal consistency. Thus, there

is a need for a ceramic standard reference material. If experimenters

were to compare their results relative to a given ceramic material an

internally consistent data set might result. In addition to this, it
Table 5

Ceramic Mechanical Characterization Parameters

Conventional Characterization:

composition (1 percent level)

average grain size

density

Augmented Conventional Characterization:

crystalline and amorphous phases

grain shape

grain size distribution

degree of preferred orientation

distribution of porosity

surface roughness

Fine Micros tructure and Other Characteristics:

trace impurities , amount and state of aggregation

dislocation concentration

twin concentration

point defect concentration

residual stresses

ambient atmosphere

radiation field

state of temper (glass)
26

would be convenient to have an industry wide standard wear reference

material. If this material is characterized and tested as a reference

material with ceramic test materials, a measure of consistency for

external comparison can be attained.

3. Better temperature control and measurement systems are needed

to effectively test ceramic materials. In many cases, contact

temperatures far exceed bulk temperatures in the ceramic material.

This factor alone may be critical to the wear of the ceramic material.

J. E. Mines, Jr.,^°'^ reported that for an alumina based ceramic

material in a "room temperature" test, a phase transformation of the

wear tip to delta alumina occurred. This indicates that surface

temperatures were 1100°C or higher.

Two solutions exist to the above problem. First, it is necessary

to develop adequate temperature modeling. Then, for a given ceramic

material in a given configuration and subjected to certain loads and

speeds, a reasonable prediction of the surface temperatures can be

made. Second, if an apparatus could measure the surface temperature

directly at high bulk temperatures and low frictional heating rates

(i.e., low loads and/or speeds), then surface activities and

reactivities of ceramics might be found. Finally, it would be useful

to measure the temperature of the ceramic contact at low bulk

temperatures taking into account the surface activities of ceramic

materials as well as temperature differences between the contact and

the bulk ceramic (i.e., differential expansion and thermal shock

properties may be important)

.
27

4. Triboiogicai testing of ceramic materials for use in engine

applications has increased rapidly in recent years. Unfortunately,

mechanisms of wear of ceramic materials in engine applications are not

generally understood. Temperatures in an engine test are not easily

measured or controlled and the engine tests themselves are usually

very time consuming and expensive. Therefore, a technology gap exists

in the area of heat engine technology. To bridge this gap, wear test

facilities should be developed that can test ceramic materials at

conditions in which high temperature mechanisms of friction and wear

can be studied. This may require high temperatures and conditions

which minimize frictional heating. As an added capability, it would

be useful if such a facility had the capability for modification to

simulate high temperature engine tests and to control, measure, and

intentionally induce vibrational transients that may cause other modes

of wear in ceramic systems.
Chapter 3

DETERMINATION OF DESIGN PARAMETERS AND SPECIFICATIONS

Ceramic tribological research is being conducted for ceramic

bearing research, tool studies, low-heat-rejection engines, and

several other critical wear processes. In designing a wear tester fo

ceramics, the end use applications should be considered. However, to

design a wear tester that will be able to test all materials at all

conditions of speed, load, temperature, atmosphere, and lubrication i

impossible. In trying to design a tester that can operate under all

loads, speeds, and temperatures, precision and sensitivity would be

sacrificed. This tester might operate under a very wide range of

conditions, but might produce high quality data only in a narrow

range. This is not acceptable in wear tester design. For any

apparatus, the researcher desires that the wear tester would:

o be as frictionally sensitive as possible so that low

frictional levels might be detected,

o maintain contact alignment so that anomalous wear problems

are not encountered,

o minimize machine vibrations so that anomalous friction and

wear can be avoided.

Therefore, the major usage of the apparatus being designed as

well as desirable secondary usages for the equipment should be
determined. Once these uses are determined, tester configuration,

temperatures, loads, speeds, and environments for the primary and

secondary applications can be specified.

The development of the design parameters and specifications for

the High Temperature Wear Facility was approached accordingly. The

purpose of this facility is to gain a basic understanding of the

mechanisms of friction and wear of ceramic materials at high

temperatures . This fundamental purpose spans a number of technical

application areas such as ceramic materials for low-heat-rejection

engines, tools, and advanced bearings. If a basic understanding of

the mechanisms of friction and wear for a number of ceramic systems

could be unraveled, then the design of materials which would minimize

friction or wear in a given application would result. Furthermore,

since many of the materials to be studied may be used as advanced hea

engine materials, it is desirable to have the future capability to

simulate motions which occur in a piston ring-cylinder liner contact

or a cam and tappet contact. This would enable the tester to test

research samples as well as to test leading ceramic candidates for

engine applications. To summarize, the major purpose of the High

Temperature Wear Facility is to effectively test ceramic materials fo

friction and wear in order to gain insight into the fundamental

mechanisms of wear in ceramic-ceramic tribo-systems . A desirable

future capability would be the ability to simulate conditions in whic

an advanced heat engine might operate.
3.1 Wear Test Configuration

The choice of wear test configuration is fundamental to wear

tester design. Many wear test configurations are used throughout the

tribological conununity, and each is designed for a specific

application. Appendix A provides a careful analysis of several major

wear test configurations. For a comprehensive review, ASLE^°^

provides a summary of over 230 wear test devices.

For the High Temperature Wear Facility, a ball-on-disk

configuration was chosen. This configuration is commonly used and is

adaptable to a number of different specimens such as pin-on-disk,

button-on-disk, pin-on-flat, and ball-on-flat without compromising

geometric integrity.

3.2 Temperature Limit

To determine the temperature ranges that are necessary to study

the mechanisms of wear of ceramic materials , the temperatures

encountered in the wear testing of ceramic materials should be

ascertained. Enomoto and Fujise,^°^ in 1983, found in an alumina too

study with cathodo- luminescence that the tool tip temperature was

nearly 1000 °C for a room temperature test. Munro^°^ found that

temperatures in a four ball contact could be higher than 800°C for a

sintered alumina material. Burton^ °® found "red hot" hot spots in a

"room temperature" steel wear test. Winer-'-^^ found that in a 20°C

bulk temperature system, flash temperatures could raise the
31

temperature for 8 up to 1100°C at asperity contacts. Earles,

Hayler, and Powell-'-^° found that surface temperatures could rise to

650°C in a room temperature test. In an advanced heat engine, we find

that temperatures at a thermal barrier pistons may be in excess of

800°C.^^^ If we were to study what might happen in a tribological

contact junction at high temperatures, a tester that has temperature

capabilities of 1450-1500°C is desireable.

3.3 Environment and Lubrication

In order to examine the wear mechanisms of materials for advanced

heat engines, the capability to test these materials under similar

conditions is necessary. Usually, the material is tested in the

presence of oxygen and oxidative atmospheres. In addition, it is

necessary to operate such an apparatus under inert gases such as

argon, helium, or nitrogen to establish an oxidation baseline. For

future studies in which the effect of water vapor on the friction and

wear of ceramic materials is to be examined, the capability to

introduce water vapor into the system is required. To stimmarize, the

atmospheres which may be used are argon, helium, or nitrogen as an

inert, and oxygen, air, or a mixture of oxygen and an inert might be

used as a test atmosphere. It would also be convenient to be able to

introduce water vapor into the wear tester for future studies.

In lubrication, it would be desirable to have the capability to

perform solid and vapor phase lubrication studies at high

temperatures. Material-material and material- lubricant interactions
32

are important and should be studied. In vapor lubrication, the

effectiveness of vapors on various materials combinations under a wide

range of conditions should be investigated to prove concept

feasibility. Current wear testers are not capable of operation under

high temperatures and vapor lubrication conditions.

3.4 Speed Ranges

In the study of surface transformations and reactions which occur

in high speed contacts, it is important to know accurately the surface

temperature at the tribological contact. This may be approached in

two ways

.

The first approach is to run the system at any speed and to

accurately measure the surface temperature in the tribological

contact. In practice this is a difficult task to perform. If this

approach is taken the questions "How will I measure temperature in the

wear contact?", "What will I use?", and "How can I be sure that I am

really measuring the temperatures in the contact?" arise.

The second approach in accurately determining the contact

junction temperature is to minimize the effects of frictional heating.

This requires very accurate, low speed control so that no appreciable

frictional heat will be generated. In this way, the assurance that

the contact temperature is very close to the bulk temperature of the

ceramic material is possible. This approach will be taken with the

High Temperature Wear Facility. With a 1750 rpm motor which has 100
33

rpm as a minimum speed, if 100:1 reduction in speed is used, then a

speed range from 1 rpm to 17.5 rpm is accessible.

Conversely, if the apparatus is to be used for the secondary

purpose of engine simulation, then much higher linear speeds are

desired. For instance, operating with a ball on a two inch disk

apparatus contacting at a radius of one inch, 1000 rpm would give a

linear speed of 8.7 ft/sec. This is closer to cam and tappet speeds

in an engine than 0.008 ft/sec (1 rpm) would be.

Therefore, the speed range is chosen to be 0 to 50 rpm with the

capability of in the future of higher linear speeds.

3 . 5 Load Range

For mechanistic studies, the load range needs to be versatile.

For chemical reaction/surface reaction studies at high temperatures

,

low loads are desirable. In this case, a 0-20 kg load range would be

more than sufficient. For simulatory motions, the average Hertzian

contact pressures need to be compared with piston ring/cylinder linear

pressures and cam/tappet pressures. In a cam and tappet contact, the

peak pressure is about 8 x 10® Pa.^^^ In a piston ring/cylinder

liner, a maximiam pressure of 3.44 x 10^ Pa for an engine operating at

136 bmep, 2200 rpm, and 2000-4000 psi cylinder pressure^ is

generated. If we take the very simple formula for Hertzian contact

radius for a ball on a flat disk:

2E
34

where L = load (in N)

R = radius of ball (in m)

u = Poisson's ratio for material under study

E = modulus of elasticity (in N/m^ ) (Pa)

a = Hertzian contact radius (in m)

For alumina ceramic material, the modulus of elasticity is 393 x 10^

Pa, and the Poisson's ratio is 0.21. Therefore for a ball radius of

6.35 mm (1/2" dia.) we have a Hertzian contact radius of (for a 10 kg

load) a = 9.113 x 10"^ m. The mean Hertzian stress would be 98N/7r(a^)

= 3.7562 X 10^ Pa. This is above the needed contact pressure by an

order of magnitude. However, at high temperatures, Hertzian contact

stresses may be less than those at low temperatures and larger loads

may be necessary.

3.6 Summary of Specifications

Table 6 presents a summary of the design parameters and

specifications for the High Temperature Wear Facility. Both needed

and desired load ranges are given in the compilation.

In addition to the above, future modification for reciprocating

motions should be possible for engine simulations.
Table 6

High Temperature Wear Facility Design Specifications

Design Target: Versatility

Configuration: Ball-on-disk

Modification for: Pin-on-disk

Bail-on- flat

Pin-on-flat

Temperature

:

Ambient to 1500°C

Control capabilities for room

temperature to 1500°C.

Load: 0-40 kg (0-394. 4N)

Capability for modification to

higher loads

.

Atmosphere

:

Oxygen

Argon, Nitrogen, Helium

Mixture of Argon, Oxygen, Nitrogen,

or Helium

Addition of Water Vapor

Lubrication: Solid Lubrication

Vapor Deposition of Lubricants

Motion: Unidirectional

Capability for modification to

reciprocating motion.

Gas Flow Rates: 0-0.15 2/min
36

3.7 Introduction to Design

A unique set of design and performance criteria exists for the

High Temperature Wear Facility. Special care should be taken in

materials selection because of the high temperature and loading

requirements of the facility. For many materials, the high

temperature strength capabilities are exceeded. These materials

cannot be used. Oxidation and environmental control at high

temperatures presents many unique design challenges. Sensitivity,

precision, and accuracy of frictional measurements coupled with high

temperature machine and contact alignment, vibration control, and

provisions for lubrication further complicate already complex design

issues

.

Novel approaches to design are necessary because of the unique

design specifications for the facility and the design challenges which

accompany them. Key technical challenges occur in heat transfer of

the system, materials selection for the facility, fluid dynamics, and

mechanical design of components. Chapter 4 describes the challenges

which occur with the heat transfer of the High Temperature Wear

Facility Furnace. Chapter 5 examines materials selection for the

facility, while Chapter 6 deals with fluid dynamics and the design of

a frictionally sensitive, high temperature alignment system. Finally,

Chapter 7 addresses several challenges which occur in the mechanical

design of components to account for differential thermal expansion,

design of a drive mechanism, and selection of materials for frame

design.
37

Chapter 4

HEAT TRANSFER

The design of a system to accommodate 1500°C temperatures

presents several complex technical challenges. Conductive, convective

and radiative heat transfer should be accounted for because of the

extreme temperature testing. The basic heat transfer equations for

such a process are quite complex and are given below as

^

pU = - (V • pUv) - (V • q) - p(V • v)

at

- (r : Vv) - (E - A) [2]

where t = time

p = density

U = internal energy per unit mass

v = mass average velocity

q = energy flux relative to mass average velocity

r = viscous stress tensor

E = local rate of photon emission per unit volume

A = local rate of photon absorption per unit volume

and
38

^ U< ' = -(V • ' ) + (E - A) [3]

at

where u' ^

' = charge of radiant energy density

q( r ) _ radiant energy flux

The operator

V • V is defined as the divergence of a vector field

ax, ax,

where v is a vector function of Xj^ , X2 , and X3 and x^^ , X2 , and X3 are

the spatial variables for a given coordinate system.

Similarly

- : Vv = L r,^ i_ V, [5]

where r,j are the components of the stress tensor field and Xj 's are

the space variables for a given coordinate system.

Such equations are difficult (at best) to solve. Therefore, a

number of simplified heat transfer models should be systematically

examined

.
4.1 Heat Transfer Calculations

There are several different methods of heating a sample to very

high temperatures. One method would be to locate a "space heater" in

near a ceramic test sample and to radiate energy directly to the

surfaces of the specimens. Initially, this type of heating was

considered. Figure 1 shows the general schematic of such a heating

system for a ball-on-disk configuration.

In this system, radiative heat transfer to heat the surfaces of

the wear test specimens is important. Therefore, to determine the

energy requirement necessary for such a system, the radiative heat

transfer equation is used:

Q(Energy/Time) = t7(€iAiFi2T/ - ^2^2 ^2 1 "^2 ^ ) [6]

where a = Boitzmann's constant

= Emissivity of specimen holder

= Emissivity of chamber

= Respective surface areas

F12 = View factor from holder

F.2 1 = View factor from chamber

= Temperature of holder

= Temperature of chamber
40

Upper Specimen

Ring

Lower Specimen

Figure 1. General Schematic of Space Heater System
41 li

Furthermore, assume total enclosure (^12 ~

'^zi ~ ) > goo^L radiators

(^1- ^2 ~ 0-8), specimen temperature of 1500°C, and surrounding

temperature of 20°C. If the area of a specimen is A^^ = 155 cm^ (24

in^ )

, the heat needed to maintain the surface is approximately

Q = 6000 W/specimen

= 23,679 BTU/hr/specimen [7]

Assuming that the space heater used would be a common molybdenum

disilicide space heater which has a heat flux density at 1500°C of

11.6 W/cm^ (75 W/in^), a surface area of 516 cm^ (80 in^ ) is needed to

provide adequate heating. This is virtually impossible in a wear

tester, because samples need to be in close proximity.

Furthermore, to complicate this problem, large thermal stresses

would be caused in this system because of very large thermal gradients

in the low conductivity ceramic materials. If this type of system were

used, thermal shocking of both specimens and sample holders might

occur

.

Therefore, an alternate approach to heating the High Temperature

Wear Facility is considered. A logical choice would be to heat the

whole specimen holder and chamber to 1500°C and to maintain the

temperature at this limit. The major technical challenge in this type

of design is the thermal insulation of the heated chamber. To obtain a

rough idea of what would be necessary for thermal insulation of such a

system, the following assumptions are made:
42

o cubic heating chamber with dimensions 15.2 cm x 15.2 cm x

15.2 cm (6" X 6" x 6")

o 6 heaters, 161 cm (25 in^) each

o heaters can supply 11.6 W/cm2 (75 W/in^

)

Figure 2 gives a general schematic of radiation heat shielding

necessary for this design.

Assuming that the heat stages are close together and the area of

each stage is approximately constant, the radiative heat transfer

equation reduces to:

"heater = ""^^^i' " T^*>

If the radiation shields are blackbodies themselves, e = 1, and

T2 = VT/ - A/aA [9]

If:

wall

= 1500°C [10]

A = 216 in^ (4 walls)

Q

= 75 W/in2 X 100 in^ (4 walls)

the following temperatures for each radiation stage result:
43

20C

Heated

Chamber

Tl = 1500C

Radiation.

Shield

Airspace

Figure 2. HeaC Staging Scheme
44

T (chamber wall) 1500°C

Stage 1 1455°C

Stage 2 1407°C

Stage 3 1356°C

Stage 4 1297°C

Stage 5 1232 C

Stage 6 1157°C

Stage 7 1068°C

Stage 8 957°C

Stage 9 803°C

Stage 10 531°C

If 500°C is the maximum allowed attachment point temperature for a

metal to ceramic interface, a minimum of 10 heat stages is needed.

This is too many stages for conventional staging.

Another approach to insulation was considered. If a porous

ceramic material were used for insulation, a greater heat insulation

would result due to multiple radiation shielding by the many pores in

the material. One side of the pore can radiate to the other thus

making each pore a radiation shield. This is illustrated in Figure 3.

If the pores are small in such a material, then many radiation

stages may be supplied per unit thickness of material. Such a material

will maximize structural strength and insulation capacity.

A configuration for the calculation needs to be specified. Top

and side views of the original high temperature wear facility furnace
Co:D:(juc:f It^d

Figure 3. Porous Insulators as Radiation Shields
46

design are given in Figures 4 and 5. If we assume uniform heat flux in

the angular (6) direction, the steady state energy transport becomes:

1 d(^^r) + 3(lz) = 0 [11]

r dr dz

where = - keff dT/dz

= - keff dT/dr

k j.^ = effective combined thermal conductivity taking into account

eff

conduction, convection, and radiation throughout the insulation

material under examination.

^ff = '^aff<T'-

and

T = T(r,z) [13

To summarize, we have a 2-dimensional , cylindrical, steady state

heat transfer equation with a temperature dependent thermal

conductivity (keff)

.

Boundary conditions may now be specified for the system. Since we

have specified that no 6 variations in heat flux, the smallest

symmetric element can be extracted from the furnace. This is shown in

Figure 6

.
47

Figure 4. Top view of High Temperature Furnace Design
Figure 5. Side View (3-D) of Insulation
J

r

Figure 6. Smallest Symmetric Element for High Temperature

Furnace Design
50

If the width of a square is one unit (arbitrary units) , the

boundary conditions become:

(ar = 4, 0<z<3, T = 1500°C

(a 4 < r < 10, z = 0, dT/dz = 0 (symmetry) [14]

(a0<r<4, z = 3, T = 1500°C

@r=0, 3<z<9, ai/ar = 0 (symmetry)

Now, specify the boundary conditions on the outer surface of the

furnace. Wo assume that the room temperature is 20 °C (68°F) which is

typical of a laboratory. Since free convection dominates heat transfer

to the laboratory at the outer surface of the furnace, we will assume a

heat transfer coefficient of 20 kcal/m^ hr °C (~4 BTU/hr ft^ °F)^^^

from the outer wall to the 68 °F room. Therefore the outer wall

boundary conditions become:

(a r = 10 0 < z < 9 q = h(T , , - T ) wall room

15

(a0<r<10z = 9q,^

= h(T , , - T ) wall room

where T = 20°C (68°F) room

h = 20 kcal/m^ hr °C (~4 BTU/hr ft^ °F)

T n = T , . (r ,z) . wall wall

Figure 7 summarizes the boundary conditions for the given

configuration. To simplify the problem, assume that the thermal

conductivity (k^^^) of the insulation is constant at its highest value
51

I -.H O

O3^

M

O

n

3

O

q = h(T , -T ) wall room

h r 20 kcal/m hr C

T= 1500 C

= 0

il-zo

a z

Figure 7. Sununary of Boundary Conditions for High Temperature

Wear Facility Furnace Design
52

of 0.4 BTU/ft hr °F. Also, assume that one arbitrary unit of distance

is equal to 2.54 cm (1 inch). This gives a minimum of 15 cm of

insulation on all sides of the high temperature chamber. Now a finite

element analysis can be conducted. Such an analysis is still quite

complex and requires computer methods. A pre -developed program for the

given configuration was supplied by the Statistical Analysis Division.

For the given configuration under specified boundary conditions, outer

wall temperatures are calculated at r=0 , z=0 , and at maximum r and z.

This provides centerline and extremity temperatures. For 15 cm of

insulation, it was found that at:

r=0, z=9 (arbitrary units), T = 100°C

r=10,z=0 (arbitrary units), T = 70°C [16

r=10,z=9 (arbitrary units), T = 20°C

These results are shown in Figure 8. If interested in temperature

distributions, Appendix B provides a finite difference analysis for the

given configuration. This analysis can be used to develop a computer

program that will provide temperature distributions. For the purpose

of feasibility analysis, centerline temperatures (maximum temperatures)

and extremity temperatures (minimum temperatures) were sufficient.

Based on the above analysis, the use of a high temperature porous

zirconia insulation such as that used by the space shuttle is viable

for furnace design. The Hign Temperature Wear Facility Furnace was

designed in this manner.
53

20C lOOC

20C

70C

1500C

1500C

J

Figure 8. Results of Finite Element Analysis

(Courtesy of Statistical Analysis Division)
Chapter 5

MATERIALS SELECTION FOR THE HIGH TEMPERATURE WEAR FACILITY

5.1 Introduction

Materials selection for wear test equipment is a complex and

difficult process. In selecting a material for a given application,

several factors are important such as:

o strength of the material

o chemical reactivity of the material the simulated environments

o fabrication technology

o fabrication cost of the material and the final machined

component

For the design of a high temperature wear tester, design is much

more complex. Components made from the materials need to maintain

mechanical stability from room temperature to 1500°C. Furthermore,

additional considerations are required such as:

o thermal expansion of materials,

o differential thermal expansion of mating materials,

o oxidative and thermal stability of materials,

o dimensional stability (shape retention) of materials,

o surface reactions and phase transformations of materials
55

There are several potential candidates from these considerations.

They are

:

o Partially Stabilized Zirconia (PSZ)

o Silicon Nitride (SijN^)

o Silicon Carbide (SiC)

A typical properties table for these materials is given in Table 7.

5.2 Partially Stabilized Zirconia

PSZ is a ceramic which is based on zirconium oxide stabilized with

magnesia or yttria. The crystalline matrix consists of cubic zirconia

with residual tetragonal and monoclinic structures at grain boundaries

and within the cubic grains. When this material is stressed, the

tetragonal phase transforms to the monoclinic phase causing a 4% volume

increase. This volume increase causes a compressive stress to be

formed at crack tips. Thus Partially Stabilized Zirconia has a

fracture toughness of 8 MPa Jm at room temperatures.

Another property which makes PSZ attractive as a structural

material is PSZ's coefficient of thermal expansion. The coefficient of

thermal expansion of PSZ is 10.1 x 10" ^/°C while that of silicon

nitride is 2.3 x 10" ^/°C. For a ceramic, zirconia has a high

coefficient of thermal expansion, and thus can mate to steels much
56

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4-) C4-1 0 in C ed CM^—' CO —<

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57

better than other ceramics. The coefficient of thermal expansion of a

typical carbon steel is ~12 x 10"V°C.^^^

However, because of the materials transformations in PSZ, several i,

problems are encountered. First, at 1240°C, all residual monoclinic

zirconia will transform to tetragonal zirconia. Accompanying this jij

transformation is a sudden 4% decrease in the volume of the

transforming zirconia. Thus, if components are overheated during

testing, phase transformations occur which change the microstructure

and density of the PSZ permanently. Figure 9 shows a typical expansion

curve of PSZ from 0 - 1500°C. Upon heating, expansion follows the

solid line. Since the microstructure has been changed, however, the

thermal contraction of the material follows the dashed line.

To compound the problem, upon cool down between 500 and 600°C,

metastable tetragonal zirconia reverts to the monoclinic phase. This

causes an ~0.4% expansion and possible failure of the bulk component.

Thus, PSZ was ruled unacceptable as a 1500°C structural material.

1

1

5.3 Silicon Nitride

Silicon nitride is widely used as a high temperature bearing

material. Table 7 shows that silicon nitride can have a hardness of 91

R45N compared to 83 R45N of alumina. Furthermore, silicon nitride also

shows no loss of flexural strength up to 1000°C. Thus, this material

is ideal for a structural material because of its high strength and

good shape retention properties.
58

Figure 9. Linear Thermai Expansion Curve of PSZ from 0-1500°C

(Source: Nilsen, Inc. Literature)
59

However, for 1500°C operation, silicon nitride is also

unacceptable. Above 1000°C, silicon nitride rapidly loses strength.

Also, oxidation of silicon nitride causes loss of nitrogen and silicon

segregation in the ceramic.

5.4 Silicon Carbide

Several different silicon carbide materials exist. These can be

divided into two categories: 1) cubic silicon carbide, and 2) hexagonal

(a) silicon carbide. Table 7 shows a comparison between these

materials. The hexagonal silicon carbide (Sohio Engineered Materials)

has a higher hardness (27.44 GPa) than any other ceramic shown in Table

7 and has approximately double the hardness value of alumina. The

flexural strength of this material is high (550 MPa) and this material

has the highest elastic modulus (410 GPa) and shear modulus (178 GPa)

of the materials shown. The tensile strength of a-SiC relative to

other ceramics is shown in Figure 10. At 1500°C, oi-SiC is the

strongest material selected and has strengths higher than silicon

nitride, partially stabilized zirconia, or alumina. Also, the tensile

strength of a-SiC increases with temperature to ~1200°C. Furthermore,

a-SiC unlike other silicon carbides has no free silicon and is

extremely oxidatively and chemically stable to temperatures over

1650°C. To summarize, a-SiC has:

o the highest hardness of the ceramics selected

o excellent oxidation stability
60

0 400 800 1200 1600

Temperature, C

Figure 10. Tensile Strength of Materials to 1500°C
61

o highest elastic modulus of materials selected

o best 1500°C tensile strength of materials selected

A typical properties compilation for this material is given in Table 8.

This material was therefore chosen for the 1500°C structural components

of the High Temperature Wear Facility.
62

Table 8

A Typical Properties Table of a-SiC

Dissociation/Melting Temperature

Property Units RT 200°C 400°C 600°C 1500°C

Thermal diffusivity cm^ /sec 0.413 0.230 0.185 0.140

(laser flash)

Specific heat cal/g °C 0.160 0.220 0.252 0.268 0.334

(drop calorimeter)

Thermal conductivity cal/cm^

sec °C 0.208 0. 160 0. 147 0. 118

Thermal-Shock Resistance Parameter

Property Units

Thermal -expansion coefficient 10"6/°C RT - 700°C ~ = 4 02

700°C - 2000°C - = 5 32

Chemical Stability

Material

HCl H2SO,

Chemical Agent

HNO3 H2 S HF NaOH

HexoloyT" SA Excellent

99% AI2O3 Fair Good Good Good Fair Fair

Tungsten Carbide Poor Fair Very Fair Fair Fair

-6% Cobalt Poor

Stainless Steel Very Very Poor Poor Poor Good

Poor Poor

Source - Sohio Engineering Materials Company.
63

Table 8 (continued)

Properties

TTni U iij. t-c U-o RTIX JL 1 ()00°C ^?oo°c i4no°n i5on°r i6so°r;

(Knoop) kg/mm^ 2800

Wet abrasion 3.4

(Riley-Stoker)

Density g/cm^ 3.1

Young's modulus Mpsi 58.9 54.9

(GPa) (406) (378)

Shear modulus Mpsi 25.8 24.5

(GPa) (178) (169)

Poisson ratio 142 .118

Flexural strength

(4 pt.)

Air kpsi 66.6 64.1 65.3 62.7 58.6

(MPa) (459) (442) (450) (432) (404)

Argon kpsi 64.8 71.6

(MPa) (446) (494)

Weilbull modulus

(2 parameter) - 12.3

Fracture toughness

(double torsion & kpsi inh 4.2 5.8

SENB) (MPa mh) (4.6) (6.4)

Oxidation not detectable

Source - Carborundum Resistant Materials Company.
64

Chapter 6

FLUID DYNAMICS

In wear tester design, two important design considerations are the

precise alignment of test specimens and the control of machine/contact

vibrations under the full speed range. Test specimen alignment is

important in wear rate measurements and uniform contact pressure

throughout the test. Misalignment during the wear test could cause

premature surface damage and change the wear mode. Friction and wear

results are dependent on the machine vibration frequency and amplitude

to a certain extent. Under excessive vibrations at a particular

sliding speed, wear mode could change due to fretting wear.

Therefore, in the design of a wear test apparatus, the designer

seeks to 1) maintain test alignment, 2) minimize vibration, 3) be

sensitive for friction measurements, and 4) maintain the structural

integrity of shafts and holders.

Traditionally, for low temperature wear test devices, the above

four conditions are satisfied by conventional methods. A precision

bearing with a lubrication system is used to control the stiffness of

the shafts. The whole system mating is controlled by high precision

machining on the rotating shafts, on the bearing packs which mate to

the shafts, and on specimen holders and other structural components.

Self -aligning contact geometry is often designed to maintain alignment

throughout the test.
65

For high temperature wear tester design, traditional methods of

maintaining test alignment and controlling vibration are no longer

effective. Conventional lubricants fail to lubricate at high

temperatures. The lubricants may thermally degrade or oxidize causing

catastrophic failure of the bearing system. Apart from lubricant

reactions, high temperature surface reactions may cause components to

fuse together. Even if parts did not fuse together, standard bearing

materials become much too soft at high temperatures causing deformation

of bearings, bearing races, shafts, and specimen holders. This will

cause loss of test alignment and will introduce vibration into the

system.

To overcome these difficulties, externally pressurized gas film

bearings were chosen because they provide:

o no contact between stationary and moving surfaces,

o good stiffness,

o good alignment capabilities,

o maximum sensitivity for friction measurements

Externally pressurized bearings of this type represent

3-dimensional compressible fluid flow problems. The basic equations

for this type of system are, in rectangular coordinates:

Equation of continuity:

dt ox 5y 3z

[17]
66

Equation of motion:

X- component

^ + v„ ^ + v„ ^ + V, : at ax ay az

3P _ _^ ary^ ^ ar^

^^^^ ^^3^

9x ax ay az

y-component

av„ av_, av„ av^

at ax ay az

^ - ^ + ^ + ^ + pg^ [19]

ay ax ay az

z- component

av_ av_ av. av_

at ax ay az

az ax ay az
67

where p = density of gas

V = velocity

x,y,z = spatial coordinates for the system

= component of gravitational force in the z direction

T = stress tensor

Such calculations are quite complex in that we have 4,

3-dimensional , coupled, partial differential equations to solve in

conjunction with the appropriate boundary conditions.

Furthermore, several other design challenges in a bearing of this

sort need to be addressed.

The first challenge in any design of this type is to design a

bearing that maintains side-to-side alignment while still allowing

freedom of movement in the other directions. Initially, the general

layout given in Figure 11 of such a self -aligning hydrostatic bearing

was considered.

Several advantages of the layout given in Figure 11 can be seen.

First, in a "U" shaped bearing design of this sort, both horizontal and

vertical stabilizations occur independently. This means that only side

to side vibrations will be compensated in the horizontal direction, and

only load variations will be compensated for in the vertical direction.

This is an ideal situation, but unfortunately, this preliminary design

had to be rejected. Because hydrostatic bearings of this type have to

be machined to very tight tolerances to maintain stiffness, it was

found that bearing designs of this type would be very unforgiving to

temperature variations between the gas supply surface and the specimen
68

Specimen Holder

Qas Supply Surface

Figure 11. Original Preliminary Hydrostatic Gas Bearing Layout

(Side View)
69

holder. A slight temperature rise in the specimen holder without a

rise in the gas supply surface would cause total bearing failure,

because of the steep angled walls (90°). The sides of the bearing

would meet due to differential thermal expansion if it occurred.

Therefore, to increase the tolerance for differential thermal

expansion, a "V" shaped bearing was next considered. This is shown in

Figure 12.

This "V" type design has the advantage of allowing for small

differences in thermal expansion that may occur between the specimen

holder and the gas supply surface. Unlike the "U" shaped bearing,

these small temperature differences will not cause bearing failure.

Unfortunately, the "V" type layout also has some strong shortcomings.

Compensation for forces in the vertical direction and the horizontal

direction are not independent. If a load on the bearing surface of

is applied as shown in Figure 12, an equal and opposite force of F^ in

the vertical direction is required to keep the bearing afloat.

However, gas bearings always exert forces which are normal to their

surfaces, so the gas bearing exerts a normal force of N. Resolving N

into its component horizontal and vertical forces, for the bearing to

stay afloat:

|Fy| = |n| cos(^) [21]

However, a side force is also generated,

[22]
70

Fif'/iro 12. " V" - .shrapcd Hydro.': r. ic Gar: Bearing Layout
71

Since these side forces are needed to maintain stiffness against

vibration, stiffness will depend upon the load applied and on the tilt

of the bearing supply surface. Table 9 gives Y^/Y^ ratios for a number

of given angles

.

If an angle of 45° were used in design and a load of ION were

applied in the y direction, a reaction of ION in the x direction would

also occur. Thus, in not having a horizontal surface to independently

compensate for load, side to side instability might occur for the

bearing. The "V" shape bearing would not be acceptable because

independent compensation in the load from a horizontal surface is not

present

.

Finally, after careful design considerations, the advantages of

both the "U" and "V" shaped bearings were combined into a new design.

Figure 13 gives top and side views of the hydrostatic bearing layout.

The next task is to design gas supply hole shape. There are two

main types of hole shapes: inherently compensated and orifice

compensated. An orifice compensated bearing and an inherently

compensated supply hole are shown in Figure 14.

The difference between these two types of designs is that in an

inherently compensated supply, the major restriction to gas flow occurs

at the bearing surface, while in an orifice compensated supply, the

major restriction to gas flow is away from the bearing surface. This

difference allows the orifice compensated bearing to be as much as 20%

more stiff, but the orifice compensated bearing may, depending on

transient gas pressures, experience a dynamic instability called
Table 9

/F^ Ratios for "V" Shaped Bearing Design

$ (degrees)

0 CO

5 11.43

10 5.67

15 3.73

20 2.74

25 2.14

30 1.73

35 1.43

40 1 . 19

45 1
73

Figure 13(b). Top View of Hydrostatic Bearing
Inherently Orifice

Compensated Compensated

Supply

Figure 14. Inherent and Orifice Compensated Supply Holes

(Sectional View)
75

pneumatic hammer (i.e. a series of high frequency surface contacts

causing bearing damage). Theoretically, inherently compensated gas

bearings should not hammer. Therefore, an inherently compensated

bearing is chosen.

For the hydrostatic bearing a three-dimensional, compressible,

unsteady state fluid flow calculation was performed. Appendix C gives

the full design calculation for the hydrostatic bearing. For this

hydrostatic bearing, the Mechanical Technology Incorporated Gas Bearing

Design Manual was used as a design guide.

To summarize the design given, it was originally decided that at

least 8 holes of 0.050" diameter should be used with argon gas. In the

calculation, worst case conditions were used. For a gas which

increases its viscosity with temperature, room temperature argon

provides the lowest load carrying capabilities. Therefore, the room

temperature properties of argon were used in the design calculation.

Since the hydrostatic bearing was to be made of an expensive

material and, since the approximation of a standard configuration

allows the design of a bearing for a non-standard configuration, it was

decided that a brass bearing prototype would be built to test the

bearing design.

Fortunately, this approach was taken in the design of this

critical component, for several interesting results emerged. The

hydrostatic gas bearing (which was inherently compensated)

pneumatically hammered. Theoretically, this should not have happened.

After contacting the author of the gas bearing design manual for thrust

gas bearings, the following was found: 1) it was assumed that for an
76

inherently compensated bearing that the bearing would not hammer and 2)

for strange configurations, a small region of instability could develop

if the wrong hole size was chosen.

Therefore, the hole size was decreased from 0.050" to 0.021". It

was found that the smaller hole hydrostatic bearing did not experience

pneumatic hammer. Therefore, in specifying 0.050" dia. holes (which

normally would be better than 0.021" dia. holes) we were merely

unfortunate in hole size selection. However, the machining of 0.021"

dia. holes of the length necessary would not be possible for current

state-of-the-art ceramic machining. The hole size had to be changed to

0.040". This was the smallest hole possible for ultrasonic machining

of the component. Another prototype bearing was made and finally an

acceptable bearing design was found. The 0.040" dia. holes did not

cause pneumatic hammer, so the bearing could be designed.
77

Chapter 7

MECHANICAL DESIGN OF COMPONENTS

There are several technical challenges in high temperature wear

tester design. They are: 1) design of ceramic/metal components to

accommodate differential thermal expansion of pieces, 2) design and

selection of drive mechanism, and 3) design of machine framework for a

high temperature facility. Each of these areas will be briefly

addressed in this section.

7.1 Design of Ceramic/Metal Components to Accommodate Differential

Therma1 Expans io

n

Thermal design of ceramic-ceramic and ceramic-metal interfaces is

critical in wear tester design. Both interfaces have unique

properties, and both require special design.

In the design of the High Temperature Wear Facility, the following

was required of ceramic -ceramic and ceramic-metal interfaces at high

temperatures

:

o interfaces cannot slip, move, or readjust themselves at high

temperatures

.

o ceramic -ceramic interfaces should remain permanent after

attachment

.
78

To meet these requirements, two approaches were considered.

Ceramic-ceramic and ceramic-metal interfaces could be attached through

the use of a high temperature refractory cement. However, refractory

cements lose bonding strength at high temperatures. For ceramicceramic interfaces which would be in the 1500°C furnace, pieces could

slip, move, and possibly mis-align themselves. Also, small particles

are usually present in high temperature refractory cements. If these

particles become entrapped between a tight fitting ceramic-ceramic or

ceramic-metal interface, high point stresses could result. This may

cause the cracking of a ceramic component.

A second way to assure that interfaces do not move is to design

ceramic -ceramic and ceramic -metal interference fits. Traditionally,

with metals, interference fitting and compression fitting are

considered to be synonymous terms. The inner component is forced, by

compression, to fit inside the outer piece, and the outer piece is

deformed elastically to make a permanent fit.

To design an interference fit, stress calculations for ceramic

materials are conducted. There are two types of stresses that may be

considered in the design of a ceramic tube to metal and/or ceramic

interface. They are tensile hoop stress on a tube (Figure 15a) and

compressive hoop stress on the tube (Figure 15b)

.

The basic design criteria for such interference fits are that if

the tube is put in tension, then the hoop stress cannot exceed the

tensile strength of the tube material, or if the ceramic tube is put in

compression, the hoop stress due to the interference fit cannot exceed
79

Figure 15(a). Tensile Hoop Stress on a Ceramic Tube Interference Fit

(Top View, Sectional)
80

Out«r Pl«c«

Compr««slv« Str«s« Csramlo Tub*

Figure 15(b). Compressive Hoop Stress on a Ceramic Tube Interference

Fit (Side View, Sectional)
the compressive strength of the tube materials. For ceramic materials,

the compressive strength is usually 1-2 orders of magnitude higher than

the tensile strength.

For Carborundums' Hexoloy a-SiC:

Tensile Strength: 350 MPa

Compressive Strength: 3900 MPa

Hoop Stress Calculations:

The general configuration of the hoop stress calculation is shown

in Figure 15(a)

:

23

where = hoop stress

q = unit loading per length of tube

r = radius of tube assuming thin wall tubes

A = area of contact of interference fit.

The hoop strain is related to the hoop stress and design parameters as

follows

:
82

where e = hoop strain

Ad = maximum diametrical interference

d = nominal diameter of tube

a - hoop stress

E = modulus of elasticity of material

For compressive stress on the tube:

G = 3900 X 10^ Pa (compressive strength of a-SiC) [25] max

E = 410 x 10^ Pa (modulus of elasticity) [26]

Ad a 3900 x 10^ Pa

e = max = max = [2/]

""^^ d E 410 X 10^ Pa

^^max = 9.512195 x lO'^ [28]

If the nominal outside diameter of the tube is 100.95 mm (3.9746"),

then

Ad = 9.6 X 10 "1 mm (3.78 x lO^^ in) [29] max '

^

However, if the contact is not perfect and only 1/10 of the area

contacts, then
83

Ad = 9.60 X 10 "2 (3.78 x 10"^ in) [30] max

Finally as an added measure of safety, an engineering safety factor is

added to insure that parts do not crack. Therefore for a tube in

compression Ad = 7.06 x 10"^ mm (2.78 x 10"^ in), max

Similarly, for tensile stress on the tube:

a = 350 X 10^ Pa [31] max

€ = 410 X 10^ Pa [32]

Ad a 350 X 10^

e - max — max = [33]

'"^'^ d E 410 X 109

Ad = 8.10 X 10'2 mm (3.19 x 10'^ in.) if all area contacts. [34] max ^

^

If 1/10 of the area contacts, then Ad = 8 . 05 x 10' ^ mm (3 . 17 x 10"^ max

in).

Another technical challenge arose in assembling the ceramic

components. With ceramic materials, the concept of compression fitting

needs to be modified because they are much more brittle than metals and

would fracture rather than elastically or plastically deform.

Therefore, the concept of shrink fitting rather than the concept of

compression fitting is used. In shrink fitting, the outside component

is heated until there is clearance between components. The components

are then assembled. Once the outer component cools, clearance will no
84

longer exist. Precise alignment of components is required so that

misalignment and fracture of components does not occur.

7.2 Design and Selection of the High Temperature Wear

Facility Drive Mechanism

In the design of a drive mechanism for the High Temperature Wear

Facility, two important factors are to minimize machine vibration and

to maintain overall machine alignment. The drive motor and alignment

system should be chosen so that vibrations are minimized and alignment

is maintained.

7.2.1 Selection of a Drive Motor

Several types of motor drives are commonly available: gear drive,

direct motor drive, toothed belt drive, v-belt drive, 0-ring drive, and

several other drive motor types. Geared drives transmit vibration

because of the engagement and disengagement of gear teeth. Toothedbelt drives produce vibrations from the engagement and disengagement of

belt teeth. Direct drive assemblies transmit motor and line vibrations

throughout the wear test assembly. Therefore, the drive that produces

the least amount of vibration, the 0-ring drive was chosen. Drives of

this type are used in precision grinding processes in which vibrational

transients are minimized.
85

7.2.2 Selection of an Alignment System

Again in this area, the maintenance of precise alignment without

the transmission and introduction of vibration is essential. In the

precision grinding industry, air bearing spindles which maintain both

precise alignment and frictionless contact are utilized. Such a

precise spindle was used in the High Temperature Wear Facility. A

general schematic of the drive system is presented in Figure 16, and a

more complete description of this part of the apparatus is given in the

Overall Design Presentation portion of this work.

Air Bearing Spindle

for Vibrationlesa

Rotation

Air Bearing Jacket

for Alignment

Variable Speed

D. C. Perm. Mag.

Figure 16. General Schematic of Drive System
86

7.3 Design of the High Temperature Wear Facility Machine

Framework

Several basic criteria dictated the design and materials selection

for the High Temperature Wear Facility. High strengths, thermal

stability, machinability , and availability were necessary for the frame

material. Two common structural materials are aluminum and steel. The

strengths properties of aluminum and steel are given below: '^^

Yield Tensile Brinell

Material Strength Strength Hardness Density Melting

(kpsi) (kpsi) (lb/in3) (°F)

Aluminum 5 13 23 0.098 1190

Steel 38 65 130 0.284 2760

Steel was chosen as a structural material because steel has a

higher strength, hardness, and melting point than aluminum. Thus,

steel provides better protection in the event of a failure in the

system insulation. To assure that the frame would provide safe and

secure support, square steel tubing and multiple design redundancies

were designed into the framework.
87

Chapter 8

DESCRIPTION OF THE HIGH TEMPERATURE WEAR FACILITY

This chapter describes, in simple schematics, the design of the

High Temperature Wear Facility. Overall schematics of the wear tester

are shown in Figures 17 to 19. This can be divided into the following

components/assemblies. They are the: 1) lower ceramic specimen holder

assembly, 2) upper ceramic specimen holder assembly, 3) high

temperature furnace, 4) air bearing spindle and loading mechanism

assembly, 5) unidirectional motor assembly, and 6) frame assembly.

Finally, after showing each of the assemblies separately, the major

assemblies will be placed into the frame of the High Temperature Wear

Facility to give the reader some sense of spatial relationships.

Figure 20 gives a general schematic of a pin-on-disk

configuration. In the High Temperature Wear Facility, a similar

configuration exists in which a rotating disk contacts a stationary pin

or ball specimen.

The disk specimen is held by the lower ceramic specimen holder

assembly. This assembly consists of 3 major components. A) a metallic

arbor which fits into a rotating spindle, B) a ceramic tube which is

connected to the arbor, and C) the ceramic disk specimen holder.

Figure 21 describes the lower assembly in a three-dimensional view.

The arbor connects a rotating spindle with the high temperature

components which protrude into the high temperature zone of the
Figure 17. Overall Schematic Diagram of the Wear Tester
(o(f centei

out ol Dags)

Air Bearing Spindia

UNIDIRECTIONAL

TEST

CONFIGURATION

OVERALL

VIEW

Air Bearing

Upper Specimen

Radiative Cooling Fins

Drive 9 en

Figure 18. Unidirectional Test Configuration
90

Rotating Stiatt

Figure 19. Unidirectional Wear Test Configuration i
91

Figure 20. Pin-on-Disk Conf iguraCion
Figure 21. Lower Ceramic Specimen Holder Assembly
93

facility. Connected to the arbor at the upper surface is a ceramic

tube. The arbor and tube are joined through careful design of an

interference fit between the tube and the arbor. Connected to the top

of the ceramic tube is the disk specimen holder. This specimen holder

is constructed out of a ceramic material and will be placed in the high

temperature zone of the facility. To connect the specimen holder to

the ceramic tube, a very carefully designed ceramic-ceramic

interference fit was designed.

The upper specimen holder assembly, which holds the pin or ball

specimen, consists of four major components. These are a) the upper

specimen sample holder (Figure 22) , B) the hydrostatic bearing mount

(Figure 23), C) the ceramic tube, and D) the Inconel topcap. The upper

specimen holder is constructed out of a ceramic material and is the

riding piece in the hydrostatic gas bearing. This bearing was designed

to maintain a frictionless contact within the high temperature zone of

the facility. Connected to the upper specimen holder is a ceramic rod

which extends outside of the high temperature zone so that friction can

be measured.

The upper surfaces of the sample holder are machined to ultra-high

precision to match the mating surface of the hydrostatic gas bearing

mount. This mount is made of the same material as the sample holder so

that the effect of differential thermal expansion is minimized.

Connected to the hydrostatic bearing mount is a ceramic tube. This

ceramic tube is designed to have an interference fit with the

hydrostatic bearing mount. Finally, to complete the upper assembly, an

Inconel topcap is connected to the tube by means of a ceramic-metal
94

Figure 22. Upper Sample Holder Diagram
Figure 23. View of Complete Upper Assembly
96

interference fit. The upper assembly is bolted to a plate which is

connected to a translation weigh which is, in turn, bolted to the frame

of the facility.

The high temperature furnace is another major component of the

facility. The ceramic upper and lower assemblies fit into the

specially designed furnace. This furnace, which can be split into left

and right halves for specimen access, was specially designed and

consists of two major components: A) the main furnace and B) the

furnace annex. The furnace contains several bayonet type SiC heating

elements specially designed to operate at temperatures to 1500°C.

Figure 24 gives a front on view of the furnace chamber, furnace annex,

and heating elements. A distinction should be made between the main

furnace chamber and the lower furnace annex for the following reason:

The furnace annex, which is located below the main furnace chamber,

remains fixed to the frame directly and cannot traverse side to side.

The main furnace chamber is attached to the translation weigh along

with the upper ceramic assembly. This allows side-to-side translation

so that off center testing can be done for the pin-on-disk

configuration.

Next comes the air bearing spindle and loading mechanism assembly.

Below the Inconel arbor in Figure 25 is the air bearing spindle and

loading mechanism assembly. This assembly allows rotation of the lower

(disk) assembly as well as applies a normal load to the specimen

contact. The assembly consists of three main parts: A) the linear air

jacket, B) the air bearing spindle, and C) the loading piston. The

linear air jacket provides an air film which maintains precise spindle
Figure 24. Furnace Chamber and Heating Elements
98

Figure 25. Design Layout - Upper and Lower Assemblies and Furnace
99

alignment while allowing vertical and frictionless translation of the

spindle. The air bearing spindle can rotate and has several internal

thrust and journal air bearings. This allows frictionless rotation of

the spindle shaft, and, since the spindle is connected to the lower

assembly, easy rotation of the disk specimen. An 0-ring drive pulley

is attached to the lower end of the spindle shaft to allow the driving

of the spindle. To facilitate vertical translation of the spindle a

pneumatic piston is attached to the bottom of the specially designed

spindle extension. This allows an axis loading of the test specimens.

Figure 26 shows the complete air bearing spindle assembly.

The unidirectional motor assembly drives the air bearing spindle

through the 0-ring drive consists of several components. These are:

A) the motor, B) speed reducer, C) 0-ring drive, D) 0-ring belts, and

E) base plates. To keep the High Temperature Wear Facility as free

from vibration as possible, the motor is mounted on its own base plate

and vibrationally isolated from the surroundings through the use of

vibration feet. A general schematic of the drive and spindle system is

given in Figure 27. A 3-D visualization of the front and side views of

the motor and spindle system are given in Figures 28 and 29.

The frame is presented in Figure 30 together with the furnace,

motor, and lower drive assembly.
100

Figure 26. Air Bearing Spindle Assembly
101

Air Bearing Spindle

for Vibrationless

Rotation

Air Bearing Jacket

for Alignment

MulUplc

0-Ring C

Drive

Variable Speed

D. C. Perm. Mag.

Motor

Figure 27. General Schematic of Drive and Spindle System
102

Figure 28. Front View - Motor Spindle System
103

Figure 29 Side View - Motor and Spindle System
104
105

Chapter 9

INSTALLATION AND INITIAL TESTING

Equipment installation and demonstration of the High Temperature

Wear Facility has been carried out. Because of the number of

components and the complexity of installation a summary checklist of

the components is given below. Initial experimental results were

presented to demonstrate the design integrity.

This section is divided into the following subsections for

convenience

:

9.1 Installation of Equipment

9.2 Testing of Equipment

9.2.1 Preliminary Testing

9.2.2 Friction Testing - Integrated Mechanical Wear

Facility Testing

9.1 Installation of Equipment

Item Installation Completed

(mo/yr)

General

High Temperature Wear Facility Mounting Feet 5/86

Setco M-50 Dovetail Slide 5/86
106

Item Installation Completed

(mo/yr)

Heating System/Temp. Meas

.

SiC Helical Heating Element 6/86

Mellen 1500°C Furnace 6/86

Mellen 1500°C Annex 6/86

Euro therm Furnace Controller 9/86

Mellen PS400 SCR Power Supply 9/86

Omega Engineering DP900 Temp. Meas. Device 10/86

Type S Pt Rd Temp. Meas. Device 12/86

Mikron M6T Infraducer 9/86

Protective Cooling Jacket 9/86

Air Purge Assembly 9/86

Mikron 60TS Digital Temp. Indicator 10/86

Motor Assembly

Motor Base Plate Assembly 7/86

Motor Leveling Brackets/Side Assembly 7/86

MWU 6 Oil Reducer 7/86

Fincor #9037509 TF 7/86

DC Permanent Magnet Motor

Motor Controller-Fincor 10/86

10 Turn Potentiometer 10/86

Assembly/Run Stop

Switch Assembly
Item

Air Bearing Spindle Assembly

Westwind Air Bearing

Spindle Model 107

0-Ring PulleyLinear Alignment Air Block

Loading Modification

0-Ring Drive

Accushnet 0-Ring EPDM

Loading Assembly

ARO Low Friction Cylinder

0415-1204-030

Entram ELF- 100 Load Button

Cylinder/Horseshoe Load Connector

Ceramic Assemblies

Spindle/Arbor Connector Rod

Arbor To Tube Assembly

Zirconia Foam Insulation Lower Assembly

Lower Sample Holder/Tube Assembly

Topcap To Tube Assembly

Tube TO Hydrostatic Mount Assembly

Gas Supply Tube & Seal System

Zirconia Foam Insulation Upper Assembly

Friction Measurement System

Installation Compl

(mo/yr)

10/86

10/86

9/86

9/86

10/86

12/86

12/86

12/86

12/86

11/86

11/86

11/86

12/86

12/86

12/86

12/86
Item

Specimen Holder/Ceramic Rod Assembly

Accelerometer/Force Transducer Bracket

Ceramic Rod/Radiation Shield Assembly

Shield Force Transducer Assembly

Laboratory Modification

Install Water Cooling Lines

Modify Existing Counter Space

Special Designed Canope Hooding

Installation

Electrical Modification/Power Supply

Installation

Canope Hood Folding Doors

Installation of 100 psi Air Lines

Installation of Argon/Compressed Air Gases

108

Installation Completed

(mo/yr)

12/86

10/86

11/86

11/86

2/86

2/86

3/86

3/86

7/86

3/86

9/86
109

9.2 Testing of Equipment

9.2.1 Preliminary Testing (OK- J)

Furnace System

- Motor System

Furnace y Motor y

Controller J Controller J

xUwcL OUpUiV 1

Temperature Measurement System Loading System & Spindle

Pyrometer & Supply J Pneumatic Cylinder J

Independent Temp . Meas

.

J Load Cell J

Linear Block J

Air Bearing Spindle J

Ceramic Components Force Transducer System

Lower Assembly J Force Transducer J

Temp. Ramp -Lower J Upper Sample Holder J

Upper Assembly J

Temp. Ramp -Upper J
110

9.2.2 Friction Testing - Integrated Mechanical Wear Facility

Testing

Room Temp

Low Load

Low Speed

Low Load

High Speed

High

All Sp eeds

High Load

All Speeds

Low L

High

oad

Speed

Low Load

Low Speed

300 C

500 C

Low Load

Low Speed

Low Load

High Speed

High Loads

All Soeeds

Low Loads

All Speeds

1000 C

Figure 31. Frictional Testing Plan

The overall frictional testing plan for the high temperature wear

facility is shown in Figure 31. Initial testing occurs at room

temperature followed by elevated temperature tests. The friction test

presented consists of a 0.5" diameter, 99.8% sintered aluminum oxide

ball contacting a 99.8% sintered alumina 2" diameter disk. Each of

these specimens were mounted in a ceramic insert which accommodates

differences in thermal expansion between the specimens and the ceramic
Ill

specimen holders. To facilitate proper seating of specimen- insert

pairs, inserts were fabricated with matching tapers to their specimen

holder counterparts.

For the test shown in Figures 36 and 37, specimen contact load was

0.502 kgf (4.92 N) , and a rotation rate of 15 rpm at a 19.05 mm contact

radius. This gives a linear speed of 29 mm/sec. Test temperature was

20°C and relative humidity at this temperature was 60% at test time.

Test specimen preparation consisted of ultrasonic cleaning of specimens

in hexane and acetone followed by Bonification in a laboratory

detergent solution. To remove residual detergent from the surface

several rinses with 10 Mil de- ionized water were done.

Figure 36 shows frictional results from the wear test. The

average frictional output level for the test is 298 mV output (0.135

kgf). Coefficient of friction for this wear test is 0.27 which is

consistent with alumina four-ball data run under same conditions with

similar cleaning procedures. Figure 37 shows a typical vibration

frequency scan for the vibration due to friction. As can be seen on

the figure, vibrational levels drop dramatically as frequency is

increased. This is to be expected in a system in which severe wear of

the specimen does not occur and is consistent with anticipated results.

This result demonstrates the mechanical soundness of the design.

High temperature wear data were not available due to the unexpected

failure of a ceramic part which was subsequently sent for repair. The

furnace and the high temperature capability was separately checked and

the results are shown in Figure 32. The system operated as expected.
Furnace Test

RT - lOOC

lOOC

100 - 250C

250C

250 - 500C

500C

500 - 600C

600C

600 - 800C

800C

800 - lOOOC

15 min

2 hrs

30 min

2 hrs

1 hr

2 hrs

30 min

2 hrs

1 hr

4 hrs

1 hr

Figure 32. Initial Furnace Testing Ramp Up
113
114

uudj
115

117 ^ N
11

Chapter 10

CONCLUSIONS AND RECOMMENDATIONS FOR FUTURE WORK

The design and development of the High Temperature Wear Facility

represents a unique task in the field of tribology. In the design of

this state-of-the-art tribological test apparatus, several new and

novel design techniques were employed in the design of the heat

transfer system, the maintenance of alignment and control of vibration,

the selection of materials for 1500°C wear tester components, and in

mechanical design. This will allow tribological testing of materials

over a wide range of temperatures, speeds, and loads.

Because of the unique capabilities of this facility, several

programs in engine research will benefit from the availability of the

High Temperature Wear Facility. Some of these programs are the

automotive gas turbine program, the adiabatic diesel engine program,

and the Stirling engine program. However, this facility should not be

viewed as a facility which is only useful for the automotive testing

industry. The study of the mechanisms of wear of ceramic materials

using this apparatus will aid the computer industry in understanding

the wear mechanisms in magneto - ceramics , it will help the tool industry

in the development of new, highly wear resistant ceramic tools, and it

will enable both the high temperature bearing industries and the

electronics industry to more fully understand the design of wear

resistant tribological components.
119

Furthermore, in an effort to make the High Temperature Wear

Facility more useful, several future modifications should be

accommodated. The first of these modifications is to expand the

capabilities of this apparatus to include the previously designed

reciprocating wear test motions for direct high temperature engine

simulation. Also, design expansions to include fretting motions should

be implemented, for this will enhance the useability of the facility.

As well as these design changes, several other simple design additions

can be implemented to enhance the attractiveness of the facility.

Modification of the apparatus to allow continuous variance of the

contact radius could allow the performance of a tribological test on

continuously fresh wear surfaces. This modification is simple and

should be implemented in the future. To accent all of the above design

capabilities, computer data acquisition, analysis, and control of the

High Temperature Wear Facility should be fully explored.

For immediate testing, the use of the High Temperature Wear

Facility to develop a model of wear for a ceramic material should be

explored. In such a model, the phase transformations due to the

tribological contact of test specimens, and the generation of new

surface and bulk chemical species due to the interactions of the tribosurfaces should be examined. In examining these parameters, a more

effective understanding of the kinetics and the thermodynamics of the

high temperature tribological contact may be possible. Since

tribological contact is a highly non-equilibrium occurrence, friction

and wear models which incorporate the non-equilibrium thermodynamic

transport equation may be useful. If such a model can be formulated, a
120

major advance in the triboiogical design of ceramic components will be

realized

.
121

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129

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.

. Friction and Wear Devices . 2nd Edition, American

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. Vol.

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.

. "Thermal Barrier Pistons and Their

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130

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pp. 23-31 to 23-40.
Appendix A

Review of Wear Test Configurations
132

A REVIEW OF WEAR TEST CONFIGURATIONS

Introduction

Wear testing of materials is one of the most diverse fields known

to engineering. Throughout the tribological testing community, many

wear test configurations exist. Since the choice of wear test

configurations is fundamental to wear tester design and development,

it is fortunate that many tribologists , more ceramic scientists,

metallurgists, materials scientists, and engineers have similar wear

test configurations. This review will attempt to classify some of

the major wear test configurations and point out strengths and

weaknesses in each of the configurations.

Wear test configurations may be categorized as follows: Class 1)

Point and multiple point contacts. Class 2) Line and multiple line

contacts, and Class 3) Area and multiple area contacts. Within each

of these classes, a wide range of motions might be implemented. For

instance, tests in which a sliding motion, a rolling motion, or a

mixture of rolling or sliding motions could be performed. A

reciprocating motion, a vibratory motion, or unidirectional motion

could be utilized in a tribological tester.

Presented below is a summary of several test configurations in

each of the three major classifications. A schematic of each of the

wear test configurations is given in conjunction with an explanation

of the strengths and weaknesses of each configuration. For a more

complete listing of wear test configurations. Friction and Wear
133

Devices . published by the American Society of Lubrication Engineers,

provides a summary of over 230 different wear test devices.

Class #1 - Point and Multiple Point Contacts

A. Single Point Contacts

1) Ball on Ball Contact

Figure 38. Two-Ball Contact

In this configuration, 2 balls contact each other as shown in

Figure 38. Both balls may rotate in either direction to simulate a

wide variety of sliding and rolling contact ranges as well as different

slide/roll ratios.
134

Strength: For metallic samples, it is reasonably easy to get well

machined, well finished metallic balls for testing. The configuration

is simple, and a wide range of rolling or sliding contacts is possible.

Limitation: At high loads, unless a very hefty and precise

alignment system is employed, contact misalignment may occur.

2) Ball on Flat/Ball on Disk Configuration

Figure 39. Ball on Flat/Ball on Disk Configuration

In this configuration, a ball is contacted with a rotating disk in

an off center position. Normally the radius of contact may be varied

to vary linear speed for a given rotation rate. Also, the possibility

exists for continuously varying the contact radius and conducting a

test in which fresh surface on the rotating flat is always contacted.
135

Strength: A wide range of linear speeds is usually possible by

changing the contact radius at a given speed.

Limitation: For new materials, samples for the rotating disk

specimen may be difficult to obtain.

3) Ball on Ring Configuration

Figure 40. Ball on Ring Configuration

In this configuration a ball is loaded against a rotating ring.

The ring turns and causes slip between the ball and the ring.

Strength: A wide range of linear speeds is possible by changing

the rpm rate of the ring.

Limitation: Stub mounted ball bearings are difficult to find
136

and/or manufacture in large quantities. Therefore, a large number of

tests should be performed per ball specimen.

4) Ball on Cylinder Configuration

In this configuration, a ball is loaded against a rotating rod or

cylinder. The cylinder is partially immersed in a lubricant test bath

if lubrication studies are desired.

Strength: For conventional materials, ball stock and rod stock

are common shapes. For new materials, the rod stock is very common.

Limitation: Analyses of wear tracks by SEM or other methods are

difficult because of the bulkiness of the cylinder specimen. For new

materials ball stock may be expensive.

Figure 41. Ball on Cylinder Configuration
5) Pin-on-Disk - Rim Loaded

137

Figure 42. Pin-on-Disk - Rim Loaded

This configuration consists of a hemispherically tipped pin that

contacts a rotating disk on the curved rim face. Note that the

configuration would be quite similar to configuration #4 - Ball on

Cylinder Configuration if a) the cylinder was very short and, if b) the

ball surface was replaced by a hemisphere on the end of a piece of rod

stock.

Strength: Almost any material can be obtained in either a disk

configuration or a rod stock shape. Pins can be machined from rod

stock and contact pressures for a given load can be varied by changing

the curvature of the hemispherical tip.
138

Limitation: It is desirable to vary the radius of contact to vary

linear speed given rotation rate. This cannot be done with a rim

loaded pin-on-disk configuration.

6) Pin-on-Disk - Face Loaded

A face loaded pin-on-disk configuration consists of a

hemispherically tipped pin contacting the flat face of a rotating disk.

Strength: The following parameters can be varied. 1. For a

given loading on disk, the curvature of the pin tip can be changed to

vary contact pressure. 2. For a given revolution rate, the linear

speed of the pin vs. disk surface may be varied by changing the contact

radius.

Figure 43. Pin-on-Disk - Face Loaded
139

As in the rim loaded pin-on-disk configuration, both types of

samples are relatively easy to obtain for new materials.

Limitation: Mounting of the disk do avoid slip may be difficult.

7) Pin-on-Disk - Non-rotating Configurations

Figure 44A. Pin-on-Disk - Fretting Motion
140

These configurations are similar to the pin-on-rotating-disk

configuration with the exception that 1) The "disk" is not necessarilycircular, and 2) The disk does not rotate. A distinction has been

made between testers that test very high frequency - low amplitude

motions which induce a fretting/fatigue type of wear versus testers

which test at larger amplitudes and lower frequencies to test

reciprocating wear. Practically, design of testers to facilitate these

motions may be quite different.
8) Cylinder on Disk Rira Configuration

141

Figure 45. Cylinder on Disk Rim Configuration

In this configuration, a cylinder or rod is loaded against the

edge of a rotating disk as shown in Figure 45.

Limitation: Perfect vertical alignment of cylinder needs to be

maintained. If this is not maintained, a line contact is formed rather

than a point contact. (In the extreme case, if the cylinder were

horizontal, a line contact would develop across the entire rim of the

disk.)
9) Crossed Cylinder Devices

142

Figure 46A. Crossed Cylinder Device - Reciprocating Cylinder
1 43

Figure 46B. Crossed Cylinder Device - Rotating Cylinder

These configurations have the same basic limitations as that of

configuration #8 - cylinder on disk rim configuration. In fact, if the

disk could be viewed as a short cylinder we have the same configuration

as the configurations 46A and 46B.

B) Multiple Point Contacts

Many multiple point contact testers exist. For the sake of

brevity, a few of the major multiple point contacts are presented below

along with a brief summary of the strengths and limitations of each

configuration.
10) Four-Ball Wear Test Configuration

144

Figure 47A. Four-Ball Wear Test Configuration

This configuration consists of 3 lower balls which are held

stationary and an upper ball which is loaded against it and rotated is

a commonly used wear tester for boundary lubrication.

Strengths: By nature, this test is a self -aligning wear test. It

is widely used throughout the lubricant industry for testing lubricants

under boundary lubrication. A common variation this configuration is

the bail on 3- flat configuration in which 3 flats are arranged at the

contact points of Figure 47A and thus a tetrahechal configuration is

preserved. Figure 47B gives a general schematic of this configuration.
145

Figure 47B. Bail on 3-Fiat Configuration

Limitation: A four-ball tester does not possess the capability of

large linear speeds unless a very high rotational speed are possible.

For new materials, ball bearings may be expensive to machine. Contact

pressures are high, even at low loads.
11) Five Ball Wear Test Configuration

146

Figure 48. Five Bail Wear Test Configuration

In this configuration, 4 lower balls are separated by a separator

and the upper ball is loaded against them and rotated. The lower balls

are free to roll and thus a simulated bearing contact is formed. A

reference where this type of configuration was used in Zaretsky, E.V.,

Parker, P.J., and Anderson, W.J., "Component Hardness Differences and

Their Effect on Bearing Fatigue," ASME Trans. Journal of Lubrication

Technology . January, 1967, p. 47.
12) Multiple Pin-on-Disk Configurations

147

V

Figure 49. Multiple Pin-on-Disk Configuration

Multiple pin-on-disk configurations have the same strengths and

limitations as single pin-on-disk configurations with the following

exception. In a configuration such as above a load applied along the

central axis of rotation does not cause uneven deflection of the disk

at very high loads.
148

Class #2 - Line and Multiple Line Contacts

A. Line Contacts

13) Disk Rim on Disk Rim

Figures 50A and SOB show two different views of this type of

configuration.

Figure 50A. Disk Rim on Disk Rim
149

Figure SOB. Implementation #2 Disk Rim on Disk Rim

Configurations of this type are commonly used to simulate pure

sliding and combinations of rolling and sliding which might occur in

tooth friction for gears. The configuration tries to simulate fatigue,

gear wear and scoring by changing the slip to roll ratio for a given

speed.
14) Rotating Cylinder on Plate Configuration

150

Figure 51. Rotating Cylinder on Plate Configuration

Configurations of this type are cominonly used. They are commonly

used in the testing of lubricants under boundary lubrication. Since a

line contact is larger than a single point contact, more wear debris is

generated under the same conditions and wear monitoring of a

configuration such as this becomes possible. Careful vertical

alignment of the rotating cylinder should be maintained in order to

prevent the tendency of the plate to twist. This would skew frictional

force measurements

.
15) Ring-on-Block Configuration

151

Figure 52. Ring-on-Block Configuration

The ring on block configuration is one of the most common

configurations for a line contact. In this configuration, a ring is

contacted with a steel block that is loaded against it. This

configuration is commonly used throughout the tribologicai world for

general sliding friction and continuous wear measurement. Lubrication

studies can be done by immersing the ring and block in lubricant or by

suspending a drop of lubricant on the bottom of the ring. The

advantage of the latter is that continuous collection of the wear

debris in the oil droplet is possible.
152

Limitations: If the block is not maintained in perfect alignment,

odd shaped wear scars result.

B. Multiple Line Contacts

16) Three Pellet on Disk Configuration

In this configuration 3 short cylinders, or pellets contact a disk

on their curved surfaces. Thus 3 lines of contact are formed. This

configuration, originally designed for solid lubrication [see Lavik,

M.T. , "High Temperature Solid Dry Film Lubrication," WADC Technical

Report 57-455, Part III, 1959]^, would probably be a poor design for

Figure 53. Three Pellet on Disk Configuration
153

metal pellets because wear debris that is generated in one contact

could be trapped in another contact thus causing 3 -body wear.

17) Ring on 2 Block Configuration

Figure 54. Ring on 2 Block Configuration

In this configuration, two blocks are loaded against a rotating

disk and two line contacts are formed. Friction is determined by

transfer of torque from the cylinder to the blocks. Continuous wear

measurement may be obtained by continuously measuring the positioning

of the blocks.
Class #3 - Area Contacts

Many different area contacts have been devised. A brief summary

of a few of these area contacts follows.

18) Pellet/Slider on Plate

Figure 55A. Pellet on Reciprocating Plate
Figure 55B. Slider on Reciprocating Plate

This configuration is quite similar to the pin-on-disk

reciprocating motion configuration (44B) with the following change:

Instead of a curved surface at the tip of the pin, a flat machined

surface is used in the testing. This allows the contact area to be

circular area rather than a "point" contact.
19) Pellet on Disk Configuration

156

Figure 56. Pellet on Disk Configuration

This configuration is similar to configuration #43 - Face loaded

pin-on-disk except that a flat has been machined on the disk end rather

than a surface of curvature.
157

Conclusions

Many different configurations exist and are widely used throughout

the tribological testing community. Each configuration is suitable to

a fixed range of conditions, speeds, loads, and environments.

Different configurations lend themselves more easily to different types

of analyses. For example, large cylinders do not usually lend

themselves easily to SEM analysis but are easily analyzed with an

optical microscope. It would be convenient to have a wear tester that

has the capability to test a number of different configurations.

Obviously, no one tester could test all configurations under all

conditions. Such a tester could test no one configuration well.

However, in certain systems, not much is lost in adding some

variability. For instance, let us look at the face loaded pin-on-disk

configuration: If we allow the variability to replace a ball for the

pin, then we have a ball-on-disk configuration. If we have the

capability to machine the pin end to a flat, then we have an area

contact -a button-on-disk contact. If the contact radius can be varied

for a given rpm rate we can test different linear speeds for these

contacts

.

If possible, such an approach should be taken in wear tester

design. This approach allows specimen variability without the cost or

sacrifice. If certain new materials come only in certain specimen

shapes, these shapes need to be utilized. It is necessary, whenever

possible, to accommodate such shapes in a wear tester. Therefore, for
the High Temperature Wear Facility, a ball-on-disk/f Lat - pin-ondisk/flat configuration has been chosen.

1. Benzig, et al

.

. Friction and Wear Devices . 2nd ed. , American

Society of Lubrication Engineers, 1976, 458 pages.

2. Lavik, M.T., "High Temperature Solid Dry Film Lubrication," WADC

Technical Report 57-455, Fart III, 1959.
159

Appendix B

Development of Finite Difference Equations for

High Temperature Wear Facility Furnace Design
160

SUMMARY OF EQUATIONS AND BOUNDARY CONDITIONS

Steady State Energy Transport Equation

i L (rq.) + L (q.) = 0 r 5r dz

dT = -Kff -

dz

ai

5r

Boundary Conditions

@r=4, 0<z<3, T= 1500°C

(ao<r<4, z = 3, T = 1500°C

(a4<r<10, z = 0, f^ = 0

(ar = 0, 3<z<9, ^-0

ar

(a r = 10, 0 < z < 9, q^

= h(T - T ) wall room
161

(a 0 < r < 10, z = 9, q = h(T - T ) wall room

T = constant

room

h = constant

Boundary Conditions Summarized

q, = h(T ,

, - T ) wail room

qr = h(T

room

T = 1500

T = 1500

81

ar

= 0

3T = 0

Case I k = constant

eft

1 a dT . a r + _

r ar ar az

ai

az = 0
162

or, in an alternate form

1 dT d^T a^T _ + + = 0

r ar ar^ dz^

Note: At r = 0 dT/dr must be zero, or the finite difference equation

would blow up to infinity. This is one of our boundary conditions.

Now finite difference equations for internal grid points and for each

boundarycan be developed. The figure below marks out each type of

boundary and internal grid point.

II

1500°C

V

1500°C

IV
163

Each of the regions I, II, III, IV, V, and internal node points will

have different finite difference equations

.

Internal finite difference equation:

Grid points

(i, j + 1)

• • • (i + 1, j) (i, j) (i - 1, j)

(i, j

- 1)

= A. (T(i - 1, j) -2T(i, j) + T(i + 1, j)

Sr^ Ar^

= J_ (T(i - 1, j) + T(i + 1, j)

dx 2Ar

r = (i - l)Ar

= J_ (T(i, j

- 1) - 2T(i, j) + T(i, j + D)

2 < i < M - 1 m...this particular ^-1 region

2 < j < N - 1
164

Finite difference equation:

T(i+l,j)] ^ [T(i-l,j) + T(i+l,j)] + J_ [T(i-l,j) - 2T(i,j) + 2(i-l)Ar2 Ar^

+ __ [T(i, j-1) - 2T(i, j) + T(i, j+D] = 0

Az^

or, gathering like terms and solving for T(i, j) we have

T(i, j)

1 1 1 2i - 1

2 ±- 1

h Ar^ 2(i-l) {T(i-1, j) + T(i+1, j)) +

hi} Ar^

J_ {T(i,

Az^

j-1) + T(i, j+1)}

Region I Finite Difference Equation

= 0 is the governing equation

Expressing 3T/3R in finite difference form:
165

Grid points for region I

(i, j + 1)

(3, j) (2, j) (1, j) Region I Grid Points

(i, j

- 1)

Since this is a special boundary, expand about Taylor series to obtain

a finite difference equation (OAr^ error) to give:

!Z = _i -I T(l,j) + 2T(2,j) - 1 T(3,j) = 0 ar Ar 2 2

where j ranges from Y to N.

If Ar 0 solving for T(l, j

) we have

T(i. j) = T(2, j) - 1 T(3, j)

3 3

Region III Finite Difference Equation

Governing differential heat transfer equation at wall

^eff

dz = h(T . . - T ) wall room

wall
166

Expressing 31 in finite difference form.

wall

Grid points for Region II:

Region II Grid Points

(i+2, N) (i+1, N) (i. N) (i-1, N)

(i+2, N-1) (i+1, N-1) (i, N-1) (i-1, N-1)

Again, this is a special boundary, expand the Taylor series about this

boundary to obtain a finite difference equation with order of error of

Lt? {QiLt? ) .

This gives the equation

aT ^ 1 5z Az

1 T(i, N) - 2T(i, N-1) + £ T(i, N-2)

2 2

If we substitute this directly into the governing differential heat

transfer equation we have:

^eff

Az

i T(i, N) - 2T(i, N-1) + \ T(i, N-2) = h(T(i, N) - T ) room
Solving for T(i,N) we have

T(i,N) =

3 hAz _ +

^ ^eff

2T(i, N-1) - i T(i, N-2) + T

2 Klf

room

i = 1 to M

Region III Finite Difference Equation

Governing differential heat transfer equation at wall

- k = h(T , , - T ) wall room

wall

Expressing in finite difference form, we obtain:

wall

Grid points for Region III

Region

III

Grid

Points

(ra, j+1)

(m, j)

(m, j-1)

(m-1,

(m-i, j)

(m-1, j-1)
168

Performing Taylor series expansions about Region III grid points, we

obtain (OAr^ )

5T ^ 1 dr Ar

1 T(m, j) - 2T(ni - 1 , j

) + 1 T(m - 2 , j

)

Substituting into the governing finite difference equation and solving

for T(m, j) we obtain:

T(m, j) =

3 hAr

+

2T(m -1, j) - 1 T(m -2, j) + T

2 k^^^ room

2 keff

j = 1 to N

Region IV Finite Difference Equations

Governing equation for this region:

= 0

Grid Points for Region IV:

(i+1, i) (i, 1) 1) (i-2, 1)

(i+1, 0) (i, 0) (i-1, 0) (i-2, 0)
169

Region IV Grid Points

Using the Taylor series to obtain an order Az^ error approximation for

dT

we obtain:

- i T(i, 0) + 2T(i, 1) - i T(i, 2)

2 2

Plugging this into the governing equation for this region:

T(i, 0) = T(i, 1) - 1 T(i, 2)

3 3

i ranges from X to M

Region V Finite Difference Equation

Governing Equation:

T = 1500°C

dT ^ 1 dz Az
170

Grid Points

(X,Y) (X-1,Y) (X-2, Y) . . . (1, Y) (0, Y)

(X, Y-1)

(X, 1)

(X, 0)

Finite difference form:

T(X, j) = 1500 j = 0, 1, . . . , Y - 1, Y

T(i, Y) = 1500 i = 0, 1, . . . , X - 1, X

Summary of finite difference equations for k^^^ = constant case

Internal Grid Points

T(i, j) =

1 1 + +

Az2 Ar^

1 ai - 1 {T(i-1, j) + T(i+1, j)}

Ar2 a(i-l)

{T(i, j

- 1) + T(i, j + 1) Az'
171

Region I

T(l, j) = T(2, j) - 1 T(3, j)

j = y, y+1, , N-2, N-1, N

Region II

T(i, N) =

3 hAz _ +

^ ^eff

2T(i, N-1) - i T(i, N-2) + T

^ ^eff

room

i = 1, 2, , M-2, M-1, M

Region III

T(M, j) =

3 hAr

+

2T(M-1, j) - i T(M-2, j) + T

^ ^eff

room

2 k

eff

j = 1, 2, , N-2, N-1, N

Region IV

T(i, 0) = T(i, 1) - i T(i, 2)
172

i = X, x+ 1 , .... , M-2 , M-1 , M

Region V

T(X, j

) = 1500, j-0, 1, ...Y-1,Y

T(i, Y) = 1500, i=0, 1, ... X- 1, X

To solve the finite difference grid use the following procedure:

1. Use the equations above to write one and only one finite

difference equation for each node point. Note that at

intersections of regions I and II, II and III, III and IV,

IV and V, and V and I, two equations for boundary conditions

could be written. Use only one boundary condition for each

of these points.

2. Once you have gotten one equation for each of the P node

points, solve all of the linear equations simultaneously for

the temperatures at the node points.

3. Change the finite difference grid size and re- iterate from

Step 1 until temperatures at the outer shell do not change

appreciably with changing grid size.
173

Example

:

For the grid given on page 3 of this appendix:

1. There are 98 nodes in the finite element grid. Thus we

would need to write 98 finite element analysis equations.

2. Solve 98 simultaneous linear equations using some technique

to solve these equations (such as matrix inversion or

diagonalization)

.

3. Increase the number of grid points and re-solve the

equation. If the outer shell temperatures have changed

appreciably, increase the number of grid points and reiterate.

Case II k = k ^^(T) eff eff

All boundary conditions (Regions I - V) remain the same as in

Case I except that in any equation where keff is used it is to be

evaluated at a node under inspection.

The Steady State Heat Transport Equation becomes

1 a , aT ^ 3 , 5T _ + r k + k =0

r ar ar az az
174

—dz

%ff —az ^ eff TZoz J-az

Further expanding we have

a , ax 2 , 1 ax a^x k-^ +^p^r + + ax ar r ar ar^

^ k

' + k = 0 ax az az''

Analysis of this equation and solution proceeds in the same manner as

in Case I except^ that now since k is a function of X iteratively

eff ^

solve the P finite element equations until 1) Xhe temperatures and

the k ^^'s for each node do not change, i.e., (estimate k ^^'s

eff ° eff

calculate X's use calculated X's to re-estimate k r-^'s, etc.).

err
175

Appendix C

Gas Bearing Design Calculation for the

High Temperature Wear Facility
Rq is the outer radius of the hydrostatic thrust bearing

is the inner radius of the hydrostatic thrust bearing

R^ is the radius of hole centers

For all of the design charts to be given, it is assumed that R^ =

VRq X R^ . This gives a maximum of load carrying capabilities with a

minimum of gas flow.

* Personal copy of RPI-MTI Gas Bearing Design Course received from H.S.

Cheng, Northwestern University, Chapter 5, Hydrostatic Thrust

Bearings, by H. S. Cheng was used in this calculation.
177

For the design given, there are four dimensionless input

parameters

:

They are

:

1. The ratio of outer to inner radius Rq/R^

2. Supply Pressure Ratio Pg/Pa

where Pg = supply pressure

P^ = pressure of the surrounding atmosphere

3. Restrictor Coefficient

6/i n a2 yRT K =

Pg Jl +

where /i = gas viscosity

n = number of feeder holes

a = orifice radius

R = ideal gas constant

T = absolute temperature

Pg = supply pressure

C = film thickness

and S = Tra^/TTcd = ratio of orifice area to inherent restriction

area.

For inherently compensated bearings 5 = «.
178

4. Feeder Hole Volume Ratio:

n V,

where n = number of feeder hole

= volume of a single hole

C = film thickness

Outputs from Design Charts (MTI , 1966)

Dimensionless load:

W

w =

W = real axial loading of system

Dimensionless stiffness:

1 + i 1 7r(R„2 . R.2) (p^ . pj

3 C

Dimensionless mass flow:
179

Dimensionless damping coefficient:

B =

c

Dimensionless angular stiffness

3 C

Dimensionless mass flow:

, 6>.RT ^

Dimensionless damping coefficient:

B =

*

R

3 M Ro

^

c

Dimensionless angular stiffness

1 + i 52 i ;r (R,2 - R.2) R^2(p^ . pj

3 C
Dimensioniess angular damping
The Design Given in Figure 57 will be Approximated bv an Annular

Bearing as Illustrated in Figure 58

181

Figure 57. Hydrostatic Bearing Design Surface
182

Figure 58. Hydrostatic Supply Surface Approximation
Because the bearing has straight slots we will assume that the

following approximation can be made

Rg - = the distance between the two air channels as shown in

Figure 58 and R^ = the shortest distance between a hole and a slot.

If this is done

Ri = 1/2"

R„ - Ri = 1.25"

so

^ = 3.5

R, = 1.75"

VRq X Ri - Rc

Use design charts for Rq/Ri = 3.0, because this is the closest

design chart for our purposes.

We want to design for maximum stiffness so we will use Figure 5.3.13

from the MTI Gas Bearing Design Course is shown in Figure 59.

185

Assume that the maximum supply pressure is going to be about 120

psia and the ambient pressure will be 14.696 psia. Therefore for the

supply pressure ratio

"° ^''^

' 8 = p./p.

14.696 psia

For anything above Vs/Va ~ ^ maximum A3 =0.9

and

1+52 C K

= 0.4

1 + I 7r(R,2 - R.2) (p^ . pj

To find load carrying capacity Figure 5.3.5 of the MTI Gas Bearing

Design Course is used assuming a Pg/Pa = 6.0 or Pg = 88.176 psia or

close to 90 psia and A3 = 0.9. This is shown in Figure 60. From the

Chart:

W

= 0.27

^(R,2 - R.2) (p^ . pj

use R, = 0.5"

Ro = 1-5"

P3 = 88 . 176 psia

= 14.696 psia

187

Plugging in the numeric values for ,

,

, and we have

W

0.27

7r(1.52 - 0.52) (88.176 - 14.696)

W = 124.65 lbs thrust

or

56 Kg

Therefore we have a load of 56 Kg,

Now a look at stiffness:

1 + 52 C K

= 0.4

3

For an inherently compensated bearing 6 = <»

Therefore 1 + 8'

2 ,2 2

1 + 52

therefore
188

(0.4) X I X n X (R„2 ,

) (Ps - Pa)

K =

3

C

and we have set 1 . 5

tl

0.5 II

Ps 88.176 psia

Pa 14.696 psia

To choose C we want the smallest possible thickness and it is

suggested that as a minimum film thickness to choose 0.5 x 10" ^ or

0.0005" as an estimate.

C = 0.5 X 10-3 in

therefore the stiffness may be given by

K = 246,233 Ib^/in

Estimate of the resonant freqeuncy of the specimen holder

1) Total weight of the specimen holder
189

Assume 2 1/2" diameter by 12" length ceramic rods +1, 4" x 6"

15" block of Carborundum's Hexoloy sintered alpha silicon carbide

which has a density of 3.1 g/crn^ or 3100 kg/m^

3100 kg 2.2046226 lb

X X m kg

0.3048 m

1 ft^

193.52668 lb

total volume

2(7r X (1/4)2 X 12)

+ (4 X 6 X 1.5)

40.7 in^ X

1 ft^

(12)in3

= 0.0235604 ft^

total mass is 4.559568 lb„

4.559568 lb„ x 32.174 ft 1 lb. sec 32.174 Ib^ ft sec
190

4.559568 Ibf x

sec 386.088 in

so

M = 1 . 180966 X 10'2 lb sec m

[N ops] resonant axial vibration

"

2n

K

M

K is in lb/in = 246,233 Ib^/in

and M is in lb sec^/in = 1.180966 x 10'^ Ibf sec^/in

N= 1 246,233 Ibf/in

27r 1. 180966 X 10-2 sec^ /in

N = 726 Hz

Therefore we

700 Hz which

get some sort of

could hinder the

resonance frequency about

performance of the bearing system.
191

Dimensionless bearing flow:

Given that As = 0.9, use Figure 5.3.9 of the MTI Gas Bearing Design

Course to obtain the dimensionless flow parameter. Figure 5.3.9 is

shown in Figure 61.

Given the assumption that

Ps/Pa = 6

we get that

6m RT

G = 0.54

^ Ps' C3

lb sec m 6/i RT

X 0.54

where /x is viscosity in lb sec/in^ (0 Argon)

(The viscosity of argon at room temperature is 0.022 cp)

R is in in^/sec^ °R

lb sec ^ 2.0886 x lO'^ ib^ sec 1 ft^ = 0.022 cp X

^ X

ft2 cp 144 in2

^J. = 3.1909167 X 10 - 9 Ibf sec

in"

C = 0.5 X 10-3 in

T = 68°F = 527.59°R

Ideal gas constant, R

193

1.9872 BTU lb mol Ar , q^.q,. ,^-2 BTU

X = 4.974965 x 10

lb mol °R 39.944 Ib^j, lb °R

To convert to the proper units of

R in

sec^ °R

a factor of 8.6051 x 10^ R. BTU

lb "R = Rsec^ "R

R

sec2 °R

= 1.7935246 x 10^

sec^ "R

Recalling

lb sec _ TT C3 p

2

= X 0.54

in 6/i RT
194

7r(0.5 X 10^3)3 X (88.176)2

r. r- , G = xO.54

6 X 3.1909167 x lO'^ 1.7935246 x 10^ 527.59°R

in^ sec^ °R

^^"^ = 9.1008891 x 10-7 ^'"f in in

G(scfm) = 2.8696969 x 10^ x G

"''^fin

G(scfm) = 0.2611679 (scfm)

Now, calculate the number of gas supply holes necessary for the

bearing. From the gas supply course:

nd = As

6/i JRT

where n = ntomber of supply holes

d = diameter

Plugging in the proven values we have:

88.176 lb, ^ ^ ^Q_3^2

in2

nd = X 0.9

6 X 3.1909167 x 10" ^ JrT
195

17935246 x 10^ in^

where R = and T = 527.59°R

sec^ °R

nd = 1.0652812 x 10"^ in

nd = 0.1065281 in

If we assume that d = 0.025, then we find that n = 4 from flow

calculations. However, from geometric considerations a minimum number

of holes is necessary.

Figure 5.3.19 of the MTI Gas Bearing Design Course gives a

procedure for finding the minimum number of holes. This chart is

shown in Figure 62.

e=lin^=lin3= 0.5493061

2 Ri 2

D = 2 = 2 yi.5 X 0.5 = 1.7320508

1 = 1.0510538

eD

d

therefore = 0.026

eD

From Figure 62

n e . = 8 n . = 14 mm mm

If we decrease hole size to 0.0125"

we need 8.5 holes ~ 9 holes

d- = 0.0131382 n e . =9

and n = 16

If we increase hole to size 0.05"

then ne.=6.1 n„._=ll mm

Calculation of failure load.

W

e = CK

C = 0.5 X 10-3

K = 246,233 Ibf/in

W = 124.65 lbs - 52 kg
198

€ = 1.01

If € > 0.4 the bearing will lock up.

However, at 20 kg = 44.09 lb

e = 0.35

This is a safe operating range. If we plug in e = 0.4 we find that

0.4 X C X K = 49 lbs

Therefore, the bearing should not be subjected to much more than area

22 kg.

Parameter sheet:

Ri = 1/2"

- Ri = 1.25"

R, = 1.75
199

say that = 1.5

Ri =0.5

maximum supply pressure 88.176 psia, 73.48 psig (5 scfm)

As = 0.9

W =56 kg, 124 lbs

cap °

K = 246,233 Ibf/in

C = 0.5 X 10" 3 in

N ... J- = 726 cps

resonance vibration frequency

mass flow rate

G = 9.1008891 X 10"^ = 0.2611679 scfm

m

diameter of feeder holes 0.050"
200

minumum number is 11

Bearing designed for 20 holes

if we look at range that bearing can handle and avoid lock up

W

€ = CK

where e = 0.4 to 0.5 (max)

therefore the very maximum that the bearing should be subjected to

avoid lock up is 28 kg load.

but

22 is a better safe estimate.
U.S. DEPT. OF COMM. BIBLIOGRAPHIC DATA

SHEET (See instructions)

1. PUBLICATION OR

REPORT NO.

NIST/SP-755

2. Performing Organ. Report No, 3. Publication Date

September 1988

4. TITLE AND SUBTITLE

THE DESIGN AND CONSTRUCTION OF A STATE-OF-THE-ART HIGH TEMPERATURE TRIBOMETER

5. AUTHOR(S)

Jeffrey P. Yellets , Stephen M. Hsu, and E. Erwin Klaus

6. PERFORMING ORGANIZATION (if joint or other t^^an NBS, see instructions)

NATIONAL INSTITUTE OF STANDARDS AND TECHNOLOGY

(lormerly NATIONAL BUREAU OF STANDARDS)

U.S. DEPARTMENT OF COMMERCE

GAITHERSBURQ, MD 20899

7. Contract/Grant No.

8. Type of Report & Period Covered

Final

9. SPONSORING ORGANIZATION NAME AND COMPLETE ADDRESS (Street. City. State. ZIP)

SAME AS ITEM #6 ABOVE.

10. SUPPLEMENTARY NOTES

Library of Congress Catalog Card Number: 88-600579

I I Document describes a computer program; SF-185, FIPS Software Summary, is attached.

11. ABSTRACT (A 200-word or less factual summary of most si gnificant information. If document includes a si gnificant

bi bliograptiy or literature survey, mention it here) High temperature ceramic tribology is one of the fastest growing and least understood

areas in tribology. The need to understand the mechanisms of friction and wear for

ceramic materials is critical. Ceramic materials are being utilized to tool

development, bearing design, materials development for low-heat-re jection diesels,

automotive gas turbines, Stirling engines, and aerospace applications. The

development of a ceramics industry capable of manufacturing high technology wear

resistant ceramics is essential in the United States, for if we do not develop such

an industry, we will lose a substantial market to an already rapidly growing foreign

market

.

This effort describes the design, development, and construction of a unique, stateof-the-art, ultra-high temperature tribometer for the mechanistic study of novel

ceramic materials. An overview of ceramic tribology is included in this work as well

as a review of many common wear test configurations. The final designs and the

construction of the High Temperature Wear Facility is presented in this work.

Recommendations for future work on high temperature mechanisms of ceramic materials

and the tribological testing of such materials are presented at the conclusion of

this work.

12. KEY WORDS (Six to twelve entries; alphabetical order; capitalize only proper names; and separate key words by semicolons)

ceramic(s); construction; design; high temperature; tribology; tribometer

13. AVAILABILITY

LxJ Unlimited

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