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Guidelines to best practices for heavy haul railway
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GUIDELINES TO
BEST PRACTICES
FOR HEAVY HAUL
RAILWAY OPERATIONS:
WHEEL AND RAIL
INTERFACE ISSUES
Click Here to View Table of Contents
IHHA ' M AY 2001
G UIDELINES T O
B EST P RACTICES
F OR H EAVY H AUL
R AILWAY O PERATIONS :
W HEEL AND R AIL
I NTERFACE I SSUES
First Edition, First Printing, May 2001©
2808 Forest Hills Court
Virginia Beach, Virginia 23454
USA
These Guidelines have been prepared by the Technical Review
Committee under the auspices of the International Heavy Haul
Association as an input to the decision making processes of heavy haul
railways. They represent the best efforts of the Technical Review
Committee authors and reviewers. The Guidelines are neither mandatory
directives nor intended to summarize and interpret the extensive heavy
haul technical literature. There are special combinations of circumstances in which
the best practices may differ from those discussed in the Guidelines. Therefore, these
guidelines are neither mandatory nor do they describe exclusive methods
to achieve optimum rail/wheel performance.
Co p y r i g h t © 2 0 0 1 I n t e r n a t i o n a l H e a v y H a u l A s s o c i a t i o n
All rights reserved.
Reproduction or translation of any part of this work
without the permission of the copyright owner is unlawful.
R e q u e s t s f o r p e r m i s s i o n o r fu r t h e r i n f o r m a t i o n should be
addressed to:
International Heavy Haul Association
2808 Forest Hills Court
Virginia Beach, Virginia 23454 USA
L i b r ar y o f C o n g r e s s C o n t r o l N o .: 2 0 0 1 1 3 1 9 0 1
Printed in the United States of America 2001
hese guidelines were prepared under the
auspices of the International Heavy Haul
Association through its Board of
Directors:
T
Australia:
Public Railways of Australia,
Brian G. Bock, Chairman IHHA
Private Railways of Australia,
Michael Darby, Director
Brazil:
Companhia Vale Do Rio Doce (CVRD),
Ronaldo Costa, Director
Canada:
Railway Association of Canada,
Michael Roney, Director
China:
China Railway Society,
Qian Lixin, Director
Russia:
The All-Russian Railway Research Institute
Alexander L. Lisitsyn, Director
Sweden/Norway:
Nordic Heavy Haul Association,
Thomas Nordmark, Director
Republic of South Africa:
Spoornet,
Harry Tournay, Director
Union Internationale des Chamis (UIC):
World Division,
V. C. Sharma, Associate Director
United States of America:
Association of American Railroads,
Roy A. Allen, Director
IHHA Chief Executive Officer
W. Scott Lovelace,
o ix
FOREWORD
Letter from the IHHA Chairman on behalf of the
Board of Directors to the readers of Guidelines to
Best Practices for Heavy Haul Railway Operations:
Wheel and Rail Interface Issues
About 50 years ago, railways in many places in the world began to
increase axle loads to provide more efficient and lower cost
transportation of bulk commodities. Serious problems with rail,
track, wheels, and cars emerged. Research was begun by numerous
companies and administrations to overcome these serious problems.
These studies were first shared at an International Heavy Haul
Conference organized in Australia and held in Perth in 1978.
Because of the overwhelming success of these meetings, a second
conference on heavy haul railway engineering and operational
concerns was organized and held in 1982 in the United States at
Colorado Springs, Colo. During that time, the delegates expressed an
interest in the establishment of a continuing organization to facilitate
sharing of information on heavy haul railway technology.
In early 1983, Dr. William J. Harris, then Vice President
Research of the Association of American Railroads, issued an
invitation to the delegates to the 1982 meeting to come to
Washington to discuss establishing a continuing organization. In the
summer of 1983, representatives of railways from Australia, Canada,
China, South Africa, and the United States formally organized the
International Heavy Haul Association. In 1994, railways in Russia
joined, followed in 1995 by the railways of Brazil. More recently, in
1999, the railways of Norway and Sweden joined as the Nordic
Heavy Haul Association. The World Division of the UIC became an
Associate Member in 1999 and participates in meetings of the IHHA
Board of Directors.
In 1991 at the conference in Vancouver, British Columbia,
Braam le Roux, then Chief Executive of Spoornet, raised the
question of developing a handbook of best practices for heavy haul
railways. The handbook was to be based on the collective knowledge
of technical presentations made at this and other IHHA conferences.
This was the beginning of the concept of a “best practice” handbook.
The IHHA has organized six international conferences and ten
specialist technical sessions to encourage the exchange of
information on heavy haul research and development. The IHHA
Board of Directors determined that while these conferences were an
extremely valuable way to disseminate state-of-the-art technology, it
was very difficult for any operating officer to gain the insights
o iii x
provided by the 16 conferences and technical sessions. Therefore,
they agreed that the publication of a handbook of guidelines, which
provided summary information on best heavy haul practices, would
be a useful contribution to the heavy haul rail community around the
world. It was with this goal in mind that the task of writing this book
was undertaken. The Directors established a Technical Review
Committee and charged it with developing guidelines for heavy haul
operations with special attention to the wheel/rail interface. The
dedication of the members of the Technical Review Committee and
the support of their organizations was remarkable. The financial
support from IHHA made the project become a reality. The funding
and support of the Russian Railway Ministry, All-Russian Railway
Research Institute, SPOORNET of South Africa, the Private and
Public Railways of Australia, and the Transportation Technology
Center, Inc. of the United States is also gratefully recognized.
The reader will find that these guidelines summarize the
technological options available in seeking the best practices on a
cost/effective basis. They are presented in a format that will make it
possible for railway operating officers to decide how best to apply
these findings to optimize their operations.
A second edition of these guidelines will be published as soon as
enough new material is available from research or field findings. The
TRC encourages readers to send comments for improvements in the
guidelines by email to Scott Lovelace, CEO of IHHA, at
ihha@erols.com.
Brian G. Bock, Chairman IHHA
o iv x
PREFACE
This handbook summarizes examples of the application of best
practices on a cost/effective basis. Findings from the research are
presented in a format for railroad operating officers to decide how to
apply them to their own individual operations.
It is clear that the wheel-rail interface is the key to the heavy haul
problem. At the interface, there must be low friction to permit
moving heavy loads with little resistance. However, there must be
enough friction to provide tractive effort, braking effort, and steering
of the train. The materials must be strong enough to resist the
vertical forces introduced by very heavy loads and the dynamic
response at the wheel-rail interface introduced by vertical
accelerations of the car induced by track and wheel irregularities.
However, neither the wear rate nor the failure rate should be so great
that cost-effective heavy haul operations are threatened.
The discussion of cost/benefit analysis in Part 5 and the case
studies presented in Part 4 indicate the variety of options available to
railroad management as it seeks to achieve optimized heavy haul
operations that are cost effective. These case studies exemplify that
the process involves a systems study in which there must be
simultaneous study of the car, the wheel, the rail, and the track. The
case studies include one case of a mine-to-port railroad operating
over a dedicated line with dedicated cars and locomotives, one case
of a heavy haul railroad with heavy haul operations being only a small
fraction of total traffic on the line, and one case of the transition of a
railroad from mixed traffic to a dedicated heavy haul operation. A
matrix of best practices is presented for railroads with a wide range of
external variables in terms of axle load, track curvature, and annual
tonnage.
The solutions chosen from the options given must also be
designed to be specific to the operating conditions of the railroad. A
dedicated rail line carrying only dedicated cars and locomotives can
consider options that are not appropriate for a heavy haul operation
that moves on a line that carries other kinds of trains
These case studies and the matrix of best practices show that
there is no one perfect solution that applies to all circumstances.
Solutions applicable to the case of the dedicated mine-to-port line
with dedicated locomotives and cars are different from those to a
railroad with mixed traffic. The matrix of options solutions
emphasizes the variety of approaches that should be examined before
arriving at an optimum decision for a particular property.
o vx
Because there are many possible solutions, Parts 2 and 3 of this
handbook include summaries of current knowledge on wheels and
cars and on rails and track.
Revisions to this handbook will be considered at regular intervals
as comments for its improvement are received and as new technology
and new field information become available.
o vi x
These guidelines were prepared and edited by the Technical Review
Committee appointed by the IHHA Board of Directors.
Dr. William J. Harris, Jr. Chairman Emeritus, IHHA, USA
Dr. Harris served from 1970 to 1985 as Vice President, Research and
Test Department of the Association of American Railroads. He
served as the first Chair of IHHA.
Prof. Dr. Willem Ebersöhn (former Chair Railroad Engineering,
University of Pretoria, South Africa)Director, Engineering Services,
AMTRAK, USA
Dr. Ebersöhn established the Chair in Railway Engineering while
serving on the faculty of Engineering Department of Civil
Engineering, University of Pretoria, South Africa. He worked
extensively on heavy haul problems with Spoornet and on high-speed
track problems with Amtrak in the USA.
Dr. James Lundgren, Assistant Vice President ,Transportation
Technology Center, Inc. USA
Following service in the engineering department of CN Rail, Dr.
Lundgren joined the Association of American Railroads representing
the railroad industry at the Transportation Test Center while under
Federal Railroad Administration operation and through the
successful transition to AAR management of TTC. He has been
associated with IHHA since its inception.
Mr. Harry Tournay, Assistant General Manager of Spoornet,
South Africa
Mr. Tournay has been associated with the design and development of
improved rolling stock and locomotives of Spoornet. He has
designed rolling stock internationally and has particular expertise in
wheel/rail interface issues.
Dr. Prof. Sergey Zakharov, Head of the Division of Tribology,
at the All-Russian Railway Research Institute. Russia
Dr. Zakharov has spent his career researching the development of
diesel electric locomotives and solving diverse railroad tribology
problems. He has concentrated recently on 11 years of study of the
wheel/rail interface issues and tribology aspects of interaction
including taking a major part in organizing IHHA-99 STSConference on Wheel/Rail Interface.
o vii x
Acknowledgements
The TRC recognizes the importance of the 16 conferences and
technical sessions sponsored by organizations represented by the
Board of Directors of IHHA. Without the rich technical literature
created by the authors of the outstanding technical papers presented,
it would have been impossible to prepare and publish these
guidelines. Recognition of each author’s work is given at the
beginning of their respective chapters.
The TRC appreciates and recognizes the contributions made by
BHP of Australia, Canadian Pacific Railroad of Canada, and CVRD
of Brazil for their distribution and willingness to share data from
their experience in the case studies presented in Part 4. The authors
of the case studies are recognized in the text.
In addition to the support provided by the IHHA Board and the
reviews of the International Review Panel, the TRC wishes to
acknowledge the outstanding help of Dr. Alexander Lisitsyn, General
Director, Member of the Board, Ministry of Railways, Russian
Federation. He and his staff at the All-Russian Railway Research
Institute sponsored a very effective and special Technical Session in
June 1999, on wheel/rail interface issues and heavy haul best
practices.
The TRC recognizes the early contributions of Wardina
Oghanna in the organization of the guidelines and his services in
helping to bring together the successful meeting in Moscow. Dr.
Oghanna worked on the interpretation of this conference as a basis
for many aspects of the guidelines.
The TRC also wishes to recognize and express its appreciation
to the China Railway Society and to the China Academy of Railway
Sciences for their continued support of IHHA goals in hosting two
conferences in China, which have produced technical papers upon
which portions of these guidelines are based.
The TRC further wishes to note the special contributions made
by Michael Roney, General Manager Engineering Services and
Systems, Canadian Pacific Railroad, who met often with the TRC and
wrote parts of the handbook, as noted in the chapters.
The TRC especially appreciates the leadership of Scott Lovelace,
Chief Executive Officer of IHHA for accomplishing this difficult
mission.
o viii x
The TRC also wishes to express great appreciation to Roy Allen,
President of TTCI, Jim Lundgren, Assistant Vice President Finance
and Corporate Development, Peggy L. Herman, Manager
Documentation, and support from each of their staff members for
undertaking the publishing of the handbook in a timely and effective
manner.
Much of the handbook’s guidelines was written by members of
the Technical Review Committee. However other authors were
invited to prepare special material. Their names are cited where
appropriate.
The Technical Review Committee established an International
Review Panel of distinguished heavy haul experts to review various
drafts of the handbook. The Review Panel offered many very helpful
comments that are reflected in the final text.
Members of the International Review Panel are:
Mr. John Elkins, President, RVD Consulting, Pueblo,
Colorado, USA
Professor Conrad Esveld, Esveld Consulting, The Netherlands
Dr. Stuant Grassie, Consulting Engineer, UK
Dr. Joe Kalousec, Principal Research Officer, National
Research Council, Research Center for Surface
Transportation Technology, Canada
Mr. Eric Magel, Associate Research Officer, National Research
Council, Research Center for Surface Transportation
Technology, Canada
Dr. Steven Marich, Consulting Services, Australia
Dr. Wardina Oghanna, Director, Australian Railway Research
Institute, Australia
Professor Klaus Reisberger, University of Graz, Austria
Mr. Mike Roney, CP Rail, Canada
Dr. Yoshohiko Sato, Nippon Kikai Hosan, Japan
Dr. Kevin Sawley, Principal Investigator, TTCI, USA
Prof. Evgeny Shur, All-Russian Railway Research Institute,
Russia
Mr. Dan Stone, TTCI, USA
o ix x
Special Note:
References for each “part” are not combined at the end of the
handbook. Instead, they are listed at the end of each respective
chapter. TRC recognizes that some readers may skip around to
different chapters, reading only certain parts pertinent to them.
Therefore, the TRC thought it would be easier for each part to have
its own list of references. The TRC will welcome any suggestions for
improving the method of presentation in the next edition of the
guidelines.
o xx
Click The Highlighted Chapter Headings to View Chapter
TABLE OF CONTENTS
PART 1: INTRODUCTION AND DISCUSSION OF
GUIDELINES TO BEST PRACTICES
1.1 Discussion of Guidelines............................................. 1-1
1.2 A Systems Approach................................................... 1-1
1.3 Discussion of the Wheel and Rail Interface ................ 1-1
1.4 Example of Cost Benefit Analysis ............................... 1-9
1.5 Discussion of Part 2 on Vehicle/Track Interactions .... 1-9
1.6 Discussion of Part 3 on Wheel/Rail Interfaces.......... 1-10
1.7 Discussion of Part 4 on Four Case Studies .............. 1-10
1.8 Discussion of Part 5 on Optimizing Heavy Haul
Maintenance Practices ............................................. 1-11
PART 2: SUPPORT TECHNOLOGIES VEHICLE TRACK
INTERACTION
2.1 Vehicle Track Interaction............................................. 2-1
2.2 Railway Wheelset and Track ...................................... 2-2
2.2.1 Vertical Forces between Wheel and Rail .......... 2-5
2.2.2 Lateral Forces between Wheel and Rail ........... 2-6
2.3 Generic Railway Vehicle Suspensions ..................... 2-11
2.3.1 Vertical Suspension......................................... 2-11
2.3.2 Inter-Wheelset and Lateral Suspension .......... 2-18
2.4 Practical Heavy Haul Vehicle Suspensions .............. 2-27
2.4.1 Heavy Haul Wagon Bogies.............................. 2-27
2.4.2 Locomotive Bogies .......................................... 2-40
2.5 Rail and Wheel Profile Design .................................. 2-41
2.5.1 Basic Considerations....................................... 2-42
2.5.2 Wheel and Rail Profiles Divided into Functional
Sections ............................................................ 2-45
2.5.3 Rail and Wheel Management .......................... 2-56
2.6 Tracking Accuracy and Tolerances........................... 2-60
2.6.1 Geometric Inaccuracies in Wheelset and Track
Geometry......................................................... 2-61
2.6.2 Geometric Inaccuracies of the Wheelset and
Suspension...................................................... 2-63
References ....................................................................... 2-69
Appendix .......................................................................... 2-70
o xi x
PART 3: WHEEL/RAIL PERFORMANCE
3.1 Application of Systems Approach to Wheel/Rail
Performance Study..................................................... 3-1
3.2 Rail Contact Mechanics .............................................. 3-4
3.2.1 General .............................................................. 3-4
3.2.2 Normal Contact Stress ...................................... 3-4
3.2.3 Creep Force—Creepage Behavior .................. 3-11
3.2.4 Influence of Traction on the Load Carrying
Capacity of the Contact Area .......................... 3-15
3.2.5 Approach to Wheel and Rail Profile Stress
Optimization .................................................... 3.17
3.3 Rail and Wheel Materials .......................................... 3-18
3.3.1 Chemical Composition..................................... 3-18
3.3.2 Microstructure.................................................. 3-19
3.3.3 Mechanical Properties ..................................... 3-22
3.3.4 Wheels............................................................. 3-26
3.3.5 General Concept of Wheel/Rail
Material Selection ............................................ 3.26
3.4 Lubrication and Friction Management....................... 3-28
3.4.1 Some Tribology Considerations ...................... 3-28
3.4.2 Rail Gauge/Wheel Flange Lubrication............. 3-29
3.4.3 Friction Control and Management ................... 3-36
3.5 Rail and Wheel Damage Modes; Mechanisms and
Causes ..................................................................... 3-41
3.5.1 Wear ................................................................ 3-42
3.5.2 Recommendations to Decrease Wheel and
Rail Wear......................................................... 3-51
3.5.3 Rolling Contact Fatigue Defects ...................... 3-52
3.5.4 Head Checks ................................................... 3-55
3.5.5 Tache Ovale (Shatter Crack from Hydrogen).. 3-57
3.5.6 Squats.............................................................. 3-58
3.5.7 Rolling Contact Fatigue Defects of Wheels—
Shelling and Spalling....................................... 3-60
3.5.8 Other Rail and Wheel Defects ......................... 3-63
3.5.9 Plastic Flow...................................................... 3-69
3.5.10 Rail and Wheel Corrugations.......................... 3-73
Acknowledgements .......................................................... 3-76
References ....................................................................... 3-77
Appendix A ....................................................................... 3-84
Appendix B ....................................................................... 3-86
o xii x
PART 4(a): HEAVY HAUL CASE STUDY: Dedicated Line with
Captive Equipment, BHP Iron Ore, Australia
4.1(a) Introduction ............................................................. 4-1
4.2(a) Wheels .................................................................... 4-2
4.3(a) Modified Profiles ..................................................... 4-3
4.4(a) Material Characteristics .......................................... 4-5
4.5(a) Lubrication ............................................................. 4-7
4.6(a) Wheel Design ......................................................... 4-7
4.7(a) Bogie Characteristics.............................................. 4-7
4.8(a) Wheel Maintenance ................................................ 4-8
4.9(a) Summary ................................................................ 4-9
PART 4(b) : CASE STUDY OF WHEEL/RAIL COST
REDUCTION ON CANADIAN PACIFIC’S COAL ROUTE
4.1(b) Nature of the Business ......................................... 4-11
4.2(b) Characteristics of the Route ................................. 4-11
4.3(b) The Consist........................................................... 4-12
4.4(b) Early Problems ..................................................... 4-12
4.5(b) Initial Attempts to Control Rail and
Wheel Wear Costs ................................................ 4-13
4.6(b) Benefits of Frame-Braced Steerable Trucks ........ 4-16
4.7(b) Premium Rail Steels and Extended Wear Limits.. 4-17
4.8(b) Increased Axle Loads and AC Traction ................ 4-19
4.9(b) Further Cost Savings ............................................ 4-20
4.10(b) The Size of the Prize .......................................... 4-22
PART 4(c) Wheel And Rail Performance at Carajás Railway
4.1(c) Introduction to Carajás Railway ............................ 4-25
4.2(c) Historical Data....................................................... 4-25
4.2.1(c) Wheels............................................................ 4-25
4.2.2(c) History............................................................. 4-26
4.2.2.1(c) During 1986 .............................................. 4-26
4.2.2.2(c) During 1987 .............................................. 4-26
4.2.2.3(c) During 1988 .............................................. 4-28
4.2.2.4(c) During 1989 .............................................. 4-29
4.2.2.5(c) During 1990 .............................................. 4-29
4.2.2.6(c) During 1992 .............................................. 4-29
4.2.2.7(c) During 1993 .............................................. 4-30
4.3(c) Improvements ....................................................... 4-30
4.3.1(c) Wheels Management Model........................... 4-30
4.4(c) Rails ...................................................................... 4.34
4.4.1(c) History............................................................. 4.35
4.4.1.1(c) 1987............................................................. 4-35
4.4.1.2(c) From 1987 to 1999 ...................................... 4-35
4.4.1.3(c) 1988............................................................. 4-35
o xiii x
4.4.1.4(c) 1990............................................................. 4-35
4.4.1.5(c) 1991............................................................. 4-35
4.4.1.6(c) From 1993 to 1996 ....................................... 4-36
4.4.1.7(c) 1997.............................................................. 4-36
4.5(c) Looking for a Solution ........................................... 4-36
4.5.1 Introduction ......................................................... 4-36
4.6(c) Methodology and Approach of TTCI's
Comprehensive Program On Carajás Railway..... 4-39
4.7(c) Implementation of TTCI's Wheel/Rail Life
Optimization Program on Carajás Railway........... 4-40
4.7.1(c) Ore Wagon Truck Performance
Evaluation and Modeling .................................. 4-40
4.7.2(c) Full-scale Testing of Standard and
Frame Braced Trucks with Load Measuring
Wheelsets.......................................................... 4-46
4.8(c) Methodology of Recommendations
for Rail Grinding Practices .................................... 4-48
4.8.1(c) Lubrication Practice ........................................ 4-52
4.8.2(c) Implementation of TRACS and
the Wheel Life-Cycle Costing Model ................. 4-54
4.9(c) Conclusions .......................................................... 4-55
References ....................................................................... 4-58
PART 4(d) Quick Reference Tables for Basic Heavy Haul
Rail System Design
4.1(d) Introduction ........................................................... 4-59
4.2(d) Using the Design Tables....................................... 4-61
4.3(d) References ........................................................... 4-61
4.4(d) General Notes....................................................... 4-62
4.5(d) Table Notes: (brief descriptions of salient
features of component classes)............................ 4-63
PART 5: Maintaining Optimal Wheel and Rail Performance
5.1 Maintaining Optimal Wheel and Rail Performance ..... 5-1
5.2 Rail Structural Deterioration........................................ 5-6
5.2.1 Management of Rail Testing to Control
Risk of Rail Fracture ................................................ 5-6
5.2.2 The Framework for Risk Management ................. 5-7
5.2.3 Defect Occurrence Rates ................................... 5-10
5.2.4 Critical Defect Sizes............................................ 5-13
5.2.5 Rail Fatigue Projection........................................ 5-15
5.2.5.1 Use of Weibull Distribution to Predict Rail Flaw
Occurrence Rates......................................... 5-16
5.2.6 Modes of Rail Testing ......................................... 5-21
5.2.6.1 Rail Testing Equipment ................................ 5-24
o xiv x
5.2.7 Ultrasonic Principles ........................................... 5-25
5.2.8 Inspection Effectiveness ..................................... 5-27
5.2.8.1 Test Probes .................................................. 5-27
5.2.8.2 Signal Processing......................................... 5-29
5.2.8.3 Displaying Indications to the Operator ......... 5-30
5.2.8.4 Operator Vigilance........................................ 5-30
5.2.8.5 Estimates of Rail Testing Reliability ............. 5-31
5.2.9 Selecting Rail Testing Intervals .......................... 5-34
5.2.9.1 Performance-Based Adjustment of
Test Intervals ................................................ 5-37
5.2.9.2 A Parametric Approach ................................ 5-38
5.2.9.3 Cluster Testing ............................................. 5-39
5.2.9.4 Special Care in Special Track Work............. 5-40
5.2.9.5 Rail Testing Intervals ¾
Canadian Pacific Approach.......................... 5-41
5.2.10 Induction Measuring Principles......................... 5-43
5.2.11 Conclusion ........................................................ 5-45
5.3 Rail Wear Measurements ......................................... 5-45
5.3.1 Rail Wear Measurement Techniques ................. 5-45
5.3.2 Rail Wear Projection ........................................... 5-52
5.4 Rail Profile Maintenance Practices ........................... 5-54
5.4.1 Rail Grinding ....................................................... 5-54
5.4.1.1 Objectives of Rail Grinding ........................... 5-54
5.4.1.1.1 Longitudinal Rail Profile Corrections ...... 5-55
5.4.1.1.2 Transverse Rail Profile Correction ......... 5-57
5.4.1.1.3 Effects of Rail Shape Parameters on
Rail Damage .......................................... 5-60
5.4.1.1.4 Grinding for Surface Condition............... 5-63
5.4.1.2 Grinding Stones and their Effects................. 5-64
5.4.1.2.1 Abrasive Stone Technology ................... 5-64
5.4.1.2.2 Surface Finish ........................................ 5-66
5.4.1.2.3 Effects of Speed and Pressure .............. 5-68
5.4.1.3 Grinding Patterns and their Use ................... 5-70
5.4.1.4 North American Grinding Practice................ 5-76
5.4.1.5 Optimizing Rail Profiles ................................ 5-78
5.4.1.5.1 Rail Profile Design.................................. 5-80
5.4.1.5.2 Rail Stresses and Pummeling ................ 5-81
5.4.1.5.3 Tangent Track ........................................ 5-84
5.4.1.5.4 High Rail Profiles.................................... 5-84
5.4.1.5.5 Low Rail Profiles..................................... 5-85
5.4.1.6 Lubrication and Grinding .............................. 5-86
5.4.1.7 Optimal Wear Rate ....................................... 5-86
5.4.1.8 Rail Grinding Strategies................................ 5-88
o xv x
5.4.1.8.1 Preventive Rail Grinding......................... 5-82
5.4.1.8.2 Preventive vs. Corrective Rail Grinding . 5-89
5.4.1.9 Transitioning from Corrective to
Preventive Grinding ...................................... 5-90
5.4.1.9.1 Preventive Gradual Grinding............... 5-91
5.4.1.9.2 Results ................................................ 5-92
5.4.1.9.3 Rail Wear ............................................ 5-93
5.4.1.9.4 Rail Surface Condition ........................ 5-94
5.4.1.9.5 Detail Fracture Rates .......................... 5-95
5.4.1.10 Advance Planning to Increase
Grinding Production ................................ 5-95
5.4.1.11 Maintaining Quality Control ..................... 5-96
5.4.1.11.1 Grinding Power ................................. 5-96
5.4.1.11.2 Ground Rail Profile............................ 5-96
5.4.1.11.3 Longitudinal Rail Profile .................... 5-98
5.4.1.11.4 Transverse Rail Profile...................... 5-98
5.4.1.11.5 Metal Removal .................................. 5-99
5.4.2 Rail Planing......................................................... 5-99
5.4.2.1 Description of the SBM 140 Rail
Planing Machine ......................................... 5-100
5.4.3 The Planing Process......................................... 5-101
5.5 Wheelset Failure Risk Management and
Maintenance............................................................ 5-103
5.5.1 Wheelset Reliability (Spoornet) ........................ 5-105
5.5.1.1 New Components ....................................... 5-105
5.5.1.2 Used Components...................................... 5-106
5.5.1.3 Condition Monitoring................................... 5-107
5.5.1.3.1 Wayside Condition Monitoring ............. 5-107
5.5.1.3.2 Run-in Inspections................................ 5-108
5.5.1.3.3 Four Monthly Maintenance Depot
Inspections........................................... 5-109
5.5.1.3.4 Workshop Maintenance........................ 5-109
5.5.1.4 Wheel Profile Monitoring ............................ 5-109
5.5.2 Wheelset Maintenance ..................................... 5-110
5.6 Wheel and Vehicle Interaction Condition Measures5-112
5.6.1 Wheel Wear Measurement Techniques ........... 5-113
5.6.2 Wheel and Vehicle Track Interaction Wayside Measuring System............................. 5-120
5.6.2.1 Weighing in Motion and Wheel Impact
Measurement (WIM-WIM) .......................... 5-121
5.6.2.2 The Low-Speed Weigh Bridges.................. 5-122
5.6.2.3 Hot-Box, Hot and Cold Brake Detectors..... 5-123
5.6.2.4 The Acoustic Defective Bearing Detection . 5-124
o xvi x
5.6.3 Design Considerations...................................... 5-124
5.7 Practical Application of Wayside Lubricators .......... 5-125
5.7.1 Friction Management ........................................ 5-126
5.7.2 Benefits of Effective Rail Lubrication ................ 5-127
5.7.3 Wayside Lubrication Capabilities and Operation................................ 5-128
5.7.4 Selecting the Most Appropriate Equipment for
Dispensing the Lubricant .................................. 5-128
5.7.5 Selecting the Optimal Type of Grease for the
Particular Operating Environment..................... 5-130
5.7.6 Measurement and Management of Lubrication
Effectiveness..................................................... 5-132
5.7.7 Positioning of Lubricators ................................. 5-136
5.7.8 Lubricator Placement Model ............................. 5-138
5.7.8.1 Track Related Factors ................................ 5-138
5.7.8.2 Traffic Related Factors ............................... 5-139
5.7.9 Case Study: Lubrication - Richards Bay Line,
South Africa ...................................................... 5-140
5.8 Optimizing Wheel and Rail Life............................... 5-143
5.8.1 Rail Optimization............................................... 5-144
5.8.1.1 The Rail Management Decision Zones ...... 5-144
5.8.1.2 Controlling Rail Wear (Maximum Rail
Wear Limits)............................................... 5-148
5.8.1.3 Rail Use Strategy........................................ 5-154
5.8.1.4 Lubrication and Curve Elevation Monitoring5-157
5.8.1.5 Transposition .............................................. 5-158
5.8.2 Wheel Optimization........................................... 5-159
5.8.3 Friction Management (Interface Optimization) . 5-163
5.8.4 A System Approach for Managing the
Wheel/Rail Interface ......................................... 5-164
5.9 Conclusion............................................................... 5-165
Acknowledgements ........................................................... 5-166
References ........................................................................ 5-166
GLOSSARY.......................................................................... G-1
METRIC CONVERSION CHART
o xvii x
Click Here To Go Back To Table of Contents
Part 1: Introduction and Discussion
of Guidelines to Best Practices
Written by Dr. William J. Harris, Chair, Technical Review
Committee (TRC), Mr. Harry Tournay, IHHA Board of
Directors and TRC member, Dr. Willem Ebersöhn, TRC
member, Dr. Sergey Zakharov, TRC member, and Dr.
James Lundgren, TRC member
1.1 Discussion of the Guidelines
These guidelines offer insights into ways to optimize a heavy
haul railway operation. In Section 1.2, there is a description of
the importance of addressing the process using a systems
approach. Section 1.3 is an extended review of the wheel/rail
interface, and Section 1.4 is an account of a cost/benefit
analysis. The introduction concludes with Sections 1.5 to 1.8,
which contain brief comments on each of the succeeding four
parts of the handbook.
1.2 A Systems Approach
The guidelines emphasize that it is no longer adequate to
change one part of the railway system without examining its
impact on the other parts of the system. Increasing car weight
can have a profound effect on track and bridges. Changing rail
properties can lead to unexpected wheel behavior. Therefore,
the balance of the guidelines in this handbook will emphasize
the interactions of components and the importance of dealing
with the wheel-rail problem as a system.
A system approach to the design and maintenance of
wheel and rail interface, in the form of best practices, can be
expected to result in minimization of rail gauge face and wheel
flange wear, avoidance of detrimental wheel and rail defects,
stable vehicle performance, including safety issues, and
minimization of noise generation.
1.3 Discussion of the Wheel and Rail Interface
The wheel and rail interface is the key to the heavy haul
problem. At that interface, there must be low friction to
permit moving heavy loads with little resistance. However,
there must be enough friction to provide tractive effort,
o 1-1 x
braking effort, and steering of the train. The materials must be
strong enough to resist the vertical forces introduced by very
heavy loads and the dynamic response at the wheel-rail
interface introduced by vertical and lateral accelerations of the
car induced by track and wheel irregularities. However, neither
the wear rate nor the brittle failure rate should be so great that
cost-effective heavy haul operations are threatened.
The remarkable ability of a steel wheel rolling on a steel
rail to carry a very heavy load seemed almost a miracle 175
years ago, when railroads first began to operate. Of course at
that time loads were low compared to those of today. The
increase in axle loads has been gradual but steady for decades.
Over 50 years ago, the rate of increase changed. Suddenly
it was necessary to improve subgrade and ballast as well as ties
and tie fasteners. Suddenly it was necessary to increase rail
hardness and improve rail quality and to introduce headhardened rail in some cases. Suddenly it was essential to
improve wheels and car suspension systems. Suddenly it was
necessary to increase inspection capabilities and frequencies to
reduce accidents and improve service.
These requirements for improved materials, designs, and
practices were based on field experience, when necessary, and
on research, when available. Whatever the nature of the
problems, continued attention to rail and wheel technology has
provided the basis for continued increases in axle loads. These
guidelines suggest options to improve the initial components
and systems as well as practices to ensure through maintenance
the continued effectiveness of the wheel/rail system to carry
increasingly heavier loads.
The technology of a steel railway wheel rolling on the rail
is eminently suitable for heavy haul, heavy axle load operations.
The unique properties of steel-on-steel contact results in
minimal deformation of both contacting bodies under load.
This results in rolling contact with minimum energy losses in
friction across the contact patch and in minimum damping
within the material of the contacting bodies. This is why the
rolling resistance associated with railroads is so low and
permits the transport of vast tonnage.
o 1-2 x
The contact patch is surprisingly small with
correspondingly high-contact stresses. Typically, contact is
made over a quasi-elliptical contact patch the size of a small
coin of 13-millimeter (½-inch) diameter (see Figure 1.1). This
implies that a 20,000-tonne train is supported over an area
equivalent to the surface of a kitchen table (1.3 m x 1.3 m or
4½ ft x 4½ ft)!
CONTACT
PATCH
CENTRALLY PLACED
sZ
sY
sX
sY
sX
sZ
sz 1400 Mpa
sX sY 800 Mpa
Yield Stress is 600 – 800
Mpa
Figure 1.1. Contact between Wheel and Rail:
Wheel Centrally Placed on the Track
Immediately beneath the contact patch in either the wheel
or rail, the steel is under tremendous pressure from all
directions as the contact pressure is “supported” by reaction
pressures from the surrounding material of either wheel or rail.
This is depicted in Figure 1.1 by the arrows converging on an
element of steel beneath the contact patch. This is termed a triaxial state of stress. Each of the “stress arrows” as shown
presses almost equally on the steel, which has no direction in
which to move or “flow” and can withstand the load. Under
these conditions, and using high strength steels, axle loads
beyond present day applications (up to perhaps 60 t or beyond)
should be possible.
The reasons that railroads have not reached these loads,
o 1-3 x
and why some railroads have trouble with prevailing axle loads,
are that these ideal-contact conditions, described above, are
not always achieved, because of the following:
•
Track and vehicle conditions can result in dynamic
loads, which are well in excess of the static and often
result in impact between wheel and rail.
•
The contact patch can be severely reduced under some
uncontrolled wheel/rail contact conditions.
•
The delicate balance of the tri-axial state of stress can
become upset by:
§ Frictional forces acting across the contact patch or
contact occurring on the edge of either wheel or rail.
§ Two-point contact occurring with gross relative
slippage over one or both contact patches with
associated accelerated wear.
Figures 1-2 through 1.6 are typical examples of adverse
wheel/rail contact conditions. Brief comments are given before
each figure.
Figure 1.2: Dynamic impact loading caused by wheel flats, rail
joints, soft rail welds, rail corrugations, and discontinuities at
switches.
Wheel
Skid
IntermediateFrequency
Impacts
RAIL JOINT
Dynamic
Load
Static Load
o 1-4 x
Figure 1.2
o 1-5 x
Figure 1.3: Intense single point contact between flange throat
and the gauge corner of the rail, which results in head-checking
and shelling.
CONTACT
PATCH
FLANGE CONTACT
SHELLING
sZ
sY
HEAD CHECKING
sX
sY
sX
sZ
Figure 1.3
o 1-6 x
Figure 1.4: Intense convex contact between the rail crown and
wheel, which can result in material flow on the field side,
shelling of the rail crown and/or wheel tread. This is
exacerbated if contact is made toward the outer edges of the
rail and wheel where there is no material to “support” the
element under the contact patch. The favorable state of stress
is “upset” and material flow occurs.
CENTRALLY PLACED
dX
CONVEX SHAPE
dZ
Figure 1.4
o 1-7 x
dY
CONTACT
PATCH
Figure 1.5: Lateral slipping between wheel and rail in curves is
a result of badly tracking vehicles. The forces tangential to the
contact patch as a result of slippage or micri-slip cause
deformation of the elements of steel under the contact patch.
This “upsets” the supporting pressures/stresses on the element
resulting in material flow and can result in intermittent crown
wear and deformation experienced as corrugations or general
material flow.
FIELD SIDE CONTACT
WORN CONDITIONS
CONTACT
PATCH 19.5
7
sZ
sX
sY
sX
Figure 1.5
o 1-8 x
sY
sZ
Figure 1.6: Inappropriate control of the contacting shapes can
limit the size and shape of the contact patch causing intense
stresses, material flow, and fatigue. Defects in rail or wheel
material in the region of the intense contact stresses exacerbate
the problem.
CONTACT
19.5
PATCH
FIELD SIDE CONTACT
WORN CONDITIONS
7
sZ
sY
sX
sX
sY
sZ
Figure 1.6
The railroad that can minimize the above mentioned issues
is the railroad that will be capable of increasing axle load or
reducing maintenance in relationship with huge advantages
over its competition.
o 1-9 x
1.4 Example of Cost Benefit Analysis
In addition to the technical issues that characterise the
wheel/rail system, there are very importance economic issues.
It is essential to take into account a cost/benefit analysis in the
course of making technical decisions. These are illustrated in
the following remarks.
Rail is the single most expensive element of the track
structure. On many railways, it is behind only labor and fuel as
an expense item. The tonnage carried by a rail before it is
condemned can range from less than 100 million gross ton to
close to 2.5 gigga gross ton.
As an example of the value of rail maintenance management,
assume that a single kilometre of rail costs $180,000 to install. Track
engineers decide that the rail has a badly fatigued surface and has
reached the end of its service life. They call for it to be replaced,
gaining a salvage value of $18,000.
But now assume that instead of replacing the rail, they did some
corrective rail grinding costing $1800 and left the rail in track. The
railway then invested the $180,000 – $18,000 - $1,800 = $160,200 in
the construction of a new customer facility at a rate of return of 20%.
This earned $160,000 * 20% = $32,000 in its first year.
The next year, the track engineers see that their rail is
approaching allowable wear limits and schedules a rail replacement,
now costing $187,200 due to cost escalation of 4%. But they have
made for the railway $32,000 – ($187,200 – 180,000) = $24,800 by
deferring replacement of rail in that kilometre, without consequence,
for an extra year. And that is why they collect a salary.
There is significant money to be made by deferring rail
replacement as much as possible without incurring risk.
Certainly it is a major responsibility of the track engineer to
ensure that he gets the most out of his rail, and rail profile
maintenance and rail testing are his most important tools to do
this.
1.5 Discussion of Part 2 on Vehicle/Track
Interactions
Part 2 of the guidelines discusses the vehicle/track interactions.
These are the components that are given significant attention
o 1-10 x
when track is laid out and freight cars and locomotives are
purchased. The contribution of the suspension systems and
other design features of the cars and the subgrade, ballast, ties,
and tie fasteners to the operations of the vehicle/track system
are discussion in Part 2 and options are described that can help
in optimizing the vehicle/track system.
1.6 Discussion of Part 3 on Wheell/Rail Interfaces
Part 3 of the guidelines addresses the wheel/rail interface
issues. It gives an overview of rail contact mechanics, wheel
and rail materials characteristics, lubrication and friction
management practices, and damage modes and their
mechanisms. It discusses the contribution of the research to
the behavior of the wheel and the rail. Through the
explanation of mechanisms, processes, and causes of damages,
it gives the justification of suggested solutions. Part 3 offers
recommendations that can be adopted by operating personnel
and an input to an optimized system.
1.7 Discussion of Part 4 on Four Case Studies
Part 4 of the guidelines offers the reader several case studies.
The first of these is based one of the great heavy haul railway
success stories, that of the experience of BHP, Australia, and
describes ways that BHP optimized heavy haul operations on a
mine to port railroad over a dedicated line with dedicated cars
and dedicated track. Under this set of circumstances, it has
been possible to fine tune the system. Since every car is the
same, or its differences are fully understood, and the track is
the same, it is possible to monitor the behavior of the system
and improve decisions for replacements as well as to achieve
optimum maintenance practices.
The second case is based on the experience of the
Canadian Pacific Railroad in which about 10 percent of the
total traffic on a specific line is heavy haul traffic and the
balance is mixed freight. This case study shows that it is
possible to achieve partial optimization by choices made
regarding the heavy haul segment of the traffic, but not to gain
the full advantages that can be gained in a dedicated railroad in
which the entire system is under the control of the heavy haul
operators.
o 1-11 x
The third case is derived from the experience of the
Companhia Vale do Rio Doce (CVRD) line in Brazil. This line
started as a mixed freight line and was gradually transformed
into what it is today, a dedicated heavy haul line. It did not start
by making the kinds of decisions that were made at BHP to
reflect the heavy haul operations in the lay out of the track.
However, the CVRD experience shows that it is possible to
make significant progress toward optimization as steps are
taken to utilize improved practices in the course of the
transition to a dedicated heavy haul railway.
The final case study presented in Part 4 describes a matrix
of options for optimized heavy haul solutions. The matrix
presents information on suggested changes in practice as
conditions change. It describes the changes appropriate for a
line with greater curvature. It discusses the options to be
considered as annual tonnage increases. It discusses the impact
of changes in axle load on rail and wheel and vehicle and track
options. This matrix of examples offers guidance as to the
direction in which changes should be considered as changes
occur in traffic and in terrain.
From study of these cases, it becomes clear that there is no
single “best practice.” There are improved practices that can be
adapted to the special circumstances of a given route, a given
traffic density, a given axle load, and other circumstances
applicable to a given railway operation. That is the reason that
the TRC has attempted to provide Part 2 of these guidelines
with insights regarding the vehicle and the track and Part 3
with insights into the wheel/rail interface for use by the
managers of a given railroad operating under specified
circumstances. That is the reason for including Part 5 with its
emphasis on maintenance.
1.8 Discussion of Part 5 on Optimizing Heavy Haul
Maintenance Practices
Part 5 addresses the issue of maintaining vehicles, track, wheels
and rail. After making sound initial decision, it is essential to
establish a maintenance procedure that is based on effective
measurement of deterioration and a set of processes to restore
o 1-12 x
wheels and rail as well as equipment and track to their desired
conditions. These inspection and maintenance practices must
also provide a basis for identifying and removing seriously
flawed components. Thus, the combination of acquisition
decisions based on an understanding of the wheel/rail interface
and vehicle track interactions with the design of a
comprehensive maintenance program can achieve the desired
optimum heavy haul operations in systems around the world.
o 1-13 x
Click Here To Go Back To Table of Contents
PART 2: SUPPORTING TECHNOLOGIES
VEHICLE TRACK INTERACTION
Written by Mr. Harry Tournay, IHHA Board of Directors and
Technical Review Committee (TRC) member
2.1 Vehicle Track Interaction
Railway vehicles form a subset of terrestrial vehicles that are
supported and receive lateral guidance from track structure.
Road vehicles form another subset where road or terrain
supports them. Drivers operate road vehicles by guiding the
steering wheel or related mechanism. This action alters the
rolling orientation of certain wheels on the vehicle, thus
changing the direction of travel.
The rail-bound vehicle reacts to the topology of the track
to follow the pre-determined path defined by the track. The
crown of the rail not only provides vertical support but also
lateral guidance of the wheels of the vehicle. The efficient
interaction between vehicle and track can support very heavy
axle loads. On the other hand, inappropriate design and
maintenance of the vehicle/track interface can lead to rapid
degradation of components within the system and can
jeopardize the profitability of the rail operation concerned.
The objective of this chapter is to describe the important
force mechanisms acting between the rail and wheel and the
influence of vehicle design on these mechanisms. Typical
symptoms of inappropriate interaction will be described so that
the reader will recognize them and be able to take corrective
action. Appropriate vehicle suspension configurations are
described together with their critical characteristics for optimal
operation. Optimal wheel and rail profiles are proposed for
world’s best practice. The influence of vehicle and track
accuracy on tracking performance is discussed.
Issues relating to vehicle/track interaction in this chapter
are described in a qualitative sense and refer to what are
considered the driving interaction mechanisms. References are
made to sources that give a more rigorous description of the
interaction mechanisms.
o 2-1 x
2.2 Railway Wheelset and Track
The railway wheelset is traditionally comprised of two steel
wheels that are fixed rigidly to a common axle (see Figure 2.1).
Wheelsets with independently rotating wheels are being used to
a limited degree on certain passenger rail vehicles, but not in
heavy haul applications. The rolling surfaces of the wheels; i.e.,
the wheel profiles, are cut to a cone angle γ. Nowadays, more
complex profiles termed “hollow,” “worn,” or “profiled”
treads are used. These have an “effective conicity” of γ, as
defined in the appendix.
ro
γ
2l
2b
Figure 2.1: Railway Wheelset
The track comprises two rails laid on sleepers at a
particular gauge, as Figure 2.2 shows. The rails are laid at an
angle, β, to the sleeper to generally match the angle, γ, of the
wheelset profile. This assists in stabilizing the rail against
rollover as the normal reaction to the contact with the wheel
passes through the foot of the rail.
When concrete sleepers are used, the connection between
the rail and the sleeper is generally made with a rail chair and a
resilient pad, which are inserted between the rail and sleeper to
attenuate high-frequency vibrations and to protect the sleeper.
Spring clips are used to fasten the rail to the sleeper. Timber
sleepers, on the other hand, give an additional degree of
inherent resilience.
o 2-2 x
Gauge
β
Rail
Pad
Rail
Chair
Figure 2.2: Track Gauge
A layer of ballast supports the sleeper. The ballast permits
alignment adjustment, as well as vertical, lateral, and
longitudinal stabilization of the track. It further provides some
vertical resilience to passing trains. The structure of the ballast
also provides protection to the track substructure by spreading
the load and by dissipating vibration energy. The voids in the
ballast permit drainage and a degree of accumulation of fine
material, without any significant change to the alignment or
resilience of the track.
Railway track is generally “banked” or superelevated in
curves to counter centripetal forces without appreciably
transferring wheel loads between outer and inner rails as the
vehicle negotiates the curves at a higher speed. In the limit,
superelevation helps prevent overturning of the vehicle.
However, inappropriate matching of superelevation to vehicle
speed can adversely influence the curving performance of the
vehicle and, in turn, the wear and stresses in rail and wheel.
The portion of track between tangent and curved track is
termed a transition curve and the vehicle experiences this as
track twisted about a longitudinal axis. This implies that the
contact patches on the different wheels may not be in the same
plane. This would lead to a loss of normal load between some
wheels and the rail if inappropriate suspension designs are
o 2-3 x
used. Furthermore, the running surface of the rail is
discontinuous at non-welded rail joints and certain types of
crossings at switches. This can cause impact loads on the wheel
and the rail and, momentarily, result in a shift in the wheel/rail
contact position on the wheel profile. This may affect wheelset
guidance.
Steel-on-steel contact produces a uniquely low rolling
resistance for railway vehicles. The geometry of the wheelset,
described mathematically as a di-cone (two cones placed backto-back having a cone angle, 2γ), imparts on the wheelset
unique properties of self-guidance; i.e., self-centering on
tangent track as Figure 2.3 shows, and the ability to negotiate
curves as Figure 2.4. shows. Hence, the railway wheelset also
has the ability to accommodate diameter inaccuracies between
the two wheels by displacing laterally on tangent track to
compensate the diameter difference. These properties result
from the rolling radius differential generated between the
wheels on the common axle. The flange is used where track
discontinuities or track geometry is so severe or the vehicle
suspension is so inadequate that the properties of self-guidance
of the wheelset are insufficient for guidance without flange
contact.
Figure 2.3: Self-Centering Motion on Tangent Track
o 2-4 x
Rc
y
cL Wheelset
cL Track
Figure 2.4: Force Equilibrium in a Curve
2.2.1 Vertical Forces between Wheel and Rail
A minimum vertical force between the wheel and the rail is
required to generate the guidance forces described above.
Failure to provide sufficient vertical wheel load can result in
derailment. A derailment is the first and most disastrous
indication of inadequate vehicle/track interaction. Minimum
values for vertical load are given, typically, in the research
results of the European Rail Research Institute.1
The maximum allowable ratio between the lateral and
vertical forces of a single wheel (known as the Y/Q force ratio)
is used as a measure of the proneness to flange climb
derailment. This ratio was originally suggested by Nadal.2 As
Nadal’s criterion is generally quite conservative, especially for
small or negative angles of attack, Weinstock defined a more
realistic criterion based on the axle sum of the Y/Q values.6
Before this limit is reached, however, either vertical track
alignment far exceeds acceptable norms or flange contact is
excessive. Indeed, flange contact is often a sign of inadequate
guidance and a source of wear and energy loss and should be
addressed.
The vehicle load, its speed, the vehicle suspension
characteristics, and the track topology determine the vertical
load over the contact interface. These are reflected in the load
o 2-5 x
on the journal bearings and in turn the load across the contact
patch.
2.2.2. Lateral Forces between Wheel and Rail
In this section a variety of forces that act in the horizontal
plane are described. The emphasis is on creep and flange
forces.
2.2.2.1 Creep Forces
The most efficient means of vehicle or wheelset guidance is by
means of creep forces. Creep forces are the forces that are
generated by the rolling of the railway wheelset, as a di-cone,
on the track as Figures 2.3 and 2.4 show. Under these
conditions, creep is produced as a result of a combination of
adhesion and micro-slip across the rail/wheel contact interface.
A more rigorous explanation is given in Part 3 and Reference
3. These creep forces are only generated when the wheelset
deviates from a pure rolling position defined by its kinematic
motion and must be reacted by forces generated at the journal
bearings. Longitudinal and lateral creep forces are explained in
more detail in the next two subsections.
2.2.2.1.1 Longitudinal Creepage
Consider a wheelset deflected laterally from a pure rolling
position by a distance y. This is referred to as the “Initial State”
in Figures 2.5 and 2.6. On straight track (Figure 2.5), the pure
rolling position is the centerline of the track. On curved track
(Figure 2.6), the pure rolling position is a position towards the
outside of the curve from the centerline where the radius
differential between the wheels allows the wheelset to
kinematically roll through the curve. On rolling forward with a
velocity, v, the deflected wheelset will want to roll to a
“Preferred State” as indicated by the chain-dotted outline of
the wheelset shown in both figures.
If the wheelset is constrained to remain in a similar attitude
to the track, as it was in the “Initial State,” creepage takes place
as the wheels roll. In the case illustrated, the outer wheels of
larger diameter slip back relative to the forward velocity of the
wheelset with the smaller diameter wheels slipping forward.
Slip forces, Fs are generated on the wheelset, which react
o 2-6 x
against the constraining forces at the journals. Forces opposite
to Fs are acting on the rail. The amount of creepage and the
creep force generated are directly proportional to the
displacement y and the cone angle γ. The constant of
proportionality for creepage is dependent, inter alia, on axle
load and contact geometry. The creepage mechanism within
the contact patch is described more fully in Part 3. The above
description is of a quasi-static form for the sake of simplicity.
Remember that a similar model may be drawn in a dynamic
sense with the inertia of the wheelset and suspension design
adding to the constraining forces.
The effects of excessive longitudinal creepage, combined
with high-contact stresses, is often seen in material flow of the
rail producing head checks and subsequent shelling (Figure
2.7), or flow of flash butt material on the gauge corner of the
rail (Figure 2.8).
FS
FJ
Actual Final State
FJ
FS
Preferred Final State
V
y
Initial State
cL Track
Figure 2.5: Longitudinal Creep on Tangent Track
o 2-7 x
FS
FJ
FJ
V
FS
Actual Final State
Preferred Final State
y
Initial Final State
cL Track
Pure Rolling
Position
Figure 2.6: Longitudinal Creep on Curved Track
Head Checks
Shelling
Figure 2.7: Head checks and Subsequent Shelling
Flow of
Flash-butt
Weld
Figure 2.8: Material Flow in Heat Affected Zone
o 2-8 x
2.2.2.1.2 Lateral Creepage
Lateral creepage may be described in a similar manner.
Consider a wheelset placed at an angle of yaw, á, on the track
as Figure 2.9 shows. On rolling forward, the preferred final
state of the wheelset is shown as chain-dotted. If the wheel is
constrained by the vehicle suspension or a flange force to be
oriented to the track in a similar position to the initial state, the
wheelset must have slipped laterally. This lateral creepage and
the associated force are proportional to the angle, á. The
constant of proportionality is dependent, inter alia, on axle load
and contact geometry. The creepage mechanism within the
contact patch is more fully described in Part 3.
High lateral creepage is reflected in lateral material flows in
the rail crown in sharp curves or at large lateral track
discontinuities as well as material flows on the wheel as shown
in Figure 2.10. This is also associated with high flange forces as
described below.
FS
FL
Final State
FS
V
Preferred Final State
Initial State
α
cL Track
Figure 2.9: Lateral Creepage on Tangent Track
o 2-9 x
Forces
Material Flows
Lateral Creep
Lateral Creep
Flange
Force
High Contact
Stresses
Flange &
Rail Wear
Figure 2.10: Worn Rails in a Curve
2.2.2.2 Flange Forces
When steering cannot be achieved by means of creep forces,
flange contact is made and a lateral flange force acts to keep
the wheelset from derailing. Flange contact is often made with
an angle of attack, á, implying the presence of lateral creepage.
A lateral force model of the wheelset under flange contact and
with lateral creepage is shown in Figure 2.10. This figure
describes typical reasons for the shape of worn rail in curves.
Associated with the flange force is a frictional component
that can lead to a reduction in load over the contact patch and
result in wheel climb and subsequent derailment. The action of
this force is included in the theory of Nadal mentioned earlier.
2.2.2.3 Other Forces
Other forces, like spin creep and the gravitational forces, do
act on the railway wheelset but are often of lesser magnitude
than those described above and do not necessarily play a
significant role. They are described in Reference 3 and other
references on vehicle dynamics.
Spin creep occurs when different parts of the contact roll
on different radii relative to the axis of the axle. Hence, a
rotational “scrubbing” action occurs at high contact angles.
This has been associated, together with high contact stresses,
with the formation of head checks. The couple associated with
spin is considered to have a minimal influence on rail/wheel
forces.
A gravitational force is generated on the wheelset when the
lateral components of the normal reaction to the contact patch
are unequal. This force occurs when the wheelset is deflected
laterally and non-conical or profiled wheels are used.
o 2-10 x
2.3 Generic Railway Vehicle Suspensions
The suspension of a bogie can be divided into the in-plane
lateral and longitudinal suspension that dictates the tracking
and curving performance, and the vertical suspension that
carries the load and has an effect on the vertical wheel rail
forces. Any practical railway vehicle requires at least two
wheelsets. The manner in which these wheelsets are coupled to
the vehicle has a significant influence on vehicle cost, the
performance of the rail and the wheel, and the guidance and
dynamics of the vehicle. Although there is a strong dynamic
coupling between vertical and lateral dynamics, vehicle
suspensions are generally discussed separately in terms of their
vertical and horizontal; i.e., lateral and longitudinal,
suspensions. Vehicle dynamics cannot be discussed without
considering the properties of the track. Hence, track geometry
and track stiffness is included in this generic discussion of
railway suspension systems.
2.3.1 Vertical Suspension
The purpose of the vertical suspension is threefold. These
generic purposes are discussed below.
2.3.1.1 Attenuation of Vertical Vehicle Vibrations
The vehicle, when moving forward on the track, experiences
vibrations of varying frequencies which excite the various
modes of the vehicle structure, body and the payload. The
dynamic modes are generally in bounce, roll, pitch, nosing, and
sway. Some of the exciting mechanisms are:
•
Long wavelength track alignment irregularities in the
vertical profile and track twist. These irregularities
typically result in vehicle input frequencies between 0.5
and 30 Hz.
•
Long wavelength track stiffness variations are also
present and activate the vehicle in similar modes and
frequencies as the alignment irregularities.
•
Short wavelength, consistent stiffness variations
associated with local sleeper support, results in vehicle
input frequencies up to 40 Hz.
•
High frequency impacts at rail discontinuities (P1
o 2-11 x
forces) often excite the vehicle body vertical modes to
induce the so-called P2 lower frequency reaction
forces.
2.3.1.2 Equalization of Wheel Loads by the Vehicle
Suspension
As the vehicle is supported on a minimum of four contact
patches on perturbed or twisted track, it is generically a
statically indeterminate structure similar to a four-legged table
on an uneven floor. As Section 2.2.1 states, sufficient vertical
load is required across the contact interface for effective
guidance. The vertical suspension stiffness must thus prevent
unacceptable wheel unloading on twisted track.
2.3.1.3 Attenuation of Vertical Vibrations to the Track
Structure
As a result of vertical vehicle dynamics, dynamic loading is
transmitted from the vehicle through the wheel into the track
super and substructure. Track elements such as the rail, the rail
pads, the sleepers, as well as the ballast and the sub-ballast
layer, are thus directly influenced by the dynamic performance
of the railway vehicle.
The typical exciting mechanisms are:
•
Vehicle body dynamics in the frequency range
between 1 and 30 Hz
•
Out-of round wheels (10 to 20 Hz)
•
Wheel flats (10 to 20 Hz)
•
Rail irregularities, such as rail joints and skid
marks
Constraints that keep the vehicle dynamists from designing
the optimum vehicle suspension are typically:
•
Limit on the minimum vertical vehicle stiffness
because of a limit on the coupler height
differential between adjacent vehicles in the tare
and loaded condition
•
Volume occupied by, as well as the stresses within,
the suspension
o 2-12 x
•
Initial cost of the suspension
• Maintenance cost of the suspension
Similarly, the track engineer is limited in what he can do to
optimize the track structure. Typical constraints are:
•
The cost of rail pads.
•
Cost constraints on the amount of ballast and
formation material and other geo-technical
materials that could be used
• Track construction and track maintenance costs
Most track dynamics analysts include the so-called
Hertzian stiffness in their analysis. This is the vertical stiffness
attributed to the deformation of the wheel and rail under load.
It is a high-order stiffness and associated with high frequency
vibrations and impacts. These are mainly of concern to track
engineers and hence this effect is included under the section on
track support structures.
2.3.1.4 Types of Vertical Suspension
Conceptually, in its simplest form, the suspension of a railway
vehicle comprises four springs vertically coupling the four
journal bearings on two wheelsets to the body (see Figure
2.11). The four springs can be designed within the space and
for the load of a relatively small and light vehicle. However, as
the vehicle becomes heavier and larger, the following factors
come into play:
•
The ability to accommodate track twist by means of
spring deflection clashes with the demands on coupler
height differential limits between a loaded and an
empty vehicle.
•
The available volume for springs and dampers in the
region of the journal bearing is limited.
•
As the carrying capacity of the vehicle increases, the
wheel base increases. This leads to increased demands
on vertical deflection to accommodate track twist.
o 2-13 x
Primary Springs
Figure 2.11: Simple Vehicle Suspension Arrangement
A solution to the above problem is the railway bogie. The
bogie is a combination of a minimum of two wheelsets within
a suspension structure, which is pivoted beneath the vehicle
body as Figure 2.12 shows. A minimum of two bogies is fitted
to a vehicle. The bogie is the equivalent of a short wheel base
vehicle with a limited but adequate vertical spring deflection to
accommodate track twist. In addition, the carrying center plate
is of limited diameter. This coupling can be designed to permit
additional track twist by means of providing sufficient side
bearer clearance.
The bogie has become standard equipment under railway
vehicles. From a vertical load bearing point of view, there are
two basic types of bogies; the rigid frame and the three-piece
bogie. They differ structurally and in the form of suspension
design.
o 2-14 x
Figure 2.12: Basic Railway Bogie
2.3.1.4.1 Rigid Frame Bogie
The rigid frame bogie acts, vertically, very much like the model
of the simple railway vehicle described above. As indicated by
the name, the single bogie frame is typically in the form of a
rigid “H” shape as Figure 2.13 shows. The load of the vehicle
body is transferred from the center pivot through the “Hframe” to the springs placed above the journal bearings. These
springs form the vertical suspension and cater for all the
requirements for the suspension as listed above.
This type of bogie has possibly not found favor in heavy
haul operations, from a vertical suspension point of view, for
the following reasons:
•
Space constraints for springs with adequate
carrying capacity and deflection in the region of
the journal
•
Cost of providing four separate spring/damper
systems on the bogie
•
Cost of the H-frame from a manufacturing
complexity and tolerance point of view
o 2-15 x
Variations of the H-frame concept include, among others,
a torsionally soft bogie frame to accommodate track twist and
a bogie in which the bending stiffness of the side frames is
used for the suspension. None of these particular concepts has
found general acceptance in practice.
Side frame
Centre Pivot
Figure 2.13: Rigid Frame Bogie
2.3.1.4.2 Three-Piece Bogie
Generically, the three-piece bogie, as implied by the name,
comprises two side frames, each resting in a longitudinal
orientation, on the journals of the wheelsets. Figure 2.14 is a
sketch of a typical three-piece bogie. The side frames support a
cross member — the third piece — termed a bolster. The
bolster is fitted with a center pivot, which couples the bogie to
the vehicle body. The three pieces, two side frames and a
bolster, are each simply supported beams. This makes the
bogie a statically determinate structure and allows the structure
to articulate under conditions of track twist without loosing
vertical wheel load. The advantages of this structure for vertical
suspension are:
•
Efficient accommodation of track twist.
•
Suspension springs are limited to two nests
offering cost advantages with respect to the
number of suspension elements.
o 2-16 x
•
Suspension springs are in a region of the structure
where more space is available than at the journal
boxes.
A disadvantage is that the side frame forms part of the
unsprung mass on the wheelset. Furthermore, the lateral
dynamics of the bogie is not optimal.
Centre plate
Bolster
Coil springs
and Damper
Side frame
Figure 2.14: Three-piece Bogie
2.3.1.5 Suspension Damping
Associated with all vertical suspensions is some means of
dissipating the energy generated as the vehicle travels over
irregular track. In heavy haul applications, this is invariably
done by some frictional means even though friction damping
has many disadvantages, such as:
•
Having a non linear force/displacement
characteristic
•
Being susceptible to stick-slip action
•
Permitting the transmission of high frequency
vibration across the suspension
• Being susceptible to wear
However, the overriding advantages of the friction damper
are its:
•
low initial cost, and
•
robustness and low maintenance cost.
o 2-17 x
2.3.2 Inter-Wheelset and Lateral Suspension
This section describes the means by which the inherent
guidance properties of the railway wheelset are utilized, and the
dynamic disadvantages of the wheelset are countered. As
Section 2.1 describes, a single unconstrained railway wheelset is
designed to permit flange free curving and self-centering on
straight track. Early on in rail vehicle development, it was
found that the property of self-centering, as Figure 2.3 shows,
is unstable for all speeds of a single wheelset. This unstable
action is termed wheelset hunting. Wheelset hunting results in
increasing lateral deflection amplitudes, intermittent cyclical
flange contact on tangent track and shallow curves, and even
derailment as the lateral acceleration of the wheelset initiates
flange climb. It was soon realized that this instability could be
countered by coupling two wheelsets in the horizontal plane.
This coupling could, however, inhibit curving and guidance on
straight track. It was also realized that the lateral suspension
stiffness between the bogie and the body has an influence on
the stability and ride quality of the vehicle.
The formulation of suspension designs to optimize both
the curving and tracking ability, the lateral ride quality, as well
as the hunting stability of railway vehicles against the constraint
of low initial costs and maintainability has challenged vehicle
designers over the years.
2.3.2.1 Vehicle Dynamics
A better understanding of railway vehicle suspension dynamics
has been achieved over the last three decades. Hunting stability
is primarily a function of the bending and shear stiffness
between two wheelsets; therefore, these stiffness terms are
further described below.
Bending Stiffness: If two wheelsets are moved relative to one
another in opposing yaw senses, as Figure 2.15 shows, the
resistance to this motion is called the yaw constraint. If this
constraint is linear, the constraint directly between the
wheelsets is termed the bending stiffness.
o 2-18 x
Figure 2.15: Bending Mode
Shear Stiffness: If two wheelsets are deflected laterally
relative to one another in opposite senses while retaining
parallelism between axle centerlines, they are said to have
moved in a shear sense (Figure 2.16). If the constraint in this
mode is linear, the constraint is termed the shear stiffness.
Figure 2.17 illustrates bogie arrangement with various
degrees of bending and/or shear constraint.
Figure 2.16: Shear Mode
o 2-19 x
Research shows that for adequate wheelset hunting
stability, the coupling between two adjacent wheelsets requires
a combination of both bending and shear stiffness. This
combination may be chosen to optimize the vehicle
characteristics being designed. A whole range of stiffness
values may be chosen. There is, however, a minimum stiffness
for each that, if chosen, requires a relatively high stiffness be
chosen for the other constraint and vice versa. Figure 2.17
shows two examples.
Shear stiffness
8
0
Bending stiffness
8
0
Figure 2.17: Various Degrees of Bending and
Shear Constraints
Example 1: If a high-bending stiffness is chosen for a bogie
design, the designer cannot afford to introduce high shear
stiffness. This is typically the case in the conventional three-piece
bogie, the rigid-frame bogie and the force-steered bogie.
Example 2: If a low bending stiffness is chosen, the designer
needs to introduce a high shear constraint for adequate hunting
stability. This is typically required in self-steering bogie designs.
o 2-20 x
A choice of optimal intermediate stiffness for both the
bending and shear stiffness can result in extremely high
stability. The reason why this type of suspension design is not
always adopted is its higher initial cost, complexity, and
maintainability. Furthermore, a design with properties optimal
in a new bogie may not maintain such an optimal state over a
practical maintenance interval. A design that can maintain its
original state between maintenance interventions, may have too
high of an initial cost. On the other hand, the service
conditions or track topology may imply a bias to a particular
design. Associated with the high constraint stiffness is the need
for greater accuracy in the tolerances of components as
inaccurate stiff suspension elements lead to tracking
misalignment of the vehicle.
Another important feature of the suspension is the lateral
coupling between the wheelsets and the body via the bogie
frame and the center plate pivot. A relatively “soft” coupling,
which is often difficult to achieve, is helpful in uncoupling the
vehicle mass from the wheelset in a manner similar to reducing
the “unsprung mass” in a vertical sense. The stiffness of this
coupling should be carefully chosen as the natural frequency of
the vehicle body in nosing is easily activated by repetitive track
irregularities, such as rail joints, and by the natural kinematic
frequency in yaw of the wheelset.
2.3.2.2 Curving and Tracking Ability
As already mentioned, more than one wheelset requires
coupling in the horizontal plane so that wheelset stability is
obtained at any practical vehicle speed. The manner in which
this is done influences the ability of the coupled wheelsets to
negotiate curved track.
2.3.2.2.1 Bogies with a High Bending Stiffness
A high bending stiffness implies that both wheelsets remain
essentially parallel to one another and hence may not attain a
radial position in a curve. There is thus a limit on the ability of
the bogie to negotiate sharp curves without flange contact.
This limit is a function of track gauge, bogie wheelbase,
wheelset conicity, gauge clearance, and bogie rotational
resistance. Curving without flange contact is shown in Figure
o 2-21 x
2.18. The “clockwise” moments resulting from longitudinal
creep are in balance with the “anti-clockwise” moments
resulting from lateral creep. Hence the bogie is kept in
equilibrium.
Pure rolling position
a
α=
Rc
c
α=
γy
C11 r
0
r0l
γRc
a
Rc
C11 γ y
r0
V
2l
y
a
2C22 R
cL Track
C11
C11 γ y
r0
γy
r0
a
2C22 R
c
2a
Rc
Figure 2.18: Curving without Flange Contact
Typically, bogies on standard gauge, with wheelbases of
approximately 1.8 m may negotiate curves of between 1500 m
and 2000 m without flange contact. Under these conditions,
and with an accurately aligned bogie, the lateral and
longitudinal creepage is low; and, minimal rail and track
damage is experienced. Side and crown wear is minimal, with a
degree of material flow to the field side of both high and low
leg after about 200 MGT. This may be corrected by grinding.
Figures 2.19 and 2.20 show the relationship between the above
variables.
o 2-22 x
LATERAL DISPLACEMENT (mm)
35
2l = 1507 mm
2a = 1830 mm
r = 457 mm
28
Conicity = 0.05
21
14
Conicity = 0.2
7
0
200
1200
700
2200
1700
CURVE RADIUS (m)
Figure 2.19: Lateral Wheelset Displacement versus
Curve Radius
90
Rc = 200m
Rc = 200m
LATERAL DISPLACEMENT (mm)
Rc = 500m
72
Rc = 1000m
54
36
Rc = 500m
18
Rc = 1000m
0
1250
1750
2250
2750
3250
3750
4250
4750
TRUCK WHEELBASE (mm)
Solid curves, conicity = 0.05
Dashed curves, conicity = 0.2
Figure 2.20: Lateral Wheelset Displacement versus
Bogie Wheelbase
In sharper curves, below a radius of 1500 m, with low
wheelset conicities and reduced gauge clearance, or at large
track discontinuities or when mis-aligned bogies are present in
the vehicle fleet, flange contact is made and the bogie takes an
attitude as Figure 2.21 shows.
o 2-23 x
Flat
Flong
Flong
P1
2l
V
P2
Flong
Flong
2a
Flat
Figure 2.21: Flange Contact in Curves
The “anti-clockwise” moments on the bogie due to the
lateral creepage resulting from the angle of attack of the wheels
are larger than those “clockwise” moments that can be
generated by the longitudinal creepage, which must thus be
supplemented by a flange force. Rail/wheel contact is similar
to that shown in Figure 2.10 with high gauge corner wear being
experienced and excessive material flow to the field side of the
low leg. Under these conditions, the most cost effective and
quickest remedy is to apply lubricant to the flange and/or
gauge corner of the high leg. Lubrication does not change the
force balance in the curve but introduces a wear-reducing
mechanism between rail and wheel.
A further investigation may reveal one or more of the
following:
•
Wheel and rail profiles with two-point contact and a low
effective conicity: This should be checked for both new
and worn rail and wheel conditions. In this case,
conformal rail/wheel profile contact combinations
are advised to support lubrication as little can be
done to improve the force situation if the other
remedies described in this section are implemented.
Limits on worn profiles may have to be set. There is
o 2-24 x
a limit on the wheelset conicity, which may be
generated. High conicity profiles can be generated to
encourage flange free curving. This is mainly done
by means of asymmetric grinding of the rail and is
often not long lasting, needing repeated and
frequent attention under heavy axle load conditions
as the effect of the lateral creepage wears the rail
crown down and negates the rail grinding action (see
Figure 2.10). The long-term effectiveness of this
measure must thus be monitored. Asymmetric
grinding can also concentrate high stresses on the
gauge corner of the high leg leading to premature
fatigue failure. This must be monitored. High
conicity wheel profiles can also lead to vehicle
instability on tangent track and any change to the
wheel profile resulting in high conicity should first
be tested on tangent track.
•
Incorrect superelevation in curves: Excessive cant will
cause the bogie to steer out of the curve to counter
the resulting inward force. This will increase the
angle of attack and lateral creepage and the resulting
flange force. The curving speed thus has to be
optimized so that the vertical forces on the left and
the right leg of the curve are almost equal. This will
prevent one side of the track to be overloaded under
heavy haul traffic. The relationship between the
lateral forces acting on the center plate and the angle
of attack is illustrated in Figure 2.22.
•
Tracking accuracy: A check should be made on the
tracking accuracy of all vehicles as some may be
“biased” to certain sense curves and thus “biased”
against others.
•
Rotational resistance between bogie and vehicle body: The
rotational resistance between bogie and vehicle body
must be checked.
o 2-25 x
Decreasing α
FLATERAL
α
Centre
Plate
Increasing α
FLATERAL
Figure 2.22: Forces Acting on the Center Plate
2.3.2.2.2 Bogies with a Low Bending Stiffness
Generally termed “Steering Bogies” these bogies use the
longitudinal creep forces generated between the wheelset and
the rail to deflect the longitudinal springs, which create the
bending stiffness. This permits the wheelsets to align to an
almost radial position to the curve, as Figure 2.23 shows. The
lateral creep forces are reduced to almost zero, eliminating the
flange force and the effect on the rail shown in Figure 2.10.
Centre of Curve
Figure 2.23: Radial Alignment in a Curve
o 2-26 x
2.4 Practical Heavy Haul Vehicle Suspensions
2.4.1 Heavy Haul Wagon Bogies
As discussed previously, heavy haul wagons are predominantly
equipped with three-piece bogies. Some railroads have also
adopted a variety of three-piece bogies with steering
characteristics. In this section, both the conventional as well as
the steering bogie designs are discussed.
2.4.1.1 Conventional Three-Piece Bogies
Although these bogies are termed three-piece bogies, they have
many other components. Furthermore, some additions have
been made to improve their running performance. These
suspension components and the additions to the basic
construction are discussed below.
2.4.1.1.1 Secondary Suspension Spring Nest
The secondary suspension spring nest of the three-piece bogie
is situated in the side frame pocket, where it rests on the side
frame spring seat and supports the bolster protruding into the
side frame window (see Figure 2.24). The spring nest, also
commonly known as the secondary suspension, consists of a
number of inner and outer hot coiled springs and a friction
wedge damping arrangement. The number of springs and their
detailed design depend on the load carrying capacity of the
particular vehicle. The friction damper typically consists of cast
wedges, wedged between the side frame and the bolster,
supported by a stabilizer spring. This suspension arrangement
is thus designed to provide friction forces between the vertical
surface of the wedge and the side frame wear plate, and to
keep the side frame and bolster square relative to each other.
The latter helps to control bogie hunting.
o 2-27 x
Primary Suspension
Secondary Suspension
Figure 2.24: Self-Steering Three-piece Bogie
The following practical design limitations influence the
optimal design of the spring nest:
•
Available space for suspension elements
•
Allowable stress limits in suspension components
•
Tare to load deflection limits
•
Vehicle tracking performance
•
Manufacturing and maintenance costs
2.4.1.1.2 Friction Damping
As mentioned in the previous section, a spring loaded friction
wedge arrangement between the bolster and the side frame
pocket is used to provide damping to the dynamic reactions of
the vehicle as it travels over irregular track. In heavy haul threepiece bogies, two types of friction wedge arrangements are
commonly in use; i.e., constant and load sensitive designs as
Figure 2.25 and 2.26 show. Constant friction damping designs
are independent of the wagon load while load sensitive designs
provide more frictional damping under heavier loads.
In friction wedge suspension designs, use is made of the
friction coefficient between steel and steel to dissipate energy.
However, under certain circumstances, such as high operating
speeds or adverse track conditions, a high wear rate between
rubbing surfaces is experienced. The resulting friction wedge
rise (Figure 2.27) between the bolster and the wear liners in the
side frame pocket causes a reduction in the frictional damping
o 2-28 x
force and a change in the vertical, lateral and warp stiffness.
This can lead to unacceptable running dynamics. On the other
hand, loss of suspension travel due to wedge binding as a result
of wedge rotation (misalignment) due to side frame wear
(Figure 2.28), results in high impact forces being transmitted
between the wheel and the rail, damaging not only the vehicle
structure but also leading to accelerated track structure
deterioration.
To prevent excessive wear as well as to prevent the wedges
from sticking, some railroads have implemented resilient
friction elements. These urethane elastomeric wear surfaces
significantly reduce wear. Hence, the available friction damping
and the warp stiffness is maintained for longer periods of
service, eliminating regular bolster slope rework.
For some heavy haul container traffic, hydraulic stabilizers
are used to control higher speed bounce dynamics, to improve
ride quality, and to minimize wheel/rail interface reactions
resulting from adverse wagon dynamics and damage to the
payload.
Bolster
Wedge
Wedge
spring
Window of
side frame
Figure 2.25: Constant Friction Damping
o 2-29 x
Bolster
Wedge
Window of
side frame
Wedge
spring
Figure 2.26: Load Sensitive Friction Damping
Friction
Wedge
Rise
Wedge Wear
Limit
Height of
Friction Wedge
Above Truck Bolster
a
b
a<b
New or “Restored”
Condition
Worn Condition
Figure 2.27: Friction Wedge Rise
o 2-30 x
Worn Side Frame
Wear Plates
Wedge
Mechanical Stops Worn
Into Slope Face Wedge
Bolster
Dynamic Vertical Force
Wedge Rotation
Wedge
Bolster
Wedge
Figure 2.28: Friction Wedge
Rotation from Empty to
Loaded Condition
2.4.1.1.3 Bearing Adapters
One of the functions of the bearing adapter is to assist in
preventing the three-piece bogie frame from losenging.
Insufficient losenging constraint can be a cause of severe
hunting at low operating speeds. Under these conditions, the
bearing adapters generally experience wear that further reduces
the losenging constraint. Therefore, in conventional threepiece bogies, it is very important to limit wear in the friction
wedge area, as well as on the bearing adapters. In bogies
incorporating a stiff rubber shear pad above the bearing
adapters (Figure 2.29), wear in this area is minimized.
o 2-31 x
Rubber Shear Pad
Bearing Adaptor
Figure 2.29: Bearing Adapter and Stiff Rubber Shear Pad
2.4.1.1.4 Shear Stiffness Enhancers
Spring planks (Figure 2.30), frame bracing (Figure 2.31) and
friction wedges all enhance the shear stiffness of the threepiece bogie while permitting a soft lateral flexicoiling stiffness.
In the absence of high lateral friction wedge forces, rocker and
pendulum designs (Figure 2.30) can be used to further soften
the lateral ride quality of the vehicle.
The so-called swing motion bogie (Figure 2.30) was
designed to improve lateral ride by providing the equivalent of
swing hangers. The bolster is supported on springs through a
spring plank, which is carried on longitudinal knife edges at the
bottom of the bolster opening in the side frames. The side
frames in turn have knife edges in the pedestal areas that
engage with the bearing adapters. Hence, the side frames
function as swing links. The spring plank interconnecting the
side frames prevent the side frames from losenging with
respect to one another which reduces the risk of bogie hunting.
One of the more successful methods of controlling bogie
hunting has been bogie side frame cross-bracing as shown in
Figure 2.31. In this design, diagonal cross-braces are added
between the side frames and resilient pads are installed
between the side frame and the axle bearing adapters.
o 2-32 x
Bogies equipped with shear stiffness enhancers are often
designed to permit a degree of inter-axle bending by fitting
either relatively stiff pads between the journal box and the side
frame or by permitting some free-play in this area. These
designs are often longitudinally too soft for high braking
forces.
Primary
Rocker Seat
Side
Frame
Spring
Plank
Secondary
Rocker Seat
Rocker
Figure 2.30: Three-Piece Bogie with a Spring Plank
and Pendulum Design
o 2-33 x
Figure 2.31: Three-Piece Bogie with Frame Bracing
2.4.1.1.5 Rotational Resistance
The rotational resistance between the bogie and the vehicle
body is important to stabilize the bogie and thus to prevent the
three-piece bogie from hunting. Rotational resistance is usually
through friction in the center plate, but can also be assisted or
replaced by a spring constraint and/or hydraulic damping. The
three basic options are described below:
•
Frictional resistance: In the case of frictional resistance,
there is an increase in the rotational moment until the
friction force has been exceeding. At this stage there is
gross slippage and thus energy dissipation through
friction. There is no restoring force.
•
Spring resistance: If a spring constraint is used, there is a
constant increase in the restoring force while the bogie
negotiates a curve. Spring constraints alone would not
be able to dissipate rotational energy and can only
influence the natural resonance modes of the vehicle
suspension.
•
Hydraulic resistance: Hydraulic rotational constraints
pose no resistance in curve negotiation and have good
properties to absorb dynamic rotational energy on
tangent track.
o 2-34 x
The first two options are relatively cheap, but there is a
compromise between hunting and curving. The hydraulic
devices are more expensive and pose no compromise between
hunting and curving. The latter devices are generally only used
in premium bogies.
All three types of constraints are good for reducing
hunting on three-piece bogies. Generally no additional benefits
are obtained by fitting any additional devices to most steering
three-piece bogies.
Constant-contact side bearings provide an added level of
hunting control that results in maintenance savings and
improved ride quality. Described below are four basic types of
side bearers.
Non constant-contact side bearers (Figure 2.32): When these type
of side bearers are used, contact between the wagon body and
the side bearer only occurs in situations of extreme track twist
and/or centripetal lateral forces on the wagon body. This
design is only used to restrict rolling motion and does not
contribute to vehicle stability. Non constant-contact side
bearers are commonly used in self-steering three-piece bogie
designs.
Resilient constant-contact side bearers (Figure 2.33): This type of
constant side bearer provides some resilience (spring
resistance) for small relative rotations between the bogie
bolster and the body frame. These designs are more suitable
for lighter payloads.
Roller constant-contact side bearers (Figure 2.34): The type of
side bearer provides almost no rotational resistance and only
acts as a constraint to the rolling motion of the wagon body.
Roller assisted constant-contact resilient side bearers (Figure 2.35):
In this type of side bearer, a resilient bearing element provides
a controlled level of hunting in both the empty and loaded
condition, while the roller bearing element, which bears the
higher bottomed loads, allows the bogies to turn with
considerable less restraint than when these high loads are
carried by non-rolling sliding friction elements.
o 2-35 x
Side Bearers
Centre Pivot
Figure 2.32: Non Constant-Contact Side Bearers
Figure 2.33: Resilient ConstantContact Side Bearers
Figure 2.34: Roller ConstantContact Side Bearers
o 2-36 x
Figure 2.35: Roller Assisted
Constant-Contact Resilient Side Bearers
It is important to prevent wheel unloading through the
correct packing of the side bearers. Furthermore, it is
important to note that more frictional resistance is created due
to the longer moment arm when using the predetermined side
bearer pitch. This has to be compensated for by the assisting
rollers and/or the resilience in the constant-contact side
bearers.
2.4.1.2 Steering Three-Piece Bogies
The conventional three-piece bogie, which has been the
standard freight car bogie for many years, has the advantage of
having low manufacturing and maintenance costs. Its
disadvantages are several; it may not have adequate vehicle
stability at higher speeds; it definitely has poorer curving
performance that results in higher wear between the wheel
flange and the rail gauge corner; and it may result in
derailments due to excessive lateral forces and a high curving
resistance.
In conventional three-piece bogies, the inter-axle shear and
bending stiffness that is required for stability is obtained from
the lateral and longitudinal stiffness of the primary suspension
which acts between the axles via the bogie frame. As the lateral
and longitudinal suspension stiffness act in series, a reduction
in bending stiffness would also reduce the shear stiffness, thus
limiting the maximum allowable operating speed. For
optimized curving performance, yaw constraints lower than
acceptable in conventional bogies are thus required. This
o 2-37 x
necessitates the inclusion of direct linkages between the
wheelsets so that an inter-axle shear stiffness that is
independent of the inter-axle bending stiffness can be
provided. A suspension arrangement is thus required between
the two wheelsets of a bogie, which ensures a virtually pure
rolling motion of the wheelset in a curve and adequate hunting
stability on straight track. Such designs are found in the socalled self-steering and forced-steering bogie designs. The main
advantages of steering bogies are reduced flange wear,
improved lateral to vertical wheel/rail force ratios, lower
curving resistance, and better derailment and hunting stability.
2.4.1.2.1 Self-Steering Three-Piece Bogies
Research has shown that a certain amount of inter-axle shear,
as well as bending stiffness is required for adequate dynamic
vehicle stability. Conventional bogies use the shear stiffness of
the bogie frame to obtain inter-axle shear stiffness. With rigid
bogie frame constructions, an effective inter-axle shear
stiffness can be obtained. However, in the three-piece bogie
arrangement, the shear stiffness is inadequate for optimal
stability. Furthermore, if the frame is used for the transmission
of shear reactions between the wheelsets, a high wheelset yaw
constraint is required. This is not possible for steering bogies.
Consequently, to optimize the curving and stability
performance, inter-axle shear stiffness, independent of the
wheelset yaw constraint is required. In this case, the
longitudinal and lateral stiffness of the primary suspension can
be selected to best suit curving performance, stability, and ride
quality.
These concepts have lead to the development of selfsteering three-piece bogies such as the cross-anchor bogie, the
bissel (or wishbone) frame bogie, and the radial-arm bogie.
These bogies are briefly described below.
2.4.1.2.1.1 Cross-Anchor Bogie
The cross-anchor three-piece bogie design (Figure 2.36) retains
the three-piece frame arrangement as well as the load sensitive
vertical and lateral damping, but introduces rubber shear pads
at the journal boxes for a controlled yaw motion of the
wheelset and provides an additional inter-axle shear stiffness
o 2-38 x
independent of the shear constraint provided by the threepiece bogie frame. This inter-axle shear stiffness is obtained by
diagonally linking the journal boxes by means of cross-anchors.
Figure 2.36: Cross-Anchor Bogie
2.4.1.2.1.2 Bissel Frame Bogie
The bissel or articulated three-piece bogie (Figure 2.37)
incorporates a pair of steering arms interconnected by an
elastomeric connection located in the center of the bogie into
the conventional three-piece bogie. Hence, this bogie design
has a reduced inter-axle bending stiffness to enhance the
curving performance and an increased inter-axle shear stiffness
to provide adequate vehicle stability.
Figure 2.37: Bissel Frame Bogie
o 2-39 x
2.4.1.2.1.3 Radial-Arm Bogie
The radial arm bogie (Figure 2.38) is a further development of
the cross-anchor bogie. Like the cross-anchor bogie, it
connects the two wheelsets of the bogie in shear. This is not
achieved, however, by cross-anchors which have to be fitted
diagonally through the bolster and be connected to sub-frames,
but by radial arms which are positioned at the outside of the
side frames. For three-piece bogies, the radial steering arms
form an integral part with the bearing adapters that rest on the
package-type journal bearings.
Figure 2.38: Radial-Arm Bogie
2.4.1.2.2 Forced Steering Three-Piece Bogies
In these bogies, steering linkages serve a stabilizing as well as
steering function. Forced steering bogies usually have the outer
and/or inner wheelset connected to the bogie frame and
vehicle body by a vertical steering lever. The inter-axle shear
stiffness remains dependent on the lateral stiffness of the
primary suspension.
2.4.2 Locomotive Bogies
Locomotive bogies are normally of a rigid frame construction.
Rigid bogie frames can be used as there is no appreciable tare
to load ratio. The advantage of a rigid frame construction is
that a stable platform is provided for the motors and the drives
so that high traction forces can be transmitted. Furthermore,
o 2-40 x
the cost of a locomotive bogie is not significant in proportion
to the total vehicle cost. Through the accurate construction of
the bogie frame and relating components, such as the horn
guides (Figure 2.39), good tracking alignment is also possible.
Primary spring
Axlebox
Horn guides
Figure 2.39: Horn Guide Axle Box Arrangement
As a result of improved AC traction equipment, advanced
computer systems, a low weight shift, and radial steering
bogies, the tractive effort and adhesions level of modern
locomotives are improving. In radial steering bogies, the
wheelsets are now pulling in a direction more closely aligned to
the direction of travel resulting in more uniform creep. Hence
more traction force is available. Some of the recent
developments include 6-axle locomotives with radial bogies.
These locomotives can provide up to 610 kN continuous
tractive effort at 35% adhesion.
2.5 Rail and Wheel Profile Design
This section will propose practical, multi-radiussed wheel
profile shapes. Much of the discussion will center on complex
interactions involving a combination of contact mechanics,
tribology, and vehicle mechanics. It will involve a discussion
on different track topologies and geography as well as a
discussion on aspects of contact mechanics that are not yet
fully understood. Designers, however, must focus on a design
o 2-41 x
that works best for them today and which, from experience,
results in optimal performance. Designers are thereby
developing a logical design approach with their present
knowledge.
2.5.1 Basic Considerations
Before embarking on the design of rail and wheel profiles,
designers must address the following basic considerations:
1. Contact is not spread over the whole rail and wheel surface:
This rather obvious statement is often ignored. The
nature of rail and wheel profiles precludes contact
over the whole profile. This contact is limited to the
regions highlighted in Figure 2.40. This implies that
the shape of both rail and wheel will change. The
question is by how much, to what shape, to what
allowable limits and at what rate.
2. Contact is not evenly distributed over the regions shown in
Figure 2.40: The incidence of contact over the rail and
wheel profile on straight track is the highest on the
center of the tread as Figure 2.41 shows. It is more
sharply defined if pure conical wheel profiles are used
on rails with high profile curvatures and less so with
profiled wheel treads on flatter rails. The contact band
is also more sharply defined when the track gauge is
more consistent. Two-point contact between wheel
tread and rail crown on straight track is to be avoided
as it produces high- conicity contact and the danger of
vehicle instability.
Figure 2.40: Potential Contact on Rail and Wheel
o 2-42 x
Conical Wheel on
Sharp Crown Radius
Profiled
Wheel
Figure 2.41: Contact Distribution on the
Wheel Tread on Straight Track
Contact on the wheel profile in curves is
invariably symmetric, if there is a balance between left
and right hand curves. Contact on the rail is
unsymmetrical and depends on the sense of the curve.
In the case of profiled wheels, the outer wheel of the
leading wheelset makes contact closer to the gauge
corner and flange fillet than the trailing wheel, as
Figure 2.42 shows. Similar differences in contact occur
on the low leg. These contact differences can be
advantageous as they reduce the number of fatigue
cycles “seen” by both wheel and rail. Contact stresses
and creepages are invariably higher in curves. For
conical wheels, contact remains centrally placed on the
top of the rail. On the wheel tread, it moves off-center
to the taping line by the amount of the lateral
deflection of the wheelset. This effect of concentrating
the contact on wheel and rail is considered detrimental
to the fatigue life of the rail. The conical shape of
wheel treads is transient as they quickly wear to a more
conformal profile and thus the use of this wheel
profile type is not pursued further.
o 2-43 x
Leading Wheelset
Trailing Wheelset
Figure 2.42: Contact between Leading and
Trailing Wheelsets, and Rails in a Curve
3. The direct association of contact and wear to creepage is
erroneous as material migration also takes place across the
profiles: Models predicting the change in form of the
wheel and rail must consider both material removal
based on a frictional work function, and the fact that
material migration occurs over the profiles. This is
evident in change in forms of both wheel and rail
shown in Figures 2.10 and 2.43, in the work of
Kalousek4 and in observations made by the author.
These will be discussed more in the following sections.
Figure 2.43: Hollow Worn Wheel Causing Track Damage
o 2-44 x
2.5.2 Wheel and Rail Profiles Divided into Functional
Sections
The functionality and design of rail and wheel profiles may be
examined in terms of the following contact regions (See Figure
2.44):
•
Region A: Contact between the central region of the
rail crown and wheel tread
•
Region B: Contact between the gauge corner of the rail
and the flange fillet
•
Region C: Contact between the field sides of both rail
and wheel
RegionC
Region A
Reg
io
nB
Figure 2.44: Functional Regions of Rail/Wheel Contact
2.5.2.1 Region A: Central Region of the Rail Crown
and the Wheel Tread
Contact is made most often in this region and occurs as the
vehicle negotiates tangent track, mild curves (non-steering
bogies), or tight curves (steering bogies). As a result of these
conditions and rail and wheel profile geometry:
•
contact stresses are the lowest stresses encountered
between rail and wheel,
•
lateral creepages and forces are low; particularly, if the
vehicles are not subject to tracking inaccuracies or
instabilities,
o 2-45 x
•
longitudinal creepages and forces are significant in
relation to lateral creepages and a dominant
consideration for vehicle stability, and
•
vehicle speeds are higher than in sharper curves.
This region is thus primarily designed to optimize vehicle
stability while providing a radius differential to curve according
to the Newland model with non-steering bogies in mild curves
and for adequate radius differential for self-steering in tighter
curves. To reduce the rate of wear across this region the
conicity should be as low as possible within the curving
requirements to “spread” the incidence of contact as wide as
possible across the wheel tread. The rail crown has a radius
over this region and a profiled wheel is preferred.
Conicity and radius differential may be calculated
according to geometrical methods or according to more
sophisticated numerical methods used in vehicle multi-body
dynamic routines. When assessing conicity, a balance should be
drawn between lower contact stresses resulting from more
conformal contact (equal wheel and rail profile curvature) and
the resulting high conicity causing vehicle instability. Twopoint contact is to be avoided at all costs because of the high
resultant conicities and because of the wear associated with
two-point contact.
Adequate gauge clearance on tangent track must be
associated with decreased conicity and the spread of contact
over the wheel promoting pummeling. This clearance may be
obtained by increasing the nominal gauge, decreasing flange
thickness (if flange wear is under control), or decreasing the
“back-to-back” dimension across flanges on a wheelset, or a
combination of all three measures.
If too soft a rail is being used, a large difference in the rail
and wheel profile radii may have to be used to counter the
“flattening” effect caused by material flow. This design action
may have to be accompanied by grinding.
The conicity of new and worn profiles must be considered.
Good tracking may result in a hollowing of the wheel tread and
in altering the initial design conicity. This should be limited to
o 2-46 x
retain conicities within vehicle stability limits. Contact toward
the field side of the rail and wheel should be encouraged by
continuing the wheel profile radius beyond the taping line to
the field side of the rail.
2.5.2.2 Region B: Contact between the Gauge Corner
and Flange Fillet
As the contact patch in this region is small, contact is often
made under the most arduous stress conditions. If two-point
contact occurs, high wear rates and material flows are present.
If single point contact occurs, high contact stresses prevail
together with spin creep and high longitudinal creep. Contact
in the gauge corner is invariably associated with high angles of
attack and lateral creepage.
Flange contact will inevitably occur at some points on the
track, in tighter curves, at locations on the track where
alignment is not good. And at locations on the track where
there are discontinuities in the running profile, such as at
points and crossings, rail joints, and skid marks. If flange
contact is not designed properly, rail and wheel damage may
occur or vehicle guidance or stability may be impaired.
There are three generic options that the profile designer
must consider when examining flange contact. These are twopoint contact, single-point contact, and conformal contact, as
Figure 2.45 illustrates.
o 2-47 x
Two Point Contact
Single Point Contact
Conformal Contact
Figure 2.45: Three Generic Forms of
Flange Contact
2.5.2.2.1 Two-Point Contact
Two-point contact is associated with gross slippage and wear if
a flange force and lateral creep are present, as is the case in
curves. Under these conditions, wheel flange wear is
accelerated until the flange shape conforms to that of the rail.
Contact is often so severe that material flows occur on the
flange of the wheel, as Figure 2.46 shows. Experience shows
that under this condition, the flange often cuts under any
lubricating film applied to the contact zone.
o 2-48 x
Figure 2.46: Material Flow under
Severe Two-Point Contact
It is often argued that two-point contact is less damaging
to the rail because the vertical load is carried away from the
gauge corner. In addition, it is often applied to rails exhibiting
gauge corner fatigue defects as a result of improper past
maintenance. It does limit, however, the amount of radius
differential and steering ability available in a curve. If taken to
its conclusion, it will result in even worse contact conditions.
First the gauge corner of the rail is removed (Figure 2.47), then
the wheels eventually wear, because of two-point contact, to
the new gauge corner (Figure 2.48). Then the gauge corner is
further removed (Figure 2.49), and so on. Where does it end?
It could possibly end in dangerous, extremely high conicity,
one-point contact between the wheel and the rail on straight
track as Figure 2.50 shows. Notwithstanding the disadvantages
described above, gauge corner relief does give a short-term
extension to rail life.
Figure 2.47: Gauge Corner Removed
o 2-49 x
Figure 2.48: Wheels Wear to New Gauge Corner
Figure 4.49: Further Gauge Corner Relief
Figure 4.50: Resulting High Conicity and
High Stress Contact on Straight Track
2.5.2.2.2 Single-Point Contact
Single -point contact is probably the most damaging to vehicle
and track. The high contact stresses occurring under high creep
conditions result in fatigue of the gauge corner.
o 2-50 x
A case may be made, however, for single-point contact
on tangent track, as it is difficult to imagine that the low angles
of attack encountered would result in excessive wear and an
alteration to a designed wheel profile. It would also help reduce
conicities and hence improve vehicle stability on straight track.
This should not be taken to an extreme, as it will impair the
ability of the vehicle to centralize itself on straight track
resulting in wheel and flange wear, which may become nonsymmetric and damage the vehicle’s tracking ability.
This, in its mildest form, produces head checks, and in
its worst form, a breaking-out of the gauge corner of the rail
(Figure 2.51). It is associated with high longitudinal creepages
causing rail material flow but, more dangerously, vehicle
instabilities in the form of hunting and associated alternating
side wear on the track.
Single-point contact occurs as a result of:
•
Incorrect wheel and rail design
•
A flattening of the railhead in service as Figure
2.52 shows
•
Excessive hollowing of the wheel tread as Figure
2.53 shows
Figure 2.51: Crushing of the Gauge Corner
o 2-51 x
Figure 2.52: Hollow Worn Wheel Mounting
the Gauge Corner on a Flattened Rail
Figure 2.53: Single-Point Contact as a result
of Hollowing of a Wheel Tread
2.5.2.2.3 Conformal Flange Contact
Conformal flange contact is observed as the gauge-corner and
flange wear to a common profile under arduous flange contact
in curves. It has been observed by the author to be of
remarkably common form under different flange contact
conditions on different railroads An example of a conformal
profile design is given in Figure 2.54. Care should be taken not
to confuse this profile with those developed on the gauge
corner as a result of single-point contact.
Little is known of the complex contact conditions that
seem to keep the contacting shapes similar; the author
postulates the following:
o 2-52 x
Typical Conformal
Geometry in Flange
Fillet to be matched
to the Rail
R10
R 40
R 70
Figure 2.54: Conformal Contact Profile
•
Relative slip increases down the contact zone, as
Figure 2.55 shows.
•
Specific pressure decreases.
•
The above two effects act in a similar manner to
the “Constant Wear, Unequal Pressure” model
associated with contact between clutch plates.
•
There seems to be a degree of material migration
as indicated in Figure 2.55.
Normal
Pressure
Lateral
Creepage
Specific
Pressure
Decreases
Spin
Material
Flows
Relative
Slip
Increases
Figure 2.55: Constant Wear Model
associated with Flange Contact
o 2-53 x
Whatever the mechanism of this form of contact, the
author has found that rail and wheel profiles produced to this
shape have maintained their shape and have performed
successfully in terms of fatigue life. The advantage of using this
profile type is:
•
It retains its shape.
•
Gauge corner fatigue is controlled under
prevailing axle loads.
•
Lubricating films are supported due to low
specific pressures.
•
Conicity is “neutral” as it would seem that the
wheelset does not experience the high conicities
associated with single-point contact.
It is recommended that wheel and rail profiles be designed
to a conformal profile as Figure 2.54 shows. Wheels and rails
may be profiled during maintenance; rails may be rolled or
profiled immediately on installation in track. What is
important in designing these profiles are:
•
Radius and lengths of the profile arcs
•
Tangential contact in blending this profile with
that of the tread to ensure the minimum of twopoint contact between the tread and flange
profiles
•
A certain liberty is available in choosing the flange
angle to suit existing maintenance standards
•
Gauge corner radii should follow the flange
profile to blend in with the rail crown profile
without producing two-point rail/wheel contact
o 2-54 x
2.5.2.3 Region C: Contact between the Field Sides of
Both Wheel and Rail
Region C is probably the most difficult to optimize because
contact between rail and wheel ends in this region, and
eventually, notwithstanding the efforts of the designer, either
high contact stresses are generated as the outer edge of the
wheel profile bears on the rail (Figure 2.56), or contact ends
before the edge of the wheel giving rise to the development of
a false flange on the field side of the tread (Figure 2.57).
Often, both effects develop simultaneously as both
contact conditions prevail at different locations giving rise to
the contact condition shown in Figure 2.58, where both high
contact stresses occur together with high longitudinal creepage
steering the wheel in the incorrect sense. This is associated
with accelerated flange wear.
Figure 2.56: High Contact Stresses
on the Field Side of the Wheel
Contacting the Low Leg
Figure 2.57: High Contact Stresses
on the Field Side of the Low Leg
Figure 2.58: A Combination of High
Stress Contact Conditions
o 2-55 x
The author suggests that the wheel profile be continued
from the tread radius design suggested until the profile is
cylindrical or to 1:40 conical. This spreads the contact as far to
the field side as possible.
Adverse field contact may be minimized or controlled by
controlling the level of maximum wheel hollowing and/or
applying a suitable field side relief to the rails
2.5.3 Rail and Wheel Management
Rail and wheel management is discussed from the viewpoint of
the impact on vehicle design and rail and wheel profile
geometry. No attempt is made to prescribe rail-reconditioning
actions based on fatigue life. It is assumed that rail and wheel
profiles have been designed according to guidelines outlined in
Section 2.4.2 with the exception of certain rail and track
geometry to be described in this section. The reader will be
taken through a thought process based on a premise that the
rail and wheel profiles must be managed so that their
properties change as little as possible during their service life. It
is assumed that their profiles are reconditioned at sensible wear
limits.
2.5.3.1 Flange Lubrication
Although an attempt is often made to design the vehicle
suspension and the wheel treads (Region A) for degree of
curving, an assessment of the Newland model reveals that the
designs are not always adequate for flange free curving. The
immediate answer is to lubricate the flanges. There is no other
answer unless a redesign of the suspension is done. Flange
lubrication is a quick and effective remedy, and the steering
forces remain unchanged.
If the wheel profile in the region of the flange is conformal
to the rail the largest possible contact patch and lowest
possible contact stress is achieved. This condition can
effectively support a lubricating film. This is a prime reason for
conformal contact as lubrication immediately reduces the rate
of flange and gauge corner wear. Experience shows this could
be by as much as six times the original un-lubricated flange
o 2-56 x
life.5 Lubrication stabilizes the rate of change of the profiles
which has an important effect on other rail/wheel contact
conditions that are explained later.
Lubrication immediately reduces the slip and creep force
regime between flange and gauge corner, reducing material
flow and surely reducing traction and fatigue effects.
Lubrication tends to be an expensive and messy
maintenance practice coupled to irritating logistics. As a result,
any practice that the maintenance departments can do which is
“half-good” is a good practice. Some heavy haul experiences
are:
•
There seems to be more success with viscous
lubricants than with oils.
•
Application by means of wayside lubricators is
messy in the immediate vicinity of the lubricators
giving rise to locomotive adhesion problems).
Furthermore, this method is ineffective beyond
100 m from the lubricator and takes a long time to
spread after grinding, even if the maintenance
department remembers to replace them.
•
Locomotive flange lubrication can work if
supported by the locomotive maintenance
department who is usually averse to the mess they
make of the locomotive.
•
Lubrication directly on the rail from a moving
vehicle would seem to be a good technical
alternative. The lubricant is spread evenly over the
rails that require lubrication. Lubricants can be
applied immediately after grinding. This method
has the disadvantage of the logistics behind the
scheduling of the vehicle.
2.5.3.2 Hollow Wear and Rail Crown Flattening
Once lubrication has been successfully implemented, or where
the tracking conditions are good due to a high percentage of
tangent track or the use of steering bogies, hollow wear of the
wheel tread becomes an important issue as the lives of the
wheels are extended. The wheel treads (Section B) have been
o 2-57 x
designed to a hollow or radiussed profile within certain
conicity limits. Hollow wear beyond the limit set and/or the
flattening of the rail crown cause immediate conformal contact
on the tread. This has the effect of:
•
Producing a high conicity for small wheelset
deflections from the centerline of the track.
•
Producing low or negative conicities for higher
wheelset deflections from the centerline of the
track due to the “false flange” on the field side of
the wheelset. This reduces the restoring action of
the wheelset to the track centerline causing flange
contact with an angle of attack resulting in
alternate gouging of the rail on curved and tangent
track.
•
Contact under hollow worn conditions results in
high contact stresses between the gauge corner
and field side of the rail and the “false flanges”
which may occur either side of the hollow worn
pattern on the wheel (Figure 2.59).
* High Contact Stresses
*
*
*
*
Figure 2.59: Regions of High Contact Stress
on Hollow Worn Wheels
Management remedies for hollow wear or rail flattening
are harder rails, rail grinding to retain a radiussed rail crown,
and limits on the amount of hollow wear allowed. In
combination with these issues, gauge variation should be
implemented on sections of tangent track to help “spread” the
incidence of contact and reduce the rate at which hollow
wheel-tread wear takes place (Figure 2.60). This is referred to
as “pummelling” by Kalousek.4. “Pummelling” may be done by
rail crown grinding or by physically widening the gauge by
using asymmetric rail pads on concrete sleepers. Care should
o 2-58 x
be taken not to induce undesired vehicle dynamics in the
process and therefore long sections of tangent track should be
set to the altered gauge. Gauge variation should not be used in
curves for the reasons described below.
Figure 2.60: Spreading Contact on the
Wheel Tread by Gauge Variation
2.5.3.3 Gauge Control in Curves
Related to the issue of hollow wear and to the field side shape
of rail and wheel is gauge control in curves. As was pointedout, contact between rail and wheel ends in Section C and
produces contact conditions shown in Figures 2.57 and 2.58.
Under these conditions, the generation of radius differential
across the wheelset is impaired, leading not only to severe
contact stresses but to higher flange forces and higher flange
wear. This adverse effect can be present even when the wheel
is within hollow wear limits. As a result, the gauge should be
controlled in curves to avoid this type of contact by:
•
Removing material on the field side of the low leg
as Figure 2.61 shows.
•
Retaining relatively tight gauge during the life of
the side wear of the high leg. This may be done by
the use of asymmetric rail pads on concrete
sleepers. As experienced, the gauge in curves
should not be left to widen beyond 10 to 12 mm.
After this, fatigue damage occurs on the low leg
and flange and side wear accelerate.
o 2-59 x
Field side
Low Leg
Figure 2.61: Treatment of the Field Side
of the Low Leg in Curves
2.5.3.4 False Flange Contact on the Gauge Corner
If all of the above recommendations for profile design and
wheel/rail contact management are followed, there should not
be a problem with this type of contact. This phenomenon
occurs primarily as a result of operating with:
•
Wheels beyond a limit of hollow wear of
approximately 2 mm
•
Side wear on the high leg beyond 12 mm.
•
A combination of hollow wear and bogie tracking
inaccuracies. This results in an unsymmetrical hollow
worn wheel profile. Depending on the direction in
which the vehicle travels, or the sense of track
curvature, the convex portion of the profile may ride
up on the gauge corner.
All the above reasons for this convex contact lead to a
fatigue damaged and excessively flattened gauge corner
deviating from the so-called conformal contact described
above. Excessive gauge corner grinding results and leads to
two-point contact and a “no-win” situation.
2.6
TRACKING ACCURACY AND TOLERANCES
As Section 2.1 describes, the mechanism of the freely rolling
railway wheelset accommodates geometric inaccuracies. There
is, however, a limit to this accommodation that is a function of
the relative geometry of wheelset and track. In addition, the
constraint of the inter-wheelset suspension elements can
combine with certain wheelset tolerances to adversely affect
o 2-60 x
the tracking accuracy of the vehicle. The stiffer, or more rigid
these elements, the finer their corresponding tolerances and
those of the wheelset must be for a particular tracking
accuracy.
The error in tracking accuracy is reflected most obviously
in lateral non-symmetrical profile wear across wheelsets of the
vehicle or, in the extreme, non-symmetrical flange wear. Less
obvious are the associated unnecessary creepages experienced,
resulting in increased energy to haul the train and increased
stresses and subsequent material flows on both wheel and rail.
There is also considerable evidence that unsymmetrical
wheel wear reduces the stability of the vehicle. This stability
can, in turn, be a function of the direction of running of the
vehicle. A vehicle may run in a most stable manner in one
direction, but run in an unstable manner in the other direction,
particularly if intermittent flange contact is being made on one
leg of the rail.
2.6.1 Geometric Inaccuracies in Wheelset and Track
Geometry
Geometric inaccuracies in the wheelset and the track geometry
is due to the relationship between the rolling radius differential
between the two wheels on a wheelset and the lateral clearance
between the flange and the gauge corner of the rail; i.e., gauge
clearance. A rolling radius differential on a wheelset can be
caused by machining two wheels to different diameters as
measured at the taping line of the wheel profile (Figure 2.62) or
by machining the wheel profile in a non-symmetrical “rotated
sense” on the wheel (Figure 2.63). If the rotation of the wheel
profiles is symmetric, only the effective conicity will be
influenced.
o 2-61 x
cL Wheels
Do + t
Do - t
∆
Taping
Line
Taping
Line
cL Rails
γy
2t=2γy
Figure 2.62: Wheelset Diameter Differential
Taping Line
Mis-orientation
Profile Rotation
Figure 2.63: Wheel Profile Orientation
When a wheelset rolls on straight track it will displace
laterally by a distance ∆ to “find” equal rolling diameters. If the
gauge clearance is insufficient to accommodate the distance ∆,
flange contact will occur and flange and gauge corner wear will
take place. This wear can be most aggressive. This error will
induce non-symmetric curving of the wheelset, which will
cause it to “favor” a particular sense of curve.
In a similar manner, non-symmetric profiling of the rail
crown will cause the wheelset to displace laterally on the track
(Figure 2.64). This can also lead to flange and gauge corner
wear, although, in the case of rail non-symmetry in curves, this
can aid the curving of vehicles if the asymmetry is in the same
sense as the curve. See Section 2.4.
o 2-62 x
cL Wheels
∆
cL Track
Gauge Clearance = ∆ max
Figure 2.64: Non-Symmetric Rail Profiles
The relationship between wheelset and rail profile
inaccuracies and gauge clearance can be established by studying
the conicity relationship. The effective conicity, γ, of a wheelset
is defined as the difference in wheel radii, divided by twice the
lateral motion caused by the radii difference. This relationship
is given in Equation 1.
γ =
rolling radius differential
2y
(1)
The above relationship is for “pure rolling” and the lateral
position of the wheelset in curves is influenced by creepages
that are, in turn, a function of the vehicle suspension. It is a
good indication of the wheelset position on straight track,
where creepages are closer to zero.
2.6.2 Geometric Inaccuracies of the Wheelset and
Suspension
As mentioned above, geometric inaccuracies are associated
with a wheelset constrained by the suspension to which it is
coupled. This coupling can be either to the bogie frame or to
another wheelset through an inter-wheelset coupling. The
couplings and errors are a function of the type and stiffness of
the inter-wheelset couplings (see Figure 2.17 and Section
2.2.2). In the next section, the errors will be discussed as errors
in the bending and shear mode.
o 2-63 x
2.6.2.1 Errors in the Bending Mode
Errors in the bending mode constrain the wheelsets to be
permanently angled in yaw in opposite senses to each other
(Figure 2.65). The stiffer the suspension in bending, the less
the generated creep forces between rail and wheel can align the
wheelsets to an accurate tracking position. This error typically
occurs with the mismatch of the wheel base across a rigid
frame or three-piece bogie, where the axle is held rigidly to the
bogie frames. Steering bogies, with low bending stiffness can
accommodate these errors and some non-steering bogies have
latterly been fitted with, albeit stiff, shear pads between the
axle-box adapter and side frame to permit some deflection and
better tracking alignment.
A more accurate matching of side frame wheel bases on
three-piece bogies, or machining axle box pedestals has been
adopted in the past few years, together with the adoption of
shear pads, as mentioned above. Typically, components to
consider as contributing to this tracking inaccuracy are:
•
Bogie side frames having dissimilar wheelbases
(Figure 2.65)
•
Unsymmetrical bearing adapters (Figure 2.65)
o 2-64 x
Side Frame + t
Side Frame - t
The Side Frame Lengths
can Vary and the Adaptors
can be Machined Wrong
R
cL Adaptor
Figure 2.65: Errors in Bending Mode
2.6.2.2 Errors in the Shear Mode
Errors in the shear mode are typically associated with
suspensions with high shear stiffness. This can happen,
typically, with steering and premium bogies that use higher
shear stiffness. The axles of the wheelsets remain parallel to
each other but are misplaced laterally, which leads to
unsymmetrical wheel wear, high stresses in the contact patch,
and vehicle instability.
o 2-65 x
Typically, components contributing to this error are:
•
Bogie side frames with spring trays
unsymmetrically placed relative to the horn guides.
Figure 2.66 shows the situation in the empty
condition whereas Figure 2.67 shows the situation
in the loaded condition. Any shear stiffness
enhancement will “unsquare” the bogie
particularly under load.
•
Unsymmetrical cross-anchors, bissels and crossbracing (Figure 2.68)
•
Unsymmetrically mounted or machined wheelsets
(Figure 2.69)
Bolster on
spring tray
Figure 2.66: Unsymmetrical Spring Tray
in the Empty Condition
o 2-66 x
Bolster on
spring tray
Figure 2.67: Unsymmetrical Spring Tray
in the Loaded Condition
Figure 2.68: Unsymmetrical Cross-Anchors
o 2-67 x
cL Bearing
cL Bearing
cL Taping line
Taping line
Taping line
lL
=
=
=
=
cL Bearing
Figure 2.69: Unsymmetrically Mounted Wheelset
o 2-68 x
REFERENCES
1. Office for Research and Experiments of the
International Union of Railways, Question B 55,
“Prevention of derailment of goods wagons on distorted
track,” Report No. 8 (Final Report), Conditions for
negotiating track twist, Utrecht, April 1983.
2. Nadal M. J.: Locomotive a Vapeur, Collection encyclopedie
scientifique, biblioteque de mecanique appliquee et genie, Vol. 186
(Paris), 1908.
3. Tournay, H.M.: “Rail/wheel interaction from a track
and vehicle design perspective,” Proceedings of International
Heavy Haul Association’s Conference on Wheel/Rail Interaction,
Moscow, Russia, 14-17 June 1999.
4. Smith, R.E. and Kalousek, J.: “A design methodology
for wheel and rail profiles on steered railway vehicles,”
Proceedings of the 3rd International Symposium on Contact
Mechanics and Wear of Rail-Wheel Systems, Cambridge, UK,
July 1990, Elsevier, Amsterdam, 1990.
5. Tournay, H.M. and Giani, J.L.: “Rail/wheel interaction:
Multi disciplinary practices developed in South Africa,”
Conference on Railway Engineering, October 1995,
Melbourne, Australia.
6. Weinstock, H.: “Wheel climb derailment criteria for
evaluation of rail vehicle safety,” Paper No. 84-WA/RT1, 1984 ASME Winter Annual Meeting, Phoenix, Az,
November 1984.
o 2-69 x
APPENDIX
The derivation of conicity:
A profiled wheel tread may be approximated by a circular arc
or a serious of circular arcs running tangentially into one
another. Rail profiles may be described in a similar fashion. For
the sake of simplicity, consider wheel and rail profiles
comprising singular arcs as shown in Figure A-1.
cL
Yr
Wheelset
Track
Yr
ro
ro
Po
Po
Rr
δo
Yw
∆R=0
Yw
Xr
Xw
D2 = d o =
Or
gr
Or
Ow
Rw
Xr
Xw
gw-gr
2
δo
Ow
D2 = do
gw
Figure A-1.
Rectangular axes Xr, Yr (their origin at the center of the
wheel profile arc) may be fixed to the wheelset and move
laterally with the wheelset relative to the Xr, Yr systems. With
the wheelsets centrally placed on the track, the relative position
of the two axis systems are shown in Figure A-1a. Contact
between both wheel and rail must occur at a point (assumed to
be the center of the contact area), where the wheel profile arc
and rail profile arc have a common tangent. It may thus be
probed that points Ow, Or and Po are collinear.
o 2-70 x
A lateral displacement y, of the wheelset from the
centerline of the track displaces the Xw, Yw, co-ordinate system
relative to the Xr, Yr, co-ordinate system. Since contact must
still occur at a point where the rail and wheel profile arcs have
a common tangent, the contact point may be determined by
extending the line Ow,-Or to intersect the rail and wheel
profile arcs at points P1 and P2 (Figure A-1b). The equations
for conicity, as a function of the lateral deflection of the
wheelset, are given in Figure A-1b and may be approximated
by:

(d 0 + y ) − 1 − (d 0 − y ) 
∆R = RW  1 −
(RW − RR )
(RW − RR ) 

Since γ =
γ ≈
∆R
:
2y
RW δ 0
(RW − RR )
(1)
Derivation of the gravitational force:
Figure A-2 illustrates both a conical and a profiled wheel
tread displaced a distance y from the centerline of the track. If
the axles remain horizontal, the angle of the contact area of the
conical wheel to the horizontal (Figure A-2a) will remain
constant and equal to the cone angle γ, of the wheel.
The horizontal components of the normal reaction
between the wheel and rail Fn, will thus remain equal and of
opposite sign to one another, for any deflection of the wheelset
from the centerline of the track. The net lateral force on the
wheelset is thus zero. Examination of Figure A-2b shows that,
for profiled wheel treads, the angle of the contact area to the
horizontal changes and differs between two wheels on the
same wheelset. The lateral components of these reactions are
thus unequal for any displacement from the centerline of the
track thus causing a lateral force to act on the wheelset. For
o 2-71 x
pure circular wheel and rail profiles, the gravitational
suspension stiffness may be expressed by:
Gr =
W
(R W − R R )
Wheelset
Track
cL
Fn
Fn
W
Fl
Fl
Figure A-2a
cL
Wheelset
Track
Deflection of wheelset
from cL of track = y
G
Fn1
W
Fl1
F n2
Fl2
Rw
Rr
D1 = (d o + y)
Figure A-2b
o 2-72 x
D2 = (d o-y)
NOMENCLATURE
a = Half wheel base
b = Half bearing center distance
C11
=
Longitudinal creep coefficient
C 22
=
Lateral creep coefficient
l = Half distance between wheelset taping lines
ro = Wheel radius at taping line
Rc = Curve radius
V = Vehicle speed
y = Lateral displacement
α = Angle of yaw or angle of attack
β = Rail angle
γ = Wheel cone angle or effective conicity
o 2-73 x
Click Here To Go Back To Table of Contents
PART 3: WHEEL/RAIL PERFORMANCE
Written by Dr. Prof. Sergey Zakharov, TRC member
3.1 Application of Systems Approach to Wheel/Rail
Performance Study
Research and operating experiences have shown that the most
effective way to obtain cost effective operation of maintaining
wheels and rails is by treating vehicle/track interaction and
wheel/rail interface as a system.1, 2, 3 About 60 factors
influence different degrees of the wheel/rail performance, and
they can be grouped into five primary categories of research
and development (Figure 3.1):
•
•
•
•
•
Wheel/Rail Dynamics
Rail Contact Mechanics
Wheel/Rail Materials
Friction Management
Wheel/rail Damage Modes
Figure 3.1: Scheme of System Approach to Wheel/Rail
1
Performance Research and Development
o 3-1 x
By applying this knowledge to the wheel and rail as a
system, it is possible to take full advantage of the synergies
created between them by understanding damage modes and
their causes, and by developing optimized strategy of
wheel/rail performance.1
In more detail these factors and their interrelation that
define, for instance, wheel/rail wear as one of the widespread
damage modes, are schematically presented in Figure 3.23 and
discussed below.
Dynamics of vehicle/track interaction
I
II
III
Linear and angular wheel
set velocities
Distribution of erlative
slippage on contact
patches
Linear and angular coordinates of a wheel set
Distribution of
friction vectors
The "third
body"
properties
Forces and moments
acting from rails on
a wheel seta
Shapeand distribution of normal
and tangential stresses on contact
patches.
Wheel flange
and rail head
profiles
Wheel/Rail Wear
Figure 3.2 Scheme of Factors Determining
Wheel/Rail Wear
Dynamics of vehicle/track interaction: There are different levels
of dynamics applicable to a wheelset, a bogie, a car, a
locomotive, or a train. The study of vehicle/track interaction
makes it possible to define vertical and lateral forces acting on
the rail, the wheel to rail angle of attack, the wheelset position
in relation to the rail, and the wheel to rail relative slippage.
Linear and angular wheelset velocity: These include the angular
velocity of the wheelset rotation, the longitudinal velocity, the
lateral velocity of a wheelset, and the velocity of the wheelset
twist around the vertical axes.
Linear and angular co-ordinates of a wheelset: The most
significant of these are the lateral displacement and the angle of
the wheelset twist. The latter is used to calculate the angle of
o 3-2 x
the attack of the wheel on the rail. Though the angle of attack
is not the only dynamic parameter responsible for slippage, it is
the most significant one, especially for large angles of attack.
Forces and moments acting from rails on a wheelset: These are the
results of the integration of the normal and tangential stresses
on the contact patches.
Distribution of slippage on the contact patches: The measure of
slippage within the wheel/rail contact patch is known as the
relative slippage or creepage. The relative slippage λ is a nondimensional value, which for the wheel tread contact is
calculated as the ratio of the velocity of relative movement of
the surfaces to the linear velocity of the surface. The relative
velocity depends on wheel and rail profiles, the angle of attack,
and such dynamic parameters as the position of a wheelset and
its instantaneous axle of rotation. Vector of the relative
slippage may be represented by three components.3,5
Distribution of friction vectors resulted from unit normal load: The
value of the friction vector at any point of the contact patch is
equal to the rolling/sliding coefficient of friction both on the
rolling and gauge side of the railhead surfaces. The direction of
this vector coincides with surface slippage.
Shape and the distribution of the normal and tangential stresses on
the contact patches: This block provides for calculation of the
shape of the contact patches on the rolling and side surfaces of
the railhead, their mutual location, and the distribution of
contact stresses on the contact patches. Contact stresses
depend on the dynamics of vehicle/track interaction,
wheel/rail profiles and materials properties.
The “third body” properties: third body refers to material layers
which change their property from the initial properties of the
material to new properties because of the process of friction.
The third body layer property has considerable influence on
wear modes and wear rate. In turn, the wear process influences
the third body properties to a great extent.
Wheel flange and railhead profiles: Wheel and rail profiles
highly influence contact stresses and slippage.
o 3-3 x
Wear: Wheel/rail wear modes and rates are defined by the
distribution of stresses and the relative slippage on the contact
patches as well as the third body properties. Thermal effects
should also be considered particularly when studying braking
and slip regimes.
This scheme of study, with corresponding changes, can be
applied to study other damage modes.
3.2 Rail Contact Mechanics
3.2.1 General
Rail contact mechanics is the study of the relationship between
stress, creepage, and geometry of rail/wheel system. As can be
seen in Figure 3.2, once linear and angular wheelset
coordinates, velocities, and forces and moments acting from
rails on a wheelset are known from the dynamics of
vehicle/track interaction, then the magnitude and distribution
of normal and tangential stresses, relative slippage (creepage)
and friction on the contact patch can be found, provided that
the rail and wheel profiles and third body properties are
known. The latter is a rail/wheel contact mechanics problem.
The problem of rolling contact for bodies having elastically
identical characteristics like wheels and rail, may be presented
separately as normal and tangential problems.4 The goals of
the normal problem are to determine the size and shape of the
contact region and the normal contact stress distribution. The
results of the normal problem are used for the solution of the
tangential problem. The results of tangential problem are the
determination of the distribution of tractive (creep) forces and
the (spin) torque over the contact regions of adhesion and slip.
3.2.2 Normal Contact Stress
The work of Hertz presented the first reliable mathematical
solution of the normal problem. The problem reads as follows.
Two bodies (wheel tread and rail rolling surface) touch at a
point. The undeformed distance can be defined if the radius of
curvatures in the region of the contact point are known. The
elastic properties of wheel and rail are the same in Poisson
ratio (ν) and modulus of elasticity (E). If the bodies are loaded
by a normal force F , then an elliptic contact area comes into
o 3-4 x
being with a longer semi-axes along the rail longitudinal axes
(Figure 3.3).
Pmax
Figure 3.3 Herzian normal contact stress distribution
over contact area [3.5]
The maximum contact pressure p can be calculated as:6
p =
3
3 FE 2
2π 3 Re (1 − ν 2 ) 2
2
,
(3.1)
where Re equivalent radii, depending on the characteristic
radii of the interacting bodies (wheel and rail).
Thus the normal contact stress on the top of the rail and
wheel tread surface depends on the wheel to rail load, wheel
and top of rail radii, and the interacting materials properties.
Hertz’s theory is valid for contacting surfaces under the
following assumptions:
•
•
•
The contacting bodies are homogeneous and isotropic
The contacting surfaces are frictionless
The dimensions of the deformed contact patch remain
small compared to the dimensions of contacting
bodies and principal radii of curvature of undeformed
surfaces
o 3-5 x
•
Linear elastic half-space theory for both bodies is used
to solve the contact problem
• The contacting surfaces are smooth
During vehicle movement the position of wheel-set in
relation to the rails changes considerably resulting in various
combinations of wheel and rail contact zones (Figure 3.4).
Figure 3.4 Potential Contact Zones of
Wheel and Rail
Though the wheel and rail profiles in IHHA countries vary
considerably, it is possible to highlight three functional zones
of rail/wheel contact (Figure 3.5): (1) contact between the
central region of the rail crown and wheel tread as (Region A),
(2) contact between the gauge corner of the rail and the flange,
(Region B), and (3) contact between the field sides of both rail
and wheel (Region C).5 Even for a constant axle load, the
normal contact-stress distribution varies considerably because
of the differences in radii of curvatures in these regions of the
contact.
o 3-6 x
R egi onC
R eg ion A
Reg
io
n
B
Figure 3.5 Functional Regions of Rail/Wheel Contact
If a continuous radius of curvature exists throughout the
contact domain, then the Hertzian solution is valid. If the
contact domain is shared by two or more separate radii of
curvature R11 and R12 (Figure 3.6a) then the Hertzian
assumption is not valid and a non-Hertzian solution is
necessary for predicting the contact patch geometry. This
approach has important application for numerous
combinations of worn wheel and rail profiles.
There are many methods and computer programs that are
used to find the normal contact stress for non-conforming,
non-Herzian contacts. The complete actual non-Hertzian
solution may be achieved by using computer program
CONTACT.7 Due to the time consuming nature of the
complete non-Hertzian solution, various approximate
techniques have been developed. For instance, using an
ellipticized non-Hertzian geometry approach, which has shown
good agreement with the exact solution (Figure 3.6 a,b).8,9
o 3-7 x
Figure 3.6a Wheel/Rail Geometry
x,y,z – coordinate system, R11 ,R12, R1',R2 –
characteristic radii
Figure 3.6b. Contact Geometry and
8
Pressure Distribution
R11 = 355.6 mm,R12 = 291.6 mm, R1'=R2 = ∝, F = 100 kN
o 3-8 x
Another approach, which is used to find the contact stress
between worn wheel and rail profiles, requires the modeling of
the contacting bodies using Winkler's elastic foundation
assumption, which states that the deformation of the surface is
proportional to the normal pressure.10,11 The size of the contact
patch, and the normal stress distribution depends on wheel to
rail normal loads, rail and wheel profiles, lateral and angular
position of the wheelset in the point of initial contact, and rail
cant.
When a wheelset is moving in a curve, if the yaw angle
becomes large, it is possible to have the wheel contacting the
rail at two different points. Two-point contact results in two
contact patches: (A) on the rail crown and (B) at the gauge side
of the railhead. Because of the wheel to rail angle of attack (α)
(Figure 3.7a), the gauge side contact patch is moved forward
(Figure 3.7b). Increasing the angle of attack results in an
increase of the distance between contact patches, the
instantaneous axis of wheelset rotation, and thus an increase in
creepage and creep forces. In the flange contact zone of the
high rail, a level of contact stress of 3000 MPa is common.
a
b
Figure 3.7a: Relative Position of the
10
Wheel to the Rail
o 3-9 x
Figure 3.7b: Location of the Contact Zones
10
under Two-point Contact Conditions
(FA=110 kN, Fb=66kN )
At the contact of severely worn rail with new or worn
wheels, the pressure at the contact patch changes its
configuration. The contact patch size decreases considerably
and shifts to the field side of the high rail resulting in
corresponding growth of the contact pressure which may reach
the yield stress and result in railhead plastic flow.
Typically, contact stress on the top of the rail running
surface (region A) range between 1300 and 1700 MPa. The
increase in wheel load increases Hertzian contact stress on the
top of rail surface as a one third power of the load (see
expression 3.1).
Hollow wear of the wheel tread (see Section 3.5) results in
high contact stress, which may occur on either side of the
hollow worn pattern of the wheel. Contact stress could rise to
6000 MPa for a 2 mm hollow worn wheel.
o 3-10 x
High contact stresses are also generated as the outer edge
of the wheel profile bears down on the rail, or contact ends
before the edge of the wheel giving rise to the development of
a false flange on the field side of the tread.
Both the magnitude and distribution of contact stress is
significantly influenced by wheel/rail profiles and whether
single- or two-point contact conditions exist. Conformal
profiles tend to result in larger contact patch having decreased
levels of contact stress as compared with non-conformal
(counter-formal) profiles.
3.2.3 Creep Force — Creepage Behavior
Results of the solution of the normal problem are used to solve
the tangential problem, which is the distribution of tractive
(creep) forces and (spin) torque over contact region of
adhesion and slip, as well as the frictional work distribution.
From the motion analysis (kinematics) of the wheelset, as
well as the forces acting it, there are three components of
creep: longitudinal and lateral creep force and (spin) torque.
Longitudinal creep and resultant forces arise because
during traction in the direction of rolling, slip occurs in the
trailing region of the contact patch (Figure 3.8a). The greater
the value of traction force, the greater is the proportion of the
slip region in the contact patch. (Figure 3.8 b) until the tractive
force reaches its maximum level when the contact patch is not
capable to absorb any additional tractive effort.
The lateral slippage depends on the angle of attack of the
wheel on the rail. Spin creepage is primarily determined by the
cone angle.
When calculating creepage (relative slippage) at contact
points of wheel and rail moving in a curve, consider the
position of the instantaneous axis of wheelset rotation.3
o 3-11 x
Direction of
rolling
Adhesion
Microslip
Traction Distribution
µN
λ
Adhesion
Microslip
Adhesion
Microslip
Adhesion
Microslip
Slip
Figure 3.8: Relationship between traction and creep [3.5]
(a) longitudinal traction forces over the contact patch
(b) creep force –creepage curve
o 3-12 x
Adhesion Problem. The maximum level of tractive force
between a locomotive driving wheel and rail depends on the
capability of the contact patch to absorb traction. This is
expressed in the form of the adhesion coefficient which is a
ratio of traction force to normal load. Normally wheel/rail
adhesion reaches its maximum at the level of longitudinal
creepage 0.01-0.02.
However, the rheological behavior of interacting layers
forming a third body in rolling/sliding contact between the
wheel thread and the rail rolling surface directly affects the
traction-creepage curve. The third body is composed of a
mixture of materials whose composition is influenced by
environmental factors and railroad operating practices.12 There
exists a flow of materials forming third body layer. Inputs to
the layer include iron oxides from oxidation of rails and rail
wear, silica mainly from locomotive sand, hydrocarbons from
oil or grease that have been applied or migrated to the top of
rail, brake shoe debris, and other contaminants.
Output from the layer include displacement of materials
from the contact patch, removal by wear or washing off rail,
and consumption by other mechanisms.13
If the surfaces are clean and dry, the adhesion coefficient
stays at the high level for higher creepage values and train
speed. When rail and wheel surfaces are contaminated,
particularly with water and especially with lubricants, the
adhesion coefficient is reduced with relative slippage (Figure
3.9) and with train speed as well.14 This behavior should be
considered when using creep force–creepage dependencies
either for simulation of track dynamics or for studies of
locomotive design.
Surface roughness influences contact stresses and the
creep force-creepage curve. It has been shown that increasing
wheel and rail surface roughness tends to increase the real
contact stress distribution as compared with the Hertzian
solution, and tends to decrease the initial slope of creep
curve.15
A model has been suggested that establishes a relationship
between the shear modulus of elasticity, plasticity, the critical
o 3-13 x
shear stress, and shear strain.13 In this model, the shear stress
initially increases as the creapage increases. As the shear stress
reaches its critical value, the initial slope of creep curve
changes. Depending on the properties of the interacting layer,
which can be controlled using friction modifiers, shear stress
may increase, decrease or not change with increase of creepage
(see Section 3.4.2).
Figure 3.9: Traction Force Creepage Curve under
Influence of Water or Lubricants
The optimal level of adhesion the railroad could utilize,
regardless of weather, is defined as the dispatchable adhesion.
Typically this level is about 22%. There has been considerable
improvements in locomotive design aimed on the increase of
the adhesion level to 35% and even more.16 The level of
dispatchable adhesion will increase correspondingly.
A number of factors besides third body properties,
influence the levels of dispatchable adhesion; particularly, the
presence of hollow worn wheels.
There are some practical methods to more efficiently
utilize and to increase the adhesion level of the locomotives,
which is very important for heavy haul operations. One of
these methods, which proved to be efficient for heavy haul
operation, is described in References 17 and 18. The suggested
method is based on the measurement of wheelset slip when
traction force exceeds adhesion forces. It employs a system of
statistical estimates of the wheelset slippage which enables it to
o 3-14 x
discern admissibility or non-admissibility of achieved adhesion
loading.
3.2.4 Influence of Traction on the Load Carrying
Capacity of the Contact Area.
According to the Hertzian theory, the maximum static
compression stress is on the surface and the maximum shear
stress is under the surface at a depth of 0.78a, where a is half of
the contact patch length. Modeling of a semi-infinite half space
subjected to Hertzian contact stress shows (Figure 3.10)5 that
immediately beneath the contact patch the material is under a
triaxial state of stress. Three stress tensors are approximately
equal resulting in shear strain and high load carrying capacity.
Farther bellow the contact patch, these stresses become less
equal and the maximum shear stress increases to a maximum.
0.5
σz
po
(Normal
stress)
σ x/po and σ /p
y
o
0
(stresses parallel to contact patch)
0.5
1.5
1.0
x/a
0.300
95 0
0.2 .29
0
0.
28
3
0 .2
67
0.2
51
0. 2
36
τ 1 (Maximum
po shear stress)
σx/po
-1.0
0.173
1.0
1.5
2.0
z
a
Figure 3.10: Stresses Beneath the Contact Patch
Under Pure Normal Loading
When tractive force is applied to the surface, the
maximum shear stresses increase and move closer to the
surface. Even if the normal strain on the surface is elastic, it
can cause plastic shear strain under the surface. Because of the
rolling motion of the wheel, there is a cyclic compressiontension shear stress behavior of subsurface layers, resulting in
an accumulation of plastic deformation under the surface and
leading to residual stresses in the material.
o 3-15 x
This behavior of materials is the cause of various rolling
contact fatigue defects in wheels and rail. There are two
significant volumes of material that exhibit deformation. One
is a very thin layer on the surface of the contact patch, and the
other is the subsurface layer in the region of the maximum
shear stresses. When the traction forces are increased, these
areas became closer and may form one area of potential failure.
Figure 3.11 shows the influence of traction on load
carrying capacity of the contact.6, 25 This diagram, sometimes
called a shakedown diagram, describes the limits of material
behavior in terms of non-dimensional normal contact pressure
P0 /k as a function of non-dimensional traction coefficient
T/N, where P0 is the normal contact pressure, k is the shear
yield strength, T is the tangential (traction) force, and N is
normal load. At relatively low traction coefficient T/N,
cumulative plastic flow occurs below the rail surface. If the
traction coefficient is high (greater than about 0.3), plastic flow
is greatest at the rail surface. The accumulation of a large
number of unidirectional plastic strain increments "ratchets"
the surface layer of material until its ductility is exhausted.19
This diagram is used to explain the mechanisms of contact
fatigue defects and the surface work hardening process. The
rate of surface deterioration depends on the friction
coefficient, the maximum contact stress, and the yield strength
of the steel.
Figure 3.11: Shakedown Diagram
o 3-16 x
3.2.5 Approach to Wheel and Rail Profile Stress
Optimization
Optimized wheel and rail profiles are those which provide for
the best performance for a given application. Performance of
assets is generally judged on the following criteria:
•
Resistance to Wear
•
Resistance to Fatigue
•
Resistance to Corrugation Development
•
Minimization of the Ratio of Lateral to Vertical Truck
Forces
•
Minimization of Noise
• Maximization of Truck Stability
In this section, we look upon contact stress factors. For a
practical application the National Research Council of Canada
has developed a Profile Optimization Model (see Part 5
"Optimizing Rail and Wheel Performance.")20
Depending on railroad conditions, wheels undergo from
about from 6x107 to 8x107 cycles before turning.19, 20 The
contact stress due to passage of a wheel over a rail can be very
large, often exceeding the plastic limit of the wheel and rail
materials. Part of these cycles ratchet the rail and wheel
material until the metals reach their ductility limits, resulting in
railhead and wheel damages (see Section 3.5.2). An excessive
amplitude and frequency of stress cycles on the gauge corner
of the railhead result in the formation of shells in that area.
By using conformal profiles in curves, the pressure
distribution in the contact patch can be decreased to compare
with non-conformal profile. However, consideration should be
given to the fact that mutual wear of conformal profiles in a
curve occurs such that the pressure distribution tends to be
concentrated in the area of the instantaneous axis of wheelset
rotation (gauge area for the rail and the flange bottom for the
wheel) thus promoting contact fatigue and plastic flow
failures).
o 3-17 x
Thus, the following recommendations are made to
optimize wheel/rail profiles in terms of the contact stresses:20
•
Avoid contact stress that is greater than three times
the strength of material in shear.
•
Distribute contact points across the wheel tread and
railhead by profile design and rail grinding so that not
only are the high rail, tangent, and low rail profiles
different, but in tangent track there is more than one
contact band
•
Vary gauge clearance in tangent track intentionally.
3.3 RAIL AND WHEEL MATERIALS
3.3.1 Chemical Composition
Rails: Rails and wheels are metallurgically similar. Both use
high carbon (0.65-0.82 %) steels that have pearlitic or near
pearlitic structure. Many types of plain carbon, alloyed, and
heat-treated rail steels are available from world manufacturers
that supply rails to IHHA countries in North America,
Australia, South Africa, and Brazil. Historically Russia and
China have developed their own standards and production of
rails.22, 23 Though rail chemical composition are close,
technology of production may vary differ considerably,
particularly at diverse metallurgical plants.
Table 3.1 shows the difference in chemical compositions
of rails used in IHHA countries for heavy haul operation.
Because of the difficulty in obtaining the required
minimum rail hardness 300 HB, rail manufacturers in North
America are permitted to vary the composition of Mn, Ni, Cr,
Mo and V within specified limits.24
Wheels: There are many grades of wheel steel depending
on carbon content. Table 3.2 lists the chemical composition of
freight car wheels used in IHHA countries for heavy haul
operation. The majority of freight cars in North American
Railways uses class C wheels with a chemical composition
shown in Table 3.2. Many railroads use class B wheels for
locomotives that differ only in carbon content (0.57-0.67%).
Russian state standards on wheels regulate non-metallic
o 3-18 x
inclusions in wheel steel by specifying the permitted size of
oxide accumulation (< 2 mm) and indices (in points from 1 to
10) of line oxides (<1), brittle silicates (<3.5), plastic silicates
(<4.0) and sulfides (<3.5).
One wear and multiple wear wheels (either wrought or
cast) are used by Railways of IHHA countries. Russian railways
use freight, passenger, and locomotive types of wheels. Freight
cars use wrought steel wheels; locomotives use tyred wheels.
Table 3.1 Chemical Composition (weight % ) of Rails
Elements
C
USA,
Canada,
Brazil
Australia
0.720.720.82
0.82
Mn
0.800.801.10
1.25
Si
0.100.150.60
0.58
S
0.037
0.025
max
max
P
0.035
0.025
max
max
Cr
0.25—
0.50
V
0.03
—
max
Ni
0.25
—
max
Mo
0.10
—
max
*standard is being changed
South
Africa
(S-60)*
0.650.80
0.801.30
0.300.90
0.03
max
0.03
max
0.701.30
—
China
Russia
GOST R
51685
(T1)
0.720.80
0.701.05
0.500.80
0.035
max
0.035
max
—
0.710.82
0.751.05
0.250.45
0.45
max
0.035
max
—
—
0.040.06
—
0.030.07
—
—
—
—
o 3-19 x
Sweden
BV50&
UIC60
0.600.82
0.801.30
0.300.90
0.025
max
0.025
max
0.801.3
—
—
—
Table 3.2: Chemical Composition (weight , %) of Wheels
Elements
C
Mn
Si
S
P
North
America,
Brazil
(Class C)
Australia
South
Africa
0.670.77
0.600.85
0.15
max
0.050
max
0.050
max
0.670.77
0.601.00
0.15
max
0.035
max
0.04
max
0.670.77
0.600.85
0.15
max
0.050
max
0.050
max
China
0.550.65
0.500.80
0.170.37
<0.040
<0.035
Russia
(freight
cars)
Sweden
(Ore
line)
0.550.65
0.500.90
0.220.45
0.045
max
0.035
max
0.670.72
0.730.85
0.20-0.40
0.020
max
0.025
max
3.3.2 Microstructure
Rails: Pearlitic steels continue to be used for heavy haul railway
track. Rails are available in either an as-rolled or a heat-treated
condition. High strength heat-treated rails are widely used for
heavy haul. The heat-treated rails are made from rail steel that
contain carbon in an amount close to the eutectoid
composition, which leads to a microstructure of pearlite. The
fine pearltic structure is promoted by the addition of alloying
elements, such as chromium, molybdenum, and vanadium, or
by accelerated cooling.
It has been shown that the optimal structure of high
strength heat-treated rails is fine lamellae of ferrite and iron
carbide called fine pearlite.
Heat treatment can be done off- or in-line during rail
production. In an off-line process the as-rolled rail is cooled to
a room temperature and its head is then reheated by various
methods, such as by an induction method, followed by
accelerated cooling of the railhead. This heat treatment process
forms fine-grained austenite in the railhead depth while the
railhead is hot. The process of austenization is controlled to
assure dissolution of the carbides and the development of finegrained austenite. It is necessary to cool rail rapidly to produce
ultra fine perlite. All processes for surface hardening of rail
require controlled heating and rapid cooling.
o 3-20 x
Another method of off-line heat treatment results in
through-hardened rail.30 Full length rails are heated in a
furnace, quenched in oil and then reheated in a tempering
process that results in forming fine pearlite through the entire
rail cross section. In-line heat treatment processes can make
use of the latent heat of the hot-rolled rail, so the entire cross
section of the railhead can be heat treated with almost uniform
distribution of the temperature and austenite grain size.
Both methods have advantages. For instance, by using offline hardening, it became possible to eliminate rail roller
straightening;31 whereas, during in-line processes, it is easier to
obtain uniform distribution of structure.
The microstructure and its changes along the rail cross
section control the mechanical and tribological properties of
rail steel.
Manufacturers produce a great variety of rail steels. For
instance, North America railroads can purchase standard,
intermediate, premium, or super grades. Standard rails are plain
carbon unheat treated, intermediate rails are alloyed, hot-bed
cooled, premium rails are plain carbon, but fully heat treated,
and super rails are micro-alloyed and head-hardened. 21 Russian
Railways can purchase three categories of rails that are
manufactured according to the state standards. 22, 23 The type
of rail selected is based on how effective it performs under
specific operating conditions. IHHA countries have selected
different types of rails that are considered to be most suitable
for the particular operating conditions for each railroad with
the objective of optimizing the operating and maintenance
practices and achieving the greatest economic return.
Improvement of the rail steel structure and cleanliness. Work is
continuing on projects to improve the quality of pearlitic rail
steels by improving the pearlitic structure and by improving rail
cleanliness. The wear resistance of pearlitic steels can be
improved by decreasing the lamella spacing of the ferrite and
iron carbide in the pearlitic structure. In addition, work is in
progress to increase the quantity of the cementite (iron
carbide) phase which aids wear resistance in the pearlite
structure. This can be achieved by the development of
o 3-21 x
hypereutectoid rail steels. For instance, steel containing 0.85 %
of carbon, micro-alloyed by 0.05% vanadium has a surface
hardness 375 HB27 whereas steel with 0.9% carbon alloyed by
0.25% chromium has a hardness of about 395 HB.33
Bainitic steels. In an attempt to improve the surface damage
resistance of rails used for high speed and heavy haul
operations, rail steels with bainitic structure have been
developed.28, 30, 31, 32 Bainitic steels contain 0.20-0.43% of
carbon. Laboratory experiments have shown that bainitic steels
have higher tensile strength and elongation than premium
pearlitic steels (for instance, 1420 MPa and 15% at room
temperature for steel with 0.35% carbon). 34 It is reported that
these rails obtain better resistance to rolling contact fatigue
defects and retain a wear resistance which gradually became
comparable to premium pearlitic steel rails. 32,34 However, the
performance of bainitic steel rails under heavy haul conditions
has not been established.
Rail cleanliness. Rail cleanliness is judged based on the
amount and distribution of soft and hard inclusions. Hard
inclusions or inclusion stringers, such as alumina-silicate(with
a composition of Al2 O3- Si O2), contribute to the initiation of
sub-surface rolling contact fatigue. Stringers of inclusions are
strongly implicated in the formation of horizontal split head
defects (see Section 3.5.3) and in the development of deepseated shells that lead to transverse defects. To quantitatively
evaluate the role of inclusion stringers in the formation of
shells, a shell index was suggested.37
Regulations (standards) on rail production limit the
content of non-metallic inclusions by specifying their size in
millimeters or by using the “inclusion” index. 21
Improved cleanliness can be achieved by the adoption of
continuous casting techniques and moving away from the use
of aluminum as a deoxidizer.
A very important factor affecting rail damages is the
concentration of gases (oxygen, hydrogen, and nitrogen) in the
rail steel. An increased concentration of gases results in a
decrease of the rolling contact fatigue and in lowered resistance
to brittle fracture. Hydrogen contributes to the formation of
o 3-22 x
shatter cracks and thus causes broken rail. High-quality rails
should contain not more than 0.00015% mass percent of
hydrogen and not more than 0.0002% of oxygen.27
Rail steel manufacturers pay close attention to clean steel
making process. One of the methods of rail steel refining is
electro-slag remelting.27 Another metallurgical method to
develop fine dispersed microstructure is to improve rail quality
by micro-alloying with nitride-forming elements, such as
vanadium and aluminum.34
3.3.3 Mechanical Properties
Rails: The mechanical properties of rails are evaluated by yield,
tensile and fatigue strength, hardness, and fracture toughness.
Yield strength is an indication of the material's plastic flow
characteristics and work hardening. Tensile and fatigue
strength is an indication of fatigue resistance of rail materials
that are characterized by laboratory or gull scale fatigue
tests.35,36
Table 3.3 lists the mechanical properties of high-strength
rail for some of IHHA countries. These rails have a yield
strength of about 760-790 MPa and a tensile strength of about
1170 MPa. Standard (as rolled) rail has lower strength
characteristics and hardness with a yield strength of about 480
MPa and a tensile strength of about 960 MPa.
Table 3.3: Mechanical Properties of High Strength Rails
USA,
Canada,
Brazil
South
Africa
China
Russia
GOSTR
51685
Yield
Strength,
MPa (min)
Tensile
Strength
MPa (min)
Elongation
% (min)
758
640
805
794
640
1172
1080
1175
1176
1080
10
9
10
6
9
Brinell
Hardness
at the
surface
340-390
340
340390
331-388
320360
Property
o 3-23 x
Sweden
Hardness. Hardness is a very important mechanical
characteristic that determines rail performance to a great
extent. Hardness and its distribution along the railhead depth
governs wear resistance, rolling contact fatigue resistance, and
plastic flow of the railhead. The types of rails available on the
market represent a wide range of surface harnesses. In North
America, there is a requirement to have a minimum hardness
of 300 HB on the rail surface. High-strength heat-treated rails
have a surface hardness in the range of 330 –390 HB. Russian
standards regulate hardness at a depth 12-20 mm below the
railhead surface, and specify that the rail hardness at this depth
should not be less than 86% of the hardness at the surface, in
particular not less than 300 HB. This requirement is necessary
to provide appropriate wear resistance when the railhead is
worn out. It is also important for resistance to the initiation of
rolling contact fatigue. There is also a requirement controlling
the rail hardness deviation along the rail length to no more
than 30 HB.
The most common methods of increasing hardness are by
increasing the volume fraction of carbide and by refining the
pearlitic structure. Refinement may be achieved by microalloying or by accelerated cooling.
Pearlitic steels have a maximum hardness level of about
400 HB. Bainitic rail steels may have hardness level of 500-550
HB. Study and tests show that bainitic steels have high
resistance to different failure modes, but at present insufficient
wear resistance. Rail manufacturers continue to work to
increase the wear resistance of bainitic steels. 34
After multiple wheel/rail interactions in operation,
hardness of wheel and rail interacting surfaces increases
compared with the initial surface hardness. It is difficult to
control work hardening during operation. However, it should
be further considered when selecting optimal rail grinding and
wheel truing practices (see Section 3.5.1).
Fracture Toughness. Fracture toughness governs the ability of
rail steel to resist propagation of brittle cracks from rolling
contact fatigue and other rail fatigue defects. Good fracture
toughness is particularly important in the prevention and
propagation of transverse defects.
o 3-24 x
Fracture toughness for rail steels may be evaluated by the
impact strength. According to the Russian standard on rails,
the impact strength of modern rail steel should not be less than
0.25 MJ/m2 at a temperature of+20oC.22
Micro-alloying of steel with vanadium together with
controlled de-oxidation with aluminum and silico-calcium
increases impact strength at low temperature,39 which is
important for railways operating at subzero temperature.
Increasing the nitrogen content in steel to 0.0015% also
enhances impact strength at sub-zero temperature.40 Increasing
phosphorus has a negative influence on fracture toughness.
Notice that fatigue initiation and crack growth
characteristics are characterized by fracture mechanics
principles and tests.
Residual Stresses: Residual stress in rail are the result of:
•
•
•
rail manufacturing process.
contact stress due to passing wheels, and
rail welding.
Manufacturing originated residual stresses arise because of
differences in the time of phase transformations in the
railhead, web, and foot during rail cooling in the rail
manufacturing process. A large contribution to the residual
stresses can be made by the cold roller straightening process. If
high roller straightening settings are used, rail failure may
occur.41 Several technique are available for evaluating the
magnitude of rail residual stresses, including strain gauging, saw
cutting or hole drilling, which are destructive methods, or
neutron diffraction and acoustic methods which are nondestructive. The simplest method is the web saw cutting
method. Depending on the type of straightening process
residual stress may be of the order of 100-300 MPa. Residual
stress measurement methods and failure criteria were
developed based on stress intensity and allowable stress
coefficient.41
Stresses in fully quenched heat-treated rails are tensile in
the railhead and base, and compressive in the rail web.
Railhead hardening usually leads to a comprehensive residual
o 3-25 x
stress in the railhead. There is a rail manufacturing process that
provides for optimized residual stress distribution. The process
involves an off-line rail treatment with controlled reheating and
cooling of the entire rail section which avoids the need for
roller straightening.31
Work hardening of the rail surface layer by passing wheels
introduces comprehensive residual stress in the railhead. These
stresses protect the surface layers by inhibiting the rate and
depth of contact fatigue crack propagation.
Rail welding results in residual stresses that are distributed
in a very complex manner with respect to their magnitude and
direction. In many cases, these stresses are the cause of rail
web failure. Use of improved welding technology and post
weld heat treatment considerably decreases the extent of weld
initiated residual stresses (see Section 3.5.3).
3.3.4 Wheels
Mechanical Properties: For North American wheels, the only
mechanical property requirement is hardness, which for Class
C wheels is 321-363 HB. Russian state standard22 requires that
tensile strength of freight car wheel steel should not be less
than 911-1107 MPa, the elongation not less than 8% , the
surface hardness not less than 290 HB, and the hardness at the
depth of maximum wear not less than 255 HB. To compensate
for the considerable differences in the hardness of wheels and
rails used in freight cars on the Russian Railways, work has
been done in:
•
improving rim hardening technology in wrought wheels,
•
increasing carbon content in wheel steel to the level in rail
steel thus obtaining surface hardness in wheels to the level
of 320-400 HB, and
•
introducing flange hardening technologies (plasma , laser,
weld-on treatments).42
Residual stresses: Residual stresses are the controlling factor
in wheel thermal failure. Improper heat treatment results in
brittle failure of wheels. To prevent thermal cracks from
developing, a rim quenching process is introduced, forming
beneficial circumferential compressive residual stresses in the
o 3-26 x
wheel rim. However, overheating of the wheels due to severe
braking may destroy the beneficial compressive residual stress.
3.3.5 General Concept of Wheel/Rail Material Selection
The material quality of rails and wheels significantly affects
their resistance to wear, rolling contact fatigue, fatigue, and
plastic flow. Improvement in rail/wheel material quality may
significantly increase the range of allowable contact stresses.
Many material characteristics are in conflict with one
another. For instance, hardness and fracture toughness are
inversely related. That is why a systems approach must be
taken to establishing the criteria for material selection.
3.3.5.1 Criteria for Rail and Wheel Material Selection
(a) There is a set of material property characteristics, the socalled constructive strength of material, which the material
should possess to meet operating requirements.43 There is a set
of laboratory, wheel/rail interaction simulated tests,32 which
help to establish these characteristics:
• Contact Fatigue Tests44
• Wear Resistance Tests45
• Static and Cyclic Resistance to Crack Propagation
Tests
• Rail Static, Cyclic and Impact Tests
The values selected after this set of tests have to be
validated first in field simulated tests (for instance at the
Transportation Technology Center’s Facility for Accelerated
Service Testing, Pueblo, Colorado, USA or Test Center of
VNIIZhT in Russia). After the simulated tests they are used in
revenue service for pilot testing
(b) The ability of material to resist surface and subsurface
flow can be checked by a shakedown diagram (see Figure 3.11).
The rate of surface deterioration depends on the traction
coefficient, the maximum contact stress, and the strength of
the steel
Recommended manufacturing techniques: The next generation of
rails and wheels will be obtained though improved steel making
facilities, hardening technologies, and inspection techniques.
o 3-27 x
Rails: Further improvement in pearlitic steel cleanliness by
diverse refining technologies includes:
•
Use of continuous casting techniques
•
Use of micro-alloyed rail steel
•
Increase minimum hardness of rails up to HB 300 in
tangent and up to HB 340 in curve track sections
•
Application of residual stress measurement technique.
Wheels:
•
Use rim quenching technology to create compressive
residual stress in rim and eliminate thermal failures and
at the same time increase the hardness of wheels to the
level of heat-treated rails.
•
Use micro-alloyed wheel steel.
•
Work on use of wheel flange surface hardening
technologies.
However, to optimize the wheel and rail system, it is
necessary to consider wheel/rail dynamics, contact mechanics,
rail/wheel material and friction management factors.
3.4 Lubrication and Friction Management
3.4.1 Some Tibology Considerations
Friction plays a major role in wheel/rail interface processes,
particularly in adhesion, braking, wear and rolling contact
fatigue damage, formation of skid flats, steering and hunting of
locomotive and car bogies, wheel squeal, and wheel climb
resulting in derailment.
Particular influence on wheel/rail interface performance
has the “third body layer.” The concept of the third body has
been introduced in tribology by Kragelsky and his schooland
developed by Godet and his supporters. 46 This concept has
been fruitfully applied to wheel/rail interaction problems to
calculate the distribution of tangential forces in the contact
zone. 47,13 The composition and properties of the third body
are influenced by the presence of lubricants, sand, wear debris,
surface roughness, environmental conditions, volume material
o 3-28 x
properties, and hardened layers resulting from wheel/rail
interaction. 48,12
There exists a flow of materials forming third body layer.
Inputs to the layer include iron oxides from oxidation of rails
and rail wear, silica mainly from locomotive sand,
hydrocarbons from oil or grease that have been applied or
migrated to the top of rail, brake shoe debris, and other
contaminants. Outputs from the layer include displacement of
materials from the contact patch, removal by wear or washingoff rail, and consumption by other mechanisms.47 The flow of
materials in third body is closely correlated with wheel-rail
adhesion.
Great influence on friction and wear in wheel/rail interface
has the surface temperature and its gradient in the direction
normal to the surfaces. High temperature and its gradients
cause chemical and structural transformations of third body
layers and their rheological properties, resulting in additional
surface stresses, changes in the coefficient of friction, and wear
rate. Consider the influence of temperature and its gradients
when studying wheel/rail interaction.49
3.4.2 Rail Gauge/Wheel Flange Lubrication
Rail or wheel flange lubrication is not new technology. Many
years ago steam locomotives were equipped with flange
lubricators. Wayside rail lubricators have also been in use for a
long time.
The importance of lubrication has increased as axle loads
and train mass have increased. The increasing axle loads and
productivity requirements to increase the tonnage of railways
heavily influence rail and wheel wear.
As Figure 3.1 shows, wheel and rail wear depends on third
body properties, particularly the friction coefficient, contact
stress, and relative slippage of contacting bodies, which in turn
depend on angle of attack, wheel/rail profiles, truck steering
abilities, and operating conditions. These form a complex
tribology system. Unbalancing this system negatively towards
unlubricated conditions results in creating high normal and
tangential wheel flange and railhead forces, and relative
slippage that leads to catastrophic wear. It is extremely
o 3-29 x
difficult, if not impossible, to return this system to normal
functioning; i.e., to the mild wear mode unless lubrication is
introduced. For instance, the Russian Railways took four years
to introduce wheel flange/rail gauge face lubrication to
decrease locomotive/car wheel flange and railhead wear rate to
the level that the rail system could operate efficiently.50
This became possible because of creating and
implementing a highly structured lubrication management
program. Criteria for lubrication effectiveness had to be
developed and used during the implementation of the
program. Criteria included quantitative measurement of the
state of lubrication, assessing the benefits of lubrication in the
form of reduced fuel consumption, effects of lubrication on
rail side wear, locomotive wheel flange wear, and the rate of
replacement of car wheelsets due to wear.
The benefits of lubrication include: 56
•
Reduction of gauge face wear in rails and flange wear
in wheels, resulting in a decrease in the rate of
replacement of wheels and rails.
•
Reduction of fuel/energy consumption associated
with wheel/rail interaction.
•
Reduction in noise associated with wheel/rail
interaction.
The optimal lubrication pattern will be the one that offers
savings in energy along with wheel and rail wear and
replacement, without causing harmful side effects to curving
and train handling.51
Types of lubrication. There are generally three types of
lubrication systems:
1. Hi-Rail based mobile lubricators. These lubricators
apply lubricant to the gauge corner of the rail. There
are many designs of applicators, from rail based
automobile units to special cars.
2. Locomotive flange lubrication. Locomotive flange
lubricators apply grease or solid lubricant to the wheel
flange, which is then transferred to the gauge side of
o 3-30 x
the railhead. The lubricant application is usually
controlled on a wheel revolutions basis. The lubricant
may be applied to either or both wheels. Grease is
applied to the wheel through a nozzle. To avoid
misaligned nozzles, an automatic control system is
used. Wheel flange lubricators for solid lubricants use
different application systems. For instance, solid
lubricant in rolled tubes are applied under pressure to
the wheel flange by a special device.
3. Wayside lubricators. There are three general types of
wayside lubricators: (1) mechanically, (2) hydraulically,
or (3) electrically driven. They all use various types of
greases. These lubricators are installed on main line
curves and at the entrance of stations. The
effectiveness of wayside lubrication is affected by the
location of the lubricator in regard to curve, the
viscosity of the grease at different (including negative)
temperatures, and the level of maintenance. Use of a
tribometer (see below) is a good practical method for
optimal lubricator location which often depends more
on the speeds and types of the trains, than on the
actual curve geometry. As temperatures in northern
countries change significantly during a year, summer
and winter types of greases may be used. During the
season, the right viscosity grease should be used.
It is very important to assure maintenance of wayside
lubricators. Common problems are improper adjustment of
applicators, leaking or failed grease hoses, or failed or
ineffective pump mechanisms. Standardization of lubricators
within maintainers territory and accessibility to lubricator
should be achieved. .
Types of lubricants. Depending on climates and environment,
there are certain requirements for lubricating materials. These
requirements should recognize the need to provide lubricating
ability up to certain contact pressure and should easily be
applied to lubricating surfaces under conditions of increased or
reduced humidity.50, 52 They should not be washed off by rain,
and they should withstand the ambient temperature from –40
to +70oC. Also, the lubricant must be capable of withstanding
o 3-31 x
the temperatures that are generated by wheel/rail interaction
and braking. Sand used to increase adhesion should not adhere
to the lubricating material after its application. It should not be
flammable, hazardous, or produce a harmful effect upon
environment. All lubricating materials intended for lubrication
in rail/wheel system should be certified according to the
appropriate country’s regulations.
Evaluation of lubrication effectiveness. For mobile rail
lubricators, there are three indices that are controlled monthly:
(1) average distance of regularly lubricated track sections, (2)
amount of lubricating materials used per one kilometer of
lubricated track, and (3) number of service runs.50 The values
of these indices depend on the characteristics of particular
railway sections, volume of traffic, type of lubricant, and
applicator design features. The measured indices are compared
to those assigned for the particular railway section, rail
lubricator, and lubricant used.
A very efficient instrument for study adjustment and
maintenance of lubricators is a tribometer (see below).
However, in operation, it is not possible to dedicate the labor
or time required to use a tribometer on a daily basis. Some
railways have developed methods of evaluation of lubrication
effectiveness based on observed conditions of rail surfaces,
including visual observation of the presence of wear debri, in
the form of expert ("eyeball") chart (Table 3.4). After using a
tribometer for a short period and summarizing the results of
visual observations, by reference to this chart, a relatively good
estimate of the coefficient of friction can be made.
o 3-32 x
Table 3.4: Expert Chart of Lubrication Effectiveness
OBSERVED CONDITIONS OF RAIL
GAUGE FACE SURFACE
53,45
EVALUATION
OF THE
COEFFICIENT
OF FRICTION.
Rough, with gouging marks of material seizure
about 0.6
Chewed up, rough
0.45 to 0.6
Smooth, with shiny unlubricated surface
0.35 to 0.45
Smooth, with lubricant covering 10 to 40% of
the surface
Smooth, with lubricant covering 40 to 60% of
the surface.
Smooth, with lubricant covering 60 to 90 % of
the surface
Thin film of lubricant covered 100 % of the
surface.
100% of surface covered by thick, black film of
lubricant
0.30 to 0.35
0.25 to 0.30
0.20 to 0.25
0.15 to 0.20
< 0.15
Another important criteria of lubrication effectiveness is
locomotive fuel or electrical energy consumption by the
locomotive fleet. There are methods to evaluate electric energy
savings due to lubrication of the locomotive fleet operating on
the particular railway section. 54 Energy savings in revenue
service on mountainous railway sections with many sharp
curves is from 6% to 12%. Energy is saved because of decrease
in friction forces at wheel flange and gauge face of the railhead,
a decrease of components of the relative slippage, and
improvement of truck dynamic behavior.
The major criteria of lubrication effectiveness is the wear
rates of wheels and rail. For instance, Russian Railways
developed norms for the wear rate of locomotive wheel
flanges in mm per 104 km of the locomotive run. These norms
depend on curves, track grades, and types of locomotive
operation. Norms on the wear rate of gauge side of the
railhead, in mm per mgt, depend on the rail and wheel steel,
and curve radius (Table 3.5)
There is also a norm on economic efficiency; i.e., the
reduction of operating expenses due to lubrication. For
instance, to be economically justified, the reduction of
o 3-33 x
operating expenses should be three times more than the cost
of lubrication.45
Table 3.5 Norms on the Railhead Gauge and
50
Wheel Flange Wear Rate
ASSETS
WEAR RATE
Railhead gauge
side wear rate,
mm/mgt
<0.66,R≤300m
≤0.05,R=300-500
≤0.04,R>500 m
≤0.025,R≥1000 m
Locomotive wheel
flange wear rate,
4
mm/10 km
≤0.45, R≤650m curves share 10%
≤0.55 for mountainous regions
3.4.2.1 Measurement of the Coefficient of Friction
Hand tribometer: 55 A tribometer has been developed by the
AAR to measure lubrication effectiveness in the field. The
tribometer was able to measure gauge face, gauge corner and
top of rail coefficient of friction. The friction forces in the
measuring wheel’s contact patch are determined by a braking
process quite similar to the automatic braking system found on
many new automobiles. The tribometer has an automated
control system that senses the point of saturation of
longitudinal creep curve. The microcomputer embedded in the
measurement head can determine the point of maximum
adhesion (creep force).
High-speed tribometer:51 A high-speed tribometer is a new
option for North American railroads. Unlike the portable
tribometer, high-speed tribometer measures the point of
saturation the creep curve by inducing lateral creep. Mobile
tribometers make it possible to measure the coefficient of
friction over long stretches of track, at speeds up to 50 km/h.
Data is collected on friction values of the top and gauge face of
both rails simultaneously and stored in a database. The data
serves various purposes; e.g., adjusting placement of wayside
lubricators and maintaining the rail lubrication system.
o 3-34 x
3.4.2.2 Problems of Lubrication
1. Excessive lubrication can provide added fuel savings, but
come at the expense of rail fatigue and decreased truck
curving performance leading to lateral loads (Figure 3.12).
2. Excessive lubrication, migration of the lubricant to the
railhead or wheel tread surface can cause wheel slip, which
reduces locomotive traction limits and can cause wheel
skid flats and rail wheel burn.
3. Human factor — overcoming the tendency for locomotive
engineers to turn off locomotive flange lubricators as a
measure to prevent locomotive wheel slip.53
4. The contamination of the track, which occurs particularly
near track mounted lubricators.56
5. Large difference in the effectiveness of lubrication of high
and low rails may promote rail rollover in curves under
conditions where bogie bolster center bowls and low rail
are inadequately lubricated.51,57
6. When the turning moment of three-piece bogies is very
high because of improper conditions of the center plate,
side bearings, and their supports, high rail lubrication will
tend to increase warping of the bogie, thus increasing the
angle of attack. When using high rail lubrication, action
should be taken to keep center plates and supports in good
condition.3
7. The adverse influence on the bogie steering characteristics
and the lateral wheelset forces occurs when the running
surface coefficient of friction of one rail in a curve
becomes very different (more 0.1-0.15).56
8. Hollow worn wheel threads can negate many benefits
achieved by lubrication.
9. The enhancement of the contact fatigue growth on the
gauge corner of rails and running surfaces of wheels.
o 3-35 x
3.4.2.3 Recommendations on Lubrication Practices
•
Form a highly structured lubrication management
program, when lubrication is introduced and maintained.
•
Establish indices of lubrication practices, designed to meet
the needs of particular railway section.
•
Establish norms for optimum lubrication of the wheel
flange and gauge face of the railhead, adjusting wear rates
on the number and radius of curves.
•
Maintain a minimum level of lubrication to assure a
smooth surface of the gauge corner of the railhead. Once
this surface becomes rough, the effect of subsequent
lubrication is greatly reduced.
•
Avoid over lubrication, particularly ahead of curves and
signals on upgrades.
•
Choose appropriate lubricant needs to take into account
the temperature range prevailing on the Railway
Standardization of lubricators within the lubricator's
maintenance territory.
•
Train all personnel involved in lubrication to reduce
human errors.
•
Keep center plate, side bearings, and supports of threepiece bogie of freight cars in good condition, when using
high rail lubrication.
•
Use the optimal lubrication pattern to save energy and to
reduce wheel and rail wear, without causing negative side
effects in curving and train handling; i.e.,economic benefits
of lubrication have to be balanced by the potential adverse
effects of lubrication.
•
Do not look at lubrication and grinding programs
separately. Form a program based on a systems approach.
3.4.3 Friction Control and Management
The coefficient of friction as it relates to railway operation falls
into three general categories: (1) low friction — 0.2 or less, (2)
intermediate friction — 0.2 to 0.4, and high friction — 0.4 or
greater.58
o 3-36 x
The objectives of friction management are to maintain the
coefficient of friction:
•
at wheel flange/rail gauge interface at the low level,
•
at the wheel tread/top of rail interface of freight cars
at the intermediate level, and
•
at the wheel tread/top of rail interface of locomotives
at high level.
Friction management is a concept that allows, by careful
selection of specific materials (friction modifiers), the
development of layers which obtain desired wheel/rail
frictional characteristics.
3.4.3.1 Friction Modifiers
Friction modifies are materials that are added into the
composite layer in wheel and rail interface to create a third
body with desired properties.
They can be divided into three categories:58
1. Low coefficient of friction modifiers (LCF), with
coefficient of friction 0.2 or less are used to reduce
friction at the wheel flange/rail gauge face interface.
Solid lubricants are examples of these types of friction
modifiers. The major difference when comparing
liquid or grease types of lubricants is that solid
lubricants, under the same pressure and relative
slippage, form thicker layers; i.e., 10–30 microns;
grease lubricants form layers of less than 5 microns
thick. This is one of the properties of solid lubricants
to form unextruded layers under given pressure and
relative slippage
2. High friction modifiers with intermediate ranges of
coefficient of friction (from 0.2 to 0.4) are used at the
wheel tread/top of rail interface to reduce rolling
resistance of freight cars, to combat growth of short
pitch corrugation, to reduce hunting, or to eliminate
wheel squeal.
o 3-37 x
3. Very high friction modifiers (friction enhancers) are
applied to increase locomotive adhesion and promote
braking effort.
All friction modifiers can be classified according to their
behavior after creepage saturation (Figure 3.13).60 If traction
decreases after saturation point, then the modifier has negative
friction characteristics. If traction increases after creepage
saturation, than the modifier has positive friction characteristic.
Depending on the rate of increase, the friction modifier is
classified as high positive friction (HPF) or very high positive
friction modifier (VHPF).
Figure 3.13: Model of Behavior of Friction Modifiers
HPF represents a family of materials that generate positive
friction characteristics in the third body layer. HPF materials
are one of the inputs in the third body interfacial layer flow
(see also 3.5.3), which can override the frictional characteristics
of the contaminants and change the characteristics of the
frictional pair from negative to positive.
HPF can be provided in either solid or liquid form,
depending on the requirements and application method.60 Solid
sticks are suitable for onboard use applied directly to the wheel
tread. From the tread, the material transfers to the rail, as the
resin material is selected to burn off under the high
temperature at the wheel-rail interface. This leaves a thin
micron scale film of the dry friction modifier, which fills in the
aspirates in the metal surfaces.
o 3-38 x
The HPF modifier is also available as a water-based liquid,
which is applied directly to the top of rail by either a hi-rail
vehicle or from the train.
3.4.3.2 Top-of-Rail Lubrication51,59
Top-of-rail lubrication friction management is intended to
further obtain energy savings and improve curving
performance of freight car trucks. Intermediate types of
friction modifiers are used and maintain the friction coefficient
at 0.35. Lubricants (friction modifiers) are applied to the top of
both rails directly behind the last locomotive. The amount and
characteristics of a lubricant should be such that most of the
lubricant is consumed by the end of the train, thus preventing
the slip of the locomotive wheels of the next train. The
coefficient of friction should range from 0.25 to 0.45 after the
last car.
Field simulated tests have shown that that average energy
savings ranged from 10 to 13 %.59 Curving forces were reduced
from 5 to 45 %, depending on curvature and the car type, and
noise levels were also reduced. Self-steering (premium) trucks
exhibited no improvement in performance.
Top-of-rail lubrication is still in the testing stage. Tests
have shown that top-of-rail systems of lubrication require
significant monitoring and adjustment. Full service braking
tests are required to evaluate performance of brakes under
conditions of use of top-of-rail lubrication. If friction
modifiers used to keep friction at the level 0.2-0.4 do not
migrate, they may work together with low friction modifiers
(lubricants) used for wheel flange/rail gauge face lubrication.
3.4.3.3 Locomotive Adhesion Enhancers
Sand, which is widely used to increase locomotive wheel
adhesion, is a simple material, but it creates many complex,
costly problems starting from maintenance of sanders and
repair of abrasive effects on wheels and rail and extending to
damages inflicted upon many components of the track
structure. VHPF modifiers, which were intended to reduce or
eliminate sanding, enhance the coefficient of friction to the
0.4-0.6 range. They are necessary for high power and high
adhesion locomotives that are required for heavy haul
o 3-39 x
operations. Many companies offer different types of these
modifiers. For instance, one VHPF is based on a thermosetting
polymer stick that is applied directly to the wheel tread of
locomotives.
Locomotive adhesion enhancers are still at the test stage.
Prior to heavy haul revenue service application, they should be
tested and if necessary improved to determine:
•
train breaking performance,
•
train rolling resistance,
•
tribology characteristics of friction enhances; in
particular, the time delay before the coefficient of
friction increases after application of the friction
enhancer to the wheel tread, and
•
track spreading forces.
Also, revenue service application requires the development
of a computer controlled dispensing system.
The self regulated friction-management system closest to
use utilizes a friction sensing and computerized dispensation of
friction modifiers, so that friction would remain within the
optimal range at all times.
3.4.3.4 Recommended Rail Friction Guidelines51
Field data and computer simulation make it possible to suggest
the following performance issues to be considered for best
practices in wheel/rail system friction control:
•
Maintain top of rail friction coefficient to a difference
of less than 0.10-0.15.
•
Maintain top of rail friction coefficient at a level
greater than 0.35.
•
If the high rail becomes lubricated, low rail lubrication
is advised to decrease the tendency of developing high
curving forces.
•
Maintain gauge face friction coefficient at a level less
than 0.25-0.30.
o 3-40 x
•
Lubrication of only the top of the rail can lead to poor
truck performance if lubricant migration occurs.
These guidelines are subject to further validation and
correction under heavy haul revenue service operating
conditions.
3.4.3.5 Use of HPF to Control Short Wave (Pitch)
Corrugation60
One of the initiation mechanisms of short pitch corrugation is
stick-slip oscillation due to negative friction under saturated
creep conditions (see Section 3.5.5). Use of HPF to alter the
friction at the wheel/rail interface from negative to positive
enables the alleviation of the formation of short wave
corrugation.
3.4.3.6 Use of HPF to Control Noise
Introducing HPF materials in the wheel/rail interface
overcome the negative friction characteristics that lead to stickslip and wheel squeal (see also Section 3.5.6).
3.4.3.7 Possible Technique of Friction Management
One possible technique of friction management is to use
onboard systems that can supply different friction modifiers
(including lubricants) and apply the required modifier to the
specific position of the rail on the line to reduce wear,
corrugation, or wheel squeal.61 The system is activated to
apply lubricant from wayside transmitters.
3.5 Rail and Wheel Damage Modes;Mechanisms
and Causes
There are numerous types of rail and wheel defects. Damage to
wheels and rails normally are specified in an individual
railroad’s manual. These documents give the defect
classification, coding description of a defect, its causes, and
instructions for action to address the problem. In this
handbook major wheel/rail damages predominantly resulting
from heavy haul operation are described and discussed.
Appendix A presents the Canadian Pacific Rail Defect
Classification with reference to the Russian Railways rail defect
o 3-41 x
coding. In Appendix B wheel defects and their coding are
presented.
3.5.1 Wear
There are two main areas of wheel and rail wear. The first one
is the top of rail and the wheel tread. The second is the gauge
face and wheel flange wear, mainly in curves. Wheel and rail
assets limits are given in corresponding railways manuals.
Wear modes and transition values: Rail and wheel wear is
generally assumed to be proportional to the energy dissipated
overcoming wheel and rail rolling resistance. As Figure 3.2
shows, wheel and rail wear is determined by the relative
slippage λ (see Figure 3.1) and stresses p at contact patches. In
turn, the relative slippage and stresses depend on dynamic
parameters of wheel/rail interaction. Wear is highly dependent
on the third body properties, which are strongly influenced by
lubrication, environment conditions (humidity, rain and snow),
and presence of sand.
Based on laboratory wear simulated tests in unlubricated
conditions, three major wear modes have been defined: (1)
mild, (2) severe and (3) catastrophic.37,3 Wear modes are
characterized by different wear rates, surfaces, and wear debris
form and size.
The wear diagram; i.e., pλ =const curves, that represents
zones with different wear modes and areas of normal and
abnormal performance is shown in Figure 3.14. This diagram
was obtained for conventional rail and wheel carbon steel of
initial hardness up to 300 HB. The curve pλ=40 is the
boundary between the mild and the severe wear modes,
whereas pλ=120 is the boundary between the severe and the
catastrophic wear modes.
The mild wear mode is characterized by the bright surface of
rollers and by the large thin metallic flakes of 1000 µm in size
and 3 µm thick, which were formed at the surface of the roller.
o 3-42 x
ÌÀÕ. PRESSURE
p , MPa
3000
2500
pλ=120
λ=120
2000
1500
2
1000
pλ=40
λ=40
500
1
0
0
0 ,02
0 ,04
0 ,06
R E L A T IV E S L IP P A G E
0 ,08
0 ,1
λ
Figure 3.14 Wear Modes Diagram for Wheel/Rail Steels
1. field of normal performance
2. boundary of abnormal performance
3
The severe wear mode is characterized by a rougher
surface. The wear rate grew intensively with increase of
pλ parameter. The wear debris was a bright flake up 500 µm in
size and 15-30 µm thick. The mean size of worn particles grew
with pλ increase.
Under the catastrophic wear mode, both worn surfaces are
very rough and show prominent score marks (Figure 3.15).
The wear debris is of different size. The larger particles are
up to 300 µm in size and 50 µm thick . The smaller spherical
particles are up to 10 µm in size.
Between severe and catastrophic wear modes, a heavy wear
mode was discovered.45 For the heavy wear mode, the
maximum and the minimum ratio of wear may differ by one
order of magnitude during a test run. The wear rate decreases
with the growth of the pλ parameter. The severe and the heavy
modes can exist at the same level of pλ and may transform into
each other. The appearance of the surface work hardening in
the heavy wear mode is closely related to the rate of the surface
work hardening processes.
o 3-43 x
Top of rail and wheel tread wear:25;62 Normally, the top of rail is
subject to high contact stress (about 1300-1700 MPa,
depending on the axle load) and relatively low (less than 0.010.015) levels of relative slippage (if there is no slip of the
locomotive or car wheels as a result of low adhesion between
wheel and rail or excessive tractive or braking effort.) In this
case, the parameter pλ is about 20 and the mild wear mode of
oxidative origin is predominant. The wear particles are mixed
with various environmental contaminants that compose a third
body layer on the top of rail. The rail wear rate is determined
by the frequency of removal of this layer from the surface and
the composition of the layer. Abrasive particles cased by sand
from locomotives or from other sources increase the wear rate
two to three times.
Over time, if not ground, rail crown causes the wheel tread
to become hollow due to wear (Figure 3.16). The hollow shape
forms when tread wear is high, and the wheel forms a false
flange at the ends of tread area:
h
Figure 3.16: Hollow Worn Wheel Profile
h — depth of hollow
Wheel hollowing:63
•
•
increases car rolling resistance and fuel
consumption
creates a false flange which causes surface damage
of rails, switches, frogs, and crossings
o 3-44 x
•
•
increases the lateral forces on the rails in curved
track, increasing track deterioration and the risk of
derailment
increases the shear stress acting toward the field
side of the low rail and the incidence of flaking
damage
Railhead gauge face64 and wheel flange wear: This wear takes
place mainly when a truck or bogie is negotiating a curve,
though it may occur for a short time on tangent track,
especially if cars are hunting. In sharp curves under dry
conditions, conventional three-piece truck negotiation lead to
the catastrophic wear mode of high intensity, resulting in a
large amount of wear particles deposited on the track (Figure
3.17) and quick change of the railhead profile (Figure 3.18).
Figure 3.17: A Large Amount of Wear Particles Deposited
on the Track in the Sharp Curve Under
Unlubricated Conditions
o 3-45 x
Figure 3.18 Worn Railhead
For mild and severe wear modes, wear rate (I) may be
considered as a liner function of tangential force (T) on the
Hertzian contact area (A) and relative slippage λ as:37
(3.2)
I=Tλ/ A
Recent laboratory simulated tests of wheel flange/rail
gauge wear have shown3;45 that particularly for severe and
catastrophic wear modes in unlubricated conditions and at a
relatively stable coefficient of friction the specific volume wear
rate in the dimensionless form can be expressed as :
I= k p λ 2,
(3.3)
where p is the contact pressure at the corresponding point
of the contact patch, λ is the relative slippage varying from 0
to 1 and k is the coefficient based on wheel and rail material’s
characteristics and the coefficient of friction.
As the relative slippage λ depends on the angle of wheel to
rail attack, which in turn for standard three-piece trucks
depends on radius of curves, then the sharper a curve is, the
greater is the wheel flange/rail side wear.
Depending on whether the wheels and rail have new or
worn profiles, there may be different contact pressures and
o 3-46 x
different relative slippage distributions on the contact patches
that will lead to different wear modes and rates of wear. In
unfavorable combinations, the contact pressure may reach
3000 MPa and the relative slippage may range from - 0.06 to0.1. Thus the pλ parameter may be over 300 which represents
the catastrophic wear mode of high intensity.
Gauge wear of the railhead takes place with a large amount
of plastic flow. Plastic deformation starts when the maximum
contact stress reaches a critical value, depending on material
properties, particularly the hardness as influenced by work
hardening. Plastic flow depends on tangential forces and the
coefficient of friction. For instance, at the coefficient of
friction of 0.6, plastic flow develops from the surface under
considerably lower contact stresses than when the coefficient
of friction is small. It was assumed that when the average
contact stress is greater than one third of the surface hardness
(measured in MPa), wear takes place with a large amount of
plastic deformation and will rapidly increase with the growth of
contact stresses.65
Factors influencing wear. As it is seen from formulae 3.2 and
3.3, wear rate is defined by the contact pressure which
depends on vertical and lateral force, rail/wheel profile and the
relative slippage, which in turn depends on axle load, dynamics
of vehicle/track interaction, and wheel and rail profiles. Other
important factors are third body properties and material
characteristics. Among the material characteristics affecting
wear resistance are material hardness, carbon content,
microstructure, and material cleanliness, particularly sulfide
content.
Influence of axle load: Field simulated tests have shown that
rail gauge face wear rate and railhead height loss (vertical wear)
rate increase when the axle load increased from 210 to 270 kN.
66 Increasing the axle load from 294 to 347 kN67;68 (the trucks
were of the same three-piece design) resulted in increased high
rail gauge face wear rate of a 350 m curve under dry operating
conditions from 0.8 to 1.3 times and from 1.2 to 1.7 times
under lubricated (contaminated) operating conditions,
depending on the rail metallurgy. The same increase in axle
load resulted in high railhead height wear under dry operating
o 3-47 x
conditions which varied from 1.3 to 3.3 times and from 1.4 to
1.9 times under lubricated (contaminated) operating
conditions. Low railhead height wear rate varied from 0.3 to
1.1 and from 0.9 to 1.1 correspondingly for dry and lubricated
(contaminated) operating conditions. Rail wear on tangent
track to a great extent is a function of the static axle load.
It should be noted that increases in wear associated with
higher axle load may or may not be economically significant ,
depending on a range of prevailing factors, including rail and
wheel maintenance procedure, rail/wheel profiles, rail and
wheel material types, bogie characteristics, and lubrication
practices.
Influence of track gauge clearance: Track gauge clearance
(clearance between wheel flanges of a wheel and a railhead
measured at determined level) increases wheel flange wear
considerably when it is less than a certain value depending on
the particular railway and operating conditions. 71, 72
Influence of wheel and rail material hardness. Hardness still
remains the most useful practical single property to
characterize material characteristics for wear studies. The most
common methods of increasing the hardness of pearlitic steels
are by increasing carbon content and by refining the
microstructure.
The influence of wheel and rail hardness is highly
dependent on the wear mode. Laboratory tests under nonlubricated conditions have shown that at a given contact
pressure and relative lateral slippage characteristic for the gauge
zone of the railhead and wheel, an increase in hardness from
HRC 30 to 50 resulted in not more than an increase of two in
the wear resistance. In the severe and the catastrophic wear
modes, the same increase of hardness resulted in up to several
times the reduction in wear rate. In the catastrophic and heavy
wear mode, the same change in hardness may result in
reduction of the wear rate up to two orders of magnitude.3
Field tests67 show that under an axle load of about 300 kN
in a 350 m curve, under non-lubricated conditions, a variation
of rail initial hardness from HB 280 to HB 380 decreases the
rail gage face wear and head height loss rate up to several
o 3-48 x
times. The field test results on the effects of rail hardness were
also reported in Reference 70.
Influence of the difference in wheel/rail hardness. Field simulated
tests64 confirmed by laboratory study3 in the range of rail to
wheel hardness ratio (HR/Hw) from 0.7 to 1.6, show that there
is no optimal rail to wheel hardness ratio providing for minimal
total wear of both wheels and rail. The wear rate of each asset
is inversely proportional to its hardness by the relation n=4-6,
that is
I
≈ (HR/Hw)n
(3.4)
Increasing the hardness of either the rail or the wheel
resulted in a decrease of the total wear rate. Laboratory studies
performed in many countries73;74 have clearly shown that an
increase in the hardness of either component will result in the reduction of
the wear rate of both components.
Influence of material structure:
•
Considerable improvement of the wear resistance of
pearlitic steels could be achieved by increasing the
carbon content (up to 0.9% ).76 Increase in the wear
resistance is explained by increase of the rolling
contact surface hardness rate (work hardening ability)
due to the refinement of microstructure of the surface
layer in the work hardening process.
• The lamellar form of carbides is more wear resistant
than the granular form.
• Another parameter is the quantity of sulfide inclusions
determined by the sulfur content and the shape of
these inclusions as determined by the alloying
elements in steel.
Working hardening: After multiple wheel/rail interactions in
operation, the hardness of the interacting surfaces increases by
comparison with the initial surface hardness. The
microhardness of a wheel flange surface may increase up to
HV0.1 600 - 800. The thickness of the hardened layer is several
tenths of a millimeter and does not exceed 0.5 mm. Wheel
tread increment in surface hardness is up to HV0.1 150-170
compared with the initial surface hardness, and the hardened
layer thickness is several millimeters thick.42
o 3-49 x
Work hardening has been observed on the running
surfaces in both standard and head hardened rails.75 The
considerable depth of the hardened layer (about 4-8 mm
depending on the rail type) develops rather quickly, but the
maximum hardness, which is achieved near the surface,
increases gradually with passed tonnage (Figure 3.19). Since the
work hardening depends on the degree of material
deformation, standard rails will work harden more than head
hardened rails. Thus, if well maintained, the standard carbon
rails will eventually exhibit similar harness near the rail surface
as the head hardened rails. However, this is not the evidence of
equal fatigue and wear resistance, which are higher for head
hardened rails.
460
440
After 50 MGT
Work Hardened Region
420
After 200 MGT
400
After 400 MGT
380
After 800 MGT
360
After 1000 MGT
340
320
300
Base Hardness
280
260
240
0
2
4
6
8
10
12
14
16
18
20
22
Depth from Contact Surface (mm)
Figure 3.19: Vickers Hardness Distribution in Standard Carbon
Steel Showing Work Hardening Developed in Tangent
75
Track at 30 to 35 Tonnes Axle Loads
When rails are ground or wheels are turned, the natural
self-hardened layers are removed and a new running-in process
starts, which leads to higher wear rates than are observed when
surfaces are work hardened.
Though work hardening always takes place in operation, it
is very difficult to control this phenomenon and to use it to
o 3-50 x
decrease wear because the work hardened layer is not stable.
One hypotheses suggests that the optimal surface hardness
should be equal to the working hardness.42
3.5.2 Recommendations to Decrease Wheel and Rail
Wear
• Metallurgical
Rails. use alloyed, fully heat treated, head hardened and
micro-alloyed steels which have hardness in the range of
340 to 388 HB and a fine pearlite structure.
Wheels: use wheel steels with similar carbon content and
alloying, and thermally hardened wheel flanges (plasma
quenching, electro-arc depositing).
! Lubrication: The presence of properly applied lubrication
on the gauge face of high rail or wheel flange causes the
wear rate of both assets to drop considerably, particularly
in sharp curves.
! Wheel/rail profiles: Establish and maintain different types of
rail wheel profiles for curves and tangent railway sections.
Preferably a conformal profile which decreases the contact
pressure
• Track structure:
- Set optimal gauge clearance (clearance between wheel
flanges of a wheelset and railhead) and its tolerances in
curves depending on the specific conditions of a
particular railway
- Maintain a track structure according to norms
• Truck (vehicle) structure: Decrease wheel to rail angle of attack
and thus the relative slippage of the components through
- the use of self steering (radial) trucks,
- maintaining conventional three-piece trucks and their
elements according to norms,
- monitoring the level of turning torque of a truck relative
to car bogie, and
- developing wayside measuring systems and criteria of
detecting cars which cause excessive assets wear.
o 3-51 x
3.5.2.1 Optimal Rail Wear Rate
There exists an optimal wear rate in which fatigue and wear are
in balance. The optimal wear rate is obtained when the surface
material wears just enough to prevent small fatigue cracks from
developing in the rail and propagating to become detail
fractures. The optimal wear rate depends on differences in
traffic type and density, axle load, rail metallurgy, and track
curvature. On the average, an optimal wear rate is estimated as
about 0.02 mm/mgt.78,79 To provide the optimal wear rate, it
is necessary to further develop and introduce friction
management technology and wear monitoring facilities.
3.5.2.2 Rail Wear Condemning Limit
There are two major approaches to the establishment of wear
condemning limits:
1. Railhead cross section loss in percent.
2. Value of railhead gauge (side) wear measured at
defined level and value of reduced wear; that is,
vertical (railhead height loss) plus half of the side wear
value.
Many railways use stress based wear limits approach. Rail
internal stresses are largely related to the depth of material left
in the railhead. A critical condition appears where the height of
the rail is reduced to such extent that the influence zone of
contact stresses and the stresses of railhead and web bending
coincide.
Values of rail condemning limits are defined based on the
type of rail, rail fatigue resistance, derailment criteria, and
railroad operating conditions (see Part 5 “Optimizing
Wheel/Rail Performance”).
3.5.3 Rolling Contact Fatigue Defects
Rails: Railways of IHHA countries are using several defect
classification systems. Rail defect classification used by the
Canadian Pacific Railway with reference to the Russian
Railways defect coding are given in the Appendix A.
Shelling:19, 24, 77 Shells occur at the gauge corner of the high
rail in curves on railways with axle loads over 200 kN (Figure
o 3-52 x
3.20). An elliptical shell-like crack with characteristic crack
growth rings propagates predominantly parallel to the rail
surface. In many cases, the shell causes metal to spall from the
gauge corner. Under certain circumstances, shells can lead to
transverse failure of rail. When the crack length reaches the
critical value (about 10 mm), the trailing edge of the crack may
turn down into the rail, giving rise to fracture of rail (Figure
3.21). Shells and the associated transverse defects represent the
principal fatigue defect of concern on heavy haul lines.
Figure 3.20: Deep Rail Shelling on the Gauge Side Of
77
the Railhead
Figure 3.21: Rail Transverse Defect resulted
77
from Gauge Shelling
o 3-53 x
The shell is a subsurface initiated defect. It initiates under
high contact initiated defect tangential stresses predominantly
at stringers of oxide inclusions in the rail steel. The critical
metallurgical factors contributing to initiation of shells are the
oxide volume fraction in percent, the stringer’s length and the
Brinel hardness (HB). These factors form a shell index38 as:
Shell Index = (Oxide volume x Stringer length)/ HB2
(3.5)
The higher the shell index, the greater is the shell defect
rate.
Shells may turn to transverse fractures initiated at hard
brittle oxide inclusion strings. Therefore, transverse defects can
be expected to decrease with the increasing use of clean steels.
Mechanical factors can also contribute to the development
of shells and transverse defects. These are normal, lateral, or
traction loads and residual stresses.
Grinding is the most generally used method to eliminate
surface defects, including spalls. The use of grinding to reduce
transverse defects is less clear. Tests on FAST have shown that
grinding to promote two point contact on high rail significantly
reduced the number of shells, but most shells turned into
transverse defects. However, many railways use rail grinding as
a means to reduce transverse defects.
The detection of small transverse cracks developed from
shells is difficult as the horizontal component of shell can
mask the vertical transverse crack during ultrasonic inspection.
Influence of axle load: Laboratory and field simulated tests
and application of linear fracture mechanics, have shown that
an increase of the axle load results in a reduction of the time
before the appearance of the first fatigue cracks and in a
reduction of the depth of longitudinal cracks and in the critical
size of the transverse cracks which may lead to cracking of a
rail.66 For instance, an increase of axle load from 210 to 270
kN resulted in the depth of the contact fatigue defect growth
from 3 to 7 mm up to from 6 to 9.5 mm for the same time of
test operations.
o 3-54 x
It was found that the rate of transverse defect
development is proportional to the ratio of axle load to the
power of 2 in new rails and 3.3 in worn rails.75 However, the
increased expected rate can be controlled by the appropriate
design of wheel and rail profiles to improve the wheel/rail
contact characteristics.
3.5.3.1 Recommended Practices to Eliminate Shelling
• Utilization of head hardened rails that have been
manufactured by clean steel making processes and are
therefore free from inclusion stringers
•
Reducing and redistributing the load on the railhead by
- using conformal or near conformal wheel and rail profiles
- using asymmetric railhead profile grinding in mild curves
(about 900 m radius and greater)
- grinding the gauge corner of the rail.
•
Keeping the rail wear close to optimal where the surface
wear rate is enough to prevent micro cracks from
propagating
•
Reduction of lateral loading by improving the curving
performance through the use of steering and self-steering
trucks.
3.5.4 Head Checks19,81
Head checks generally occur on the gauge corner of the high
rail in curves with a radius from 1000 to 1500 m, when rail
wear rate is rather low (Figure 3.22). Head checks also may
occur in tighter (less than 1000 m ) curves near the gauge
corner of the high rail. Flakes similar to head checks, which are
usually slightly inclined or parallel to the direction of travel, can
also be found on the field side of a low rail. Head checks may
branch up towards the surface of the rail, giving rise to spalls.
On heavy haul lines, head checks may lead to rail fractures.
These fractures are prevalent under conditions of high thermal
or residual stresses. Head checks, in addition, lead to the
development of corrugations.
o 3-55 x
Figure 3.22: Head Checks on Gauge Corner
of Railhead
Head checks initiate at the rail surface as a consequence of
large unidirectional plastic strains. These strains arise from an
incremental accumulation of plastic strain during each cycle of
loading of the strain hardened material. Accumulation of a
large number of unidirectional plastic strain increments
“ratchet” the surface layer of material until its ductility is
exhausted. Microscopic cracks then develop at a small angle to
the surface (Figure 3.23).
Figure 3.23: Ratcheting in Rail Steels Associated
25
with Contact Fatigue
o 3-56 x
The critical conditions for initiation of cracks at the rail
surface are a high ratio of the normal Hertzian contact stress to
the material shear strength and a high ratio of the tangential to
the normal load (T/N), as seen from the “shakedown” diagram
(Figure 3.11). Cracks initiated by ratcheting grow perpendicular
to the prevailing direction of the traction force. In mild curves,
longitudinal traction is dominant and cracks grow primarily
perpendicular to the direction of travel. In sharp curves
traction forces are primarily in the lateral direction and cracks
are primarily parallel to the direction of travel. For mixed
longitudinal and lateral traction, the cracks can grow at an
angle about 45 degrees to the direction of travel.
Water and lubricants trapped in the crack increase the
speed of crack propagation.80
3.5.4.1 Recommended Measures to Decrease Head
Checks
•
Use of head hardened and premium rails that raise the
threshold at which ratcheting and crack initiation occurs
•
Redistribution of the number and intensity of wheel
contacts by profile grinding the rail
•
Use of improved and self steering trucks to decrease
traction forces
•
Lubrication of the gauge corner to reduce traction forces,
although care is required since lubrication can also
exacerbate crack growth
•
Use of rail grinding
3.5.5 Tache Ovale19 (Shatter Crack from Hydrogen)
These are defects which develop about 10-15 mm below the
railhead from longitudinal cavities caused by the presence of
hydrogen. These cavities may exist in the parent rail steel or
they can arise in welds from poor welding practice. Transverse
fracture occurs under a dynamic load when the crack becomes
sufficiently large (Figure 3.24). Development of tache ovale is
influenced by thermal and residual stresses from roller
straightening.
o 3-57 x
77
Figure 3.24 Rail with Tache Ovale Defect
3.5.5.1 Recommended Practices to Decrease the
Defect
• Improvement of steel quality by reduction of hydrogen
content in the rail steel (vaccum treatment of melted steel
and special heat treatment)
•
Control of welding procedure to prevent water entrapment
in rails.
3.5.6 Squats19,81
“Squats” occur on tangent track and in curves of large radii on
the rolling surface of the railhead and are characterized by the
darkened area on the rail (Figure 3.25). Squats are surface
initiated rolling contact fatigue defects. Each squat consist of
two cracks, a leading one which propagates in the direction of
travel and a trailing crack, which propagates in the opposite
direction and is several times longer than the leading track and
contains numerous branches. One of these branches may
propagate transversely across the rail. Ultrasonic inspection of
these defects is difficult since the transverse defect is shielded
by the shallow, horizontal primary crack.
o 3-58 x
Figure 3.25: Squat on Running Surface of Rail in
71
Tangent Track
Such cracks may initiate from a white etching martensitic
layer on the surface of the rail. Other mechanism of squat
formation are linked to the longitudinal traction of locomotive
wheels, whereby a surface layer of rail material ratchets until an
individual crack develops at the center of the rail.
The trailing crack propagates faster then the leading one. If
the trailing crack is allowed to reach a length of about 20 to 50
millimeters, one of the branching cracks will turn into
transverse crack.
The “squat” type cracks are surface breaking cracks with
mouths opened upwards and exposed to the action of liquid.
Modeling of squat crack development has shown that
propagation of squats under the rail surface is feasible under
condition of the fluid entrapment effect exerting hydrostatic
pressure at the crack tip. 70
Squats are predominantly found on lines with mixed
passenger and freight traffic.
3.5.6.1 Recommendations to Prevent Formation of
Squats
•
Use of preventive grinding
•
Use of harder rails that increase the threshold at which
ratcheting and crack initiation occur.
o 3-59 x
3.5.7 Rolling Contact Fatigue Defects of Wheels —
Shelling and Spalling21, 24,82,84
Appendix B contains the wheel defect coding used by the
North American and Canadian Railways. The visual
appearance of shelling and spalling are usually
undistinguishable
3.5.7.1 Shelling
This is a subsurface rolling contact fatigue defect developed on
the wheel tread under action of normal contact and shear
stresses (Figure 3.26 ). The mechanism is similar to formation
of shelling in rails.
Figure 3.26: Wheel Thread Deep Shelling
Non-conformal wheel/rail profiles considerably increase
the contact stress. The contact stress often doubles when a
hump develops between running bands on the wheel tread
surface. The hump is formed most frequently with steerable
trucks. Also a false flange of hollow wheels causes the stress to
increase. Another causes of excessive contact stress is dynamic
overloading due to unbalanced load, impact from skid flats,
out-of round wheels, and track irregularities.
o 3-60 x
3.5.7.1.1 Recommendations to Decrease Wheel
Shelling
• Retruing of wheels when a hump or false profile appears
on the wheel tread surface
•
Schedule more frequent light truing of wheel tread surfaces
to eliminate surface defects
•
Provide for conformal wheel/rail profiles
•
Use wheel steel with improved cleanliness
• Use self steering trucks
3.5.7.2 Spalling
Cracks can be initiated by thermal shock produced on the
wheel tread by the rapid heating and cooling when the wheel
slides (skid flats) on the rail during intended or unintended
braking.84,85 When the wheel experiences gross sliding on a rail,
large frictional energy is generated instantaneously, causing the
wheel surface temperature to rise above the austinization limit
(about 720 C) The damage to the wheel is much more severe
than to the rail, since the energy dissipated in the wheel contact
zone, while it is distributed over the sliding length of the rail.
When the austenite is quenched, martensite is formed.
Martensite is a hard, brittle steel phase. After formation, this
brittle phase will easily fracture under loading cycles, initiating
cracks at the surface which eventually result in spalling (Figure
3.27). Dynamic overloading also contributes to high loading
cycles.
Figure 3.27: Wheel Spalling on Wheel
24
Thread with Marks of Martensite Layer
o 3-61 x
3.5.7.2.2 Recommendations to Decrease Wheel
Spalling
• Prevent wagons from being moved with hand brakes
applied.
• Maintain brake valves to enhance uniform braking
throughout the train.
• Select proper empty/load devices on light tare weight
wagons.
The rate of spalling is inversely proportional to car weight.
The less a car weight, the greater the percentage of wheels
removed for spalling.85
The most effective and controllable means for preventing
spalling is to ensure that the wheel is subject to the proper
brake force.86 This is achieved by improved and properly
maintained car braking equipment, particularly for cars
equipped with empty/load devices.
Another possibility for reducing the tendency for spalling
is to increase the temperature at which martensite forms.84
Seasonal variation in shelling and spalling rate. Winter time in
North America and Russia increases wheel shelling damage
considerably compared with summer time. This is evidently
because of the increase of the track stiffness and thus impact
of track distortions on forces between the wheel and the rail.
Another cause of this phenomenon is the influence of liquid.
Water in the form of rain or melted snow enhances crack
propagation rate considerably due to the hydrostatic effect of
liquid trapped in the crack.80 The worst conditions occur when
a dry period (when cracks are initiating) is followed by a wet
period when water enhances crack propagation.48
In contrast to the situation with rails, wheel shelling does
not result in wheel fracture. However, shelled wheels may
contribute considerably to track dynamic overloading, which
may in some circumstances cause rail breakage.
o 3-62 x
3.5.8 Other Rail and Wheel Defects
Rail: Cracks outside the contact zone are about 10% –20% of
all rail defects.24 The most common of these defects is bolt
hole cracking, horizontal split head, and head and web
separation. There is also a vertical split head defect.
Web cracks: Rail web cracks are developed under the
railhead in the longitudinal direction. Cracks may develop from
bolt holes or from the transition from the railhead and web
zones, or from rail welding.
Bolt hole failures: The typical bolt hole failure has cracks that
propagate along a plane at 45 degree to the vertical plane
(Figure 3. 28). This failure is associated with web shear stress
caused by battered rail joints and the stress concentration at
bolt holes caused by poor drilling and beveling.
Recommendations to Eliminate Bolt Hole Failure:
•
Use of bolt hole expansion by drawing a tapered
mandrel through the bolt hole in order to generate a
protective circumferential residual stress around the
hole (sleeve expansion).
•
Proper use of manual track tools
•
Proper maintenance of rail joints.
Figure 3.28: Rail Bolt Hole Defect
o 3-63 x
77
Web/head/foot separations.:These cracks are often developed
in the railhead and web or rail foot and web transition zones.
Corrosion processes strongly influence the process of crack
initiation and development.77 Improper maintenance of rail
joints (for instance, over tightening of bolts) increases the
probability of crack initiation. Recommendations to eliminate rail
web/head/foot separation:
•
Careful inspection of rails
•
Proper maintenance of rail joints
•
Measures to avoid corrosion
Horizontal split head (Figure 3.29): This defect is initiated
from strings of oxide inclusions in the rolling direction.77, 28
This defect is associated most commonly with worn head
hardened rail. Recommendations to eliminate horizontal split head
defect: Proper inspection of rails during manufacturing and in
operation
•
Use clean steel
•
Use deep head hardening
Figure 3.29: Horizontally Split Railhead
77
Vertical split head (Figure 3.30): The damage is associated
with defects of microstructure in the railhead. Also, cracks
originate under the rail surface where high tensile stress appear
as a result of intensive plastic flow and formation of hardened
part of the railhead. In addition, intensive plastic flow may due
o 3-64 x
to overloading of low rail in curves and use of rail with
inadequate hardness. Recommendations to eliminate vertical split head
defect:
•
Improve the production quality control of rails
•
Avoid overloading of the low rail
•
Replace rail found to have this defect
Figure 3.30: Vertically Split Head of Rail
77
Rail foot (base) defects: This defect may be a cause of rail
fracture. Cracks may appear because of manufacturing defect,
uneven rail foot support, corrosion (particularly in tunnels),
resulting in corrosion fatigue rail failure. Recommendations to
eliminate rail foot defects:
•
Improvement of the rail production quality control.
•
Designing and maintaining the rail superstructure on
the level that provides for low bending tensile stress of
the rail base and thus high resistance to the slot
corrosion.
•
Careful handling and installation of rail to avoid base
damage.
Rail weld defects: Problems arising from rail welding may
contribute to about 20% of rail failures.24 These defects may
be divided between defective plant welding and field welding.
o 3-65 x
There are several methods of plant and field welding: contact
electric flash-butt, thermite and gas-pressed welding. IHHA
member countries use various combinations of welding
technologies for plant and field welding. North American
Railways predominantly use electrical flash-butt for plant
welding, thermite for field welding because of a much lower
cost in the field compared with mobile flash-butt welding
equipment.
Russian Railways mostly use contact electric flash butt
welding for plant, and also for field rail welding. To avoid
reduction of hardness of the rail in the weld zone, a technology
that provides mechanical and thermal treatment of the
hardened area was introduced in 70s30 and later was upgraded
to form differential thermal treatment technology. This
technology has been applied to continuos welded and
conventional rails.
Contact flash-butt welding technology with post-welding
induction heat treatment has been successfully applied to fulllength hardened rails of 75 kg/m developed in China for heavy
haul railway operation.87
Flash butt welds have developed horizontal split webs.
Thermite welds have experienced many defects ranging from
shrinkage porosity or inclusions, which led to horizontal split
web failures and subsurface shelling of the railhead leading to
rapid deterioration of the gauge side of rail.76
For field welding, a significant problem is the hardness
difference between the parent rail and weld joint. A weld that
is harder or softer than the rail creates a bump or a dip in the
rail respectively. This is one of the causes of the initiation of
corrugation. Both hard and soft welds are subject to contact
fatigue.
Field welds made using thermite welding have several
problem areas. One of the problems is a stress concentration
in the rail base from the welding process that may result in the
rail failure. Studies are under way to improve stirring of the
weld metal in the mould.
o 3-66 x
To reduce risk of weld failure another solution is to use
the wide-gap (40-80 mm instead of 25 mm) weld technology of
thermite welding . This larger gap allows many rail defects to
be repaired directly with a single-wide gap instead of a plug rail
insert and two conventional welds.21
Influence of axle load. Increasing axle load increases rail weld
failure considerably.
Recommendations to decrease rail weld defects:
•
Improve plant and field flash-butt welding technology
by the introduction of rail weld thermal treatment
technology.
•
•
Improve technology to control defects in field welds.
Apply positive results in the development of thermite
field welding technologies (wide-gap welding, post
weld heat treatment, etc).
Wheel (engine) burn: Slip of a driving wheel on rail causes
wheel ("engine") burns of the rail surface. Because of high
friction, temperatures of the rail can rise above the temperature
required for transformation of eutectoid pearlitic steel to
austenite which may lead to the formation of a martensitic
layer. In unfavorable conditions, this may result in transverse
defect (Figure 3.31). Recommendations to eliminate wheel burn defects:
•
Improve braking equipment
•
Minimize the running surface contamination of rails
which occurs with inefficient lubrication practice
Wheel thermal cracks:24,98 Thermal cracks are due to
formation of tensile residual stresses from repeated cycles of
heating and cooling as occur during hard braking. Thermal
cracks are distinguished in appearance from rolling contact
fatigue by their length and orientation. These cracks extend
vertically into the surface material and will not propagate by
rolling contact. Failure will not occur until a pre-existing crack
and a residual circumferential tensile stress are present. Impact
may cause crack growth to be explosive and may cause the
wheel to break into pieces. Recommendations to decrease wheel
thermal cracks: Avoid drag or stretch braking
o 3-67 x
•
When manufacturing wheels, perform rim quenching
to form high circumferential residual tensile stresses in
the wheel which prevent crack growth.
Shattered rim defect. This defect is caused by large fatigue
cracks that grow parallel to the tread surface about 10 mm
below the tread (Figure 3.32).24,98 It is suggested that a
shuttered rim defect forms from porosity and alumina
inclusions in the wheel tread. This defect is expected to be
found more frequently as axle loads increase. Recommendations to
eliminate shattered rim defect:
•
•
•
Improve manufacturing control of wheels.
Improve inspection of wheels to detect shattered rim
defects at early stage.
Use cleaner wheel steel.
Figure 3.32: Shattered Rim Defect
Out of Round Wheels: This is the result of wheel/rail/brake
shoe interactions. Out of round wheels, when exceeding
certain limits, cause high impact loads on rail.
Slid flat wheel: This defect is caused when the wheel slides
instead of rolls on the rail. A flat wheel may be formed during
hard braking, when the wheel slides on the rail. Flat wheels
may generate very high impact loads on the rail and will
eventually evolve into an out-of-round wheel. Recommendations
to decrease flat wheel defect:
o 3-68 x
•
Avoid hard (drag) braking.
•
Avoid defective empty car load devices.
•
Many railways are introducing wheel impact detectors
and developing condemning limits for out of
roundness and flat wheels based on an impact load
value criteria and retruing the wheel that exceeds the
condemning limits.
Tread metal buildup: Wheel tread metal buildup occurs as a
result of rail, wheel and brake shoe pick up of debris. Wheel
metal buildup is another source of impact loads on rail. It is
reported that metal pick-up occurs under wet conditions,
particularly in the winter time when rails are covered with
snow. 48,89 Recommendations to decrease wheel tread metal buildup:
•
Improve wheel inspections.
•
Perform wheel truing in time.
•
Test brake valve for leakage or proper operation.
There are other wheel defects which may be a cause of
wheel change: cracked or broken rim, grooved tread, thin rim,
thin flange, vertical flange. There are instruments to measure
corresponding defect and condemning limits for each of
defect, which are described in various railroad manuals.
Appendix B contains a table of wheel defects and their coding.
3.5.9 Plastic Flow
Railheads: Railheads are subject to plastic flow. The depth of
plastic flow can vary from a fraction of millimeter to 10-15
mm. The contour of the railheads that has undergone plastic
flow can vary considerably. However, there are three major
regions of intensive plastic flow on the high and low rails.
Sometimes under heavy cars, the low the railhead is crushed
(see also Figure 3.30). The field side of the low rail can suffer
from plastic flow because of overloading of the low rail (Figure
3.33) and inadequate railhead material resistance to the rail
loading conditions encountered in service. Low rail
overloading in curves may occur because of an unbalance
between train speed and superelevation; i.e., trains moving at a
lower speed than is appropriate for the designed superelevation
o 3-69 x
for a particular curve. Sometimes low rail crushing may occur
because of track gauge widening when the low rail is subjected
to a combined high stress and lateral traction. This is made
worse by hollow wheel profiles. Wheels with false flange may
also contribute to low rail plastic flow.
Figure 3.33: Low Rail with Plastic Flow
There are also two potential zones of plastic flow on the
high rail. One is the gauge side of the high rail which appears
as a result of severe two-point contact and traction forces from
components of wheel to rail slippage. Another zone, on the
field side of the high rail may form because of high lateral
forces resulting from lateral displacement of the wheelset.
Recommendations to decrease plastic flow:
•
Use hardened rails with enhanced yield strength.
•
Perform periodic rail grinding to retain designed
railhead profile.
•
Perform proper control of wheel profiles.
•
Ensure that train speed conforms to designed
superelevation.
•
Implement measures to resist gauge widening above
prescribed limit.
o 3-70 x
•
Implement measures to reduce lateral/vertical load
ratio through introducing improved trucks and
lubrication (friction management).
Rail joints and welds: There may be plastic deformation of
the railheads in the region of rail joints. Rail ends are subject to
high dynamic forces from passing wheels. If rail end resistance
to plastic deformation is not adequate to withstand the forces
or the rail joints are not properly maintained, then plastic
deformation may occur.90 This defect is characteristic for non
head hardened rails. There also may be plastic deformation of a
"saddle" type just after the rail joint in the direction of train
movement. The cause of this defect is the dynamic action of
wheels passing rail joint and the lower hardness of the railhead
at rail ends. Plastic flow may appear in the region of rail welds.
Depending on weld technology and subsequent heat treatment
there may be a zone of softer weld metal or two zones one on
each side of the rail weld, in regions affected by the welding
where the rail material after welding becomes softer.
Recommendations to decrease rail welds (see Section 3.5.8):
Stress raisers: Wheel/rail contact may have many high
stress areas and suffer from plastic flow resulting from
unmatched wheel and rail profiles.5 For instance, contact under
hollow wheel conditions and high axle load results in high
contact stresses between the gauge corner and the field of the
rail and the false flanges which may occur on either side of the
worn pattern of the wheel (Figure 3.34). Recommendations to avoid
stress raisers:
Figure 3.34 Contact under Hollow Wheel
Conditions Resulted in High Stress
5
(Stress Raisers)
o 3-71 x
•
Use of preventive rail grinding.
•
Perform wheel turning when hollow wear is beyond a
limit.
•
Apply measures to eliminate stress raiser formation.
• Maintain optimal wheel and rail profile.
Wheel plastic flow. There are two main zones of wheel plastic
flow; i.e., tread surface and flange surfaces. When severe twopoint contact takes place, material flows in direction toward
the flange top (Figure 3.35 ). When the railhead is severely
worn there may be occasions when only top of flange contact
takes place, resulting in strong plastic flow of the top of flange
region. This occurs if the height of wheel flange is considerably
less than the railhead.21,50
Figure 3.35: Wheel Flange Plastic Flow under
Severe Two-point Contact
Wheel tread plastic flow towards the flange is due to high
lateral creepage forces. Other causes of wheel tread plastic flow
are connected with stress raisers, like the false flange of
hollowed wheels. Recommendations to avoid wheel plastic flow:
•
•
•
•
Use wheel truing to eliminate stress raisers.
Use harder wheel material.
Apply rail grinding to avoid severe side wear of the
railhead.
Maintain optimal wheel\rail profiles.
o 3-72 x
3.5.10 Rail and Wheel Corrugations
Rail Corrugation: Corrugation is a periodic irregularity of the rail
surface of one or both rails and is measured by special
instruments. The basic feature of corrugation is that it initiates
on one side of the rail first in curved and tangent track. Rail
corrugation results in loosening of rail fastenings, slippage,
ballast deterioration, and other track element problems. It also
negatively affects the performance of some rolling stock
elements.
Corrugation is initiated by railhead irregularities caused by
rail manufacturing, rail joints, welds, or contact fatigue defects.
Generally, six types of corrugations were identified. However
for freight traffic three types of corrugation predominate:91
short wave corrugation (30-80 mm long), medium wave (200600 mm) and long wave (about 1.5m).
The causes of each types of corrugation are different.
Short wave corrugation: Causes of short wave length
corrugation (short pitch) are considered to be the frictional self
excited stick-slip vibration of a wheelset.77,92,93,95 This occurs in
areas of high traction or intensive braking, where oscillating
longitudinal traction develops due to excitation of the
fundamental torsional resonance of the wheelset.96 The
primary difference in the corrugation formation mechanisms
rests in wavelength fixing and initiation mechanisms. The
wavelength fixing mechanism is usually related to natural
frequencies of the wheel or wheelset and rail and rail tie
coupling. The initiation mechanism can either be related to the
rail-wheel surface irregularities, or the traction creepage
relationship of a particular section of the track. When the
wheel/rail contact patch operates in the vicinity of saturated
creepage, the wheel rolls forward while the traction force
builds up towards the maximum of the traction creep curve.
Once this maximum is achieved, the force can not be increased
any further because of the negative friction characteristics, and
slip occurs.60
Medium wave length corrugation: It is considered that medium
type corrugation with wavelengths of about 200-300 mm is a
characteristic of heavy haul operation (Figure 3.36). Causes of
o 3-73 x
this type of corrugation are explained by the resonance theory.
The excitation of the vibration of unsprung mass of cars on
track having a certain stiffness initiates resonant vibrations
resulting in high dynamic forces that lead to an initial
corrugation process. The wavelength of rail corrugation
depends on resonance frequency of vehicle-track system. The
flexibility of roadbed is an important factor influencing the rail
corrugation. Enhancing the flexibility of roadbed reduce the
rail corrugation. After initiation of corrugations, rolling stock
and track interacting forces lead to plastic flow in the
corrugated areas and intensify the development of corrugation.
Plastic deformation is the leading mechanism of the
development of corrugations. It leads to work hardening
process resulted in the longitudinal hardness distribution of the
corrugated rail became periodic and follows the corrugation
wavelength.97 Contact fatigue defects on rails, like head checks
and shelly spots, also initiate corrugation. It is considered that
shelly spots excite resonant vibration resulting in a wavelength
of about 150-450 mm.25
Long wave corrugation: Long wave corrugation is initiated at
periodic waves of manufacturing origin. The latter is due to the
vibration of the rolling mill mechanism. Though rails are
straightened by roller levelers, some irregularities may be
present. The development of long-wave corrugation is defined
by the rail bed characteristics.
Influence of lubrication: Lubrication greatly affects the
formation and development process of corrugation. High rail
lubrication considerably reduces low rail corrugation. 92, 93,94
Influence of axle load. Corrugation can develop in lower axle
load operations under adverse wheel/rail contact conditions.
Increased axle loads result in an increased growth rate of
corrugation, including the corrugation depth per tonnage
passed,68 and the vertical loads at the bottom of the
corrugation wave. Recommendations to eliminate corrugation:
•
Corrugation is combated by grinding.
•
Short wave corrugation is combated by use of friction
modifiers at the wheel tread/top of rail.58
o 3-74 x
•
Use of higher strength rails.
•
Possible use of softer rail pads to reduce vertical
stiffness characteristics.
Wheel corrugation: wheel corrugation is reflected in wheel
tread profile irregularities. Wheel tread profile can be measured
by a longitudinal wheel profilometer. It is considered that
thermomechanical interaction between the brake block and the
wheel tread is a main contributor to the development of tread
corrugation. High temperature produced during braking
initiates periodical wheel tread wear. The wave length and the
amplitude of corrugations are highly depend on braking block
size and material in combination with wheel material
properties. To decrease the corrugation of wheel treads, it is
recommended that a block material with low elastic modulus
be adopted.99
Influence of axle load. A change in axle load from 29 to 35
metric tonnes showed that corrugations were similar in spacing
and severity. Recommendation: Perform wheel truing timely,
when the amplitude of corrugation reaches the condemning
limit based on impact load criteria.
Wheel Squeal: Wheel squeal is a problem for many railways,
including railways with heavy haul operation, that travel in
urban areas. Wheel squeal originates at the top surface of the
rail and the wheel tread in curves. Wheel squeal is generated by
a combination of rubbing plus stick-slip.100,101 Rubbing arises as
a consequence of the relative movement of the wheel over the
rail, particularly lateral sliding. When the ability of the
wheel/rail interface to accommodate creepage is saturated (at
about 1% of creepage), pure sliding starts. As slipping starts,
friction is reduced. The sliding action ceases when the force
generating the creep is dissipated by the sliding movement and
then the process begins again. This is known as the stick-slip
phenomena.102 Energy dissipated by the stick-slip process
excites squeal in the wheel/rail system.
o 3-75 x
The stick-slip process occurs when the friction
characteristic is negative; i.e, the friction force decreases after
certain level of creepage. Actual properties of the third body
that forms between wheel and rail depend on wheel and rail
materials and environmental conditions.
Friction modifiers that have positive friction
characteristics; i.e., increase friction force with increasing
creepage, decrease or eliminate squeal. Recommendations to
eliminate wheel squeal: Application of friction modifiers to top of
the rail on the low rail in curves.2,60
Acknowledgements
Appreciation is expressed to Evgeny Shur, Head of
Laboratory, All-Russian Railway Research Instate, Stephen
Marich, Marich Consulting Services, Australia, Harry
Tourney, Assistant General Manager, Spoornet, South Africa,
Robert Harder, Associate Professor of Engineering, George
Fox University, USA, James Lundgren, Assistant Vice
President, Danial Stone, Chief Metallurgist and Kevin Sawley,
Principal Investigator, TTCI, USA, for their assistance and
valuable comments.
o 3-76 x
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o 3-83 x
APPENDIX A
RAIL DEFECTS AND CODING
Type of Defect
Transverse Fissure
Compound Fissure
Detailed Fracture from
Shelling
Detailed Fracture from
Head Check
CPR Coding
(Contractor
Designation)
TDT(CO)
TDT(L)
TDT(M)
TDT(S)
TDD(CO)
TDD(L)
TDD(M)
TDDD(S)
DFS(CO)
DFS(L)
DFS(M)
DFS(S)
DFC(CO)
DFC(L)
DFC(M)
DFC(S)
Transverse Defect under
Weld Repair
DFW or
TDW(CO)
DEW or TDW(L)
Broken Rail or Ordinary
Break
BR
Defective Field Weld
Transverse Defect
DWF(CO)
DWF(L),DWF(M)
DWF(S)
DWP(CO)
DWP(L),DWP(M)
DWP(M)
Defective Plant Weld
Russian Railways
Coding [3.77]*
20.1-2
21.1-2
21.1-2
Not observed
Not observed
79.1-2
26.3
Engine Burn Fracture
EBF
27.1-2
Bolt Hole Crack
BHO (outside of
joint)
BHJ (in joint area)
53.2
Vertical Split Head
VSH
Horizontal Split Head
VSJ
HSH
Head and Web Separation
HSJ
HWO
53.1
30B.1-2
30G.1-2
52.1-2
o 3-84 x
HWJ
Split Web
SWO
55
SWJ
Broken Base
PBO
60.1-2; 62.1-2
PBJ
PRO
Piped Rail
50.1-2
PRJ
* There are other rail defects in the classification77
o 3-85 x
APPENDIX B
WHEEL DEFECTS AND CODING
Wheel Defect
AAR Why
Made
Defect
98
Code
Russian
Railways
Coding
Thin flange
41-60
14
Vertical flange
41-62
15
High flange
41-64
—
Hollow tread
10,11
Thin rim
41-73
17
Wheel out-of
round
41-67
—
41-69
—
41-74
32
Shattered rim
41-71
26,27
Shelled thread
41-75
22-1,222,22-3
Thread buildup
41-76
21
Tread slid flat
41-78
20
Thermal
cracks
extending into
plate
Thermal
cracks
o 3-86 x
Defect Cause
Wheel flange worn to the
condemnable limits.
A vertical flange is worn
to a point that the inside
surface of the flange is
flat or vertical to the wheel
tread.
Defect occurs from thread
wear.
Wear between rail crown
and wheel tread.
This defect is caused
from long service life.
This defect is the result of
wheel/rail/brake shoe
interaction.
Thermal cracks are due to
formation of tensile
residual stresses from
repeated cycles of
heating and cooling as
occur during hard braking.
This defect is caused by
large fatigue cracks that
grow parallel to the tread
surface.
Shelling and spalling of
wheel tread is caused by
the rolling contact fatigue.
A build up tread is caused
by metal from the tread,
rail or a cast iron brake
shoe. It then becomes
attached around the
wheel tread.
This defect is caused
when the wheel slides
instead of roles on the
rail.
Click Here To Go Back To Table of Contents
PART 4(a): HEAVY HAUL CASE
STUDY: Dedicated Line with Captive
Equipment; BHP Iron Ore, Australia
Written by S. Marich, member of the International Panel of
Reviewers, A. Cowin and M. Moynan, past members of the
Technical Review Committee (TRC) based on the presentation,
“Managing the Wheel/Rail Interface under Very High Axle Loads
and Tonnages at BHP Iron Ore, Australia.”
4.1(a) Introduction
The operating conditions at BHP Iron Ore, in North West
Australia, are amongst the most severe in the world. Under the
current action of 37.5 tons nominal axle loads and with over 90
million gross tons of annual traffic over a primarily single line
track, the appropriate management of resources, including
wheels (and rails), is critical to achieving the operational
objectives. The line is a dedicated mine to port operation with
capture locomotive and car fleets.
Railroad operations at BHP Iron Ore (BHPIO) began in
1969. The initial railroad was designed to carry about 8 million
gross tons (mgt) of iron ore per year at nominal axle loads of
30 tons. However, within 10 years the haulage rate had
increased to over 40 million tons per year, and currently is
close to 100 mgt per year. Such a rapid expansion led to
unprecedented and major technical problems on the railroad,
including very severe wheel and rail wear, which threatened to
curtail production and hence profitability. These conditions
were ideal motivators for management to invest in both short
and longer term development activities, and accept and
implement technical changes exhibiting the potential for more
cost efficient practices.
BHPIO's operation consists of two mine to port single
track iron ore lines. The Mount Newman line is a 426 km
route with long passing loops. For the most part, the line is
tangent with light curvature. However, moderate (minimum
radius 528 meters) curves are surmounted in the route through
the Chichester range. The line is generally 68 kg/m CWR with
o 4-1 x
a preference for head hardened rail on curves and tangents.
Over one third of the line is concrete sleeperd, with an
ongoing program to be fully concrete sleeperd within a few
years. Loading and unloading loops are on relatively tight
curvature. The Yarrie line is 217 km of similar topography.
BHP Iron Ore has directed much of its technical effort
toward the area of wheel/rail interaction with the aim of
maximising the wheel (and rail) life.
Various wheel and rail initiatives have included:
•
implementation of higher strength and improved
quality wheel and rail materials, to improve their wear,
deformation and fatigue lives,
•
revision of bogie tolerances and workshop practices,
to improve the bogie dynamic characteristics,
•
introduction of modified wheel plate designs, to
reduce the sensitivity to thermal damage,
•
development and implementation of modified wheel
and rail profiles, and
•
wheel monitoring and maintenance procedures, to
optimize wheel life.
These long term “system” developments have led to a six
fold increase in the wheel life, while increasing nominal axle
loads from 30.0 tonnes to 37.5 tonnes, and annual gross
tonnnages from less than 10 million gross tonnes to over 90
million gross tonnes.
4.2(a) Wheels
The major modes of wheel deterioration experienced at
BHPIO are:
•
•
•
•
•
flange wear,
tread wear and hollowing,
surface cracking and spalling,
subsurface cracking (shelling),
thermal cracking, and
o 4-2 x
•
other modes, such as wheel burns, wheel flats,
mechanical damage.
Initially flange wear was by far the most important factor
determining wheel machining cycles, leading to a wheel life of
only about 340,000 km. Consequently, the initial
developments concentrated on this aspect.
4.3(a) Modified Profiles
In the early years, a major problem was the lateral instability or
hunting on tangent track of the ore cars using standard 3-piece
bogies. Hunting caused accelerated cyclic rail wear and
considerable wheel flange wear. Because of excessive flange
wear, a considerable portion of the tread would be machined
away to re-establish the required profiles. Multiple-wear
wheels were being machined only two or three times before
discarding them.
At the same time, rail wear was so severe that the system
would soon be unable to continue operation. An initial
strategy adopted was to transpose a large proportion of the rail,
particularly in tangent track, and re-profile the crown of the rail
by grinding to ensure that wheel/rail contact would be
maintained near the centre of the running surface. Asymmetric
rail profiles were also introduced in curved track to improve
the wheelset steering characteristics. These measures had
immediate benefits on the rail wear, extending the rail life by
about 30% in the sharper curves and 50% in the shallower
curves. The rail profiling also led to some improvements in
the wheel life.
With the severe rail wear addressed, the next major step
was the development and implementation of modified wheel
profiles. This work, consisting of both modelling analysis and
field trials, clearly indicated that the original conical wheel
profiles were unsuited to heavy axle load operations.
An immediate change was to reduce the wheel flange
thickness, increasing the wheel/rail clearance from about 5 mm
to 9 mm. This measure led to a reduction in wheel flange wear
of about 20-25%.
o 4-3 x
The next step was the development and implementation of
a modified narrow flange wheel profile, which together with
the modified rail profiles, incorporated a number of aspects
that were essential for improving wheel and rail life, including:
•
•
•
A relatively conformal contact between the wheel
throat and the gauge corner region of the high rails in
the sharper curves, to increase the effective conicity
and hence improve the wheelset/bogie curving
performance, and to reduce wheel/rail contact stresses
and hence contact fatigue defects.
The creation of adequate wheel/rail clearance and an
increase in the tread curvature, which was established
by the use of an elliptical tread section, to reduce the
wheel tread hollowing, wheel/rail contact stresses and
hence contact fatigue defects.
A relatively large clearance at the rail gauge
corner/wheel throat region in the tangent rails,
resulting in two-point contact, to ensure that a relatively
broad (up to 30 mm) wheel-rail contact occurs near
the centre of the running surface of the rails, which in
turn reduces the effective conicity between wheels and
rails and hence the sensitivity to vehicle hunting.
The modified wheel (and rail) profiles are illustrated in
Figure 4-1(a). Initial field trials showed that the new profiles
led to:
•
•
•
•
reductions in the wheel flange wear of up to 70%,
reductions in tread wear of about 50%,
reductions in the rate of development of uneven wheel
wear within wheelsets and bogies, and
reductions in fuel consumption of 6-9%
The advantages of having conformal rather than two-point
contact in the gauge corner region of the high rails have been
confirmed, with field data showing that the two-point contact
can lead to an increase in the rail gauge face wear rate (and
hence the wheel flange wear) of up to 50%.
o 4-4 x
bhpio new
Conformal
Conformal
bhpio new
High Rails/New Wheels
Low Rails/New Wheels
bhpio new
Tangent Rails/New Wheels
Two Point
Two Point
Figure 4.1(a): Rail/Wheel Contact Conditions for
Curved and Tangent Track
The full scale implementation of the modified wheel
profiles has been one of the major factors in increasing the
average wheel life to over 2 million kilometres, even with the
increases in the nominal axle loads that were implemented in
1981 (to 32.5 tonnes) and 1992 (35 tonnes). The modified
profiles have been so successful that in the last 5-10 years
wheel flange wear has no longer been the main reason for
machining wheels at BHPIO.
4.4(a) Material Characteristics
At BHPIO the original wheels were manufactured according to
AAR-C specifications, with a hardness in the range 321-363
HB. In 1989 the specified hardness range was modified to 341375 HB, which led to:
•
an increase in the mean hardness level, and hence a
reduction in the overall wear rate, but possibly more
importantly
•
a reduction in the hardness range, and consequently a
reduction in the uneven wheel wear within wheelsets
o 4-5 x
and bogies, keeping in mind that wheelsets are
machined on the basis of the faster wearing wheel.
The development and introduction of micro-alloyed
wheels by 1991 provided additional benefits, including:
•
•
a further increase in the hardness range to 362-401
HB, which led to additional reductions in the wear
rate, and
considerable improvements in the plastic deformation
characteristics, particularly under monotonic and cyclic
compression loadings, which led to reduced rates of
wheel tread hollowing.
With the implementation of modified wheel and rail
profiles, tread hollowing had become the major reason for
wheel machining rather than flange wear. Reductions in the
amount and frequency of wheel tread removal led to an
additional problem: the development of sub-surface fatigue
defects were found to initiate at impurities in the material.
Further modifications to the wheel steel specification with
emphasis on steel cleanliness were implemented, keeping a
balance between performance and cost effectiveness.
Continued assessment of wheel performance also showed
the benefits of having a fully pearlitic microstucture in the
wheels, rather than a bainitic microstructure which led to much
greater wheel hollowing rates. This was associated with the
increased wear and deformation exhibited by bainitic materials,
at hardness levels equivalent to pearlitic materials.
From 1978-1979 higher strength; higher hardness rails
were used to replace the original standard rails. The popular
belief was that these rails would adversely influence the wheel
wear rate. The opposite has been actually observed namely: the
use of higher hardness rails led to a reduction in the wear rate of wheels.
The improved material characteristics and the improved
wheel/rail interaction characteristics were the major factors
that also allowed the condemning diameter of the wheels to be
reduced from 895 mm to 880 mm, which represented one
additional wheel machining cycle.
o 4-6 x
4.5(a) Lubrication
Early on lubrication of the wheel flange and rail gauge face was
found to result in a marked reduction in wear and also in fuel
consumption. However, in 1986, the results of a detailed
cost/benefit analysis showed that with the implementation of
the modified profiles and the improved material characteristics,
lubrication could no longer be justified, on the basis of several
cost factors, including:
•
•
•
•
the reduction in traction, and hence the need for
additional tractive effort,
the cost of maintaining lubricators,
the enhancement of rail contact fatigue defects that
occurs in the vicinity of lubricators, and
the development of wheel slip, particularly on grades,
and associated rail and wheel defects.
Consequently, lubrication has been discontinued (with the
exception of locomotive wheels), with no major consequence
on wheel life which continued to increase.
4.6(a) Wheel Design
Up to the mid 1980’s, about 10% of the wheels were
prematurely removed from service because of overheating and
thermal damage, associated with abnormal braking conditions.
To reduce the susceptibility to thermal damage, the wheel
plate design was modified, firstly to a curved section and then
to a low stress (S-plate) section. Laboratory dynamometer and
field testing showed that the new design allowed much greater
flexibility of the wheel plate, and hence reduced the adverse
changes in residual stress levels on overheating.
4.7(a) Bogie Characteristics
These improvements considerably reduce the economic
benefits associated with new, more sophisticated and much
more expensive bogie designs. Considerable work has been
undertaken to improve the characteristics and maintenance of
the current 3-piece bogies, including:
o 4-7 x
•
•
•
•
•
•
•
Matching mounting wheels with similar hardness
values on axles.
Pressing wheels on axles to the minimum back to back
dimensions (within a tolerance of 2 mm).
Machining wheels consistently to a flange number of
zero.
Having a maximum wheel diameter difference within
axles of 0.5 mm, and within bogies of 9 mm.
Having a maximum sideframe length difference of 2
mm.
Taking particular care in maintaining the bogie friction
wedges, to maintain the required bogie lozenging
stiffness and damping characteristics.
Ensuring that appropriate lubrication is applied to the
bogie center plate region, noting that currently plastic
center plate liners are used to control the friction.
•
The retrofitting of all bogies with constant contact
side bearers.
These procedures have been successful in reducing the
sensitivity of vehicles/bogies to hunting, for speeds up to 85
kph, with the consequential improvement in wheel life.
4.8(a) Wheel Maintenance
To manage the wagon fleet a vehicle/bogie/component
tracking system was developed and introduced which provided
detailed information on a wide range of parameters, including
the wheel diameters, defects, flange readings and distance
travelled.
A new high production lathe allows all wheelsets to be
machined between centres and accurate profiles to be
reproduced onto the wheels, and also allows a tight control on
the wheel diameters.
BHPIO developed condition monitoring procedures to
assess the characteristics of vehicles in terms of their dynamic
performance, in particular their hunting behavior.
o 4-8 x
This led to redefining the allowable hollowing limit of the
wheel tread and initiates actions to be taken on those wagons
which exhibit relatively poor dynamic behavior.
4.9(a) Summary
Every railroad operator looks for the ultimate life from wheels
and rails. However not all owners or operators of freight
wagons have control over the maintenance of both the wheels
and rails, even within private rail systems. This is partly due to
the traditional structures that railroads operate under, the lack
of understanding by the track maintainers and wheel
maintainers of each other’s needs, and the incentives to
independently reduce their operating costs.
The optimization of wheel and rail life can be obtained
only when there is a coming together of the two disciplines, to
operate under a set of rules with the primary common aim of
achieving what is best for both.
At BHP Iron Ore, the management and control of both
wheels and rails for the benefit of the overall system, has been
able to increase the wheel life from 340,000 km to over 2
million kilometres, while still increasing axle loads and
eliminating lubrication. This has been possible by the
development and application of a wide range of strategies.
Further increases in the wheel life up to 2.5-3.0 million
kilometres, are believed to be possible by controlling the wheel
machining cycle, within the current limits, and the amount of
metal that is removed during each cycle. As BHPIO's diligent
efforts illustrate, combining science, engineering, management
and supervisory control appropriately can create a world class
heavy haul railway exceeding in productivity and cost
effectiveness. The best practices of the BHPIO operation are
summarized in tabular form, Table 4.1(a).
o 4-9 x
Table 4.1(a) Best Practices developed by BHPIO Rail Operations for
Very High Traffic Density, Extreme Axle Loads Dedicated System
TRAFFIC MIX: DEDICATED HH AXLE LOAD: 37.5
TERRAIN: MODERATELY DIFFICULT TRAFFIC DENSITY: 100 mgt
RAIL
TRAFFIC
DENSITY
o 4-10 x
TYPE
Premium
100 mgt
WHEELS
TYPE
Micro-alloy
Premium
Multi-wear
970 mmØ
(895 mm
condemning
limit)
S-plate
365+ HB
PROFILE
modified
narrow
flange
CROSSTIES
WEIGHT
68 kg/m
BOGIES
3-piece
constant
contact SB
Bogie
tolerances:
see Part 2,
Section 2.4.
Plastic
center plate
liners
lubricated
FASTENERS
Concrete
monobloc
Elastic
PROFILES &
MTCE
WHEEL/RAIL
curve:
conformal
tangent:
modest 2point
LUBRICATION
WHEEL/RAIL
locomotive
wheels only
no rail
lubrication
BALLAST
M/L SWITCH &
CROSSING WORK
Crushed
Rock
WEAR
LIMITS
wheels:
condemn
@ 880
mmØ
NOTES
*Frequent, mileage based preventative
maintenance teardowns
*Maintenance of tight tolerances and
proactive maintenance programs
*Continuing research and monitoring
program
*Automated capture of key component
wear information
*Full component tracking throughout
life cycle
*World class system performance
evaluation and corrective action
protocols
Click Here To Go Back To Table of Contents
PART 4(b) Case Study of Wheel/Rail
Cost Reduction on Canadian Pacific’s
Coal Route
Written by Mr. Michael D. Roney, member of the IHHA Board of
Directors
4.1(b) Nature of the Business
Since the late 60’s, one of the most important lines of business
for the Canadian Pacific Railway has been the transportation of
metallurgical coal from the mines of southeastern British
Columbia to tidewater at Vancouver. The coal is of good
quality and is competitive on world markets, and as such, it
finds its way into the furnaces of Pacific Rim steel mills.
But this is in spite of the challenge of moving this coal
from mines more than 1100 km. (700 mi.) from tidewater, and
lifting it over two mountain ranges and through some of the
toughest railroading country in the world. CPR carries more
than 25 million metric tonnes of coal per year over this route
in unit trains with payload capacity ranging from 11,000 to
13,250 metric tonnes, powered by three 4400HP AC traction
locomotives, when negotiating the steep grades.
4.2(b) Characteristics of the Route
The route is predominantly single track with 46% of the
routing traversing curves tighter than 3492 m radius (1/2
degree) and 133 km. (80 miles) of curves less than 312m radius
(over 6 degrees). Maximum curvature is 170m radius (11
degrees). The controlling grade westbound is 1.1% and the
route passes through several tunnels, including the Mt.
Macdonald tunnel, at 14.6 km. the longest in the Western
Hemisphere.
Weather is another factor in the operation. Temperature
extremes in the Thompson River valley range from +43C
(110F) to –34C (-30F). The routing also passes through 132
potential avalanche paths and encounters annual snowfalls
through the Rogers Pass of 1220 cm. (40 ft.).
When the coal business started in earnest, the track was
o 4-11 x
laid with 66kg/m rail. Rail used in curves was a 1.3%
chromium steel with a hardness of 325 BHN minimum, with
conventional 260 BHN standard carbon rail in tangents.
Sleepers were typically a 2438-mm long (8-ft.). Douglas Fir
softwood tie fastened with cut spikes.
4.3(b) The Consist
Canadian Pacific’s coal unit train operations started in 1970,
using remote control mid-train power and up to 13
locomotives, or 39,000 HP on the steepest grade. The 111 cars
in the consist were a bathtub gondola capable of carrying 95
tonnes (105 tons) of coal, for an axle load of 30 tonnes (33
tons). They were equipped with standard three-piece Barber S2 trucks and rotary couplers.
4.4(b) Early Problems
As tonnages built, it became apparent that rail and sleeper
damage was occurring at a rate that was eating up profit
margins. Rail in curves tighter than 468 m (4 degrees) was
lasting an average of only 200 million gross tonnes (212 mgt).
Whereas in the past, rail would work harden in service, curve
low rail had begun to flow plastically. With excessive plastic
flow came head checking and development of corrugations in
3 years or less. The gauge face of curve high rails began to
abrade and white flakes of metal would sprinkle the right-ofway. Of more concern was the onset of deep gauge corner
shelling, initiating around 9.5 mm (3/8 in.) below the gauge
corner. These shells were associated with an increased risk of
transverse defects and rail fracture.
Timber sleepers began to display excessive cutting of the
tieplate into the tie on the outside of curves. This would then
lead to wide gauge conditions.
CPR uses two-wear Class C wheels with a hardness
between 321 and 363 BHN. These wheels were found to last
283,000 km (170,000 mi.), with the prime cause (63%) of
changeouts being shelled treads. Crack initiation was caused
by a combination of lateral tangential forces due to frequent
curve negotiation in conjunction with the heat inputs of
braking. Once initiated, the moisture present from blowing
o 4-12 x
snow would provide the hydraulic action that caused such
cracks to grow.
4.5(b) Initial Attempts to Control Rail and Wheel
Wear Costs
It was determined that there were several main factors leading
to high rail costs:
1. The 655 MPa (95,000 psi) yield strength of the alloy
rails was inadequate for heavy axle load unit train
service.
2. Wheels were traversing curves with a great deal of
slippage or creep and lateral forces were
correspondingly high.
3. The rail grinding being performed was ineffective at
preventing the re-occurrence of corrugations.
4. Low rail damage was being affected by wide gauge.
Domestic rail manufacturers found that it was very
difficult to make a conventional alloy steel with greater yield
strength without making it too brittle. By 1984, CPR was
substituting a 350-390 BHN chromium-alloyed head hardened
rail from Japan that was head hardened through heat
treatment. This steel was also metallurgically cleaner, having
lower densities of brittle oxide-inclusions which were often
found at the initiation points of transverse defects. Rail was
also converted to 68 kg/m (136RE) as curves were changed
out to take advantage of an additional 5 mm (3/16in) vertical
wear allowance.
The periphery head hardened rails from Japan showed
reduced wear and plastic flow, yielding a 25-100%
improvement in rail life. In addition, the cleaner steels did not
show the usual infant mortality phenomenon, characterized by
an early appearance of a population of transverse defects. On
the other hand, some of the head hardened rails developed
deep gauge corner shells that were attributed to the fact that
the hardened rail gauge corners did not quickly conform to the
throat of the wheel and were subjected to a “point” load.
Wide gauge was attacked by converting curves to 274-cm
o 4-13 x
long (9-ft.) hardwood ties in curves. The 36-cm (14-in.) tieplate was changed out for a 41 cm. (16 in.) plate eccentric to
the field to resist overturning.
A research study on rail grinding, prompted by success in
the Australian Pilbara with profiling to asymmetrically reshape
rail, found that the rate of regrowth of corrugations could be
reduced by 40% by preferential grinding of the field side of the
low rail. This provided some relief from contact with the sharp
reverse curvature of hollowed wheel treads when gauge
widened. On the high rail, it was found that grinding down the
high rail gauge corner preferentially reduced the rate of
occurrence of shelling. Both actions were made easier by
initially grinding the rail to a 200 mm (8 in.) head radius.
Grinding cycles were also progressively tightened down to 16
million gross tonnes (18 mgt) between grinds, with the goal of
providing better control of rail shape and surface cracking with
a single pass, if possible. This action, in conjunction with the
phasing in of metallurgically cleaner, harder rail steels, was
successful at eliminating corrugations as a reason to change out
rail.
On the wheel side, CPR began testing different wheel
profiles in the late-70’s, with the theory that a “worn” wheel
profile when machined onto a new wheel, would not exhibit
the wearing-in that foreshortens the life of the AAR 1 in 20
coned profile used at the time. Professor Heumann of
Germany had introduced this “worn wheel” profile into
Europe in 1934, suggesting that a profile that provides a single
point of contact in curving would offer better performance
than the usual two-point scenario. This idea was appreciated
by engineers at the Canadian National Railway, who in early
1970’s introduced a wide-flange Heumann profile for
locomotives in mountain service. When their tests found that
wear life (primarily limited by flange wear) was doubled over
the conventional thinner flange AAR 1:20, the “CN wide
flange Heumann” profile (Profile A) was adopted as their
standard. The AAR followed the CN progress closely and
based on a different sample of wheels and rails developed a
similar “worn wheel profile,” presented as the interchange
standard AAR1B profile in 1986.
o 4-14 x
Wheel profiles were subsequently converted from the CN
Profile “A” to the new industry standard, the AAR 1B. The
1B brings many of the benefits of “worn wheel” profiles and is
a good balance between what is needed for good curving and
for reduced track hunting; albeit without as much metal in the
throat area as the CN Profile A.
According to testing subsequently done on several
railroads in Canada, the AAR1B wheel had shown some lack
of matching to their rail. The CN found that the AAR1B wheel
provides a two point contact in sharp curves that undergoes
high rates of slip induced wear until after about 60,000 km of
run-in. The CN Heumann A wheel required about 30,000 km
of run-in, with subsequent designs reducing that number to
lower than 8,000 km. At the Quebec Cartier Mining Co.
railway, the AAR1B was associated with the wheel version of
“infant mortality” – a high incidence of contact fatigue induced
wheel-shelling failures within the first 20-30,000 kilometers of
life. A QCM Heumann wheel profile was designed and in
“head-to-head” field testing, the new profile reduced the
incidence of wheel shelling by about 60 percent. The success
of an optimized wheel profile at QCM led the Railway
Association of Canada in 1996 to commission the
development of the NRC-ASW, “anti-shelling” wheel profile
for Canada’s captive grain and coal fleets.
The NRC-ASW is a “worn” wheel profile designed to
minimize creepage and contact stresses that contribute to
rolling contact fatigue shelling of steel wheel treads. This wheel
profile provides the following geometrical features compared
with the AAR1B (see Figure 4-1(b)):
•
The addition of 1.6 mm of metal in the flange
root, which significantly improves steering
performance, reducing creepage and wear.
•
The 1:20 cone angle in the tread contact region
(same as the AAR1B) leaves unchanged the
wheel’s resistance to hunting in standard gauge,
tangent track.
•
A 20” field-side roll-off is another notable change,
since it further improves the wheelset steering
o 4-15 x
moment, and increases significantly the time to
development of a false flange.
Figure 4-1(b): The NRC's Second Generation Anti-Shelling Wheel
(ASW) Profile Compared with the AAR1B Wheel
Besides the NRC-ASW profile, CPR has been testing
wheels of different chemistry and steel cleanliness, with up to
eight different wheels being tested. While wheels with micro
alloying elements look promising, the evaluation continues.
4.6(b) Benefits of Frame-Braced Steerable Trucks
CPR became interested in steerable radial trucks after field
tests with the Barber-Sheffel and DR-2 trucks in the early 80’s
showed flange wear reduced to ¼ that of control cars. Further
extensive field tests with 240 steering truck equipped cars
confirmed the saving and added the observation that wheel
tread shelling was reduced by 2/3, simply because of the
reduction in lateral slippage through curving. Important from
an economics point of view was the fact that retrofittable
versions showed a performance that was close to comparable
to the new trucks. Frame braced trucks became the standard
for the coal fleet in 1989
Recent analysis indicates that Frame Brace (FB) has
improved wheel performance on the coal fleet by
approximately 40%, increasing average wheel life from
325,000 km (195,000 mi.) to 453,000 km (272,000 mi.), while
gross vehicle weight has increased from 120 tonnes (263,000
lbs. to 130 tonnes (286,000 lbs.).
In addition, previous controlled fuel testing indicated that
the use of FB resulted in 5.8% fuel savings over the length of
the coal loop.
Besides improved wheel performance and fuel efficiency,
o 4-16 x
recent teardown of trucks with FB after 1.4 million miles
showed minimal wear , with cars able to go many more miles
without attention.
To control the risk of rail, wheel and bearing failures due
to tread shells, CPR installed a flat wheel detector at two one
locations along the route in 1987. This detector removes any
impacts exceeding 140 kips from the mix, setting out the
offending cars for wheel changeout. There are now 11 such
flat wheel detectors strategically located across the entire CPR
trackage.
4.7(b) Premium Rail Steels and Extended Wear
Limits
The experience with periphery-hardened Japanese rails had
illustrated the benefits of clean steel-making practices and of
increased hardness and yield strength, but it had also shown
the importance of managing the contact stresses between
wheel and rail, largely the result of the degree of conformance
between the profiles. The bottom line was that if the risk of
rail failure could be reduced, rail could be allowed to wear to
reduced sections.
Canadian Pacific then embarked upon a program to extend
rail wear limits to get extended like from the new, cleaner rail
steels. The pillars of the program were:
1. Frequent reprofiling of rail to ensure good contact
stresses.
2. Regular laser-based measurement of rail cross sections
to ensure that the railway knew the status of wear of
each curve on the system.
3. More frequent ultrasonic inspection to manage the risk
of internal defects.
4. Computerized analysis of wear rates to project the
optimal time for rail renewal.
5. Improvements in the metallurgy of rail steels at deeper
depths within the railhead, in anticipation of stressing
the base metal to greater limits.
Parameters were set on the tolerances on the maximum
o 4-17 x
deviations of the rail profile from design templates and these
were used to program rail grinding to reduce damage from
high contact stresses. At this time, the profile templates were
changed to be more conformal with the worn wheel and the
extent of gauge corner undercutting was reduced to balance
the need to promote good steering with the need to control
gauge corner fatigue. Different profiles were used for
tangents, low curvatures, mild curvature and high curvature.
The sharper the curve, the more the gauge face undercutting
and the more the field side relief of the low rail. A frequent
rail grinding cycle was maintained to concentrate on
preventively dressing surface cracks before they cracked out to
form surface spalling. By controlling surface cracking, the rail
surface layers were prevented from gross weakening that
contributed to formation of corrugations.
Tightening up on ultrasonic rail flaw detection further
mitigated the risk of internal rail fatigue. This increased the
probability that an internal defect would be detected in the
time between reaching a detectable size and growing under
traffic to the size that represents a significant risk of fracture.
The third factor that helped to reduce rail fatigue was the
adoption in 1987 of a deep head hardened heat-treated alloy
rail from Japan. Deep head hardened rail had rail surface
harnesses in the 370-390 Brinell range, an improved yield
strength, and was able to retain hardness in the 340 BHN
range to depths of 18 mm (3/4 in.) into the rail head. CPR
also adopted an intermediate grade rail for tangent and mild
curves with a 325-340 BHN range now available in clean
steelmaking practice from domestic sources. This rail raised
the surface yield strength above the threshold that had
previously succumbed to heavy rail flow and corrugation. CPR
began to purchase all rails to a 200-mm (8-in.) head radius,
which is how they were being ground in the field.
With the three factors of frequent single pass profile rail
grinding, frequent risk-based rail testing, and deeper hardness,
higher yield rail steels, the reason for changing out rail in
curves shifted overwhelmingly to wear. This left CPR poised
to extend rail wear limits with immediate and large savings to
the rail budget.
o 4-18 x
The new rail wear limits moved the average percent head
loss from 25% of the head to 35-40%. They were based upon
finite element analysis of the rail and were designed to ensure
that internal rail stresses did not exceed 2/3 of yield at depths
in the rail head where catastrophic defect types like vertical
split heads were seen to initiate.
It was found that the new extended wear limits did not
increase the risk of failure, but that rail would wear rapidly
beyond these limits. If they were significantly exceeded, rail
fracture could and did occur.
The key element in controlling risk of extending wear
limits was knowing the wear condition of rail accurately, and
being able to project the right time to change out the rail.
Optical rail measurement technology proved to be the answer.
CPR started with LITESLICE optical rail measurement on
their Track Evaluation Car in 1992, and converted to the more
accurate laser-based LASERAIL measurement system in 1994.
CPR now measured rail wear three times per year and
developed computer programs that projected the observed
wear up to five years into the future. This also permitted
centralized planning of rail programs with the computer
projections being the prime input to rail planning.
4.8(b) Increased Axle Loads and AC Traction
As always in railroading, nothing stands still and the improved
control of rail and wheel wear was the catalyst to reduce
operating costs by increasing axle loads. Increased car
capacities with improved net-to-tare reduce car-miles and
operating costs. If the track is in good shape and designed for
the traffic, increased maintenance of way costs, constituting
only 12-14% of variable costs of traffic, are overwhelmed by
fuel, crewing and motive power cost savings.
CPR indexed up gross vehicle weight to 125 tonnes
(275,000 lbs.) and then 130 tonnes (286,000 lbs.) on 4 axles by
1995. Again, radial frame-braced trucks controlled damage in
the high curvature environment. Newer coalsets were
purchased with aluminum car bodies, which had overcome the
weld-fatigue problems of earlier aluminum construction.
These cars raised the weight-to tare ratio of the 130 tonne
o 4-19 x
(286,000 lb.) GVW cars to 5.2.
Canadian Pacific converted to 4400HP AC traction units
in 1996. These new units initially put the rail top surface to the
test as algorithms were perfected to control wheelslip. With
the available adhesion increased from 19% to 38% on normal
running and 48% on lifting the train from a start, the rail takes
a high longitudinal tangential force, but what is critical is that
adhesion in the contact patch does not become slippage. With
the perfection of computerized microslip control, early
indications of surface damage disappeared and the AC units
have not produced any incremental damage that is not handled
by regular rail grinding; 6000 HP units with radial trucks are
poised to become the new standard for the coal service and
have contributed substantially to operating cost savings.
4.9(b) Further Cost Savings
It has been observed in the field that rail deterioration
increases when gauge is permitted to progress to greater than
13 mm ( ½ in.) wide. This is likely due to the greater tendency
of wheel tread hollowing to contact the low rail field side.
While predominantly a timber-sleepered railway, CPR became
concerned with the continued use of cut spikes in the sharper
curves of the western corridor under AC traction and heavier
axle loads. A program was undertaken to replace cut spikes
with a rolled plate held down by 5 screw spikes with spring
washers, and Pandrol e-clips. This fastening system was
designed to give extra safety against rail rollover and to reduce
timber sleeper deterioration by greatly controlling tieplate
movement and the subsequent abrasion of the wood fibers in
the rail seat. But as a very important side benefit, it has been
found that rail life is double in such curves because gauge is
typically controlled and rail does not rotate as freely and
maintains good contact conditions with the wheel.
Another good news story has been the effect of continued
installation of harder, cleaner rail steels in curves. It has clearly
been shown through rail profile measurements that the harder
steels maintain their lower surface stress shapes longer, are less
susceptible to surface cracks, and have the strength to resist
gross plastic flow and corrugation. As a result, CPR reduced
o 4-20 x
from four to three annual grinds of the coal route, again
predominantly single passes. The cycle between grinds is now
in the 23 million gross tonne (25 mgt) range and the annual
saving is $676,000 Cdn. Rail grinding templates have also been
adjusted to more conformal design rail profile shapes, which
has reduced metal removal through grinding from 60% of
average vertical head loss to 40%., while controlling surface
fatigue.
CPR has recently installed min. 370 BHN low alloy
hypereutechtoid rail steels in curves on the coal route. These
steels have a higher carbon content and thicker cementite in
the matrix. In field tests, they have shown a 10%
improvement in wear life on low rails and a 25% improvement
on curve high rails. They appear to be more resistant to
surface cracks and work harden well. The railway has also
installed some rails to the AREMA 141 section (71 kg/m).
This rail is attractive as it offers an additional 5.6 mm (7/32 in.)
vertical wear and is completely compatible with the 136RE
section already in place.
In 1999, CPR contracted for a survey of the status of rail
lubrication using a new hi-rail equipped tribometer car. The
survey uncovered major opportunities to improve rail life
through better and more consistent protection of the friction
coefficient of the rail gauge face.
CPR attempted to implement the optimal friction
guidelines developed by the TTCI in the US in a 50-mile test
zone. These were designed to both protect the rail gauge face,
reduce lateral gauge widening and improve fuel efficiency. The
guidelines were:
•
Maintain top of rail friction coefficient differential, left
to right <0.1 µ
•
Maintain top of rail friction>0.30 µ
•
Maintain gauge friction coefficient of <0.25 µ
New electronic lubricators were carefully spaced to achieve
the desired result. They incorporated two 55-inch wiper bars
o 4-21 x
with 48 lubricant ports on each rail. A more expensive rail
grease was also applied. In spite of a doubling of the lubricator
spacing over what had existed, these lubricators were
successful in meeting the 0.25 µ gauge face criterion. On the
other hand, the top of rail friction was difficult to meet.
Notwithstanding, the results showed that lateral rail wear was
reduced by 75%, high rail vertical wear was reduced by 10%,
but low rail vertical wear was increased by 15%. With further
improvement in delivering lubrication to the top of rail, it is
hoped that the full fuel benefits will be achievable in the near
future.
Another new development has been the installation of low
alloy hypereutechtoid rails in sharp curves, starting in 2001.
The new rail steels have a higher carbon content, improving
hardness and yield, but without the loss of ductility usually
expected of increased carbon. This is achieved by thickening
the cementite layer. In early testing, these rails have shown a
10% improvement in low rail life and a 25% improvement in
high rail life over low alloy deep head hardened rails.
There also appears to be good economics to changing the
rail standard from the 136RE rail section to the new AREMA
141 section. The 141 has 20% more available vertical wear at a
net cost increase including installation of only 5-7%. This is
under test as the new standard for the western mainline.
On the mechanical side, studies of wheel tread hollowing
on cars in the coal fleet have sought to determine whether
there are benefits to reprofiling wheels that are hollowed to
greater than 3mm, at which time, studies have indicated a
significant deterioration in curving performance. It was found
that the coal wheels had only 4% of wheels in this category,
less than the 6% found for a sampling of the US rail industry
by TTCI.
4.10(b) The Size of the Prize
Strategies to improve wheel and rail wear control on Canadian
Pacific Railway’s coal route have been an evolutionary practice
that has progressively balanced the wear mechanisms of rail,
risk management, wheel and rail profile matching, and the
changing demands of the operating environment. The job will
o 4-22 x
never end as competitive pressures in the world coal market
demand a continuous stream of operating cost efficiencies.
On the CPR, wheel and rail wear control strategies have
proven to be of great value not only in reducing wheel and rail
costs, but in facilitating operating cost efficiencies such as
increased axle loads and fuel consumption improvements.
Looking back, in 1970, the average wheel lasted 280,000
km (170,000 mi.), while today wheels average 453,000 km
(272,000 m.). Average rail life on the highly curved coal route
has improved from 293 million gross tonnes (325 mgt) to 612
million gross tonnes (680 mgt). At the same time, the new AC
traction consists and heavier axle loads have reduced car mile
costs by 25% and fuel costs by 15%.
o 4-23 x
(blank)
o 4-24 x
PART 4(c) Wheel And Rail
Performance at Carajás Railway
Written by Ricardo Schmitt Martns and Ronaldo José Costa, CVRD
4.1(c) Introduction to Carajás Railway
Carajás Railway is part of an integrated mine/railroad/port
project called Carajás Iron Ore Project which belongs to
CVRD/EFC - Companhia Vale do Rio Doce. In this Railway
is offered freight and passengers transportation service at
north and northeast of Brazil. It counts with 892 km of a
single track with 1.60 m gauge from Carajás - PA to São Luís MA being iron ore the main product hauled. The typical
consist to the iron ore transportation is formed by 02
locomotives and 206 cars being each car with the total weight
of 122 metric ton.
4.2(c) Historical Data
4.2.1(c) Wheels
On the original project it was to use 38”, one wear (1W)
wheels on GDT (Gondola Dual) wagons which transport iron
ore. The expected wheels’ mean life was 500.000 km without
truing them, it was all based on the track design that has 73%
of tangents and 27% of curves being the sharpest one of 783
m radius. But just after starting operation, in 1985, surface
defects appeared on the wheels and for this reason their life
should be much less than the expectation. In fact, wheels’
mean life dropped to approximately 350.000 km as shown in
Figure 4-1(c).
At that time, 10% of the wheels presented surface defects
and they appeared on both forged and cast wheels in the same
proportion and periodicity.
o 4-25 x
Survival curve - One wear wheel (1W)
100
% not scrapped
90
80
70
60
50
40
30
20
10
1500
1440
1380
1320
1260
1200
1140
1080
960
1020
900
840
780
720
660
600
540
480
420
360
300
240
180
60
120
0
0
Kilometers (km x 1000)
Figure 4.1(c) - Survival Curve - One Wear Wheel (1W)
In the beginning, it was believed that the defects were
caused only by brake efforts, attributed to the excessive
number of stops on the sidings because there were no
signaling available in the railway.
Some actions were taken, as listed, in order to minimize
the damage effects the railroad was facing.
4.2.2(c) History
4.2.2.1(c) During 1986
In 1986, CVRD/EFC contacted wagons, brake equipments
and brake shoes manufacturers to discuss about the problem
was happening. The representatives decided to form a
technical group to collect data to a further analysis of the
problem.
Results: There was no success with this group and also
there were no recommendations.
4.2.2.2(c) During 1987
In 1987, CVRD/EFC contracted a Brazilian university to
investigate the problem. The proposed work included an
instrumented bogie, an instrumented wheel set with
thermocouples in order to evaluate bogie dynamics and heat
generation between wheel and brake shoe. Both aspects were
studied during a complete trip.
o 4-26 x
Yet during 1987, CVRD/EFC took some actions:
A - Increased use of dynamic brake
B - Started truing wheels with defects
C - Diminished number of train stops during a trip
D - Removed C-PEP
E - Started using thin flange on the wheels to propitiate
more freedom between wheel set and rail
F - Started turning the consists from time to time
G - Measured wagons brake ratio
H - Tested wagons with isolated brake
I - Contracted an international engineering company. The
following activities were developed:
I.1 - Design of a field instrumentation:
I.1.1 - Wheel profile measurement
I.1.2 - Impact loads evaluation
I.1.3 - Angle of attack measurement
I.1.4 - Contact loads measurement
I.2 -Residual stress on wheels with different defects
stages measurement
I.3 -Thermo graphic measurements on wheels during
brake application
I.4 - Monitoring with data acquisition system
I.5 - Data processing and sample analysis
I.6 - Final report
J - Tested imported brake shoes
K - Material investigation
Results: From the Brazilian university, the conclusion
pointed that the heat developed between brake shoes and
wheels was not high enough to cause damages to the wheels
and the efforts generated by bogie dynamics were considered
acceptable.
By truing the wheels, it was possible to increase wheels'
lives.
The international engineering company, conclude it was
necessary to develop a new wheel profile.
All the others actions did not aggregate any substantial
improvement.
o 4-27 x
Click Here To Go Back To Table of Contents
The main worries were:
A - A high number of wagons stopped in the shop,
leading to a low availability.
B - Irreversible damage to the bearings, caused by high
impacts.
C - Damage to the track, caused also by high impacts;
D - Broken wheel.
4.2.2.3(c) During 1988
In 1988, CVRD/EFC decided to reduce the GDT gross load
in 10 metric ton to evaluate the impact of axle load and also
started using two wear wheels (2W).
Again, CVRD/EFC contracted an international institute
company, the same that suggested a new wheel profile. The
continuation of the first contract main objective was to
develop a new wheel profile. In fact, there were two modified
profiles.
Results: The load reduction did not cause any difference
on the wheels performance.
With two wear (2W) wheels it was possible to reprofile the
wheels more time and it represented an increasing on wheels’
life as shown in Figure 4-2(c), because there was more material
to be removed.
The proposed modified wheels' profiles showed no
improvement compared to 1:20 wheels profile used.
Survival curve - Two wear wheel (2W)
100
% not scrapped
90
80
70
60
50
40
30
20
10
1500
1440
1380
1320
1260
1200
1140
1080
960
1020
900
840
780
720
660
600
540
480
420
360
300
240
180
60
120
0
0
Kilometers (km x 1000)
Figure 4-2(c): Survival Curve - Two Wear Wheel (2W)
o 4-28 x
4.2.2.4(c) During 1989
In 1989, CVRD/EFC contracted another international
institute. It was supposed that the problem was related to the
wheel/rail interaction and the conclusion would be based on
the contact mechanics theory. This conclusion would lead to
keep contact stress with low values and use wheels of high
hardness, as detailed:
A - Contact stress with low values
A.1 - Controlling wheel geometry
A.1.1 - Reducing wear: Eliminating hunting, lubricating
locomotives wheels’ flanges, improving curving
A.1.2 - Preventive wheel truing
A.2 - Controlling rail geometry
A.2.1 - Machining and planning rails: Initially correctively and
then preventively
A.2.2 - Reduction of plastic flow by means of lubrication
A.2.3 - Eliminating gage corner wear: lubricating and
improving curving
B - High hardness
B.1 - Avoid wheels of low hardness
B.2 - Avoid excessive wheel heating
Results: Some of the recommendations helped to improve
the wheel/rail system, mainly the rails.
4.2.2.5(c) During 1990
In 1990, CVRD/EFC bought a portal lathe machine to initiate
a wheel truing program.
Results: With a new lathe machine, it was possible to
initiate a wheel truing program to avoid increasingly damages
to the railway and also get more from the wheels.
4.2.2.6(c) During 1992
In 1992, CVRD/EFC changed the minimum rim thickness
from 30 mm to 25 mm.
Results: It was possible to remove more material;
consequently, wheel life was increased.
o 4-29 x
4.2.2.7(c) During 1993
In 1993, CVRD/EFC started using multiple wear (MW)
wheels.
Results: With multiple wear (MW) wheels it was possible
to remove more material representing another increment on
wheels’ life as shown in Figure 4.3(c).
Survival curve - Multiple wear wheel (MW)
100
% not scrapped
90
80
70
60
50
40
30
20
10
1500
1440
1380
1320
1260
1200
1140
1080
960
1020
900
840
780
720
660
600
540
480
420
360
300
240
180
60
120
0
0
Kilometers (km x 1000)
Figure 4.3(c). - Survival Curve - Multiple Wear Wheel (MW)
Effectively, the main actions adopted, were:
A - Wheel truing, as a palliative action;
B - Use of wheels with thicker rim (MW). Firstly two
wear (2W) and then multiple wear (MW) wheels.
4.3(c) Improvements
By the end of 1993, with results considered insufficient, it was
decided to separate the focus in two: The first focus was to
build a wheel management model in order to obtain the best
cost effectiveness from the wheels and processes involved and
the second was to look for a solution or attenuation to increase
the period the wheels could run without defects. The second
focus drove, necessarily, to an involvement of rail part.
4.3.1(c) Wheels Management Model
The first thing to be done was to build a data base so it would
be possible to understand better the process that was taking
place. After a certain time it was noticed that:
o 4-30 x
A - Wheels were trued in average at every 7 (seven) months
but the standard deviation was almost 3 (three) months
B - Approximately, 20% of the wheels on the fleet presented
defects
C - The average depth of material removed was 6 (six) mm
with a standard deviation of 2.5 (two point five) mm
With some information extracted from the data base a
matrix matching the relationship between cost and
performance was built to show where CVRD/EFC was placed
and to which position it would be possible to go. Part of the
matrix is shown in Figure 4.4(c).
Truing
period - [Month]
4
5
6
7
8
9
10
11
12
3,0
4,5
5,0
1,0
1,2
1,0
4,0
5,7
6,0
8
6
5
Variables
Average depth of material - [mm]
5,0
5,5
5,5
Average wear - [mm]
1,5
1,0
1,5
Average loss of material - [mm]
6,5
6,5
7,0
Number of truing
5
5
5
6,0
6,0
6,5
1,0
1,5
1,0
7,0
7,5
7,5
5
4
4
km accumulated
Cost with truing - [R$]
Cost with wheel replacement - [R$]
Total cost - [R$]
Unit cost - [R$/1000 km]
km accumulated
Cost with truing - [R$]
Cost with wheel replacement - [R$]
Total cost - [R$]
Unit cost - [R$/1000 km]
km accumulated
Cost with truing - [R$]
Cost with wheel replacement - [R$]
Total cost - [R$]
Unit cost - [R$/1000 km]
km accumulated
Cost with truing - [R$]
Cost with wheel replacement - [R$]
Total cost - [R$]
Unit cost - [R$/1000 km]
km accumulated
Cost with truing - [R$]
Cost with wheel replacement - [R$]
Total cost - [R$]
Unit cost - [R$/1000 km]
km accumulated
Cost with truing - [R$]
Cost with wheel replacement - [R$]
Total cost - [R$]
Unit cost - [R$/1000 km]
km accumulated
Cost with truing - [R$]
Cost with wheel replacement - [R$]
Total cost - [R$]
Unit cost - [R$/1000 km]
km accumulated
Cost with truing - [R$]
Cost with wheel replacement - [R$]
Total cost - [R$]
Unit cost - [R$/1000 km]
km accumulated
Cost with truing - [R$]
Cost with wheel replacement - [R$]
Total cost - [R$]
Unit cost - [R$/1000 km]
Figure 4.4(c): Cost Matrix
Once the desired place to go was known, the thing to be
done was to diminish the percentage of wheels with surface
defects which caused damages to the rails and required a
removal of great quantity of material to rebuild the original
profile. Thus, it was initiated a very deep and detailed
accompaniment to verify when the defects appeared and how
long they took to develop into a critic situation. Based on these
facts it was possible to plan a better wheel truing program.
o 4-31 x
After a certain time the results were not the expected, then
checking the data base it was noticed that the right amount of
wheels planned to be machined was being achieved but not the
right ones, it means that it was good on the quantitative aspect
but bad on the qualitative one. Then internal procedures had
to be changed and results were achieved.
With a reasonable percentage of wheels with defects, it
was possible to maximize other two important and dependent
variables: period to machine the wheels and depth of material
removed.
From Figure 4.5(c) to Figure 4.11(c) can be noticed the
numbers obtained for certain parameters that were considered
important during the development and implementation of this
management model.
Figure 4.5(c), shows the evolution of surface defects on
the wheels.
Set/98
Mai/98
Jan/98
Set/97
Mai/97
Set/96
Jan/97
Mai/96
Set/95
Jan/96
Mai/95
Set/94
Jan/95
Mai/94
Set/93
Jan/94
Set/92
Jan/93
Mai/93
Mai/92
Set/91
Jan/92
Mai/91
Set/90
Jan/91
Jan/90
Mai/90
Set/89
% Defects
Surface defects on wheels at EFC - From 1989 to 1998
50
45
40
35
30
25
20
15
10
5
0
Month
T3
E1
E2
E3
TOTAL
Figure 4.5(c): Wheels Defects - From 1989 to 1998
Figure 4.6(c) is a zoom on Figure 4.5(c) and shows better
how the defects were from 1994 until 1998, period on which
they diminished dramatically.
Surface defects on wheels at EFC - From 1994 to 1998
9
6
3
T3
E1
Month
E2
E3
Set/98
TOTAL
Figure 4.6(c): Wheels Defects - From 1994 to 1998
o 4-32 x
Nov/98
Jul/98
Mai/98
Mar/98
Jan/98
Set/97
Nov/97
Jul/97
Mai/97
Jan/97
Mar/97
Nov/96
Jul/96
Set/96
Mai/96
Jan/96
Mar/96
Nov/95
Jul/95
Set/95
Mai/95
Jan/95
Mar/95
Set/94
Nov/94
Jul/94
Mai/94
Mar/94
0
Jan/94
% Defect
12
Figure 4.7(c), shows the distribution of period for truing
wheels comparing the figures in 1994 and 1998. The average
increased but the most important is the standard deviation
diminishment which helps on previsions.
As in Figure 4.7(c), Figure 4.8(c) shows an improvement
on the average and also on the standard deviation for the
depth of material removed. Both together formed the actual
expected wheel life as shown in Figure 4.9(c).
Distribution - Period for truing the wheels
PER94
Percentage
PER98
Period - [Month]
Figure 4.7(c):. Distribution - Period for Truing the Wheels
Distribution - Depth of material removed
DEPTH94
Percentage
DEPTH98
Depth - [mm]
Figure 4.8(c): Distribution - Depth of Material Removed
o 4-33 x
Survival curve - Multiple wear wheel (MW)
100
% not scrapped
90
80
70
60
50
40
30
20
10
1500
1440
1380
1320
1260
1200
1140
1080
960
1020
900
840
780
720
660
600
540
480
420
360
300
240
180
60
120
0
0
Kilometers (km x 1000)
Figure 4.9(c): Survival Curve - Multiple Wear Wheel (MW)
Finally, Figures 4.10(c) and 11(c) show the consumption
of 38” wheels at CVRD/EFC. It is still irregular but better
than in the past and the amount is less, due to higher wheel
life, consequently costs with wheel acquisition are less too.
Quantity
Consumption of 38" wheels at EFC - From 1987 to 2000
14000
12000
10000
8000
6000
4000
2000
0
1987
1988
1989
1990
1991
1992
1993
1994
1995
1996
1997
1998
1999
2000
Year
Figure 4.10(c): Consumption of 38”at CVRD/EFC – Annually
Jul/00
Jan/00
Jul/99
Jan/99
Jul/98
Jan/98
Jul/97
Jan/97
Jul/96
Jan/96
Jul/95
Jan/95
Jul/94
Jul/93
Jan/94
Jan/93
Quantity
Consumption of 38" wheels at EFC - From 1993 to 2000
1600
1400
1200
1000
800
600
400
200
0
Month
Figure 4.11(c): Consumption of 38”at CVRD/EFC -Monthly
4.4(c) Rails
During Carajás Railway construction, it was used
approximately 120,000 metric ton of rails from four different
manufacturers.
o 4-34 x
The gradual increment in volume transportation, since the
beginning of Carajás Railway operation, has affected directly
rail life, causing a high tonnage of rail substitution.
Chronologically, many occurrences have been taking place,
according to what is manifested as follow.
4.4.1(c) History
4.4.1.1 (c) 1987
CSN rails started showing internal defects from manufacturing
process and also external superficial defects.
4.4.1.2(c) From 1987 to 1999
It was removed 39,766 metric ton (53.7%) of CSN rails being
33% because of internal and superficial defects and 67%
because of fatigue.
4.4.1.3(c) 1988
NKK and NSC rails, type NHH, started showing significant
superficial defects, leading CVRD/EFC to start changing
them.
4.4.1.4(c) 1990
CVRD/EFC started re-profiling process to remove superficial
defects and also adopted the use of ultrasonic inspection to
identify internal rails defects in order the improve rail life and
reduce operational risks. The estimated life of CSN rails was
450 MGT but the defects appeared when they accumulated
only 200 MGT.
Voest Alpine rails started showing superficial defects.
Twenty-thousand metric ton of Japanese rails, type DHH,
with different mechanical and metallurgical characteristics
from the previous one, were acquired and installed but the
same kind of defect seen before happened again.
4.4.1.5(c) 1991
CVRD/EFC began rail grinding process and thus it was
possible to reach 800 MGT for the rail life.
o 4-35 x
4.4.1.6(c) From 1993 to 1996
CVRD/EFC removed all Japanese rails.
4.4.1.7(c) 1997
CVRD/EFC installed 5,600 metric ton of Sydney Steel
Corporation rails without any heat treatment. Performance has
been good.
4.4.1.7. 1998
CVRD/EFC started installing Huta Katowice rails.
4.5(c) Looking for a Solution
In 1992, it was formed a multi functional group with people
from rolling stock maintenance, track maintenance and
operation and its main objective was to study the problem
related to wheel/rail interaction and get in touch with what
was being done all over the world.
The efforts of this group and the knowledge accumulated,
led CVRD/EFC to contract an international technology
center, Transportation Technology Center, Inc. (TTCI) in
1996 to work on CVRD’s heavier axle load implementation
program. But, before going to heavier axle loads it was
important to evaluate the impacts of this implementation to
assure the decision would represent a better result on the
bottom line. And the situations analyzed should be
accompanied by recommendations to attenuate impacts or to
maximize performance.
4.5.1. Introduction
The results and conclusions of several investigations
conducted on Carajás Railway formed the cornerstone of
subsequent engineering analyses on the benefits and cost
tradeoffs essential to business decisions based on the
economic consequences of moving to a heavier axle load.
During many years of research in North America on the
impact of heavier axle loads on train operations, vehicles and
track structure, many sound technical approaches have been
developed to optimize the investigations focused on 32.5- and
35-metric ton axle loads. TTCI has studied their effect on track
o 4-36 x
degradation and the optimal material selections, maintenance
practices and operating techniques that will achieve improved
safety, increase productivity, and minimize ton-kilometer costs
of transport.
Initially, wheel and rail life were optimized by balancing
the two damage mechanisms of wear and fatigue on the
CVRD system. The dynamic performance of the ore cars was
also optimized with use of analytical modeling and
instrumented wheel sets. Train handling practices are being
validated with the TOES (Train Operation and Energy
Simulation) model, with recommendations as appropriate to
reduce dynamic train forces for longer equipment lives,
reduced track maintenance costs, and enhanced operating
safety. The impact of heavier axle loads and increased train
weights on the track structure was subjected to evaluation
through TTCl's economic analysis software package. The
systems based analyses of the interaction between operating
practices, track structure impacts, and vehicle loading will
enable CVRD/EFC to choose appropriate engineering
strategies to accommodate safe, cost-beneficial migration to
32.5 metric ton axle.
This program's initial task was to address the various
concerns on the performance of wheel and rail under current
axle loads of 30.5 metric ton. TTCI addressed the problem of
wheel/rail life optimization on Carajás Railway based on the
AAR's experience in assisting North American railroads to
operate under heavy axle loads with minimal wheel and rail
deterioration problems. The AAR's approach focused on the
two damage mechanisms of wear and fatigue to be balanced on
the CVRD/EFC system. The Carajás Railway experienced
serious wheel shelling problems during the first ten years of
operation and took several measures to optimize wheel
performance prior to the program initiated in October 1996
between TTCI and Carajás Railway.
CVRD's wheel and rail life was optimized by evaluating:
axle alignment, lateral/longitudinal creepage, wheel wear and
false flange development, rail wear and profile grinding,
wheel/rail fatigue damage, and controlled track lubrication.
o 4-37 x
The dynamic performance of Carajás ore cars were
evaluated through a combination of test data and analytical
modeling. The truck performance was analyzed with reference
to axle alignment, steering and other truck parameters. The
truck design optimization was achieved through a combination
of analytical modeling and the use of instrumented wheel sets.
The impact of heavy axle load implementation on car
equipment, maintenance, and track components/maintenance
was considered in the study from the overall systems point of
view. The North American experience with respect to increase
in axle load from 26 to 33 tons (short ton) was marked by
significant increase in premature car component fatigue failure
such as center plate, bolster, and spring failures. To address
these problems, TTCI promoted the development of fatigue
analysis methods in North America contributing to improved
component designs and better cost estimation for maintenance
planning. TTCI proposes to use similar fatigue analysis
techniques for Carajás Railway to evaluate the life expectancy
of the current fleet with the suggested modifications by
col1ecting revenue service load data (truck bolster and side
bearing loads) planned in the program which must be
concluded during 2001.
From the track component point of view, the North
American experience with respect to increase in axle loads was
handled predominantly by increased maintenance efforts
pertaining to joints and fatigue related defects in rail.
TTCI used its proprietary Total Right-of-Way Analysis
and Costing System (TRACS) software to determine the
effects of heavy axle load traffic on the existing track structure
of Carajás Railway. TRACS combines the engineering-based
deterioration models with life-cycle costing techniques to
estimate track maintenance and renewal costs as a function of
Carajás' route geometry, track components, and traffic volume.
The predictions of TRACS consist of the costs and
maintenance requirements due to fatigue, wear and failure of
rail, wood ties, ballast, fasteners, and turnouts. The TRACS
results can be used by CVRD/EFC to optimize life-cycle
component costs in response to axle loads and traffic volume.
o 4-38 x
4.6(c) Methodology and Approach of TTCI's
Comprehensive Program On Carajás
Railway
The following are tasks related to wheel/rail life optimization
on Carajás Railway resulting in various recommendations for
improvements in car equipment, rail profile grinding, and
lubrication practices toward the implementation of higher axle
loads:
A - Onboard ore wagon truck performance measurements for
current design and existing axle loads (30.5 tonnes).
B - Wayside wheel/rail force measurements.
C - Ore wagon characterization tests and truck design review.
D - NUCARS (TTCI's vehicle dynamics model) simulations
of ore wagons under existing and increased axle loads.
C - NUCARS simulation with various arrangements of truck
side bearings.
D - Modification of existing truck design with warp resistance
retrofit for improved performance.
E - Full-scale testing ore wagons with truck modifications
supported by wayside wheel/rail force measurements and
onboard truck performance measurements.
F - Study of current rail grinding practices on Carajás
Railway.
G - Installation of two wayside rail lubricators on an
experimental basis.
H - Recommendations for controlled rail profile grinding on
an experimental basis at selected locations.
I - Supply of a pair of high precision load measuring wheel
sets.
J - Full-scale testing of standard and modified trucks with
instrumented wheel sets under 30.5- and 32.5-tonne axle
loads.
K - Implementation of TRACS and wheel life-cycle costing
model for the economic analysis of heavy axle loads and
improved truck suspension trucks on Carajás Railway.
o 4-39 x
4.7(c) Implementation of TTCI's Wheel/Rail Life
Optimization Program on Carajás Railway
This part of the paper specifically focuses on the truck
performance evaluation, track improvements, and economic
analysis related to the wheel/rail life optimization on Carajás
Railway.
4.7.1(c) Ore Wagon Truck Performance Evaluation
and Modeling
The initiation of this task was propelled by the fact that Carajás
Railway was experiencing severe wheel/rail shelling problems
under the current axle loading of 30.5 metric ton . A team of
TTCI's experts visited Carajás Railway and inspected the wheel
and rail defects, which are presented in Figures 4.12(c) and
13(c).
Figure 4.12(c): Wheel Shelling Typically Found
on the CVRD Ore Wagon Wheels
o 4-40 x
Figure 13(c): Photograph of Spalling on the
Surface of Rail Typically Found on the Carajás Railway
It was concluded that the main defect was the
development of small transverse surface cracks on the wheels
and rails from "ratcheting" strains. Ratcheting strain refers to
the accumulation of plastic strain on the wheel or rail surface
that is caused by creep forces generated by uni-directional
rolling/sliding. The surface cracks found on the wheels and rail
of the Carajás Railway had the topographical appearance of
cracks associated with ratcheting strains. The direction of
plastic flow of the rail material in the vicinity of the surface
cracks was consistent with the direction of longitudinal creep
forces that would normally be applied by the loaded car wheels
to the rails.
In brief, ratcheting strains cause cracks in the rail surface
by deforming the rail material near the surface such that the
carbide laminated and grain boundaries are aligned parallel to
the rail surface. As illustrated in Figure 14, the material
between the laminates experiences ductility exhaustion and
cracks as the plastic strains continue to accumulate. This
results in a series of shallow transverse cracks on the surface of
the rail.
o 4-41 x
Direction of longitudinal
creep forces applied to
rail by wheels
Flakes/Cracks
Ratchetting
Strain Deformation
Rail Material
Microstructure Laminates
Figure 14.(c): Ratcheting Strains in Rail Material Caused by Large
Longitudinal Creep Forces Between Wheel and Rail
Ratcheting strains are formed when large
longitudinal/lateral creep forces are applied to the wheels or
rails consistently in one direction. Because the Carajás Railway
is practically uni-directional, that is, the loaded trains travel
almost entirely from the Carajás mine to São Luís, and the
rolling directions of the wagons are changed only every 3
months, the longitudinal creepages applied to the wheel and
rails may be considered essentially uni-directional. The rate at
which ratcheting strains accumulate depends on several
factors, including the magnitude of the longitudinal and lateral
wheel/rail creepages and the level of wheel/rail friction.
A series of wayside measurement tests were conducted to
measure the vertical and later wheel/rail forces and wheel set
angles of attack associated with the CVRD/EFC ore wagons
to investigate the theory that ratcheting strains were causing
the surface cracks observed on the wheels and rails of the
Carajás Railway. Large longitudinal/lateral creep forces and
wheel set angles of attack are frequently associated with
wagons that have bogie turning or alignment problems.
The lateral wheel/rail forces of the CVRD/EFC ore
wagons that were measured during the wayside measurement
tests were found to be larger, on average, than similar wagons
on similar tracks in North America. This supported the idea
that large longitudinal lateral creep forces might be a factor
contributing to ratcheting strains. Also, the wheel set angles of
attack were found to be somewhat larger, on average, than
o 4-42 x
similar wagons on similar tracks in North America, especially
in tangent track. This supported the idea that the ore wagon
bogie turning resistance might be higher, on average, than
"normal" unit train North American wagons. During the
wayside measurement tests, several ore wagons were
characterized by exceptionally large lateral wheel/rail forces
and wheel set angles of attack. These wagons were selected for
the onboard measurement tests.
The onboard measurement tests were designed to measure
the bogie turning and warp characteristics directly as the
wagon traveled on the Carajás Railway. Bogie warp refers to
the distortion of the three-piece bogie such that the side
frames skew relative to the bolster and allow the wheel sets to
develop large angles of attack. The bogie bolster to wagon
body measurement indicated that the bogie did not turn
properly to maintain its wheel sets in alignment with most
curves and some tangent track section of the railway. Also, the
bogie warp measurement indicated that the bogie warped in
most curves and had a tendency to remain skewed in some
tangent track sections of the railway. Together, these results
suggested that the turning resistance of the bogie was too high
and prevented the bogie from running proper alignment with
the track. Figure 15 presents the bolster rotation and warp
angles measured in several curves during the on-board
measurement tests.
NUCARS modeling of Carajás Railway ore wagons was
carried out after the completion of characterization tests.
These tests involved measurement of natural vibration
frequencies for loaded and empty wagons in the following
modes: bounce, pitch, yaw, lower center roll, and upper center
roll. The data obtained in the form of natural frequencies were
converted into: roll inertia, pitch inertia, yaw inertia, vertical
truck stiffness, lateral truck stiffness, and center of gravity for
the loaded and empty wagons. The NUCARS analysis included
the weights of the bolsters into the secondary suspension
springs and the spring weight of the wagon. The above
information was used to develop NUCARS models for
simulating the performance of the ore cars with various
suspension modifications.
o 4-43 x
10.0
20.0
5.0
0.0
15.0
10.0
-5.0
-10.0
5.0
0.0
-5.0
-15.0
-20.0
-10.0
-15.0
-25.0
-30.0
-20.0
0
1000
2000
3000
4000
Time (sec.)
Figure 15(c): Bogie Bolster Rotation and Warp Angle
Measured in Several Curves during On-board Test
TTCI designed several suspension modifications, with a
number of constant contact side bearing (CCSB) arrangements
and a warp resistance truck design. AlI the variations of CCSB
modifications (preload of 2,500 lbs., 1,100 lbs., 3,000 lbs.
versus standard 3,600 lbs.) were evaluated by NUCARS
modeling and full-scale testing. The performance
improvements with CCSB modifications were found to be
marginal and as a result, TTCI proposed truck frame bracing
with roller side bearings and primary shear pads as the best
solution for Carajás Railway to reduce truck warping. The
results of full-scale testing with standard trucks and with warp
resistance retrofits are shown in Figures 4.16(c) and 17(c).
Figure 16(c): Warp and Bolster Roll Angles for
Loaded CVRD Standard Truck at 65.4 km/h
o 4-44 x
Figure 17(c): Warp and Bolster RoIl Angles for
Loaded Frame Brace Truck at 65.4 km/h
The previous results show that the standard trucks on
Carajás Railway were still warped and have not straightened
out properly on tangent sections after negotiating a curve;
whereas, frame bracing has significantly reduced the warp
angle and bolster rotation in both the curving and tangent
sections. The wayside wheel/rail force measurement stations
established by TTCI on Carajás Railway also indicated that the
standard trucks exhibited higher lateral loads in the curve and
tangent test sections when compared to warp resistance
design.
A prototype wagon with warp resistance design was put
into operation after the completion of the above tests. The
wheels of the test wagon were monitored since August 1997
and no defects were noticed on the wheels tread even though
they have completed in excess of 320,000 km. The normal life
of wheels between re-profiling for standard wagons is 160 to
170,000 km.
o 4-45 x
4.7.2(c) Full-scale Testing of Standard and Frame
Braced Trucks with Load Measuring
Wheel Sets
A pair of high precision load measuring instrumented wheel
sets (IWS) were specially constructed, calibrated, and
commissioned by TTCI on Carajás Railway. The main purpose
of IWS was to evaluate the differences in performance
between standard and warp resistance trucks to implement
higher axle loads on Carajás Railway. Both the load measuring
wheel sets were successfully commissioned by operating them
under loaded wagons at different speeds. The vertical and
lateral force measurements from IWS were compared to the
wayside wheel/rail force measurements and good correlation
between the two measurement systems.
The standard CVRD/EFC truck and a modified truck
equipped with frame bracing were tested over-the-road under
30.5 and 32.5-tonne axle loads. All the test runs were made
over a distance of 40 km. on Carajás Railway to evaluate the
wagon performance with and without truck modifications
under existing and increased axle loads. Figure 4.18(c) presents
the lateral force exceedance plot for all test configurations
between km posts 13 and 18. The lateral forces were plotted
for each configuration as a function of percentage of
exceedance. The standard CVRD/EFC truck's lateral forces
are greater than those obtained from the warp resistance
design under similar conditions of operation. As expected, the
lateral forces of standard truck under increased axle load were
far greater than warp resistance truck under similar operating
conditions. Figure 4.18(c) also indicates that the standard truck
under increased axle load has a problem of negotiating tangent
and curved sections as shown from the large negative lateral
force percentages seen here.
The frame-braced truck exhibited lower lateral forces
under increased axle load when compared to standard truck
under existing axle load (30.5-metric ton) conditions. There
seems to be a consistent offset of 1.500 lbs of total lateral
force between the two truck types.
o 4-46 x
Figure 4.18(c): Comparison of Total Truck Lateral Loads
Figure 4.19(c) presents the longitudinal force exceedance
plots for the lead axle of both truck types, under normal and
increased axle load conditions. Frame braced truck exhibited
far less longitudinal force for lead axle compared to standard
truck under all loading conditions, demonstrating that warp
resistance truck required less steering force to negotiate the
curves and tangent sections. TTCI's modifications for CVRD
truck clearly indicated that the strain ratcheting that is
happening with standard trucks is greatly reduced with warp
resistance design.
Figure 4.19(c): Comparison of Axle 1 Longitudinal Forces
o 4-47 x
From these results, it can be clearly seen that the warp
resistance modifications are producing lower lateral and
longitudinal steering forces when compared to the standard
bogie arrangements. From the percent exceedance plots, the
warp resistance over load and standard loaded result shows
lower lateral and longitudinal forces than the standard bogie
arrangements.
The main reason for this is the warp resistance bogie
arrangement does not warp; whereas, the standard bogie does.
This increased warp causes an angle of attack between the
wheel and rail and therefore results in higher lateral forces. The
larger lateral forces in turn causes a larger steering force to
compensate for the larger lateral forces. The wheel set is trying
to turn the axle back to a radial position, which results in
higher longitudinal creep, which in turn causes the ratcheting
effects seen in the wheel and rail surfaces.
4.8(c) Methodology of Recommendations for Rail
Grinding Practices
TTCl's experts visited Carajás Railway and noticed that apart
from wheel shelling problems. CVRD was experiencing a
serious rail defect and wear problem. Rail fatigue defects were
noticed to occur in several forms such as spalls, head checks.
squats, and shells. Battered welds were another source of
concern. Figures 4.20(c) and 21(c) present some of the rail
defects such as spalls and squats occurring on Carajás Railway.
Figure 4.20(c): SpaII Defects Occurring on Rails
in all Track Geometries
o 4-48 x
Figure 4.21(c): Typical Squat Defects
Under TTCI's comprehensive program for Carajás
Railway, various forms of rail defects were received including
the current rail maintenance practices. Under this program
TTCI and CVRD/EFC engineers developed a new rail profile
grinding practice combined with limited wayside lubrication.
This section of the paper describes the methodology
developed under TTCI's program for Carajás Railway to
minimize rail defects and optimize rail maintenance practices.
The main reason for rail fatigue defects occurring on
Carajás Railway was that most of the wheels in service were reprofiled within 100,000 km of wagon operation and conformal
shapes of wheel and rail could never be achieved. In this
scenario, the contact band remains narrow and the contact
stress remains high, ultimately causing the rail to fatigue in the
narrow contact region. Even though Carajás Railway has
mostly mild track curvatures, the wheels attain flanging as
shown in Figure 4.22(c), which illustrates new CVRD/EFC
standard (AAR 1:20, narrow flange) and a worn wheel profile
overlaid on a new 136 RE rail (recently installed on Carajás
Railway). Both new and worn wheels contact the rail almost in
the same location, when the flange nears the rail, just above the
gage corner. This may be the main reason for rail fatigue
defects, apart from the fact that excessive longitudinal creep
forces are developed by the warping of CVRD standard trucks.
o 4-49 x
Contact Location
on New Rail
CVRD Rail
Worn
CVRD Rail
Theoretical
136-14
136-14
Figure 4.22(c): Wheel/Rail Contact on New 136 RE Rail,
Concentrated on the Gage Side of the Rail with both
CVRD's New and Worn Wheel Profiles
TTCI evaluated the two primary concerns of rail profile
grinding shape of rail profile on tangent location and curves
and frequency of grinding. The main purpose of TTCI's
recommended rail grinding and profile maintenance practices
is reduce rail fatigue and rail wear with an optimum balance
while trying to achieve the following goals:
A - Promote proper vehicle dynamics by reducing hunting and
improving curving performance;
B - Protect rails from fatigue by grinding relief into those areas
of the railhead that can experience excessively high
contact pressures;
C - Minimize the contact pressures necessary to achieve Goal
Number 2 to protect rails and wheels from rolling contact
fatigue;
D - Minimize the wear rates necessary to protect rails from
fatigue.
Based on the above requirements, rail profile shapes were
re-engineered for Carajás Railway with a different philosophy
of rail grinding methodology verified at the Facility for
Accelerated Service Testing (FAST), Pueblo, Colorado, USA.
After a thorough review of current operating conditions
and the proposed changes under TTCI's program to be
implemented on Carajás Railway's rolling stock, a new rail
profile grinding methodology was recommended for curves
and tangent sections (Figures 4.23(c) and 24(c)). This method
o 4-50 x
was implemented at selected sites with success. Those profiles
were designed to promote curving, reduce hunting, protect
rails from high contact stresses, and protect rail from fatigue
and to minimize wear rates.
Figure 4.23(c): TTCI'S Recommended Profiles for
Grinding Maintenance of CVRD Rails
o 4-51 x
a
Km 10.0 Siding, Tangent
TTC Recommended Low Rail and Tangent Profile
b
Km 427 No Grind Test on CSN Rail, Low Rail
TTC Recommended Low Rail and Tangent Profile
c
Km 146.465 after rail planing, Low Rail
TTC Recommended Low Rail and Tangent Profile
Figure 4.24(c): TTCI Recommended
Low Rail/Tangent Profile
4.8.1(c) Lubrication Practice
TTCI's experience at FAST suggests that in conformal
wheel/rail situations, much of the wheel/rail surface fatigue,
wheel/rail wear, rail corrugation, and weld batter problems are
due to misaligned axles. The true source of the problem lies in
the creep forces that result from misalignments. When rails at
FAST were well lubricated, creep forces were reduced and all
the problems mentioned were reduced. When improved trucks
were installed and creep forces reduced even more, the
o 4-52 x
problems mentioned were further reduced, and in some cases
complete eliminated. There are some benefits to be gained by
lubrication, and more to be gained with improvement of truck
performance. Re-designing bogies is a long-term solution as it
will take time to learn the best way to modify the bogies and
institute the changes. TTCI and CVRD/EFC engineers
decided to investigate effects of creep force reduction through
a limited lubrication study.
In order to estimate effects of lubrication, NUCARS was
used to predict the contact forces that develop at the
wheel/rail interface with a loaded wagon, ground rails, and
current bogie arrangement. Longitudinal forces were predicted
since much of the wheel and rail damage observed along the
railway was due to excessive longitudinal forces.
NUCARS predicted that the longitudinal contact forces on
lubricated rail in tangent track are about one half that of dry
rail.
TTCI and CVRD/EFC engineers realized this is not as a
complete solution as modifying bogies can be. Reduction in
creep forces achieved through lubrication can give some
indication of what effect improved bogies will have.
In August 1997, TTCI and CVRD/EFC engineers
installed a wayside hydraulic rail lubricator system near curve
on Carajás Railway to begin a study of lubrication effects on
rail performance. The rail section adjacent to the lubricator was
ground to the TTCI recommended practice (both curve and
tangent sections). For comparison, another similar curve and
tangent sections were ground elsewhere to the TTCI
recommended practice also. This gave a back-to-back
comparison of the new TTCI practice under "lubricated" and
"dry" conditions. It was demonstrated with correct rail profile
grinding and lubrication a reduction in rail wear with
improvement in surface fatigue conditions.
o 4-53 x
4.8.2(c) Implementation of TRACS and the Wheel
Life-Cycle Costing Model
To assess the engineering and economic implications of
implementing lubrication, increased axle loads, and framebraced suspensions on the Carajás Railway, two sophisticated
models are used: TRACS and the Wheel Life-Cycle Costing
Model.
The TRACS Total Right-of-Way Analysis and Costing
System is a proprietary software package of track component
degradation models used to assess the engineering and
economic effects of degradation to rail, wood ties, ballast, and
turnouts. Its flexibility allows it to be used for a number of
analyses including:
A - Determining the engineering effects of traffic on track;
B - Assessing specific maintenance policies such as adding
track lubrication and changing to alternative component
technologies;
C - Specific costing analysis, such as determining the
increased track costs if additional cars are hauled on a specific
route;
D - Budgeting, such as how many miles of rail will have to
be replaced over a specific time period.
The Wheel Life-Cycle Costing Model is used to estimate
wheel lives and equivalent uniform annual costs (EUAC) for lwear, 2-wear and multi-wear wheels. This model can evaluate
the alternative costs of the three wheel types, as well as the
benefits of reduced wheel maintenance requirements
associated with the reduced rolling resistance achieved with
track lubrication and the implementation of warp resistance
trucks.
In the lubrication study, the TRACS Rail Wear Model was
used to determine the decreased rail wear effects of installing
lubricators on the São Luís-Carajás route. The analysis was
performed to estimate the rail wear reduction between "dry
rails" and "lubricated" rails. Since rail wear and wheel wear are
directly related, the rail wear improvements can be used as a
o 4-54 x
measure of expected wheel life improvements. Hence, the
interpreted results of the TRACS Rail Wear Model were used
as life improvement input data to the Wheel Life-Cycle Model.
The analysis estimates rail wear reduction ranging from
14,55 percent in tangent track to as much as 43,14 percent (839
m radius curved track). The estimated weighted average
reduction in rail wear and wheel wear for the entire São LuísCarajás route is 17,73 percent. In the analysis of the effects of
frame-braced suspensions, even greater improvements are
expected. Due to the elimination of wheel shelling, it is
expected that reductions of over 50 percent of current wheel
maintenance costs will be achieved.
4.9(c) Conclusions
For its Carajás Railway, CVRD has undertaken a multi-faceted
systems approach to wheel and rail life optimization under
heavy haul traffic. To lower maintenance costs and assume
reliability it was necessary to solve serious wheel and rail
degradation under 30.5-metric ton axle loads before the
productivity increases achievable through the introduction of
higher axle loads could be undertaken.
Specifically, the application of research and analyses has
been able to develop solutions for optimizing wheel and rail
life by managing the dynamic performance of the ore car fleet.
Recommended practice for Carajás operations have been
shown to achieve a pragmatic balance between wear and
fatigue of the rolling contact elements.
Rapid development of wheel shelling tread defects has
been eliminated in a test wagon. Frame brace design trucks
with roller side bearings and primary shear pads have shown
no tread defects after 320,000 kilometers, around two-fold
increase in service between re-profiling as compared to the
standard three-piece truck (160,000 - 170,000 km between reprofiling).
Another important point is that the good results with warp
resistance trucks encouraged CVRD/EFC to evaluate the
performance of this design in a sample of 24 wagons and,
again, the good results led CVRD/EFC to decide for the
o 4-55 x
installation of warp resistance in part of GDT fleet. At this
moment there are 750 wagons with warp resistance trucks and
warp resistances will de installed in more 820 wagons. It was
also possible to evaluate another alternatives like Swing Motion
truck which were installed in 20 wagons and good results were
obtained, too. In Table 4.1(c), there is an actual comparison
among standard truck, warp resistance truck and Swing
Motion truck performances.
Table 4.1(c): Comparison Among Standard Truck, Frame Braced
Truck and Swing Motion Truck Performances
Types of Trucks
Item
Unit
Re-profiling
period
Material
removed
Wheels’
average life *
Percentage of
defects
10 x
km
mm
3
3
10 x
km
%
Standard
160~170
Frame
Braced
300
Swing
Motion
260
4,0~5,0
5,0
5,0
1,250
2,200
1,900
<1
<1
<1
* Expectative for the new trucks design. The 20 wagons
equipped with Swing Motion trucks still running and figures still
changing.
Reductions in wheel and rail wear are direct benefits of the
improved curving performance; the modified trucks show
significant reductions in lateral and longitudinal steering forces
on curves and greatly improved tracking on tangents.
Once the severe wear environment imposed by the poorly
performing truck is eliminated, the severe rail wear and fatigue
damage consequences will diminish. This will permit
CVRD/EFC to adopt rail maintenance practices to achieve
more conformal wheel and rail profile matching. A rail profile
restoration practice tailored to the Carajás operation has been
developed. Selected test sites have demonstrated lower peak
contact stresses, facilitated curving, reduced hunting on
tangents, and a reduction in wear rate.
o 4-56 x
In combination with the rail profile restoration activity, a
judicious application of track lubrication has been shown to
reduce rail wear and substantially improve surface fatigue
conditions. The implementation of track lubrication has been
estimated to achieve an average rail wear reduction and wheel
wear-life improvement of 17 percent. The additional
improvements due to implementing frame-braced trucks are
expected to exceed this with rail wear reductions greater than
20 percent and wheel wear-life increases up to 100 percent.
With a program to retrofit the ore car fleet with improved
trucks, aggressive rail profile restoration and maintenance and
the introduction of rail friction level control through precision
lubrication, the Carajás Railway will be able to virtually
eliminate the wheel and rail problems now experienced under
30-metric ton axle loads and confidently implement 32.5- or
35-metric ton axle loads for a highly productive, safe, reliable
and low cost heavy haul system.
The efforts conducted by CVRD/EFC brought some
good results and many things were learned and now they are
part of a routine but much more can be achieved (and must
be).
Joining recommendations that come from research,
accumulated knowledge, new technologies and a good
management model, the results can be better, specially now, in
a global world on which costs must be reduced, efficiency and
quality maximized to offer the best product/service to the
clients aiming the company prosperity. The best practices of
the CVRD operation are summarized in tabular form in Table
4.2(c).
REFERENCES
1.
Fassarela L. J. V., Oliveira Neto L. E., Barros Filho A.,
Martins R. S., Costa R. J., Coelho Filho J. R., Rajkumar B.
R., Urban C. L. - Wheel/Rail life optimization with the
implementation of increased axle loads on Carajás Railway,
Brazil.
o 4-57 x
2.
3.
4.
5.
Souza D. J. , Martins R.S. - Ações para elevação da vida útil
de rodas ferroviárias na EFC, Segundo Seminário de
Tecnologia Ferroviária da CVRD.
Barros Filho, A - Ensaios comparativos em rodas ferroviárias.
Kalousek J., Magel E. - Optimizing wheel/rail interaction,
Railway Track and Structures, January 1997.
Stone D. H., Moyar G. J., Guins T. S. - An interpretative
review of railway.
o 4-58 x
Click Here To Go Back To Table of Contents
PART 4(d): Quick Reference Tables for
Basic Heavy Haul Rail System Design
Prepared at the Transportation Technology Center, Inc., USA.
Written by Dr. James Lundgren, member of the Technical Review
Committee
4.1(d) Introduction
The concepts embodying the primary components of the track
and vehicle systems characteristic of heavy haul operations has
been introduced with the case studies presented for the
BHPIO, CPR and CVRD examples. The demands on the
components comprising the wheel and rail system are highly
dependent upon the total environment the transport demand
places on them. This includes not only topography, climate,
native soil conditions, but also traffic levels, axle loads, train
lengths, track gauge, annual tonnages, vehicle design and
similar factors.
The objective of these reference tables is to present in
concise form, reference charts that may be used to gain a "firstcut" or preliminary concept of what might constitute an
appropriate starting point for a reliable and economical heavy
haul operation. As illustrated in the case studies, there are
many opportunities to address refinements and adjustments to
the base cases. In most instances the operation will grow from
the starting point in response to specific conditions and
modifications and will evolve into a more specific system
"tuned" to the unique features of the environment, traffic
conditions and maintenance options. Consequently, much
greater productivity may be achieved from the fine tuning of
system operations beyond these base case recommendations.
As illustrated in the BHP Iron Ore case study, much can be
achieved by continually observing, measuring and
incrementally improving a particular system and its
performance characteristics.
In light of the wide spectrum of conditions and
environments likely to impact any specific heavy haul
operation, the authors have chosen to present the general
guidelines by referencing axle loads, terrain and traffic density.
o 4-59 x
Four axle load categories of 35+, 30-34, 25-29, and 20-24
metric tonnes; two terrain types represented by curvatures less
or greater than 875 meters are intended to encompass severe
operating conditions (curvature and gradient) and less
challenging, more gentle terrain by lines having generous
curvature design; and three traffic severity levels have been
chosen for cataloging the fundamental cases with generic
guidelines.
Although more categories and finer divisions of the
operating characteristics could be made, the current set is
believed to provide a useful guide to positioning any particular
operation. From these starting points, the individual
characteristics of the operation need to be reviewed and
appropriate adjustments made. The table recommendations
may be relied upon to provide a safe and dependable
operation, but not necessarily the optimum achievable with
further study and adjustments based on experience. To
achieve world class performance, any heavy haul system will
require close study, hands on control and precise, diligent
maintenance practices. Much of this assistance is available
directly through recognized consulting bodies, or more
indirectly through the knowledge and experience base
contained within the collection of IHHA proceedings, papers
and conference documents. The CVRD case study presented
earlier illustrates this Progressive stepwise approach to an
optimal system.
Although the tables of recommended practice concern
themselves with dedicated heavy haul services, they are
generally applicable to mixed traffic lines (as represented in the
Canadian Pacific Railway case study). The heavy haul traffic
will generally drive the robustness of the track structure chosen
and is likely to be the major cause of wear and degradation.
Additional traffic applied will likely bring a number of
other demands: ride quality as an example. Rigid geometry
standards required for higher speeds may be imposed. More
frequent inspection and maintenance may well be essential for
passenger and dangerous cargo traffic using the line. In any
situation with mixed traffic types, the non heavy haul traffic
will bring increased complexity of interactions between wheel
o 4-60 x
and rail. Full optimization will be handicapped by the
inclusion of other vehicle types and their performance
attributes and demands on the system. The degree of
optimization is likely to be somewhat less than that achievable
in the closed loop systems dedicated to a single heavy haul
operation.
4.2(d) Using the Design Tables:
As indicated, the various tables are initial starting points for the
design and operation of a heavy haul line that meets the traffic
criteria targeted. The component recommendations will
provide for a safe and reliable operation, but may not
necessarily be the best long term solution. This can be
achieved only through a careful measurement and observation
of the wear and degradation that occurs under traffic
experience. As illustrated in the specific case studies discussed
earlier continuous improvement in response to the
maintenance demands and wear conditions will lead to a finely
tuned, highly efficient operation for the local conditions and
the specific needs of the operation.
The General Notes section of the tables should be
reviewed in conjunction with their use. Further definitions and
explanations of key features of the components,
recommendations and descriptions are provided in the Table
Notes section. Both precede the tables.
4.3(d) References
For many components, specific detailed technical
specifications or engineering recommendations are available
for consultation. Examples of these include the American
Railway Engineering & Maintenance-of-Way Association
(AREMA) Manual of Recommended Practice for track
construction (more detailed specifications for rail, tie, fastener,
ballast, etc.); the Association of American Railroads (AAR)
Mechanical Division Manual of Standard Practice (wheels,
axles, bearings, vehicle fatigue, etc.); the International Union of
Railways (UIC) codes; and specific standards developed by the
various heavy haul railways as sound practice.
o 4-61 x
4.4(d) General Notes:
1. Heavy Haul (HH) Operation: HH traffic is defined as 25
tonne or greater axle loads, 20 million gross tonnes (MGT)
annual traffic on the line or the operation of trains in
excess of 5,000 gross tonnes.
2. Table Values are recommended minimum values or
practices.
3. Recommendations are based on an assumption of nominal
railway track and equipment design convention: separate
locomotive units hauling 4-axle twin wagons with track
gauges of 1000 to 1600 mm.
4. Table recommendations are to be considered appropriate
starting points for system design with appropriate
modifications to compensate for local operating
conditions, environmental factors and equipment
deficiencies.
5. Table entries are common practice norms; however, other
components or designs achieving comparable engineering
performance may be substituted.
6. Track and vehicle components work as a system. In
special circumstances, particularly robust components may
allow relaxation in strength/performance of others (e.g. a
superb formation, ballast and sleeper system may permit
the use of lighter rail sections). However, the overall
economy of the system may be compromised.
7. Particular environments and maintenance practices may
suggest modifications to the recommended values (e.g.
severe contamination conditions, such as blowing sand
may require avoidance of lubrication).
8. Experimental, new, or untried designs, configurations, or
materials are not recommended for widespread adoption
or introduction into heavy haul lines until their long-term
service life has been demonstrated.
9. Metric units: metric units are assumed in the tables where
not specified: e.g. axle loads are metric tonnes.
10. Rail selection: the tables reflect general application
recommendations. Optional choices include using a
heavier or lighter section to achieve lower maintenance or
o 4-62 x
lower construction costs respectively. This may be judged
in light of the expected life of the line (e.g. ore body) the
tolerance for and cost of increased maintenance activity,
and the funds available, short or long term. Depending on
the length and spacing of the various curve and tangent
sections of the line, it may be feasible to use different rail
quality on various line segments (e.g. premium on curves,
standard in tangent) to achieve the economic optional
solution. The principle could be applied to other track
components as well to create several different "standards"
within a line. This practice is not generally considered
economic except for rail selection or for line characteristics
having long uninterrupted tangents as maintenance and
inventory complications.
11. Longitudinal rail forces arising from train operation or
thermal effects must be accommodated within any track
structure design. The selection of appropriate fastener
systems, including "anchoring" or rail in conventional
spike fastened wood tie systems, plays a significant part of
a robust track system. The ballast shoulder nominal
recommendations in the tables contribute significantly to
the lateral resistance of the track structures to thermal
buckling or "sun-kinks." Expected maximum temperature
differentials, the appropriate selection of a rail laying
neutral temperature and anticipated longitudinal train
forces all play a role in the selection of a track structure
"lateral" design strength that will provide safe, economic
service.
4.5(d) Table Notes: (brief descriptions of salient
features of component classes)
1. TRACK STRUCTURES:
For heavy haul service it is desirable to avoid curvature less
than 500 m; gradients in excess of one percent; geometric
quality appropriate FRA Class 5 or equivalent should be
achieved.
2. RAIL:
§ Super Premium = premium rail chemistry with
hardening (heat-treated, very fine pearlitic):
o 4-63 x
BHN ≥ 388; Rc≥ 42
§ Premium = "premium" rail chemistry and/or hardening
(micro-alloy, fine grained, pearlitic):
BHN = 341-388; Rc≥ 36.5
§ Standard = standard carbon rail chemistry:
BHN > 300-340; Rc≥ 32
All HH applications require CWR for economic maintenance.
3. CROSSTIES:
§ Monoblock concrete: AREA (AREMA) specifications
or equivalent
§ Premium = Treated hardwood or approved softwood:
approximate dimensions (W x D x L) of 0.12 x 0.16 x
(1.7-1.8) ratio of track gauge
4. FASTENERS:
§ Premium: elastic: spring clips or equivalent providing
resilient vertical clasping force and longitudinal
constraint. (Many proprietary designs are available.)
§ Nominal wood spikes: "nails" driven into wooden
crossties as in traditional cut or "dog" spike. A more
robust wood tie fastening may be achieved with "lag" or
coach screws or threaded drive spikes.
§ Pads should be designed with appropriate stiffness and
damping properties for traffic loads.
5. BALLAST:
§ Crushed rock: angularity, gradation, abrasion resistance
and cementing characteristics should be carefully chosen
for "high" end performance.
6. WHEELS:
Premium = heat treated (HT) (e.g. the AAR Class "C" rim
quenched wheel represents a nominal standard, having a
curved plate design to be compatible with rail profile
selected). An in-depth discussion of rail and wheel profile
matching is found in Part 2. The AAR 1B or an equivalent
simulated "worn" profile can be used as a starting point for
RE rail sections; the UIC standard should be used with
UIC sections as the initial selection. For the heavier axle
o 4-64 x
loads (35 plus tonnes) and higher traffic densities (over 50
MGT), it is recommended that specifically designed wheel
and rail profiles be developed for the given operation
through careful monitoring and experimentation. The
basic principles are explored in Part 2 pages 45 to 59.
7. BOGIES:
§
§
§
Standard 3-piece: with tight assembly tolerance
Improved standard: shear pads, cross-braced, spring
tray or plank designs
Premium: radial or steering trucks of various
proprietary designs
8. PROFILE MAINTENANCE
Part 5 suggests monitoring protocols
§ Wheel: hollow, roundness corrugation, flange wear
tolerances chosen for compatibility with service
demands
§ Rail: refer to Part 5, Section 5.8 for optimizing system
(wheel/rail)
9. LUBRICATION:
Managed appropriate lubrication application methods and
lubricants. Special circumstances may dictate no
lubrication (e.g. sand). Part 5, Section 5.7 addresses
lubrication.
10. SWITCH & CROSSING WORK:
Mainline 1:20. Moveable frogs, undercut switch points as
minimum specification for heavy services.
11. SPEED: operating speeds of HH are in the speed range
up to 80 kph (40-50 mph). Higher speeds or mixed traffic
operations may dictate variances from table
recommendations.
12. WEAR LIMITS: refer to Part 5, Section 5.8
13. FLAW DETECTION: refer to Part 5, Section 5.8
14. CONDITION MONITORING:
§ Wayside systems to track performance trends of
individual vehicles is recommended particularly at
higher tonnages with axle loads.
o 4-65 x
§
Automated onboard inspection vehicle measurement
systems (geometry, rail profile, etc.) recommended for
track structure monitoring particularly at higher
tonnage and loads
15. GEOMETRY AND RAIL FLAW INSPECTION:
§ Recommended intervals are for track in good
conditions.
§ Intervals should be adjusted (shortened) for older
infrastructures and unusual conditions.
o 4-66 x
Table 4.1(d) TRAFFIC MIX: DEDICATED HH
AXLE LOAD: 35+
TERRAIN: < 875 meter radius
TRAFFIC DENSITY: 50+ MGT
TRAFFIC
DENSITY
>50 MGT
TYPE
Premium
in
tangent
Super
premium
in curves
WEIGHT
136 RE
or
UIC 60
WHEELS
TYPE
PROFILE
Prem HT
curve plate
AAR Class C
1000mm ø or
equivalent
BALLAST
M/L
SWITCH &
CROSSING
WORK
GEOMETRY
INSPECTION
FLAW
INSPECTION
350 mm
select
crushed
rock
+ 200 mm
sub-ballast
300 mm
shoulders
Premium
rail
tangential,
movable or
closing
point frog
~ 3-6 month
intervals with
rail profile
monitoring
~3 month
intervals
RAIL
specifically
designed
(See
Section 2.5)
CROSSTIES
PREM:
WOOD or
CONCRETE
MONOBLOC
Nominal
spacing
490 mm wood
600 mm
concrete
BOGIES
Improved
Standard 3piece or selfsteering
FASTENERS
ELASTIC
with
elastomeric
pads
PROFILES & MAINTENANCE
WHEEL
RAIL
limit hollow
wear to 2 mm
Periodic grind to remove
corrugation and surface
damage and to treat rail
joints (material flow)
LUBRICATION
WHEEL/RAIL
Curves: Gauge face
ì<0.25-0.30
Rail head ì<0.350.40
Äì<0.10-0.15 L-R
Tangent: rail head
ì>0.35
WEAR
LIMITS
Measure
frequently to
ensure
economic
optimum
Table 4.2(d) TRAFFIC MIX: DEDICATED HH
AXLE LOAD: 35+
TERRAIN: < 875 meter radius
TRAFFIC DENSITY: 30-49 MGT
TRAFFIC
DENSITY
30-49
MGT
TYPE
FASTENERS
BALLAST
M/L SWITCH
&
CROSSING
WORK
GEOMETRY
INSPECTION
FLAW
INSPECTION
ELASTIC with
elastomeric
pads
350 mm
select
crushed
rock
+ 200 mm
sub-ballast
300 mm
shoulders
Premium
rail
Tangential,
spring point
premium
frog
~ 4-6 months
with profile
monitoring
~4 month
intervals
RAIL
TYPE
Premium
in
tangent
Super
premium
in curves
WEIGHT
136 RE
or
UIC 60
WHEELS
PROFILE
Prem HT
curve plate
AAR Class C
1000mm ø or
equivalent
specifically
designed
(see Part 2,
Section 2.5)
CROSSTIES
PREM: WOOD
or CONCRETE
MONOBLOC
Nominal spacing
490 mm wood
600 mm
concrete
BOGIES
Improved
Standard 3piece or selfsteering
PROFILES & MAINTENANCE
WHEEL
RAIL
limit hollow
wear to 2 mm
Periodic grind to remove
corrugation and surface
damage and to treat rail
joints (material flow)
LUBRICATION
WHEEL/RAIL
Curves: Gauge face
ì<0.25-0.30
Rail head ì<0.350.40
Äì<0.10-0.15 L-R
Tangent: rail head
ì>0.35
WEAR
LIMITS
Measure
frequently to
ensure
economic
optimum
Table 4.3(d) TRAFFIC MIX: DEDICATED HH
AXLE LOAD: 35+
TERRAIN: < 875 meter radius
TRAFFIC DENSITY: 20-29 MGT
TRAFFIC
DENSITY
20-29
MGT
BALLAST
M/L SWITCH
&
CROSSING
WORK
GEOMETRY
INSPECTION
350 mm
select
crushed
rock
+ 200 mm
sub-ballast
300 mm
shoulders
Premium
rail
Tangential,
spring point
premium
frog
~ 6 months
with profile
monitoring
RAIL
TYPE
Premium
in
tangent
Super
premium
in curves
WEIGHT
136 RE
or
UIC 60
WHEELS
TYPE
PROFILE
Prem HT
curve plate
AAR Class C
1000mm ø or
equivalent
specifically
designed
(see Part 2,
Section 2.5)
CROSSTIES
FASTENERS
PREM: WOOD
or CONCRETE
MONOBLOC
Nominal Spacing
490mm wood
600 mm
concrete
BOGIES
Improved
Standard 3piece or selfsteering
ELASTIC —
curves
Elastic or
spikes
—tangent
PROFILES & MAINTENANCE
WHEEL
RAIL
limit hollow
wear to 2 mm
Periodic grind to remove
corrugation and surface
damage and to treat rail
joints (material flow)
LUBRICATION
WHEEL/RAIL
Curves: Gauge face
ì<0.25-0.30
Rail head ì<0.350.40
Äì<0.10-0.15 L-R
Tangent: rail head
ì>0.35
FLAW
INSPECTION
~6 month
intervals
WEAR
LIMITS
Measure
frequently to
ensure
economic
optimum
Table 4.4(d) TRAFFIC MIX: DEDICATED HH
AXLE LOAD: 30-34
TERRAIN: < 875 meter radius
TRAFFIC DENSITY: 50+ MGT
TRAFFIC
DENSITY
TYPE
Standard
tangent
>50 MGT
premium
in curves
WEIGHT
136 RE
or
UIC 60
WHEELS
TYPE
PROFILE
Prem HT
curve plate
AAR Class C
900mm ø
BALLAST
M/L
SWITCH &
CROSSING
WORK
GEOMETRY
INSPECTION
FLAW
INSPECTION
300 mm
select
crushed
rock
+ 200 mm
sub-ballast
300 mm
shoulders
Premium
rail
Tangential,
spring point
premium
frog
~ 3-6 months
with profile
monitoring
~3 month
intervals
RAIL
specifically
designed
(see Part 2,
Section 2.5)
CROSSTIES
PREM: WOOD
or CONCRETE
MONOBLOC
Nominal Spacing
500mm wood
600 mm
concrete
BOGIES
Improved
Standard 3piece or selfsteering
FASTENERS
ELASTIC
—curves
Elastic or
spikes
—tangent
PROFILES & MAINTENANCE
WHEEL
RAIL
limit hollow
wear to 2 mm
Periodic grind to remove
corrugation and surface
damage and to treat rail
joints (material flow)
LUBRICATION
WHEEL/RAIL
Curves: Gauge face
ì<0.25-0.30
Rail head ì<0.350.40
Äì<0.10-0.15 L-R
Tangent: rail head
ì>0.35
WEAR
LIMITS
Measure
frequently to
ensure
economic
optimum
Table 4.5(d) TRAFFIC MIX: DEDICATED HH
AXLE LOAD: 30-34
TERRAIN: < 875 meter radius
TRAFFIC DENSITY: 30-49 MGT
TRAFFIC
DENSITY
RAIL
TYPE
Standard
tangent
30-49 MGT
premium
in curves
WEIGHT
136 RE
or
UIC 60
WHEELS
TYPE
PROFILE
Prem HT
curve plate
AAR Class C
900mm ø
AAR 1B or
equivalent
CROSSTIES
PREM: WOOD
or CONCRETE
MONOBLOC
Nominal Spacing
500mm wood
600 mm
concrete
BOGIES
Improved
Standard 3piece or selfsteering
FASTENERS
ELASTIC
—curves
Elastic or
spikes
—tangent
BALLAST
300 mm
select
crushed
rock
+ 200 mm
sub-ballast
300 mm
shoulders
M/L
SWITCH &
CROSSING
WORK
Premium
rail
Fixed point
premium
frog
PROFILES & MAINTENANCE
WHEEL
RAIL
limit hollow
wear to 3 mm
Periodic grind to remove
corrugation and surface
damage and to treat rail
joints (material flow)
GEOMETRY
INSPECTION
FLAW
INSPECTION
~ 4-6 months
with profile
monitoring
~4 month
intervals
LUBRICATION
WHEEL/RAIL
Curves: Gauge face
ì<0.25-0.30
Rail head ì<0.350.40
Äì<0.10-0.15 L-R
Tangent: rail head
ì>0.35
WEAR
LIMITS
Measure
frequently to
ensure
economic
optimum
Table 4.6(d) TRAFFIC MIX: DEDICATED HH
AXLE LOAD: 30-34
TERRAIN: < 875 meter radius
TRAFFIC DENSITY: 20-29 MGT
TRAFFIC
DENSITY
RAIL
TYPE
Standard
tangent
20-29 MGT
premium
in curves
WEIGHT
136 RE
or
UIC 60
WHEELS
TYPE
PROFILE
Prem HT
curve plate
AAR Class C
900mm ø
AAR 1B or
equivalent
CROSSTIES
PREM: WOOD
or CONCRETE
MONOBLOC
Nominal Spacing
500mm wood
600 mm
concrete
BOGIES
Improved
Standard 3piece or selfsteering
FASTENERS
ELASTIC
—curves
Elastic or
spikes
—tangent
BALLAST
250 mm
select
crushed
rock
+ 100 mm
sub-ballast
300 mm
shoulders
M/L
SWITCH &
CROSSING
WORK
Premium
rail
Fixed point
premium
frog
PROFILES & MAINTENANCE
WHEEL
RAIL
limit hollow
wear to 3 mm
Periodic grind to remove
corrugation and surface
damage and to treat rail
joints (material flow)
GEOMETRY
INSPECTION
~ 6 months
with profile
monitoring
LUBRICATION
WHEEL/RAIL
Curves: Gauge face
ì<0.25-0.30
Rail head ì<0.350.40
Äì<0.10-0.15 L-R
Tangent: rail head
ì>0.35
FLAW
INSPECTION
~6 month
intervals
WEAR
LIMITS
Measure
frequently to
ensure
economic
optimum
Table 4.7(d) TRAFFIC MIX: DEDICATED HH
AXLE LOAD: 25-29
TERRAIN: < 875 meter radius
TRAFFIC DENSITY: 50+ MGT
TRAFFIC
DENSITY
TYPE
Standard
tangent
>50 MGT
premium
in curves
WEIGHT
132 RE
or
UIC 60
WHEELS
TYPE
PROFILE
Prem HT
curve plate
AAR Class C
900mm ø
BALLAST
M/L
SWITCH &
CROSSING
WORK
GEOMETRY
INSPECTION
FLAW
INSPECTION
300 mm
select
crushed
rock
+ 100 mm
sub-ballast
300 mm
shoulders
Tangential
Fixed point
frog
~ 3-6 months
with profile
monitoring
~4 month
intervals
RAIL
AAR 1B or
equivalent
CROSSTIES
PREM: WOOD
or CONCRETE
MONOBLOC
Nominal Spacing
500mm wood
600 mm
concrete
BOGIES
Improved
Standard 3piece or selfsteering
FASTENERS
ELASTIC
—curves
Elastic or
spikes
—tangent
PROFILES & MAINTENANCE
WHEEL
RAIL
Periodic grind to remove
corrugation and surface
damage and to treat rail
limit hollow
joints (material flow)
wear to 3 mm
Differential between
curve and tangent
to treat rail joints
LUBRICATION
WHEEL/RAIL
Curves: Gauge face
ì<0.25-0.30
Rail head ì<0.350.40
Äì<0.10-0.15 L-R
Tangent: rail head
ì>0.35
WEAR
LIMITS
Measure
frequently to
ensure
economic
optimum
Table 4.8(d) TRAFFIC MIX: DEDICATED HH
AXLE LOAD: 25-29
TERRAIN: < 875 meter radius
TRAFFIC DENSITY: 30-49 MGT
TRAFFIC
DENSITY
30-49 MGT
RAIL
TYPE
Standard
WEIGHT
132 RE
or
UIC 60
WHEELS
TYPE
PROFILE
Prem HT
curve plate
AAR Class C
900mm ø
AAR 1B or
equivalent
CROSSTIES
PREM: WOOD
or CONCRETE
MONOBLOC
Nominal Spacing
500mm wood
600 mm
concrete
BOGIES
Improved
Standard 3piece or selfsteering
FASTENERS
ELASTIC
—curves
Elastic or
spikes
—tangent
BALLAST
250 mm
select
crushed
rock
M/L
SWITCH &
CROSSING
WORK
GEOMETRY
INSPECTION
FLAW
INSPECTION
Fixed point
frog
~ 4-6 months
with profile
monitoring
~4 month
intervals
300 mm
shoulders
PROFILES & MAINTENANCE
WHEEL
RAIL
Periodic grind to remove
corrugation and surface
damage and to treat rail
limit hollow
joints (material flow)
wear to 3 mm
Differential between
curve and tangent
to treat rail joints
LUBRICATION
WHEEL/RAIL
Curves: Gauge face
ì<0.25-0.30
Rail head ì<0.350.40
Äì<0.10-0.15 L-R
Tangent: rail head
ì>0.35
WEAR
LIMITS
Measure
frequently to
ensure
economic
optimum
Table 4.9(d) TRAFFIC MIX: DEDICATED HH
AXLE LOAD: 25-29
TERRAIN: < 875 meter radius
TRAFFIC DENSITY: 20-29 MGT
TRAFFIC
DENSITY
20-29 MGT
RAIL
TYPE
Standard
WEIGHT
132 RE
or
UIC 60
WHEELS
TYPE
PROFILE
Prem HT
curve plate
AAR Class C
900mm ø
AAR 1B or
equivalent
CROSSTIES
PREM: WOOD
or CONCRETE
MONOBLOC
NOMINAL:
500mm wood
600 mm
concrete spacing
BOGIES
Improved
Standard 3piece or selfsteering
FASTENERS
ELASTIC
—curves
Elastic or
spikes
—tangent
BALLAST
250 mm
select
crushed
rock
M/L
SWITCH &
CROSSING
WORK
GEOMETRY
INSPECTION
Fixed point
frog
~ 6 months
with profile
monitoring
FLAW
INSPECTION
~6 month
intervals
250 mm
shoulders
PROFILES & MAINTENANCE
WHEEL
RAIL
Periodic grind to remove
corrugation and surface
damage and to treat rail
limit hollow
joints (material flow)
wear to 3 mm
Differential between
curve and tangent
to treat rail joints
LUBRICATION
WHEEL/RAIL
Curves: Gauge face
ì<0.25-0.30
Rail head ì<0.350.40
Äì<0.10-0.15 L-R
Tangent: rail head
ì>0.35
WEAR
LIMITS
Measure
frequently to
ensure
economic
optimum
Table 4.10(d) TRAFFIC MIX: DEDICATED HH
AXLE LOAD: 20-24
TERRAIN: < 875 meter radius
TRAFFIC DENSITY: 50+ MGT
TRAFFIC
DENSITY
>50 MGT
RAIL
TYPE
Standard
WEIGHT
115 RE
or
UIC 54
WHEELS
TYPE
PROFILE
Prem HT
curve plate
AAR Class C
830mm ø
AAR 1B or
equivalent
CROSSTIES
PREM: WOOD
or CONCRETE
MONOBLOC
NOMINAL:
spacing 500mm
wood 600 mm
concrete
BOGIES
Improved
Standard 3piece
FASTENERS
ELASTIC
—curves
Elastic or
spikes
—tangent
BALLAST
250 mm
select
crushed
rock
M/L
SWITCH &
CROSSING
WORK
GEOMETRY
INSPECTION
FLAW
INSPECTION
Tangential,
Fixed point
frog
~ 6-8 months
with profile
monitoring
~6 month
intervals
300 mm
shoulders
PROFILES & MAINTENANCE
WHEEL
RAIL
limit hollow
wear to 3 mm
Periodic grind to remove
corrugation and surface
damage
LUBRICATION
WHEEL/RAIL
Curves: Gauge face
ì<0.25-0.30
Rail head ì<0.350.40
Äì<0.10-0.15 L-R
Tangent: rail head
ì>0.35
WEAR
LIMITS
Measure
frequently to
ensure
economic
optimum
Table 4.11(d) TRAFFIC MIX: DEDICATED HH
AXLE LOAD: 20-24
TERRAIN: < 875 meter radius
TRAFFIC DENSITY: 30-49 MGT
TRAFFIC
DENSITY
30-49 MGT
RAIL
TYPE
Standard
WEIGHT
115 RE
or
UIC 54
WHEELS
TYPE
PROFILE
Prem HT
curve plate
AAR Class C
830mm ø
AAR 1B or
equivalent
CROSSTIES
WOOD or
CONCRETE
MONOBLOC
NOMINAL:
spacing 500mm
wood 600 mm
concrete
BOGIES
Improved
Standard 3piece
FASTENERS
ELASTIC
—curves
Elastic or
spikes
—tangent
BALLAST
250 mm
select
crushed
rock
M/L
SWITCH &
CROSSING
WORK
GEOMETRY
INSPECTION
Fixed point
frog
~ 8-10
months
with profile
monitoring
300 mm
shoulders
PROFILES & MAINTENANCE
WHEEL
RAIL
limit hollow
wear to 3 mm
Periodic grind to remove
corrugation and surface
damage
LUBRICATION
WHEEL/RAIL
Curves: Gauge face
ì<0.25-0.30
Rail head ì<0.350.40
Äì<0.10-0.15 L-R
Tangent: rail head
ì>0.35
FLAW
INSPECTION
~8 month
intervals
WEAR
LIMITS
Measure
frequently to
ensure
economic
optimum
Table 4.12(d) TRAFFIC MIX: DEDICATED HH
AXLE LOAD: 20-24
TERRAIN: < 875 meter radius
TRAFFIC DENSITY: 20-29 MGT
TRAFFIC
DENSITY
20-29 MGT
RAIL
TYPE
Standard
WEIGHT
115 RE
or
UIC 54
WHEELS
TYPE
PROFILE
Prem HT
curve plate
AAR Class C
830mm ø
AAR 1B or
equivalent
CROSSTIES
WOOD or
CONCRETE
MONOBLOC
NOMINAL:
spacing 500 mm
wood 600 mm
concrete
BOGIES
Improved
Standard 3piece
FASTENERS
ELASTIC
—curves
Elastic or
spikes
—tangent
BALLAST
250 mm
select
crushed
rock
M/L
SWITCH &
CROSSING
WORK
GEOMETRY
INSPECTION
Fixed point
frog
~ 8-10
months
with profile
monitoring
250 mm
shoulders
PROFILES & MAINTENANCE
WHEEL
RAIL
limit hollow
wear to 3 mm
Periodic grind to remove
corrugation and surface
damage
LUBRICATION
WHEEL/RAIL
Curves: Gauge face
ì<0.25-0.30
Rail head ì<0.350.40
Äì<0.10-0.15 L-R
Tangent: rail head
ì>0.35
FLAW
INSPECTION
~8-10
month
intervals
WEAR
LIMITS
Measure
frequently to
ensure
economic
optimum
Table 4.13(d) TRAFFIC MIX: DEDICATED HH
AXLE LOAD: 35+
TERRAIN: ≥ 875 meter radius
TRAFFIC DENSITY: 50+ MGT
TRAFFIC
DENSITY
>50 MGT
RAIL
TYPE
Premium
WEIGHT
136 RE
or
UIC 60
WHEELS
TYPE
PROFILE
Prem HT
curve plate
AAR Class C
1000mm ø
specifically
designed
CROSSTIES
PREM: WOOD
or CONCRETE
MONOBLOC
NOMINAL:
spacing 500 mm
wood 600 mm
concrete
BOGIES
Improved
Suspension
standard 3piece
FASTENERS
ELASTIC
or dog
spike
BALLAST
350 mm
select
crushed
rock
+200 mm
sub-ballast
250 mm
shoulders
M/L
SWITCH &
CROSSING
WORK
GEOMETRY
INSPECTION
FLAW
INSPECTION
Premium
rail
tangential,
movable
point frog
~ 3-6 months
with profile
monitoring
~3 month
intervals
PROFILES & MAINTENANCE
WHEEL
RAIL
limit hollow
wear to 3 mm
Periodic grind to remove
corrugation and surface
damage
LUBRICATION
WHEEL/RAIL
Curves: Gauge face
ì<0.25-0.30
Rail head ì<0.350.40
Äì<0.10-0.15 L-R
Tangent: rail head
ì>0.35
WEAR
LIMITS
Measure
frequently to
ensure
economic
optimum
Table 4.14(d) TRAFFIC MIX: DEDICATED HH
AXLE LOAD: 35+
TERRAIN: ≥ 875 meter radius
TRAFFIC DENSITY: 30-49 MGT
TRAFFIC
DENSITY
30-49 MGT
TYPE
Premium
WEIGHT
136 RE
or
UIC 60
WHEELS
TYPE
PROFILE
Prem HT
curve plate
AAR Class C
1000mm ø
BALLAST
M/L
SWITCH &
CROSSING
WORK
GEOMETRY
INSPECTION
FLAW
INSPECTION
350 mm
select
crushed
rock
+200 mm
sub-ballast
250 mm
shoulders
Premium
rail
tangential,
spring point
premium
frog
~ 4-6 months
with profile
monitoring
~4 month
intervals
RAIL
specifically
designed
CROSSTIES
PREM: WOOD
or CONCRETE
MONOBLOC
NOMINAL:
spacing 500 mm
wood 600 mm
concrete
BOGIES
Improved
Suspension
standard 3piece
FASTENERS
ELASTIC
or dog
spike
PROFILES & MAINTENANCE
WHEEL
RAIL
limit hollow
wear to 3 mm
Periodic grind to remove
corrugation and surface
damage
LUBRICATION
WHEEL/RAIL
Curves: Gauge face
ì<0.25-0.30
Rail head ì<0.350.40
Äì<0.10-0.15 L-R
Tangent: rail head
ì>0.35
WEAR
LIMITS
Measure
frequently to
ensure
economic
optimum
Table 4.15(d) TRAFFIC MIX: DEDICATED HH
AXLE LOAD: 35+
TERRAIN: ≥ 875 meter radius
TRAFFIC DENSITY: 20-29 MGT
TRAFFIC
DENSITY
20-29 MGT
TYPE
Premium
WEIGHT
136 RE
or
UIC 60
WHEELS
TYPE
PROFILE
Prem HT
curve plate
AAR Class C
1000mm ø
BALLAST
M/L
SWITCH &
CROSSING
WORK
GEOMETRY
INSPECTION
350 mm
select
crushed
rock
+200 mm
sub-ballast
250 mm
shoulders
Premium
rail
spring point
premium
frog
~ 6 months
with profile
monitoring
RAIL
specifically
designed
CROSSTIES
PREM: WOOD
or CONCRETE
MONOBLOC
NOMINAL:
spacing 500 mm
wood 600 mm
concrete
BOGIES
Standard 3piece or
Improved
Suspension
standard 3piece
FASTENERS
ELASTIC
or dog
spike
PROFILES & MAINTENANCE
WHEEL
RAIL
limit hollow
wear to 3 mm
Periodic grind to remove
corrugation and surface
defects
LUBRICATION
WHEEL/RAIL
Curves: Gauge face
ì<0.25-0.30
Rail head ì<0.350.40
Äì<0.10-0.15 L-R
Tangent: rail head
ì>0.35
FLAW
INSPECTION
~6 month
intervals
WEAR
LIMITS
Measure
frequently to
ensure
economic
optimum
Table 4.16(d) TRAFFIC MIX: DEDICATED HH
AXLE LOAD: 30-34
TERRAIN: ≥ 875 meter radius
TRAFFIC DENSITY: 50+ MGT
TRAFFIC
DENSITY
>50 MGT
TYPE
Standard
in
tangent
Premium
in curves
WEIGHT
136 RE
or
UIC 60
WHEELS
TYPE
PROFILE
Prem HT
curve plate
AAR Class C
1000mm ø
BALLAST
M/L
SWITCH &
CROSSING
WORK
GEOMETRY
INSPECTION
FLAW
INSPECTION
300 mm
select
crushed
rock
+200 mm
sub-ballast
250 mm
shoulders
Premium
rail
tangential,
spring point
premium
frog
~ 4-6 months
with profile
monitoring
~4 month
intervals
RAIL
specifically
designed
CROSSTIES
WOOD or
CONCRETE
MONOBLOC
NOMINAL:
spacing 500 mm
wood 600 mm
concrete
BOGIES
Standard 3piece or
Improved
Suspension
standard 3piece
FASTENERS
ELASTIC
or dog
spike
PROFILES & MAINTENANCE
WHEEL
RAIL
limit hollow
wear to 3 mm
Periodic preventive grind
to restore conformal
contact
with manual monitoring
LUBRICATION
WHEEL/RAIL
Curves: Gauge face
ì<0.25-0.30
Rail head ì<0.350.40
Äì<0.10-0.15 L-R
Tangent: rail head
ì>0.35
WEAR
LIMITS
Measure
frequently to
ensure
economic
optimum
Table 4.17(d) TRAFFIC MIX: DEDICATED HH
AXLE LOAD: 30-34
TERRAIN: ≥ 875 meter radius
TRAFFIC DENSITY: 30-49 MGT
TRAFFIC
DENSITY
30-49 MGT
TYPE
Standard
in
tangent
Premium
in curves
WEIGHT
132 RE
or
UIC 60
WHEELS
TYPE
PROFILE
Prem HT
curve plate
AAR Class C
900mm ø
BALLAST
M/L
SWITCH &
CROSSING
WORK
GEOMETRY
INSPECTION
300 mm
select
crushed
rock
+200 mm
sub-ballast
250 mm
shoulders
Premium
rail
Fixed point
premium
frog
~ 6-8 months
with profile
monitoring
RAIL
AAR 1B or
equivalent
CROSSTIES
WOOD or
CONCRETE
MONOBLOC
NOMINAL:
spacing 500 mm
wood 600 mm
concrete
BOGIES
Standard 3piece or
Improved
Suspension
standard 3piece
FASTENERS
ELASTIC
or dog
spike
PROFILES & MAINTENANCE
WHEEL
RAIL
limit hollow
wear to 3 mm
Periodic preventive grind
to restore conformal
contact
with manual monitoring
LUBRICATION
WHEEL/RAIL
Curves: Gauge face
ì<0.25-0.30
Rail head ì<0.350.40
Äì<0.10-0.15 L-R
Tangent: rail head
ì>0.35
FLAW
INSPECTION
~4-6
month
intervals
WEAR
LIMITS
Measure
frequently to
ensure
economic
optimum
Table 4.18(d) TRAFFIC MIX: DEDICATED HH
AXLE LOAD: 30-34
TERRAIN: ≥ 875 meter radius
TRAFFIC DENSITY: 20-29 MGT
TRAFFIC
DENSITY
20-29 MGT
TYPE
Standard
WEIGHT
132 RE
or
UIC 60
WHEELS
TYPE
PROFILE
Prem HT
curve plate
AAR Class C
900mm ø
BALLAST
M/L
SWITCH &
CROSSING
WORK
GEOMETRY
INSPECTION
FLAW
INSPECTION
250 mm
select
crushed
rock
+100 mm
sub-ballast
250 mm
shoulders
Premium
rail
Fixed point
premium
frog
~ 6-8 months
with profile
monitoring
~6 month
intervals
RAIL
AAR 1B or
equivalent
CROSSTIES
WOOD or
CONCRETE
MONOBLOC
NOMINAL:
spacing 500 mm
wood 600 mm
concrete
BOGIES
Standard 3piece or
Improved
Suspension
standard 3piece
FASTENERS
ELASTIC
or dog
spike
PROFILES & MAINTENANCE
WHEEL
RAIL
limit hollow
wear to 3 mm
Periodic preventive grind
to restore conformal
contact
with manual monitoring
LUBRICATION
WHEEL/RAIL
WEAR
LIMITS
Lubrication as
required where
appropriate on curves
Measure
frequently to
ensure
economic
optimum
Table 4.19(d) TRAFFIC MIX: DEDICATED HH
AXLE LOAD: 25-29
TERRAIN: ≥ 875 meter radius
TRAFFIC DENSITY: 50+ MGT
TRAFFIC
DENSITY
>50 MGT
TYPE
Standard
in
tangent
Premium
in curves
WEIGHT
115 RE
or
UIC 54
WHEELS
TYPE
PROFILE
Prem HT
curve plate
AAR Class C
900mm ø
BALLAST
M/L
SWITCH &
CROSSING
WORK
GEOMETRY
INSPECTION
300 mm
select
crushed
rock
+100 mm
sub-ballast
250 mm
shoulders
Tangential
Fixed point
frog
~ 6-8 months
with profile
monitoring
RAIL
AAR 1B or
equivalent
CROSSTIES
WOOD or
CONCRETE
MONOBLOC
NOMINAL:
spacing 500 mm
wood 600 mm
concrete
BOGIES
Standard 3piece or
Improved
Suspension
standard 3piece
FASTENERS
ELASTIC
or dog
spike
PROFILES & MAINTENANCE
WHEEL
RAIL
limit hollow
wear to 4 mm
Periodic manual
monitoring
FLAW
INSPECTION
~4-6
month
intervals
LUBRICATION
WHEEL/RAIL
WEAR
LIMITS
Lubrication as
required where
appropriate on curves
Measure
frequently to
ensure
economic
optimum
Table 4.20(d) TRAFFIC MIX: DEDICATED HH
AXLE LOAD: 25-29
TERRAIN: ≥ 875 meter radius
TRAFFIC DENSITY: 30-49 MGT
TRAFFIC
DENSITY
30-49 MGT
RAIL
TYPE
Standard
WEIGHT
115 RE
or
UIC 54
WHEELS
TYPE
PROFILE
Prem HT
curve plate
AAR 1B or
AAR Class C
equivalent
900mm ø or
equivalent
CROSSTIES
WOOD or
CONCRETE
MONOBLOC
NOMINAL:
spacing 500 mm
wood 600 mm
concrete
BOGIES
Standard 3piece
FASTENERS
ELASTIC
or dog
spike
BALLAST
250 mm
select
crushed
rock
M/L
SWITCH &
CROSSING
WORK
GEOMETRY
INSPECTION
FLAW
INSPECTION
Fixed point
frog
~ 6-8 months
with profile
monitoring
~6 month
intervals
250 mm
shoulders
PROFILES & MAINTENANCE
WHEEL
RAIL
Establish good profile to
start and then periodic
limit hollow
maintenance with
wear to 4 mm
manual measurements
and inspection
LUBRICATION
WHEEL/RAIL
Lubrication as
required where
appropriate on curves
WEAR
LIMITS
Measure
frequently to
ensure
economic
optimum
Table 4.21(d) TRAFFIC MIX: DEDICATED HH
AXLE LOAD: 25-29
TERRAIN: ≥ 875 meter radius
TRAFFIC DENSITY: 20-29 MGT
TRAFFIC
DENSITY
20-29 MGT
RAIL
TYPE
Standard
WEIGHT
115 RE
or
UIC 54
WHEELS
TYPE
PROFILE
Prem HT
curve plate
AAR 1B or
AAR Class C
equivalent
900mm ø or
equivalent
CROSSTIES
WOOD or
CONCRETE
MONOBLOC
NOMINAL:
spacing 520 mm
wood 630 mm
concrete
BOGIES
Standard 3piece
FASTENERS
ELASTIC
or dog
spike
BALLAST
250 mm
select
crushed
rock
M/L
SWITCH &
CROSSING
WORK
GEOMETRY
INSPECTION
FLAW
INSPECTION
Fixed point
frog
~ 6-8 months
with profile
monitoring
~6 month
intervals
250 mm
shoulders
PROFILES & MAINTENANCE
WHEEL
RAIL
Establish good profile to
start and then periodic
limit hollow
maintenance with
wear to 4 mm
manual measurements
and inspection
LUBRICATION
WHEEL/RAIL
Lubrication as
required where
appropriate on curves
WEAR
LIMITS
Measure
frequently to
ensure
economic
optimum
Table 4.22(d) TRAFFIC MIX: DEDICATED HH
AXLE LOAD: 20-24
TERRAIN: ≥ 875 meter radius
TRAFFIC DENSITY: 50+ MGT
TRAFFIC
DENSITY
>50 MGT
RAIL
TYPE
Standard
WEIGHT
115 RE
or
UIC 54
WHEELS
TYPE
PROFILE
Prem HT
curve plate
AAR 1B or
AAR Class C
equivalent
830mm ø or
equivalent
CROSSTIES
WOOD at
±610mm
CONCRETE
MONOBLOC
at ±680 mm
BOGIES
Standard 3piece
FASTENERS
ELASTIC
or dog
spike
BALLAST
250 mm
select
crushed
rock
M/L
SWITCH &
CROSSING
WORK
GEOMETRY
INSPECTION
FLAW
INSPECTION
Fixed point
frog
~ 6-8 months
with profile
monitoring
~6 month
intervals
250 mm
shoulders
PROFILES & MAINTENANCE
WHEEL
RAIL
Establish good profile to
start and then periodic
limit hollow
maintenance with
wear to 4 mm
manual measurements
and inspection
LUBRICATION
WHEEL/RAIL
Lubrication as
required where
appropriate on curves
WEAR
LIMITS
Measure
frequently to
ensure
economic
optimum
Table 4.23(d) TRAFFIC MIX: DEDICATED HH
AXLE LOAD: 20-24
TERRAIN: ≥ 875 meter radius
TRAFFIC DENSITY: 30-49 MGT
TRAFFIC
DENSITY
30-49 MGT
RAIL
TYPE
Standard
WEIGHT
115 RE
or
UIC 54
WHEELS
TYPE
PROFILE
Prem HT
curve plate
AAR Class C
830mm
AAR 1B or
equivalent
CROSSTIES
WOOD at
±610mm
CONCRETE
MONOBLOC
at ±680 mm
BOGIES
Standard 3piece
FASTENERS
ELASTIC
or dog
spike
BALLAST
250 mm
select
crushed
rock
M/L
SWITCH &
CROSSING
WORK
GEOMETRY
INSPECTION
Fixed point
frog
~ 8-10
months
with profile
monitoring
250 mm
shoulders
PROFILES & MAINTENANCE
WHEEL
RAIL
Establish good profile to
start and then periodic
limit hollow
maintenance with
wear to 4 mm
manual measurements
and inspection
LUBRICATION
WHEEL/RAIL
Lubrication as
required where
appropriate on curves
FLAW
INSPECTION
~8 month
intervals
WEAR
LIMITS
Measure
frequently to
ensure
economic
optimum
Table 4.24(d) TRAFFIC MIX: DEDICATED HH
AXLE LOAD: 20 TO 24
TERRAIN: ≥ 875 meter radius
TRAFFIC DENSITY: 20-29 MGT
TRAFFIC
DENSITY
20-29 MGT
RAIL
TYPE
WEIGHT
Standard
115 RE
UIC 54
WHEELS
TYPE
Prem HT
curve plate
AAR Class
C 830mm
PROFILE
AAR 1 B
CROSSTIES
WOOD at
±610 mm
Monobloc
CONCRETE
at ±680 mm
BOGIES
Standard 3piece
FASTENERS
Dog spike or
ELASTIC
BALLAST
250 mm
depth with
250 mm
shoulders
M/L
SWITCH &
CROSSING
WORK
GEOMETRY
INSPECTION
fixed point
frog
once per
year
PROFILES & MAINTENANCE
WHEEL
limit hollow wear to 4
mm; establish good
profile to start and then
periodic maintenance;
manual measurements
and inspection
RAIL
Establish good
profile to start and
then periodic
maintenance with
manual
measurements and
inspection
LUBRICATION
WHEEL/RAIL
lubrication as required
where appropriate on
curves
FLAW
INSPECTION
once per
year
WEAR
LIMITS
Click Here To Go Back To Table of Contents
Part 5: Maintaining Optimal Wheel and Rail
Performance
Written by Mr. Michael D. Roney, member of the IHHA
Board of Directors and Professor Willem Ebersöhn, member
of the Technical Review Committee.
5.1
Maintaining Optimal Wheel and Rail
Performance
Rail is the single most expensive element of the track structure.
On many railways, it is behind only labor and fuel as an
expense item. The tonnage carried by a rail before it is
condemned can range from less than 100 million gross tons to
close to 2.5 gigga gross tons.
As an example of the value of rail maintenance
management, assume that a single kilometre of rail costs
$180,000 to install. Track engineers decide that the rail has a
badly fatigued surface and has reached the end of its service
life. They call for it to be replaced, gaining a salvage value of
$18,000.
But now assume that instead of replacing the rail, they did
some corrective rail grinding costing $1800 and left the rail in
track. The railway then invested the $180,000 – $18,000 $1,800 = $160,200 in the construction of a new customer
facility at a rate of return of 20%. This earned $160,000 * 20%
= $32,000 in its first year.
The next year, the track engineers see that their rail is
approaching allowable wear limits and schedules a rail
replacement, now costing $187,200 due to cost escalation of
4%. But they have made for the railway $32,000 – ($187,200 –
180,000) = $24,800 by deferring replacement of rail in that
kilometre, without consequence, for an extra year. And that is
why they collect a salary.
There is significant money to be made by deferring rail
replacement as much as possible without incurring risk.
Certainly it is a major responsibility of the track engineer to
ensure that he gets the most out of his rail, and rail profile
maintenance and rail testing are his most important tools to do
this.
o 5-1 x
The catch is that the above statements assume that the
aging rail does not have a direct impact upon other expenses,
such as wheel wear or fuel consumption. Furthermore, there
must not be a significant increase in risk of rail fracture.
To optimize rail and wheel life cycle, it is imperative that
these two components are managed jointly, as it is the
wheel/rail interaction that determines the performance of each
component.
In the case of wheel and rail, rail maintenance holds the
key, as rail is static in its location along the track layout and
more accessible for maintenance. Although the wheel
requirements of a vehicle is more variable in the sense that it
moves over a variety of rail layouts and conditions, its
requirements must match those of the rail to ensure optimal
wheel rail interaction.
The joint strategies of rail re-profiling, friction
management (lubrication), control of gauge, and rail condition
measuring can protect rail-caused cost impacts.
Rail re-Profiling: Rail and wheel life is reduced if the
wheel/rail contact conditions deteriorate through migration of
the contact band or flattening of the rail. Regular metal
removal can also control surface fatigue, which is ultimately
associated with internal rail defects. Both are controllable
through regular re-profiling.
Friction Management: Maintaining the rail surface friction
within specified limits on the rail gauge face, but also on the
top of rail, can reduce rail and wheel wear and improve fuel
consumption.
Gauge Control: As the gauge face wears, widening gauge
will change wheel/rail contact locations, which can accelerate
wear and fatigue. If the rail is allowed to spread (rollover)
under dynamic gauge widening, rail contact fatigue can result.
Lubrication and attention to fastener condition act together to
control these cost impacts.
o 5-2 x
Rail Condition Measuring: Rail fatigue under high axle loads
combined with the rail section reductions through wear and
grinding is an ideal environment for condition based
maintenance practice with the measuring devices available
today. The rate of occurrence of some internal rail defects
increases and can be detected with a rail defect-measuring plan.
The cost impact does not need to be high, however, as long as
such occurrences are detected, rail defects can be changed out
in a production fashion behind the rail testing operation.
Technology advances have made rail profile measuring
accurate and repeatable to the extent that rail wear rate can be
reliably determined and used for maintenance planning.
In the same manner, wheel profile and to a certain extent
defect detection technology is also at the point where
maintenance can move from a routine based maintenance
practice to a condition based maintenance practice.
Unfortunately, wheel risk management and maintenance
practices are not at the same level as rail risk management. As
an example, one heavy haul railway reports that for the year
2000 track related failures accounted for 23% of all accidents,
where wheel failures represented 5% of the accidents. But
wheel related accidents are characteristically more severe and is
reflected by the cost of these accidents as being 31% of the
total accident costs. (FRA reported Safety Statistics shows a
ratio of 36% track vs. 3% wheel related accidents with 11% of
the track-related derailments due to broken rails1).
Considering the high cost of wheel related derailments, it
seems that there is a discrepancy in our management of failure
risk of rails versus wheels. As typically indicated, a wheel that
lasts 5 years on a 30 mgt line would likely be tested and
profiled 2 times, whereas a rail in the same period would likely
be re-profiled and ultrasonically tested 16 times.
Nevertheless, over the past decade, railways have increased
their control of wheel/rail contact, although not always in an
integrated manner, between the wheel shops and the track
maintenance teams.
If rail sees frequent maintenance attention, as discussed
above, how long can it stay in track? Depending on operating
o 5-3 x
conditions, costs can certainly be minimized when the rail is
ultimately removed for loss of railhead, as opposed to any
other cause such as fatigue. This gives rise to seriously
rethinking the maximum allowable rail wear limits.
Whereas rail limits in the past have been associated with
the ultimate contact of a high wheel flange with the joint bar,
continuous welded rail has virtually eliminated this barrier.
Another concern in the past, particularly for railways with 30ton and greater axle loads and/or standard carbon rails, has
been that the rail would be misshapen by the time such limits
were being approached, as a result of plastic flow that the
whole track structure would be suffering from the poor vehicle
tracking.
In an environment where dynamic wheel/rail interaction
and profiles are controlled, railway engineers can start to work
to get the full life potential from the yield strength of the rail
section.
The achievement of full utilization of rail generally requires
a deliberate plan to put the following supporting strategies in
place:
1. Develop target rail profiles that are seen to achieve
low fatigue and wear. Because of the heavy influence
of contact fatigue on rail, these target profiles would
typically incorporate some conformity with wheel
profiles. A mechanical representative should be on
the team to advise on implications for wheels, and to
explore the potential for joint optimization of wheel
and rail profiles. This part is addressed in previous
chapters.
2. Measure rail and wheel conditions to determine
maintenance needs.
3. Develop rail and wheel wear projection methodology.
4. Develop rail and wheel fatigue life projection
methodology.
5. Perform economic evaluation of different premium
rail options based upon condition monitoring and
execute plans to progressively balance rail strength
o 5-4 x
with service environment. Determine the use of
premium and intermediate rail steels in different
tonnage and curvature classes based on minimum life
cycle costs. Perform analyses to determine
transposition and re-use policies.
6. Perform extensive rail re-profiling to the target profile
and to correct existing rail spalling, corrugations, and
head checks.
7. Re-profile wheels to target profile.
8. Install lubricators in locations with past history of
higher gauge face wear rates.
9. Develop new rail wear limits and supply new design of
joint bar to support extended wear limits in secondary
lines.
10. Develop new wheel wear limits.
11. Implement regular and frequent ultrasonic rail
inspection as well as regular rail wear and profile
condition measurement. Quality assessment is an
important part of this strategy. Correlate rail
deterioration with track geometry condition and gauge
widening to develop joint strategies.
12. Implement regular and frequent wheel flaw inspection
as well as regular wheel profile condition
measurement.
13. Implement frequent maintenance rail re-profiling
(grinding) on regular cycles. The objective should be
to move to single pass rail grinding, with speeds
adjusted to grind as fast as possible to control rail flow
and fatigue occurring between grinding cycles.
14. Adjust profiling standards and rail-testing intervals to
match needs of rails approaching extended wear limits.
15. Implement a condition based maintenance plan for
wheel re-profiling.
Not all of these steps have to be in place to achieve major
savings in costs. It is suggested, however, that quantum
improvements in rail and wheel life require some attention to
each of the steps. Critical to the implementation of this
strategy is that all people involved clearly understand both the
o 5-5 x
end objective and the role of each of the elements. This will
undoubtedly require some education of track managers
through senior engineering managers of such aspects as role of
rail and wheel profile designs, fatigue mechanisms, and wear
rates.
5.2
5.2.1
Rail Structural Deterioration
Management of Rail Testing to Control Risk of
Rail Fracture
The occurrence of internal defects in rails is an inevitable
consequence of the accumulation of fatigue under repeated
loading. To maximize rail life, heavy haul railways live with
controlled rates of defect occurrences, relying on regular
ultrasonic or induction rail testing and strategic renewal of rail
that is obviously showing evidence of fatigue.
The consequences of internal flaws can be serious. An
unrecognized defect can result in rail breakage with
interruption of service and the potential risk of catastrophic
consequences. At the least, an isolated case means a repair
cost and introduces two unwanted welds, while a series of
defects can condemn a whole rail length.
FRA Reported Safety Statistics for 1999 1 shows 11% of the
accidents were rail and joint bar related. Figure 5.1 shows the
distribution of accidents for rail and joint bar defect types for
1999.
Train Accidents Reported for Rail and Joint Bar Defect Type (%)
Transverse/compound Fissure
21%
17%
Broken Base of Rail
15%
Verrtical Split Head
Head and Web Seperation (outside joint)
13%
12%
Detaied Fracture from Shelling
4%
Other Rail and Joint Defects
4%
Broken Field Weld
3%
Horizontal Split Head
3%
Joint Bar Broken
3%
Worn Rail
Head and Web Seperation (inside joint)
Mismatched Rail-head Contour
2%
1%
Bolt Hole Crack or Break
1%
Joint Bolts Broken or Missing
1%
Figure 5.1: Distribution of Accidents Per Rail and
Joint Bar Type
o 5-6 x
The challenge is to avoid the occurrence of service failures
due to undetected defects. Service failures are more expensive
to repair and can lead to costly line disruptions or even
derailments. The role of rail defect testing is therefore to
protect service reliability while avoiding overly conservative rail
renewals. To illustrate the scope of rail testing’s contribution,
consider that North American heavy haul railways detect an
average of 0.4 defects per track km. (0.6 rail defects/mi) each
year while inspecting at intervals of 18 mgt (20 mgt) and
experience 0.06 service failures/km (0.1 service failures/mi.).
One service defect in two hundred leads to a broken rail
derailment. Rails are typically replaced when total defects are
occurring at a sustained rate of 1-2/rail km (2-3/mi.).2
In controlling risk, the most basic control variable is the
test interval. A railway administration must decide upon the
frequency of passage of the rail testing equipment that will
balance the cost of testing and rail change-out with the
expected derailment cost to minimize the net cost of the risk.
In this delicate equation, the reliability and operating speed of
the testing system play an important role.
5.2.2 The Framework for Risk Management
The practice of rail testing has a simple objective of reducing
the annual costs incurred as a result of broken rails. But there
are many variables involved. Figure 5.2 shows the most
important of these.
o 5-7 x
Figure 5.2: Factors Controlling the Risk of
Broken Rail Derailments
The direct cost of undetected rail breaks is the difference
between the cost of replacing broken rails on an emergency
basis, and the cost of the orderly replacement of detected
defects. The cost of derailments caused by undetected broken
rails is an indirect cost of poor inspection reliability. The
derailment cost is the annualized cost of the rare but high cost
occurrence of a derailment. As the probability would be
derived statistically from past records, the annualized
derailment cost is also called the “expected derailment cost.”
This cost is also related to specific characteristics of the railway
o 5-8 x
such as the remoteness of the routing, the severity of the
terrain, the type of lading, and the size and speed of the trains.
The number of service failures is intimately related to the
effectiveness of the inspection. In a risk management
approach, high inspection reliability is required where long
trains are travelling fast along a watercourse in proximity to
population centers. While heavy haul lines typically have long
trains and high derailment costs, train frequencies may be less
than on mixed freight lines, defect growth rates may be more
uniform, and tonnage is easier to track for the purpose of
planning test intervals. Many heavy haul operators have
dedicated rail-testing vehicles, which they may use at frequent
intervals, even monthly.
To ensure an effective rail testing program, the test
equipment must be properly designed and calibrated to reliably
indicate defects, the equipment logic must present to the
operator only those indications that could be a rail defect, and
the operator must be experienced and diligent.
In addition, test frequencies must be matched to the
growth rate of critical defects so that at least one test, and
preferably more inspections, are made in the interval between
the development of a rail defect to a minimum detectable size
and its growth, to a size that represents a significant rate of
rapid fracture.
In practice, the growth rate of rail defects is both highly
variable and rarely known with any certainty. Rail testing on
heavy haul operations often presents some specific problems,
for as traffic loading is high, defect growth is accelerated and
the time scale for intervention is compressed. The tendency
for each wheel passage to stress the rail in a similar pattern can
increase defect growth rates. At the same time, heavy axle
loads can lead to a fatigued rail surface that may present
confusing indications from testing equipment. The use of rail
of various metallurgical qualities further complicates the task.
Most heavy haul operators attempt to control risk by
monitoring of the reliability of the test through evaluation of
failures occurring soon after testing and by comparing ratios of
service to detected rail defects.
o 5-9 x
5.2.3 Defect Occurrence Rates
The driving factor determining the risk of rail fracture is the
rate at which a population of internal flaws develops in the rail.
Internal flaws in rail have a period to initiation and a period
during which a crack will propagate. The risk is introduced
when cracks remain undetected during their growth to critical
sizes. This occurs when the period between the times the
crack reaches detectable size is significantly shorter than the
testing interval. Figure 5.3 presents the example of a rail flaw
with a long period of exposure before failure and one with a
short exposure time. An example of a long exposure time
might be a transverse fissure, which is detectable at a small size
due to its central location in the head, and which may grow
slowly. At the other extreme might be a defective weld with
poor fusion in the web area. The web cracks would typically
be large at first detection and could be expected to propagate
rapidly.
In revenue service, rail in a given routing would be
expected to have a broad population of defects of different
sizes, each growing at a different rate.
In a typical heavy haul line, the population of flaws of
different sizes can be assumed to be distributed according to
an exponential distribution (Figure 5.4), where there are many
very small flaws, but very few large defects. As fatigue cycles
accumulate in the rail from high contact stresses, more flaws
are initiated and those already present continue to grow.
Critical to the risk of a rail break is the number of defects in
the right tail of the distribution. The area of the right tail of
the distribution would represent the number of rail flaws that
are of sufficient size that they could fracture suddenly. It is a
fact that the distribution of internal defects by size varies from
location to location. A highly stressed track segment or one
laid with a dirtier steel, will have a population distribution
shifted to the right and should, in theory, be tested more
frequently to achieve the same risk level.
o 5-10 x
Figure 5.3: Size Distribution of Flaws in
Rails in a Typical Line
Figure 5.4: Transverse Defect Growth Rates
Measured under 39-Ton Axle Loads at
FAST/Heavy Tonnage Loop
Because rail flaw detection is quite effective at detecting
large defect sizes, the distribution of defects in track at any one
time is skewed to the smaller sizes. The critical objective of rail
testing programs is to both eliminate the right tail high risk
defects in the distribution at the time of testing while
o 5-11 x
attempting to detect all defects from the distribution which,
through growth, will have reached the high risk level by the
time of the next rail test.
Hence both testing reliability and test intervals are
important. But most importantly, test reliability and testing
intervals must be matched.
Presently, there is little hard evidence on either the growth
rates of different defects or their critical sizes. One notable
exception is the defect growth relationship determined in
studies by the Transportation Technology Center, Inc. (TTCI)
a subsidary of the Association of American Railroads (AAR) at
the Facility for Accelerated Service Testing Facility (FAST),
Pueblo, Colorado USA. By monitoring the growth of
transverse defects under the controlled conditions of a unit
train cycling over the test loop, TTCI measured a wide range
of different growth rates.
Figure 5.5 plots the progression of transverse defects that
developed under a consist with 35 ton (39 Ton) axle loads.3,4
The tonnage required to initiate a defect was found to be very
difficult to predict, but once initiated, transverse defects were
found to grow non-linearly with tonnage, as would be
predicted from fracture mechanics theory. Under the uniform
heavy loading conditions of the FAST consist, some defect
growth rates were found to be quite rapid. Rapid growth rates
could also be expected where tensile residual stresses are
present in the railhead, and in low temperatures in continuous
welded rail where the rail is again in tension.
Part 3, Appendix A presents Canadian Pacific rail system
criteria for the protection of defective rail in track.
o 5-12 x
Figure 5.5: Transverse Defect Growth Rates Measured Under
39-Ton Axle Loads in Fast Heavy Tonnage Loop
5.2.4 Critical Defect Sizes
Experience has shown that rail can fracture suddenly from
transverse defects as small as 10% of the railhead. Generally,
risk is significant when a transverse defect is larger than 35%
of the head. A bolt hole crack is known to start to grow
rapidly when the length exceeds about 13 mm (½in.), and rapid
fracture can usually be anticipated from a 25 mm (1 in.) crack.
In general, railways have relied upon experience to distinguish
between fractures which present a substantial risk and those
which may safely remain in track for a specified period of time.
For example, Part 3, Appendix A: Canadian Pacific Rail
Defects5 presents the mandatory Protection Codes imposed by
Canadian Pacific Rail System on trains passing over detected
rail defects prior to their removal from track. This table
presents one heavy haul railway’s assessment of the risk
associated with different sizes and types of defects.
A study of dynamic fracture of rails conducted at Queen’s
University at Kingston, Canada,6 has shed some more light on
the dynamic load capacity of rails, and hence the risk of
fracture under heavy axle loading. The study involved
dropping dynamic impact loads typical of those imposed by
shelled wheel treads, out-of-round wheels or wheel flats on
rails, which had been removed from track because of detected
defects of different types. The rail specimens were pulled
o 5-13 x
longitudinally to simulate tensile stresses from low
temperatures, and some specimens were tested at down to –
20°C.
It was found that:
1. Impact loading was far more likely to fracture defects
in the transverse plane;
2. The tensile stresses imposed by temperatures
substantially less than the neutral temperature were
important in causing rail fracture;
3. Where the rail is shelling excessively, sudden rail
fracture will occur at lower, more frequent impact
level;
4. The residual, thermal and dynamic stresses imposed by
traffic contributed equally to total stress intensity;
5. The size of the flaw is a more important risk factor
than the percent of the railhead that has fractured. In
fact, a larger railhead may fracture more easily under
dynamic loading. Because a greater rail mass must be
rapidly “moved aside” under a high frequency impact,
i.e. has greater inertia, a larger railhead is less
compliant and may absorb more energy in impact.
Through observation of the conditions under which a rail
with a known defect could fracture suddenly, an equation was
developed from this work, which calculates the peak dynamic
load at fracture, Pdyn, stated in kilopounds (kips):
p dyn =
4.83 K
b
I C − 1.38 ∆ T − 6.46 σ
(1)
Where:
K
is the fracture toughness of the rail steel. This
value IC is typical 38.5 MPa for standard rail steels
and 20% higher for premium rail steels;
∆T
is the variation of the rail temperature from its
neutral or stress-free temperature in degrees
Fahrenheit.
o 5-14 x
σr
is the residual stress determined from the opening
that develops in a saw cut test. A value of 15.7
kPa/mm (14.3 ksi) is a good estimate from the
Queen’s University tests.
For example, using the above empirical equation, the
following combinations of conditions could cause sudden rail
fracture:
For a transverse defect covering 8% of the railhead:
• a rail temperature of 56° Celsius (100° Fahrenheit)
below the neutral temperature and a dynamic wheel
load of 356 kN (80,000 lbs).
For a transverse defect covering 10% of the railhead:
• a rail temperature of 56° Celsius (100° Fahrenheit)
below the neutral temperature and a dynamic wheel
load of 311 kN (70,000 lbs).
For a transverse defect covering 18% of the railhead:
• a rail temperature of 39° Celsius (70° Fahrenheit)
below the neutral temperature and a dynamic wheel
load of 311 kN (70,000 lbs).
For a transverse defect covering 40% of the railhead:
• a rail temperature of 56° Celsius (100° Fahrenheit)
below the neutral temperature without any wheel
loading.
See Part 3, Appendix A: Rail Defects (Canadian Pacific
Rail & Spoornet)
5.2.5 Rail Fatigue Projection
Most railways performing regular projection of rail life use the
Weibull methodology for projecting rail fatigue occurrence
rates. The Weibull methodology is useful in identifying
locations where trends are sustained vs. the case where defects
have remained constant. The situation where rail defect
occurrence rates are increasing is more critical, as this may
signal a mature fatigue process. These projections are used to
identify consistent trends in rail defect occurrences that could
be cause for a rail renewal program.
o 5-15 x
Attention to trends identified through regular use of
Weibull projections may guide selection of a strategy to correct
a defect trend by tamping up rail joints, building up rail ends by
welding, relieving the gauge corner, or attending to flat wheels.
Rail should be changed out when the annual cost of
repairing rail defects exceeds the value of deferring the renewal
for another year. At a repair cost of only $2500 per defect, and
an annual value of $18000 in interest savings if you leave the
rail in track, it requires a strong trend line to justify a rail
replacement for defect occurrences alone. But if a significant
number of these rails are failing in service, this introduces the
possibility that leaving the rail in track may incur the high cost
of line outages during emergency rail replacements and broken
rail derailments.
The key therefore is in maintaining effective rail testing.
As shown, service reliability requires both effective testing
systems and frequent rail testing. Attainment of long rail
service lives in a heavy haul environment similarly requires a
strategy to support rail economics with effective rail testing.
5.2.5.1
Use of Weibull Distribution to Predict Rail
Flaw Occurrence Rates
β T − γ 
f (T ) =


η  η 
( β − 1)
e
T − γ 
−

 η 
β
(2)
The Weibull probability density function is given by:
ƒ(T) > 0, T > γ, β > 0, η > 0, -∝ < γ < ∝
Where:
β
= Shape parameter
γ
= Location parameter
η
= Scale parameter
T
= Time, Tonnage etc.
o 5-16 x
The Weibull reliability function is given by
−T
−γ

 η
R(T ) = e



β
(3)
and the Weibull failure function
F (T ) = 1 − R(T )
−
=1− e
T −γ

 η



β
(4)
The failure function is manipulated into the following form:
−T
1 − F (T ) = e
1
=e
1 − F (T )
ln
−γ

 η



T −γ

 η



β
β
T − γ
1
=
1 − F (T )  η



β


1
 = β ln (T − γ ) − β ln (η )
ln  ln
F
T
−
1
(
)


(5)
This linear relationship is used for constructing Weibull
probability paper. β ln (η ) is constant for a given situation.
The Weibull failure rate, λ (T), is given by
λ (T ) =
β
η
T − γ

 η



(β − 1)
(6)
In rail failure analyses one of two avenues for the calculation of
reliability or failure rates can be followed:
1. A maximum number of failures, defects or
occurrences, per distance of track, of a certain nature
can be decided upon beforehand. Once this level of
failures has been reached it is assumed that 100% of
occurrences had been experienced and some action
like replacing of the rail is taken.
o 5-17 x
2. No previous decision regarding the number of defects
that is allowable in the track has been taken. Here use
is made of the so called Median Rank to allocate a
value of F(T) to failures. The Median Rank will, in this
case, again be based on a unit length of track.
In order to obtain relevant results from a Weibull analysis
of rails the track must be divided up in homogeneous units.
Information required for analysis includes:
1. The type of defect (Classification of failure);
2. Tonnage to failure;
3. Time to failure;
4. History of repairs and maintenance;
5. Infrastructure data. Position in track etc.
Lengths of rail in a unit may vary upon conditions. In
general lengths from 5 km to 50 km may be used. The
considerations for lengths of rail to be identified, tested and
analysed will be discussed later.
The following example illustrates the typical use of the
Weibull function.
Failure data for heavy haul-line 20 to 40 km:
Failure type
Line length (km)
: Kidney shaped crack
: 20
Max. defects per km : 5 (Has to be decided on as policy)
Table 5.1 shows data from a spreadsheet program used for
the calculation of the Weibull parameters.
o 5-18 x
Tonnage
(mgt)
100
200
300
400
500
600
700
800
900
Table 5.1: Results from Weibull analysis
Failures Ave. Failures % Failed
Years
per km
per period
2
4
6
8
10
12
14
16
18
5
8
4
8
4
6
5
9
8
0.25
0.65
0.85
1.25
1.45
1.75
2.00
2.45
2.85
5
13
17
25
29
35
40
49
57
The columns in Table 5.2 are:
F(mgt)
% of failures of the Kidney shaped crack type.
Based on the max. defects allowable per km.
Tonnage (mgt)
Actual gross load carried by rails.


1

Y = ln  ln
 1 − F ( MGT ) 
x = ln (Tonnage)
Weibull parameter calculation
:0
Location parameter, γ
F(mgt)
5
13
17
25
29
35
40
49
57
Table 5.2: Calculation of Weibull parameters
Tonnage (mgt)
Y
X
100
-2.9702
4.6051
200
-1.9714
5.2983
300
-1.6802
5.7037
400
-1.2459
5.9914
500
-1.0715
6.2146
600
-0.8421
6.3969
700
-0.6717
6.5510
800
-0.3955
6.6846
900
-0.1696
6.8023
Regression Output: (Linear regression done on Y and X
columns)
o 5-19 x
Constant
-8.50235
Std Err of Y Est
R
0.088236
0.991049
No. of Observations 9
Degrees of Freedom 7
X Coefficient(s)
1.207463
Std Err of Coef.
0.043373
Shape parameter, β = X coefficient = 1.207463
β ln η
= 8.502351 (Constant)
ln η
= 7.041499
η
= 1143.099
From Table 5.2 the values of the Weibull parameters were
obtained:
η
= 1143.1
β
= 1.207
γ
=0
Using the Weibull parameters obtained above the
following typical calculations are now possible:
Reliability at certain life
T − γ
R (T ) = e − 
 η



β
 2 000 − 0
R (2 000) = e − 
 1143.1
R (2 000) = 0.14 = 14%
1.207



In terms of our model of five allowable kidney shaped
defects per km the rail will after carrying 2000 mgt have a 14%
reliability; i.e., only 0.14 x 5 = 0.7 defects per km will be
allowable or 4.3 defects per km will already exist.
o 5-20 x
Failure rate at a certain life
λ (T ) =
β
η
T − γ

 η



( β − 1)
(1.207 − 1)
1.207  1500 − 0 


1143.1  1143.1 
= 0.00111 failures / defects per MGT per km
λ (1500) =
When a certain defect level will be reached
Should it be decided to start ordering rails when defects
have reached a level of 4.5 defects per km:
F (T ) =
4.5
= 90%
5.0
T − γ
= 1 − e 
 η
−



β
1.207
 2280 − 0 
= 1 − e− 

 1143.1 
(By means of replacement of T in computer model).
= 0.90 = 90%
This means that the defect level of 4.5 defects/km will be
reached by the time 2280 mgt has passed over the rail.
Further refinements to this model by for instance adding
confidence limits is possible.
5.2.6 Modes of Rail Testing
An effective rail test is one in which all defects which could
present a hazard to the safe passage of trains, at the time of
testing and projected forward to the time of the next test, are
located and sized to an accuracy that permits a valid decision to
be made on its removal. A skilled test operator can perform a
reliable test at a spot location with hand probes within a few
minutes. The problem is in knowing where to look for the
defect. The solution is to use a machine to locate potential
defects at speed.
o 5-21 x
Thus, there are two distinct components to rail testing:
1. Location of defects by machine.
2. Description of defects by the operator.
The basic elements that are necessary to ensure that a
defect is correctly located and identified are:
1. The equipment used for detection must represent the
size of the defect sufficiently accurately to produce a
recognizable indication(s) on the operator’s display
under all rail surface conditions that could be
expected.
2. The indication presented to the operator must be
clearly identifiable as a rail defect.
3. The operator must have sufficient training, experience,
and vigilance to respond to a defect indication with
very high reliability and to correctly perform hand
testing to identify the defect.
The two aspects of detection, involving machine and
operator, explains why rail testing is carried out in a variety of
ways:
“Non-stop” hand testing, where an operator pushed
trolley-mounted equipment along at a walking pace and carries
out a pedestrian sweep, stopping to explore and confirm
indications. This offers the advantage that the information is
presented to the operator at a slow speed and he may perform
a simultaneous visual check of rail surface conditions. The
major disadvantage is the high cost per test kilometre due to
slow test speed and the need to test each rail separately.
While some railways still do out-of-face rail testing with
ultrasonic hand trolleys, their use is waning due to high labor
costs, the need for a chase vehicle anyway to regularly recharge
water supplies and safety considerations. Therefore, further
comments will focus on flaw detection by vehicle.
“Stop and confirm” machine testing, where the test
vehicle stops at each indication and the hand operator gets out
to verify and mark the defect.
o 5-22 x
Advantages are:
• Less sophisticated equipment means lower capital
cost.
•
Rail is marked for renewal at detection.
•
Detection can run in sync with rail repair.
•
Test vehicle may be hi-rail equipped, increasing its
flexibility to move between sites over roads and to set
on and off track at road crossings.
Disadvantages are:
•
Slower operation than non-stop testing.
•
Test equipment may have to run under rules for track
equipment to permit backup moves.
•
Crew of hi-rail equipped ultrasonic cars must travel to
other lodging.
“Non-stop” machine testing, where a multi-probe machine
travels the section at typically 35 – 40 km/h. Real time
computing attempts to recognize signal indications, which
could possibly be defects, and to paint the location for follow
up hand testing.
Advantages are:
•
Less interference with traffic.
•
Lower unit cost of testing.
•
Higher productivity.
•
In some territories, signal systems will not allow railbound equipment to back up, leaving a long walk if
immediate verification is required.
Disadvantages are:
•
Higher capital cost of equipment.
•
Ultrasonic car may leave behind more defects than can
be fixed in a day, leading to slow orders.
o 5-23 x
•
Longer time interval between detection and
verification.
•
Rail surface fatigue can cause excessive indications.
•
Real time detection by computer must necessarily be
conservative leading to a tendency to paint too much
rail.
•
Hand test results are not available to recognize for
recalibration..
“Tandem machine testing” is a variation on non-stop
testing. In this approach, a “chase vehicle” working in the
same track possession follows the principal testing vehicle.
The led test car works non-stop as fast as the rail condition will
allow, while the satellite car regulates its average speed, stops
included, to match the site advance. This method is capitalintensive, but addresses many of the disadvantages of non-stop
testing.
5.2.6.1 Rail Testing Equipment
Various types of non-destructive methods have been employed
for testing rail in track, the main ones being:
•
Induction, where a low voltage eddy current is passed
along the rail between two moving probes, inducing a
strong magnetic field around the rail. An internal rail
defect causes distortions in the field around the rail,
which are picked up by search coils (see 5.2.10).
Induction methods can detect railhead defects and
certain web defects outside of the joint bar area.
•
Residual magnetic, where the rail is magnetized, and
search coils generate a weak current at irregularities in
the rail.
•
Ultrasonic, where ultrasonic waves are beamed into
the rail and the echoes are studied for irregularities.
As ultrasound is the technology most frequently used by
heavy haul railways, future comments will deal with ultrasonic
rail flaw detection only.
o 5-24 x
5.2.7 Ultrasonic Principles
While ultrasonic is a very specialized field usually left to the
experts, it is helpful for the railway user to have a grasp of the
principles involved. The test tool is a beam of electro-acoustic
energy with a frequency in the region beyond the hearing
range. The beam – which can be linked to that of an electric
torch or flashlight – is some 20 mm (1 in.) in diameter at its
origin and diverges from cylindrical at 3–5 degrees. The beam
is pulsed – switched on then off – at a set distance along the
rail, usually 2, 4, or 5 mm.
Ultrasonic transmitter crystals may be fitted in sliding
shoes or in rotating wheels. The sliding shoes are in closer
contact with the rail and afford a good angular stability, as the
mounting is flexible and adaptable. The wheel probes deal
better with irregular rails and offer a broad base for scanning
the railhead. A combination of both systems exist on some
machines.
Like light, the transmitted ultrasonic energy is refracted
upon changing from medium to medium, and reflected upon
meeting a suitable surface that is roughly perpendicular to its
propagation. In the flaw detection operation, two phenomena
are of interest. A beam hitting a discontinuity can reflect back
and disclose an “appearance” that indicates a potential defect.
And a beam masked from an expected end-echo cannot reflect
back, and thus gives a “disappearance” that indicates a
potential defect.
In practice, one is looking for various types of defects,
each with its own characteristics. The major characteristic
distinguishing different defect types will be the defect’s plane
of propagation. Thus, a transverse defect will be situated on a
plane across the railhead and sloping at some 70 to 90 degrees
to the vertical, while horizontal split heads and vertical split
heads describe themselves. In order to achieve reflection, the
search beam must meet the defect plane at about right angles.
From here, it becomes clear why rail-testing cars are fitted with
several probes on each rail.
The beams are centered on the longitudinal axis of the rail
by mechanical means with reference to the gauge face. As the
o 5-25 x
beams descend the web, and cannot radiate out from that path,
there are zones in the rail foot hat are not tested by the
ultrasonic method.
The passage of the ultrasonic energy is not as clean as one
would like. In particular, there is a disturbed zone of about 10
mm at the interface between the probe and the rail. The
passage from one medium to the other is assured by a film of
water that acts as a couplant. Nevertheless, the first several
millimetres of the entry into the rail cannot be exploited.
Other parasite effects occur also. Thus the first stage in the
recovery of the test information is to filter the returning energy
to remove the misleading effects. The second stage is to set
adjustable gates to select the areas in the rail that are of
interest.
At this point, the energy reflected can be visualized on a
cathode ray oscilloscope, an illustration referred to a an A-scan.
Defects will be recognizable from a set shape of the
oscilloscope trace. They are said to have a “signature.”
In principle, this information suffices to locate potential
defects. In practice, the traces are lively and require great
attention for interpretation, and there are simultaneously
several channels of information for each rail. The problem is
now one of information technology to assure the recognition
of potential defects among the mass of tested data that will
flow through the system. For example, some 10 information
points are generated every few millimetres along the rails, while
travelling at 15 – 25 km/hr. Another visual aid may be a
pictogram. Known as a B-scan, this is a picture of a rail
section, with diode lights or computer graphics given a "“quick
glance” view of suspicious echoes. An audio tone signal can
give a supplementary indication.
All of the above indications are usually presented to the
operator in real time, but are usually stored on a multi-channel
line recorder for later consultation in cases of controversial
findings.
o 5-26 x
5.2.8 Inspection Effectiveness
Rail flaw detection reliability is a system in which total
reliability is the joint result of the performance of test probes,
data processing clarity of information displays to the operator,
the operator himself, and the management of rail testing
intervals. Where there is a weak link in the system, it must be
compensated by the other elements of the system.
For example, weak performance from the ultrasonic
probes can be compensated somewhat by more sophisticated
processing. Or overly complicated or confusing information
displays must be compensated by an experienced operator.
Most importantly, poor overall performance from rail testing
can be compensated somewhat by very frequent testing
intervals.
The following discusses some of the factors affecting railtesting reliability.
5.2.8.1 Test Probes
To locate an internal defect in rail, ultrasonic or induction
energy must be transmitted along a pathway and at an energy
level sufficient to produce a clearly anomalous reflection from
the fracture surface. This reflection must be distinguishable
from the base of the rail itself and from railhead surfaces, and
the signal received from the fracture surface must be
substantially greater than the overall noise level introduced by
the grain structure of the rail or from the probe itself. And the
signal received must be sampled sufficiently frequently to have
captured sufficient pulses or “echoes” to have “seen” the
defect.
Fortunately, the larger the defect, the greater the signal
reflected. The difficulty arises in the detection of threshold
defect sizes. The pulse count from a small transverse defect,
for example, can be very similar to the “noise floor,”
particularly in older, less clean rail steels.
Furthermore, the full size, or length of the defect, is almost
never seen. This is because the full size of the defect is only
seen if it happened to be oriented at exactly 90 degrees to one
of the search beams. As there are an infinite number of
o 5-27 x
possible defect orientations and a finite number of probe
orientations, this is rare. What is usually seen as the reflecting
surface is therefore the projection of the fracture surface onto
the plane of the probe.
True sizing of the defect only occurs when the defect is
linear and oriented at 90 degrees to one of the search beams,
As there are an infinite number of defect orientations and a
finite number of probe orientations, this is rare. In the
example shown in Figure 5.6, a 32 mm (1-1/4 in.) curvilinear
bolt hole crack emanating at 45 degrees from the bolt hole is
seen as a 25 mm (1 in.) defect by the ultrasonic probe
arrangement. A particular problem is experience in those rare
circumstances where the crack emanates vertically from the
bottom of the bolt hole. This type of crack is susceptible to
pull-apart, but is transparent to the 0 degree probe and underrepresented by 43% by the 35° probe.
Figure 5.6: Length of a Bolt Hole Crack 32 mm Long and
Growing at 45° as seen by 35° and 0° Ultrasonic Probes
Another example of undersizing by ultrasonic occurs when
defects are located in the extreme corners of the railhead, the
initiation point for detail fractures from shell (Figure 5.7).7 8
Here the problem is due to the diverging angles of the
ultrasonic beam. A centrally-located ultrasonic beam, 20 mm
thick and diverging at 3–5 degrees, is unable to illuminate the
parts of the crack surface near the gauge corner and in the
o 5-28 x
head-web fillet area. Flaws larger than 65% of the head area
are therefore characteristically undersized. This can be
counteracted to some degree by the addition of 70 degree
probes on field and gauge side, but the lateral separation of
such probes is constrained by the possibility of “taking air”
when encountering a severely worn rail gauge corner.
Figure 5.7: Illustration of Characteristic under Sizing of a Large
Transverse Defect by a Single 70° Ultrasonic Probe
5.2.8.2 Signal Processing
The electronic signals received by the test probes must be
adjusted to discern a recognizable signal, while eliminating
spurious “noise.” Signal processing may affect overall testing
reliability if filters are not adjusted to recognize returning signal
thresholds that could constitute valid defects. For example,
processing logic that does not lower the threshold level for
defects found deeper in the railhead whose reflected energy
will be less than a shallow defect, would potentially miss
smaller defects deeper in the rail.
o 5-29 x
5.2.8.3 Displaying Indications to the Operator
Most testing systems use a “gating logic” whereby an ultrasonic
echo is divided into reflections from head, web, base, and
possible defect. An indication of a potential defect is
presented to the operator only if sufficient signal pulses or
echoes have been received within a time interval that would
constitute the “defect zone.” System test reliability depends
upon the successful definition of the time interval through
which echoes from the probes should be counted as an
indication of both the presence and size of a defect.
Recent developments have also sought to assist the
operator by recognizing patterns typical of rail ends, bolt holes,
bond pin drillings, etc. The objective is to present to the
operator only those patterns, which cannot be, explained by
typical track features. This prevents the operator from being
flooded with information that he must mentally process.
5.2.8.4 Operator Vigilance
Current testing systems continue to be operator sensitive. The
ideal operator can maintain mental vigilance over extended
periods of time, using his training and experience to identify
suspicious pen indications or patterns, in spite of various
distractions within the test car. Such operators exist, but there
are an equal number of excessively conservative operators who
frequently stop and hand test and may mark a rail for
unnecessary removal where they do not recognize the pattern
of indications. On the other hand, some operators are
production-oriented, or are perhaps too quick to attribute
unusual indications to a rail surface condition. Operator
performance should be reviewed regularly be selecting random
samples of recorded signal indications, ideally in territories
where the number of detected defects has changed dramatically
between tests. These recordings should be reviewed with the
operator to locate areas where he may have missed a potential
rail defect.
o 5-30 x
5.2.8.5 Estimates of Rail Testing Reliability
The effect of defect size on the probability of finding a defect
is well illustrated in Figure 5.8.8 It was compiled by the
Transportation Systems Center and the AAR, and is an
assessment of typical testing capabilities of current contractors’
equipment. The AAR model estimates that a defect covering
60% of the railhead has a 90% chance of detection in a given
test. At the same time, a flaw covering 10% of the head is
likely to have a probability of detection of only 45%.
Figure 5.8: Estimated reliability of conventional
rail test equipment
The American Railway Engineering Association is more
demanding in their recommended minimum performance
guidelines for rail testing. These guidelines, summarized as
Table 5.3,9 define the minimum acceptable percentage of
defects that must be detected by a single ultrasonic test. The
detection rate recommended by the AREA as indicative of a
fair to good ultrasonic inspection varies with the type of defect,
its size range and the class of track. For example, the AREA
recommends that rail testing services be considering to be
operating below acceptable performance if more than 65% of
o 5-31 x
transverse defects in the 5-10% range go undetected. On the
other hand, 98% of transverse defects covering 60% of the
railhead must be correctly identified and marked for removal.
The specification also defines the minimum sizes of defects
that are considered both worthy of reporting and within the
size range for reliable detection.
Such a specification invites questions to as how the
required testing performance can be verified. The best method
is to have a section of test track containing defects of know
sizes. This tests the capabilities of the test equipment, but is
not a realistic test of the vigilance of the operator in normal
service. Some railways run tandem tests where two test cars
will alternate running in the trailing position.
Defects found by one test car and operator and not the
other, after verification by breaking open or lab inspection,
would be considered missed defects for the purpose of
verifying performance to the specification.
o 5-32 x
Table 5.3: Minimum Performance for Rail Testing
Size
(Length or % of head Area)
Defect Type
Transverse Defects in the Rail
Head eg. transverse fissure
compound fissure engine
burn/welded burn fracture
Detail Fracture from Shelling
or Head Check
Defective welds
– Plant Welds (Head)
- Plant Welds (Web)
- Field Welds (Head)
- Field Welds (Web)
Longitudinal Defects in the
Rail Head eg. horizontal split
head vertical split head
Web Defects * eg. head and
web separation split web
Piped rail
Category
I
II
5-10%
65%
55%
11-20%
21-40%
85%
90%
75%
85%
41-80%
98%
95%
81-100%
10-20%
99%
65%
99%
55%
21-40%
85%
75%
41-80%
95%
85%
81-100%
3-5%
6-10%
11-20%
98%
65%
75%
85%
95%
65%
75%
21-40%
90%
85%
41-80%
95%
95%
81-100%
99%
99%
12-25 mm
75%
65%
25-50 mm
95%
90%
more than 50 mm
99%
95%
5-10%
75%
65%
11-20%
80%
70%
21-40%
85%
80%
41-80%
95%
90%
81-100%
99%
95%
12-25 mm
75%
65%
25-50 mm
90%
85%
more than 50 mm
99%
95%
50-100 mm long
80%
70%
100 mm - 1 m
95%
95%
more than 1 m
99%
99%
50-100 mm
95%
more than 100
more mm
than 200
99%
mm any
Any size with non vertical orientation,
evidence of bulged web or progression into weld.
Web Defects in
Joint Area *
eg. bolt hole crack
head and web separation
Reliability Ratio (% of such
defects properly indicated as
flaws in any single test)
85%
85%
90%
95%
75%
12-25 mm
75%
65%
25-50 mm
75%
65%
85%
50-100 mm
90%
more than 100
99%
99%
mm more than halfway through the web.
* defects must have progressed
o 5-33 x
5.2.9 Selecting Rail Testing Intervals
It can be seen from the above that management of rail testing
incorporates two abstract disciplines:
•
Risk management: an outlay of known prevention
costs to achieve a hypothetical reduction in the
probability of damage.
•
Statistical performance: the evaluation of detection
success rates in relation to probability tolerances.
Furthermore, the subject matter is not definitive. The
growth rates of known types of defects are not narrowly
predictable. There are sources of rail breakage that are clearly
not predictable, such as defects in the rail foot, and there are
random events such as infrequent but significant impacts from
wheel defects.
The well documented defect growth experience of the
FAST Heavy Tonnage Loop presents a unique opportunity to
explore the theoretical relationship between rail testing
frequency and the possibility of a service break. For illustrative
purposes, assume that the 11 defects shown in Figure 5.5
represents the full population of defects initiated and growing
over a one year period in a 20 km line carrying 40 million gross
ton annually. Based upon practical experience, it can further
be assumed that a transverse defect left in track with a size
greater than 60% could represent a significant risk of sudden
fracture, and that it would be the objective of ultrasonic testing
to prevent this eventuality.
It can be seen that in the absence of testing, from 1–3 of
these defects could be expected to have covered more than
60% of the railhead by 20 million gross ton of accumulated
traffic; 2–4 more would reach this threshold by 30 million
gross ton. By the time 40 million gross ton had passed over
this rail, it could be speculated that two broken rails would
have been experienced, representing a 1 in 100 chance of a
broken rail derailment using U.S. average statistics.
This case study can be used to illustrate the reduction in
risk that can be expected with rail testing. For example, if
testing at 9 million gross ton intervals (10 mgt), the rail flaw
o 5-34 x
detector car would pass over the defect identified with the
asterisk at the time when it would cover 23% of the railhead.
According to AREA specs in Table 5.3, it would have a 90%
chance of detecting and marking this defect for removal. If the
defect were missed and the flaw detector care were to again
pass over the site at 13 million gross ton (15 mgt), the flaw
would now cover 55% of the railhead and should be detected
with 98% probability. The net probability of detecting this
particular defect before it reaches the 60% size is therefore
calculated as 0.90 + 0.10 (.98) = 0.998. Of course, this makes
the perhaps gross assumption that there is no particular
recurring condition that is preventing detection.
Using this same methodology, one can calculate the net
effect of different test intervals on the probability that one of
these defects will reach the 60% level before being detected
and marked for removal. Using the AREA Minimum
Performance Guideline as an assessment of the typical
detection performance of the test car, the results shown in
Table 5.4 are obtained for the year.
Table 5.4: Effect of Test Interval on Expected Number of
Undetected Defects in a Hypothetical 20 km Line with
Transverse Defect Growth Rates as Measured in the FAST
Heavy Tonnage Loop
Test Interval
Expected No. of Defects that
will reach the 60% Level
Undetected
36 mgt (40 mgt)
11
18 mgt (20 mgt)
0.700
9 mgt (10 mgt)
0.234
4 mgt (5 mgt)
0.004
It can be seen that the probability of leaving a transverse
defect undetected to a size representing a high risk of failure is
very dependent upon the testing frequency. Again, this
assumes that the probability of success is independent from
test to test. On the surface it would appear that very frequent
rail testing is economical, but there is a case of diminishing
returns.
o 5-35 x
Assume for example that it costs $2,500 to replace a rail
detected behind a rail flaw detector car costing $50 for the test.
An emergency replacement, on the other hand would involve
delay of trains and could cost $10,000 per occurrence. For
transverse defects, perhaps 1% of service failures may be
expected to lead to derailments costing an average of $400,000.
Using these cost numbers, the value of the different test
intervals can be calculated from the above probabilities for the
hypothetical 20 km. The results are tabulated in Table 5.5.
This example would indicate an optimal testing interval of
5–18 million gross ton, with 9 million gross ton as potentially
the most economical.
To illustrate the value of reliable testing, assume that a rail
flaw detector vehicle is used that does not meet the AREA
Performance Guidelines, but instead performs as predicted by
Figure 5.9. When testing for defects in this line at 9 million
gross ton (10 mgt) intervals such a car would be calculated to
leave an expected 2.02 defects that would progress to the 60%
level. A comparison of the economics of the two vehicles is
included as Table 5.6.
Test
Interval
40 mgt
20 mgt
10 mgt
5 mgt
Table 5.5: Economics of Test Frequency in a Hypothetical Line
Cost of
Expected
Total Cost
Annual Cost of Defect
Testing
Annual
per year
Repairs
per year
Broken Rail
Detected
Service
Derail. Cost
$1000
$2500
$111,000
$44,000
$158,500
$2000
$30750
$7000
$2800
$44,550
$4000
$31915
$2340
$936
$39,191
$8000
$32490
$40
$16
$40,546
Figure 5.9: Decision Matrix to Direct Risk Reduction
o 5-36 x
Table 5.6: Economics of Rail Testing Performance in a Hypothetical
Line at 9 mgt Testing Interval
Specification Cost of Annual Cost of Defect
Expected
Total
Defining Car Testing
Repairs
Annual
Cost per
Performance
per
Broken Rail
year
Detected
Service
year
Derail. Cost
AREMA
$4000
$31915
$2340
$936
$39,191
TSC/AAR
$4000
$27443
$20,228
$8092
$59,763
Model
This example calculates that the less accurate test vehicle
would cost this heavy haul line an additional $20,000 each year
in this 20 km line segment, or $1000/km. There is clearly
value in maintaining good quality control on testing. In this
example the difference is chiefly due to the differences in the
testing performance in the 30 – 60% size. As a general
statement, if a rail flaw detection vehicle is to be cost effective,
it must be very good at detecting defects such as transverse
defects in the 30 – 80% size range, as this likely will represent
that last time the defect is seen by the detector car before crack
out unless test intervals are exceedingly tight.
5.2.9.1
Performance-Based Adjustment of Test
Intervals
In light of the inexact nature of the science, most heavy haul
railways control risk by monitoring the occurrence of both
detected and service defects.
In North America heavy haul railway practice, risk is
typically judged to be sufficiently high to merit tightening test
intervals when:
•
Service defect rates exceed 0.17 service failures/km
(0.1 service failures/mi/yr).
•
Service plus detected rail defects exceed 0.04
failures/km/million gross ton (0.06 failures/mi./mgt).
•
The ratio of service to detected defects exceeds 0.2.
In fact, risk can be the result of track condition that is not
matched to service demand, test intervals that are not matched
to the reliability of testing systems, or both. The appropriate
o 5-37 x
course of action can be determined by comparing service
failure statistics and the number of detected failures.
Figure 5.9 illustrates the decision matrix that could be used
to direct an effort to reduce risk. For example, if service
failures are exceeding 0.04/km/million gross ton (0.06
failures/mi./mgt), it is apparent that the railway property is
living with a significant risk of a broken rail derailment. The
obvious question is whether ultrasonic testing is reliable and is
being performed frequently enough to find the defects. Should
it also be the case that service failures represents more than
20% of all rail defects recorded, it can be surmised that testing
intervals must indeed be tightened as a first step to reducing
risk. On the other hand, if service failure rates are low, but
detected defects are high, it can be presumed that rail testing is
effectively compensating for a track that may have stress
problems or cumulative fatigue damage.
5.2.9.2 A Parametric Approach
In 1991, Committee 4 of the American Railway Engineering
Association developed a quantitative guidance for specifying
recommended ultrasonic testing intervals. The emphasis was
not on specifying the intervals themselves, but on illustrating
how different railways have perceived the relative effects of
different parameters on risk, and hence the resultant test
interval.
The results represent the experience of two major US
railroads that have developed inspection interval planning
equations. The multipliers suggested to account for different
conditions are given in Table 5.7:
Table 5.7: Inspection Multipliers per Parameter
Significant Parameter
Parameter Range
Inspection
Interval
Decrease
Annual Tonnage Rate
10:1 ratio increase
70-80%
Track Class (as defined by U.S. FRA Class 1 to 6
60%
maximum allowable freight
(16 km/h – 133 km/h
speed)
Existence of Passenger
vs. exclusive freight
50%
Trains
line
Rail section Size
68 kg/m vs. 45 kg/m
70%
Prior Rail Defect Rate
10:1 ratio increase
50-70%
o 5-38 x
5.2.9.3 Cluster Testing
When addressing the risk of a rail-caused derailment, it is wise
to look at the rail plant comprising a routing as a series of
shorter sections of track, with different defect-producing
potential. In older railway lines, lack of homogeneity is a
natural result of the relaying of sections of track in different
years, resulting in rails with different accumulated service
tonnage.
Curved track also generally produces more defects due to
the additional stresses imposed by lateral loading. This occurs
even if the sections of track has rail with the same accumulated
service tonnage with uniformly good track support conditions.
Finally, variations in track and sub-grade support
conditions, rail metallurgical cleanliness and rail weld quality
can have profound influences on defect occurrence rates. As
an example of the impact of metallurgical cleanliness in the
period 1972 – 1980, Canadian Pacific Rail found that fully 38%
of transverse defects had occurred in rails from the “A” or top
position of the ingot, which potentially has the greatest density
of non-metallic inclusions. “A” ingot rails would have
constituted only 18% of the population of rails in track.8
It follows then, that the greatest risk reduction pay-off
from rail testing will result from tests in those locations with
higher defect occurrence rates. The practice of scheduling
additional tests in high defect locations is called “cluster
testing.” To reduce the additional cost of testing intervals
governed by high defect locations, rail-bound rail testing cars
may deadhead without testing over intervening track segments
with acceptable defect occurrence rates.
Many railways schedule “cluster testing” on the basis of
some known characteristic of track. For example, a high
curvature section of an otherwise tangent routing might receive
an extra test, as could a length of older rail, or jointed rail
within a continuous welded rail routing. Other railways
monitor service and detected rail failures and schedule testing
based upon the limits of high defect rate locations.
o 5-39 x
Hi-rail based cars are particularly adapted to cluster testing,
where road access and frequent level crossing permit easy
access to spot locations without tying up track in deadheading.
As there is a cost to deadheading between sites to be
cluster tested, for scheduling purposes, high defect locations
should be 10–20 km in length, with sufficient defect
occurrence rates, when averaged over this length, to trigger the
selected threshold for an additional test.
Therefore, when selecting test intervals for a routing, these
should be somewhat related to the potential for a service
failure, in turn leading to a derailment. Test intervals should
target specific longer track sections with significantly different
characteristics of rail age or quality, rail weight, jointed vs.
welded rail, track support quality and curvature. When, after
tailoring testing intervals to these characteristics of track,
defect occurrence rates in any homogeneous grouping of track
are still high, an additional test should be performed in the
offending location to control overall risk.
5.2.9.4 Special Care in Special Track Work
The turnout area represents a particularly difficult area to test
due to change in the cross-section of the rails and castings.
This means that the ultrasonic echo will return at a different
time than expected or that some probes will contact the
running surface at unusual angles. As a further complication,
castings have a considerably coarser grain structure than the
surrounding steel leading to different base echoes. As a
general rule, only the standard rail cross-sections within the
turnout area and railway diamonds are effectively tested with
flaw detection vehicles.
In at grade crossings, the fouling of the rail surface by
road-borne materials, particularly salt, can obstruct a good
ultrasonic indication. This can be overcome by sweeping out
the crossing in advance, slowing down the test and reversing if
an unusual indication is seen. Welds are another problem.
Because of the change in grain structure and the fact that
fractures can propagate rapidly from very small cracks or stress
raisers, welds are very difficult to test either with ultrasonic or
induction. One possibility is to have automated ultrasonic
o 5-40 x
recognition of the weld upset, which could trigger a change in
the signal gain and the use of tighter inspection tolerances.
Most heavy haul railways ensure more careful testing
through special track work. Some have retained a program of
hand testing, however hand testing typically uses the same
probes that are used by rail flaw detector cars.
5.2.9.5
Rail Testing Intervals – Canadian Pacific
Approach
Canadian Pacific Rail System uses a risk management approach
whereby rail-testing intervals are adjusted according to
different categories of risk. The approach used is to first
segment all tracks into homogenous sections with the same
tonnage, weight of rail, type of traffic and rough levels of past
defect occurrence rates. These segments must be at least 16
km (10 miles) long to be practical for an additional test.
The approach is to first select a basic testing interval that is
dependent upon tonnage, which is a proxy both for the rate of
accumulation of fatigue in the rails and the probability that a
service failure will be encountered by a train. As Table 5.8
shows, there are six basic testing intervals based upon the level
of tonnage.
The testing interval for each track segment may then be
upgraded to the test frequency corresponding to the next
highest risk class if there is an additional element of risk
associated with the track segment. The factors that will qualify
for a more frequent risk are:
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Higher Risk Traffic: Line carries passenger trains
Line carries hazardous materials
Lower Standard Rail:
Non control cooled rail is being used in a
line
Carrying more than 1 million gross ton
per year 50 kg/m (100 lb/yd.) or lighter
rail is being used in a line carrying more
than 2.7 million gross ton/year.
50 kg/m (100 lb/yd.) or lighter rail is
being used in a line where train speeds
exceed 67 km/h (40 mph).
Evidence of Rail Fatigue:Detected rail defects exceed 0.7 defects
per km
(1.2 defects/mi.) per test
Evidence of Low
Inspection Effectiveness:Service failures exceed 0.12 failures per
km
(0.2 failures per mile) per year.
Table 5.8: CP Rail Testing Standards are Based Upon Eight
Testing Frequencies
Traffic Density +
Traffic +
Rail Type +
Defects =
Test Class
(mgt/yr.)
< 0.5
5 yrs.
Hazardous
0.5 – 2.7
3 yrs.
Materials
Non –
> 0.7/km/yr.
Service/detect
2.8 – 7.2
2 yrs.
Cooled
ed ratio > 0.2
7.3 – 13
annual
Passenger
< 50 kg/m
14 – 27
2/yr.
> 70 km/h
> 27
3/yr.
4/yr.
5/yr.
If any three of the above factors are in evidence in the line
segment, the testing interval is tightened by two classes.
Therefore, in the example of Table 5.9, a line carrying 10
million gross ton per year would be tested once per year. But
if it carried hazardous goods a well, it would be tested twice
per year. If in addition to this the line was laid with lighter
than 50 kg/m (100 lb/yd) rail and had a service to detected
defect ratio of 0.25, it would be tested three times per year, or
after every 3 million gross ton of traffic.
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Table 5.9: Test Interval Class is Tightened Based
Upon Risk Factors
Traffic Density + Traffic +
Rail Type +
Defects =
Test
mgt/yr.
Class
7.3 – 13
7.3 – 13
7.3 – 13
annual
Hazardous
Materials
Hazardous
Materials
2/yr.
< 50 kg/m
Service/detected
ratio > 0.2
3/yr.
5.2.10 Induction Measuring Principles
The induction testing technique requires the injection of a
direct current into the rail. The current is generally around
3600A. The injection takes place through the application of
two sets of brushes that are placed on the railhead. The spacing
between the brush sets is of the order 120cm (4ft). The current
flows into the rail through the leading brush set and out
through the trailing brush set. The rail thus becomes part of an
electrical circuit.
Once motion is introduced, a magnetic field associated
with the current flow in the rail is induced. The magnetic field
is the means by which information about the condition of the
rail is coupled to the sensor unit. The sensor unit is located
between the two sets of brushes. The sensor unit is set up to
maintain a constant lift-off between the underside of the unit
and the surface of the railhead. If this is not done, the data
recorded will be noisy and thus very difficult to interpret.
The mechanism by which rail condition is inferred starts
with the current. In general for modern rail weights, only the
head and the top part of the web is “filled” with current. In the
past with smaller rail sections, the whole rail section has been
filled with current. As the current flows through the rail, if any
features such as a defect block the current path, the current
will take the shortest possible route to get around the
obstruction. This distortion of the current flow will also lead to
a distortion of the associated magnetic field. It is this distortion
of the magnetic field that is detected by the sensor unit (see
Figure 5.10).
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Figure 5.10: Distortion of Induced Magnetic Field around
Two Types of Flaws
The sensor unit itself houses multiple coils or Hall Effect
devices. Often the arrangement is differential in nature to help
keep the number of false indications down. By differential it is
meant that two identical sensors located next to each other
across the railhead will be wired together. Thus it is only when
one sensor sees a disturbance and the other doesn’t that a
signal will be sent to the test system. For example, a rail end is
essentially a gross transverse defect, both sensors will see the
rail end so no signal will be sent to the test system. A
transverse defect will generally only be seen by one sensor, so
the asymmetrical disturbance will send a signal to the test
system. Multiple sensors are used to allow the detection of all
of the components of the field disturbances.
Considering the current flow through the rail, as it is
longitudinal, current distortion will not occur as a result of
longitudinal features in the rail. The features that will produce
the most current disturbance are those that are transverse in
the railhead. Unlike the ultrasonic technique, the induction
technique does not have trouble with inspecting right to the
top surface of the railhead. The nature of the current flow is
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such that it is the very center of the railhead that is likely to be
missed if the system is unable to fill the railhead with energy.
The signals sent to the system are generally observed to
determine if they exceed a set threshold. If they do, a count is
started. The number of threshold exceedances then determines
whether the data is presented to the inspecting operator as a
potential defect or not. With increasing computer power new
analysis algorithms, some combining information from
different channels (both induction and ultrasonic), are
becoming more common.
The data can be presented in many different formats. Most
often it is a combination of processed (counted) data and raw
analog data side-by-side. The processed data is often the
mechanism that highlights the problem area and then the
subtle features of the indication can be extracted from the
analog waveform.
5.2.11 Conclusion
While recent research has provided some clues to assessing
risk, it is not yet possible to develop a solid mechanistic
relationship between the risk of derailment and the frequency
of tests. To control the likelihood of service failures, rail can
either be tested very frequently with lower accuracy equipment,
as it will be likely that the rail is scanned while the defect is
larger and more easily detectable. Alternatively, a more
accurate test system can be employed on a longer cycle, as it is
likely that even if the test happened to coincide with the early
appearance of the defect it will still be detected. In fact, heavy
haul operators have found that the costs of poor service
reliability are such that it is profitable to both use very effective
testing systems, with particular emphasis on high reliability in
detection of larger defects, while also maintaining frequent
testing.
5.3 Rail Wear Measurements
5.3.1 Rail Wear Measurement Techniques
It is true in the rail industry as in any other that “What Gets
Measured Gets Managed.” A regular program of rail cross
sectional/wear measurements is critical to the ability to
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proactively plan maintenance and renewal of specific lengths of
rail.
The cross sectional profile measuring techniques can be
categorized as:
•
Manual mechanical feeler and tracing gauges.
•
Manual electronic tracing gauges.
•
Non contact-optical measurement systems.
The manual rail-wear gauge is the most common technique
used to obtain a spot measurement of both the crown and
gauge wear as referenced from the field side bottom crown
corner. Figure 5.11 defines the measurement locations on the
rail cross section relative to the original profile.10
Figure 5.11: Wear Measurements
One widely used system is the MINIPROF portable rail
measurement system. It consists of a notebook computer
connected to a rail-measuring unit that magnetically clamps to
the rail-running surface with a gauge bar resting
perpendicularly on the opposite rail. Data communication
between notebook and measuring unit takes place via
dedicated electronics built into a small extension box
connected to the parallel printer port of the notebook
computer. The sensing element consists of a small magnetic
wheel, with a diameter of about 12 mm, attached to the
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extremity of two joint extensions. This magnetic wheel ensures
contact during measurement with the rail surface. By moving
the magnetic wheel manually the extensions rotate. The
measuring system uses a polar coordinate system with 2
degrees of freedom. The two angles are measured with the aid
of optical encoders, having accuracy in the order of microns,
see Figure 5.12.11
reference
l
1
ø1
optical
encoders
l
2
ø2
magnetic
wheel
profile
Figure 5.12: MINIPROF Measurement Principle
The computer samples the transducer data, which are in
polar coordinates, and calculates the profile in Cartesian
coordinates. By averaging closely spaced values accuracy is
further improved. After these calculations the true profile is
displayed on the computer screen together with a reference
profile and some characteristic parameters. Besides, digital data
are calculated and stored in ASCII format for later processing.
These data consist of x and y coordinates. The measured
profile is overlaid and aligned with a reference profile and
crown gauge and side wear is calculated as shown in Figure
5.13
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Figure 5.13: MINIPROF Wear Calculations
The use of the MINIPROF has been expanded to
incorporate wheel, turnout, point, and frog profile
measurements. Figure 5.14 shows the measuring set-up.
An example of successive cross sectional Frog profile
measurements is shown in Figure 5.15.
While the MINIPROF is widely used there are several
other Systems (either electronic or laser based) that are used to
conduct the same functions as the MINIPROF.
While it is feasible to annually measure in every curve by
hand, automated systems developed over the last 20 years are
now capable of measuring the rail profile to accuracy of +/0.127 mm at track speeds. These systems became feasible due
to the development of real-time processing hardware and highspeed personal computers. The technology is currently widely
accepted as an efficient method for the collection of rail
profiles and determining rail wear characteristics for most large
heavy haul railways and numerous other general freight as well
as high-speed passenger and rapid transit applications.
The available systems use optical measurements, normally
mounted underneath a track geometry car, grinding machine or
hi-rail vehicle. Cross section of the rail is typically illuminated
by a laser and captured using a high-resolution video cameras.
One example of such a system is the ORIAN (Optical Rail
Inspection & Analysis) system set-up is shown in. Figure 5.16.
12
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Figure 5.14: Examples of MINIPROF on Wheel, Rail,
and Turnout
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Figure 5.15: Examples of MINIPROF frog measurements
Figure 5.16: ORIAN system key components
This system consists of four key components:
1. Two rail measurements sensor heads;
2. Control Electronics;
3. Central Computer; and
4. ORIAN Network Interface for data communications
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Each of the two rail measurement sensor heads consist of
two high resolution CCD cameras and two laser modules. The
measurement sensor heads are attached indirectly to the axle of
the vehicle to allow for the system to measure in any degree of
curvature and provide accurate rail inclination and gauge
measurements.
While the vehicle is moving encoder pulses are generated
at precise intervals (i.e. every 3 m). These encoder pulses tell
the video cameras and lasers to snap a picture of the rail. Each
rail image acquired by the system is translated into real-world
X-Y coordinates by the computer and synchronized to a
unique track location. From this coordinated rail profile data
the rail dimensions and type can be determined and the wear
and metal flow can be calculated by comparing the current
profile to the original profile. Additionally the rail inclination
and track gauge can also be determined from the relative
position and orientation of rail profiles. The rail inclination or
rail seat cant and gauge can also be used to identify locations
where poor sleeper conditions may contribute to increased rail
wear.
The rail profiles and wear measurements are computed in
real-time which allows for the immediate display and reporting
of rail wear parameters. Rail wear can be reported in many
ways. Automated systems allow for the collection of vast
amounts of data due to the frequency of measurements.
Typically railways will configure these systems to measure the
rail every 2-5 metres and collect data on their rail condition
from 1 – 4 times a year. Crown rail wear side wear rail cant and
track features can be plotted in the format as shown in Figure
5.17
Figure 5.18 shows the display of both rail’s cross sectional
profiles.
o 5-51 x
Figure 5.17: Display of Measurements
Figure 5.18: Cross Sectional Display
5.3.2 Rail Wear Projection
When comparing the priorities for rail renewal between many
different locations, some railways use a summary report, which
identifies each curve or length of tangent as a single Rail
Quality Index. The usual index number is expresses as a
percentage of the maximum wear limit. With the use of
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automated rail wear measurements, it is possible to use a
statistical average of several rail wear measurements in the
curve or tangent. Use of the 90th percentile rail wear
measurement, representing the wear measurement exceeded by
10% of measurements is used as the planning standard on
Canadian Pacific Railway. Figure 5.19 presents one such report
plot.
In this display, each bar represents an individual curve or
1/10 mi. (0.16 km) of tangent rail. Often, these reports include
a projection of the time for renewal or transposition of the rail.
In Figure 5.19, the cross-hatching represents wear projected 2
years into the future. The TQI = 100 is the 100% wear limit
level.
Rail wear projections can be on the basis of statistical
studies of rail wear, engineering estimates, or empirical
projections based upon the current wear level measured in the
curve, divided by the accumulated tonnage since new. With the
use of an accurate rail wear measurement (within 0.4 mm), it is
more reliable to make projections of rail wear specific to each
individual curve. This of course requires a good database on
the year of installation and accurate track locations.
Once a rail wear projection procedure has been
established, a rail renewal estimate can be drawn up. Rails
identified for renewal or transposition can be identified for
both left and right rail if the rail wear limit will be reached
within 5 years of the rail measurement. Such a report should
also integrate similar projections of the progress of rail defect
occurrences to a specified trigger limit.
o 5-53 x
Figure 5.19: Canadian Pacific Railway – Rail Wear Index Report
5.4 Rail Profile Maintenance Practices
Two techniques are used to maintain the required rail profile.
The more commonly used grinding practice where the amount
of metal to be removed is limited and rail planing generally
used to salvage rail that has been neglected to such an extent
that large volumes of material must be removed to return the
correct profile
5.4.1 Rail Grinding
5.4.1.1 Objectives of Rail Grinding
The natural processes of wear and deterioration of rail steel can
proceed at a pace that results in a long service life, or they can
result in rapid condemnation of a rail. The difference lies in
the applied contact stresses and the yield strength of the steel
itself. When these are not in balance, the rail economy suffers.
o 5-54 x
Then grinding of rails has evolved as a maintenance
technique to insert some control of the processes of rail
surface fatigue and plastic flow.
Heavy haul lines in particular are characterized by frequent
occurrences of both surface and shape distortion through
plastic flow. On mixed freight lines ride quality and noise
abatement may be the major concerns, while track occupancy
times may be very limited. In spite of very different practices
on different railways around the world the objectives are the
same:
1. Attempt to maintain a balance between all wear
mechanisms to prevent premature replacement due to
surface or subsurface fatigue of rail.
2. Facilitate steering and dynamic stability of vehicles.
3. Control rail surface conditions that give rise to higher
dynamic loadings and track vibrations.
4. Control the wheel/rail contact and interaction
characteristics.
5.4.1.1.1 Longitudinal Rail Profile Correction
The vast majority of rail grinding is performed by rail-bound
machines using rotating grinding stones. The stones are
annular, the flat side being applied to the rail as opposed to the
edge of the stone, as in machining applications.
The removal of metal occurs through the abrasion and
gouging action of the rotating cutting grains. Metal removal is
dependent upon the characteristics and condition of the
abrasive and the application pressure on the grinding wheels
(Figure 5.20) as well as the grinding speed and the angle
between the grinding stones and the surface being ground.
The removal of rail corrugations and other longitudinal
irregularities in the rail surface occurs through:
1. Ensuring that the stone bears down on the peaks of
the corrugations in preference to the valleys.
2. Surging power and pressure on the stone to produce a
differential cut when encountering a corrugation peak.
o 5-55 x
The diameter of the stone itself acts as a baseline that
ensures that corrugation peaks of shorter wavelength are
bridged by the stone and are thus subject to preferential
grinding. The usual stone diameter is 250 mm, so that the
grinding stone easily works to establish a level plane over this
span. For corrugation wavelengths longer than the wheel
diameter, it is very important that the individual grinding
motor is not permitted to follow the profile of the corrugation,
for to do so would result in the same amount of metal removal
from the valleys of the corrugation as from the peaks.
One way of extending the baseline is to block two motors
together so that they will rise and fall in unison. This causes
them to bridge over all irregularities in rail surface over a
longer rail length (Figure 5.21). Another method is to control
the rate at which the grinding stone will be permitted to rise
and fall in following the rail surface using a hydraulic cylinder
or active feedback based upon changes in motor torque.
Figure 5.20: Grinding Metal Removal is Dependent Upon
Pressure and Angular Orientation of Stone
The profiles shown are exaggerated in the vertical scale and
shifted to line up
o 5-56 x
Figure 5.21: Bridging of Grinding Motors and Damping
Facilitate Longitudinal Profile Correction
The damping in the motor suspension means that any
bump in the rail surface with a wavelength of less than 1 m will
result in an increase in the contact pressure between the
grinding stone and the rail. This will increase the torque on the
motor and increase the current draw. If the resulting power
surge is permitted, the energy input will increase the depth of
cut at the leading edge of the corrugation. Most high
production rail grinders will permit a timed overload of the
electric motor, while regularly grinding on straight level rail at
the rated capacity of the motor.
Similarly, as the rail surface drops away at the trailing edge
of the corrugation, the motor torque drops off, and with it the
depth of cut.
As corrugations have a sinusoidal shape, grinding of the
peaks initially produces substantial reduction in the amplitude
of the corrugation. But as the peaks are ground, the area of
metal in direct contact with the grinding stone on subsequent
grinding passes increases and the depth of cut is proportionally
less.
5.4.1.1.2 Transverse Rail Profile Correction
Railways with a more mature grinding practice will typically
devote most grinding to the preventive practice of reshaping
the transverse profile. Rectifying the profile in the transverse
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plane can improve the contact geometry between the wheel
and the rail. Producing conformity between the worn wheel
and rail reduces the contact stresses. High contact stresses are
the cause of plastic flow and surface fatigue such as spalling,
shelling and head checks.
Internal stresses which give rise to rail defects within the
railhead, such as transverse defects, are also directly related to
the level of contact stresses. Internal stresses are also increased
when the wheel is permitted to ride heavily on the gauge
corner of the rail, or far to the field side of the rail, due to the
torsion of the railhead.
So from the point of view of rail stresses, rail grinding
should seek to maintain a rail shape that conforms well to the
worn wheel, while ensuring that the wheel's reaction is
supported through the web of the rail.
Another objective of transverse rail re-profiling is to
support good vehicle behavior, particularly in curving. In the
practice of "profile grinding," the head of the rail is re-profiled
to control the location of contact between the wheel and the
rail. This is illustrated schematically in Figure 5.20. By
concentrating grinding on the field side of the high rail and on
the gauge side of the low rail, the wheel contact is shifted from
point A to point B in Figure 5.22. The outer wheel is thereby
forced to ride toward the gauge side of the high rail leg and the
inner wheel to ride toward the field side of the low leg. The
differential in the rolling radius causes the wheelset to steer
toward the smaller wheel radius, i.e. the inner wheel. This
produces limited self-steering up to 873 - 700 m radius of
curvature. Note in Figure 5.22 that the gauge corner of the
high rail has also been relieved slightly to avoid contact with
the throat of the wheel. This dresses gauge corner head
checking and eliminates a high stress contact. By grinding to
form a radius of curvature on the railhead that is slightly
sharper than the radius of curvature of the head hollow on
typical worn wheels, the contact band could be tightly
controlled.
o 5-58 x
The objectives of maintaining conformal contact,
minimizing loading eccentricity and maintaining loading
reaction through the web of the rail can be at odds with one
another. In contrasting the re-profiling practices of different
rail most of the differences can be seen to be related to the
relative importance placed upon control of rail fatigue vs.
wheel and rail wear in curve negotiation.
In fact, flat annular grinding wheels can only produce an
approximation of the "optimal" rail profile. Worn rail and
wheel shapes are comprised of intersecting curves of varying
radii. But each grinding stone grinds a flat facet into the rail.
By sequencing a series of stones at different angles along the
length of the grinder, each subsequent stone grinds on a new
plane (Figure 5.23). As a result, the as-ground rail profile is
actually a polygon approximation to the curvatures of the
desired rail profile.
Figure 5.22: Asymmetric Profile Grinding Increases Rolling
Radius Difference by Shifting Both High and Low Rail
Contacts to Inside of Curve
o 5-59 x
Figure 5.23: Typical Geometries of Cross Sectional Areas
Removed During (A) And (B) the First and (C), (D) and (E) the
Following Grinding Machine Pass(Es)
5.4.1.1.3 Effects of Rail Shape Parameters on Rail
Damage
Simulation studies indicate that a corrugated rail surface will
significantly increase dynamic wheel loads (Figure 5.24). The
increased loading is non-linear with both the depth of
corrugation and the train speed. It is generally felt that
corrugations need to be controlled to within 0.5 mm in slow
speed track to avoid noticeable foreshortening of the life of
track components. This limit needs to be tightened to 0.25
mm in track with speeds of 80 km/h or greater.
o 5-60 x
Figure 5.24: Estimated Effect of Corrugation Depth and Speed
on Dynamic Vertical Contact Forces
Contact too close to the gauge corner of the outer rail
(Figure 5.25), while ideal to promote steering, produces a stress
situation that is difficult to manage under conditions of heavy
axle loads or high curvature. The reasons are:
1. The radii of curvature of the wheel throat and the rail
gauge corner, while of the same orientation, are small,
so that any mismatch has a large relative effect on the
size of the contact area. This amplifies contact
stresses.
2. Single-point wheel rail contact at the gauge corner
results in a single contact area supporting all vertical
and lateral forces, as opposed to having these spread
between the wheel flange and the wheel tread.
3. The contact area is not planar to the rotational axis of
the wheelset, causing a component of spin creep. This
spin creep contribute to material flow.
4. Contact near the gauge corner occurs in an area with
very little supporting metal. An incremental shear
collapse of the corner can result from excessive
loading, analogous to a progressive slope failure in a
railway embankment.
o 5-61 x
Figure 5.25: Single Point Contact between Wheel and Rail
Finite element studies show that compressive stresses in
the railhead are concentrated just under the gauge corner and
are more than double the case where the gauge corner has
been undercut by grinding.
Contact of a wheel with a worn hollow on the field side of
the rail is also generally avoided through rail grinding and good
control of gauge. (Figure 5.26). Contact between the wheel
and the far field of the rail would typically occur at a point
where the wheel tread profile has developed a reverse tread
curvature. Again, a high contact stress can be expected.
Hollowed wheel treads produce high stresses and overturning
moments on the field side of the low curve rail.
Figure 5.26: Removal of field side metal from curve low leg
reduces eccentric loading of rail
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In addition. contact with the far field of a rail produces an
eccentric loading condition. The resulting torsion of the
railhead can lead to vertical split heads or "shear breaks,"
particularly if the rail is worn to condemnable limits.
5.4.1.1.4 Grinding for Surface Condition
Grinding is also performed to correct rail surface conditions
that will lead to further rail deterioration. Cooper 13classifies
this type of grinding as:
1. Preparative: Cleaning mill scale or nicks introduced in
construction from newly laid rail, smoothing high
welds, to ensure a good start to revenue service for the
rail.
2. Preventative: Removing layers of fatigued metal before
micro cracking leads to more serious damage.
3. Curative: Recovering rail that has been damaged by
engine burns, ballast pressed into the rail surface.
Preventive grinding for removal of fatigued metal may
require regular removal of only 0.15-0.4 mm of metal from the
surface. Surface cracking is seen to reduce the rail's resistance
to flow and is observed to accelerate under heavy axle loads or
poor contact geometry. Of course, all grinding will involve
some rectification of a fatigued surface. However, for effective
preventative grinding, the treatment must be done frequently,
at intervals of roughly 12 million gross ton in curves,
depending on a wide range of conditions, including: rail type,
severity of track curvature, grinding equipment, axle loads, and
wheel/rail profiles.
In spite of the reason for the grind, any grinding pass does
contribute to the removal of fatigued surface materials. In
addition, the regular removal of surface metal through grinding
may postpone the initiation of rail shelling defect initiation.
Theoretical studies of the effects of vertical wear rate on
rail life suggest that the fatigue life of rail can be extended by
progressively moving the point of maximum fatigue damage
down through the railhead by vertical wear and metal removal.
The longest net rail life under traffic with 30 ton axle loads was
o 5-63 x
found to result when the rail gauge corner is worn at a rate of
0.02-0.05 mm per million gross ton of traffic. This is 3-4 times
the natural wear rate to be expected in the absence of grinding.
Therefore, in territories where rail is removed because of rail
shelling or transverse defects as opposed to achievement of
wear limits, it may be economical to supplement natural wear
with artificial wear in the form of rail profile grinding.
It is shown in this work that it is possible to actually wear
away rail shells before they grow to detectable size. On the
other hand, it has been demonstrated in the same study that
excessive rates of metal removal may contribute to vertical split
heads, emanating at 10-12 mm below the contact surface. As
there is the obvious penalty in early achievement if vertical
wear limits, this strategy would require a large section rail
and/or extended wear limits, and would only be merited in a
location with a history of surface fatigue defects.
5.4.1.2 Grinding Stones and their Effects
5.4.1.2.1 Abrasive Stone Technology
The performance of the rail grinding stone is the key to
effective and productive rail grinding. Grinding stones are
specifically engineered to perform within a given range of
contact pressures, revolutions per minute and heat inputs. In
effect they are engineered to balance good cutting performance
for a given range of energy input to the cutting surface- and to
maintain its performance over a long service life. Proper
matching of the grinding wheel/abrasive and the grinding
equipment is an important feature in an efficient grinding
operation.
Production grinding wheels consist of a disc of varying
thickness, starting between 50 and 75 mm in thickness. These
discs are made up of a matrix of thousands of abrasive grains
held together in a synthetic resin-bonding agent (Figure 5.27).
Each grain acts as a cutting tool and the bond material is the
"tool post.” A good abrasive is one which:
1. fractures along many different cleavage lines, each
fracture producing a sharp cutting surface;
2. is resistant to abrasive wear for long life; and
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3. has a moderate to high fracture toughness to prevent
premature fracture.
Figure 5.27: Fracture and Release of Cutting Grains at
Grinding Wheel Interface Maintains Cutting Performance
As the wheel rotates at around 3600 RPM, the cutting
grain gradually dulls. A good abrasive will then continue to
fracture along its cleavage planes to expose new cutting edges.
For example a zirconia (fired) alumina abrasive will continue
microfracturing to remain sharp through 80% of its life. For a
regular aluminium oxide grain, the equivalent figure is 30%.
At some point in its life, the grain becomes too small to
regenerate a cutting surface and will remain dull. As this
continues, the friction on the grain builds up and it absorbs
heat. The bond is formulated to respond by ultimately
permitting the entire grain to dislodge from the matrix,
exposing a fresh grain beneath. And in fact the abrasive grains
are spaced in the stone's structure to specifically permit this
without clogging the grinding interface. Hence the process
starts again.
If the contact pressure at the grinding stone interface are
less than the design envelope, "friction plowing" occurs rather
than the desired cutting action. The abrasive grains may not
dislodge on dulling and the stone will glaze up. This causes the
wheel to heat up resulting in discoloration of the rail. Good
grinding performance requires a stone selected for the
operating parameters and good grinding power and pressure
control and dynamic motor stability. Section 5.5 will discuss
several ways to recognize the quality of the work done at the
critical grinding wheel interface.
o 5-65 x
Stone life is a major concern to the grinding contractor as
it represents a high proportion of his costs. The railway
authority, on the other hand, is concerned with the volume of
metal removed per hour. As a compromise, stone life should
be at least the length of the longest track occupancy (e.g., 6-8
hrs.).
5.4.1.2.2 Surface Finish
There are two aspects of the surface finish of the as-ground
rail; the ridges left by the facets and the surface roughness left
by the grinding marks or scratches.
The ridges left by the polygon approximation of grinding
to the desired rail profile are related to the sequence of the
stone angles, in particular the last 6-8 stones to grind the rail.
Pattern design targets controlling the difference in the angles
of subsequent cuts to within specified limits to avoid large
ridges, which may become the sites of flow over or surface
fatigue.
The action of grinding stones produces a saw-tooth shape
of surface profile. These grinding "scratch" marks trace the
rotation and forward advance of the cutting grains. The rough
nature of the as-ground rail surface is dependent upon the
stone grit size and the grinding motor control. Surface finish
may or may not be a concern to a railway authority. The stated
concerns are:
1. If the ridges fold over under traffic, the valley of the
ridge may act as an initiation point for a micro-crack.
2. The heat of grinding forms martensite of the surface.
3. The ground surface will result in an audible whine
under passage of trains for several days after grinding.
The particular opinion of the railway authority on the
importance of the surface finish has very important
consequences for the cost of rail grinding. If a rougher surface
finish can be tolerated, a considerably more aggressive grinding
stone can be used. Aggressive stones are characterized by a
zirconia alumina cutting grain, which produces a swarf-like
particle. The payoffs are big, as there is at least a six fold
difference in the volume of metal removed per pass at the
o 5-66 x
same grinding speed between the aggressive stones used in
North America and the finishing stones common to European
practice (Figure 5.28).
Figure 5.28: Metal Removal Per Pass Aggressive vs.
Finishing Stones
A smoother rail finish is not required of heavy haul lines
but is sometimes specified for passenger lines or in urban
areas. In this case a finer grit stone can be specified.. Grit size
refers to the physical size of the abrasive grain particles.
Production rail grinders generally use grain sizes between a 14
and 18 grit. Grit sizes can be used in rail grinders up to a 32
grit for a fine finish, with a corresponding reduction in metal
removal rates, and some risk that the grinding stone can
become contaminated.
Tests on Canadian Pacific Railway with rail sections
removed from freshly ground rail14 have confirmed that some
of the heat of grinding is indeed absorbed in the uppermost
layers of the steel, transforming it to hard martensite.
Fortunately, the martensite is brittle and is confined to the
peaks of the ground surface. As a result, it is sheared off with
the passage of a few long trains. Any cracks that may be
formed in the process are significantly smaller than those that
o 5-67 x
develop through contact fatigue, and within days, the running
surface of the rail is usually restored to its original surface
roughness. It is not apparent, however, that this benign
mechanism can be extrapolated to European experience, as the
"disappearance" process for scratch marks may be different
under light axle loads.
5.4.1.2.3 Effects of Speed and Pressure
Grinding stones that operate at the corners of the railhead will
be subject to higher contact pressures and will tend to cut
more deeply on a ratio of perhaps 4 or 5 to 1 versus a stone
grinding on the top surface of the rail. Some grinding
contractors use different grades (hardness) of stones to achieve
good stone life for stones confined to work on the far field and
gauge of the rail. Hardness is related to the amount of bond
material used, which in turn determines the ease with which
grains are released from the matrix. If the bond material is too
hard, the wheel can overheat; if too soft, the wheel life is
short. In spite of the use of harder stones, there is nonetheless
a tendency for a "cupping" to develop on the stone and it is
good practice to alternate stone angles to "dress" the cutting
surface.
Similarly, repeat patterns of the rail grinder should alter
stone settings to avoid repeated grinding on the same planes.
Grinding production is dependent upon always grinding at an
angle different from the previous cut.
o 5-68 x
Figure 5.29: Relationship between Metal Removal Rate (MRR)
and Grinding Power (kW)
The cut rate delivered by an individual grinding stone is
directly proportional to the applied grinding (Figure 5.29). On
the other hand, grinding wheel life is reduced with increasing
power. Therefore it is important for the railway supervisor to
ensure that grinding motors are drawing the desired amperage
to ensure a good production rate. Some railway authorities
also require grinding contractors to submit records of stone
consumption.
A rail grinder with a good versatile stone composition and
dynamically stable mounting of grinding motors will exhibit a
linear inverse relationship between depth of metal removal and
the speed of the grind. Within the operating speed range of
the grinder, which may be 3-20 km/h, it can usually be
assumed that a doubling of speed will halve the depth of metal
removal. In fact, recent research has indicated that at higher
speeds, some of the energy of the forward advance of the
grinder may result in proportionally greater cutting rates at
higher speeds. Again, this assumes that there continues to be a
balance between the stone composition and the energy in the
system.
o 5-69 x
The pivot point of the rail grinder motor is also important
to ensure that the grinding stone maintains contact within the
rail on an axis that passes through the center of the inside
diameter of the stone under all angle positioning. Where this
does not occur, the stone will contact on its side instead of the
leading edge of the stone. This reduces the length of the
contact path for each cutting grain and may result in cupping
of the stone.
5.4.1.3 Grinding Patterns and their Use
A grinding pattern refers to a combination of grinding motor
angle settings and accompanying pressures, which enable the
sequence of grinding stones passing over the rail to produce a
given net reshaping of the rail.
Corrective grinding patterns can be described by the
relative percentage of grinding motors deployed in each of the
six key grinding zones of Figure 5.30.
Figure 5.30: Rail Grinding Patterns are Described by the
Relative Numbers of Stones Displayed in Six Key Segments
o 5-70 x
The following describes the effect of directing grinding
effort in each of these zones.
Increase Relative Grinding Effort by:
Moving Stones to:
If Desired to:
Far Field
Dress field side lips on low rail.
Provide false flange clearance on low
rail for locations with good gauge, e.g.
on concrete ties.
Increase the rolling radius riding the
high rail to improve steering in shallow
curves by moving contact to gauge.
Near Field
Sharpen head radius on excessively flat
low rails.
Provide false flange clearance for
locations with wide or varying gauge.
Center
Broaden a contact band showing
evidence of excess stress such as one
over peaked by frequent profile grinding.
Remove corrugations that have
developed to their deepest on rail
center.
Remove head checks in the contact
band.
Dress engine burns or high/dipped
welds.
Near gauge
Reduce the rolling radius riding the low
rail to improve steering in shallow
curves by moving contact to field.
Flatten a former high rail that has been
transposed to the low rail position.
Remove gauge corner cracks and spalls
that are progressing towards the center
of the rail.
Mid gauge
Remove gauge corner cracks and spalls
in territories with good gauge.
o 5-71 x
Remove gauge corner flow in low rails.
Dress a flow lip and associated
subsurface cracks transposed to the
gauge corner of a high rail from the field
side of a former low rail.
Provide conformal contact or relief from
contact between the gauge corner and
throat of the wheel.
Deep gauge
Remove spalls and offload shells in
severe shelling conditions, or where
poor tie conditions and/or sharp curve,
permit rail to hollow.
Remove gauge side flow lips.
Figure 5.31 presents a sample diagram detailing a grinding
pattern for the 44 stones grinding one rail for an 88-stone
production rail grinder. The dots on the grid at the top of the
page plot the angle setting for each motor. The numbers from
top to bottom refer to the carriage or module number. Each
module controls 3-4 grinding motors. Each radial line plotted
on the rail cross-section below indicates the center of where
each grinding facet would fall on a new 68 kg/m rail on a 1 in
20 rail cant. This particular pattern will perform preferential
removal of metal from the gauge corner of the rail to work on
gauge corner cracking or spalling.
o 5-72 x
Figure 5.31: Example of a Grinding Pattern
Note in Figure 5.31 that the dot map of motor angle
positions makes an "X" shaped pattern across the rail. This is
done deliberately to specify that each stone will be angled
relative to the last facet for good grinding production. This
improves the cut rate and ensures that the facets are separated
and distinct. The angular difference between subsequent
motors is important to establish the pitch of the ridges left
o 5-73 x
after grinding. On the gauge corner, a tight spacing of facets is
necessary to reform the gauge without sharp ridges.
Table 5.10 lists a set of standard patterns used by Canadian
Pacific Railway to deal with different rail conditions. These
consist of both corrective patterns designed for heavy grinding
on the gauge, field or center of the rail, profiling patterns
designed for slight alteration of the rail profile or to shift
moderate grinding effort to different locations on the railhead,
and maintenance patterns, designed to cover all areas of the
railhead without changing the rail shape that already exists.
Note that the list contains patterns providing different levels of
attention to each of the zones shown in Figure 5.30.
Table 5.10: Sample Patterns used by a North American
Heavy Haul Railway
Pattern Type
Pattern Description
Corrective
Severe Gauge Relief
Patterns
Severe Field Relief
Centre-Concentrated
Extreme Field and Gauge
Concentration
Profiling Patterns
Tangent profiling
High Rail - Mile Curve Standard
Profiling
High Rail - Sharp Curve
Low Rail - Standard profiling
Tangent - Heavy Gauge Cut
Maintenance
High Rail Profile Maintenance
Patterns
Low Rail Profile Maintenance
Equal Metal Removal
The usual patterns used by Canadian Pacific Rail for the
initial profiling of both high and low rail leg in mild curvatures
are illustrated in Figure 5.32 in terms of both histograms
showing the distribution of grinding effort, and the usual depth
of grinding at 4 km/h.
o 5-74 x
Figure 5.32: Standard Canadian Pacific Railway Grinding
Settings and Results for Reprofiling of Intermediate Curves
Corrective Grinding Patterns
Corrective grinding profiles are designed for heavy
concentration of grinding effort on field, gauge or center of the
rail. They are used as initial passes when the longitudinal
profile is very corrugated, or if the transverse profile departs
substantially from the target profile. They are also used to
address plastic flow lips. Corrective patterns are also called
"production" patterns.
Profiling Patterns
Profiling patterns are designed to apply a specified feature
to a rail profile, while providing full coverage of the rail ball.
o 5-75 x
They might provide, for example, a slight additional field side
or gauge relief, as judged to be desirable as a target rail profile.
Maintenance or Finishing Patterns
Maintenance patterns are designed to retain the existing
rail shape, or to add slight correction. When these are used, it
is judged that the rail has already attained the optimal shape.
They are typically used where rail is being ground frequently, to
remove light surface fatigue. Maintenance patterns are geared
more toward equal metal removal across the full railhead.
They are also useful as a final higher speed pass after multiple
passes using corrective grinding patterns. As a finishing pass,
they help in blending facets left from previous passes, and in
ensuring that the full head has received some grinding.
5.4.1.4 North America Grinding Practice
North American freight railways make extensive use of rail
grinding primarily to prevent premature rail replacement from
surface fatigue. Axle loads of 30 ton and greater, plus the
imperfect rail support conditions offered by timber sleepers
and a wet environment combine to place a premium on the
fatigue resistance of rail. North American tolerances on rail
surface condition are typically broad, but are tightening with
adoption of more frequent grinding. A typical specification on
rail surface conditions requiring grinding is at least 50% of
corrugations in a curve exceeding 0.25 mm in depth. When
grinding, the transverse profile might typically be left within
0.12 mm of a template gauge in main lines and 0.25 mm in
secondary lines.
Historically, the control of rail corrugations has been the
impetus for rail grinding. In the past, however, grinders have
operated with fixed stone positions and an inability to grind at
sharp angles. As a result, grinding for corrugations would
typically flatten the rail while long grinding intervals would
ensure that fatigue cracks were left in the valleys of the
corrugations to weaken the rail's resistance to plastic flow.
Grinding was seen as a temporary measure, as corrugations
would return quickly to the same locations.
The first major breakthrough in the productivity of rail
grinding in North America occurred when it was determined
o 5-76 x
that rail corrugations were initiated by high contact stress
between the "false flange" of a worn wheel tread and the field
side of the rail, particularly under wide gauge conditions.
Preferential grinding of the field side of the rail (Figure 5.33)
was found to reduce the occurrence of high stress contacts
toward the field side of the rail, with the result that corrugation
growth rates were retarded.
It has also been found that the problem of rail gauge
corner checking, spalling and ultimately shelling could be
reduced markedly with aggressive relief of the rail gauge
corner.
Figure 5.33: Field Side Grinding of Low Leg to Retard
Corrugation Re-Growth Rates
The trends recently have been to move away from the
aggressive relief of the low leg field side and high leg gauge
corner toward a more conformal contact between the worn
wheel and rail. The logic has been to spread fatigue damage
over a broad area of the rail crown. But railways that have
departed too much from the themes of high rail gauge corner
relief and low rail field have seen an increase in internal rail
defects.15
Railways in North America have now virtually
standardized upon regular undercutting of the field side of the
low leg. Most have also standardized upon undercutting of the
high leg gauge corner, however there is some variance in the
extent of the undercutting.
o 5-77 x
Figure 5.34: Differences in Treatment of the High
Rail Gauge Corner
Figure 5.34 illustrates the differences in treatment of the
high rail gauge corner. A conformal contact is advisable to
promote steering. The extent of undercutting of the gauge
corner is dependent upon just what is required to control
gauge corner fatigue. Extreme undercutting of the gauge
corner should be avoided as it increases lateral forces.
5.4.1.5 Optimizing Rail Profiles
North American railways have standardized on the National
Research Council of Canada's eight 200 mm 8” railhead
profiles (Figure 5.35) that were presented in 1991. Beginning
with the TT profile designed for (tangent track profile), the
H1-H4 provide progressively larger gauge corner relief, in steps
of about 0.5 mm, while the L1 to L3 provide progressively
larger field side relief. The graduate relief allows this family of
profiles to be applied to a variety of track conditions, including
wide gauge, large dynamic rail rotations and soft metallurgy.
The eight-inch railhead was a compromise – the “ideal” value
was recognized as 10 inch. But given the rates of wear and
plastic flow over a typical grinding interval, a rail that starts
with an 8-inch head was found to flatten to a 12 inch,
providing the favored 10 inch as an average for that grinding
cycle.
o 5-78 x
Figure 5.35: NRC’s Complete Set of Eight
200 mm (8 in.) Profiles
The NRC templates were developed, and application
guidelines provided, to maintain a two point conformal contact
with respect to the average worn wheel. This was required to
prevent gauge-corner collapse under excessive one-point
contact loads. Proper gauge face lubrication is recommended
to minimize the impact on wheelset steering performance.
Improved rail steels, elastic fasteners, high horsepower railgrinders and accurate profile measurement systems are
prompting further evolution of rail grinding. Railways that
have made the commitment to rigorously maintaining a
preventive rail-grinding interval are finding it possible to
control gauge corner fatigue with less aggressive relief of the
gauge corner. It is proving possible to retain control over rail
fatigue while promoting good steering action. An example of
how template selection has evolved in the last decade is shown
in Table 5.11.
Table 5.11: The evolution of NRC template application on the BNSF Pacific
Northwest (new starting 1998).
LOCATION
Sharp corrective
Preventive > 7°
Preventive 3.5° to <7°
Preventive 1.5° to < 3. 5°
Preventive < 1.5°
Tangent
HIGH RAIL
OLD
NEW
H4
H4
H4
H3
H2
TT
H3
H2
H2
H1
TT
TT
LOW RAIL
CURVATURE GAUGE
≥3.5°
≥3.5°
≥3.5°
<3.5°
<1.5°
o 5-79 x
> 25 mm
< 13 mm
< 13 mm
< 13 mm
All
NEW
L3
L2
L1
TT
TT
More recent rail grinding programs on BNSF and
Canadian Pacific have included the development of optimized
rail profiles as an integral element of the modernized program.
These profiles are custom designed to promote a healthy
interaction with the specific wheels and car types running over
that territory. A separate high-rail profile for mild and sharp
curves is common and one or two low rail profiles are usually
required. On both BNSF and Canadian Pacific, two tangent
rail templates are being applied to spread contact across the
wheel tread to inhibit hollowing, a practice that has proven
successful at Spoornet and QCM.
5.4.1.5.1 Rail Profile Design
In the 1990’s, rail profiles were developed from the average
worn wheel profile, using computer aided design packages.
Considerable emphasis was necessarily placed on developing
the proper average wheel. The goal of the design exercise was
to generate rail profiles that maximized stability in tangent
track, but minimized contact stress without overly
compromising steering in curves. In heavy haul environments,
a 2 point conformal contact was usually sought, while in
transits the objective was a one point conformal contact.
The NRC uses the term conformal to refer to the general
condition where (as per the Webster's definition) the wheel and
rail profiles have "similar shapes.” Figure 5.36 shows NRC's
definitions for conformity between the wheel and high-rail
profile at an L/V of approximately 0.6, for various new and
worn rail profile combinations. One and two-point contact
conditions are shown. Be it a single or two-point contact
scenario, a contact is closely conformal if the gap d or s between
the un-deformed wheel and rail is approximately 0.1 mm
(0.004 inch) or less. Upon loading, elastic deformation of the
wheel and rail will cause that gap to be closed, resulting in a
wide contact ellipse that spans an appreciable portion of the
wear band, e.g. 1.0 to 1.5 inches. A larger gap, up to 0.4 mm
(0.015 inch), provides a contact that is still conformal but only
becomes closely conformal after appreciable wear or plastic
flow. For values of d or s exceeding 0.4 mm (0.015 inch), the
contact is considered non-conformal, since the profiles are now
fully separated and do not take advantage of the reduced
o 5-80 x
contact stresses available by employing more conformal
geometries. While some apply the term “conformal contact” to
both the one and two point conformal contacts, the NRC has
found that the difference in stress, creepage and overall
performance warrants the “complication” of retaining the
more precise definition.
Figure 5.36: Conformity between the Wheel And High Rail
Profile at an L/V of Approximately 0.6 (Slight Rail Rotation)
5.4.1.5.2 Rail Stresses and Pummeling
Contact fatigue of the rail surface is the result of excessive
contact stress and creepage. Both contact stress and creepage
are governed by the wheel/rail contact geometry, which in turn
depends not only on the initial, unworn geometry of each
component, but also the changes in geometry that occur with
wear, fatigue and plastic flow. As an example, wheel/rail
creepage tends to pull metal from the high-rail shoulder into
the gauge-corner, filling in any initial relief. The high rail then
takes the shape of the average wheel. The 50% of wheel
population that has more flange-root metal than average will
steer well through the curve but apply a very high normal
stress to the gauge corner, contributing to subsurface fatigue.
These are the 1-point conformal and non-conformal contacts.
The other half of the wheel population - the 2 point contacts continues to steer poorly and plastically deform and wear the
surface.
o 5-81 x
Modeling software has proven valuable for the design of
optimal rail profiles that in practice must deal with a
distribution of wheel profiles – from unworn to very worn,
new to hollow, wide flange and thin flange. The NRC for
example has developed a Profile Optimization (Pummeling)
Model that applies measured worn wheel profiles to a truck
characteristic of that fleet and derives distributions of contact
stress, fatigue damage, stability and curving performance.
Through an iterative process, the model is used to engineer rail
profiles that optimize the wheel/rail interaction in tangent and
curved track.
The accumulated normal contact stress between wheel and
rail for the population of wheel profiles analyzed may be
plotted in a Pummeling Diagram.
Figure 5.37 shows photographs and pummeling diagrams
for typical preventive and corrective rail profiles. The rail
surface cracks are highlighted with the use of dye penetrant.
Figure 5.37a and b are for a 6.5 degree curve, with a track
gauge of 12 mm (0.47") wide, maintained at preventive
intervals of 13.6 mgt (15 mgt) to NRC templates H2 / L2. The
rail surface condition shows visible but very shallow cracks.
These cracks can be removed and the profile restored to the
NRC template in one pass at a speed of 12.8-kph using a highproduction rail grinder.
Figure 5.37c and d are of a 6.5 degree curve, with a track
gauge of 22 mm (0.87") wide, maintained at corrective intervals
of 54.4 mgt (60 mgt) to NRC templates H4 / L2. The rail
surface cracks on the high-rail gauge-corner and low-rail fieldside are very deep. These cracks are caused by high contact
stresses from a large percentage of wheels contacting the highrail gauge-corner and false-flange contact on the field side of
the low-rail. This surface condition requires three to eight
passes on the high-rail and five to nine passes on the low at 10
kph will restore profile and remove the cracks.
o 5-82 x
A
B
C
D
Figure 5.37: Rail Surface Fatigue and Pummeling
Diagrams on Preventive (a and b) and Corrective (c and d)
Ground Rail in the Lakeside Test Site
o 5-83 x
5.4.1.5.3 Tangent track
Most railways conscientiously apply a central 200-mm radius
running band to all tangent rail. As geometry cars, ties and
fasteners improve, this means that more and more wheels are
running continuously at the same contact band on the wheel
tread, a practice that promotes wheel hollowing. The NRC
developed new templates that provide two distinct running
bands, separated by about 12 mm – one biased towards the
gauge (TG) and the other biased towards the field (TF). Both
profiles were designed to avoid excessive reshaping of the rail
by grinding, minimize contact stress, minimize surface damage,
improve curving and minimize the potential for hunting.
Figure 5.38 illustrates the impact of the new profiles on the
distribution of contact – together the profiles broaden the
pattern of wear on the wheel tread, reducing both the number
of hollow wheels that develop and the rate at which they
hollow. The benefits of this profile strategy will be increased
rail life in curves and tangent track, reduced grinding effort,
lower lateral track forces (through better steering overall),
increased wheel life and reduced fuel consumption.
Figure 5.38: The Distribution o Contact Across the Wheel Tread
when the TG and TF Rail Profiles are Applied, Compared with that
for the NRC-TT Alone
5.4.1.5.4 High Rail Profiles
Optimal high rail profiles must avoid concentrations of stress
and fatigue but also maximize the vehicle curving performance.
Pummeling analysis showed that the NRC high rail templates
exhibited excessive contact-stress and poor curving when
mated with a particular group of measured worn wheels.
Improved profiles were developed for the high rail of mild
o 5-84 x
(HM) and sharp curves (HS). Figure 5.39 illustrates the
substantial reduction in expected surface fatigue damage for
the new profiles compared with the current NRC templates.
The amount of metal that must be removed to re-profile the
rail is reduced by 0.38 mm and 0.63 mm for mild and sharp
curves respectively. As well, the steering performance of
vehicles was substantially improved.
(a)
(b)
Figure 5.39: Plot of (A) Surface Fatigue Distribution for the
NRC-H1 and HM in Mild Curves and (B) the Distribution of
Expected Internal Fatigue Damage for NRC-H2 and HS Profiles
in Sharp Curves
5.4.1.5.5 Low Rail Profiles
The presence of hollow wheels on heavy haul railroads has
dictated heavy field side grinding in the past. Wide gauge
further exacerbates that problem. Under preventive grinding
and with modern steels, less plastic flow takes place and less
metal needs to be removed. A pummeling analysis was carried
out on the three NRC low rail templates, and an improved
design, called the L10, was eventually developed. Instead of
continually relieving the field side to avoid low-rail/false-flange
contact, the contact stress was instead reduced by increasing
the head radius to 250-mm from the usual 200-mm. The final
design reduced the surface damage index by 30% for mild
curves and 23% for sharp curves. The steering moment was
little affected by the low rail design. The L10 reduces metal
removal by 75% compared to the NRC-L2 (Figure 5.40) which
translates into improved rail grinding production and increased
o 5-85 x
rail life. The L10 will be applied to all low rails of mild and
sharp curves, rather than the three profiles currently used.
Figure 5.40: A Comparison of the New L10 and Previous NRCL2 Rail Templates
5.4.1.6 Lubrication and Grinding
Rail surface fatigue cracks grow fastest when contaminated by
water and somewhat slower when contaminated with a mixture
of water and lubricant. On the other hand, lubrication
substantially reduces the tractive stress at the wheel/rail surface
and therefore reduces the number of contact cycles that
contribute to fatigue. For this reason, preventive rail grinding
(where surface cracks are eliminated) in combination with
lubrication can significantly increase rail life. Conversely, the
application of lubricants to damaged rail can increase the rate
of crack growth. Since lubrication also minimizes gauge-face
wear of the high-rail and controls lateral forces in the curve, an
effective lubrication program is essential to a successful
preventive grinding program as well as to maximize rail life.
5.4.1.7 Optimal Wear Rate
The optimal wear rate is the rate of wear required to just
control rail surface fatigue. Insufficient wear results in rail
fatigue, while excessive wear reduces rail life. The optimal wear
rate will vary across a railway property and depend on
differences in tonnage and axle load, type of traffic, rail
metallurgy, track curvature, environment / season, track gauge,
lubrication standards, etc. Under preventive grinding, the
grinder removes only a thin skin of fatigued and micro-cracked
metal from the rail surface, artificially controlling the wear rate
but leaving behind a healthy work-hardened layer.
o 5-86 x
One process for determining the optimal wear rate starts
first by analyzing service worn rail samples to determine the
fatigue crack growth rates and direction of propagation. For
example, a micrograph from a sharp curvature (6º), high-rail
(Figure 5.41) shows a gauge corner free of fatigue and only
short, approx. 0.35-mm, perpendicular cracks on the ball of the
rail. This demonstrated that the current grinding interval and
metal removal rate on sharp curves was adequate to control
fatigue. The same study found that at 27 mgt (30 mgt), surface
fatigue was more severe - 1 mm (0.040 in) oblique cracks
populated the rail.
As examples, the optimal metal removal rate for sharp
curve high rails at a 13.6 mgt (15 mgt) grinding interval was
found to be 0.1 mm (0.004 in) on the center ball area and 0.25
mm (0.010 in) on the gauge corner. On another railway with 23
mgt (25 mgt) cycles grinding intervals, it was determined that
0.18 mm (0.007 in) should be removed from the gauge
shoulder and the gauge corner. These metal removal depths
were adopted by their respective railroads as minimum metal
removal targets for all curves on preventive cycles.
Figure 5.41: Distribution of Surface Cracks in a Sharp-Curve
High-Rail under Preventive Grinding
o 5-87 x
Grinding Patterns
Implementation of the optimal metal removal rate requires
accurate knowledge of the metal removal for each grinding
pattern at various grinding speeds. Today’s high production
equipment regularly grind track in the preventive mode at
speeds ranging from 9.6 to 22.4 kph (6 to 14 mph). Data
collected is utilized in a table for each pattern giving the
maximum grinding speed to ensure the optimal metal removal
rate. These maximum grinding speeds are carried out on all
grinding territories.
These patterns were fine-tuned when created to match the
existing rail condition to the new NRC BAR Gauge profiles.
As grinding machine configurations changed over the years,
the patterns were automatically mapped to the new equipment
configurations, introducing minor variations from the original
pattern at each iteration. The typical rail shape also changed
over time due to changing traffic, loads and wheel profiles.
The two factors combined to result in patterns which often
exhibited ridges at various locations on the rail surface, and
were not well suited to meeting the rail profiles desired.
A redesign of the grinding patterns is necessary to improve
the efficiency of the preventive grinding strategy. Rail and
profile specific patterns concentrate the metal removal where it
is needed most to address profile and rail surface conditions
without wasted metal removal on areas of the rail which don’t
need it. Improved patterns also reduce crack growth rates
through closer conformance to the desired profile and better
geometric smoothness.
5.4.1.8
Rail Grinding Strategies
5.4.1.8.1 Preventive Rail Grinding
The current trend in North America is to continue to push
production rail grinding speeds as high as possible while
maintaining control over rail fatigue. The advantages are
reduced grinding costs and reduced loss of rail metal. Railways
such as the Burlington Northern, Canadian Pacific Rail and
Canadian National now have many kilometres of rail where
they have achieved the desired rail profiled through grinding.
o 5-88 x
Grinding is usually completed in a single pass at grinding
intervals of between 7 and 22 million gross ton.
5.4.1.8.2 Preventive vs. Corrective Rail Grinding
As demonstrated on the BNSF Canadian Pacific and CN and
many other railroads, even the best premium rail cannot
prevent surface fatigue from developing in the outermost layer
of the rail steel under today's traffic and axle loads. The
growth of surface (and subsurface) fatigue cracks is governed
by the contact stress and slip. From studies conducted by the
NRC, micro-cracks develop at the most stressed portion of the
rail surface within 4.5 to 7.3 mgt (5 to 8 mgt). In their early
stage, the microscopic cracks grow very slowly. Since cracks
grow faster as they get longer, their growth rate accelerates
with time. The preventive grinding strategy is designed to
address the damaged surface of the rail before the micro-cracks
enter their stage of rapid growth. By completely removing all
short cracks, the preventive mode takes advantage of the crack
initiation phase and period of slow growth. Removing the thin
skin of the rail surface that contains the micro-cracks can be
accomplished with a single, high-speed pass of the grinder. At
the same time, the "optimal" profile is maintained on the rail
and a good, protective layer of work-hardened material
retained. Under preventive grinding, the rail surface is
maintained to control contact stress and promote wheelset
steering, while at the same time retaining resistance to crack
initiation and growth by virtue of its work hardened layer.
Corrective grinding results in the rail being subjected to
higher contact stresses for longer intervals. Even the toughest
premium rail cannot withstand this assault. Corrective
grinding therefore must apply many passes at low speed to
address very deep cracks. This heavy metal removal from the
rail strips away the work hardened layer, while at the same time
usually fails to eliminate the deepest cracks. Corrective
grinding is thus associated with larger overall metal removal
rates, and therefore contributes to shorter rail life. In addition,
the failure to regularly address the profile results in greater
lateral forces to the track structure and trucks, leading to
excessive strain on fastening and truck components. The
potential for truck hunting also increases considerably. Failure
o 5-89 x
to regularly address welds and other surface irregularities
contributes to ballast and tie deterioration.
Table 5.12: Summary of Differences between Preventive and
Corrective Rail Grinding Strategies
Preventive Grinding
Corrective Grinding
Grinding frequency
Cumulative Traffic (mgt)
Cumulative Traffic (mgt)
Sharp curves
7.3 to 18.2 (8 to 20)
36.4 to 72.7 (40 to 80)
Mild curves
14.5 to 36.4 (16 to 40)
54.5 to 109.1 (60 to 120)
Tangent track
21.8 to 54.5 (24 to 60)
72.2 to 181.8 (80 to 200)
Grinding speed
9.6 to 19.2 km/h (6 to12 mph)
4 to 9.6 km/h (2.5 to 6 mph)
Grinding passes
1
3 to 9
Characteristics
Grinding interval depends on
curvature
Interval depends on traffic levels
(mgt)
grind even if there are no visible
surface defects
all surface cracks removed
usually out-of-face
crack initiation period available
Existing cracks start to propagate immediately
work hardened layer removed by many grinding
passes
profile deteriorates within
about 18.2 mgt (20 mgt).
work hardened layer retained
Optimal profile always maintained
(lower contact stresses, better
stability in tangent track and
steering through curves)
welds addressed regularly
5.4.1.9
usually time based (e.g.
annual grinding)
grind rail with visible/severe
deterioration
Deepest cracks not removed
welds addressed infrequently. Weld dipping leads to
fastening, tie and ballast
deterioration.
Transitioning from Corrective to Preventive
Grinding
Traditional methods of implementing preventive grinding
programs required significant short-term increases in grinding
resources to first restore all of the rail to a good profile and
clean surface condition before preventive cycles could be
implemented. But many railroads that recognize the benefits of
preventive grinding are simply unable to acquire sufficient
additional budget for the initial "clean-up" phase. The
“preventive-gradual” grinding strategy was devised by the NRC
to address this scenario.
o 5-90 x
5.4.1.9.1 Preventive-Gradual Grinding
The preventive-gradual grinding strategy involves embarking
straight onto preventive grinding intervals from the current
corrective scheme without first undertaking the expensive task
of “cleaning” all the rail. The rail is transitioned to the desired
profile and crack-free state on a gradual basis. This strategy
starts with frequent one-pass grinding as with traditional
preventive grinding, but with additional metal removal each
pass – a method that is only feasible with today’s modern highperformance grinding equipment. The objective is to
immediately gain the benefits of an optimized preventive
grinding strategy while gradually catching up to the profile and
surface cracks.
Figure 5.42 shows the staged profiling and crack removal
process. The desired rail profile is achieved in Stage 1 of the
strategy within one to three passes. Stage 2 includes the next
one to three cycles, which gradually stop the initiation of new
cracks. The final stage consists of a further one to three cycles
to remove the remaining inactive cracks to produce a clean rail
surface. The entire process typically takes three passes on
tangent and shallow curves, and up to nine passes on sharp
curve low-rails.
Figure 5.42: Staged Crack Removal with the PreventiveGradual Strategy
o 5-91 x
Two essential components of the preventive-gradual
strategy are effective lubrication and proper track gauge.
Lubrication significantly reduces lateral forces in a curve,
essential to maintaining contact stresses on the rail at
manageable levels. Wide gauge in curves causes the false flange
on hollow wheels to contact the running area of the low rail,
resulting in very high contact stresses and poor wheelset
steering.
This strategy was first implemented by the BNSF on its
Pacific Northwest (PNW) corridor in February 1998.16 This
territory consists of 8,300 track-km (5160 track miles) with
annual tonnage over the core routes of 27 mgt (30 mgt) to 81
mgt (90 mgt). A significant proportion of the track consists of
sharp curves on concrete ties with heavy mountain grades.
Rail in sharp curves is predominantly 136RE deep headhardened premium rail. One 88 stone rail grinder was assigned
to this corridor. The PNW was selected because it was the
most demanding of BNSF’s four grinding territories and most
likely to rapidly demonstrate the success or failure of the
preventive-gradual strategy.
Grinding cycle intervals were established of 13.5 mgt (15
mgt) on sharp curves 3.5 degrees or greater, 26 mgt (30 mgt)
on mild curves and 41 mgt (45 mgt) on tangent track.
The preventive-gradual grinding strategy had begun
demonstrating significant benefits by the end of the first year.
Visible rail surface defects had decreased, 98% of the rail was
at the desired profile, and test site measurements verified that
rail wear and grinding costs were reduced compared to other
grinding strategies.
5.4.1.9.2 Results
Through adoption of the preventive-gradual grinding strategy
BNSF has achieved significant productivity gains in its rail
grinding program. Comparing 2000 performance to 1997,
grinder utilization improved by 31%, grinding passes per curve
per year decreased from 3.9 to 2.4, and the mean cycle interval
declined from 56.2 to 24.5 mgt (62 to 27 mgt). These gains
directly translated into the ability to cover more territory with
the same amount of grinding resources, at a lower cost.
o 5-92 x
5.4.1.9.3 Rail Wear
A 8 km (5 mile) long test area was established on the PNW
corridor to measure the effects of different grinding strategies
on rail wear. The test site consisted of 10 curves between
5°51’ and 6°31’ curvature, and two mild curves. Train speeds
averaged 48 kph (30 mph) at under balanced speed, on
concrete tie track. Annual tonnage during the test varied
between 55.3 and 60.7 mgt (61 and 67 mgt).
The following grinding strategies were applied to specific
curves in the test site to determine the relative merits of each
approach:
•
No Grind – Correctively ground prior to the
beginning of the test to remove all visible surface
defects, then left unground for the duration of the test.
•
Maintenance – Correctively ground at 27 mgt (30
mgt) intervals
•
Corrective – Correctively ground at 54.5 mgt (60 mgt)
intervals
•
Preventive-Gradual – 1 pass preventive-gradual
intervals of 13.5 mgt (15 mgt)
•
Preventive-Immediate – Correctively ground at the
start of the test, then 1 pass preventive grind at
intervals of 13.5 mgt (15 mgt)
Rail surface conditions were extremely poor at the
beginning of the test. All curves except the preventive-gradual
received 3-5 passes on the high-rails and 5-9 passes on the low
rails to remove all visible surface defects and cracks.
Rail wear test results after the first year of the PNW
preventive-gradual initiative were presented in.16 The test sites
continued to be monitored over the second year of the
program until 113 mgt (125 mgt) of traffic and 8 grinding
cycles of 13.5 mgt (15 mgt) intervals had been completed.
The second year rail wear data is shown in Figure 5.43.
Although the no-grind scenario exhibited low wear, the
development of severe spalling and corrugation on the low rail,
o 5-93 x
and heavy checking and shell development on the high rail
precluded this approach as a viable option.
The effects of wide gauge on low rail wear can been seen
in the second year preventive immediate results. False flange
contact on the center of the low rail caused rapid flattening,
fatigue crack formation, and accelerated rail wear.
Figure 5.43: Grinding Strategies in the Test Site, showing Total
Wear from Grinding and Traffic after 113.6 mgt (125 mgt)
(bottom segment of each bar shows results
at end of first year)
The test results show that the preventive-gradual method
is the most effective strategy for minimizing total rail wear. As
a result of the preventive-gradual strategy and improved rail
lubrication practices, BNSF’s 2000 curve rail relay program was
44% lower than 1997 levels.
5.4.1.9.4 Rail Surface Condition
In the test site the preventive-gradual and preventiveimmediate curves demonstrated the best and most consistent
rail surface condition, while corrective and no-grind yielded the
worst, with significant checking, shells, spalls and corrugation.
Rail surface condition on BNSF has improved dramatically
on its preventive-gradual territories. Premature rail relay
because of rail surface condition in 2000 was 53% lower.
Additionally main track rail detection exceptions, where poor
rail surface condition prevents ultrasonic rail flaw inspection,
have decreased from 238 locations in 1998 to 5 in 2000.
o 5-94 x
5.4.1.9.5 Detail Fracture Rates
Tangent detail fractures in 2000 declined 16% from 1999
levels, the first significant reduction in 8 years. Detail fractures
on curves were at a 10 year low. BNSF believes these
reductions are a direct result of the grinding optimization
methods instituted at the end of 1999.
5.4.1.10 Advance Planning to Increase Grinding
Production
A key to a productive grind is to have a regular grinding cycle
and a grinding production plan. The grinding production plan
should be available in advance of the arrival of the grinder on
site. The plan should be based upon an advance survey of the
territory identifying the types of conditions that are being
targeted for correction. Surveys should use a grinding template
and a straight edge with taper gauge. The Track Manager
should accompany to discuss his requirements.
The grinding plan should specify the patterns to be used,
the number of passes and the grinding speed. It is
recommended that they be done on a track profile chart to
relate the curve locations to such features as signals and
crossings. The objective is to develop a production approach
to grinding that enables the grinding operator to set up
patterns in advance and to use grinding speed to advantage to
cut only as deep as is required. The railway’s grinding
supervisor is therefore freed to concentrate on required
deviations from the plan and on quality control. Grinding
Production Plans make sure that all involved know what the
plan is. In fact, they are often used to inform the Train
Dispatcher of the track occupancy requirements of the
grinding operation. With this knowledge, he may be able to
work out improved track blocks.
Figure 5.44 is a sample grinding plan used by Canadian
Pacific Railway. These plans are done on a PC and are
displayed on computer onboard the grinder. The program
used also estimate elapsed times for use in work block
planning.
Some railway authorities are now contracting for a
completed track mile and paying the contractor on basis of
o 5-95 x
dollars per kilometre ground to within specifications. The
contractor therefore both surveys rail conditions and performs
the grind. In some cases, (Mitchell et al., 1989),17 the
contractor also performs his own post grind measurements.
This can be an effective way to operate as it encourages the
contractor to plan and innovate and can improve the railway’s
control of the grinding expense. At the same time, it requires
that the railway have a good understanding of rail conditions
and well written contract.
5.4.1.11 Maintaining Quality Control
As discussed before, good grinding performance requires that
each grinding motor is operating at the correct pressure and
angle. Maintaining a regular inspection of the grinding
operation can yield premiums in grinding effectiveness. The
following checks should be performed by a railway supervisor
as least once per track occupancy of two hours or more.
5.4.1.11.1 Grinding Power
Check the grinder control panel to see that at least 9 out of 10
grinding motors are running at their rated capacity. This is
typically 80 percent of rated capacity for motors grinding on
the top surface of the rail and 65 percent of rated capacity for
motors grinding sharper angles on the gauge and field side.
5.4.1.11.2 Ground Rail Profile
The grinding supervisor should spend as much time as possible
walking behind the rail grinder to check that his plan has
produced an effective treatment of the rail. Where an
additional pass is required he may communicate this need by
radio to the grinder operator.
o 5-96 x
Figure 5.44: PC-Based Preparation of Rail
Grinding Production Plan
o 5-97 x
While inspecting the ground rail, he should look for signs
of grinding wheel malfunction. The following are possible
problems:
•
•
•
•
•
•
There are gaps in the rail surface that are not being
ground. This indicates some grinding wheels are not
working.
There are sharp ridges left after grinding. This indicates
that some stones are not working or grinding angle
positions are not correct.
Some grinding facets have grinding marks diagonal to the
rail or in line with the rail’s longitudinal axis. This
indicates that an inoperative stone is being dragged or that
it is not positioned laterally and is grinding outside of its
inside diameter.
There are extensive black marks on the rail. This indicates
overheating of the rail. Consistent brown or black marks
may indicate excessive downward pressure on the motor.
The rail surface finish consistently shows deeper scratch
marks that are still in evidence on the contact band within
1 – 2 weeks of the grind.
The rail surface shows evidence of gauges due to a
grinding carriage miss-adjustment, causing contact of the
corner of the grinding stone with the rail.
5.4.1.11.3 Longitudinal Rail Profile
The removal of corrugations can be checked with a straight
edge and taper gauge. A minimum length of 600 mm is
recommended. Each railway’s standards on the extent of
corrugation removal may differ. Corrugations that exceed 0.25
mm (0.010”) are known to re-grow fairly rapidly.
5.4.1.11.4 Transverse Rail Profile
Transverse rail profile should be measured with a template
gauge mounted on a bar that will sit on the plane of the top of
the two rails. This is important to correct for rail cant. The
templates should cover portions of the gauge corners of the
high and low leg to ensure correct shaping of these critical
locations by higher angle stones. A taper gauge or gap gauge
can be used to judge the extent of the fit of the as-ground rail
o 5-98 x
profile to the template. As considerable grinding time can be
spent in achieving a perfect fit, a tolerance level is
recommended.
Use of the template gauge should correctly locate the
contact band, except in locations where there is significant loss
of rail cant under loading.
This can be verified by spray painting the rail and
observing the width and location of the contact band as well as
whether one or two-point contact has been achieved.
5.4.1.11.5 Metal Removal
There are several methods for periodically verifying the depth
of the cut achieved after one or more grinding passes. Some
railways apply indentations to the rail with a hole punch. The
depths of these are measured before and after grinding using a
Starret Dial Gauge. Dial Gauges, displacement transducers
and lasers are also used to measure the distance of the rail
surface to a measuring frame that is affixed to the rail.
For regular verification of grinder performance, the use of
a template gauge on a bar is probably sufficient. An
experienced supervisor can see by the change in the gaps as
measured with his taper gauge whether or not the metal
removal is within specification and properly distributed across
the rail. The careful application of the measuring equipment
discussed in Section 5.3 may also be used to measure metal
removal.
5.4.2 Rail Planing
When a rail profile is allowed to wear or deform due to plastic
deformation to such an extent that large volumes of material
must be removed to return the correct profile, grinding may
prove uneconomical. In such a case a rail planing machine can
be used to correct the rail profile. Such a situation developed
on the South African Iron Ore export line and the Plasser &
Theurer SBM 140 Rail Planer was used to correct the rail
profiles. The SBM 140 is capable of removing large volumes of
material and restoring the design rail profile rail within the
required tolerances.18
o 5-99 x
5.4.2.1
Description of the SBM 140 Rail Planing
Machine
The SBM 140 rail planing machine is made up of two sections,
the four-axle planing machine and the single-axle swarf
collector. Weighing approximately 80 ton, the SBM 140 is
driven on all four axles to give the machine the necessary
thrust force required to remove the metal from the rail crown.
The metal shavings produced during the planing work are
removed from the track by two chain type electro magnets on
drums on either side of this trailing swarf collector. They are
approximately 95% effective provided the cribs are properly
filled with ballast.
Two bi-directional planing units are positioned between
the bogies and are supported on the rail by a number of tool
guide rollers that are hydraulically pressed onto the rail with a
force of 6 ton. The tool guide rollers will prevent the tools
from following the surface profile of the rail where short
defects occur thus ensuring the removal of corrugations and a
fine planed surface on the rail.
Each planing unit has a tool holder for securing the tool
inserts. Four different tool inserts are being used. Some of
these tungsten carbide tool inserts are standard commercial
tool bits whereas others where specially developed for the
planing machine. These tool inserts are used in different
configurations depending on the tool holder used. Tool
holders can be changed within seconds due to their dovetail
design.
Accurate transverse positioning of the tools on the rail
surface is achieved using datum rollers on the field side of the
rails. Accuracy of ±1 mm from the design transverse position
was consistently achieved. To produce the rail profile required
the blending in of the various cuts is achieved by an onboard
control system that enables the tools to be adjusted in
increments of 0.1 mm.
Figure 5.45 shows the general arrangement of the planer,
as well as the planing unit.
o 5-100 x
Figure 5.45: SBM 140 Planing Machine
5.4.3 The Planing Process
Figure 5.46 shows a typical worn profile as well as the required
profile after planing.
Figure 5.46: Typical Worn Profile and Restored Profile
o 5-101 x
Figure 5.47a: Planer Cut Sequence
o 5-102 x
Figure 5.47b: Planer Cut Sequence (continued)
5.5 Wheelset Failure Risk Management and
Maintenance
Wheels are not managed independently but as wheelsets and
because of the large number of wheelsets present in heavy haul
trains, the reliability of these components have to be extremely
high. The failure of one wheel or bearing could lead to a
o 5-103 x
derailment costing the railway millions of dollars in damage of
infrastructure and rolling stock as well as loss in production.
FRA Reported Safety Statistics for 19991 shows 3% of the
accidents were wheel related and 2% were axle and journal
bearing related. Figure 5.48 shows the distribution of accidents
reported for wheel defects per type for the last 5 years.
Figure 5.48: Distribution of Accidents per Wheel Type
The wheelset failure risk management and maintenance
discussion below is given from a Spoornet perspective with
specific reference to the practices on the COALlink coal
export line. The required reliability for wheels is as high. To
achieve high reliability levels the wheelset is seen as a subsystem of the total train. The subsystem reliability depends on
the reliability of the wheelset as well as the reliability of the
condition monitoring system, as depicted in Figure 5.49. In
simple terms the reliability of the wheelset sub-system can be
expressed as:
R wheelset sub-system = 1 – [(1 – R wheelsets) x (1 – R Condition monitoring
system)]
o 5-104 x
For example,
if the reliability,
R wheelsets = 0.999 995
and the reliability,
R Condition monitoring system = 0.9,
then,
R Wheelset sub-system = 0.999 999 5.
Improvement of the wheelset subsystem can thus be
substantially enhanced by the reliability of the wheelset
condition monitoring system.
Figure 5.49: Wheelset Sub-System Reliability Model
5.5.1 Wheelset Reliability (Spoornet)
5.5.1.1 New Components
High wheelset reliability depends on the reliability of the
wheels, axles and journal roller bearings. From a wheel point of
view reliability is ensured through meticulous wheel
specifications, covering all aspects of the wheel from design
stress levels, to effects of heat input. Material properties and
allowable flaw sizes in different parts of the wheel are specified
to ensure high reliability and wheel tread wear resistance.
The wheel specification includes non-destructive testing
requirements to ensure acceptable wheels. Destructive tests
and material property tests are performed on a regular basis to
ensure conformance of new wheels to the specification.
Each wheel manufactured is tested, ultrasonically, in the
rim area to ensure that no flaws larger than 1 mm in diameter
is present in the highest stress areas. Further magnetic particle
inspection is done on the whole wheel to ensure that no
surface cracks are present in the wheel. Finally the wheel bore
is visually inspected during final machining when the wheel is
fitted to the axle. Only limited flaws are allowed in the wheel
bore.
o 5-105 x
Similar specifications and test methods are used to ensure
high reliability of new axles.
As far as roller bearings are concerned, there are a number
of factors that determine the roller bearing life and reliability.
These factors range from axle load, wheel and track condition,
and environmental factors through to seal type and lubrication
used. Because of the complexity of these factors replacement
cycles for roller bearings are usually refined as service
performance data becomes available. The choice of new versus
remanufactured bearings is also influenced by wheel wear rates,
costs and required reliability.
Typical bearing cycles for COALlink are:
•
1 million km for class F bearings on 26-ton axle load
wagons.
•
700 000 km for class D bearings on 20-ton axle load
wagons.
These bearings operate on short turnaround cycles
between high altitude, dry environment to low altitude, marine
environment. Moisture ingress and corrosion related damage
have a major influence on the performance of these bearings.
5.5.1.2 Used Components
Wheels are normally only replaced when their minimum
diameter is reached. Standard minimum diameters as per AAR
recommendations are used.
Axles are recycled. When wheels reach their minimum
diameter, they are removed from the axle and the axle is
reused. In addition to the normal ultrasonic testing of axles,
axles being recycled are also tested by means of magnetic
particles during the recycling operation.
Bearings are removed at the end of their cycle and
returned to bearing manufacturers for re-manufacturing to
Spoornet’s own bearing specification.
o 5-106 x
5.5.1.3 Condition Monitoring
5.5.1.3.1 Wayside Condition Monitoring (see Par 5.6.2)
Realizing the importance of a reliable condition monitoring
system, COALlink has been investing in the development of
an Integrated Condition Monitoring System (ITCMS) over the
last 3 to 4 years. A schematic of the COALlink ITCMS is
shown in Figure 5.50. Wayside measuring points are connected
to a computer network along the track, either directly or
through modems. An automatic vehicle identification system
reads wagon numbers from ID tags. The measured data,
together with the wagon numbers are stored in a database. In
addition measured values are compared to preset warning and
train stop limits. Any critical alarm is relayed to Centralised
Traffic Control from where a command to stop a train is
issued. The information in the database is used to analyse
wagon condition and for research purposes. An example of
such an analysis is to identify performance trends and
components wear rates. The information can also be used to
learn more about inter relationships that exist between
different measurements on the same wagon. For example
bogie couple forces can be related to wheel diameter
differences and or load distribution on a wagon.
Figure 5.50: Schematic of the Spoornet COALlink ITCMS
Measuring points operational on the ITCMS are:
•
Hot bearing detectors; These systems are seen as short
term warning systems as a bearing normally overheats
very rapidly only a short distance before final failure.
•
Acoustic defective bearing detection system (system
under development); This system is seen to be an early
o 5-107 x
warning system as bearings emit sounds indicating
eminent failure for long distances before final failure.
•
Brake efficiency measurement – wheel temperature;
These measurements are done in a location on the
track where brakes are normally applied. Wheel
temperatures give an indication of the efficiency of the
braking system throughout the train. Too high wheel
temperatures may indicate sticking brake systems,
hand brakes on or leaking valves. Low wheel
temperatures may indicate brakes not working for any
of a number of reasons.
•
Brake problem measurement – wheel temperature;
These measurement are done in a location on the track
where brakes are normally not applied. Sticking brake
systems or hand brakes that were left on will in this
case cause high wheel temperatures.
•
Wheelset lateral force and bogie couple forces; The
purpose of this measuring system is to identify
wheelsets on wagons that are exerting high lateral
forces or couple forces to the track. It is used to
identify bogies (trucks) that require maintenance.
•
Weighing In Motion – Wheel Impact Measurement
(WIM-WIM); This system is used to identify wagons
with load distributions out of allowable limits. Further
wheel impacts are measured to identify skidded wheels
that require attention.
•
A measuring point for the measurement of wheel
profile and wheel diameter is currently being planned.
5.5.1.3.2 Run-in inspections
The wayside condition monitoring system is currently still
being enhanced by run-in inspections. These inspections are
visual and audio inspections of wagon condition at the arrival
yard. Major success has been achieved with these inspections
to identify, among others, wheel and roller bearing problems.
At this stage the run-in inspections still enhances the total
condition monitoring system reliability, but it is foreseen that
the wayside system will in future reduce the need for run-in
inspections on wheelsets.
o 5-108 x
5.5.1.3.3 Four Monthly Maintenance Depot
Inspections
Every wagon in the COALlink fleet is inspected, every four
months, at a maintenance depot. During this inspection the
condition of wheels and bearings are visually inspected. Wheel
treads are examined for the presence of thermal cracks on the
tread. If these cracks are present, the depth of cracks is
measured. If crack depths exceed a certain limit, the wheels are
removed for reprofiling.
During these inspection opportunities, components such
as brake shoes are replaced.
When the need arises for special test programs, these tests
are also performed at the maintenance depot during the
inspection cycle. This normally happens when a particular
high-risk problem has been identified on a certain group of
wheels.
5.5.1.3.4 Workshop Maintenance
During workshop maintenance of wheelsets wheels and axles
are tested ultrasonically. Tests on the axles are concentrated on
high stress areas and areas where fretting may occur. Attention
is also given, during this inspection, to areas where service
damage may occur, such as damage caused by dragging brake
rods.
On-wheels-tests are concentrated on the tread area. A
surface wave ultrasonic test is done on reprofiled wheels to
ensure that all cracks have been removed from the wheel tread
area.
The wheelset failure risk management approach described
here is very successful in improving wheelset reliability and
wheelset problems are currently at a low level.
5.5.1.4 Wheel Profile Monitoring
Wagons are inspected with every round trip in the
arrival/departure yards (run-in inspections), during this inspection
the wheels are inspected. If found to be close to the limits the
wheels are gauged with a go/no-go gauge and those that fail are
sent to a maintenance depot where the wheelsets are changed-out
and reprofiled. Typical wheel life cycles are as follows:
o 5-109 x
1. Number of cuts before 20 ton/axle wheels are scrapped
New diameter
= 863 mm (new)
Minimum diameter
= 812 mm (min)
Available for cuts (863-812)/2
= 25.5 mm
A typical cut for 2 mm tread wear will result in a 5 mm
diameter loss (2 mm for wear and 3 mm undercut). This will
result in 5 cuts or 6 cycles in a wheel life.
2. Number of cuts before 26 ton/axle wheels are scrapped
New diameter
= 915 mm (new)
Minimum diameter
= 870 mm (min)
Available for cuts (915-870)/2
= 22.5 mm
A typical cut for 2 mm tread wear will result in a 5 mm
diameter loss (2 mm for wear and 3 mm undercut). This will
result in 4 cuts or 5 cycles in a wheel life.
On OREX Spoornet’s Iron Ore export line a monitoring
system that measure the wheel flange height on a continuos
basis was installed. These measurements give Spoornet the
wheel wear rate for the fleet which is used for wheel
maintenance planning purposes. The measured wear rate is
confirmed on a smaller sample of wagons with "MINIPROF"
measurements. As the wheels on this line only experience tread
wear and no or little flange wear, there is a direct relationship
between this flange height measurement and tread wear.
Therefore this measurement (flange height) can be used with
confidence to predict tread wear, and maintenance cycles.
5.5.2 Wheelset Maintenance
Reprofiling of 2-wear or multi-wear wheels for thin flange is a
usual practice on heavy haul railways worldwide. Recently,
TTCI has reported on some very interesting work on the
economic costs of permitting wheel tread hollows to develop.
A hollow tread situation develops when tapered wheels are
permitted to wear in the center, particularly where there is
heavy brake shoe contact. Unlike Europe, Australia and South
Africa, there is no standard in North America on how deep a
wheel tread may be permitted to hollow. This is the flipside of
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typical rail grinding standards which call for grinding when the
low rail becomes too flat. TTCI’s survey of 6,500 wheels
found that 6% had more than 3mm of wheel tread hollowing.
They also predicted that wheel with 3 or more mm (> 0.12 in.)
of wheel hollowing had a large effect on train resistance and
hence fuel consumption, and imparted higher forces in tangent
and low curvature track, and truing at this point was a net
positive cost. A large cost of damage to special track work is
also speculated due to the high stress impacts of the wheel’s
“false flange.” Inclusion of wheel hollowing as a wheel
maintenance standard is being pursued by TTCI through the
AAR.
Figure 5.51: Heavy Haul Railways with Wheel Lives
Considerably Longer than the Norm in North America Maintain
Maximum Limits on Wheel Tread Hollowing
The Australians and South Africans have bragging rights
on their average wheel lives, and indeed these are impressive by
North American standards. Spoornet reported that 1 million
miles was the average life for a 2 wear wheel, with many
exceeding 1.5 million miles. This compares to about a
320,000-mile average in North America. These long lives are
achieved through regular turning of wheels on a lathe to
restore the profile, akin to rail profile grinding. Very little
metal needs to be removed at each turning. So, for example,
the South Africans will make 5 cuts of a wheel over its life,
with an obvious benefit in maintaining a good profile. A prime
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deterrent to this practice in North America at present is the
regulation in the AAR Wheel and Axle Manual that new or
reconditioned bearings must be applied at each visit to a shop,
including for wheel reprofiling.
In North America it is typical19 to replace the candidate
wheelset from a freight car when it needs reprofiling, replacing
it with one from the inventory. The wheel shop is often some
distance from the location where the wheelset was removed.
This repair technique makes frequent reprofiling to control
wheel shape costly.
Many heavy haul railways use under floor or above floor
wheel lathes that do not require the wheelset to be removed.
The most recent machines are called “centerless.” They
support some of the weight of the vehicle on the bottom of
the roller bearing housing, and therefore between axle centre
and tool. They are attractive in that they do not require
removal of the bearing end caps. They use contour milling
cutters to re-profile two wheels on one wheelset
simultaneously. Some machines can handle two wheelsets
simultaneously. Underfloor lathes are expensive at $2-3
million but are more productive than wheelset removal, taking
30 to 60 minutes per car.
Again the downside is that the bearings are not necessarily
replaced. On the other hand, there have been recent
improvements in roller bearing reconditioning and mounting
practice to improve bearing performance. Acoustic detectors
would assist in alleviating any concern about bearings being left
on for the life of the wheel.
Above floor lathes maybe another possibility. Here the
foundation is less expensive and the setup for reprofiling is
faster. But whatever the technique that makes most sense for a
particular railway, it is clear that regular reprofiling to address
wheel hollows is essential best practice for heavy haul railways.
5.6 Wheel and Vehicle Interaction Condition Measures
As the demand for railroads to haul heavier loads and faster
trains continues to grow, so does the need to ensure that and
wheelsets and vehicles are in safe operating condition. Unsafe
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operation of rolling stock can lead to increased costs especially
in the Heavy haul environment. By introducing reliable
measuring devices to measure wheel and vehicle component
conditions and interaction performance the risk increased cost
and derailments can be reduced. The most basic application of
the digital wheel and rail profiles is that it allows the rail profile
to be matched against the prevailing wheel profile to examine
the wheel/rail contact characteristics.
5.6.1 Wheel Wear Measurement Techniques
The wheel profile undergoes very large dimensional changes
through its life (Figure 5.52), usually more so than the rail.
Wheel profile measurements quantify the wear in the zones as
specified.
Figure 5.52: Wheel Wear Zones
Techniques to dimensionally measure the wheel profile
condition can be categorized as:
•
Manual mechanical feeler and tracing gauges
•
Manual electronic tracing gauges
• Non contact-optical measurement systems
Presently, wheel inspections are performed by car
inspectors utilizing a variety of hand held gauges in accordance
with Association of American Railroads (AAR) specifications
(Examples as used by Canadian Pacific Rail are shown in
Figure 5.53 and Figure 5.54). However, such gauges are not
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often used. More often a wheel is visually assessed by an
inspector to determine if a gauge measurement should be
made. As a result, complete wheel evaluations are rarely
performed.
Figure 5.53: Example 1 of Manual Mechanical Feeler Gauge
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Figure 5.54: Example 2 of Manual Mechanical Feeler Gauge
A variation of the electronic tracing MINIPROF gauge
exist that measure a wheel profile as shown in Figure 5.55.
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Figure 5.55: MINIPROF Wheel Profile Measurement
When a hand held evaluation is performed, the
measurements generally made are vertical flange (flange
height), flange width, and rim thickness. While shelling,
cracking, flange radius, and other parameters are also
important, these parameters are not generally measured
quantitatively as part of the wheel inspection process. Even
though only a few measurements are routinely performed, the
process is still time consuming, prone to inaccuracies, and
performed inconsistently due to differences in the subjective
visual assessment of the inspector. Some railroads perform
measurements based on the mileage of a car, which is hard to
accurately quantify, while other railroads inspect wheels
erratically due to the inconsistent availability of inspectors.
The railway industry has long sought an economical roll-by
system to measure wheel wear parameters and detect wheel
defects. The development of non-contact optical measuring
principles has only recently been applied in a practical roll-by
wheel wear parameter measuring system. The WheelScan12
system is an example that has been put in service to measure
flange height, flange width, tread width, and rim thickness.
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The WheelScan incorporates an optical/laser-based
technique to determine the extent of the wheel rim wear and
wheel flange wear as a wheel passes over the system, see Figure
5.56. The optical instrumentation consists of a laser-based light
sources and CCD video cameras.
Figure 5.56: WheelScan System Overview
The light source produces a plane of laser light at the exact
time the wheel comes into the field of view of the CCD
camera, see Figure 5.57. This plane of light illuminates a
narrow band on the wheel across the wheel tread and flange.
The CCD cameras capture the images and subsequently
transfer this information to a microprocessor, which calculates
the critical dimensions of the wheel profiles.
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Figure 5.57: WheelScan Laser/Video Camera System
The source of illumination is a laser synchronized with the
passing of the wheel. The laser is utilized to freeze the motion
of the wheel at track speeds and provide a reliable light source
for field applications. Special optics are utilized to allow the
laser light to be spread across the entire wheel width. By
utilizing a combination of custom filters and high-speed CCD
cameras, the system acquires sunlight free images of the wheel
profile.
Once the wheel enters the cameras field of view, the
WheelScan system sends a signal to the video cameras to
acquire two video images of each wheel of a given axle of the
train. The video images are time stamped to allow for the
correlation of train tag information and axle identification.
Each video signal is then transferred in analog form to the
WheelScan host computer. The computer’s internal parallel
processors convert the analog signal to digital information,
Figure 5.58(a).
After the correction of the digital data, the image is
scanned to locate the edge of the rim and the inside of the
wheel flange. These two reference points allow for the
determination of the wheel’s center line, see Figure 5.58(b).
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After locating the wheel’s center line, the WheelScan
system scans for a maximum flange height. The flange height
is calculated by subtracting the vertical component of the
wheels center line from the maximum flange height, see Figure
5.58(c).
C
(a)
(b)
Tread Width
Vertical
Flange Height
3/8”
C
(c)
Flange
Thickness
(e)
C
(d)
C
Rim
Thickness
(f)
Figure 5.58: Wheel Wear Parameters
The WheelScan system computes the tread width by
measuring the lateral distance between the flange and the rim.
The distance measured is referenced from the face of the rim
to a point on the flange which is 3/8" vertically above the
centerline of the tread, see Figure 5.58(d).
The WheelScan system computes the flange thickness by
locating the transition point where the wheel flange interfaces
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with the tread and then subtracting the outside flange, see
Figure 5.58(e).
The WheelScan also computes the rim thickness by
locating the most extreme point on the digital rim data. Using
the vertical component of the wheel’s centerline, it computes
the rim thickness, see Figure 5.58(f).
5.6.2
Wheel and Vehicle Track Interaction –Wayside
Measuring system.
To integrate damage control strategies from different
disciplines in railway engineering a vehicle track interaction
measuring system to evaluate the performance of the vehicles
moving over the track is an important ingredient for the
success of a heavy haul operation. Such a system provides an
essential part of the information to determine if synergy exist
between the vehicle, track and the profitability of the heavy
haul operation.
As an example the Spoornet Integrated Condition
Monitoring System (ITCMS) will be described below. This
system monitors the COALlink fleet as it passes a number of
wayside measuring points occurring regularly along the line.
Refer to Section 5.5.1.3. As each measuring point have unique
measuring requirement its location next to the track will be
determined by the track layout and train operating
characteristics at that point.
These measuring points comprise of the following:
1. Visual vehicle inspections placed at the entrance to
large yards with camera/video equipment to identify
loose or missing components, e.g. springs, hoses,
brake pads etc.
2. Hot bearing detectors (called Hot Box detectors).
3. Brake efficiency measurements (called Hot Brake
detectors).
4. Brake problem measurement (called Cold Brake
Detectors).
5. Acoustic defective bearing detection.
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6. Dragging Equipment Detectors (called DED’s) are
spaced every 10km to detect shifted wheels or
dragging equipment.
7. Weighing In Motion and Wheel Impact Measurement
(called the WIM-WIM).
8. Skew bogie detection measures the induced lateral
forces to identify misaligned bogies.
9. Flange height detection provides an indication of the
wheel wear.
10. Automated Vehicle identification. (AVI). These
devices are not necessary at each measuring point.
5.6.2.1
Weighing In Motion and Wheel Impact
Measurement (WIM-WIM)
Figure 5.59: Track Mounted WIMWIM and Analysis Systems
The WIMWIM system measures the induced loads to
determine skew and overloading as well as impacts caused by
wheels with flats or out of round shapes. Spoornet has
invested in 6 WIMWIM systems on its heavy haul lines.
This is a strain gauged based system capable to measure
high speed traffic in excess of 250 km/h. All instrumentation
can be fitted to rails without removal of the rails or
interruption to the traffic flows. To accommodate different
wheel radii of 730 to 1220 mm some re-arrangement of
sleepers/ties, is required to obtain full coverage.
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The WIMWIM system is able to measure and determine
the following:
1. Wheel impacts in ton-force
2. Dynamic angle of attack – as a function of lateral
forces expressed in a couple force.(ton)
3. Mass of all vehicles inclusive of skew loading.
Various reports are generated on the system. These include
stop train alarms, maintenance reports and a full report for
archiving. Bad actors are eliminated when impacts exceed
45 ton (99kips) by stopping the train and 27 ton or above to
mark a wheel for maintenance.
The mass measuring is used to warn of skew loading
patterns or overloading. Overloading can cause major
problems for the allocated traction, which contributes to
skidding of under certain conditions. Skew loading must be
controlled to ensure that loads are within the applicable design
axle load of the line.
5.6.2.2
The Low-speed Weigh Bridges
Figure 5.60: Low Speed Weigh Bridge with
Individual Weighing Cell
Low speed weigh bridges are mass measuring systems
certified by the South African Buro of Standards as a device
approved for the use in trade to measure weights. The weigh
bridges are used by the commercial department for revenue
measurement purposes.
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These weigh bridges have a maximum speed limit of
8km/h and could create problems for high density lines due to
the time factor. (20-35 minutes for a 200 wagon train).
5.6.2.3
Hot-Box, Hot and Cold Brake Detectors
Figure 5.61: Hot Box Detectors and the Processors Within a
Wayside Container
Hot bearing detectors placed at regular intervals to
measure journal bearing temperature. The hot box systems are
regarded as safety-critical systems. Exceptions immediately
generate an alarm within the local CTC (Central Traffic
Control) Office. The temperature of each axle bearing is
captured for trending purposes. Presently the system is
sensitive to the ambient temperatures, resulting in problems
with trending the data. Upgrading of the systems will lead to
more accurate and reliable systems.
Brake efficiency measurements (called Hot Brake detectors)
measures wheel temperature at a location on the track where
brakes are normally applied; i.e., (after steep downgrade).
Brake problem measurement (called Cold Brake Detectors)
measures wheel temperature at a location on the track where
brakes are normally not applied; i.e., (after a long upgrade).
The hot and cold brake systems are derivatives of the
original hot box system, except that the measuring eyes are not
aimed at the bearing housing, but at the brake block locations.
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5.6.2.4
The Acoustic defective Bearing Detection
Figure 5.62: Acoustic Bearing System on Track and Inside the
Wayside Container
This measuring system was developed by Transportation
Test Center Inc. (TTCI) Pueblo CO U.S.A. as part of the proactive bearing management program to detect defective
bearings. It uses a microphone system to listen to the bearings.
The signal from each bearing is compared to the unique signal
template of a healthy bearing and exceptions are generated.
From the photo it can be seen that single track is used as multilane traffic cause difficulties with signal analysis.
5.6.3 Design Considerations
Figure 5.63 shows a diagrammatic representation of the
ITCMS At each measuring point, data is collected and passed
to a Field Measuring Station (FMS). Each FMS can
accommodate up to four measuring points (e.g. 2 hot box
detectors and 2 wheel impact detectors – or- 2 weigh bridges
and a skew bogie detection measuring point).
The FMS’s then sends the data from the field to an office
machine known as the “Data Concentrator.” This is usually
located at a main CTC offices. From the CTC’s the data is then
passed through to the main server at a head office where the
analysis system resides. Vehicle numbers are attached to the
data received. All exception parameters are generated and Emails would be addressed to the proper departments to deal
with any exceptions. All measurements are kept in a central
database for analysis purposes.
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Figure 5.63: Diagrammatic Representation of the ITCMS
5.7 Practical Application Of Wayside Lubricators
North American railroads have been applying lubrication to
the wheel/rail interface for many years to control wheel and
rail wear, reduce lateral forces in curves and produce
substantial savings in train energy (fuel) consumption. The
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traditional method of applying lubricant to the rail was by
means of wayside lubricators. In recent years substantial
improvements in wayside equipment technology has improved
equipment reliability, reduced maintenance requirements,
increased the track miles covered by each lubricator and
minimized wastage of the lubricant. Wayside systems however
do not always provide the maximum benefit of effective
wheel/rail lubrication. Results on Canadian Pacific Rail
demonstrate that wayside systems alone cannot produce the
recommended friction levels for the top of the rail. Wayside
systems must be supported by on board systems, currently
under development, such as hi-rail or locomotive systems, to
provide effective top of rail friction management in addition to
gauge face lubrication. Best practice friction management
guidelines for wayside systems are provided in this chapter.
5.7.1 Friction Management
Friction Management is the process of controlling the
frictional properties at the rail/wheel contact to values that are
most appropriate for the particular operating conditions20. In
general terms, this means that the goals are:
•
Lubrication of the gauge face of the rail to minimize
friction, wear and curving resistance (µ between 0.1
and 0.25).
•
Provide an intermediate friction coefficient (µ between
0.30 and 0.35) at the top of the rail under rail cars to
control lateral forces in curves, and rolling resistance
in both curved and tangent track. A special class of
products is generally required to achieve the
intermediate friction conditions21, 22, 23 - lubricants are
generally not suitable since they compromise
locomotive traction and safe braking of trains.
•
Improve traction under driven locomotive wheels (and
possibly under emergency braking situations) through
the application of adhesion enhancers. Sand is most
commonly used to improve adhesion but other
products, including alumina [JAPAN high speed] and
solid stick products24 are also used.
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5.7.2 Benefits of Effective Rail Lubrication
Benefits from effective wheel and rail lubrication have been
reported in many recent studies with wayside lubricators and
top of rail friction modifiers. Some of the benefits from
effective lubrication have been reported as follows:
1. J.deKoker 25 reports on tests on Spoornet which have
demonstrated 51% reduced energy required to traverse
a 200 metre radius curve, 28% less energy used by
trains on the Richards Bay Coal Line, and a 6 fold
increase in wheel life.
2. J.deKoker 25 reports lubrication studies by Sante Fe,
Conrail and ICG Railroads where energy savings of
25% to 30%, 24% and 17.5% respectively were
achieved.
3. Reiff26 documents the reductions in fuel consumption
at FAST of 30% with generous lubrication compared
to dry conditions. Also numerous lubrication tests in
the field on Class 1 railroads with long tangents, sharp
curves and grades have demonstrated fuel savings of
5% to 15%. Also a lubricated top of low rail and
generous high rail gauge face lubrication significantly
reduces curve lateral forces.
4. TTCI28 NUCARS analysis demonstrated energy
savings of: 15% with wayside lubricators, 39% with
Top of Rail friction modifiers alone and 65.5% with
top of rail and good wheel flange (gauge face)
lubrication.
5. J.Rucinski27 of Queensland Rail reports energy savings
on their narrow gauge coal lines of 4.3% for loaded
trains and 1.4% for empty trains.
6. Canadian Pacific Rail have found in recent studies that
improved wayside lubrication with preventive rail
grinding has increased rail life on average by 80% for
the high rail gauge face and 50% for the high rail top
surface. Canadian Pacific are working on top of low
rail lubrication strategies to improve the life of the low
rail.
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New systems are being tested on board locomotive and on
hi-rails to improve “top of rail” friction management through
the use of lubricants with modified coefficients of friction.
This section will focus on the application of wayside
lubrication equipment.
5.7.3
Wayside Lubrication – Capabilities and
Operation
Wayside lubrication systems have the potential to provide
substantial savings to railroads through reduced wheel and rail
wear, minimize track deterioration, and reduce fuel
consumption. Proper application includes:
1. Selection of the most appropriate equipment for
dispensing lubricant
2. Selecting the optimal type of grease for the particular
operating environment
3. Measurement and management of lubrication
effectiveness
4. Positioning of lubricators
5. Lubricator placement model
5.7.4
Selecting the Most Appropriate Equipment for
Dispensing the Lubricant
The performance of grease in the track is dependent on the
climate, the railway operating conditions, the dispensing
equipment utilized for the task and using a dedicated lubricator
maintainer. Field trials are required to determine the suitability
of the lubricant and the lubricator hardware for the territory.
Narrowing down the selection of the final grease for field trials
can be best achieved with laboratory test to determine
performance of the grease against key performance
characteristics as described in Section 5. New equipment
technology is available today, which has greatly improved
wayside lubrication effectiveness. Each railway has its own
specialized application. Overall, the choice of the best
lubricator system for the railway is dependent on the following:
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1. Ease of installation and simplicity of operation
2. Reliability of performance and ease of maintenance
3. Availability of spare parts
4. Availability of lubricant to be used
5. Financial considerations
The majority of wayside equipment in use today is the
mechanical contact or hydraulic system. The newer technology
lubricators are electronic which provide improved reliability
and serviceability. The systems employ a non-contact rail
mounted sensor, which detects the passing of wheels and
signals the electric motor to dispense grease. Control box
settings can be adjusted to regulate the volume of grease
dispensed to minimize wastage and maximize the distance
covered by each unit.
The wayside lubricator wiping bars vary in length from 61
cm (24 in.) with 18 grease ports to 140 cm (55 in.) with 48
grease ports. The longer bars will dispense grease over the
entire circumference of the wheel and minimize grease flingoff and wastage. Usually 2 bars per rail are installed in a tangent
location preferably adjacent to low to medium curvatures (less
than 3 degrees curvature), allowing the grease to carry for
greater distances. Reiff22 reports that Norfolk Southern
Railway introduced longer and improved lubricator bars and
found 107% improvement in grease carry distance for gauge
face protection, 67% reduction in grease consumption, and
57% reduction in grease wastage. Improvements in lubricator
efficiency reduced the number of lubricators from 49 to 20 in
an 80-mile mountainous territory. Also train stalls were
completely eliminated.
Implementing an effective wayside lubrication strategy
requires a trained and dedicated lubricator maintainer.
Canadian Pacific and Canadian National Railways have
employed full time lubricator maintainers to greatly improve
the reliability and efficiency of the lubricators. This strategy
ensures grease is on the rail all the time and reduces rail/wheel
wear and fuel used by the locomotives.
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5.7.5
Selecting the Optimal Type of Grease for the
Particular Operating Environment
The three key characteristics of lubricants that impact
performance in wayside systems are:
1. Lubricity refers to the lubricant’s capacity to reduce
friction, with poor lubricity being associated with
higher wear rates. But since the rates of wear under
“dry” conditions are orders of magnitude greater than
those under lubricated conditions, the key is to ensure
that there is lubricant where needed at the wheel/rail
interface. Of less importance is whether the lubricant
provides a friction coefficient in the field of 0.10 or
0.25. Recent test results from various manufacturers
show greases have about the same lubricity.
2. Retentivity is a measure of the time (or number of
wheel passes, or mgt) that the lubricant is able to
retain its lubricity. For example, grease, which operates
under boundary lubrication conditions, can only
partially separate the surfaces of the wheel and rail.
The individual microscopic protrusions on each
surface – referred to as asperities – are engaged with
each other. The frictional heat generated by any such
encounter can yield a “flash temperature” measured in
hundreds of degrees Celsius. Flash temperatures of
600 to 800ºC are typical of rail/wheel contacts. The
lubricant is consumed at these microscopically
localized hot spots, literally being “burned up.” Once
all the lubricant is gone, the coefficient of friction
quickly climbs from its “lubricity” coefficient of
friction of 0.05-0.1 to its “dry lubrication” coefficient
of friction of about 0.6. Retentivity is a function of
load and creepage. Laboratory tests show that
retentivity decreases with increasing load and
increasing lateral creepage (angle of attack). The
practical implication of this is that loaded trains
consume (“burn”) grease at a much higher rate than
empties, and sharp curves consume grease much faster
than mild curves.
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3. Pumpability: The importance of technical variables
and actions to ensure a continuous delivery of
lubricant to the wheel/rail interface cannot be
overemphasized. Preventing gauge face wear in curves
thus depends greatly on the pumpability (which is
affected by, among other things, the stirability of the
grease) and ensuring that lubricators are not allowed to
go dry or be shut down for prolonged periods of time.
The performance of lubricants at various temperatures
in track depends on their ability to be pumped at all
temperatures experienced on the railway system. For
example on the Canadian Pacific, the operating
temperature range is – 40° to + 60° Celsius. Testing of
the lubricant in a cold chamber at temperatures of –
40°C shows that grease become stiff, while at hotter
temperatures of say + 60°C the grease tends to
separate and slump from the rail.
The performance of greases is often stated in terms of
Timken or 4 ball performance tests, which are high pressure,
high-speed evaluations of the ability of the lubricating
constituent to be drawn from the carrier to perform its
function in a millimetre sized contact zone. These tests
correlate poorly with field tests of railway lubricant
performance.
The rail/wheel contact occurs over a dime-sized patch and
is macroscopic when compared with the thickness of the
lubricant film. At the wheel/rail interface, the lubricating
constituent (e.g. graphite or moly) is taken into the interface
along with the carrier (e.g. soaps) to provide the final
performance. Laboratory wheel/rail simulations, using fullsized and smaller scale test rigs, have proven effective in
evaluating the comparative performance of various greases at
the wheel/rail interface. But ultimately, the performance must
be measured under real operating conditions.
Canadian Pacific used the laboratory testing of various
greases to selected a candidate grease with high retentivity,
good gauge face lubricity, was suitable for summer and winter
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operation in their northern climates and was within their
budget.
5.7.6
Measurement and Management of Lubrication
Effectiveness
In order to achieve savings from an effective lubrication
strategy railroads have to properly manage wayside lubrication
strategies. Typical track locations have to be measured and
monitored at regular intervals. Lubrication effectiveness has
been measured by a hand-operated tribometer for a number of
years (Figure 5.64). This equipment applies a known load to a
wheel which is set to run on the rail gauge corner or top rail
surface. The wheel is initially free running and rolling
resistance gradually increases until slippage occurs. The
resistance (force) is proportional to the coefficient of friction
between the wheel and the rail. This equipment is useful when
applied to short distances of track for monitoring purposes. A
Hi-Rail tribometer (Figure 5.65) allows a railway to measure
large distances of track at speeds of up to 30 mph. Data is
collected simultaneously from the top and gauge corner of
both rails. Examples of data collected on Canadian Pacific
Railway by both systems in the years 2000 and 2001 are shown
in Figure 5.66 and Figure 5.67
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Figure 5.64: Hand Operated Tribometer used
on Canadian Pacific Rail
Figure 5.65: Portec Hi-rail Mounted Tribometer used on
Canadian Pacific Rail
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Figure 5.66: Friction Data from the Hi-Rail Tribometer on
Canadian Pacific Thompson Subdivision
between Milepost 0 and 50
In October 1999 Canadian Pacific conducted a hi-rail
tribometer run over their system. Figure 5.66 shows the
coefficient of friction on the top surface and gauge face over a
50-mile section of the Thompson Subdivision. At that time 18
hydraulic lubricators were used in this section of track. This hirail tribometer run demonstrated to Canadian Pacific how poor
their lubrication conditions were. Although, the section forces
were spending considerable time maintaining these lubricators
and continuously replacing parts the lubricators and the grease
were not effective. In October 2000, Canadian Pacific installed
10 new electric lubricators with a new grease in this 50 miles of
track and a dedicated lubrication maintainer was appointed for
the entire sub-division. The grease was selected from
performance in laboratory tests and the cost from the
manufacturer. The effectiveness of the lubrication improved
significantly (refer to Figure 5.67).
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Figure 5.67: Friction Data obtained from the Hand Operated
Tribometer on Canadian Pacific Thompson Subdivision in
2000, between Milepost 10 and 14.5
The Thompson Subdivision between milepost 10 and 14.5
consists of a series of back to back, sharp curves, of up to 11
degrees in curvature. In January 2001, a measurement of
lubrication effectiveness, Figure 5.67, demonstrated the
improved gauge face lubrication achieved in a representative
section of the subdivision. The coefficient of friction on the
gauge corner is below the target level of 0.25. Previously three
hydraulic lubricators were used in this section of track and now
there are two. Note that the target top of rail coefficient of
friction has not been achieved using wayside lubricators alone.
Canadian Pacific has established best practice targets for
lubrication so that the process can be better managed and
improved in the future. Canadian Pacific is also working on a
formulae (refer to Section 5.9) to optimize the location of
lubricators based on various performance parameters. Having a
formulae allows a railway to compare the performance of new
greases and new equipment introduced in future years.
The coefficient of friction guidelines that Canadian Pacific
has adopted for lubrication management are as follows:
1. Maintain top of rail friction coefficient differential, left
to right <0.1µ
2. Top of Rail friction 0.3 ≤ (µ) ≤ 0.35
o 5-135 x
3. Gauge face of high rail coefficient (µ) ≤ 0.25
Although top of rail friction coefficients are not being
achieved, Canadian Pacific is investigating better strategies at
this time. Lubricators are being continuously evaluated,
adjusted and fine-tuned to provide the optimal placement and
optimal settings.
5.7.7 Positioning of Lubricators
Every railway territory is different. Variations are experienced
in curve radii, tangent lengths, track gradients, traffic type and
wear state, train speed and braking requirements, axle loads,
rail grinding strategies, climate, etc. All these factors influence
the migration and retentivity of the lubricant on the rail. The
optimal placement of lubricators must consider these
environmental and operational factors, within the constraint
that the lubricator must be maintained.
Controlled field testing can establish the reliability and
efficiency of wayside lubricators based on the following
factors:
1. Prevention of wastage by fling-off and build up on the
top of the rail
2. Monitor grease burn-off with the passage of trains
3. Measure the distance covered by each lubricator
4. Good pumpability at all temperature ranges
5. Prevention of plugging of the lubricator ports
6. Ensuring no washing of grease with rain and snow
7. Ensuring the grease remains on the gauge corner, does
not contaminate the top of rail and does not slump off
the gauge corner with high ambient temperatures
8. Other factors not related to the grease or the
lubricator are: ensure that the rail gauge corner on the
high rail after grinding is smooth with no deep
grinding facets (this prevents the transfer and spread
of grease); ensure that track gauge is within +/- 1mm
at the lubricator site; ensure that trains are not hunting
at the lubricator site
o 5-136 x
The consequences of poor wayside lubrication are as follows:
1. Locomotive wheel slip and loss of braking capability
on grades
2. Poor train handling
3. Prevents ultrasonic rail flaw inspection
4. Gauge corner shelling on infrequently ground rail
5. Wastage of lubricant
6. High lateral forces in curves and increased degradation
of track
Tests were conducted on Canadian Pacific Rail using
electric lubricators in the Thompson Subdivision to determine
the optimal setting for reduced lubricant wastage for long and
short bars. Lubricators either side of one electric lubricator
were turned off for 3 days with traffic of 30 trains per day to
dry down the influence of surrounding lubricators. The track
between the rails was covered with clear plastic for 15 metres
either side of the lubricator and lubricator control box settings
were increased with the passage of each train varied until flingoff occurred. With long bars in place the optimal setting for
the Canadian Pacific lubricator was ¼ second of motor
operation for every 16 wheels. At this setting, the volume of
lubricant dispensed was measured in ounces per wheel. The
distance covered by the spread of grease on the gauge corner
of the rail was measured to the point where the coefficient of
friction exceeded 0.22.
At the same time all the factors listed in items 1 to 8 were
evaluated during the winter and summer months. The
lubricator maintainer was trained by the equipment supplier in
the operation and maintenance of the new technology
electronic lubricators.
In the past Canadian Pacific had a formulae for lubricator
placement. This added the product of curvature times the
length of the body of the curve (including half the transition
length) not to exceed 600 degrees. On Spoornet,25 wayside
lubricators on bi-directional lines, i.e. carrying traffic in both
directions, were spaced about 6 km of high leg curve distance
o 5-137 x
apart, or between 300° and 360° of deflection angle of the curve.
The Richards Bay Coal Line had lubricators installed at 87° of
deflection angle or 1.5 km of accumulative high leg rail apart.
5.7.8 Lubricator Placement Model
Spoornet, has developed criteria and an equation for
positioning wayside lubricators.29. The approach has also been
applied to the Canadian Pacific specific traffic conditions.
The placement of lubricators is influenced by a number of
factors listed and discussed below. In general, the length of
track being considered for lubrication is adjusted by a number
of track related factors. The adjusted length is then divided by
a number of traffic related factors to determine the placement
increment. These factors known to influence the carry distance
of the grease will be discussed.
5.7.8.1 Track Related Factors
Length of the curve; this factor is determined by each curve
length in the length of track adjusted for its degree of
curvature. This factor is considered to be the primary track
related factor in the final equation.
Length of straight between curves; this factor is added to
the primary length of the curve, and in effect extends the
effective length of curvature. On straight track the play
between the wheel flanges and the side of the rail is between
15-35 mm depending on wheel flange wear. "Hunting" or
lateral sinusoidal movement occurs when the trucks or wagons
oscillate between the two rails, resulting in contact between the
wheel flanges and the rail. The grease on the wheel flanges is
wiped off due to this contact and it should be taken into
account when calculating the position of rail lubricators. The
straight track contact should be determined for each site
considered.
Gradient; the influence of the gradient is difficult to determine
but it is known that steep gradients are usually associated with
sharp and long curves, few and short straight sections, slow
train speed and continuous application of brakes on the trains.
These factors should be considered on their own merit.
o 5-138 x
Type of grease used; each railway usually standardizes on one
type of grease for an extended period of time or would use
special types of grease for certain sections of track. If a change
is made from one type of grease to another, the carry distance
of the new grease should be determined by laboratory and field
testing. The necessary adjustments to the distance between
lubricators should be made when the new type of grease is
introduced for general use on a track section. If this factor is
neglected, over or under lubrication could result.
Applicator bar configuration; if shorter applicator bars are
used, less grease is applied to the wheels, with the resultant
reduction in grease available for carry down the track and
reduced length of curves that are lubricated.
5.7.8.2 Traffic Related Factors
Direction of traffic; on bi-directional lines trains run in both
directions on the same track and the lubricators move grease
towards each other. On lines carrying traffic only in one
direction, this does not happen and double the number of
lubricators is needed. For Canadian Pacific Rail the traffic
patterns on bi-directional track are equivalent to directional
traffic. Several trains are scheduled to run back to back in one
direction followed by several trains in the opposite direction.
The gauge face has been measured to dry down when
lubricator are placed at a spacing suitable for bi-directional
traffic. Thus traffic patterns are site specific.
Type of bogie (or truck); the design of the bogies traversing
the line plays a major role in the usage of grease. The equation
has to be adjusted based on the curving capability of the
bogies. If most of the bogies are self steering there is less need
for lubrication. In general for Canadian Pacific frame braced
bogies can steer in 5 degree curves and 3 piece bogies can steer
in up to 2 degree curves. For Spoornet, self-steering bogies can
steer through 8 degree curves.
Axle loading of the rolling stock; as the axle loadings on the
trucks and locomotives increase, the lateral component of the
flange force will also increase, thus axle loading and
distribution should be taken into account.
o 5-139 x
Locomotive bogie wheel base; the longer the wheel base the
more likely that the wheel of the locomotive will contact the
rail.
Speed of the train; as the speed of the train increases, the
dynamic forces created by the train also increase, which
adversely affects the lubrication distribution.
Angle of attack between the wheel flange and rail; the
angle of attack determines the forces acting on the grease film
applied on the rail. The more acute the angle, the higher the
lateral forces.
Miss-aligned and skew bogies; tests indicate that significant
additional train resistance is generated by trucks with
misaligned bogies and axles, mainly on tangent track. It was
reported that a misalignment of 4 mrad could double the
rolling resistance of a truck on tangent track. Canadian Pacific
reported an average misalignment of 1 mrad on all bogies
(trucks). Thus bogie conditions should be considered.
Braking of the train; the application of brakes on the train
causes the wheels to heat up and the grease is burnt off where
severe braking takes place. In areas where this happens the
distance between wayside lubricators should be reduced.
5.7.9
Case Study: Lubrication: Richards Bay Line:
South Africa
Locomotives, operating the Richard’s Bay line between Ermelo
and Richard’s Bay experienced extreme flange wear (between
50 and 100 000 km between wheel reprofiling). Lubrication
was applied to the gauge corner of the high leg in all curves of
less than 1 000 m radius by means of apparatus mounted to a
motorized rail trolley. The improvement in flange wear, in
relation to the amount of grease applied, is illustrated by Figure
5.68. It may be concluded that wheel life is a function of the
amount of lubrication and of the vehicle type (axle load, bogie
wheelbase).
o 5-140 x
Figure 5.68: Wheel Life and Grease Consumption vs. Time
The track was laid with 60 kg/m manganese rails (1065
mm gauge) with curvature distributions as shown in Figure
5.69.
The locomotives type 11E, 7E1, 7E3 are of Co-Co
configuration with characteristics as shown in the table.
Locomotive Type
11E
7E1
7E3
Nominal Axle load
28
21
21
Distance between
Outer Wheels of Truck (m)
4.4
4.4
4.06
What is interesting to note from Figure 5.68 is a 2.5 times
increase in the amount of grease applied to the line resulted in
a tenfold increase in the wheel life of type 7E locomotives.
The application of less than 1 000 kg of grease per month
o 5-141 x
however had little affect on wheel life. The amount of grease
applied was to a degree related to the number of curves which
were lubricated on the line and led to the following hypothesis:
Figure 5.69: Spoornet Coal Line Curve Distribution
Consider a locomotive running on a track of length
(say) 100 km. After one trip the flanges have worn by 10
mm. This wear is represented by the graph indicated
“worst” in Figure 5.70.
Figure 5.70
o 5-142 x
A mechanism (for instance lubrication) is introduced on
the whole line which reduces the rate of wear of the flanges by
a factor of 10. The wear rate is now represented by the graph
“best.” If the lines is now cleaned of all lubricant and only the
first 50% of the track is lubricated the wear may be
represented by the graph indicated by the term “50%.” This
process may be repeated for various “percentages” of lubricant
applied or curves lubricated. Now, if resultant wheel life is
plotted against the percentage of the line lubricated, the result
shown in Figure 5.71 is obtained. This indicates that the
maximum returns on lubrication is achieved only with
complete lubrication and may go some way to explaining the
vast differences experienced by railroads in applying
lubrication to combat wear. This hypothesis has not been
verified.
Figure 5.71
5.8 Optimizing Wheel And Rail Life
The wheel and rail together are responsible for transmitting the
static and dynamic loads from the car body through to the
track structure. At the contact patch between the wheel and
rail, the vertical load of the wheel must be supported, and the
steering, braking and traction forces transmitted.
The interactions that occur at the wheel and rail contact
play a pivotal role in the end performance of the vehicle/track
system. Looking at the rail, wheel, and its interface
o 5-143 x
independently as well as managing the wheel/rail interface as a
system will optimize the performance of a heavy haul
operation.
5.8.1 Rail Optimization
The rail is the static component of the wheel/rail interface and
thus easier to measure and maintain. To optimize the rail life
the limits of the rail fatigue and rail wear should be reached
simultaneously, preferably removing the rails for wear before
the risk for rail fatigue becomes too high.
When rail fatigue is brought under control through regular
rail re-profiling, rail in any curve or tangent section can be
targeted for replacement at the full limits of wear.
5.8.1.1 The Rail Management Decision Zones
Experience has shown that rail performance is by no means
uniform, and a mature rail line will see varying levels of rail
wear dependent upon track curvature, lubrication regime and
even sub-grade type. It makes good economic sense to look at
rail planning units as individual legs of individual curves and at
300–400m (0.18-0.25 mi.) lengths of tangents. In fact, the
overwhelming influence of rail material costs over the cost of
rail installation makes it usually uneconomic to replace both
legs of a curve in the same year, if there is at least one year of
additional life in one of the legs.
A rail’s passage through various stages of wear can be
plotted on a wear diagram such as that shown in Figure 5.72.
The two-dimensional representation with vertical wear on the
y-axis and gauge face wear on the x-axis is in recognition of the
fact that both the geometrical constraints on rail wear and the
build-up of internal stresses, are jointly related to the loss of
railhead height and lateral wear.
The representation is useful to plot a rail’s wear progress.
The progress of the rail wear through its life can then be
visualized as a vector. The objective is to chart the pathway
that results in the maximum rail service life, stopping just short
of incurring risk or of affecting other cost areas, including
foreshortened life of other track components. Projections of
the progress of rail wear can follow the same vector direction.
o 5-144 x
On this diagram, there are a number of decision points
that invite an improved service life. Assuming that fatigue is
being controlled, wear will progress to the point where it must
be considered whether the rail in its current condition would
better serve the railway in a secondary line, called the “Zone of
Possible Relay.” This decision depends upon the ratio of the
demand for second position rail to new rail, a function both of
the railway’s tonnage distribution over the network and the
policy on the traffic density that merits laying with new rail.
The objective is to relay a rail that has some remaining wear
life into a curve with a greater need for rail of that quality.
Often the relay decision is contingent upon the value of using a
newer, premium steel in the main line. Another common
cause of relay is a critical need of rail in a secondary line.
Where partly worn continuously welded rail is relayed to a line
previously laid with jointed rail, there is a benefit gained in
improved track performance in the secondary line.
Figure 5.72: Rail Management Decision Zones
o 5-145 x
In jointed track, wear can only progress to that
combination of vertical and lateral wear that ultimately results
in contact between the flange of tread worn wheels and the top
of the joint bar (Line A). New joint bar designs that remove
material from the location of the interference (Figure 5.73)
open the way for extended wear limits. These designs make
full use of modern finite element techniques and improved
understanding of the relative contribution of shear and flexural
action in the joint bar under dynamic loading. With the use of
joint bar assemblies with improved flange clearance, emergency
rail repairs can also be made in heavily worn continuous
welded rail.
Figure 5.73: New High Clearance Joint Bar Design
The joint limit threshold may also be retained as a wear
limit for a rail that has shown a history of past internal defects,
as it has probably sustained such accumulation of subsurface
fatigue that it should not be taken to extended wear limits.
o 5-146 x
On the other hand, a rail in continuously welded rail
territory, which has been regularly reprofiled by grinding, can
be worn to an extended limit (Line B). But it must be certain
that the renewal has taken place by the time these stresses have
reached the point where the risk of sudden fracture is
unacceptable (Line C).
In a mixed freight railway, rail between Line A and Line B
may be removed prematurely to serve as maintenance rail,
replacing individual lengths of rail removed because of rail
defects. This rail needs to have some service life left and to
match the cross section of existing rails. Rail between Line B
and Line C may be suitable for slow speed use in a yard. Rail
beyond Line C must be scrapped. By using well defined riskbased wear limit standards like these, again assuming that
grinding and testing have managed the fatigue issues, can
maximize the utilization of the expensive rail asset on a railway.
The wear condition of each individual curve can be plotted
in the manner as shown in Figure 1. At various points in the
progress of this wear vector, the track engineer is faced with
one of the following decisions:
•
Re-profile rail by grinding
•
Adjust lubrication
•
Adjust cant
•
Re-gauge
•
Transpose rail
•
Replace and use in a lower tonnage line
• Replace and scrap rail
With good, regular condition assessment and a clear
understanding of the maximum limits of rail deterioration, it
should be possible to recognize the value of these various
options and to proactively manage each individual curve or
length of tangent to a long life in track.
There is frequently a value to transposing a curve rail with
a high ratio of lateral to vertical wear to get full use of its
o 5-147 x
vertical wear limits. Transposing can involve either exchanging
a high and a low rail in the same location to expose a new
gauge face, setting the high leg down to the low leg and putting
new steel into the high leg, or exchanging worn with another
curve or tangent. The “Transpose Zone” would represent the
stage of wear at which it may be feasible to consider a
transposition. To be economical, transposition requires
somewhat greater lateral wear than vertical wear. Because the
optimal shapes of the high and low rail are different, and
transposition reverses the direction of flow, timely rail
reprofiling by grinding is an important element of this strategy.
Some railways do not do strict transposals for this reason, but
would set a used high rail down to the low rail position (a
“setdown”). To result in an overall net saving relative to
leaving the rail in its initial condition, transposition must be
done at the right time. Wear studies6 have shown that the
“window of opportunity” may be quite short.
5.8.1.2
Controlling Rail Wear (Maximum Rail Wear
Limits)
The objective of extended wear limits is to replace rail just
before its expected cost of remaining in track exceeds the value
of deferring its replacement. If there is not a major distortion
of the rail’s shape with age, or substantial loss of gauge, this
economic limit is likely correlated with the build up of an
unacceptable level of rail stresses internal to the rail. High
internal rail stresses will ultimately lead to higher rates of rail
failure, which have clear economic consequences.
Therefore, it makes business sense to determine maximum
limits of internal rail stresses for setting targets for planned
renewal of rail. Stress-based wear limits must of course be
backed up with frequent grinding and rail inspection practices.
Rail internal stresses are largely related to the depth of
material left in the railhead. The highest stresses come from
gauge corner loading, but are controllable with rail grinding. A
critical condition exists where the height of the rail is reduced
such that the influence zone of contact stresses and the
stresses of railhead and web bending coincide. Rail web
stresses are generally below 207 MPa (30 ksi).
o 5-148 x
These influences were investigated in a study sponsored by
CPR and BC Rail (Igwemezie, 1992). The study utilized Linear
Elastic Finite Element Analysis to examine stresses in 100
lb/yd. (50 kg/m) Canadian Pacific Re, 115 RE (57 kg/m) and
136 RE (68 kg/m) rail at different wear levels and loading
configurations.
Typical peak vertical loading of 173 kN (19.5 tons) was
assumed in the analysis. These were applied in the model in
conjunction with an 87 kN (9.8 tons) transverse frictional force
simulating peak curving forces and an 87 kN (9.8 tons)
longitudinal force to simulate support, however the difference
in internal stresses were minimal.
Experience in the past with excessively worn rail has
shown increased rates of occurrence of vertical split heads.
Therefore, the criterion set for maximum wear was that level
of wear that causes unacceptable stresses in a zone of exclusion
defined by investigating the envelope of crack initiation point
for vertical split heads found in track. This zone of exclusion is
shown in Figure 5.74.
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Figure 5.74: Maximum von Mises Stress for all Load Cases vs.
Total/Combined Railhead Wear for 115 lb/yd RE Rail
Figure 5.74 plots the maximum von Mises stress against
total wear for three wear patterns:
1. Wear that is exclusively vertical wear.
2. Wear that is 50% vertical and 50% lateral wear.
3. Wear that is exclusively lateral wear.
The latter is not practical in the field, but was required to
establish a wear envelope. By selecting a given stress level, and
reading off the intersection points with the curves established
for each of the three wear patterns, it was possible to define
the wear envelope of Figure 5.75. In this figure, wear levels
corresponding to different maximum stresses in the zone of
o 5-150 x
exclusion are plotted against the former rail wear limits used by
CPR and BC Rail.
Figure 5.75: Wear Interaction Curves for 115 lb/yd RE Rail
It was determined that the maximum safe stress level for
internal rail defects was 67% of the yield strength of the lowest
yield rail used (517 MPa (75 ksi) standard carbon rail), or 345
MPa (50 ksi). This yielded a basic Line A outline for new wear
limits which was modified to consider the need to limit rail
gauge face wear to control gauge. The resulting rail wear limits
for 136 RE (68 kg/m) and 115 RE (57 kg/m) rail are shown in
Figures 5.76 and 5.77.
Line C represents the point at which trains must be slow
ordered until the rail is replaced. In fact, it is targeted to plan
replacement to correspond with reaching Line B. Line B
represents a stress level of 276 MPa (40 ksi).
o 5-151 x
Figure 5.76: Rail Management Decision Zones for 68 kg/m (136 lb. RE)
o 5-152 x
Figure 5.77: Rail Management Decision Zones for
57 kg/m (115 lb. RE)
Figure 5.78 is an accurate representation of the total rail
wear of a 68 kg/m rail (136RE) at Line C.
o 5-153 x
Figure 5.78: Accurate Representation of the Total Rail Wear of
a 68 kg/m Rail (136RE) at Line C
5.8.1.3 Rail Use Strategy
The strategy for achieving maximum rail economics is to
regularly re-assess the progress of a rail through its wear vector.
Regular rail grinding with advance rail condition assessment is
designed to prevent premature replacement due to excessive
rail surface fatigue. Re-profiling also controls field side
wheel/rail contact that would introduce eccentric loading.
Excessive gauge corner damage is also controlled as a strategy.
In mild to intermediate curves, the objective is to plan to
replace rail at Line B (Figure 5.79). In the high legs of sharp
curves, transposal is targeted to achieve maximum rail
utilization if and when the rail passes through the Transpose
Zone (Figure 5.80).
o 5-154 x
Figure 5.79: Rail Renewal Strategy –
Mild to Intermediate Curves
Parallel ultrasonic rail testing and defect projections are
performed to assess the need to replace the rail due to internal
rail failure.
Using harder and cleaner steel has a number of advantages
for the rail/wheel interaction. In terms of rail grinding, another
positive influence of harder, cleaner rail steels is that they
should require less grinding. When ground to templates that
reflect natural wear patterns, i.e. conformal contact, their shape
remains stable and requires less grinding to compensate for
plastic flow. Harder rails require tighter tolerances on the
profile, but with the right shape and the right yield strength
and cleanliness, alloyed heat treated rail steels have the capacity
to withstand today’s loadings without premature fatigue. If
o 5-155 x
they must be reshaped frequently, the grinding template is
wrong.
Figure 5.80: Rail Renewal Strategy – Sharp Curves
o 5-156 x
5.8.1.4 Lubrication and Curve Elevation Monitoring
A good rail wear measurement system is also useful in
monitoring whether a site is effectively lubricated, permitting
early correction. If there is a systematic variation in the rail
gauge face wear between one end of a curve and the other, this
may indicate ineffective lubrication distribution through the
curve. If there is a correlation between rail wear variations and
cant deviations as seen in track geometry car traces, this may
indicate a need to correct cant.
In addition, comparison of reported ratios between gauge
face and vertical wear against standards of the track curvature
can be used to identify priority locations for correction or cant
or lubrication. Canadian Pacific Rail has developed the gauge
face to vertical wear rates typical of well maintained curves
(Table 5.13) for use in identifying locations that may require
adjustments to either lubrication or curvature to achieve
maximum rail life.
Table 5.13: Gauge Face to Vertical Wear
Ratios for Well Maintained Track on CP
Rail (95% Conventional bogies)
RADIUS IN
HIGH LEG
METRES
1746 R
0.20
873 R
0.29
582 R
0.31
437 R
0.35
349 R
0.45
291 R
0.50
249 R
0.52
< 249 R
0.60
Such measurements for the high leg can be compared with
average elevation through the curve compiled with the rail
wear files as the track geometry car measures both rail wear
and cant. If the ratio is out of standard and the curve is not
seen to be under-elevated, lubrication effectiveness is
suspected. The equivalent figures for the low leg might
indicate over-elevation.
o 5-157 x
5.8.1.5 Transposition
Rail transposition can be an economical practice in some lines.
The range of situations where it is beneficial is becoming
increasingly restricted with effective lubrication and use of
steerable bogies. On the other hand, modern rail profiling has
reduced past problems with poor profiles for rails moved from
high to low leg positions and vice versa, as well as the contact
fatigue problems attending redirection of the plastic flow. A
stress analysis on transposed rail (Igwemezie, 1993) showed
that it is beneficial to transpose only if the gauge face wear is
considerably greater than the vertical wear. Referring again to
Table 5.13, this should not be the case except for very sharp
curves.
Figure 5.81 shows the results of the stress analysis for 115
RE (57 kg/m) rail. In Figure 5.81, the von Mises stress
increases from 156 MPa (23 ksi) when the rail is new to 345
MPa (50 ksi) with 22 mm of total wear, comprised of 11 mm
vertical wear and 11 mm gauge face wear. If the rail were to be
set down to the low rail position at point B, i.e. with 7 mm of
vertical wear and 7 mm of gauge face wear, the rail in the low
rail position would jump up to position D and continue along
to point E. This would represent a stress level in the zone of
influence of 283 MPa (41 ksi). In the low rail position, the rail
would reach the 345 MPa (50 ksi) stress threshold after an
additional 6 mm of vertical wear. The total wear on the rail if
it were left to run to maximum wear limits in its original
position is therefore 22 mm, which is greater than the 20 mm
total wear for the transposed rail.
o 5-158 x
Figure 5.81: Maximum von Mises Stress vs. Total Wear for
Transposed 57 kg/m (115 lb/yd) Rail
5.8.2 Wheel Optimization
As the wheel negotiates high rails, low rails and tangent track,
at narrow and wide gauge, the contact patch moves all across
the wheel profile. The worn wheel shape is therefore an
“envelope” (Figure 5.82) of all the rails that it contacts. At
portions of the profile where there is a greater frequency of
contact, more slip and higher stresses, wear will be greater than
at other sections of the profile. A freight railway with good
track alignment and primarily tangent track can be expected to
wear mostly at the wheel tread and encounter little flange wear.
Territories with considerable curved track will encounter gauge
face wear and often fatigue at the field-side of the wheel.
The design of a wheel profile is thus specific to its
operating environment; for a captive fleet the design is easier
than for a railway that combines heavy haul and general traffic.
Some railways like BHP Iron Ore in Australia, Spoornet in
South Africa and others around the world have worked at
developing profiles optimized for their service.
o 5-159 x
Figure 82: The Worn Wheel Profile is an "Envelope" of All the
Rails that it Contacts
The optimization of the wheel profile in North America
has been influenced by the unique nature of the North
American Railway industry. The concept of the “worn wheel”
profile appeared in the 1970’s and has gone through various
iterations. Currently the AAR1B is the current NA interchange
standard. It has been recognized that this wheel is not a good
representation of the North American service worn wheels. A
number of developments have recently taken place where
wheel profiles are designed for specific applications. Such an
example is the NRC –ASW.
The NRC-ASW is a “worn” wheel profile designed to
minimize creepage and contact stresses that contribute to
rolling contact fatigue shelling of steel wheel treads. This wheel
profile provides the following geometrical features compared
with the AAR1B (see Figure 5.83):
•
The addition of 1.6mm of metal in the flange root,
which significantly improves steering performance,
reducing creepage and wear.
•
The 1:20 cone angle in the tread contact region (same
as the AAR1B) leaves unchanged the wheel’s
resistance to hunting in standard gauge, tangent track.
•
A 20” field-side roll-off is another notable change,
since it further improves the wheelset steering
moment, and increases significantly the time to
development a false flange.
o 5-160 x
Figure 5.83: Comparison between the NRC-ASW and
AAR1B Wheel Profile
The rail and wheel profiles have to be designed as a system
(refer to Part 2.4) and the way the wheel profile mates with its
rail has a direct bearing on a variety of performance issues
including:
1. Wheel and rail wear
2. Wheel and rail fatigue
3. Wheel climb and L/V derailments
4. Track hunting
The flange face must be at an angle sufficiently large
enough to inhibit wheel-climb derailments. The flange root
must minimize contact stress with the gauge corner of the rail
but simultaneously provide adequate curving. The wheel tread
must provide stability. A false flange on the field side of the
wheel must be avoided since it compromises steering and is
responsible for very high contact stresses at the low rail.
The establishment of a hollow wheel limit is the subject of
recent intense research in North America. Some railroads,
such as Spoornet in South Africa and the Cartier Railway
Company in Canada have already adopted a wheel-hollowing
limit for their operations, re-turning wheels at about 2 mm of
hollowing. As captive operations, they are able to quantify and
capture the benefits of an improved wheel/rail interaction,
which include reductions in wheel shelling, curving resistance,
wear and fatigue. Most North American Class 1’s are waiting
for an AAR ruling to emerge before considering a similar
practice. It appears that a 3-mm hollowing limit may be
established for interchange service based on current economic
models.
o 5-161 x
The wheel profile undergoes very large changes through its
life, usually more so than the rail, which in most railroads is
regularly re-profiled through grinding. The unworn freight
wheel (Wide and Narrow flange standards) generally starts with
a flange of width 30-37 mm that can narrow to about 24 mm
before being condemnable. Tread wear, meanwhile, is limited
to about 11 mm, since at greater levels, the “high flange” can
impact joint bars, wayside lubricators, turnout components etc.
All of the above have to be taken into account including
the environment and layout in which the service is provided.
Tournay30 suggest the following “recipe” for profile design:
•
Recognize that the lateral creep force is the most
damaging mechanism to wheel and rail
•
Recognize the symptoms in the form of wear and rail
and wheel damage attributable to the lateral creep
force
•
The lateral creep force may be appreciably reduced by
the use of steering bogies
•
Steering can be enhanced by conformal flange contact
•
Conformal flange contact is an optimum condition for
non-steering vehicles and supports lubrication
•
Lubrication is a most effective “panacea” in
preventing rail and wheel wear and damage
•
Hollow wear control will improve wheel and rail
damage and prevent the impairment of the vehicle
steering properties
•
Hollow wear rates may be extended by gauge variation
on tangent track
•
Gauge control in curves will improve tracking and
reduce wear of the high leg and contact fatigue on the
low leg
o 5-162 x
5.8.3 Friction Management (Interface Optimization)
The benefits of lubrication have been known for quite a while
and have been highlighted as a “best practice.” It is only in
recent years that substantial improvements in wayside
equipment technology have improved reliability, increased the
track miles covered by each lubricator and minimized wastage
of lubricant. Wayside systems must be supported by other
systems to provide top of rail friction management to improve
lubrication effectiveness benefits. New systems are being
tested on board locomotive and on hi-rails to improve “top of
rail” friction management through the use of lubricants with
modified coefficients of friction.
TTCI predictions using their NUCARS model peg fuel
savings up to 13% when the top of rail is lubricated. The high
rail and low rail friction must be in balance, however. High
lateral loads were predicted when the gauge face is lubricated,
but the top of both rails is dry, or when the top of the high rail
is lubricated, but the low rail is quite dry.
TTCI’s guidelines can be stated as:
•
Maintain top of rail friction coefficient differential, left
to right <0.1µ
•
Maintain top of rail friction>0.30 µ
•
If the top of the high rail becomes lubricated, lubricate
the low rail too.
•
Maintain gauge friction coefficient of <0.25 µ
Recently, hi-rail powered tribometers are available that can
survey both top of rail and gauge face friction continuously
over a subdivision.
The “best practice” should include:
1. Periodic measurement of top of rail and gauge face
friction and managing the friction in correlation with
rail wear rates.
2. Selection of the optimal type of grease by laboratory
testing.
o 5-163 x
3. Selection of the best equipment (wayside, moving or a
combination) for lubricant application in-track.
4. In-track testing of the optimal lubricator setting, which
minimize rail wear and grease wastage.
5. Develop the formulae for the optimal location of
lubricators on a specific territory.
5.8.4
A System Approach for Managing the
Wheel/Rail Interface
If each of the disciplines responsible for maintaining the
vehicle and the track independently pursue optimization
actions for the wheel and the rail, these actions can counteract
any possible improvements and in some instances have
disastrous consequences. A good example is the introducing of
harder rails without taking the effect on the wheels into
account.31
The biggest challenge is to get the departments responsible
for the vehicle and the track to work together and define a
wheel rail interaction strategy to obtain synergy between the
vehicles, track and the long term profitability of the heavy haul
operation.
Referring to an existing operation: to put such a strategy
into place measures have to be introduced to define the
existing conditions from which standards can be developed to
guide maintenance procedures taking long term profitability
into account. These are:
1. Vehicle
•
Wheel profiles
•
Wheel wear limits
•
Wheel out of round and flats limits
•
Bogey alignment standards
•
Impact limits
2. Track
•
Rail gauge
o 5-164 x
•
Rail hardness
•
Rail profiles
•
Rail wear limits
3. Friction control standards
With these standards in place, the wheel and rail should be
corrected to the required profiles. The measures should then
be used in an integrated manner to achieve conformity
between the wheel and the rail and to move to a preventive
mode of maintenance.
An important part of this approach is the information
system that needs to be in place that can make the information
available to all the role players that need to manage the
wheel/rail interface.
5.9 Conclusion
Rail performance cannot be optimized without a consideration
of wheel profiles and bogie performance. The reverse is also
true. Optimizing wheel and rail performance means
optimizing the wheel/rail system. This involves:
•
Recognizing that lateral creep is the most damaging
mechanism for wheel and rail and taking steps to
improve curving.
•
Controlling the risk of rail failure through rail grinding
and testing so that rails will always be removed for
wear.
•
Managing the friction between wheel and rail through
the use of lubrication.
•
Regularly reprofiling rails to shapes that conform to
the wheel and reduce high stress contacts.
•
Avoiding wheel tread hollowing through reprofiling.
•
Maintaining gauge to less than 15 mm wide.
•
Using clean, harder rail steels where they make
economic sense.
o 5-165 x
Click Here To Go Back To Table of Contents
Acknowledgements
Most of the content of this chapter is based on course notes
(“Rail/Wheel Interaction and Metallurgy”) presented as part of the
Chair in the Railway Engineering Program at the University of
Pretoria. Additional inputs were also received from Peter Sroba and
Eric Magel, Canadian Centre for Surface Transportation Technology,
Johan Marais, Principal Engineer, Spoornet, Michael D. Tomas,
Senior Technologist, Spoornet, Daniel L. Magnus, KLD Labs
Incorporated, Robin Clark, Sperry Rail Incorporated, John Stanford,
Burlington Northern Santa Fe Railroad, Leon Zaayman, Product
Specialist, Plasserail, South Africa, and a number of IHHA and
related publications.
1.
References
Railroad Safety Statistics, Annual Report 1999, US
Department of Transportation, Federal Railroad
Administration, August 2000.
2.
D.D. Davis, M.J. Joerms, O. Orringer and R.K. Steele,
“The Economic Consequences of Rail Integrity,” in
proceedings of the Third International Heavy Haul
International Heavy Haul Association Conference,
Vancouver, Canada, 1986.
3.
R.K. Steele, “Overview of the FAST/HAL Rail
Performance Tests,” in proceedings of a Workshop on
Heavy Axle Loads, Pueblo, CO., 1990.
4.
P. Clayton, “Fatigue Behaviour of Rail Steels in a 33-kip
Wheel Load Experiment at FAST,” Bulletin 731 of the
American Railway Engineering Association, May, 1991.
5.
CP Rail System, “Standard Practice Circular 27, Rail
Testing,” Montreal, Canada, 1992.
6.
J. Igwemezie, S.L. Kennedy, X. Feng, and W. Rowan,
“Dynamic Rail Fracture Under Dynamic, Thermal and
Residual Stresses,” The 5th International Heavy Haul
International Heavy Haul Association Conference, Beijing,
China, June, 1993
7.
A.M. Zarembski, “Misreading Rail Flaw Size,” in RT&S,
March, 1986
o 5-166 x
8.
O. Orringer, et al, “Detail Fracture Growth in Rails:Test
Results,” Theoretical and Applied Mechanics, Volume 5,
No. 2, 1986.
9.
American Railway Engineering Association, “Manual for
Railway Engineering,” Chicago, IL, 1992.
10.
“Rail-Wheel Interaction and Metallurgy” Course, Chair In
Railway Engineering, University of Pretoria, South Africa,
1993.
11.
C. Esveld, L. Gronskov, “Rail Profile 2: Progress in
Wheel and Rail Measurement” The 6th International Heavy
Haul International Heavy Haul Association Conference,
Cape Town South Africa, 1997.
12.
D. L. Magnus, “Track Speed Rail and Wheel Inspection
Technology for Preventative Maintenance Planning,”
Conference on Railway Engineering, Rockhampton,
Queensland, Australia 6-9 September 1998.
13.
J. Cooper 1993 1 “Rail Flaw Detection: A Particular
Challenge,” The 5th International Heavy Haul International
Heavy Haul Association Conference, Beijing, China, June,
1993
14.
J. Kalousek, P.S. Sroba, C. Hegelund, 1989 1”Analysis of
Rail Grinding Tests and Implications for Corrective and
Preventative Grinding” The Institution of Engineers,
Australia National Conference Publication No. 89/13,
Brisbane Australia.
15.
S. Linn D. Abell, J. Kalousek, 1993 1 “Planning of
Production rail Grinding on the Burlington Northern
Railroad,” The 5th International Heavy Haul Association
Conference, Beijing, China, June, 1993.
16.
J. Stanford, P. S. Sroba, E. Magel, “Burlington Northern
Santa Fe Preventive – Gradual Grinding Initiative”
AREMA, Chicago, IL September 1999.
17.
R. Mitchell, P. J. Stewart, P.S. Sroba, 1989 1 “Rail Grinding
from a Contractors and Operators Perspective” The
institution of Engineers, Australia National Conference
Publication No. 89/13, Brisbane Australia.
o 5-167 x
18.
H. Höne “Rectification of Rail Profiles on the SishenSaldanha Iron-Ore Export Line with the Rail Planing
Machine,” International Heavy Haul Association, Special
Technical Session, Moscow Russia, July 1999.
19.
G. S. Hamilton,., “Alternate Means of Re-Profiling Freight
Car Wheels,” Transportation Technology Center, Inc.,
Pueblo, Colorado, April 2000.
20.
J. Kalousek and E. Magel, “Managing Rail Resources,”
AREA, Vol. 98, Bulletin 760, May 1997, pp. 139-148
21.
R. Runyon, “Recent Developments in Top-of-Rail
Lubrication,” Advanced Rail Management’s Wheel/Rail
Interface Seminar, Chicago, May 4-5, 1999
22.
R. Reiff and S. Gage, "Evaluation of Three Top of Rail
Lubrication Systems,” TTCI report No. R-936, December
1999
23.
D.T. Eadie. J. Kalousek and K. Chiddick, "The role of
high positive friction (HPF) modifier in the control of
short pitch corrugations and related phenomena,”
Proceedings of Contact Mechanics and Wear of
Rail/Wheel Systems, 5th International Conference, Tokyo
July 2000, p. 42.
24.
S. Gage and R. Reiff, "Evaluation of Century Oil
Lubrication Products,” TTCI report P-91-107, July 1991.
25.
J.DeKoker., “Rail and Wheel Flange Lubrication” Read to
South African Permanent Way Institute, Oct 1993.
26.
R.Reiff and D.Creggor “ Systems Approach to Best
Practice for Wheel and Rail Friction Control” International
Heavy Haul Conference 1999
27.
J.Rucinski, J.Powell “Assessment of Wheel and Rail
Lubrication Strategies at Queensland Rail”
28.
AAR Annual Research Review 1998 and 2000. Pueblo
Colorado
29.
J.DeKoker., "Development of a Formulae to Place Rail
Lubricators,” Fifth International Tribology Conference,
27-29 September, 94.
o 5-168 x
30.
H. M. Tournay, “Rail/Wheel Interaction from a Track and
Vehicle Design Perspective,” International Heavy Haul
Association, Special Technical Session, Moscow Russia,
July 1999.
31.
A. Durham, “Case Study: The Coal Line Wheel and Rail
Interaction Strategy,” International Heavy Haul
Association, Special Technical Session, Moscow Russia,
July 1999.
o 5-169 x
Click Here To Go Back To Table of Contents
GLOSSARY
Association of American Railroads (AAR) An industry
association whose responsibilities include safety standards
(including design standards and approval), maintenance,
operations, service and repair standards, and car service rules.
AAR Manual of Standards and Recommended Practices
(MSRP) Publication containing the technical specifications
Acceleration Rate of change of speed miles per hour (change)
per second or miles per hour per minute.
Adhesion Coefficient of friction between wheel and rail for
acceleration and retardation. When this force is exceeded,
wheel slipping or sliding takes place.
Adhesion Coefficient The percent of the total weight on the
driving wheels of a locomotive that is available for traction. It is
largely dependent on the condition of the rail, and can vary from
a low of 10% (.10) on wet rail to a high of 40% (.40) on dry
sanded rail. Average coefficient of adhesion is about 0.25.
Adhesion Limited Speed A speed at which adhesion (friction)
between wheel and rail limits the acceleration possible from the
available locomotive tractive effort horsepower. Attempting
greater acceleration causes the locomotive wheels to slip.
Adhesion (of Drivers) A measure of the ability of locomotive
driving wheels to accept rotational force without slipping on rails,
usually expressed as a percent of the total weight on the drivers.
Alignment The position of track in the horizontal plane
expressed as tangent or curve.
Alloy Steel Steel with added silicon, manganese, nickel, or
other elements to give greater strength, or to impart other
desirable properties for a particular use.
o G-1 x
American Railway Engineering & Maintenance-of-Way
Association (AREMA) Professional organization whose
membership is comprised of Railroad Maintenance-of-Way
officials. The AREMA develops and establishes material
Specifications and Track construction standards.
Ampere The fundamental unit of measure for electric current.
One ampere is defined as the current the flows when a potential
of one volt is impressed on a resistance of ohm.
(Ohm is a unit of electric resistance. One ohm is equal to that
resistance required to cause a one volt drop in potential when
the current is one amp.)
Anchor, Rail A device installed on the rail base preventing
longitudinal rail movement and build-up of axial force. Anchors
have been available in a variety of designs throughout the history
of the rail industry. Rail anchors available today include the
following: Unit Rail Anchor Co. (Unit Spring Anchor and Unit-IV
Drive-On), Portec Inc. (Improved Fair), Woodings-Verona
Transportation Products (Woodings). True Temper Railway
Appliances Inc. (Chenneloc and Trueloc), and Rails Co.
(Compression Rail Anchor).
Axle The steel shaft on which the car wheels are mounted. The
axle holds the wheels to gage and transmits the load from the
journal bearing to the wheels.
Axle Seat The cylindrical surface of a car wheel which comes in
contact with the axle (also called “the wheel bore”). The
corresponding part of an axle is called “the wheel seat.” Both
surfaces are critical for a proper wheel fit on the axle.
Axial Force The lengthwise force resulting from train movement
and/or thermal contraction/expansion in rail.
Balance Speed The speed of a train on a curve when the wheel
loads are evenly divided between both rails. Also see
superelevation, balanced.
Ballast Material selected for placement on the roadbed to hold
the track in position, distribute weight, dissipate force, and
provide drainage.
o G-2 x
Ballast Depth The vertical depth of ballast between the bottom
of a tie and the sub-ballast.
Ballast Section The cross-section profile of ballast in a track.
Ballast, Broken Stone Any ballast consisting of crushed quarry
stone.
Ballast, Cemented Ballast that has lost all drainage capability
due to hardened fines or mud between stones.
Ballast, Crushed Rock See ballast, broken stone.
Bogie The running gear of a highway semi-trailer which may be
removable or longitudinally adjustable. Also, the European
railway term applied to railway freight and passenger car trucks.
Brake The whole combination of parts by which the motion of
the locomotive car or train is retarded or arrested.
Brake Shoe A block of friction material formed to fit the curved
surface of the tread of a wheel, and riveted or otherwise bonded
to a steel backing plate having provision for quick and positive
securement to the brake head.
Braking Force The pressure of the shoe against the wheel.
Braking Power A term used to describe the ability of a car to
stop during a brake application.
Braking Ratio The relation of the weight of the car or
locomotive to the braking force by the weight of the car or
locomotive.
Brinell Hardness (Bhn) The numerical expression of a metal's
hardness; the harder the metal, the higher the number. For
example, standard or plain steel rail has a Bhn of approximately
260.
Broken Rail Term commonly used to describe any rail defect
rendering a rail unfit for normal operation.
o G-3 x
Buckled Track A short length of track that is radically out of its
desired alignment. This track defect usually occurs at locations
with continuous welded rail and is caused by sub-standard
conditions or deficiencies coupled with high rail temperatures,
high axial forces, and the dynamic loads of moving trains. See
sun kink.
Buff A term used to describe compressive coupler forces.
Cant To lean or tilt an object, such as a rail, slightly from level.
Cant is usually expressed as a rate of inclination, such as 1 in
40, etc.
Capacity As applied to a freight car, the nominal load in pounds
or gallons which the car is designed to carry.
Car Body The main or principal part in or on which the load is
placed.
Carbon Steel Steels alloyed with carbon, manganese and
silicon; the properties of which are due essentially to the
percentage of carbon in the steel.
Cast Steel Wheel A railway wheel made by pouring molten
steel into a mold under well controlled conditions, followed by
appropriate cleaning and heat treating.
Center Sill The center longitudinal structural member of a car
underframe, which forms the backbone of the underframe and
transmits most of the buffing shocks from one end of the car to
the other.
Circuit (Electrical) A complete path of an electric current
including the generating device.
Coach Screw A spike with a round shank and threads.
Code (Rules) A general term used to describe any set of
regulations dealing with some specific subject, such as
interchange of freight cars or per diem.
o G-4 x
Code of Federal Regulations Regulations issued by various
branches and agencies of the federal government under the
authority of statues.
Coefficient of Friction The measure of friction in percentage,
between the brake shoe and the wheel.
Cold Working The process of rolling and shaping metal without
using heat to shape or increase hardness. The railhead
becomes hardened, after time, due to the passing of rolling stock
wheels.
Compression A general term used to describe forces which
have a tendency to squeeze together.
Compressive Strength The maximum compressive stress
which a material is capable of sustaining without permanent
deformation.
Compression of a Train The bunching of cars in a train caused
by run-in of slack from the rear end.
Contact Patch The band of contact between a wheel tread and
the rail tread.
Contaminant Any physical, chemical, biological, or radiological
substance or matter that has an adverse affect on air, water, or
soil.
Continuous Welded Rail (CWR) Rail lengths welded end to
end into rail strings providing a track without rail joints; also
called welded rail or ribbon rail.
Controlled Cooled The process of eliminating hydrogen gas in
steel by controlling the cooling rate of hot steel.
Corrosion The deterioration or eating away of the surface of
metal through chemical action.
o G-5 x
Corrugated Rail A rail flaw consisting of the wave-like wearing
of the rail tread visualized as peaks and valleys. There are many
causes of this condition, and it is generally accepted that each
location experiencing it has a unique cause. Short-wave
corrugation or roaring rail or simply corrugations, has a
wavelength of 1 to 3 inches. This type of condition is common
on heavy rail transit, light rail transit, and high speed passenger
operations. Short-wave, or intermediate wave corrugation, or
undulations, has a wavelength of 3 to 24 inches. This type of
condition is common on heavy freight operations. Very long
wave or long wave corrugations, has a wavelength greater than
24 inches and is common on very high speed operations.
Furthermore, this type of corrugation has a very shallow depth
between peaks.
Corrugation A wave-like rail tread caused by uneven rail wear.
Coupler A device located at both ends of all cars and
locomotives in a standard location to provide a means for
connecting a locomotive units together, for coupling cars
together or for coupling cars together to make up a train. The
standard AAR coupler uses a pivoting knuckle and an internal
mechanism that automatically locks when the knuckle is pushed
closed, either manually or by a mating coupler. A manual
operation is necessary to uncouple two cars whose couplers are
locked together.
Crack Separation of material extending partially but not
necessarily completely through the cross section of the plate.
Cross Level The vertical relation between the top of the two
rails of a track.
Cross tie See tie.
Cut Spike A spike consisting of steel nail-like device. The cut
spike has a square shank and a chisel end with the point
perpendicular to the wood fibers, thereby reducing splitting of the
tie during driving of the spike. The head of the spike hooks over
the rail base. In North American, the cut spike is the most
common type of rail fastener in use.
o G-6 x
Damaged Rail A rail damaged or broken due to a derailment,
improper handling, or other causes.
Dead Load In car design calculations, the weight of the carbody
with all attachments and appurtenances that will be supported by
the trucks. See Live Load.
Density Weight per unit of volume, generally expressed as
pounds per cubic foot or pounds per gallon in the English
system; or kilograms per cubic meter, or kilograms per liter, in
the metric system.
Derailment Anytime the wheels of a car or engine come off the
head of the rail.
Detail Fracture A rail defect consisting of a fracture of the
railhead caused by surface imperfections. This condition usually
arises from shells, head checks, flaking, or welded bond wire
connections.
Direct Fixation Method of attaching rail directly to bridges or to
concrete slabs without using ties or ballast. See Slab track.
Resilient fasteners are also commonly called direct fixation.
Draft A term used to describe forces resulting in tension in the
coupler shank. The term "draft" means the opposite of the term
"bluff."
Draft Gear A term used to describe the energy-absorbing
component of the draft system. The draft gear is installed in a
yoke which is connected to the coupler shank and is fitted with
follower blocks which contact the draft lugs on the car center sill.
So-called "standard" draft gear use rubber and/or friction
components to provide energy absorption, while "hydraulic" draft
gear use a closed hydraulic system consisting of small ports and
a piston to achieve a greater energy-absorbing capability.
Hydraulic draft gear assemblies are generally called "cushioning
units."
o G-7 x
Draft System The term used to describe the arrangement on a
car for transmitting coupler forces to the center sill. On standard
draft gear cars, the draft system includes the coupler, yoke, draft
gear, follower, draft key, draft lugs and draft sill. On cushioned
cars, either hydraulic end-of-car cushion units and their
attachments replace the draft gear and yoke at each end; or a
hydraulically controlled sliding center sill is installed as an
integral part of the car underframe.
Drawbar The mechanism for coupling together cars and
locomotive units. A term formerly used synonymously with
coupler. It has been used indiscriminately to designate both the
old link and pin drawbar and the modern automatic car coupler.
Drawbar Force The force exerted through the couplers by the
locomotive on coupled cars, by one car upon another, etc. This
force is usually greatest at that coupler between the last
locomotive unit and the first coupled car in the train.
Drawbar Pull The tensile coupler force. Locomotive pulling
power is sometimes expressed in terms of "pounds of drawbar
pull."
Empty Weight see light weight.
False Flange The flange on the overhanging portion of a wheel
tread; it is caused by wear on the wheel tread in the area where
the wheel and rail normally make contact.
Fastener, Elastic See fastener, resilient.
Fastener, Rail A general term describing the method of
attaching the rail to the tie or tack bed.
Fastener, Resilient Any type of rail fastener other than cut
spikes that provide a more positive connection between the rail
and tie or slab track. A variety of proprietary designs are
currently available from many manufacturers. Two common
examples of resilient fasteners are the Lineloc Rail Fastener and
the Pandrol Clip.
Field Weld See weld, thermite.
o G-8 x
Flaking A rail flaw consisting of the gouging of metal on the
railhead; it is indicated by small chipping and cavities.
Flange (1) The portion of a wheel that protrudes down from the
rail tread to guide rolling stock along a track; (2) one side of a rail
base; (3) a projecting edge of any structural object, for
strengthening, guiding, or securing.
Frog The portion of a turnout or track crossing where wheels
cross from one track to another; named because of its
resemblance to a frog (animal).
Frog Number The ratio between the theoretical frog heel length
and frog heel spread of a frog or half the cotangent of half the
frog angle.
Frog, Bolted Rigid A frog made of tee rails milled and fitted to
form an assembly held together with frog bolts and filler blocks.
Frog, Cast Manganese A frog consisting entirely of cast
manganese steel.
Frog, Movable Point (MPF) A frog with movable rails at a
shallow angle which form a continuous path. The movable point
frog is used in diamond crossings and slip switches and/or high
tonnage routes.
Frog, Rail Bound Manganese A frog with a manganese
casting fitted between and into tee rails and held together with
frog bolts.
Frog, Spring A frog without fillers between the frog point and
one wing rail, and with springs holding the wing rail up against
the frog point. Main track traffic travels on the side of the frog
with the uninterrupted surface for the passage of wheels. The
diverging traffic opens the sprung wing rail when each wheel
passes. Spring frogs are right and left-hand depending on which
track requires the unbroken path.
Frog, Swing Nose A frog in a turnout with a movable frog point
connected to a switch machine for positioning relative to the
switch position. The types of swing nose frogs include Welded
V, Bolted V, Cast V, or Forged V.
o G-9 x
Gauge In general terms, any device used for measuring an
independent quantity such as pressure rate of flow, volume,
length, area, etc. An instrument with a calibrated scale or dial for
measuring or indicating quality. An analog typed readout.
Gauge Corner The edge of a railhead on the gauge side.
Gauge Line The spot on the side of the railhead 5/8 inch below
the rail tread, where track gauge is established. Gauge lines
other than 5/8 inch are found on light rail transit.
Gauge, Track Measured at right angles, the distance between
running rails of a track at the gauge lines.
Gauge, Wide Any track gauge greater than a nominal design
standard as a result of improper installation or track component
deterioration.
Geometry Car A car equipped with electro-mechanical sensors
used to automatically detect and record track geometry over long
distances. The geometry car may be either self-propelled or
pulled by a locomotive.
Grade, Percent of The rise or fall of track over a distance of
one hundred feet. For example, a rise of one foot in 100 feet
equals one percent.
Gradient The difference in pounds per square inch between
brake pipe pressure on the locomotive and maximum obtainable
on the rear of the train. The direct result of brake pipe leakage.
The rate of inclination of track in relation to the horizontal.
Hard Conversions When products are hard converted, the
design of the item in SI metric units result in a physical change
and the produce is not interchangeable with earlier products built
to conventional unit design.
Head Checks A rail flaw consisting of shallow surface cracks in
the railhead usually found on the gauge corner. Head checks
generally run at a 45 degree angle to the axis of the rail; they
usually occur on the high rail of curves.
o G-10 x
Head-Hardened Rail A rail with only the railhead heat treated to
provide a harder steel for locations of extreme service, such as
curves.
Heat -Treated Heating and cooling a metal or alloy in such a
way as to obtain desired conditions or properties. Heating for the
sole purpose of not working is excluded from the meaning of this
definition.
Heat Treatment The process of altering the properties of a
material, usually steel, by specific heating and cooling
operations. Heat treated track components are good in locations
requiring high strength and durability.
Joint The junction of members or the edges of members that
are to be joined or have been joined.
Journal That part of an axle or shaft on which the journal
bearing rests or a roller bearing is applied.
Journal Bearing The general term used to describe the load
bearing arrangement at the ends of each axle of a railcar truck.
So called plain journal bearings are block of metal, usually brass
or bronze, shaped to fit the curved surface of the axle journal,
and resting directly upon it with lubrication provided by free oil
contained in the journal box. Journal roller bearings are sealed
assemblies of rollers, races, cups and cones pressed onto axle
journals and generally lubricated with grease. Vertical loads are
transferred from the journal bearing to the truck side frame
through the journal bearing wedge (in plain bearing designs), or
through the roller bearing adaptor in roller bearing trucks.
Lateral Force Any sideways force in the track.
Lateral Motion Sideways movement of a railcar and/or its
components, resulting in large measure from dimensional
clearances between parts of the truck assembly. Excessive
lateral motion in truck assemblies is a major cause of premature
wear of the truck and car body components.
o G-11 x
Light Weight Empty weight or tare weight (of cars or of the
train). The empty weight of a railroad car or of train including its
trucks and any other appurtenances considered standard to the
car. The light weight is stenciled on every freight car in
conjunction with the capacity and load limit stenciling, and is
abbreviated Lt. Wt.
Limit Gauge A term applied to many forms of gauges which are
used for determining whether pieces exceed or fall below a
certain specified range of dimensions. Limit gauges are
sometimes called "Go-No-Go" gauges.
Live Load In car design, the live load is the load imposed on the
car structure by outside forces such as the lading and any other
specified supplemental loads such as accelerations due to
vertical irregularities in the track structure.
Load Limit The maximum weight of lading that can be loaded in
a car. For cars meeting standard AAR design criteria, the load
limit is equal to the maximum allowable gross weight on rails
(determined by axle and wheel size) less the light weight of the
car.
Lubricant Any liquid or grease employed to coat a surface upon
which another surface rotates or slides in order to reduce the
friction.
Lubricator, Rail A device (mechanical or hydrostatic) used to
supply oil to parts of the compressor and compressor governor
under pressure and to reduce friction between wheel flanges and
the railhead.
L/V Ratio The L/V ratio is defined as the ratio of the lateral force
to the vertical force of a car or locomotive wheel on a rail. It is an
important factor affecting the tendency to turn over under load,
and is often a point of discussion in evaluating the cause of a
train derailment. When the lateral force is greater than the
vertical force, wheel climb is imminent. When the ratio is
approximately .64, unrestrained rail can turn over. An L/V ration
of 1.29 may cause a wheel to climb new rail. Note: Studies by
the AAR show there are some situations where an L/V of about
0.29 can turn the rail over.
o G-12 x
Magnetic Particle Testing A non-destructive test method for
identifying cracks or discontinuities in castings or machined
parts.
Maximum Gross Weight on Rails For a single car, the
maximum permissible weight of both car and lading permitted for
operation in unrestricted interchange service.
Meter One of the standard length measurements in the metric
system, equivalent to 39.368 inches in the English system.
Metric System A decimal system for measuring length,
capacity, surface and weight, using the meter as the unit of
length, the liter as the unit of volume, and the gram as the unit of
weight.
MGT An abbreviation for Million Gross Tonnes, MGT is a
measurement of track and rail wear. It is the number of tonnes
of traffic load, expressed in millions, that have passed over a
given section of line in a specified time span. Usually, MGT is
calculated on an annual basis.
Modulus of Track The vertical stiffness of a track. Measured
by instrumented measurements, modulus is the amount of
vertical deflection under a train.
Movable Point Frog (MPF) A frog with movable rails at a
shallow angle which form a continuous path. The movable point
frog is used in diamond crossings and slip switches and/or high
tonnage routes.
Multiple-Wear Wheel A steel railway wheel made with sufficient
original rim thickness to permit turning ufll flange and tread
contours at least twice during the life of the wheel.
Narrow Gauge Railroads built to less than the standard 4’ 81/2”
between rails gauge.
Net Force A force which causes an object such as a train to
accelerate. Total force applied to the train less drag forces.
Net Weight The weight of only the contents of the car.
o G-13 x
NFL Bearing A factory lubricated journal roller bearing
assembly made with superior seals and requiring no field
lubrication during its normal service life. NFL bearings can be
identified by the absence of grease fitting in the end cap.
One-Wear Wheel A steel railway wheel designed with a rim
thickness such that full flange and tread contour cannot be
restored by turning.
Plain Bearing As distinguished from a journal roller bearing; a
journal bearing arrangement whereby a brass or bronze bearing
is held in place against a polished axle journal, and lubricated by
free oil in a journal box fed to the bearing by a lubricating device.
Preferred Rail Laying Temperature (PRLT) The optimum
temperature at which continuous welded rail is installed and
anchored to reduce thermal stresses when hot or cold.
Pressure A unit force generally measured in terms of pounds
per dquare inch (or kilograms per square centimeter) created by
the action of a compressed gas or fluid in a confined space.
Preventive Maintenance Inspection to discover if something
needs repairing before it fails and performing the necessary work
in order to stop or slow that failure.
Profile (1) A graph of a longitudinal length of a track depicting
humps and dips. (2) The surface uniformity of a rail measured at
the mid-point of a chord.
Quenching To quickly cool a hot metal to produce a hardened
surface. Oil quenching is a common treatment for track
materials, such as track bolts and rail.
Rail As used in car construction, any horizontal member of a car
super-structure. The term is usually used in combination with
some additional identifying word such as "belt rail" or "hand rail."
As used in track, a rolled steel shape, commonly a T-section,
designed to be laid end to end in two parallel lines on cross ties
or other suitable supports a form a track for railway rolling stock.
Auxiliary rails include guardrails and third rails.
o G-14 x
Rail Anchor A device installed on the rail base preventing
longitudinal rail movement and build-up of axial force. Anchors
have been available in a variety of designs throughout the history
of the rail industry. Rail anchors available today include the
following: Unit Rail Anchor Co. (Unit Spring Anchor and Unit-IV
Drive-On), Portec Inc. (Improved Fair), Woodings-Verona
Transportation Products (Woodings), True Temper Railway
Appliances Inc. (Channeloc and Trueloc), and Rails Co.
(Compression Rail Anchor).
Rail Clip Generic term used to name various rail hold-down
devices other than common spikes. There are many proprietary
designs of rail clips currently available. See fastener, resilient.
Rail Creep The occasional lengthwise movement of rails in
track. Rail creep is caused by the movement of trains or
temperature changes. It is common practice to stop the effect of
creeping by the use of rail anchors or resilient fasteners.
Rail Defect Rail condition consisting of a complete fracture or
sufficient fissures that may render the rail unfit for normal
operation. See rail flaw.
Rail Flaw Imperfections in the surface or interior of the rail.
Imperfections are not considered dangerous in themselves, but
can propagate into rail defects.
Rail Grinding The removal of surface metal on the railhead.
The removal of metal occurs by use of production equipment
with rotary abrasive grinding stones.
Rail Grinding, Corrective Rail grinding that eliminates incipient
cracks, shells, engine burns, and corrugation.
Rail Grinding, Maintenance Cyclical rail grinding to eliminate
incipient cracks, shells, engine burns, and corrugation.
Rail Grinding, Profile Rail grinding that reshapes the railhead
to a desired contour to optimize the contact patch.
o G-15 x
Rail, Heavy An electric railway with the capacity for a "heavy
volume" of traffic and characterized by exclusive rights-of-way,
multi-car trains, high-speed and rapid acceleration, sophisticated
signaling and high platform loading. Also known as "rapid rail,"
"subway," "elevated (rail-way)," or "metropolitan railway (metro)."
Rail Lip The overhanging metal at the corner of the railhead. A
rail lip results from the metal flow of steel that occurs on the
railhead.
Rail Lubrication Any type of lubricant placed on the gauge line
to reduce the friction between the railhead and wheel tread.
Rail Lubricator A device designed to apply grease to the gauge
side of the rail head at the beginning of a curve in order to
minimize wear of the rail and wheel flange, or to eliminate noise.
Rail Neutral Axis The point in the rail web in which internal
pressure is compressive (pushing) above and tensile (pulling)
below during vertical loading of the rail.
Rail Neutral Temperature (RNT) The temperature when there
is no axial force in the rail.
Rail Neutral Temperature Shift The difference between an
existing rail neutral temperature and its original or adjusted
neutral temperature. A downward shift to cooler temperatures is
common due to maintenance activities and train movements.
Rail Tread The top portion of the railhead where rail/wheel tread
contact occurs. Also called running surface.
Rail Wear A rail flaw consisting of reduction of the railhead as a
result of abrasive action between the steel wheel on the steel
rail. Wear on the top of the railhead — caused by the wheel
tread — is called "top wear," "tread wear," "vertical wear," or
"head height loss." Wear on the side of the rail — caused by
wheel flanges — is called "side wear," "gauge wear," "horizontal
wear," or "gauge face loss."
Rail Web The vertical member of a rail that provides bridge or
beam strength to carry the rolling stock loads from tie to tie.
o G-16 x
Rail, Alloy Rail containing special metal elements for increasing
the hardness of rail. Alloy rails are used in locations of extreme
service.
Rail, carbon Rail made of steel containing 10 points of carbon.
Rail, Continuous Cast Rail rolled from steel manufactured by a
process in which molten steel is drawn without interruption from
a special casting machine.
Rail, Continuous Welded (CWR) Rail lengths welded end to
end into rail strings providing a track without rail joints. It is also
called welded rail or ribbon rail.
Rail, Controlled Cooled Rail with steel having hydrogen gas
eliminated by a controlled process of temperature reduction.
After rolling the steel into the rail shape, the hot rail is reduced to
air temperature over a specified time. The controlled cooled
process was introduced in 1936.
Rail, Corrugated A rail flaw consisting of the wave-like wearing
of the rail tread visualized as peaks and valleys. There are many
causes of this condition, and it is generally accepted that each
location experiencing it has a unique cause. Short-wave
corrugation or roaring rail or simply corrugations, has a
wavelength of 1 to 3 inches. This type of condition is common
on heavy rail transit, light rail transit, and high speed passenger
operations. Short-wave, or intermediate wave corrugation, or
undulations, has a wavelength of 3 to 24 inches. This type of
condition is common on heavy freight operations. Very long
wave or long wave corrugations, has a wavelength greater than
24 inches and is common on very high speed operations.
Furthermore, this type of corrugation has a very shallow depth
between peaks.
Rail, Head-Hardened A rail with only the railhead heat treated
to provide a harder steel for locations of extreme service, such
as curves.
Rail, Heat Treated Rail that has been hardened by a heating
Rail, High Carbon A rail with extra carbon added to the steel
during the manufacturing process to increase its hardness.
o G-17 x
Rail, Work-Hardened Rail that has a hardness greater than
when manufactured, as a result of the cold working of the steel
by repeated traffic loading.
Railhead (1) The top of the rail in which rolling stock wheels
are guided. The railhead also accepts the weight from rolling
stock in a very small area at each wheel/rail contact oint. (2) The
end of a railroad line.
Railroad The entire system of track together with the stations,
land, rolling stock, and other property used in rail transportation.
All forms of non-highway ground transportation that run on rails
or electromagnetic guideways, including (1) commuter or other
short-haul rail passenger service in a metropolitan or suburban
area, and (2) high-speed ground transportation systems that
connect metropolitan area, without regard to whether they use
new technologies not associated with traditional railroads. Such
term does not include rapid transit operations within an urban
area that are not connected to the railroad system of
transportation.
Raised-Wheel Seat Axle See Axle.
Regenerative Braking The retardation system on electric cars
or locomotives which can return power developed by traction
motors acting as generators to the third rail or catenary for use
by other units.
Resilient Fastener Any type of rail fastener other than cut
spikes that provide a more positive connection between the rail
and tie or slab track. A variety of proprietary designs are
currently available from many manufacturers. Two common
examples of resilient fasteners are the Lineloc Rail Fastener and
the Pandrol Clip.
Resistance In general, resistance denotes opposition to
movement or flow. In mechanical systems, any force that
opposes motion such as friction or action of a spring could be
termed as resistance. In electrical circuits, resistance is
opposition to the flow of current, and is measured in units called
ohms.
o G-18 x
Rim On a railway car wheel, that portion around the outer
circumference that forms the edge of the tread. The thickness of
the rim is a measure of the amount of wear remaining in the
wheel, and when this dimension reaches a given limit (as
measured with the AAR steel wheel gauge), the wheel must be
scrapped.
Rim Quench Following heat treatments, all rims are quenched
to harden wheel treads.
Rollability A term generally applied in classification yards
pertaining to the characteristics of individual cars and their ability
to roll.
Roller Bearing The general term applied to any group of journal
bearings that employ hardened steel rollers to reduce rotational
friction. Roller bearings are sealed assemblies that are
mechanically pressed onto an axle, and transfer the wheel loads
to the truck side frames through a device known as a roller
bearing adapter that fits between the bearing outer ring and the
side frame pedestal.
Roller Bearing Adapter A casting that fits between a freight car
roller bearing and the truck side frame to transfer the load from
the side frame to the bearing.
Roller Side Bearing A side bearing fitted with rollers to reduce
the friction in curving. See Side Bearing.
Rolling Equipment Includes locomotives, railroad cars, and
one or more locomotives coupled to one or more cars.
Rolling Radius Differential The different radius contact points
between the wheel tread and rail tread on the low rail vs. the
high rail, accomplished by tapered wheels. When in curves, the
wheel flange on the high rail is up against the gauge line, with
the wheel flange on the low rail pulled away from the low rail
gauge line. This action results in a longer radius contact point on
the wheel contacting the high rail, thereby inducing a steering
affect of wheel sets through curves. In addition, wheel wear and
rail wear is minimized due to a reduction in wheel slip.
o G-19 x
Rolling Stock (1) Any on-track wheeled equipment. (2) A
general term used when referring collectively to a large group of
railway cars. The vehicles used in a transit system, including
buses and rail cars. Includes locomotives, railroad cars that
carry commodities and passengers, and one or more
locomotives coupled to one or more cars.
Running Gear A general term used to describe the group of
parts whose functions are related to movement of the car.
Running gear includes the wheels, axles bearings, suspension
system and other components of the trucks.
Running Surface The top portion of the railhead where
rail/wheel tread contact occurs. Also called rail tread.
Screw Spike A spike with a round shank and threads.
Section Modulus The bending strength of a particular rail
section.
Shatter Crack Minute shallow cracks that occur in steel after
exposure to the intense heat, which occur during torch cutting or
locomotive wheel slipping.
Shelling A wheel defect characterized by pieces of metal flaking
out of the tread surface, and caused by fatigue failure of the
metal in the tread. Very shallow shells are called spalls.
Side Bearing A loaded bearing component, located either on
the truck or body bolster, and arranged to absorb vertical loads
arising from the rocking motion of the car. There are various
types of side bearings ranging from simple flat pads to complex
devices which maintain constant contact between the truck
bolster and car body. See Truck Side Bearing.
Side Bearing Roller A solid steel cylindrical roller which fits
loosely in a rectangular retainer or "cage," fastened to the truck
bolster, and contacts the body side bearings during lateral
rocking of the car. The rollers are used either singly or in pairs,
and have the advantage of decreasing resistance to truck
rotation when the car is rounding a curve.
o G-20 x
Side Bearing Truck A truck in which the weight of the car is
transmitted at the side instead of the center.
Side Frame In the conventional three-piece truck, the heavy
cast steel side member which is designed to transmit vertical
loads from the wheels through either journal boxes or pedestals
to the truck bolster.
Side Frame Key A short steel retainer bolted to the bottom of a
pedestal type side frame to prevent roller bearing assemblies
from becoming dislodged from the side frame pedestals.
Sometimes called a roller bearing key.
Slab Track Track constructed without ties using a concrete
base. The rails are connected to the concrete with direct fixation
fasteners.
Slid Flat Wheels Railcar wheels with flat spots resulting from
sliding on the rail; generally, due to over braking or failure to
release the car handbrake. Flat spots larger than the AAR
maximum allowed are a cause for wheel replacement.
Slide Wheel sliding is where the wheel does not rotate on its
axle and motion exists between the wheel and the rail.
Speed, Maximum Authorized The highest speed trains may
travel over a given segment of railroad as established by
operating rules and timetable special instructions, unless
otherwise restricted. Sometimes referred to as normal speed.
Speed, Overbalance The speed of rolling stock on a curve at
which wheel loads place more force on the low rail than the high
rail. This occurs when trains are operating slowly on a curve
with superelevation, or there is too much superelevation in a
curve.
Speed, Underbalance The speed of rolling stock on a curve at
which wheel loads place more force on the high rail than the low
rail. This occurs when trains are operating faster than balance
speed. Curves often are intentionally under-elevated to force the
wheel up against the high rail to achieve a smooth ride, or in high
speed track where a high amount of superelevation would be
undesirable.
o G-21 x
Spike Any device that fastens a rail, tie plate, switch stand or
other object to a tie.
Spike, Cut A spike consisting of steel nail-like device. The cut
spike has a square shank and a chisel end with the point
perpendicular to the wood fibers, thereby reducing splitting of the
tie during driving of the spike. The head of the spike hooks over
the rail base. In North America, the cut spike is the most
common type of spike in use.
Spike, Drive A spike consisting of round nail-like device with
fluted (thread-like) sides. The drive spike is driven into predrilled holes and used to fasten a variety of items to the track,
such as guard logs, spacing straps, etc. Also called drive lag.
Spike, Screw A spike with a round shank and threads.
Spring A general term referring to a large group of mechanical
devices making use of the elastic properties of materials to
cushion loads or control motion.
Spring Frog A frog without fillers between the frog point and
one wing rail, and with springs holding the wing rail up against
the frog point. Main track traffic travels on the side of the frog
with the uninterrupted surface for the passage of wheels. The
diverging traffic opens the sprung wing rail when each wheel
passes. Spring frogs are right and left-hand depending on which
track requires the unbroken path.
Spring Group Any combination of standardized coil springs
used in each truck side frame, and selected to match car
capacities and obtain desired vertical suspension characteristics.
Cars are often stenciled to show the number of specific springs
of various designations, e.g., 5 D5 outer 3 D5 inner, that make
up the spring group standard to the car.
Spring Nest Two or more coil springs of different diameters,
one fitting inside the other and acting in combination. Truck
springs are commonly made up of standardized outer and inner
coils, nested and arranged in various spring groups.
o G-22 x
Spring Plank A steel plate fitting under each end of the truck
bolster on older trucks to provide a bearing surface for the spring
group.
Spring Seat A cup-shaped piece of cast steel, or cast wrought
iron, on which the bottom of a spring rest. Also called "spring
plate." They are further distinguished by the name of the spring
for which they serve such as bolster spring seat, equalizer spring
seat.
Spring Switch A switch in which the switch rails are connected
to a spring mechanism allowing trains to make a trailing
movement with the switch set in the improper position, thereby
forcing the switch rails back to the original position after the
passage of each wheel. Also called Vaughn Spring Switch,
invented by D.F. Vaughn of the West Jersey and Seashore
Railroad, circa 1880.
Standard Gauge The standard distance between rails of North
American railroads, being 4 feet 8.5 inches measured between
the inside faces of the rail heads.
Stress A term used in engineering to denote force per unit of
area in structural members, expressed commonly in units of
pounds per square inch, (psi). Stresses are classified by the
type of reaction they produce in the fibers of the material they
affect. Tensile stress tends to pull apart, compressive stress
tends to press together, and sheer stress tends to slide parallel
adjacent surfaces against each other in opposite directions.
Stress Relief Commonly used as a synonym for post weld heat
treatment.
Stress Relieving Heating to a suitable temperature, holding
long enough to reduce residual stresses, and the cooling slowly
enough to minimize the development of new residual stresses.
Surface (1) The geometrical condition of the track at the rail
tread over a given distance. Surface is best described as the
smoothness of the track over distances. (2) See Profile.
o G-23 x
Suspension The resilient system through which a car body is
supported on its wheels. Suspension systems involve the use of
hydraulic devices, friction elements and coil, elliptic, rubber or
pneumatic springs.
Swing Bolster A truck bolster suspended by hangers or links so
that it can swing laterally with relation to the truck and thus lower
the effects of lateral impact received through the side frames and
wheels. Trucks equipped with swing bolster are known as swing
motion trucks.
Swing Nose Frog A frog in a turnout with a movable frog point
connected to a switch machine for manipulation relative to the
switch position, swing nose frog types include Welded V, Bolted
V, Cast V, or Forged V.
Swing Hanger Bars or links, attached at their upper ends to the
frame of a swing motion truck, and carrying the spring plank at
their lower ends. Also called "bolster hanger."
Swing Motion Truck A truck with a bolster and spring plank
suspended on swing hangers so that they can swing laterally in
relation to the truck frame.
Switch The component of a turnout consisting of switch rails
and connecting parts providing a means for making a path over
which to transfer rolling stock and on-track equipment from one
track to another.
Tamp or Tamping The process of compacting ballast under ties
to provide proper load bearing.
Tare Weight The weight of the empty car. The sum of the
empty weights of the cars is the tare weight of the train.
Temperature Effect Change in temperature, produced by the
expansion or contraction of air due to variation in pressure.
Temperature, Preferred Rail Laying (PRLT) The optimum
temperature at which continuous welded rail is installed and
anchored to reduce thermal stresses when hot or cold.
o G-24 x
Temperature, Rail Neutral The temperature which there is no
axial force in the rail.
Testing, Non-Destructive Rail Any method of inspecting for
internal fissure in rail without physically damaging it.
Testing, Ultrasonic Rail The process of testing for fissures in
rail by passing ultra high frequency sound through the rail.
Sound waves that reflect off fissures are detected and measured
to determine the location and size of the fissure.
Thermal Cracking A wheel defect characterized by fine cracks
running transversely across the tread and caused by excessive
heat generated at the tread surface during heavy prolonged
braking. Undetected thermal cracks will propagate through the
flange or rim into the wheel plate and cause failure.
Thermite or Thermit Weld The process of welding the ends of
two rails together with a foundry-like process. The thermite
process consists of iron oxide and aluminum powder. This
material is ignited with a magnesium charge, and a reaction
occurs resulting in molten steel, which is poured between the rail
ends causing fusion. Also see continuous welded rail.
Tie The portion of the track structure generally placed
perpendicular to the rail to hold track gauge, distribute the weight
of the rails and rolling stock, and hold the track in surface and
alignment. The majority of ties are made from wood. Other
materials used in the manufacture of ties include concrete and
steel. Also called cross tie.
Tie Pad A material made from rubber or fiber placed between
the tie plate and tie to reduce abrasion, inhibit plate cutting, and
cushion load impact.
Tie Plate A metal plate placed between the rail base and tie to
distribute the weight of trains over a larger surface, thereby
reducing tie damage. Tie plates also give lateral stability to the
rail by restraining movement.
o G-25 x
Tie, Concrete A tie made of concrete with reinforcing rods. The
rods are pulled into tension during the curing cycle of the
concrete then cut when the concrete cures to place the tie into
compression.
Track An assembly of fixed location extending over distances to
guide rolling stock and accept the imposed dynamic and static
loads. See track structure.
Track Defect An anomaly in any part of the track structure
requiring repair or other action, such as a reduction in speed.
Track Inspection Car A passenger car used in regular consists
from which the track may be inspected but not necessarily
containing track geometry evaluation equipment.
Track Modulus The vertical stiffness of a track. Modulus is
determined by instrumented measurements of the amount of
vertical deflection under rolling stock loads.
Track Structure The railroad track comprised of rail, tie plates,
fasteners, cross ties, rail anchors, guardrails, ballast, and
subgrade.
Track-Train Dynamics The study of the motions and resulting
forces that occur during the movement of a train over a track
under varying conditions of speed, train makeup, track and
equipment conditions, grades, curves and train handling.
Track Twist Any change from point to point from the designated
cross level or superelevation in the track. Also called warp or
change.
Tractive Effort The force (at the wheel/rail interface) at a given
speed the locomotive exerts to move itself and any coupled cars.
Train An engine or more than one engine coupled, with or
without cars, displaying markers. For practical purposes, a train
is a group of coupled cars hauled by a locomotive.
Train Brake The automatic brake. The combined brakes on
locomotive and cars that provides the means of controlling the
speed and stopping of the entire train.
o G-26 x
Train Consist The composition of the complete train excluding
the locomotive. The cars in a train.
Train Mile The movement of a train a distance of one mile
measured by the distance between terminals and/or stations and
includes yard switching miles, train switching miles, and work
train miles. Yard switching miles may be computed on any
reasonable, supportable, and verifiable basis. In the event
actual mileage is not computable by other means, yard switching
miles may be computed at the rate of 6 mph for the time actually
engaged in yard switching service.
Train Resistance A force which resists or opposes movement
of a train. Resistance to motion along the track, attributed to
bearings, wind and air resistance, flange contact with rail, etc.
Transverse Defect A group of rail defects that have transverse
components, such as transverse fissures, compound fissures,
and detail fractures.
Transverse Fissure A rail defect consisting of a break in the
crosswise plain of the rail, beginning at or near the center of the
railhead. The fracture begins at a hydrogen inclusion and
spreads outward. The resulting break will usually show the
hydrogen inclusion as a dark spot in the center growing into an
oval or round smooth spot within the railhead.
Tread In car wheel nomenclature, the slightly tapered exterior
running surface of the wheel that comes in contact with the top
surface of the rail, and also serves as a brake drum on cars with
conventional brake arrangements. In rail nomenclature, the top
surface of the rail which contacts the vehicle.
Tread Brakes Those wherein the brake shoes bear against the
tread of the wheel when brakes are applied.
Truck The complete assembly of parts including wheels, axles,
bearings, side frames, bolster, brake rigging, springs and all
associated connecting components, the function of which is to
provide support, mobility and guidance to a railroad car.
Truck Bolster The main transverse member of a truck
assembly that transmits carbody loads to the side frames
o G-27 x
through the suspension systems. The ends of the bolster fit
loosely into the wide openings in the side frames and are
retained by the gibs, which contact the side frame column
guides. Truck bolster contact with the carbody is through the
truck center plate, which mates with the body center plate and
through the side bearings.
Truck Center Plate The circular area at the center of a truck
bolster, designed to accept the protruding body center plate and
provide the principal bearing surface for carbody support on the
truck bolster. Truck center plates are often fitted with a
horizontal wear plate and a vertical wear ring to improve wearing
characteristics and extend bolster life.
Truck Centers On a single car, the distance between the truck
center pins as measured along the center sill from the center line
of one body bolster to the center line of the other.
Truck Frame A structure made of cast steel in one piece, to
which the journal boxes or pedestals, springs and other parts are
attached, and which forms the skeleton of a truck. One piece
truck frames are not generally used for freight cars, but are often
found on locomotives and passenger cars.
Truck Hunting A lateral instability of a truck, generally occurring
at high speed, and characterized by one or both wheelsets
shifting from side to side with the flanges striking the rail. The
resulting motion of the car causes excessive wear in car and
truck components, and creates potentially unsafe operating
conditions. For freight vehicles, the phenomenon occurs
primarily with empty or lightly loaded cars with worn wheelsets.
Trucks, Radial Railroad freight car trucks whose wheelsets are
suspended in the truck frame in such a way as to allow the
wheelsets to align themselves to the radius of a curve in the
track. This is usually accomplished by low yaw constraints
between the wheelsets and truck frame, together with interconnections between the wheelsets for good hunting stability.
o G-28 x
Truck Side Bearing A plate, block roller or elastic unit fastened
to the top surface of a truck bolster on both sides of the center
plate, and functioning in conjunction with the body side bearing
to support the load of a moving car when variations in track cross
level cause the carbody to rock transversely on the center plates.
See Body Side Bearing, Side Bearing and Side Bearing Roller.
Truck Springs A general term used to describe any of the
several types of springs used in the suspension system of trucks
to provide a degree of vertical cushioning to the car and its load.
Truck Wheel Base The horizontal distance between the centers
of the first and last axles of a truck.
Underframe (1) The term used to refer to the entire structural
framework of the car below the floor, including the center sill,
side and end sills, bolster, cross members, stringers and other
attached components. (2) The supporting matrix in the
understructure. Generally the body bolsters, sills, platforms,
floor stringers, cross bearers and cross ties that support the
superstructure.
Vehicle Passenger car, freight car, locomotive, motor car, etc.;
a general name to include any of the various means of hauling or
the power units used on the railroad trains.
Vertical Bounce An instability at high speed where the vehicle
oscillates vertically on the suspension system.
Visual Inspection A non-destructive test method using ordinary
vision to detect surface imperfections in materials and welds.
Voltage Drop The decrease in voltage to a current carrying
conductor.
Weld A seam where two members are joined, formed by any of
several heating processes resulting in melting and fusing
together of the metal on either side of the joint, often with the
addition of filler metal to improve the properties of the joint.
Welding processes are extremely important in modern car
construction.
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Weld Bead A weld deposit resulting from a pass.
Weld, Electric Butt The process of fusing together rail ends by
use of electric current. The process consists of passing high
electric current between the rail ends, drawing an arc, and
thereby heating the steel to a plastic state. The heated rails are
forced together resulting in fusion. This type of rail welding is the
process used in modern welding or rail into continuous lengths.
Generally, this operation is performed at a central location with
long strings of rail transported to the construction site. Some
welding machines are attached to on-track equipment for on-site
welding. Also see continuous welded rail.
Weld, Thermite or Thermit A method process of welding the
ends of two rails together with a foundry-like process. The
thermite process consists of iron oxide and aluminum powder.
This material is ignited with a magnesium charge, and a reaction
occurs resulting in molten steel, which is poured between the rail
ends causing fusion. Also see continuous welded rail.
Welder A person who joins metal parts by the application of a
fusion process, either manually or with semi-automatic welding
equipment.
Welding The process of joining two members by the addition of
heat and filler metal.
Weldment An assembly whose component parts are joined by
welding.
Weld Pass A single progression of welding or surfacing along a
joint or substrate. The result of a pass is a weld bead, layer or
spray deposit.
Weld Root The point at which the back of the weld intersects
the base metal surfaces.
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Wheel The specially designed cast or forged steel cylindrical
element that rolls on the rail, carries the weight and provides
guidance for rail vehicles. Railway wheels are semi-permanently
mounted in pairs on steel axles, and are designed with flanges
and a tapered tread to provide for operations on track of a
specific gauge. The wheel also serves as a brake drum on cars
with on-tread brakes.
Wheel Base The horizontal distance between centers of the first
and last axles of a locomotive or car.
Wheel Bore The hole through the hub of the wheel which is
machined to precise dimensions to create a press fit on the axle
wheel seat.
Wheel Burn Damage to the tank shell or jacket due to frictional
contact with a rotating wheel, resulting in metal flow and/or
discoloration due to frictional heat.
Wheel Creep An operating condition wherein the wheel is not
purely rolling on the rail, yet, it is not purely slipping on the rail.
The coefficient of friction between wheel and rail is greatest in
this transition between purely rolling and purely slipping.
Wheel Flange The tapered projection extending completely
around the inner rim of a railway wheel, the function of which, in
conjunction with the flange of a mate wheel, is to keep the wheel
set on the track by limiting lateral movement of the assembly
against the inside surface of either rail.
Wheel Grinding A process of refining the tread contour of steel
railway wheels by rotating them against a grinding wheel under
precise controls.
Wheel Peener A process to induce compressive stresses on
plate surface of wheel.
Wheel Plate The part of a railway wheel between the hub and
the rim.
Wheel Rolling The wheel rotating on its axle theoretically
without motion existing between the wheel and the rail at the
area of contact.
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Wheel Seat The term used to describe a pair of wheels
mounted on an axle.
Wheel Sliding The wheel not rotating on its axis and motion
existing between the wheel and rail at the area of contact.
Wheel Slip An operating condition wherein there is wheel
rotation on its axis with motion of the wheel at the point of
contact with the rail. Wheel rotation speed during wheel slip is
greater than it is during rolling.
Wheel Slipping The wheel rotating on its axle with motion
existing between the wheel and rail at the area of contact.
Wheel Tread The slightly tapered or sometimes cylindrical
circumferential surface of a railway wheel that bears on the rail
and serves as a brake drum on cars with conventional truck
brake rigging.
Wide Gauge In general usage, the distance between the heads
of the rails of a railroad when it is significantly greater than 4 feet
8.5 inches (4' 8 1/2"), in contrast to broad gauge, which means a
material increase, as to 5 feet 6 inches (5' 6") or 6 feet..
Wrought Steel Wheel A railway wheel made by hot forging and
rolling as opposed to the pressure casting process.
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