U.S. Department of 

Transportation

Federal Railroad 

Administration

Safety Evaluation of High-Speed Rail Bogie 

Concepts

Office of Research

and Development

Washington, DC 20590

DOT/FRA/ORD-13/42 Final Report

October 2013
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1. AGENCY USE ONLY (Leave blank) 2. REPORT DATE

October 2013

3. REPORT TYPE AND DATES COVERED

Final Report

4. TITLE AND SUBTITLE

Safety Evaluation of High-Speed Rail (HSR) Bogie Concepts:

Comparative evaluations of HSR bogie design concepts for operation up to 125 mph

5. FUNDING NUMBERS

6. AUTHOR(S) and FRA COTR

Adrian Gorski, Dennis Richter, Andrew Smith, Jingjun Zhang , David Watson, and

Robert Spinelli Jr.

7. PERFORMING ORGANIZATION NAME(S) AND ADDRESS(ES)

GE Transportation

2901 East Lake Road

Erie, Pennsylvania 16531

8. PERFORMING ORGANIZATION

REPORT NUMBER

9. SPONSORING/MONITORING AGENCY NAME(S) AND ADDRESS(ES)

U.S. Department of Transportation 

Federal Railroad Administration

Office of Research and Development

Washington, DC 20590

10. SPONSORING/MONITORING

AGENCY REPORT NUMBER

DOT/FRA/ORD-13/42

11. SUPPLEMENTARY NOTES

COTR: S.K. (John) Punwani

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This document is available to the public through the FRA Web site at www.fra.dot.gov, or by 

calling (202) 493-1300.

12b. DISTRIBUTION CODE

13. ABSTRACT (Maximum 200 words)

The study defines the basic design concepts required to provide a safe, reliable, high-speed bogie for the next generation PRIIA 

passenger locomotive. The requirements and conditions for the U.S. market create unique design challenges that currently cannot 

be addressed by a U.S.-based manufacturer. A global survey identified dominant design trends focused on reducing P2 forces and 

optimizing bogie stability, curving performance, and ride quality. A likely design configuration with Bo-Bo arrangement was 

defined for the Task 2 stability analysis. The design of experiments, response surface techniques, and design space method used 

for the stability analysis demonstrated that a feasible design space of suspension variables could be identified for the conditions 

evaluated. The coinciding requirements for a higher horsepower engine, Tier 4 emissions, and crashworthiness equipment will add 

significant weight to the next generation locomotive. In order to meet the P2 limit of 82,000 pounds force at 125 mph, the next 

generation passenger locomotive will require a frame-hung traction motor, most likely with fully suspended drivetrain, and a disc 

brake system added to the already existing tread brake and dynamic brake systems. A preliminary safety FMEA concluded the 

new drivetrain would likely not increase safety risk.

14. SUBJECT TERMS

High-Speed Rail (HSR), bogie, concepts, Bo-Bo arrangement, Pareto plot, stability analysis, P2 

force, PRIIA locomotive

15. NUMBER OF PAGES

158

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ii

Contents

Executive Summary........................................................................................................................ 1

1. Introduction ................................................................................................................. 4

1.1 Background ................................................................................................................. 4

1.2 Objectives.................................................................................................................... 5

1.3 Overall approach ......................................................................................................... 5

1.4 Scope ........................................................................................................................... 6

1.5 Organization of the report ........................................................................................... 6

2. Reports......................................................................................................................... 8

2.1 Task 1 - Technology Survey of High-Speed Bogie Concepts..................................... 8

2.2 Task 2 - HSR Bogie Running Stability Summary Report......................................... 27

2.3 Task 3 - Dynamic Load Summary Report................................................................. 65

2.4 Task 4 - Braking Capability Summary Report.......................................................... 76

2.5 Task 5 - Wheel and Axle Summary Report .............................................................. 91

2.6 Task 6 - Locked Axle Safety Risk Assessment Summary Report ............................ 99

3. Report Conclusions ................................................................................................. 104

4. References ............................................................................................................... 106

Appendix A. Truck Design Fundamentals................................................................................. 107

Appendix B. Survey results tables – Truck design parameters ................................................. 120

Appendix C. Basic configurations of surveyed trucks............................................................... 125

Appendix D. Task Modeling Results......................................................................................... 129

Abbreviations and Acronyms ..................................................................................................... 150
iii

Illustrations

Figure 2.1.1-a: Interaction between wheel and rail geometries...................................................... 8

Figure 2.1.3-a: Bo-Bo wheel arrangement.................................................................................... 12

Figure 2.1.3-b: Surveyed locomotive truck speeds....................................................................... 12

Figure 2.1.3-c: Surveyed locomotive truck speeds....................................................................... 13

Figure 2.1.3-d: Wheel diameters................................................................................................... 14

Figure 2.1.3-e: Types of fabricated truck frames.......................................................................... 15

Figure 2.1.3-f: Surveyed wheel bases........................................................................................... 15

Figure 2.1.3-g: Axle guiding device ............................................................................................. 16

Figure 2.1.3-h: Traction links....................................................................................................... 16

Figure 2.1.3-i: Motor suspensions ................................................................................................ 18

Figure 2.1.3-j: Survey results of motor suspensions..................................................................... 19

Figure 2.1.3-k: Relation between unsprung mass and locomotive weight for P2 force limit of 

82,000 lb (track parameters are from PRIIA spec)............................................................... 20

Figure 2.1.3-l: Brake arrangements .............................................................................................. 21

Figure 2.1.3-m: Wheel-mounted disc brake ................................................................................. 21

Figure 2.1.3-n: Axle mounted disc brake...................................................................................... 22

Figure 2.1.3-o: Hollow-shaft (quill) mounted disc brake (Bombardier ALP45 DP).................... 22

Figure 2.1.3-p: Discs mounted on separate shaft (Siemens SF1) ................................................. 22

Figure 2.1.3-q: Truck/Body traction connections......................................................................... 24

Figure 2.2.2-a: High-speed truck used for Task 2 analysis: frame-hung motor and fully 

suspended drivetrain ............................................................................................................. 28

Figure 2.2.2-b: ADAMS/Rail Models for dynamic analysis........................................................ 29

Figure 2.2.2-c: MCAT track definition for simulations................................................................ 30

Figure 2.2.2-d: Wheel and rail profiles used for the simulations.................................................. 33

Figure 2.2.2-e: Parameter definition for contact geometry calculation ........................................ 34

Figure 2.2.2-f: Contact geometry of new Amtrak standard wheel profile and new AREMA136 

rail profile.............................................................................................................................. 34

Figure 2.2.2-g: Contact geometry of new Amtrak standard wheel profile and worn rail profile on 

straight track.......................................................................................................................... 35

Figure 2.2.2-h: Contact geometry of new Amtrak standard wheel profile and worn rail profile on 

curved track........................................................................................................................... 36

Figure 2.2.2-i: Contact geometry of worn wheel profile and worn rail profile on straight track . 37
iv

Figure 2.2.2-j: Contact geometry of worn wheel profile and worn rail profile on curved track .. 38

Figure 2.2.2-k: Parameters to be used for DOE analysis.............................................................. 40

Figure 2.2.2-l: Comparison of the responses on curved track between MCAT track inputs and 

measured random track inputs.............................................................................................. 45

Figure 2.2.2-m: Comparison of the responses on straight track between MCAT track inputs and 

measured random track inputs – Measured class 5 track irregularities versus MCAT class 5 

inputs, frame-hung motor, Loco Speed 95 mph (MCAT Class 5 inputs on curved track are 

used for straight track) .......................................................................................................... 47

Figure 2.2.2-n: Strategy of determining feasible design space..................................................... 49

Figure 2.2.3-a: Regression functions of the response surfaces..................................................... 50

Figure 2.2.3-b: Pareto plot ............................................................................................................ 51

Figure 2.2.3-c: Sensitivity plots.................................................................................................... 51

Figure 2.2.3-d: Stability analysis of the selected concept with frame-hung motor ...................... 52

Figure 2.2.3-e: Lateral displacement of leading axle – truck with axle-hung motor.................... 53

Figure 2.2.3-f: Lateral displacement of leading axle – truck with frame-hung motor.................. 54

Figure 2.2.3-g: Pareto and Sensitivity analysis of linear critical speed – 2-axle truck with framehung motor, worn wheel, and worn rail profiles................................................................... 55

Figure 2.2.3-h: Example of simulation results – frame-hung motor, Class 7 MCAT straight track 

inputs, new wheel-rail profiles, 125 mph, middle points (level 2) of the suspension 

parameters............................................................................................................................. 56

Figure 2.2.3-i: Sensitivity analysis results for truck with frame-hung motor, new wheel and worn 

rail profiles, Class 7 track, 125 mph ..................................................................................... 57

Figure 2.2.3-j: Comparison of sensitivity results between different wheel-rail profile 

combinations (frame-hung motor, Class 7 track, 125 mph) ................................................. 59

Figure 2.2.3-k: Example of two-dimensional design space.......................................................... 60

Figure 2.2.3-l: Minimization process of lateral car body acceleration in m/s^2 (frame-hung 

motor, straight track, new wheel, worn rail, MCAT Class 7 track, speed 125 mph) ........... 61

Figure 2.2.3-m: Minimization process of lateral truck acceleration in m/s^2 (frame-hung motor, 

straight track, new wheel, worn rail, MCAT Class 7 track, speed 125 mph)....................... 61

Figure 2.2.3-n: Minimization process of L/V of leading axle (frame-hung motor, 1.1 degree 

curve, 6” superelevation, new wheel, worn rail, MCAT Class 7 track, speed 125 mph)..... 62

Figure 2.2.3-o: Minimization process of track shift force of leading axle in N (frame-hung motor, 

1.1 degree curve, 6” superelevation, new wheel, worn rail, MCAT Class 7 track, speed 125

mph)...................................................................................................................................... 62

Figure 2.3.3-a: P2 force for PRIIA loco compared to existing diesel rolling stock ..................... 68

Figure 2.3.3-b: P2 force for PRIIA loco: loco weight versus unsprung weight, 125 mph .......... 69

Figure 2.3.3-c: P2 force for PRIIA loco: loco weight versus unsprung weight, 110 mph .......... 70
v

Figure 2.3.3-d: Main effects plot for parameters used in the P2 force equation for 125 mph 

operation ............................................................................................................................... 71

Figure 2.3.3-e: Comparison of potential P2 forces on different classes of track.......................... 72

Figure 2.3.3-f: Comparison of P2 forces on Class 4 track, at 60 mph, for freight and passenger 

locomotives........................................................................................................................... 73

Figure 2.3.3-g: P2 force – Passenger versus Freight for Class 4 track operation......................... 73

Figure 2.4.1-a: Rail network divided by area of the country........................................................ 77

Figure 2.4.2-a: Approach for utilizing the Gordon and Orringer study to draw conclusions about 

brake requirements on PRIIA locomotive ............................................................................ 79

Figure 2.4.3.1-a: Energy absorbed by each wheel for passenger and freight locomotives........... 80

Figure 2.4.3.2-a: Locomotive brake types.................................................................................... 82

Figure 2.4.3.2-b: Locomotive wheel energy reduction resulting from disc brake use ................. 83

Figure 2.4.3.2-c: Locomotive wheel energy reduction resulting from use of blended brake ....... 84

Figure 2.4.3.2-d: Wheel power and Heat Flux for a 285,000-pound PRIIA locomotive in Air only 

Emergency Brake application based on approach used in the Gordon and Orringer study . 87

Figure 2.4.3.2-e: Assessment of expected wheel temperature for a PRIIA loco utilizing tread 

brakes only in emergency stop from 125 mph based on the wheel temperatures from the 

Gordon and Orringer study ................................................................................................... 88

Figure 2.5.3-a: Comparison of loading assumptions used around the world for axle design....... 93

Figure 2.5.3-b: Rotating bending fatigue tests for press-fitted axles (Spec C, E represent 

induction hardening processes)............................................................................................. 95

Figure 2.5.3-c: Comparison of allowable alternating stress as a function of ultimate strength.... 96

Figure 2.5.3-d: Failure mode as a function of wheel seat diameter ratio...................................... 97

Figure 2.6.1-a: Axle-hung motor design....................................................................................... 99

Figure 2.6.1-b: Frame-hung motor HS coupling design............................................................. 100

Figure 2.6.1-c: Frame-hung motor LS coupling design.............................................................. 100
vi

Tables

Table 2.1.2-a: Parameters quantified in survey ............................................................................ 11

Table 2.2.2-a: MCAT geometry, tangent track, Class 7 track ...................................................... 30

Table 2.2.2-b: MCAT geometry, curved track, Class 7, CD = 6 inches....................................... 31

Table 2.2.2-c: Configuration of curved track................................................................................ 31

Table 2.2.2-d: Wheel and rail profile specification ...................................................................... 31

Table 2.2.2-e: 10 major design variables for modeling ................................................................ 39

Table 2.2.2-f: MCAT performance parameters ............................................................................ 41

Table 2.2.2-g: DOE table .............................................................................................................. 42

Table 2.2.2-h: 3 levels of suspension parameters......................................................................... 43

Table 2.2.2-i: Combinations of track and wheel-rail inputs for response surface analysis.......... 48

Table 2.2.3-a: Feasible stable and safe design space of truck with frame-hung motor................ 63

Table 2.4.3.1-a: Comparison of kinetic energy between freight and passenger locomotives . 80

Table 2.4.3.1-b: Comparison of kinetic energy between Class 43 and PRIIA locomotives......... 81

Table 2.4.3.2-a: Comparison of heat flux input between Arrow-III cars and diesel locomotives

............................................................................................................................................... 86

Table 2.5.3-a: Matrix of high-speed locomotive applications and relevant application data ....... 92

Table 2.5.3-b: Comparison of material standards for axles.......................................................... 94

Table 2.6.2-a: Severity scoring definition................................................................................... 101
1

Executive Summary

Background

The Federal Railroad Administration (FRA), through its Office of Research and Development, is 

facilitating the development of safe, efficient high-speed (HS) passenger rail systems in the 

United States. To this end, FRA will identify and support key innovative technologies in a range 

of rail industries, including track and structures, train control, human factors, passenger cars, and 

passenger locomotives. 

With respect to passenger locomotives, the “Passenger Rail Investment and Improvement Act” 

(PRIIA) specifies the need for a diesel-electric passenger locomotive capable of operating at a 

sustained speed of 125 mph. Perhaps the most critical subsystem for maintaining safe, stable, 

and comfortable operation at increasing speeds is the locomotive truck or bogie (the words truck 

and bogie are used interchangeably in this study). 

In order to understand the truck design concepts and tradeoffs required to ensure safe, reliable 

operation while meeting PRIIA requirements, FRA has contracted GE Transportation (GET) to 

conduct a “comparative evaluation of different HSR bogie concepts for operation up to 125 

mph.” This document reports the methods, results, and major conclusions of that study. It

focuses only on bogie system designs. 

Next generation PRIIA unique operational & design requirements

A meaningful assessment of high-speed bogie design concepts for operation in the United States

cannot be provided or understood without first defining the U.S. operating conditions and design 

requirements that make the next generation PRIIA-specified application unique and challenging.

The PRIIA specifications emphasize that a vital part of the U.S. intercity passenger rail network 

will continue to consist of Amtrak routes which share right-of-way with freight railroads. 

Therefore, next generation HS passenger locomotives must continue to have the capability to run 

on all classes of track. This will likely result in suspension design tradeoffs, as a suspension 

system optimized for high speeds on class 7 track will probably not be optimal for lower speeds 

on lower classes of track.

The high horsepower diesel engine required for 125 mph speed adds significant weight to the 

locomotive compared with most other high-speed locomotives, which are almost all powered by 

electricity only. Additionally, the coming Tier 4 emissions regulation will add significant weight 

to the engine, as will the additional structural material needed to meet the crashworthiness 

requirements that will also be part of the next generation PRIIA locomotives. Therefore, the U.S.

passenger locomotive that meets PRIIA, Tier 4, and crashworthiness standards will have an 

estimated axle weight of between 30 and 33 metric tons (MT). This compares with an axle load 

of 22 MT or less for the high-speed locomotives surveyed in this report. 

Finally, the PRIIA specifications will require a P2 force limit of 82,000 lbf. This is the same as 

the P2 force of an F40 locomotive at 110 mph. At 125 mph, the current F40, which weighs 

260,000 lbf, will exceed the PRIIA limit. GE estimates the next generation PRIIA capable of 

125 mph will weigh from 270,000 to as much as 300,000 lbf. 
2

Summary of Results

In order to provide a thorough evaluation of HSR bogie concepts and meet FRA objectives, GET 

scoped and conducted the study as the six separate tasks listed below: 

Task 1 – Global survey of high-speed rail (HSR) concepts

Task 2 – Running stability analysis

Task 3 – Dynamic loads on track of P2 force and track panel shift force

Task 4 – Braking capability

Task 5 – Fatigue reliability of wheel and axle

Task 6 – Drivetrain arrangement safety risk assessment

Key study highlights:

The study first defines the design requirements for the next generation PRIIA locomotive. The 

requirements and operating conditions create a unique design challenge. Currently, no U.S.

manufacturer provides a passenger locomotive that meets the requirements, which include

meeting the P2 limit of 82,000 lbf. The global survey identified dominant HS design trends 

focused on reducing P2 forces and optimizing stability, curving performance, and ride quality. A 

likely truck design configuration, designated Concept 1, was developed from the survey data to 

be used in the Task 2 running stability analysis. The design of experiments, response surface 

techniques, and design space method used in Task 2 demonstrated that a feasible design space of 

suspension variables could be identified for the conditions evaluated. The report also concludes

that the next generation passenger locomotive will require a frame-hung traction motor (even at 

110 mph) and a disc brake system added to the already existing tread brake and dynamic brake

(DB) systems. The report concludes that, based on GET’s current knowledge, the new drivetrain 

system expected to be required for a high-speed truck with frame-hung motor will likely not 

create additional safety risks. Finally, it was found that North American axle design methods are 

conservative compared with European and Japanese methods. U.S. designers may need to 

develop more accurate methods to calculate axle stress based on input data. 

Detailed summary:

The comprehensive global survey identified a number of dominant design trends in current HS 

bogies manufactured and operated in Europe, Japan, and China. Not surprisingly, the primary 

design characteristics focus on two general areas: 1) reduction of unsprung mass and total 

weight in order to minimize P2 forces, and 2) optimization of stability, curving performance, and 

ride quality.

The survey results also verified that the mostly all-electric locomotives in overseas markets 

operate at 22 MT per axle or less, much less than the anticipated 32 MT or more for future PRIIA 

locomotives. From the global survey results, two promising truck design configurations were 

identified for this study. The first of these, or Concept 1 (see below), was chosen as the 

configuration to be used in the Task 2 running stability analysis. 

In order to understand and quantify the effects of the truck suspension variables on bogie

performance, a bogie and locomotive dynamics model was developed using ADAMS/Rail. The

locomotive truck design configuration, identified as Concept 1, was used for the evaluation. 

This configuration included a two-axle truck with Bo-Bo arrangement, frame-hung motor, 

fabricated frame, and combination of tread and disc brake systems. MCAT track conditions for 

both straight and curved tracks were used with new and worn conditions for both wheel and rail. 
3

The impact and effect on the safety related performance of the truck suspension parameters was

investigated using design of experiments and response surface methods. A design space method 

was developed and it successfully demonstrated that a feasible combination of suspension 

parameters can be identified for the truck configuration and variables analyzed. For a complete 

verifiable design, more conditions and variables than could be included in this study should be 

evaluated.

The Task 3 study showed that in order to stay within the P2 force limit of 82,000 lbf, the trucks 

for future PRIIA passenger locomotive would have to incorporate a frame-hung motor to reduce 

unsprung mass. No current U.S.-produced locomotives can meet the limit at 125 mph. The 

relatively light F40 will meet the limit at 110 mph, but not at 125 mph. Both the alternating 

current (AC) and direct current (DC) motor versions of the Genesis locomotives exceed 82,000 

lbf at 110 mph. 

Braking requirements at 125 mph were evaluated. Braking energy is a function of locomotive 

mass and speed. Increasing speed to 125 mph increases the amount of energy that must be 

absorbed by 56 percent. Stopping a four-axle PRIIA diesel powered locomotive at 125 mph 

requires a bogie with an additional disc brake system combined with the already existing tread 

brake and DB system. For the required emergency stopping event with air-only brake operation, 

neither a tread nor a disc brake system alone was found to be sufficient. 

As unsprung mass is reduced in U.S.-high-speed bogies, the axle will become one of the more 

significant remaining components. The study compared U.S. and Non-U.S. design practices in 

the areas of calculation methods for axle stress and input definitions and assumptions for those 

calculations. We found that the U.S. approach to axle design is more conservative. This is felt 

to be partly due to the slower running speeds that result in lower P2 forces even with the heavier 

axle, and perhaps more significantly due to the lack of reliable data with which to accurately 

characterize axle loading. Going forward, North American axle designers may find it helpful to 

define axle loads less conservatively with the aid of more accurate input data. In addition, it may 

be necessary to consider higher performance materials and improved axle processing techniques 

to better improve fatigue resistance. Industry and government cooperation to develop better data 

and joint design standards should be considered as a way to ensure safe, reliable, and optimized 

high-speed axle designs. 

An assessment was made of the safety risks associated with the three basic drivetrain designs 

most likely to be used for 125 mph applications in North America. An FMEA was conducted for 

each of the design cases to assess the potential for increased risk. It was concluded that locked 

axle has historically not led to derailment and is therefore not considered a safety concern. 

Based on GE’s best current knowledge, it was considered unlikely that specific additional safety 

requirements in the design of the frame-hung traction motor concept would be required. In other 

words, although there may be additional reliability concerns associated with high-speed designs, 

those concerns were not believed to be tied to an increase in safety risk. This is another area in 

which industry standards, jointly developed by industry and Government for certain drivetrain 

components and systems, should be considered in order to ensure safe optimum designs.

The results indicate that the goal of meeting the PRIIA requirements will have a significant 

impact on bogie complexity, cost, and maintenance. A study is recommended to determine the 

cost-benefit tradeoffs between realizing a specific P2 limit and potentially incurring additional 

purchase and operating costs. 
4

1. Introduction

1.1 Background

The Federal Railroad Administration (FRA), through its Office of Research and Development, is 

researching new technologies that will benefit the U.S. high-speed rail system. The primary 

goals of these efforts will be focused on achieving safe, efficient, and effective deployment of 

integrated passenger rail systems in the United States. To this end, FRA will identify and 

support key innovative technologies in a range of rail industries, including track and structures, 

train control, human factors, passenger cars, and passenger locomotives. 

With respect to locomotives, the “Passenger Rail Investment and Improvement Act” (PRIIA) 

specifies the need for a diesel-electric passenger locomotive capable of operating at a sustained 

speed of 125 miles per hour (mph). This report focuses specifically on the bogie system designs

that will be required for a locomotive to safely and reliably meet the PRIIA requirements. 

The PRIIA specifications also state that a vital part of the U.S. intercity passenger rail network 

will continue to be the Amtrak long distance routes which share right-of-way with freight 

railroads. Therefore, an additional requirement for the next generation HS passenger 

locomotives is the capability to operate on all classes of track.

Unique North American (NA) Operating Environment for Passenger Rail

The requirement to develop High-Speed diesel locomotives capable of 125 mph on class 7 track, 

and also capable of operating throughout the Amtrak long distance routes on all other classes of 

track, presents unique challenges not found in other nonU.S. high-speed rail applications.

In addition to the requirement of being able to operate on all classes of track, the diesel engine 

and its required support systems significantly increase the weight of the locomotive compared 

with a conventional all-electric locomotive. The diesel engine weight in North America is 

further increased by the Tier 4 engine emissions requirement, as well as by new crashworthiness 

standards, both of which add significant additional weight to the locomotive. The estimated axle 

loads for the U.S. application will be 30 metric tons (MT) or more, compared with the 22 MT or 

less loads currently in place for nonU.S. applications.

Finally, the new PRIIA P2 force limit of 82,000 pounds (lb) is more stringent than past 

requirements. Most existing passenger locomotives do not meet this requirement even at the 

significantly slower speed of 100 mph (the relatively light F40 locomotive, at 3200 hp, is the 

exception).

Currently, the authors are unaware of any U.S.-based manufacturers with locomotive truck 

technology capable of operating safely at 125 mph.

Purpose and focus of this Study

In order to define the basic requirements for a safe HS passenger locomotive truck, FRA 

commissioned this study to conduct a comprehensive evaluation of the different truck design 

concepts with respect to North American operating conditions and requirements, and to 

determine the best configuration of components and design parameters for High-Speed Rail 

(HSR) bogies that can safely and reliably guide a diesel-electric locomotive on all classes of U.S.

track, and at a sustained speed of 125 mph on class 7 track.
5

Several important areas for further study that may benefit the rail industry were identified in this 

report.

1.2 Objectives

The following list presents the key objectives of this study:

• Highlight the differences and unique requirements that define the North American operating 

environment and locomotive requirements. 

• Conduct a global survey of the major locomotive truck design concepts and components.

• Determine a feasible configuration of components and design parameters for HSR bogies that 

can safely and reliably guide a diesel-electric locomotive on all classes of U.S. track, at a 

sustained speed of 125 mph on class 7 track.

• Include a proactive assessment of safety risks associated with HSR locomotive trucks and 

mitigation recommendations if needed. 

1.3 Overall approach

The study was scoped and planned by GET as six separate tasks designed to achieve the study 

objectives. The tasks with brief summaries, as stated in the original scope of work (SOW), are 

listed below. As the study progressed, it became apparent that some tasks needed to be 

modified, either because the changes better met the FRA objectives, or because the tasks could 

not be completed meaningfully within the limitations of this study. These changes will be 

identified within the specific report sections that describe each task. Note also that the title for 

Task 6 was changed as explained below.

Task 1 – Literature survey of HSR bogie concepts

Evaluate various HSR bogie concepts; understand their applicability, performance range, relative 

costs, and impact on safety. 

Task 2 – Running stability

Investigate the impact of suspension system design including primary, secondary suspensions, 

connections of the bogie to car body, alternative configurations such as frame-hung motors, and 

wheel base dimension, as well as passive and active steering; understand the engineering 

feasibility and limitation of the suspension system. 

Task 3 – Dynamic loads on track including P2 force and track panel shift force

Evaluate the impact of locomotive weight, motor suspension type, and the suspension parameters 

on P2 and track shift force.

Task 4 – Braking capability

Evaluate the capabilities and limitations of different brake arrangements including tread brakes, 

disc brakes, and a combination of tread and disc brakes. 

Task 5 – Fatigue reliability of wheel and axle

Evaluate the impact of high-speed application on wheel-axle fatigue life. Investigate the duty 

cycle of axle load under U.S. track conditions and the effects of reduced axle weight.
6

Task 6 – Drivetrain arrangement safety risk assessment

The title for this task has been changed from the original contract proposal and document. The 

original title was “Locked axle.” The original task description stated “Evaluate bogie response 

to locked axle conditions and other drivetrain component failures.” Since the study considered 

the entire drivetrain, not just the locked axle condition, the more correct title was used. 

The terms truck and bogie are used interchangeably in this report.

1.4 Scope 

This study is focused on passenger locomotive truck design concepts required to safely operate at 

125 mph and navigate the unique operating conditions of the United States, as well as meet the 

requirements defined in the recent PRIIA specifications. The following summarizes the scope in 

more detail:

• This study is specifically concerned with passenger locomotive truck design concepts 

required to safely and reliably operate at 125 mph on Class 7 track and in other conditions as 

defined below.

• The design parameters for this study are focused on the U.S. operating environment which 

includes Amtrak long distance routes which share right-of-way with freight railroads; highspeed bogies must be capable of operating on those routes and on all other classes of track. 

• The design requirements for this study specifically include a diesel-powered locomotive 

system capable of sustaining 125 mph.

• Except where it impacts safety, reliability is not generally within the scope of this study.

Though GE recognizes the advantages to the industry of standardizing certain truck components 

and systems, this is not within the contracted scope. Should FRA coordinate and support further 

study or discussion in this area, GE would like to be involved and contribute to the success of 

such activities.

1.5 Organization of the report

The report is organized and formatted in the standard FRA technical report format. 

In the body of the report, each of the six tasks described above is reported as a separate report 

section with definition, approach, results, and conclusions for each. 

In order to provide a complete but also readable report, many of the detailed tables, truck design 

configurations, and graphic results from Tasks 1 and 2 are provided as Appendices 2–4. 

A detailed section on the basics of truck design is provided in Appendix 1. An understanding of 

the concepts presented here is critical to understanding the report’s approach, results, and 

conclusions. Appendix 1 is summarized at the beginning of the Task 1 report.
7

Notice

The conclusions stated in this report are based on the knowledge gained during this project, 

GE Transportation’s past experience and best current knowledge. While GE is confident in 

the general conclusions drawn from this limited study, it is possible that a more thorough 

effort, such as an actual truck design and development project, would result in new 

knowledge that could modify these conclusions. The results, therefore, should not be 

considered final design recommendations. 

Notice on use of graphics

Much of the truck related graphics reviewed during this study were obtained from 

manufacturers’ brochures and Web sites. Due to IP concerns, we are not able to include 

these in a public document. Where possible and allowed, we have provided links to some of 

these resources in the reference section. 

Nomenclature

Within this report, the words truck and bogie are used interchangeably. An acronym and 

abbreviation definition list is provided at the end of the report.
8

2. Reports

2.1 Task 1—Technology Survey of High-Speed Bogie Concepts

2.1.1 Definition

The purpose of Task 1 is to conduct a review of worldwide high-speed truck design practices. 

Comparisons are also made between the typical overseas operating conditions of high-speed rail 

and the unique conditions required for the PRIIA application in the United States. A thorough 

survey of available design information was conducted and the most significant design trends 

relevant to this study were identified. 

The search was conducted by personnel at GE Research (GER), GET, and a contract engineer 

with over 20 years of truck design experience. Although specific design information was not as 

available as originally hoped, we were able to review dozens of truck designs for major highspeed passenger locomotives throughout the world and identify a number of important design 

trends. These trends focus on reducing weight and unsprung mass, as well as controlling 

stability, curving performance, and ride quality. 

A comparison of the relative costs of the key products identified was also a goal of the original 

project scope, but useful cost information was generally unavailable, and as a competitor, it 

would be inappropriate for GET to request or acquire cost and price data.

Truck Design Requirements & Performance Characteristics

The truck is the interface between the locomotive and the track. As illustrated in Figure 2.1.1-a 

below, the interactions between wheel and track geometries provide the dynamic inputs to the 

truck and locomotive. A complete understanding of the correct and relevant inputs and system 

responses is fundamental to designing a safe and reliable passenger locomotive truck.

Figure 2.1.1-a: Interaction between wheel and rail geometries
9

Truck functional requirements

The truck subsystem provides the following basic functions:

1) support the car body weight and distribute the locomotive weight to the axles

2) guide and steer the locomotive on the tracks safely and with stability under all operating 

conditions 

3) provide traction and braking effort between the wheels and rails

4) transmit traction and braking force to the car body

5) reduce and isolate vibrations and impact loads from wheel-track interactions

6) mount and support auxiliary components and equipment

Key performance measures

The basic functional requirements stated above must be met by a system that will provide an 

acceptable combination of the following six key truck performance measures in a cost effective, 

reliable, and safe package. 

1) Stability under all operating conditions

2) Derailment prevention in all conditions

3) Wheel-rail dynamic forces

4) Braking capability

5) Ride quality and comfort

6) Component and structural reliability and life

Understanding these functional requirements is important to understanding design tradeoffs and 

the results of this study. All six requirements are described in significantly more detail in 

Appendix A.

Comparison of Operating Conditions

To draw meaningful conclusions from a survey of worldwide high-speed truck design concepts 

requires an understanding of the operating conditions in the overseas applications and how they 

compare with U.S. operating conditions.

The differences in U.S. and overseas requirements and conditions are discussed in the 

introduction section and elsewhere in this study. Axle load and track quality are summarized 

again here because these are particularly relevant to the U.S. design requirements and report 

conclusions.

Locomotive weight and axle load – Although some of the truck manufacturers claim that 

their designs could also be used for applications of diesel locomotives, most of the surveyed 

locomotives are electric. Electric-powered locomotives tend to be significantly lighter than

diesel locomotives. The diesel engine and supporting equipment (e.g., the cooling and oil 

systems) add considerable weight to the locomotive. In North America, the impending Tier 4 
10

regulations will force the addition of still more weight in order to provide the necessary 

emissions reduction systems. Current best estimates are that the future PRIIA and Tier 4 

certified passenger locomotive will have an axle load of approximately 33 MT or higher.

To meet the specified P2 force limit, this high axle load will drive a lot of design changes. 

What could work well in the existing truck design which is used in other countries may no 

longer be applicable for U.S. high-speed locomotive—for example, wheel-mounted disc 

brake, truck frame, and partially suspended drivetrain. A fully suspended drivetrain will be 

necessary for U.S. design to meet the P2 force requirement. 

Track quality – The track geometry and quality standards in the United States are different 

from those in other countries. In general, the U.S. track classes have much larger allowable 

geometrical deviations than other countries. This means that the U.S. tracks will have much 

more and higher excitations.

2.1.2 Approach

A thorough survey of available literature and information was conducted to identify current HSR 

bogie design characteristics and trends in all the major world markets and applications. 

The following areas were specifically targeted:

 Focus on systems and components that specifically relate to the six key performance 

characteristics described in the previous truck design section 2.1.1.

 Designs specifically for high speeds, defined for this study as 125 mph or higher

 Axle loads—significant impact on P2 forces

 New developments and trends

The following major locomotive builders and their products were included in the survey:

• Siemens (Germany)

• Bombardier (Canada)

• Alstom (France)

• Voith (Germany)

• Vossloh (Germany)

• Talgo (Spain)

• CAF (Spain)

• CSR/NSR (China)

• Ansaldo Breda (Italy)

• Hitachi Rail (Japan)

• Antranz (Sweden)

• EMD (USA)

• MPI (USA)

• GE (USA)
11

In order to cover the most recent development trend in high-speed truck design, the following 

high-speed train manufacturers and their products were reviewed as a reference:

• Siemens (Germany)

• Bombardier (Canada)

• Alstom (France)

• Talgo (Spain)

• CSR/CNR (China)

• Kawasaki (Japan)

• Nippon Sharyo (Japan)

• Ansaldo Breda (Italy)

• Hyundai-Rotem (S. Korea)

For information collection, the Internet and other public documents and references were used. 

Due to limitations of the public information sources, some critical technical data may not be 

available or complete. 

Using the strategy set up above, the survey was conducted following specific designs and 

parameters for evaluation of performance as shown below in Table 2.1.2-a:

Table 2.1.2-a: Parameters quantified in survey

1 Truck type (wheel arrangement)

2 Speed range

3 Axle load

4 Wheel diameter (new vs. worn)

5 Bogie frame design

6 Wheelbase

7 Axle guiding design

8 Primary suspension (vertical)

9 Secondary suspension

10 Motor suspension design

11 Brake design/arrangement

12 Traction link design between truck and car body

13 Builder

14 Drivetrain design

15 Parking brake design/arrangement

16 Axle design

17 Max TE/Axle

18 Power per axle

19 Gauge

20 Minimum curve

21 Bogie weight

22 Application

Items 1–12 are reviewed in detail in the following results section. Items 13–22 are not discussed 

in the results section, but are listed in Appendix B as part of the tabular data review.
12

2.1.3 Results

Detailed technical data of the survey results for some typical truck products are provided in the 

tables in Appendix C. Some representative configurations of the truck designs are sketched in 

Figure C.1a to C.1g. In the following sections, significant findings in each design area will be 

discussed.

Wheel arrangement

Unlike typical North American freight locomotives, the majority of the surveyed trucks for 

passenger applications have a Bo-Bo arrangement (a locomotive with two bogies, each bogie 

having two powered axles), as shown below in Figure 2.1.3-a. In a Bo-Bo wheel arrangement, 

each of the two locomotive trucks has two powered axles. Such an arrangement provides a 

symmetrical truck layout and component design and is suitable for high-speed operation. Its 

wheel base is generally smaller than the total wheel base of the three-axle truck. Therefore, it 

may provide better curving performance.

Figure 2.1.3-a: Bo-Bo wheel arrangement

Speed range

The speeds of the surveyed locomotive trucks are illustrated in Figure 2.1.3-b. They range from 

140 kph (87 mph) to 357 kph (221 mph). The majority of the higher speed (200 kph (125 mph) 

or more)) applications are found in Europe.

Figure 2.1.3-b: Surveyed locomotive truck speeds

1 2 3 4 5 6 7 8 9 10 11 12 13 14

Siemens, 

SF1 

Siemens, 

SF2

Siemens, 

SF3

Siemens, 

SF4

Bombardier, 

Flexx 

Power 

140/200 

Bombardier, 

Flexx power 

250

Bombardier, 

Flexx power 

350

Alstom, 

Prima II

Vossloh, 

Euro 

3000AC Talgo

Bombardi

er, ALP 

45 DP

GE, 

Genesis

MPI/GE 

MBTA

Ansaldo 

Breda, 

E403
13

Axle load

While the axle loads in U.S. applications are approximately 30~33t (67,000~72,750 lb), the axle 

loads in the rest of the world are much lighter. As shown in Figure 2.1.3-c, all of the nonU.S.

trucks surveyed are designed for axle loads of 22 MT or less. The high-speed applications of 

350 kph or more have the lowest loads. The axle load of Bombardier Flexx power 350 (truck #7 

in Figure 2.1.3-c) is only 17 MT (37,478 lb). The smaller axle load helps keep the P2 force at an 

acceptable level, even at the much higher speeds.

Figure 2.1.3-c: Surveyed locomotive truck speeds

Wheel Diameters

Figure 2.1.3–d below compares wheel diameters of the trucks surveyed. The speed comparison 

chart is also provided for reference. 

The wheel diameters of the surveyed trucks ranged from 1010~1250 mm (39.76~49.21 inches

(in)). The selection of the best wheel diameter depends on a number of factors, which include 

locomotive speed, tractive effort, traction motor rpm and power, required bottom rail and 

locomotive clearances, unsprung mass, and wheel life requirement. 

For a high-speed application, reducing wheel diameter can reduce unsprung mass and P2 force in 

turn. But a comparison of the speeds of the surveyed locomotives with the wheel diameters 

below does not show a consistent link between smaller wheel size and higher speed. This 

affirms the assertion that wheel diameter selection is complicated and dependent on many 

factors.

1 2 3 4 5 6 7 8 9 10 11 12 13 14

Siemens, 

SF1 

Siemens, 

SF2

Siemens, 

SF3

Siemens, 

SF4

Bombardier, 

Flexx 

Power 

140/200 

Bombardier, 

Flexx power 

250

Bombardier, 

Flexx power 

350

Alstom, 

Prima II

Vossloh, 

Euro 

3000AC Talgo

Bombardi

er, ALP 

45 DP

GE, 

Genesis

MPI/GE 

MBTA

Ansaldo 

Breda, 

E403
14

Figure 2.1.3-d: Wheel diameters

Truck frame

Except for the MBTA locomotive truck (#13 on the chart), which is designed by the U.S.-based 

company, MotivePower, Inc., (MPI), all other surveyed truck frames are fabricated. Fabricated 

frames are widely used for high-speed applications because they are lighter and easier to 

manufacture than castings. Through improved welding technology and quality, a fabricated 

truck frame can meet the strength and reliability requirements. By using the finite life design 

method based on load duty cycles, a fabricated truck frame can be optimized to meet the required 

weight and strength requirements. To further reduce weight, some designs even use different 

materials for the transoms (e.g., Ansaldo-Breda truck in which the entire transom is made of 

different material and bolted on to the side frame).

There are different shapes of fabricated truck frames. Figure 2.1.3-e below shows four typical 

truck frame designs. Type A (e.g., Flex Power 140/200 truck of Bombardier) and type B (e.g.,

GE Genesis truck, Vossloh design) were used in the earlier designs and for relatively lower 

speeds. Type C (e.g., Vossloh Euro 3000AC) and Type D, in particular, (e.g., Siemens SF1, 

SF2, SF3 and SF4, Bombardier Flexx Power 250 and Flexx Power 350, Alstom Prima II 

locomotive truck, Ansaldo Bread truck) are more widely used for higher speed applications. 

This is, in part, because the taller coil springs are used for the secondary suspension and their 

longer operating length needs more space between car body and truck frame.
15

Figure 2.1.3-e: Types of fabricated truck frames

Wheel base

Figure 2.1.3-f below summarizes the wheel base of different truck designs with ranges from 

2600~3000 mm (102.36~118.11 in). A longer wheel base generally improves stability but may 

result in poor curving performance. There may also be a tradeoff between the desired truck 

length and packaging requirements in order to adequately contain all the other truck equipment 

within the available space. A speed chart is again included for reference. Note that, as with 

wheel diameter, there is no clear correlation among the surveyed trucks between wheelbase and

speed.

Figure 2.1.3-f: Surveyed wheel bases

1 2 3 4 5 6 7 8 9 10 11 12 13 14

Siemens, 

SF1 

Siemens, 

SF2

Siemens, 

SF3

Siemens, 

SF4

Bombardier, 

Flexx 

Power 

140/200 

Bombardier, 

Flexx power 

250

Bombardier, 

Flexx power 

350

Alstom, 

Prima II

Vossloh, 

Euro 

3000AC Talgo

Bombardi

er, ALP 

45 DP

GE, 

Genesis

MPI/GE 

MBTA

Ansaldo 

Breda, 

E403
16

Figure 2.1.3-g: Axle guiding device

As noted previously, the axle guiding device is important for controlling the locomotive’s critical 

speed (hunting stability), which must be well above actual operating speed. The survey results 

show that except for the MBTA truck, which has a pedestal leg design, all other trucks have 

elastic axle guiding designs. Figure 2.1.3-g above illustrates three typical axle guiding designs. 

In Figure 2.1.3-g (a), the journal box is connected to the truck frame through a traction link. The 

traction link may have two different types: three-point link, such as Siemens SF1~SF4, or a twopoint link, such as Bombardier Flexx Power trucks, as shown in Figure 2.1.3-h below. With 

proper selection of the bushings in the link, both three-point and two-point link designs can 

provide enough longitudinal and lateral guiding stiffness to meet both stability and curving 

performance requirements. This is especially true when a relatively soft stiffness of the rubber 

bushing, which still meets the straight track stability requirement, is chosen. Such soft elastic 

guiding can utilize the maximum self-steering properties of the tapered wheel profiles. 

For an axle guiding design like that shown in Figure 2.1.3-g (b) (Alstom Prima locomotive truck

- #8), the two-point link in Figure 2.1.3-h (c) below will be used. The guiding device in Figure 

2.1.3-g (c) is integrated with the journal box—as is the case with the Ansaldo Breda truck. It is 

elastically connected to the truck frame by using a rubber bushing at one end. 

The design life of the rubber bushing can be as long as 8 years or more. This makes such

designs maintenance free within the design life.

Figure 2.1.3-h: Traction links
17

Primary and secondary suspensions

For high-speed applications, the primary and secondary suspensions almost uniformly use coil 

springs (see Figures C.1a to C.1g in Appendix C). Coil springs will mainly provide stiffness in 

the vertical direction, but little stiffness in lateral and longitudinal directions. While the elastic 

axle guiding devices will mainly control the wheelset stability (hunting), the coil suspensions 

will mainly reduce and isolate the vibrations introduced by the track irregularities. The coil 

springs will provide large displacement (travel) of motions between axle and truck frame and 

between the truck and car body. Together with the necessary dampers, the design can provide 

good ride quality and passenger comfort. 

The coils springs will provide relatively low yaw stiffness between the truck and the car body. 

This can benefit the curving performance and reduce track shift force and wheel wear. 

Hydraulic dampers are used to provide damping for vibration control. Yaw and lateral dampers 

are used to control hunting. 

An antiroll bar can be used to control the roll motion of the car body, but it is not used very often 

on locomotives. Among the surveyed trucks, there is only one truck design that uses the antiroll 

bar (see Talgo’s truck design in Figure C.1f in Appendix C). 

Motor suspension

The three basic motor suspension methods utilized by the surveyed trucks are: (1) axle-hung 

(nose) suspension, (2) frame-hung (with drive train either partially suspended or fully 

suspended), and (3) body-hung suspension (drivetrain partially suspended).

Schematics of the three configurations are shown below in Figure 2.1.3-i:
18

Figure 2.1.3-i: Motor suspensions
19

Figure 2.1.3-j below shows the motor suspension methods of each of the surveyed 

manufacturers. For reference, the speed plot is repeated next to it. As indicated, for speeds 

above 140 km/h, all but one of the trucks outside the United States use either frame-hung or 

body–hung motor suspension. The only exception is the #9, Vossloh Euro 3000 AC truck which 

is designed for 200 kph (125 mph), but still uses an axle-hung motor suspension. The axle load 

for this design, 22 MT, is well below the U.S. application. In contrast, two of the three surveyed 

high-speed locomotives in the United States use axle-hung motor suspension; only one uses 

frame-hung motor suspension.

Figure 2.1.3-j: Survey results of motor suspensions

The unsprung mass has a large impact on the magnitude of the P2 force between the wheel and 

rail. For a given track structure, the P2 force will solely be a function of the locomotive weight 

(axle load), unsprung mass, and locomotive speed. As locomotive speeds increase, the 

locomotive weight or unsprung mass, or both, must be reduced to maintain the same P2 force. 

Figure 2.1.3-k below shows how, as speed increases, the locomotive weight and unsprung mass 

must be reduced to maintain a specified constant P2 force. If locomotive weight increases, 

unsprung mass must be significantly reduced to maintain the desired P2 force limit. In the case 

of Figure 2.1.3-k, the constant P2 limit is 82,000 pounds force (lbf) for the track structure 

specified by PRIIA. 

Among the three types of motor suspensions shown in Figure 2.1.3-i, the axle-hung suspension 

will have the maximum unsprung mass for the same wheelset design and, for that reason, is 

normally used for lower speed applications, such as Siemens SF2, Bombardier Flexx Power 140, 

Vossloh Euro 3000 AC (Euro Light). The frame-hung suspension with fully suspended 

drivetrain, shown in Figure 2.1.3-i (c) (e.g., Siemens SF1 & SF4, Bombardier Power 250, and 

Ansaldo-Breda E403 locomotive truck), will provide the smallest unsprung mass. It is often 

used for very high-speed truck designs or the application of a very heavy locomotive running at 

higher speeds. 

1 2 3 4 5 6 7 8 9 10 11 12 13 14

Siemens, 

SF1 

Siemens, 

SF2

Siemens, 

SF3

Siemens, 

SF4

Bombardier, 

Flexx 

Power 

140/200 

Bombardier, 

Flexx power 

250

Bombardier, 

Flexx power 

350

Alstom, 

Prima II

Vossloh, 

Euro 

3000AC Talgo

Bombardi

er, ALP 

45 DP

GE, 

Genesis

MPI/GE 

MBTA

Ansaldo 

Breda, 

E403
20

The configuration of frame-hung suspension with partially suspended drivetrain (as shown in 

Figure 2.1.3-i (b)) will have an unsprung mass between that of the axle-hung suspension and 

frame-hung suspension with fully suspended drivetrain (Siemens SF3, Bombardier Flexx Power 

350, Alstom Prima II locomotive truck). Body-hung motor suspension in Figure 2.1.3-i (d) in 

which the motor is suspended on the carbody may not necessarily offer the minimum unsprung 

mass since the gear box on the axle mass adds to the unsprung mass.

Figure 2.1.3-k: Relation between unsprung mass and 4-axle locomotive weight for P2 force 

limit of 82,000 lb (track parameters are from PRIIA spec)

Brake design

Figure 2.1.3-l shows the survey results of air brake arrangements. For operation at speeds of 140

kph or higher, the majority of the truck designs, especially the trucks designed by European 

companies, use disc brakes. The two locomotives designed by U.S. companies are equipped 

with tread brakes alone. The trucks of New Jersey Transit’s (NJT) APL 45 DP locomotive 

designed by Bombardier are provided with both tread brake and disc brake.

There are two types of disc brake arrangements widely used: wheel mounted discs and axle 

mounted discs. As illustrated in Figure 2.1.3-m, wheel-mounted disc brakes are normally used 

for powered axles because of drivetrain arrangement and space requirement for packaging gear,

gear case, and motor, among other things. Adding wheel-mounted disc brakes therefore

increases the unsprung mass. 

Axle-mounted disc brake arrangements have different variations. Figures 2.1.3-n through 2.1.3-

p show three typical axle-mounted disc brake designs. For unpowered axles, like the application 

on an Electrified Multiple Unit (EMU) high-speed train, the design in Figure 2.1.3-n is 

commonly used. Depending on the requirement for the brake capability, multidiscs can be 

directly mounted on the axle. Again, such a design will increase the unsprung mass. To reduce 

the unsprung mass, a hollow shaft (quill) mounted disc brake arrangement was developed by 
21

Bombardier for their fully suspended drivetrain design (Figure 2.1.3-o). Since the quill and the 

discs mounted on the quill need more space than just the axle itself, this will require a longer 

distance between the frame-hung motor and the axle. As a result, the gear transmission may 

need “idlers” in the gear case, which may require an increase in wheel base length.

Figure 2.1.3-l: Brake arrangements

Figure 2.1.3-m: Wheel-mounted disc brake
22

Figure 2.1.3-n: Axle mounted disc brake

Figure 2.1.3-o: Hollow-shaft (quill) mounted disc brake (Bombardier ALP45 DP)

Figure 2.1.3-p: Discs mounted on separate shaft (Siemens SF1)
23

Another variation of the axle-mounted disc brake arrangement, intended to reduce the unsprung 

mass, is the Siemens SF1 truck design. In this design, a separate shaft mounted with brake discs 

is driven through the drivetrain, as shown in Figure 2.1.3-p. The separate shaft is mounted on 

the frame, and the overall drivetrain is fully suspended. Compared with the Bombardier design 

in Figure 2.1.3-o, this arrangement may have the advantage of somewhat reducing the 

wheelbase. 

The combination of tread and disc brakes may be a tradeoff between brake capacity, design 

complexity, and wheel reliability. Tread brake can be used as a supplement to the disc brake 

within the thermal capability of the wheel. Periodic application of the tread brake can clean the 

wheel tread to improve adhesion. 

Secondary Traction Links 

The connection in longitudinal direction between truck frame and car body will transmit tractive 

and brake efforts. Its design and location may impact the truck configuration, wheel base, and 

some of the performance traits, especially weight transfer. Although a lot of today’s locomotives 

use individual axle control and can maximize the utilization of the transferred weight, it is still 

good practice to reduce the weight transfer as much as possible, especially for locomotives with 

truck control for traction. Therefore, a good practice is to make the “traction point” as low as 

possible (as close as possible to the rail top).

In addition to providing force transmission in the longitudinal direction, the connection between 

the truck and car body should provide enough flexibility for lateral, vertical, and rotational 

motions. 

Figure 2.1.3-q shows several typical traction connections between truck and car body. The 

center pivot (or traction pin) design in Figure 2.1.3-q (a) is a conventional design and is still used 

very commonly in many applications like the Siemens SF series truck and the Vossloh truck 

because of its simplicity. The center pin in this application does not take any vertical load. The 

pin can move freely in the lateral direction, relative to the truck frame, until the lateral bump 

stops are encountered.
24

(a) Center Pivot (b) Center Pin - Traction Links

Figure 2.1.3-q: Truck/Body traction connections

The combination of traction pin and traction links, shown in Figure 2.1.3-q (b), is another type of 

widely used connection. The “Z” type arrangements of the links provide the necessary 

compliance in lateral direction. The application of the rubber bushings between the links avoids 

wear and is maintainence free. The traction rod design in Figure 2.1.3-q (c) is very commonly 

used on trucks produced by Bombardier, Alstom, Talgo and Ansaldo. The traction point of this 

configuration can be much lower and close to the rail top. For application on a diesel 

locomotive, special consideration should be given to the arrangement of the traction rod in order 

to avoid interference with the fuel tank.

2.1.4 Conclusions and Recommendations

Conclusions

The primary design characteristics of the high-speed trucks surveyed are focused on two general 

areas: 

1) Reducing unsprung mass and total weight in order to minimize P2 forces

2) Optimizing stability, curving performance, and ride quality

The following list summarizes the dominant design trends of the high-speed trucks surveyed: 

• Two axle truck with Bo-Bo arrangement for high stability and curving performance. 

Wheel base is typically between 2.6~3 meters, and the wheel diameter ranges from 1.01~1.25 

meters.
25

• Fabricated truck frame for light weight and manufacturability (1 truck had a cast frame)

• Frame-hung motor suspension with either partially suspended drivetrain or fully suspended 

drivetrain to reduce unsprung mass (and therefore reduce P2 force). For fully suspended 

drivetrain design, quill (hollow shaft) design is most common.

• Disc brakes or combination of disc and tread brakes to increase brake capacity and reduce 

possible damage to the wheels. Wheel mounted discs are very commonly used. But to 

further reduce unsprung mass, suspended quill mounted discs or separate axle mounted discs 

are used.

• Elastic connections (rubber bushings) between journal box and truck frame for better axle 

guiding performance to increase locomotive hunting stability and curving performance

• Coil springs for both primary and secondary suspensions for optimum ride quality and 

comfort

• Center pivot or end traction rod for secondary link (connection between truck and 

carbody)

• Application of active yaw control to improve curving performance 

• Low axle loading to reduce P2 force—all but one of the nonUS surveyed locomotives have 

an axle load of 22 MT or lower. The only exception is the Bombardier Flex 250 designed for 

U.S. application that has an axle load of 32.6 MT. 

• Solid axle for locomotive applications

Recommendations

Impact of high axle loads

The high axle loads of 30–33 MT or higher for the U.S. high-speed truck will make meeting the 

P2 force limit extremely challenging and will drive a number of design changes. What works 

well in the overseas applications may not be adequate for the heavier U.S. applications. Wheelmounted disc brake and partially suspended drivetrain, for example, will likely not be adequate 

in the United States. Some preliminary analysis, discussed later in more detail, indicates that a 

fully suspended drivetrain, as well as combined disc and tread brake systems, will be necessary 

for the U.S. design. 

Impact of U.S. track conditions

The suspension systems in the existing foreign locomotive trucks will have to be designed

differently for U.S. track conditions. The dynamic loads experienced by the truck and 

components due to track excitations are very different, necessitating a different design duty 

cycle/load spectrum than that used for truck design in other countries. The truck structure and 

components for the U.S. design will also have to be designed differently. In short, there may be 

no existing truck designs from other countries that meet the performance requirement, reliability 

and life requirements, and safety requirements for operation on U.S. track. This has been 

demonstrated in past experience and tests. Developing new high-speed trucks for the PRIIA 

specified U.S. application will require leveraging both existing technology and new innovations.
26

Proposed design concepts

Due to the extreme axle load requirement for truck design to reach speeds of up to 125 mph, the 

following two concept examples are considered the most promising options:

Concept 1

 Two axle truck with Bo-Bo arrangement – stability and curving performance

 Fabricated truck frame – weight

 Frame-hung motor with fully suspended drivetrain – unsprung mass 

 Combination of disc and tread brakes by using either quill mounted discs or separate 

disc axle 

Concept 2

 Three-axle truck with A1A arrangement 

 Frabricated truck frame

 Frame-hung motor with fully suspended drivetrain

 Combination of disc and tread brakes by mounting discs on center axle

Concept 1, with only four axles per locomotive, still faces the challenge of high axle load. To 

meet the P2 requirement, the fully suspended drivetrain design and the arrangement of the disc 

brake are critical. Both will increase the complexity of the component design and packaging 

requirements. 

Concept 2 reduces axle load by distributing weight over six axles rather than four. The center 

axle can be used for axle mounted disc brakes. This may lead to a simpler drivetrain system 

design due to the decreased demand of brakes on the quill. A disadvantage may be that the 

overall truck wheel base will be larger than the base for the Concept 1 design. 

Configuration for Task 2 modeling 

Concept 1 was chosen as the preferred concept for U.S. application. This concept was chosen

for the modeling analysis described in Task 2, which is described in more specific detail in the 

next section. Four-axle locomotives with frame-hung traction motors are the dominant and 

proven trend in passenger locomotives. It is believed that this concept is likely to be the most 

successful technology and also the most commercially viable concept for the next generation 

passenger locomotive.
27

2.2 Task 2—HSR Bogie Running Stability Summary Report

2.2.1 Definition

The design of the suspension system will critically impact the safety related locomotive/truck 

dynamic performance in the areas of stability, derailment safety (curving), dynamic wheel-rail 

forces, ride quality, and dynamic loads on the truck components. The purpose of Task 2 is to use 

the selected truck design concept to study and analyze the effects of truck suspension system on 

truck dynamics performance under U.S. track conditions.

2.2.2 Approach

For the analysis, the following approaches were taken: 

• Develop a truck and locomotive dynamics model in ADAMS/Rail by using the truck concept 

with frame-hung motor as proposed in Task 1

• Use track conditions as specified in MCAT (Minimally Compliant Analytical Track) for 

dynamic simulations on straight and curved tracks

• Consider both new and worn wheel and rail conditions

• Investigate impact and effects of suspension parameters on safety related performances by 

using Design of Experiments (DOE) and response surface method

• Determine feasible design space of suspension parameters to meet allowable safety limits of 

dynamic performances

Truck/Locomotive Dynamics Model

The dynamic analysis for this task was conducted using ADAMS/Rail, an advanced program 

specifically developed to model and analyze railway vehicle dynamics. 

The truck model developed for this analysis is based on the truck design Concept 1 described at 

the end of the previous section (Task 1). It is shown in Figure 2.2.2-a below. It consists of 

frame-hung motor, fully suspended drivetrain (quill drive), three-point traction link with rubber 

bushings for axle guidance, primary and secondary coil springs, primary vertical dampers, 

secondary lateral, vertical and yaw dampers, as well as a center pin design for the connection 

between truck and car body. The locomotive is assumed to weigh 282,000 lb. The truck center 

pin distance is 510.6 in (12.97 m) and wheel base is 118.26 in (3 m). The new wheel diameter 

was set at 40 in. With the frame-hung motor and fully suspended drivetrain, the unsprung mass 

is 6,733 lb.
28

Figure 2.2.2-a: High-speed truck used for Task 2 analysis: frame-hung motor and fully 

suspended drivetrain

Figure 2.2.2-b below shows the ADAMS models of the truck and locomotive. The locomotive 

model is a multibody dynamic system model. It includes 109 “parts” (bodies) and has 486 

degrees of freedom. The primary and secondary suspension systems are modeled as linear. The 

motor suspension and the quill suspension are modeled as rubber parts with nonlinear 

characteristics. Necessary hard stops in both primary and secondary connections are built in by 

using either impact or bistop functions. The traction motor is modeled with a stator and rotor. 

Gear elements are used to simulate the torque and motion transmission between the rotor and 

wheelset.
29

Figure 2.2.2-b: ADAMS/Rail Models for dynamic analysis

Track Inputs

The dynamic responses of the locomotive are functions of the track inputs. For the purposes of 

this study’s analysis, the track conditions specified in Appendix D of CFR 49 Part 213 (MCAT 

Simulations Used for Qualifying Vehicles to Operate at High-Speeds and at High Cant 

Deficiencies) were used. MCAT specifies a track model containing defined geometry 

perturbations at the limits permitted for a class of track and level of cant deficiency, as shown in 

Figure 2.2.2-c. Class 7 track, with a disturbance wavelength of 62 feet (ft), was chosen for the 

simulations on both straight and curved tracks. The corresponding parameters of the MCAT
30

track setup are shown in Table 2.2.2-a and 2.2.2-b. The track configuration for the curving 

analysis is detailed in Table 2.2.2-c, where a 1.1-degree curve with a super-elevation of 6 in is 

used. A running speed of 125 mph is used for the analysis on both straight and curved tracks. 

This speed will create a 6-inch cant deficiency for the selected 1.1-degree curve.

Figure 2.2.2-c: MCAT track definition for simulations

Table 2.2.2-a: MCAT geometry, tangent track, Class 7 track
31

Table 2.2.2-b: MCAT geometry, curved track, Class 7, CD = 6 in

Table 2.2.2-c: Configuration of curved track

Wheel and Rail Profiles

The combination of the wheel and rail profiles will significantly impact the running stability and 

dynamic response of the locomotive. To understand the influence of possible variations of the 

wheel-rail profile during the normal operation, the following typical wheel and rail profiles are 

used for the wheel-rail contact geometry calculation:

Table 2.2.2-d: Wheel and rail profile specification

in m in m ft m

Hunting Perturbation a1 0 0 10 3.048 d1 1000 304.8

Gage Narrowing a2 0.5 0.0127 62 18.8976 d2 1000 304.8

Gage Widening a3 0.5 0.0127 62 18.8976 d3 1000 304.8

Repeated Surface a9 0.75 0.01905 62 18.8976 d4 1500 457.2

Repeated Alinement a4 0.375 0.009525 62 18.8976 d5 1000 304.8

a11 0 0 62 18.8976

a10 1 0.0254 62 18.8976

a6 0 0 62 18.8976

a5 0.5 0.0127 62 18.8976

Short Wrap a12 0.5 0.0127 62 18.8976 d8 1000 304.8

a7 0.333 0.0084582 62 18.8976

a8 0 0 62 18.8976

a13 0.667 0.0169418 62 18.8976

Single Surface d6 1000 304.8

d7

Disturbance segments Disturbance Amplitude a Wave Length λ Segment length d

Single Alignement 1000 304.8

MCAT Curving, Class 7, CD=6 inches

Combination Perturbati d9 1000 304.8

Track Distance Curvature Superelevation Speed

(ft) (degree) (in) (miles/h)

0 0.0 0 125

917 0.0 0 125

1022 1.1 6 125

11423 1.1 6 125

11528 0.0 0 125

26247 0.0 0 125

New Worn

Wheel Amtrak Standard Measured

Measured on straight track

Rail AREMA136 Measured on curved track
32
33

Figure 2.2.2-d: Wheel and rail profiles used for the simulations

Wheel-Rail Contact Geometry

The calculation of wheel-rail contact geometry determines the following contact parameters as a 

function of wheelset motions: contact patch location, contact patch dimensions, rolling radius at 

contact patch, and contact angles at contact patch. These contact parameters are used to 

determine the normal and tangential (creep) forces, based on rolling contact mechanics between 

the wheel and rail.

Another important contact geometry parameter is the equivalent conicity which is important for

determining the truck hunting speed (critical speed). Larger equivalent conicity may cause lower 

critical speed. 

Different wheel-rail profile combinations will result in different contact parameter functions. 

They will have different impacts on the kinematics and mechanics between the wheel and rail. 

Therefore, they will have different effects on the locomotive running performances.

For the contact geometry calculation, the following parameters were used: a standard gauge of 

56.5 in, rail inclination of 1/40, wheel back-to-back distance of 53.188 in, wheel diameter of 40 

in, and wheel flange angle of 75 degrees.
34

Figure 2.2.2-e: Parameter definition for contact geometry calculation

Contact geometry results

Figures 2.2.2-f through 2.2.2-j show some of the results of the contact geometry calculations. 

Note that, in general, the contact geometry is nonlinear. This is especially true for worn wheelrail profile combinations where the contact geometry parameters are highly nonlinear functions 

of the wheelset lateral displacement (see Figure 2.2.2-i and Figure 2.2.2-j). In order to accurately 

capture the nonlinear characteristics, a generalized nonlinear wheel-rail contact geometry module 

must be used for the dynamic simulations.

Figure 2.2.2-f: Contact geometry of new Amtrak standard wheel profile and new 

AREMA136 rail profile
35

Figure 2.2.2-g: Contact geometry of new Amtrak standard wheel profile and worn rail 

profile on straight track
36

Figure 2.2.2-h: Contact geometry of new Amtrak standard wheel profile and worn rail 

profile on curved track
37

Figure 2.2.2-i: Contact geometry of worn wheel profile and worn rail profile on straight 

track
38

Figure 2.2.2-j: Contact geometry of worn wheel profile and worn rail profile on curved 

track
39

Wheel-Rail Contact Mechanics

To describe the forces generated between the wheel and rail, the nonlinear rolling contact 

mechanics method was used. The creepage-creep force relations were determined by using 

Kalker theory. For calculation of the creep forces, a friction coefficient of 0.5 was used between 

the wheel and rail. 

Simulation of Design of Experiments (DOE) and Response Surface

A DOE simulation and response surface analysis is used to understand the sensitivity of the 

locomotive running performances to the suspension parameters and identify the critical design 

parameters for high-speed truck design. The regression functions of the response surfaces can be 

used for selection of design parameters and performance optimization.

Design variables of DOE

To study the impact of the suspension parameters on the performances of the selected truck 

concept, a simulation DOE was used to find the response “surfaces”—performances as functions 

of the characteristic parameters of the suspension system. Table 2.2.2-e below lists the 10 major 

suspension parameters selected as design variables for the DOE study. Figure 2.2.2-k shows 

where these parameters are located on the truck.

Table 2.2.2-e: 10 Major Design Variables for Modeling

Parameters Variable Comment

Primary Suspension—vertical stiffness/spring PS_Vert 2 per journal box

Primary Suspension—lateral stiffness/spring PS_Lat 2 per journal box

Primary Suspension—vertical damping/damper Pri_Damp 1 per journal box

Secondary Suspension—vertical stiffness/spring SS_Vert 2 per truck side

Secondary Suspension—lateral stiffness/spring SS_Lat 2 per truck side

Secondary Suspension—vertical damping/damper Ver_Damp 1 per truck side

Secondary Suspension—lateral damping/damper Lat_Damp 1 per truck side

Secondary Suspension—longitudinal (yaw) 

damping/damper

Long_Damp 1 per truck side

Traction link radial lateral & longitudinal 

stiffness/bushing

Rad_Bush 3 per link (journal 

box)

Traction link cocking stiffness/bushing Cock_Bush 3 per link (journal 

box)
40

The lateral stiffness of a coil spring can be determined by using a relation to its vertical stiffness. 

To simplify the DOE analysis, the lateral stiffness of the primary and secondary coil springs in 

Table 2.2.2 (see also Figure 2.2.2-k) was determined by

PS_Lat = αp* PS_Vert

SS_Lat= αs* SS_Vert

Thus, the independent design parameters for the DOE in the above table can be reduced to eight.

Figure 2.2.2-k: Parameters to be used for DOE analysis

MCAT performance parameters used as responses

Table 2.2.2-f lists the nine major performance parameters along with their specified limit values.
41

Table 2.2.2-f: MCAT Performance Parameters

Performance parameter Limit Comment

Wheel load reduction rate dQ/Q * ≤ 0.85 MCAT: V

V0

≥ 0.15

dQ = V0 − V,Q = V0

Wheel L/V * ≤ 1.13 = tan(δ)−0.5

1+0.5 tan(δ)

, δ = 75o

Net Axle Lateral Force * ≤ 0.47

kips

= 0.4 + 5/Va with Va= static 

axle load

Truck side L/V ≤ 0.60

Car Body Lateral Acceleration ** – Transient:

peak-to-peak * ≤ 0.75 g

for locomotive

Car body lateral Acceleration ** – Sustained: 

RMS ≤ 0.12 g

for locomotive

Car Body Vertical Acceleration ** – Transient: 

peak-to-peak * ≤ 1.0 g

Car body Vertical Acceleration ** – Sustained: 

RMS ≤ 0.25 g

Truck frame lateral Acceleration *** – RMS * ≤ 0.30 g

Notes:

*: These performances will be used as objectives during optimization.

**: The accelerometers shall be placed on the floor of the vehicle as near the center of a truck as 

practicable.

** The accelerometer shall be mounted on a truck frame at a longitudinal location as close as practicable 

to an axle’s centerline (either outside axle for trucks containing more than two axles), or, if approved by 

FRA, at an alternate location.

To simplify the DOE simulations, the list above was reduced to the following six critical 

performance parameters to build the response surface: 

• Max. wheel load reduction rate dQ/Q of leading axle

• Max. wheel L/V of leading axle

• Net axle lateral force of leading axle

• Car body lateral acceleration—transient peak to peak

• Car body vertical acceleration—transient peak to peak

• Truck frame lateral acceleration—transient peak to peak
42

Three parameters in Table 2.2.2-f were not included in the simulations. They are:

• Truck side L/V

• Car body lateral acceleration **—Sustained: RMS

• Car body vertical acceleration **—Sustained: RMS

These parameters are related to three other selected parameters: maximum wheel L/V of leading 

axle, car body lateral acceleration (transient peak to peak), and body vertical acceleration 

(transient peak to peak). After the suspension parameters are finalized, the three performance 

parameters not included can be confirmed against the specified limits to make sure these 

performances also meet the requirement. 

DOE Design

For the DOE design, GE DFSS (Design for Six Sigma) III 6-sigma optimization tools were used. 

In order to capture the nonlinear relations between the responses (performances) and the 

suspension parameters, a third-order response surface model was used. The DOE was determined 

by using the face-centered CCD (Central Composite Design) method based on an eight-factors, 

three-levels fractional factorial table. Each DOE contains 157 runs to establish the multiple 

inputs-outputs matrix, as shown in Table 2.2.2-g.

Table 2.2.2-g: DOE Table
43

Table 2.2.2-h shows the three levels of each suspension parameter in the fractional factorial table 

described above. For each case, the three levels are given by the minus 40 percent value, the 

nominal value, and the plus 40 percent value.

Table 2.2.2-h: 3 Levels of Suspension Parameters

Track disturbances used for DOE analysis

In addition to the track configuration (straight and curved tracks) and wheel-rail profile 

combination (new and worn profiles), track quality (track irregularities) also largely impacts the 

dynamic responses of the locomotive. The track irregularities specified in MCAT are 

deterministic disturbances whose amplitudes are the maximum allowable safety values for the

corresponding track classes. In reality, the track irregularities are random. 

To understand whether the MCAT track inputs can sufficiently represent the response 

characteristics of the locomotive, a comparison between the simulation results—using measured 

track inputs and the MCAT inputs—was conducted. The comparison results are shown in 

Figures 2.2.2-l and 2.2.2-m below. For this comparison, the random track irregularities measured 

on a Class 5 track were used. The comparison results show that for the same track class, the 

responses under MCAT inputs are much more severe than the responses caused by the measured 

random track irregularities. Therefore, it was felt that the MCAT inputs alone would be 

sufficient to represent the track qualities for the response surface study.
44
45

Figure 2.2.2-l: Comparison of the responses on curved track between MCAT track inputs 

and measured random track inputs

(1.9-degree curve, 6-inch superelevation, measured Class 5 track irregularities versus MCAT 

Class 5 inputs, frame-hung motor, loco speed 95 mph)
46
47

Figure 2.2.2-m: Comparison of the responses on straight track between MCAT track 

inputs and measured random track inputs

(Measured Class 5 track irregularities versus MCAT Class 5 inputs, frame-hung motor, loco 

speed 95 mph (MCAT Class 5 inputs on curved track are used for straight track))
48

Table 2.2.2-i shows the combinations of track configuration, track disturbances, and wheel-rail 

profiles for the DOE and response surface analysis. As explained above, the MCAT track inputs 

were deemed sufficient, and the wheel-rail input combinations identified with an X in table 

2.2.2-i below were chosen to provide sufficient range of combinations and fit within the budget 

and timing of this study.

As indicated in the discussion of the DOE design, each of the combinations marked with “X” 

below will require 157 simulations (runs) to find the response surfaces for the six specified 

performances. All together, this means that at least 6 x 157 = 942 simulations should be 

conducted. 

As the MCAT inputs are individual track irregularities, the responses will be transient, and a 

RMS evaluation is not possible. Hence, the outputs of the DOE simulation will focus on the 

evaluation of transient responses, meaning only peak values will be considered.

Table 2.2.2-i: Combinations of Track and Wheel-Rail Inputs for Response Surface Analysis

Optimization and Feasible Design Space

Based on the response surfaces and their regression functions, as determined by the DOE 

simulations, an optimization of the performance responses will be conducted to find the best 

suspension parameter combinations. As indicated before, this is an optimization which handles 

multiple design parameters and multiple objectives. Moreover, the optimization has to cover the 

different track configurations (straight and curved tracks), as well as different wheel-rail profile 

combinations. The suspension parameters which result in optimized performances on straight 

track may not necessarily be good, or even acceptable, for curved track. Similarly, the 

suspension parameters which can provide optimum performance for new wheel-rail 

combinations may not necessarily be the best suspension parameters for worn wheel-rail 

combinations. Therefore, a simple approach that involves looking for the common parameter 

space which can meet all the performance requirements under all track conditions and wheel-rail 

combinations will be used, as shown in Figure 2.2.2-n.
49

Figure 2.2.2-n: Strategy of determining feasible design space

2.2.3 Results

Presentation of results 

The following plot formats are used to describe the results:

Response surface fit

In general, the response surfaces are fitted first by using third-order polynomial functions of the 

design variables. Depending on the characteristics of the responses, the effect of going with a 

higher order of some of the design variables or their combinations may not be significant. In that 

case, the fit is then approached with a lower order, like second-order or linear functions to 

capture the major effects of the design parameters. The quality of the fit is controlled by the Rsq, as indicated in Figure 2.2.3-a. In general, the response surfaces are close to or higher than 97

percent.
50

Figure 2.2.3-a: Regression functions of the response surfaces

Pareto plot of coefficients of regression functions

A Pareto plot is used to visually demonstrate the effect of different design parameters on the 

selected performance. According to the design of the GE DFSS GEN III tool, the three levels of 

each design variable are normalized to (-1, 0, 1) and the vertical axis of the Pareto plot represents 

the value of 2*abs (coefficient) of the corresponding design parameters or their combinations on 

the horizontal axis, as shown in Figure 2.2.3-b. It is a simple indication of which term (parameter 

or parameter combinations) in the regression function(s) may have the greatest effect on (or 

contribution to) the specific performance for the specified variable range (in the normalized case, 

the range is 2 for each variable). It can help easily identify the most important parameters for the 

suspension design. It should be noted that the Pareto charts plot only the absolute value of the 

coefficients.
51

Figure 2.2.3-b: Pareto plot

Sensitivity plot

The sensitivity plot is used to demonstrate how the change of an individual design parameter will 

impact the performance. It is a quasi-2D or 3D plot of the response surface in which all the other 

parameters will be fixed to their center-point (level 2) values in the DOE table, as shown in 

Figure 2.2.3-c.

Figure 2.2.3-c: Sensitivity plots

Results of Stability Analysis

The proposed concept of a two-axle truck with frame-hung motor was analyzed first for its 

stability. A linear stability analysis—using the equivalent conicities that correspond to the three

wheel-rail profile combinations—was conducted to predict the preliminary critical speeds. As the 

results in Figure 2.2.3-d show, the linear critical speeds of the frame-hung concept are well 

above the axle-hung concept, and more importantly, well beyond the required maximum 

operation speed of 125mph (55.9 m/s). The illustrated curves of critical speed vs. equivalent 

conicity were calculated based on the values of the center levels of the suspension parameters. 

For reference purpose, the results of the truck with axle-hung motor are also showed in the 

figure. Its critical speed is much lower than the frame-hung motor concept.
52

Figure 2.2.3-d: Stability analysis of the selected concept with frame-hung motor

As indicated in the discussion of wheel-rail contact geometry in Section 2.2.2, the contact 

parameters are nonlinear functions of the lateral displacement of the wheelset. In general, they 

are not constant. This is the case, especially, for worn wheel-rail profiles. The results of the 

linear stability analysis are only a prediction of the critical speed for real wheel-rail profile 

combination. To confirm that the critical speed of the concept meets the speed requirement, 

nonlinear analysis was conducted by simulating the locomotive with real wheel-rail profiles 

through a track section with a single lateral track disturbance (half sine wave with amplitude of 5

mm). Figures 2.2.3-e and 2.2.3-f show the results of the nonlinear analysis. The nonlinear critical 

speed for truck with axle-hung motor is at least 156.5 mph (75 m/s), and at least 161 mph for 

truck with frame-hung motor. Both are well above the prediction results of the linear stability 

analysis.
53

Figure 2.2.3-e: Lateral displacement of leading axle—truck with axle-hung motor
54

Figure 2.2.3-f: Lateral displacement of leading axle—truck with frame-hung motor

Figures 2.2.3-g(a) and 2.2.3-g(b) below show the Pareto and sensitivity analysis results of the 

critical speeds using linear stability analysis. As indicated, the critical speed is relatively 

sensitive to the longitudinal and lateral stiffness (Rad_Bush) of the axle guiding device (threepoint primary traction link), yaw damping (Long_Damp), and the secondary vertical damping 

(Vert_Damp).
55

Figure 2.2.3-g: Pareto and Sensitivity analysis of linear critical speed – 2-axle truck with 

frame-hung motor, worn wheel, and worn rail profiles

Results of other performances:

Derailment coefficient L/V, wheel load reduction rated Q/Q, track shift force, vertical and 

lateral car body accelerations, lateral truck acceleration

Based on the designed response surface DOE in Table 2-7 and the parameter levels in Table 2-9, 

simulations were conducted for each of the track and wheel-rail profile combinations in Table 2-

9. Figure 2.2.3-h shows examples of the simulation results of some of the performances in time 

domain in which the corresponding MCAT inputs are also indicated for correlation and 

understanding. As indicated, the responses of each performance parameter to different track
56

inputs vary. To develop the response surface of the specific performance, only the maximum 

value of the response amplitudes on the specified track will be taken for the DOE point. For 

example, the maximum value of the lateral car body acceleration happens at the location of the 

“single surface” input which introduces both vertical and cross level disturbances to the 

locomotive. That maximum value will be taken as the response of the lateral car body 

acceleration for the simulated DOE point (in this case, it is the DOE point at which each of the 

suspension parameters has the value of level 2).

Figure 2.2.3-h: Example of simulation results—frame-hung motor, class 7 MCAT straight 

track inputs, new wheel-rail profiles, 125 mph, middle points (level 2) of the suspension 

parameters

The Pareto and sensitivity analysis results based on the response surfaces of the six performance 

parameters for various wheel and rail profiles, as well as the track conditions, are shown in detail 

in Appendix D. In general the response surfaces are nonlinear, and the effects of different design 

parameters on the performances vary. The performances may be sensitive only to some of the 

design parameters. Figure 2.2.3-i illustrates some examples of the sensitivity analysis results for 

the new wheel and worn rail profile combination. Figures 2.2.3-i (a) and (b) are lateral car body 

and lateral truck accelerations on straight track, and Figures 2.2.3-i (c) and (d) are wheel L/V and 

track shift forces on curved track. Note that increasing the radial stiffness of the bushings in the 

axle guiding devices (equivalent to longitudinal and lateral stiffness of the guiding device) can 

reduce the lateral car body and truck accelerations, and therefore improve the stability on straight 

track. But such an increase would negatively impact the derailment coefficient on curved track. 

One very interesting observation in Figures 2.2.3-i (a) and (b) is that the lateral car body and 

truck accelerations are relatively sensitive to primary and secondary vertical stiffness. This 
57

seems strange at first glance, but by looking at the simulation model, it is understood that this is

due to the setup of the relationship between the lateral stiffness and the vertical stiffness in the 

primary and secondary suspensions: lateral_stiffness/vertical_stiffness = c. Increasing the 

primary and secondary vertical stiffness will also increase their lateral stiffness and therefore 

impact the lateral vibrations of the locomotive. This special setup in the model should be 

carefully considered while interpreting the results in order to avoid any confusion.

Figure 2.2.3-i: Sensitivity analysis results for truck with frame-hung motor, new wheel, and 

worn rail profiles, class 7 track, 125 mph
58

It seems that the secondary vertical damping (Vert_Damp) has great influence on the 

performances, especially on the lateral car body acceleration, due to its influence on the roll 

motion of the car body. Increasing the secondary vertical damping can reduce both the lateral car 

body and truck accelerations. It can also help reduce the L/V ratio and the track shift force. But it 

will increase the vertical car body acceleration and wheel load reduction rate, as shown in Figure 

2.2.3-i (e) and (f).

As indicated before, different wheel-rail profile combinations significantly impact dynamic 

performances. Figures 2.2.3-j (a) to (h) below show comparisons of some of the sensitivity 

analysis results between profile combinations of new wheel and worn rail, and worn wheel and 

worn rail. Due to different wheel-rail profile combinations (and the different contact geometry 

parameters, see Figures 2.2.2-g to 2.2.2-j), the sensitivity of the performances to the design 

parameters is also different. The effect of the secondary yaw damping (Long_Damp) becomes 

more significant for the worn wheel and worn rail combination, as indicated below in Figures 

2.2.3-j(b), (d), (f), and (h).
59

Figure 2.2.3-j: Comparison of sensitivity results between different wheel-rail profile 

combinations (frame-hung motor, class 7 track, 125MPH)

The secondary lateral damping also has great influence on both the lateral accelerations on 

straight track and the curving performance on curved track. Some of the parameters even change 

or reverse the trends of the influence on the performances. Examples of this reversed trend

include the secondary lateral damping in the above Figures 2.2.3-j (c) and (d), the secondary 

vertical damping in (e) and (f), and the secondary vertical stiffness in (g) and (h). 

Some of the conclusions observed for the new wheel and worn rail profile combination may no 

longer be applicable to the worn wheel and worn rail profile combination because the 

performance parameters are so dependent on the wheel-rail profiles, which change constantly 

during the normal operation. The selection of the design parameters is much more challenging. 

To meet all the dynamic performance and running safety requirements, a systematic tradeoff and 

optimization between the performances and design parameters for the various wheel and rail 

profile combinations will be necessary.
60

Feasible Design Space

To develop the feasible design space of the suspension parameters, an optimization process is 

used. It considers the six selected dynamic performances as optimization objectives and the 

suspension parameters as design variables. The allowable lower and upper performance limits, as 

specified in MCAT, and the engineering feasible parameter variation ranges of +/-60 percent of 

the nominal values in the DOE analysis (see Table 2.2.2-h) are used as constraint/boundary 

conditions. The design space of the suspension parameters will be determined by finding all the 

possible parameter combinations when minimizing and maximizing the performances 

(objectives) within (or up to) the allowable performance boundaries. 

As indicated in the previous sections, the performance responses largely depend on the track 

configurations (straight or curve) and wheel-rail profile combinations (new or worn). Therefore, 

the above stated procedure is performed for each of the track and profile combinations. The 

design space which will meet all track and profile combinations is then determined by the 

method, as illustrated graphically in Figure 2.2.2-n. 

In general, the allowable variation range of one design parameter within the feasible design 

space will depend on the ranges of the other parameters. They interact with each other through 

nonlinear functions, as illustrated in Figure 2.2.3-k. For simplicity and engineering applicability, 

only a fixed range of each parameter will be reported in this study.

Figure 2.2.3-k: Example of 2 dimensional design space
61

Figures 2.2.3-l through 2.2.3-o provide examples of the minimization process of several 

performance parameters and the corresponding variations of some of the suspension parameters. 

The design space determined to meet both critical speed (stability) and safety performance 

requirements on straight and curved track is shown in Table 2.2.3-a below.

Figure 2.2.3-l: Minimization process of lateral car body acceleration in m/s^2 (frame-hung 

motor, straight track, new wheel, worn rail, MCAT class 7 track, speed 125 mph)

Figure 2.2.3-m: Minimization process of lateral truck acceleration in m/s^2 (frame-hung 

motor, straight track, new wheel, worn rail, MCAT class 7 track, speed 125 mph)
62

Figure 2.2.3-n: Minimization process of L/V of leading axle (frame-hung motor, 1.1 degree 

curve, 6-inch superelevation, new wheel, worn rail, MCAT class 7 track, speed 125 mph)

Figure 2.2.3-o: Minimization process of track shift force of leading axle in N (frame-hung 

motor, 1.1 degree curve, 6-inch superelevation, new wheel, worn rail, MCAT class 7 track, 

speed 125 mph)
63

Table 2.2.3-a: Feasible stable and safe design space of truck with frame-hung motor

2.2.4 Summary

The dynamic running performance of the locomotive depends not only on the suspension 

parameters, but also on the track conditions and wheel and rail profile combinations. To meet the 

running safety requirements, the suspension design of the truck system should consider all the 

possible wheel and rail profile variation ranges for the specified operation conditions on both 

straight and curved tracks. Due to the conflicting requirements of suspension parameters between 

stability on straight track and derailment safety on curved track, a thorough tradeoff of the design 

parameters must be conducted. The tradeoff can be achieved through optimization of the critical 

safety related performance responses by setting realistic constraint and boundary conditions for 

both the design parameters and performance requirements. 

2.2.5 Conclusions and Recommendations

The design space method developed in this study by using multi-objective optimization can 

provide both optimized concept of the suspension parameters and the feasible concepts which 

meet all the running safety requirements for both straight and curved track conditions, as well as 

specified wheel-rail profile ranges. This offers flexibility for engineering design. Different 

design concepts for the suspension system can be selected within the design space to achieve 

specific requirements. 

For the curving performance analysis in this study, a large curve radius (1.1 degree) based on the 

MCAT specification was used in order to meet the 125 mph speed requirement. In practice, there 

may be sharper curves with smaller radii, either on a high-speed line where the train will slow 

down, or in cases where the high-speed locomotives may run on freight lines with smaller 

curves. It is recommended that future analyses include smaller curves to evaluate the derailment 

safety and track shift forces for a more complete range of operating conditions. 

In this study, the primary lateral stiffness and secondary lateral stiffness are coupled with their 

corresponding vertical stiffness based on the physical products. This coupling assumption may 

limit the capability of the optimization, and the interpretation of the sensitivity results of such 

analysis is difficult and sometimes confusing. It is recommended that in future analyses the 

lateral stiffness be decoupled from its vertical stiffness.
64

As demonstrated in this study, truck stability and safety performance are significantly influenced 

by the worn wheel-rail profile combinations. Depending on how severely worn the profile is and 

which worn profile combination is used, the response behavior of the locomotive could be very 

different. There are a huge number of possible profile combinations and no standardized 

combinations within the industry. It is difficult, then, to determine which profiles should be 

used—and impossible to compare designs based on different profile combinations. It is 

recommended that consideration be given to developing an industry “standardized” worn wheel 

and rail profile databank based on typical measured profiles. 

It is also recommended that consideration be given to establishing a maximum allowable 

equivalent conicity for specified speed ranges. This would ensure that the wheel and rail profiles 

are maintained within a range that assures safe operation.

Additional recommendations are as follows:

(1) Develop and define standardized track irregularity inputs for corresponding track classes 

based on real measured track data.

(2) Develop standardized dynamic load (g-load) inputs for strength design of truck frame and 

attached components, including maximum dynamic loads and fatigue loads.
65

2.3 Task 3—Dynamic Load Summary Report

2.3.1 Definition

The intent for Task 3 was to evaluate the impact of locomotive weight, motor suspension 

method, and other suspension parameters on P2 force and track lateral shift force. This task was

divided into two sections: the first for the P2 force evaluation and the second for track lateral 

force. During the Task 3 analysis it became apparent that certain limitations of the Task 2 

analysis prevented adequate exploration of P2 force and track lateral force. For the reasons 

explained at the end of this Task 3 report, the lateral track shift force analysis was removed from 

the study results.

P2 Force

P2 force is a vertical force between the wheel and rail. The limit for the next generation diesel 

passenger locomotive is set by the PRIIA 305-005 specification at 82,000 lbf. This limit will be 

much more challenging to meet than past limits, especially since it must also be met at 125 mph. 

The limit applies to defined track parameters also specified by PRIIA 305-005. In order to make 

a meaningful P2 force comparison between past applications and projected PRIIA applications, 

the P2 force of existing or past applications must be recalculated using the PRIIA track 

parameters. 

The P2 force equation is defined in British Railways Board Group Standard GM/TT0088 Issue 1, 

Rev. A.

( ) T u

T u T

T

u t

u k m

k m m

c

m m

m P P v 















+ − + = +4 2 0 2 1 π

α

Where

P0 Static wheel load in pounds

α Dip angle in radians

ν Vehicle speed in inches/second

mu Unsprung mass per wheel in lbf/in/sec2
66

mt Track mass per wheel in lbf/in/sec2

kt Track stiffness per wheel in lbf/in

The PRIIA 305-005 specification provides track parameter values as follows:

α = 0.0085 Total dip angle in radians based on 1⁄2 degree on both sides of the dip

mt = 1.1335 Track mass per wheel in lbf/in/sec2 for concrete tie track

ct = 671 Track damping per wheel in lbf/in/sec for nominal track conditions

kt = 392,900 Track stiffness per wheel in lbf/in for nominally stiff concrete tie track 

(corresponding to track modulus of 5,100 lb/in/in, assuming a track 

deflection of 0.084 in under a 33,000-pound wheel load)

2.3.2 Approach

Since P2 force limit is tied to specific track parameters defined in the PRIIA specification, the 

vehicle builders can only influence two parameters: the overall vehicle weight and the unsprung 

weight of each wheel set. Coinciding with the more stringent P2 force limit for the next 

generation passenger locomotives will be the additional requirements of meeting the Tier 4 

emissions standards and stricter crash worthiness standards, each of which will add significant

additional weight to the locomotive. The locomotive manufacturers will have to reduce the 

unsprung weight of each wheel set. Since approximately one-third of that weight comes from 

partial support of the traction motors, removing the traction motor from the axle is the most 

practical way to achieve a significant reduction in unsprung weight.

This section reviews the PRIIA P2 force specification and presents the choices available to the 

OEM locomotive builders.

PRIIA Loco Assumptions

Locomotive weight

PRIIA locomotives are assumed to need Tier IV compliant engines which are expected to require 

“after treatment solutions” and will increase the overall weight of the vehicle. Additionally, new 

crashworthiness requirements necessitate additional weight in the car body structure.

Existing passenger, 4-axle, locomotives used in North America range from 260,000 lb to 

295,000 lb, as indicated by Technology Vehicle Report issued in August 2011 by the Next 

Generation Equipment Committee. Considering the above weight challenges, it is reasonable to 

evaluate a range of locomotive weights between 260,000 lb and 300,000 lb.

Unsprung Weight

Unsprung weight is the portion of the wheel set weight in direct contact with the rail. Portions 

which are supported by the primary suspension are sprung weight. Unsprung weight is typically 

calculated from the following components for each wheelset:

Basic

2 Locomotive wheels

1 Locomotive axle

2 journal bearings

2 journal bearing housings

4 Primary springs – 1⁄2 weight (if used)
67

Optional

For Axle-Hung motor designs 

1 Gear

2 motor suspension bearings

1 motor suspension bearing housing (u-tube)

1 Traction Motor (unsprung portion)

Disc Brake rotors (if used)

For Frame-Hung motor designs 

1 Gear (if mounted on the axle)

1 Gearbox (unsprung portion – if mounted on the axle) includes bearings

1 Quill Drive (unsprung portion – if used)

Optional components need to be considered depending on the concept used: axle-hung versus 

frame-hung traction motors, drivetrain type and brake type. Considering the components 

described above it is likely that the unsprung weight will range from 5500 lb to 8500 lb.

2.3.3 Results

Resulting Unsprung Weight Requirements

Using the P2 force formula we can quickly calculate the expected P2 force for different weight 

locomotives and different unsprung weights for proposed PRIIA locomotives. Additionally, a 

comparison can be made between P2 forces on proposed locomotives and existing locos. The 

value of such a comparison can be appreciated when considering that past P2 force calculations 

have used different track parameters. Since, typically, P2 force limits are based on some vehicle 

which is considered acceptable, using different track parameters will result in a different P2 force 

on that vehicle and therefore a different P2 force limit.
68

Figure 2.3.3-a: P2 force for PRIIA loco compared with existing diesel rolling stock

Amtrak currently utilizes both AC and DC Genesis, while MBTA will soon be taking delivery of 

new HSP-46 AC locos. These locomotives can be operated at a maximum of 110 mph while 

PRIIA locos are expected to run at 125 mph. Figure 2.3.3-a shows that by utilizing the weight 

ranges described in the previous section for both the locomotive overall weight and the unsprung 

weight along with the track parameters provided in the PRIIA 305-005 specification, the existing 

locomotives do not meet the PRIIA P2 force limit even at the slower operational speed of 110

mph. Historically, Amtrak has used the F40 locomotive as a baseline for P2 force calculations. 

It appears from the figure above that the current P2 force limit is based on the P2 force generated 

by an F40 loco operating at 110 mph. It should be clear that future PRIIA locomotives weighing 

260,000 lbf (F40), or more, will need an unsprung weight of approximately 7000 lbf, or less, to 

stay below the P2 force limit. This likely includes all conceivable future PRIIA locomotives.

So, the obvious question is “What unsprung weight can be tolerated for a given locomotive 

weight?” Since the PRIIA standard provides no guidance for benefits of P2 forces lower than the 

limit, it is reasonable to assume that, at least from design standpoint, it is sufficient to only meet 

the specification rather than attempt to exceed it. Given the current technology and challenging 

design tradeoffs facing North American locomotive builders, just meeting the minimum 

requirements will be difficult and expensive.

For purposes of analysis then, it makes sense to calculate the maximum allowable unsprung 

weight that will result in a maximum P2 force of 82,000 lbf for a given locomotive weight. 

Figure 2.3.3-b, does this for 125 mph operation using the PRIIA track parameters.
69

Figure 2.3.3-b: P2 force for PRIIA loco: Loco Weight versus unsprung weight, 125 mph

Closer inspection of this chart suggests that if the lightest unsprung weight for an axle-hung 

motor design is 7500 lbf, then the overall locomotive weight needs to be approximately 244,000 

lbf. This is not considered a likely solution, as there is currently no viable concept for a diesel 

powered, Tier 4, and crashworthiness compliant locomotive that can achieve such a light weight.

As discussed in Task 4 of this study, it is likely that because of the higher braking energy 

requirements, the new heavier and faster PRIIA locomotive will need to add a disc brake system

to the tread brake system currently already on typical U.S. passenger locomotives. On an axlehung design, this addition will further increase the unsprung mass, causing the P2 forces to 

increase beyond the values of existing stock. This further adds to the need for an even lighter 

locomotive.

Based on the preceding, we conclude that the next generation 125 mph passenger diesel will 

have to use frame-hung motors in order to reduce the unsprung weight.

Class 6 or Class 7

The target speed of 125 mph can only be achieved on class 7 track. Currently, class 7 track in 

the United States is limited to short sections in the North East Corridor. The fastest diesels are 

limited to 110 mph on class 6 track. There is some belief in the industry that due to lack of 

infrastructure, the PRIIA specification should be tailored to 110 mph operation, or phased in later 

as high-speed infrastructure becomes available. If this were the case, and the P2 force limit was 

maintained at 82,000 lbf, then Figure 2.3.3-b could be modified for 110 mph operation. Now, as 

shown in Figure 2.3.3-c, an axle-hung traction motor design becomes more plausible. Of the 

locomotives compared previously in Figure 2.3.3-a, only the 260,000-pound F40 locomotive 

meets the P2 limit, even at 110 mph.
70

Figure 2.3.3-c: P2 force for PRIIA loco: Loco weight versus unsprung weight, 110 mph

In light of the earlier conclusion that an axle-hung motor design is not conceivable for a 125 mph 

locomotive, the remaining comparisons will assume that an axle-hung motor design is based on a 

110 mph locomotive. For tracking purposes, this locomotive will have 260,000 lbf overall 

weight with 8,500 lbf unsprung weight, just like the F40.

Note that the F40PH has a horsepower rating of 3200 while the PRIIA locomotive power 

requirement is estimated to be as much as 4500 HP to achieve and maintain 125 mph. This will 

almost certainly result in a heavier engine and engine support systems. 

Effect of track parameters on P2 force

The previous section discussed the effect of locomotive parameters on P2 force. It may be of 

interest to compare the effect of locomotive weight and unsprung weight with the effect of track 

parameters that impact P2 force: Dip angle, track mass, stiffness, and damping. Different values 

have been used over the years and so it is important to understand the effect of changing these 

values.

Dip Angle

The dip angle selected in the PRIIA specification is 0.0085 radians. It could be argued that this 

is a conservative number especially for 125 mph operation on class 7 track. Regardless, other 

publications use both higher and lower values ranging from 0.003 to 0.01. Higher values of dip 

angle have a tendency to increase the P2 force.

Effective Track Mass

The effective track mass in the PRIIA specification is 1.1335 lbf-sec2

/in. This value is consistent 

with track made with concrete ties. Values in literature range from 1.1335 to 2.592 lbf-sec

2

/in. 

Higher values of track mass have a tendency to reduce the P2 force.
71

Effective Track Stiffness

The effective track stiffness in the PRIIA specification is 392,900 lbf/in. This value is defined as 

a value for track made with concrete ties. Upon closer inspection, this value appears to be 

somewhere in between values for wood and concrete ties used in the past by Amtrak. 

Values used in the past have been in the 265,000–650,000 lbf/in range for wood and concrete tie 

tracks, respectively. Higher values of Kt have a tendency to increase the P2 force.

Effective Track Damping (Ct)

The effective track damping in the PRIIA specification is 671 lbf-sec/in. This value is two to 

five times higher than values used in the past. Values used in the past were 121 to 205 lbf-sec/in. 

The example in the GMTT/0088 is 55.4 x 103 Ns/m which is equivalent to 316 lbf-sec/in. 

Higher values of Ct have a tendency to reduce the P2 force.

Figure 2.3.3-d shows a Main Effects plot derived from 729 combinations of 6 parameters, each 

of which has three levels (36=729). The P2 force is calculated for 125 mph operation. The 

center value for each parameter is the PRIIA value while the other two enveloping values are 

from the ranges described above. (For parameter symmetrical ranges, the center point will lay 

on the red dotted line which represents the 82,000 lb, while unsymmetrical ranges like the dip 

angle and Ct will have the center point offset from the line even though those center points 

represent the limit.) By utilizing the ranges described above for each parameter, it becomes clear 

that the dip angle has the most far reaching effect on P2 force on the track side, while unsprung 

weight is the parameter to optimize on the vehicle side.

Figure 2.3.3-d: Main Effects plot for parameters used in the P2 Force equation for 125 mph 

operation

Tracks with concrete ties will have more mass and may exhibit more damping, but will also be 

typically stiffer than tracks with wood ties. 

It is reasonable to assume that the dip angle will vary with class of track due to tighter geometry 

standards, with smaller angles on higher classes of track. As such, perhaps it is worthwhile to 

compare the P2 forces exerted by a given vehicle on class 7 track at 125 mph with class 4 track 

at 80 mph. Figure 2.3.3-e shows such a comparison where classes of track correspond to the 

maximum allowable speed and the rail joint dip angle gets smaller as the track gets better.
72

Figure 2.3.3-e: Comparison of potential P2 forces on different classes of track

Assuming that class 6 and class 7 tracks are dedicated for passenger transport, it is conceivable 

that these tracks could be maintained to higher track standards. If this were the case, operation 

on class 4 track, at 80 mph, could produce values higher than the high-speed operation. Since 

class 4 track is most likely going to be shared by freight traffic, it is reasonable to want to 

understand which type of vehicle will produce higher P2 forces.

The comparisons in Figures 2.3.3-f and 2.3.3-g below are based on the following locomotive 

configurations:

Loco Overall Weight (lbf) Unsprung Weight (lbf) HP (reference)

F40 260,000 8,500 3,200

Genesis DC 268,000 9,500 4,250

Genesis AC DM 275,000 8,500 3,200

GE EVO DC Freight 436,600 10,725 4,400

HSP-46 290,000 8,500 4,650

PRIIA Axle-Hung (110 mph) 260,000 8,500 4,500

PRIIA Frame-Hung 285,000 6,500 4,500

Keep in mind the following comparisons are for purposes of analysis and discussion of options. 

The authors do not know of a viable path to reduce the PRIIA locomotive weight, with 

crashworthiness structure and Tier 4 engine components included, down to 260,000 lbf. 
73

Figure 2.3.3-f: Comparison of P2 forces on class 4 track, at 60 mph, for freight and 

passenger locomotives

Figure 2.3.3-f shows that at the same 60 mph speed, both of the AC and DC Genesis units, 

because of their lower weight, have lower P2 forces than the heavy-haul Evolution freight 

locomotive. But maximum speed for passenger vehicles is higher than for freight vehicles so a 

comparison should be made at the maximum speeds allowed.

Figure 2.3.3-g: P2 force – passenger versus freight for Class 4 Track operation

Figure 2.3.3-g shows that passenger locos with axle-hung traction motors exert higher P2 forces 

than freight locos on the same track, mostly due to differences in operating speed. The new 
74

PRIIA locomotive with axle-hung motors will be better than the old Genesis because the AC 

motor is lighter than the DC motor, resulting in lower unsprung weight. The overall loco weight 

would have to go down to 256,000 lbf on the axle-hung design in order to match P2 forces 

generated by heavy haul freight locomotives. In contrast, the 125 mph frame-hung motor design 

exerts lower track forces on class 4 track than the 110 mph axle-hung motor design, as well as 

lower forces than the freight locomotive. This further reinforces the conclusion that frame-hung 

motors will be required on passenger vehicles to meet the PRIIA specifications. This will also 

be true at 110 mph. As stated above, we do not know if a viable path exists to reduce the PRIIA 

locomotive weight, with crashworthiness structure and Tier 4 engine components, to below 

260,000 lbf. At this time, it seems certain the next generation PRIIA certified locomotive will 

weigh considerably more than 260,000 lbf. 

2.3.4 Conclusions and Recommendations (P2 force analysis)

The current PRIIA P2 force specification uses F40 as a baseline locomotive. This locomotive 

has a relatively lighter 3200 HP engine and also light locomotive weight of 260,000 lbf. The 

Genesis locomotives, which make up the bulk of the Amtrak diesel fleet, are 4000HP to 4200HP 

(P40 and P42, respectively) and weigh 268,000 lbf. These locomotives will not meet the PRIIA 

required 82,000 lbf limit at 125 mph, nor at 110 mph.

Locomotive builders can influence two of the six parameters (not including speed) that affect P2 

forces, total locomotive weight and unsprung mass. Total weight is likely to increase, despite 

manufacturers’ efforts to lighten the total vehicle. In order to meet the new crashworthiness and 

emissions standards, as well as achieve the higher horsepower required to reach 125 mph, the 

new PRIIA locomotive will likely be heavier than the Genesis. This places additional pressure 

on reducing unsprung mass. We conclude that a frame-hung traction motor design will be 

required to limit the P2 force to 82,000 lbf on the next generation passenger locomotive. 

2.3.5 Lateral track shift force

Regarding lateral track shift force evaluation, Task 2 of this report included the track shift force 

evaluation [Figures 2.2.3-j (g) and 2.2.3-j (h) in Task 2] for a specific high-speed condition of a 

1.1 degree curve at 125 mph. This evaluation shows that the most sensitive parameters, the most 

notable being yaw damping, are those which affect dynamic behavior. At higher speeds,

dynamic effects begin to dominate the classical quasi-static behavior associated with curving 

analysis in tighter curves at lower speeds. Additionally, worn wheels, which tend to create more 

flange clearance and result in lower forces in tight curves, will drive the opposite effect at highspeeds, making the vehicle less stable.

The original intention for this section was to piggyback off the Task 2 analysis to draw 

conclusions about the effect of suspension design on track panel shift forces. During the Task 3 

analysis it became apparent that the limitations of the analysis in Task 2 also prevent an adequate 

exploration of this topic in Task 3. Additional analysis was performed to explore the question 

about performance in tighter curves, and although all simulation resulted in values that were 

below the proposed limit for Net Axle L/V, the analysis failed to sufficiently drive a conclusion

or any recommendation and therefore was not included in this report. 

As frame-hung motor designs enter the market, they affect either wheelbase or car body 

connection designs because the distance between the traction motor and the wheel set increases 
75

compared with axle-hung traction motor designs of the same capability. These changes play a 

role in the overall dynamic performance of the vehicle and performance of the track (lower panel 

shift forces). 

A recommendation for future work in this area is to explore the effect and interactions of 

wheelbase changes, car body connection type (traction pin versus traction link), and suspension 

design. The impacts of these design tradeoffs should be evaluated across a wider range of 

curvatures, including those utilized by freight traffic, as well as with new and worn conditions of 

wheel and rail profiles.
76

2.4 Task 4—Braking Capability Summary Report

2.4.1 Definition

Braking capability is a critical safety factor for a high-speed passenger locomotive. As operating 

speed increases, designing and packaging the brake components becomes more challenging. The 

energy that must be absorbed and converted to heat by the brake system increases with the 

square of the velocity. The need to absorb this additional energy necessitates adding material 

and weight to a bogie system that, in order to reduce P2 forces, requires a reduction in total 

weight and/or unsprung weight. 

Task 4 evaluates brake design requirements and likely characteristics of a brake system for a 

PRIIA specified diesel passenger locomotive operating at up to 125 mph.

Current Design Trends

Rail vehicles in North America primarily use air operated tread brakes in which the brake pad is 

applied directly to the tread of the locomotive or car wheel. Locomotives generally combine 

dynamic braking (DB) with the friction tread brakes, reducing the amount of thermal energy the 

friction brake must absorb. However, safety regulations require that the locomotive be able to 

make an emergency stop utilizing air brakes only. This study evaluates the requirements needed 

to meet the air brake only case, with no DB contribution.

Brakes convert kinetic energy into heat and the tread brake applies this heat directly to the 

wheels. At higher speeds, the wheels can reach temperatures which can change their 

metallurgical structure, resulting in wheel cracking and other wheel defects. Although 

transformation temperatures are above 1,300 degrees oF, generally, for design purposes, 

temperatures have been limited to something less than 1,000 degrees oF. To reduce wheel 

defects, this limit should be even lower. Since today’s existing rolling stock already reaches 

these temperature limits, in order to ensure acceptable braking temperatures, tread brakes must 

be supplemented, or replaced, by other means.

DB has been used on diesel locomotives for over 50 years with good success, but always as a 

secondary system which supplements the critical air operated friction brake. The friction brake 

system has to be designed to provide sufficient retardation for emergency stop, though such stops 

are very rare. From a duty cycle standpoint, DB still plays an important role in emergency 

stopping and its inherent wheel slip control provides opportunity for braking optimization.

In Europe, China, and Japan, disk brakes have been used extensively, especially on high-speed

trains. However, there are major differences between those vehicles and the one required for 

PRIIA. The most significant difference, of course, is weight, which requires larger discs on the 

heavier North American locomotives, further feeding the vicious cycle of performance strength 

and weight. Performance drives large traction motors which occupy space between the wheels,

leaving little to no space for axle mounted disc brakes. Heavier locomotives drive higher 

dynamic track forces which force the use of frame-mounted traction motors. This in turn creates 

the need for a flexible coupling between the traction motor and the wheel set, again adding 

weight to the vehicle. All this makes it difficult to simply replace tread brakes with disk brakes.

Diesel Locomotive

The diesel locomotive uses a diesel engine to produce electricity, which in turn is used to power 

electric traction motors. Although, originally, DC motors were applied to locomotives, AC 
77

motors have slowly become more popular, mostly due to absence of ground brushes which wear 

over time and need to be replaced. New generation passenger units will utilize AC traction 

motors. Compared to an electric-only locomotive, the diesel engine and its support systems,

such as cooling and lube oil, can double the total locomotive weight.

Comparison between passenger and freight

The differences in operating characteristics between passenger and freight diesel locomotives 

dictate differences in design requirements. Freight operations vary from heavy haul (coal, iron 

ore, etc.), which is typically lower speed and tighter curves, to intermodal (general merchandise, 

mail, etc.), which is generally faster—up to 80 mph in the United States—and with less tight 

curves. In short, freight locomotive bogies are designed to provide the maximum pulling 

capacity for the train, while retaining the ability to traverse tight curves.

Passenger locomotive bogies are ideally designed to operate at higher speeds with less emphasis 

on the ability to pull long trains with many trailing tons. The objective is generally to pull six to 

eight cars filled with passengers to keep scheduled arrivals and departures. In the U.S. Amtrak 

environment, of course, the truck and locomotive must be designed to operate on all classes of 

track and speeds, not just at high-speed and higher class tracks. 

Passenger Rail can be divided into commuter (< 50 mi) and intercity. Commuter trains make 

many stops while intercity make fewer. The duty cycle for the passenger locomotive braking 

system is very different from the freight duty cycle.

Comparison between European and North American passenger 

The North American Passenger Diesel locomotive will be twice as heavy as its European 

counterparts, due to infrastructure differences, crashworthiness requirements, and the heavy 

diesel engine. 

Europe has by far the most densely populated rail network in the world and utilizes passenger 

rail on a much larger scale than North America. Figure 2.4.1-a shows a map of rail network 

density around the world.

Figure 2.4.1-a: Rail network divided by area of the country

While the United States has more miles of track than any other country, this track is primarily 

built and used for freight service. The rail infrastructure limits freight traffic to 80 mph and 
78

passenger traffic to 90 mph. Several special corridors exist where the rail infrastructure is 

dedicated to passenger service and allows speeds up to 110 mph. There are a few relatively short 

sections that are electrified with the infrastructure to support speeds up to 150 mph (ACELA). 

Prior to World War II the United States had electrified over 20 percent of its rail network, but 

today electrification is limited to the North East Corridor and some metropolitan areas like New 

York City and Chicago. 

In contrast, almost 50 percent of Germany’s rail system is electrified. This allows use of EMU

which has advantage over head-end-powered trains because it distributes the required train 

power across more vehicles and more axles. With respect to truck design, electrification enables 

use of smaller traction motors which results in lower vehicle weight and lower unsprung weight.

Existing 125 mph head-end diesel locomotives

So where is there a precedent for 125 mph service for a head end diesel locomotive? Prototypes 

of Spanish Talgo XXI and Russian TEP80 locomotives were built but never mass produced. 

British Rail Class 43 HST is the only diesel put into actual service and still running today. 

Though 197 units were produced, these units have some major differences from the proposed 

PRIIA locomotive. The units weigh less than 155,000 lb (17.5 metric tons per axle) thanks 

mostly to a lightweight high-speed engine and lack of current FRA crashworthiness 

requirements. As a result, less energy is required to stop the Class 43 HST than is required to 

stop its future U.S. counterpart.

In the United States, NJT has purchased 26 APL-45DP (dual mode) locomotives which weigh 

288,000 lbf (33 MT) and are the best example of a working configuration for North American 

passenger locomotives. Though they operate at only 100 mph in diesel mode, they are capable 

of 125 mph in electric mode. These units utilize both tread brakes and disc brakes and this 

section of the report describes in some detail why that is the case.

2.4.2 Approach

The Task 4 study compares braking energy needs for the PRIIA locomotive with other existing 

rail vehicles in North America and Europe. Additionally the task uses prior work (Gordon and 

Orringer) in the area of wheel heating to make comparisons between the vehicles used in that 

study and the future PRIIA locomotive, as well as between existing freight and passenger diesel 

locomotives.

Braking Energy Fundamentals

The basic premise of a locomotive air brake is that the kinetic energy of the moving vehicle has 

to be converted to friction heat. As such, one can calculate the kinetic energy of a vehicle based 

mainly on its speed and mass since Kinetic Energy (KE) is:

KE = 1⁄2 mv2

Where 

m = mass of the vehicle 

v = vehicle speed
79

Heat Flux

Knowing the total kinetic energy that needs to be converted to heat is not enough to predict 

wheel temperatures. Number of wheels, wheel diameter, and brake shoe width determine how 

much surface area will be used to absorb the heat. The rate of application will determine the 

power going into the wheels while the surface area will allow conversion of this power into heat 

flux. Figure 2.4.2-a shows the approach undertaken in this paper.

Figure 2.4.2-a: Approach for utilizing the Gordon and Orringer study to draw conclusions 

about brake requirements on PRIIA locomotive

The existing Gordon and Orringer research is extrapolated to reach conclusions about the 

braking considerations for the new PRIIA diesel locomotive. This research utilized a finite 

element thermal model to predict wheel tread temperatures based on some given heat flux input. 

Although the maximum wheel temperature depends on the rate of heat flux input, since 

deceleration rates are typically controlled for passenger comfort (1.6–2.0 mph/s) and since the 

highest heat flux will occur at the highest speed, calculating the maximum heat flux still allows 

the reader to draw the same conclusion.

The focus of this portion of the study is on brake energy dissipation and heat flux input. No 

actual wheel temperatures were calculated, as that is beyond the scope and funding of this report. 

It is felt that the earlier benchmarking sufficiently drives the conclusion.

While we believe that the general conclusions drawn here are well founded, more detailed 

analysis and calculations could provide valuable insight into specific design parameters and 

details that will have to be met and provided in a working system. This could be a helpful area 

for additional study.

2.4.3 Results

2.4.3.1 Brake Energy

Passenger versus Freight 

Though modern freight locomotives are always heavier than their passenger counterparts, they 

may also have more axles and operate at lower speeds. As the kinetic energy calculations below 
80

in Table 2.4.3-a indicate, freight locomotives require less total energy dissipation per locomotive. 

When the higher number of axles is taken into account, the difference is even more significant on 

an energy per wheel basis.

Table 2.4.3.1-a: Comparison of Kinetic energy between Freight and Passenger locomotives

Weight, 

max

Number of 

Axles Speed, max Kinetic energy

Freight locomotive 436,600 lbs 6 80 127 MJ

Existing Passenger locomotive 278,000 lbs 4 110 153 MJ

PRIIA Passenger locomotive 285,000 lbs 4 125 202 MJ

Typical freight locomotives weigh less than 420,000 lb and travel at speeds below 75 mph. 

Figure 2.4.3-a shows that, compared with the typical freight locomotive, the 125 mph PRIIA 

loco has to dissipate twice as much energy per wheel when stopping from maximum speed. The 

figure also shows O&G Arrow-III, a passenger car from NJT used in the Gordon and Orringer 

study discussed later in this section.

Figure 2.4.3.1-a: Energy absorbed by each wheel for passenger and freight locomotives

European versus NA passenger

European locomotives in general are lighter, faster, and typically set up in distributed power 

rather than head-end power. For a closer comparison, it is useful to consider existing diesel 

locos which run at 125 mph. One such loco is the British Rail Class 43:

http://en.wikipedia.org/wiki/British_Rail_Class_43_(HST)

Still, even this diesel unit is considerably lighter than the North American locomotives. Utilizing 

the same approach shown in Table 2.4.3-b, it is no surprise that the North American units, which 

are almost twice the weight of their UK counterparts, will need almost twice the braking force 

since kinetic energy is directly proportional to the weight of the vehicle.
81

Table 2.4.3.1-b: Comparison of Kinetic energy between Class 43 and PRIIA locomotives

Weight, 

max

Number of 

Axles

Speed, 

max

Kinetic 

energy

UK – Class 43 154,280 lbs. 4 125 109

North America – PRIIA Loco 285,000 lbs. 4 125 202

It is clear that North American and European operations are quite different. Technology used in 

Europe is not directly applicable to North America. In addition to the huge difference in axle 

load, there is also a difference in track quality, which leads to a fundamental difference in 

suspension design approach. In Europe, the standard is to provide a relatively stiff primary 

suspension and a relatively soft secondary suspension. In North America, which is dominated by 

freight traffic, an alternate approach consisting of a relatively soft primary suspension and a

considerably stiffer secondary suspension has been adopted. The soft primary suspension is 

more adaptable to rough track sections, while the stiff secondary helps with weight transfer,

which is important in heavy haul applications. This arrangement performs relatively well at slow 

freight speeds but at higher speeds may not be the optimal solution. This topic is covered in 

section 2 of this report.

2.4.3.2 Comparison of brake types and thermal dissipation

The general premise of any braking system is that the kinetic energy of the speeding vehicle 

must be converted to heat. This is typically done via friction brake or through DB, where the 

electricity produced by the traction motors working in DB mode is directed into resistor grids 

which dissipate the heat away from the wheels.
82

Figure 2.4.3.2-a: Locomotive brake types

Tread brakes

Tread brakes are pneumatically (air pressure) powered and are the most common method of 

stopping railway stock. Friction material in the form of a brake shoe is applied directly to the 

tread of the wheel, thus creating a retarding force directly at the wheel. One benefit of tread 

brake use is that it “conditions” the tread of the wheel, or to put another way, cleans it. 

The downside of this brake configuration is that the dissipated heat is mostly transferred into the 

wheel. Though quite acceptable at lower speeds typically found on freight service, for passenger 

operation, as speeds rise, so do the temperatures, and when the temperatures reach beyond 800 

oF, the wheels begin to run the risk of experiencing metallurgical changes which generally have a 

negative impact.

Wheel defects are a big concern for any railroad since broken wheels cause delays and may be a 

precursor to derailment. At higher speeds, consequences of derailment grow exponentially, 

especially on passenger trains. 

Generally, the Association of American Railroads (AAR) defines several classes of wheels, with 

softer wheels aimed at high tread brake applications and harder wheels at low tread brake 

applications (AAR M107/M208). There is evidence that harder wheels are more resistant to 

rolling contact fatigue (RCF) which generates micro cracks in the tread surface of the wheel and,

if gone unchecked, can develop into shattered rims.

Over the last decade, the railroads have been replacing their tracks with harder rail and also 

harder wheels, partially due to these findings on RCF.
83

Disk Brakes

Disc brakes come in several varieties based on where they are applied in the drivetrain. The 

rotor can be mounted directly on the locomotive axle, it can be housed around the web of the 

locomotive wheel, or it can reside on a dedicated shaft which is connected to the axle through a 

gearbox. The obvious benefit of a disc brake is that it reduces the thermal load on the wheel and 

typically provides a larger surface area for heat dissipation. Conversely, it is typically harder to 

package and increases not only overall weight but, more importantly, the unsprung weight (in the 

case of axle and wheel mounted rotors). On diesel locomotives, the traction motors are typically 

large so space is at a premium.

Figure 2.4.3.2-b: Locomotive wheel energy reduction resulting from disc brake use

Figure 2.4.3-c above compares braking energy absorbed by the wheels for a number of relevant 

applications. The blue bars are for the case of tread brake only. Note that the PRIIA application 

has the highest braking energy due to its higher speed. The red bar for that case shows the 

energy going into the wheel for the case of tread brake combined with disc brake.

Dynamic Brake

DB has gained popularity over the years, especially on AC locomotives which provide additional 

capability at lower speeds. Traditionally, DB was used as an auxiliary brake which reduced the 

demand on the friction brakes. Today it is still used in this fashion, but has gained more 

acceptance in the industry as a primary means of slowing down freight locomotives. Though 

freight locomotives generally have separate controls for dynamic and “air” brakes, certain 

passenger locomotive models utilize ‘blended brakes” where the operators only have a single 

control for the locomotive brakes (on top of the train line brake, which controls the brakes on the 

entire train). Figure 2.4.3-d shows that even though the air only operation has high energy 

dissipation demand on each wheel on the passenger application, when blended brakes are utilized 
84

the demand goes down below typical freight locomotive levels. This explains why the existing 

passenger Genesis units can utilize tread brake only solution at 110 mph. Even so, although 

emergency braking happens rarely, and the likelihood that DB failure will occur during an 

emergency application from 110 mph is remote, it is conceivable that such operation could 

induce some damage to the wheels, especially if one focuses only on the brake energy required.

Figure 2.4.3.2-c: Locomotive wheel energy reduction resulting from use of blended brake

Estimate of potential wheel damage from braking energy

It is beyond the scope of this paper to utilize finite element modeling to predict wheel 

temperatures under passenger applications. Instead, an alternative approach is used to draw 

conclusions without actually calculating the temperature. In 1995, Gordon and Orringer wrote a 

paper dealing with wheel heating on New Jersey Arrow-III cars, where they did utilize a finite 

element model to predict wheel temperatures. The Arrow-III cars are multiple unit powered cars 

which distribute traction and braking to multiple axles in the train consist. During the conversion 

from DC to AC motors, the railroad decided to reconfigure each married pair of cars to have 

three powered trucks and one unpowered truck, meaning that the powered trucks had DB and the 

unpowered truck did not. The study focused on increase in operation speed from 90 mph to 100 

mph and evaluated impact of auxiliary brakes in the form of both disc brake and DB. One of the 

inputs to the model was the heat flux going into the wheel tread surface. 

The study predicted temperatures of 945 degrees oF for braking from 100 mph on the unpowered 

truck and 1,157 degrees oF on the powered truck with DB disabled. When the brake was 

augmented by DB on three out of four trucks, the temperatures reduced to below 800 degrees oF. 

The tread brake retarding force at full brake pipe pressure is 7,000 lbf, However, the original 

configuration limited the retarding force to 5600 lbf. With the use of augmented auxiliary DB on 

three of the four trucks, it appears that the average retarding force from the tread brake was 
85

reduced although the peak force at 100 mph actually went up. The DB provided 26 percent

augmentation but was applied at a 3⁄4 ratio since only three of the four trucks were powered; in 

other words, the DB reduced the thermal strain on the wheels by 20 percent. 

It is interesting to observe that the DB augmentation reduced the tread brake demand under 

normal operation, which naturally includes the use of DB in a blended mode; however, it 

obviously did not change the brake demand for air only emergency stop. This suggests that 

emergency stopping with air only was not the main driver for the design decisions at NJT. In 

fact, disc brakes were also evaluated and actually provided best reduction of thermal input into 

the wheels, including air only emergency stops, but in the end were not selected.

Regardless of the above conclusion, the Gordon and Orringer study can be utilized to extrapolate 

the expected heat flux for the future PRIIA locomotive. Heat flux for the given case is 

dependent on the brake shoe force and the contact area. As such, this approach is a better way of 

comparing demand for thermal dissipation in a brake system than simply looking at brake 

energy. Since the Arrow-III cars uses smaller diameter wheels and smaller shoes than those used 

on a diesel locomotive, the relative heat flux going into those wheels is actually higher. Table 

2.4.3-c shows the comparison of parameters and relative heat flux between the Arrow-III cars 

and diesel locomotives.
86

Table 2.4.3.2-a: Comparison of Heat Flux input between Arrow-III cars and diesel 

locomotives

Arrow-III Genesis PRIIA Freight Hi-ad 

Vehicle Weight (lb) 139,000 278,000 285,000 436,600

# of axles 4 4 4 6

speed (mph) 100 110 125 80

wheel diameter (in) 32 40 40 42

brake shoe width (in) 2.5 3.38 3.38 3.38

contact area (in^2) 251 425 425 446

brake retarding force per truck (lb) 7,000 12,800 14,000 30,000

energy (MJ/wheel) 7.9 19.1 25.2 10.6

energy ratio 1 242% 320% 134%

heat flux (MW/m^2) 2.147E+06 2.555E+06 3.176E+06 2.766E+06

heat flux ratio 1 119% 148% 129%

Table 2.4.3-c above compares diesel locomotives and the Arrow-III cars from the Gordon and 

Orringer paper. The diesels are represented by an Amtrak Genesis Dual Mode locomotive, the 

potential PRIIA locomotive as defined in section 3 of this paper, and a heavy haul GE ES44AC 

freight locomotive. 

The Genesis Dual Mode unit is used in this example because it represents the heaviest Amtrak 

locomotive which operates at 110 mph. As such, it likely demonstrates the highest current 

braking demand for a diesel locomotive in the United States today.

The freight locomotive is included in this comparison because it represents the majority of the 

U.S. rail diesel traffic, or at least it represents the right tail of the distribution from a weight 

perspective. The speed of 80 mph is actually faster than the actual operating speed, as most 

railroads limit their traffic to 70 mph, and few attempt 75 mph. Because it is the speed limit for 

freight traffic on FRA class 5 track, 80 mph is used in this case. 

Returning to the comparison in Table 2.4.3-c, the diesels are heavier and have larger wheels so 

they require more energy dissipation but have more surface area to distribute the heat. The 

passenger locomotives have a higher operating speed and higher weight, while the freight 

locomotive has more axles and larger wheels, but also has a much higher retarding force per 

wheel than the passenger units. This higher retarding force is made possible by the lower 

operating speed which keeps the heat flux in a workable range. Though freight locomotives do 

not utilize blended braking, they do have an independent DB.

The brake retarding force for the diesels was calculated assuming a coefficient friction of 0.4. 

Although slightly high, it does not impact the overall conclusion.

As discussed in the Gordon and Orringer paper, the maximum wheel temperature is reached 

early in the braking event, before the locomotive slows down to 60 mph and below. It is 

therefore important to understand the highest values of heat flux which will occur at the highest 

speeds. Table 2.4.3-c compares the total energy required as well as the heat flux generated 

during initial brake application. Figure 2.4.3-e shows heat flux as a function of time during 
87

emergency stop from 125 mph; this figure is similar to ones used in the Gordon and Orringer 

study.

Figure 2.4.3.2-d: Wheel power and heat flux for a 285,000-pound PRIIA locomotive in aironly emergency brake application based on approach used in the Gordon and Orringer 

study

The resulting heat flux on the Genesis passenger and freight locomotives is higher than on the 

Arrow-III cars. The Genesis unit generates 119 percent of the heat flux produced by the NJT 

cars, while the PRIIA locomotive is likely to produce 148 percent. It is reasonable to assume 

that for “air only” emergency stop, the temperature would likely be even higher than the 1,100 oF 

observed on the NJT cars. Figure 2.4.3-f shows the heat flux and wheel temperature from the 

Arrow-III cars, as well as the expected temperature and the presumed heat flux of approximately 

3 MW/m2.
88

Figure 2.4.3.2-e: Assessment of expected wheel temperature for a PRIIA locomotive

utilizing tread brakes only in emergency stop from 125 mph, based on the wheel 

temperatures from the Gordon and Orringer study

Though this does not appear to be an issue on the existing Genesis fleets (likely due to the 

blended braking), the PRIIA locomotive would likely push the temperatures towards the 

transformation temperatures and increase chances for the formation of untempered martensitic 

structure which is brittle and can break out of the wheel tread surface. In most cases, the blended 

brake operation will drive the temperature to less than half that value.

The PRIIA locomotive, stopping from 125 mph, will generate heat flux higher than 3 MW/m2

with 40-inch wheels—and as a result, higher wheel tread temperatures. It is true that increasing 

wheel diameter will reduce the heat flux, in this case 60-inch wheels would match the heat flux 

in the Arrow-III cars, but that would likely still produce temperatures above 1,000 oF.

2.4.4 Summary

Braking energy is a function of the locomotive mass and square of the velocity. Increasing speed 

from 100 mph to 125 mph increases brake demand by 56 percent (1252

/1002=1.56). 

Additionally, the new Tier 4 emissions standards combined with crashworthiness standards are 

expected to result in an increase of mass in the new passenger locomotives compared existing 

locomotives. With this additional brake demand, air only braking can no longer rely solely on 

tread brake as this type of configuration would drive wheel temperatures beyond 1,000 oF, which 

is a major contributor to wheel tread cracks. As concerns over RCF continue to push 

specifications for harder wheels, it becomes more important to reduce the thermal load on wheels 

through the auxiliary brake systems. 
89

When comparing passenger and freight thermal dissipation needs, it becomes obvious that the 

operating speed is the main driver for the difference in these two operations. Since kinetic 

energy is a function of velocity squared, the energy dissipation needs at 125 mph become 

significantly higher: three times more kinetic energy at 125 mph on a PRIIA locomotive as 

opposed to 70 mph on a typical freight locomotive. Approximately the same ratio is reflected in 

Figures 2.4.3-c and 2.4.3-d.

As much as speed is the main driver in the comparison between freight and passenger operations 

in North America, comparisons with Europe make weight the clear differentiator. Lower weight 

in Europe means lower braking demand. Use of distributed power typically in the form of EMU 

and DMU allows the use of smaller traction motors or, in some cases, motors on just some axles, 

which frees up more space for disc brakes. On axles with no traction motors, three to four rotors 

can be placed on a single axle, thus eliminating the need for tread brakes. 

In normal braking situations, the DB will carry a sufficient portion of the brake load to 

adequately offload the heat input into the wheels from the tread brake. For safety reasons, DB is 

still considered a secondary system, meaning the vehicle must meet stopping distance 

requirements while utilizing only air brake. 

Based on weight, performance, and packaging constraints on the PRIIA locomotive, the disc 

brakes alone system is impractical. It would need to dissipate 50 percent more energy than 

currently used arrangements and as a result would need to occupy considerably more space. 

Removing tread brakes would also mean that the wheels lose the benefit of being cleaned by the 

thread brake shoe.

2.4.5 Conclusions and Recommendations

The new PRIIA Diesel Locomotive designed to operate at 125 mph will need to increase the 

number of braking systems from two to three. All three of the following systems will be 

required: 

1. Tread Brake – currently used on Amtrak Genesis

2. Dynamic Brake – currently used on Amtrak Genesis

3. Disc Brake – currently not used on Amtrak Genesis but used on NJT APL-45 Commuter

The new PRIIA passenger diesel loco will have both tread and disc brakes as a primary brake 

system but will rely on DB to reduce thermal input into wheels and disc brakes.

Because heat input into the wheels and brake discs is dependent on the duration and frequency of 

application, it may be helpful to define a braking duty cycle for the PRIIA locomotive. This area 

has potential for much debate between designers when one considers the impact that DB has on 

the air brake demand. The air brake has to be designed to stop the locomotive during emergency 

in case of DB failure without sustaining damage to the equipment. But how many times does it 

have to do this? Indefinitely? How often should emergency brake be applied consecutively? 

Considering the time it takes the train to get up to speed, it should be several minutes, but how 

likely is that scenario? It may be worthwhile to explore the tradeoff between cost of replacing 

damaged rotors or wheels in the unlikely event of multiple emergency stops with no DB against 

the cost of designing and maintaining a brake system which can sustain indefinite emergency 

applications with no DB. 
90

There is evidence that the railroads may prefer a simpler brake system that is capable of stopping 

the train safely in the few rare instances that, for whatever reason, DB may not be available. NJT 

appears to be one of these railroads, considering that it chose not to install disc brakes which 

would reduce wheel temperatures even in the air only emergency brake application. Experience 

may suggest that emergency braking should be defined as a one-time occurrence as related to 

energy absorption and thermal calculations. Typically, commuter train duty cycles are the most 

demanding, from heat generation standpoint, because of the relatively short distances between 

their stops, but an intercity train may not be subject to such demand. Therefore, utilization of the 

same brake equipment as on commuter trains may lead to unnecessary locomotive weight 

increases and complexity. Standardizing these design inputs may allow more optimal designs 

and ensure a consistent safety margin among designers.
91

2.5 Task 5—Wheel and Axle Summary Report

2.5.1 Definition

The axle is a critical component in a high-speed bogie system. Axle failures could result in 

derailment. Axle reliability is therefore essential to the safe operation of a high-speed bogie. In 

addition, the axle mass comprises a significant proportion of a rail vehicle’s “unsprung” mass. 

As previous sections have discussed, unsprung mass has a key influence on a rail vehicle’s 

dynamic axle loading and is of particular significance during high-speed operation. 

A significant amount of effort has been directed towards axle design for rail vehicles in highspeed service with the objective of reducing axle mass without compromising reliability. This 

survey reports on some of the major areas that impact high-speed truck safety and performance.

2.5.2 Approach

This study evaluates several key areas of current axle technology and the design of safe highspeed applications. Some key global design practices are compared with the U.S. approach, 

particularly in areas that impact axle weight and size. As in Task 1, major design approaches are 

summarized but with a more specific focus on axles. Inspection and material technologies, 

where known, were included. Certain key design and material approaches are discussed in more 

detail, again, in just those areas that have the biggest impact on the size and weight of axle 

design.

2.5.3 Results

Axle Design Trends in Existing High-Speed Applications

There are a number of locomotive applications in Europe, Japan, and China, which safely 

operate at speeds of 125 mph and above. A data set was developed to study the characteristics of 

those applications, and in particular, the application parameters that are relevant to the axle 

design. These parameters include vehicle speed, static axle load, unsprung mass per axle, the 

configuration of the wheel set in the bogie, what axle technology is employed, and what type of 

inspection technologies are employed, if any. 

The data set in Table 2.5.3-a demonstrates that most high-speed applications have a static axle 

load of 22 MT or less, and the fastest applications limit axle loads to 13 MT. One notable 

exception is the Bombardier ALP 45 DP which can operate at 200 kph (125 mph) in electric 

mode, with an axle load of 32.66 MT. While actual figures for unsprung mass and specific 

details of the axle design are typically unavailable, it is certain that most manufacturers are 

developing configurations designed to minimize unsprung mass through the use of fully 

suspended drive systems or truck frame suspended motors with axle-hung gearboxes.
92

Table 2.5.3-a: Matrix of High-Speed Locomotives Applications and Relevant Application Data

Company 

name

Type of 

application

Axle/Bogie 

design

Material 

Technology

Inspection 

technology

Unsprung mass

Bombardier

ALP 45 DP – Diesel 

Electric

Axle Load : 32.66t

Max Speed: 210kph 

(electric)

160kph (diesel)

Bogie designs same as 

ALP-46 bogie with 

changes in axle size & 

other components.

Fully suspended drive 

(ALP 46-A).

4-axle loco. Hollow 

axles designed with 

EN13103 (trailer) and

EN13104 

motorized).Quill 

design.

Axle-hung & nose 

suspended drive

Axle design: UIC A4T

A4T steel 24 

CrMo 5

Alternating Current 

Field Measurement 

(ACFM); Advanced 

eddy current 

inspection; UT Phased 

array with 3-D 

imaging. Tecnatom

has supplied automated 

axle inspection system 

to bombardier. Also, 

developed UT data 

analysis software.

Standard: M-101/90-

A (USA)

Possible ways for 

reducing unsprung mass 

from this company 

strategy.

Laterally suspended axles.

Suspending TM on truck.

Traction rod minimizing 

load transfer b/w axles.

Soft vertical and lateral 

suspension provides low 

lateral and vertical forces.

Alstom Prima II

Axle load: 22.5t

Max Speed: 200 kph 

(Passenger);

140kph (Freight);

120kph (HD freight)

Bo-Bo (86-90t);

Co-Co (129-135t)

Axle-hung nose 

suspended traction 

motor.

Collaboration with 

Euraxles for innovative 

axle designs and 

protection.

Estimating POD from 

response versus size 

method. Size measured 

by Alternating Current 

Potential Drop 

(ACPD), Phased array 

UT, TOF diffraction.

Siemens Axle load: 21.5–22t

Max Speed: 230kph 

(electric);

201kph (diesel)

Change in axle 

dimensions for next 

ICE3.ICE 3/Valero is 

nose-suspended drive. 

Max axle load for 

ICE3 is 17t.

Max speed: 330kmph

Mechanized UT 

inspection. Earlier 

inspection was done 

for every 3lakh kms 

(6months) to 

60,000kms (6weeks) to 

now 30,000kms 

(3weeks). Periodic 

inspection helps to 

detect small cracks.

Standard: EN13261-

2004

Hyundai Rotem HEMU-400X

Static axle load: 13t

Max speed: 400km/h

Four axles powered by 

synchronous induction 

motors.

Kinki Sharyo Shinkasen

Max axle load: 11.39t

Max speed: 285kmph
93

Axle design standards and methodologies

Various axle design methodologies are employed throughout the world. In Europe, powered 

axles are designed in conformance with EN 13104 standard, and in Japan railway axles are 

designed according to Japanese Industrial Standard E 4502. North America does not have a 

design methodology standard to govern axle design practice. North American axle designers are 

currently tasked with the challenge of following prudent axle design practice in the absence of 

the guidance of a design standard.

Axle design standards consist primarily of three parts:

1) Input definitions, including axle loading assumptions 

2) Calculation method of axle stress 

3) Design limits

The design standards can have a significant effect on axle size and weight. There are significant 

differences between the European and Japanese standards discussed here and the known U.S.

design practices. 

Axle loading definitions

Actual axle loading is a function of vehicle speed, track conditions, and vehicle architecture. 

Figure 2.5.3-a shows a comparison of the basic loading assumptions employed in European and 

Japanese axle design standards. Note that the European standard does not consider axle loading 

to be a function of vehicle speed. The Japanese standard utilizes multiple relationships for 

horizontal acceleration in order to provide different limits for different classes of track.

Figure 2.5.3-a: Comparison of loading assumptions used around the world for axle design

In U.S. design practices, the vertical axle loading component is considered to be solely a function 

of vertical car body acceleration. In international standards, the vertical loading component is 

0

0.5

1

1.5

2

2.5

0 50 100 150 200 250 300 350 400

Sprung Mass Aceleration (gs)

Locomotive Speed (km/hr.)

Japan Vertical

Japan Horizontal

EN Horizontal

EN Vertical
94

the summation of vertical car body acceleration and the vertical loading reaction loads from carbody roll resulting from horizontal acceleration. In the case of vertical acceleration, North 

American assumed stress is similar to the international design standards at speeds of less than 70

mph. At greater speeds, a very significant departure in design practice exists in North America, 

resulting in a much more conservative approach to axle design at speeds over 70 mph. This is in 

part because the slower speeds in North America have allowed such a conservative approach, but 

is primarily due to a lack of sufficient data to permit the proper characterization of axle loading 

based on the wide variety of North American track conditions and higher speeds. As North 

American applications reach higher speeds, dynamic loading from unsprung mass becomes 

increasingly significant. Maintaining such conservative design practices may result in 

unnecessarily heavy axle designs, with higher P2 forces as the penalty.

Axle Material

One of the critical aspects of axle design is the material specification. Table 2.5.3-b provides a 

comparison of various standard axle material grades. Asterisk marks indicate the most 

commonly employed grades. Note that the various standards from around the world have 

converged on two general approaches to heat treatment, the first being a normalizing process and 

the second being quench and temper.

Table 2.5.3-b: Comparison of Material Standards for Axles

Standard 

Axle Grade

Processing Selected Mechanical Properties Selected Composition Data

Yield 

Strength 

(MPa)

Ultimate 

Strength 

(MPa)

Min. % 

Elongation 

C Mn Si

AAR Grade 

F*

Double Normalized and 

Tempered

345 607 min 22 .45–.59 .60–.90 .15

AAR Grade G Quenched and Tempered 345 586 min 19 - .60–.90 .15

AAR Grade H Quenched and Tempered 414 724 min 18 - .60–.90 .15

EA1N* Double Normalized 320 550–650 22 .40 1.20 .50

EA1T Quenched and Tempered 350 550–700 24 .40 1.20 .50

EA4T* Quenched and Tempered 420 650–800 18 .22–.29 .50–.80 .15–.40

GOST 31334 

A1

Normalized or Normalized and 

Tempered

300 520–650 22 .40 1.20 .50

Quenched and Tempered 350 550–700 24

GOST 31334 

A2

Normalized or Normalized and 

Tempered

360 600–750 17 .50 1.20 .50

Quenched and Tempered 390 620–770 19

GOST 31334 

A3

Normalized or Normalized and 

Tempered

420 650–800 19 .40 1.60 .50

Quenched and Tempered

GOST 31334 

A3

Normalized or Normalized and 

Tempered

420 650–800 19 .40 .80 .50

Quenched and Tempered
95

Higher strength alloy steels such as 30NiCrMoV12 have been successfully employed in highspeed applications in Europe. This material has a minimum ultimate strength of 932 MPa which 

permits a further reduction in axle mass, given similar loads. 

Additional material processing has been employed in several applications to further improve the 

axle’s fatigue properties. Examples of additional material processing include induction 

hardening, sub-critical quenching and cold working axle fillets. The most prevalent and 

successful example of additional heat treatment processing can be found on the Japanese 

Shinkansen bullet train network where axles have been induction-hardened since the network 

began service in 1964. Figure 2.5.3-b provides an illustration of the improvements in fatigue 

properties gained through the employment of an induction-hardening process vs. a quenched and 

tempered material.

Figure 2.5.3-b: Rotating bending fatigue tests for press-fitted axles (Spec C, E represent 

induction hardening processes)1

European axle design standards permit axle designs to operate at a dynamic stress level 

approximately 20 percent higher by employing the quenched and tempered grade EA4T. 

Improving material capability in high-speed applications enables lighter axle designs and thus is 

often employed in higher speed applications where unsprung mass is critical.

Figure 2.5.3-c below provides a comparison of standard steel grades used around the world and 

the alternating stress allowances permitted by the specified design standard of the region. The 

permissible alternating stress, as a proportion of alternating stress, decreases when quench and 

tempered axles, as opposed to normalized axles, are employed. This results from an increase in 

notch sensitivity at higher strength, thus decreasing the realizable benefit from employing a 

higher strength steel alloy.

 

1 K Hirakawa1, K. Toyama and M Kubota, The analysis and prevention of failure in railway axles, 

International Journal Fatigue Vol. 20, No 2 pp.135–144, Elsevier Sercices ltd. 1998.
96

Figure 2.5.3-c: Comparison of allowable alternating stress as a function of ultimate 

strength

*In the case of the AAR material grade, no corresponding design standard exists. Examples of known 

values used in U.S. design were input.

**Source 2

Figure 2.5.3-d illustrates the allowable axle stress in two categories: (1) the regions of the axle 

where the axle body is not in contact with a mating component and the main failure mode is 

bending fatigue, and (2) the regions of the axle located under a fitted component where fretting 

fatigue is the primary failure mode. Note a convergence with a relatively similar value for 

allowable stress from various sources around the world. Furthermore, it is interesting to see the 

difference in allowable stress in regions of the axle where components are fitted versus regions 

without a mating component. K. Hirakawa1 and M. Kubota make the case that axle failure mode 

can be controlled simply with the axle geometry; a dataset in Figure 2.5.3-d presents an 

illustration of axle failure modes as a function of the ratio of the axle diameter under the 

wheelset versus the diameter in the axle body.

 

2 K Hirakawa1 and M Kubota, On the fatigue design method for high-speed railway axles, Proc Instn Mech Engrs 

Vol 215 Part F, IMechE 2001

0.00 0.20 0.40 0.60 0.80 1.00

AAR Grade F*

AAR Grade H*

EA1N

EA4T

SFA55**

Proportion of Ultimate Strength.

Alternating Stress

Permitted In Axle

Body

Alternating Stress

Permittesd Under

Fitted Componet
97

Figure 2.5.3-d: Failure Mode as a function of wheel seat diameter ratio

Axle Inspection Technologies 

Axle Inspection technologies are an important part of high-speed locomotive axle life cycle 

management. Table 2.5.3-a shows that many applications employ ultrasonic testing techniques 

while the axle is installed on the locomotive. Additionally, Japanese and European high-speed

trains require wheel sets to be disassembled and a surface inspection to be performed at specified 

intervals. 

Axle inspection procedures, the possible benefits, and impacts on cost are beyond the scope of 

this study. This is recommended as a possible area for further industry study. 

2.5.4 Conclusions and Recommendations

Minimizing unsprung mass through the use of fully suspended drive systems or truck frame 

suspended motors is the trend in high-speed truck design. This study previously concluded that 

U.S. manufacturers will also have to follow this trend to achieve 125 mph for the next generation 

passenger locomotives. Reducing unsprung mass in this way leaves the axle mass as one of the 

more significant remaining portions of a rail vehicle’s “unsprung” mass. Possible options to 

reduce axle weight should be considered and evaluated with sound engineering judgment.

Axle designers in North America may benefit by adopting a definition for axle loading similar to 

the one provided in the European and Japanese axle design standards in which axle loading is 

defined as a function of speed and track class. 

Overseas methods to predict axle loading are generally less conservative than U.S. methods. The 

U.S. manufacturers could benefit from better high-speed loading data. Moving forward, the 

industry should consider developing and providing better design data in conjunction with 

industry design standards and methods, with the goal of providing efficient, safe, and reliable 

axles that may also reduce weight.

As more efficient axle designs are required due to higher speed applications in North America,

axle designers should consider higher performance material and additional axle processing to 
98

improve fatigue performance. As new materials are specified, the acceptable failure criteria may

need to be carefully reviewed and changed as needed.

The North American passenger industry may need to consider adopting more advanced

inspection capabilities similar to those employed by its global counterparts.
99

2.6 Task 6—Locked Axle Safety Risk Assessment Summary Report

2.6.1 Definition

This assessment is to determine the safety risks associated with the three basic drivetrain designs 

that would most likely be applied to high-speed rail (limited to 125 mph) applications in North 

America. The three designs were chosen based on current freight and passenger rail axle-hung 

motor designs, as well as two variations of frame-hung motor designs found in North America 

and globally. The designation of these designs in this report will be “Axle-Hung” (figure 2.6.1-

a), “Frame-Hung HS (High-Speed) Coupling” (figure 2.6.1-b) and “Frame-Hung LS (LowSpeed) Coupling” (figure 2.6.1-c). The high- and low-speed designations refer to the location of 

the shaft line on which a flexible coupling resides. For example, the high-speed coupling design 

has the coupling on the motor shaft line (figure 2.6.1-b) and the low-speed coupling design has 

the coupling between the wheel and the quill shaft along the same shaft line as the main axle 

(figure 2.6.1-c). The assessment of these particular designs is motivated by the requirement for 

the reduction of unsprung mass—for a heavier diesel powered locomotive operating in the 

Amtrak environment—as the speed increases (see previous background and Task 1 discussions).

Figure 2.6.1-a: Axle-hung motor design
100

Figure 2.6.1-b: Frame-hung motor HS coupling design

Figure 2.6.1-c: Frame-hung motor LS coupling design

2.6.2 Approach

The approach is based on using Failure Modes and Effects Analysis (FMEA) methodology. GE 

Transportation assesses risk of a design using this method and specifically defines a safety risk 

using the severity score within this system. Although the FMEA also includes assessment of risk 

to reliability that includes scoring “occurrence” and “detection,” the reliability assessment was 

not considered a part of this study. Only the line items within the FMEA that scored high (9 or 

10) on severity are discussed. Table 2.6.2-a shows the definition of the severity scores. Severity 

scores of 9 or 10 are consider safety risks; scores of 8 or below are not.
101

Table 2.6.2-a: Severity Scoring Definition

Each design was evaluated as a drivetrain subsystem within the combo system. The drivetrain 

subsystem consists of components from the traction motor to the wheels. Five functions of the 

drivetrain subsystem were defined as follows:

1) Transmit tractive effort and braking torque from the motor and/or disk brakes to the 

wheels.

2) Operate for 20 year life and meet service reliability requirements.

3) Support the weight of the locomotive (wheel to axle to journal bearings into truck)

4) Transmit dynamic and steering loads (wheel to axle to journal bearings into truck)

5) Locate wheels and spacing

Within each of those functions, failure modes were defined and scored for severity using the 

scale listed in Table 2.6.2-a. Score levels for each failure mode were based on prior experience 

on North American freight and passenger locomotives.

Failure modes within the drivetrain system generally fall under failure or degradation of specific 

components within the drivetrain. Components that are not deemed as safety risks are grouped

together within the FMEAs, and the specific components that result in safety risk are singled out 

with separate line items in the FMEAs.

Effect Severity of Effect Ranking

Hazardous without 

warning

Potential failure mode affects safe system operation and/or 

involves noncompliance with government safety regulation 

without warning.

10

Hazardous with 

warning

Potential failure mode affects safe system operation and/or 

involves noncompliance with government safety regulation 

with warning.

9

Very High System/item inoperable with loss of primary function, but 

safe. (Such as a locomotive mission failure)

8

High System/item operable but at reduced performance level. 

Customer dissatisfied.

7

Moderate System/item operable, but comfort/convenience item 

inoperable. 

6

Low System/item operable, but comfort/convenience item 

compromised. 

5

Very Low Defect noticed by most customers. 4

Minor Defect noticed by average customer. 3

Very Minor Defect noticed by discriminating customer. 2

None No effect 1
102

2.6.3 Results

Of the five functions listed above, 3 and 4 are the only two that were identified as leading to a 

potential safety risk. The other three functions had failure modes that would only result in 

reliability degradation and so they will not be discussed further in this report.

The wheels, axle, and main axle bearings are the drivetrain components used to support the 

weight of the locomotive and transmit dynamic loads as well as steering loads. In our 

experience, the failure of main axle bearings can lead to locked axles but does not cause 

derailment and has not been considered a safety concern. The two components of concern are 

the main axle and wheels. Failure of either of these two components could cause the loss of 

function to support the locomotive and result in an inability to transmit dynamic and/or steering 

loads to the truck frame. This failure could lead to derailment; the main axle and wheels are 

therefore considered safety critical components.

In the axle-hung motor designs for both freight and passenger rail applications, these components

are flagged as safety critical components but have long standing controls such as internal design 

guides for axles and AAR certifications for wheels used in risk mitigation. Based on industry 

experience, the risk of failure of these components within their existing speed ranges is 

considered to be well understood.

As speeds increase and it becomes necessary to frame hang the motor, the next most cost 

effective design is to use a flexible coupling on the motor shaft line (Figure 2.6.1-b). 

Misalignment in the flexible coupling could result in additional risk to the drivetrain system 

through additional induced vibration and reliability of the coupling itself. Failure of the coupling 

would result in loss of tractive effort and possible locked axle, but the risk to wheel or axle shaft 

failure is deemed to be very low. There are additional reliability risks associated with a load 

bearing gear case and increased number of oil seals, but these components do not increase risks 

that would lead to wheel or axle failure.

As speeds increase even further, and the unsprung weight requirement is reduced, a frame-hung 

motor with a low-speed coupling design may be necessary (figure 2.6.1-c). Additional reliability 

risks were identified in the coupling and quill shaft, but additional safety risks are low. Lowspeed coupling failure will result in loss of torque but not a derailment. Coupling or quill shaft 

failure could lead to contact with axle, but the operator has sufficient warning to avoid further 

damage. Additional inspection of axle shaft and wheels may be required in the event of coupling 

or quill shaft failure.

2.6.4 Conclusions and Recommendations

Axle and/or wheel failure is the only traditional safety concern that could lead to derailment. 

Occurrence of axle failure has been historically low.

Locked axle has historically not led to derailment and, within the industry, has not been 

considered a safety concern.

For locomotive weight and speed combinations that are within the capability of the axle-hung 

and frame-hung high-speed coupling designs, specific additional safety requirements in the 

design are not currently recommended.

Additional reliability concerns may be associated with the high-speed designs, but with good 

design practice additional safety risk is low.
103

For frame-hung low-speed coupling designs, consideration for axle shaft and wheel inspection 

should be given in the event of coupling and/or quill shaft failure.
104

3. Report Conclusions

Summary:

An evaluation of bogie concepts for future high-speed passenger locomotives at speeds up to 125 

mph was conducted. It was first necessary to define the U.S. operating conditions and 

requirements that will define the bogie design specifications. There are a combination of 

requirements and conditions that make the U.S. high-speed bogie application a unique design 

challenge. Specifically, the combination of high axle load with the PRIIA specification that 

limits the P2 force to 82,000 lbf, and the necessity to operate on all classes of track and speed 

ranges, result in difficult design-tradeoffs. Currently, no U.S.-based manufacturer can provide a 

truck that will meet the next generation passenger requirements. 

The Task 1 comprehensive global survey identified a number of dominant design trends in 

current high-speed bogies manufactured and operated in Europe, Japan, and China. The primary 

design characteristics are focused on two general areas: (1) reduction of unsprung mass and total 

weight in order to minimize P2 forces; and (2) optimization of stability, curving performance, 

and ride quality. All of the non-U.S. high-speed locomotives surveyed operate at 22 MT per 

axle, or less—much less than the anticipated 30–32 MT, or more, for next generation 

locomotives in the United States. 

In Task 2, a truck and locomotive dynamics model was developed using ADAMS/Rail. From 

the global survey results, two likely truck design configurations were identified and described. 

The first of these, or Concept 1, was chosen as the configuration to be used in the Task 2 

modeling analysis. Concept 1 includes a two-axle truck with Bo-Bo arrangement, frame-hung 

motor, fabricated frame, and a combination of both tread and disc brakes. MCAT track 

conditions for both straight and curved tracks were used with new and worn conditions for both 

wheel and rail. The impact and effects on safety related performance parameters of the truck 

suspension parameters were investigated using design of experiments and response surface 

methods. 

To develop a feasible design space of possible suspension parameters, an optimization process, 

or design space method, was developed. This method successfully demonstrated that a feasible 

combination of suspension parameters could be determined for the specific and limited truck 

configuration and conditions analyzed. For a thorough analysis and actual design, more 

conditions and variables should be evaluated.

Task 3 showed that in order to stay within the P2 force limit of 82,000 lbf, the future PRIIA 

passenger locomotive trucks will have to incorporate a frame-hung motor to reduce unsprung 

mass. No current U.S.-produced diesel locomotives can meet the limit at 125 mph. The 

relatively light F40 (260,000 lbf) will meet it at 110 mph, but not at 125 mph. Both the AC and 

DC motor versions of the Genesis locomotives exceed 82,000 lbf at 110 mph.

Task 4 investigated braking requirements for a 125 mph passenger locomotive. Braking energy 

is a function of locomotive mass and speed. Increasing speed to 125 mph increases the amount 

of energy that must be absorbed by 56 percent. To stop a four-axle high horsepower diesel

locomotive from 125 mph will require an additional disc brake system combined with the already 

existing tread brake and DB system. For the safety required case of air-only brake operation, 

neither a tread nor a disc brake system alone will be sufficient. The addition of a disc brake 

system will add to the design challenge by adding weight and packaging difficulties. 
105

The Task 5 evaluation discusses possible impacts on axle design resulting from a possible need 

to lighten this safety critical component. As unsprung mass is reduced in high-speed U.S.

designs, the bogie axle will become one of the more significant remaining components of 

unsprung mass. There will be pressure to lighten this component while maintaining reliable and 

safe operation. 

Both European and Japanese axle designs are guided by standard practices defined, in the case of 

Europe, in conformance with EN 13104, and in Japan according to Japanese Industrial Standard 

E 4502. North America does not have a standard to govern axle design practice. The study in 

Task 5 compared U.S. and non-U.S. design practices from the perspective of the three basic 

areas of axle design standards: (1) input definitions and assumptions, (2) methods of calculating 

axle stress, and (3) definition of failure criteria. A review and comparison of axle materials was 

also discussed.

With respect to items 1 and 2, the study showed that the U.S. approach to axle design is more 

conservative. This is likely due to slower speeds which have allowed heavier axles, and, more 

importantly for high-speed applications, the lack of sufficient data to permit more accurate 

characterization of axle loading. Maintaining such conservative practices may result in 

unacceptably heavy axle designs as speeds increase. 

Industry and government cooperation to develop better data and joint design standards should be 

considered as a way to ensure safe, reliable, and optimized high-speed axle designs.

Another area reviewed in Task 5 was axle material. It was concluded that it may be necessary to 

consider higher performance materials and improved axle processing techniques to improve 

fatigue resistance. Further study with industry cooperation is recommended for this area. 

In Task 6, an assessment was made of the safety risks associated with the three basic drivetrain 

designs most likely to be applied to 125 mph applications in North America. An FMEA was 

conducted for each of the design cases to assess the potential for increased risk. It was 

concluded that locked axle has historically not led to derailment and therefore has not been 

considered a safety concern. Based on GE’s best current knowledge, it was considered unlikely 

that specific additional safety requirements in the design of the frame-hung traction motor 

concept would be required. In other words, while there may be additional reliability concerns 

associated with high-speed designs, they were not felt to be related to an increase in safety risk. 

However, since safety is the overriding concern in high-speed bogie design, during an actual 

high-speed bogie design and development program, this area would be investigated much more 

rigorously than the scope of this study allowed. 

Industry standards jointly developed by industry and government for certain of the critical 

drivetrain components and systems may also be important for ensuring safe optimum designs 

and possibly reducing costs through standardization.
106

4. References

1. ADAMS – User manual

2. Bombardier Bogies: http://www.bombardier.com/files/en/supporting_docs/BT-BogiesFLEXX_Power.pdf; http://www.bombardier.com/files/en/supporting_docs/BT-BogiesFLEXX_Portfolio.pdf

3. CIA World Fact Book https://www.cia.gov/library/publications/the-worldfactbook/geos/gm.html

4. GE DFSS III tools

5. Gordon J.E., Orringer O., 1995. “Investigation of the Effects of Braking System 

Configurations on Thermal Input to Commuter Car Wheels.” Volpe National Transportation 

Systems Center. Report no. DOT/FRA/ORD-95/DOT-VNTSC-FRA-95-7

6. ISO 2631 Evaluation of human exposure to whole-body vibration

7. Kalker, J.J.: Three dimensional Elastic Bodies in Rolling contact, Kluwer Academic 

Publisher, Dordrecht, 1990

8. MCAT – Department of Transportation, Federal Railroad Administration: 49 CFR Parts 213 

and 238 Vehicle-Track Interaction Safety Standards; High-Speed and High Cant Deficiency 

Operations; Proposed Rule

9. Montgomery, Douglas C. 2005. Design and Analysis of Experiments: Response surface 

method and designs. New Jersey: John Wiley and Sons, Inc.

10. Siemens Bogie catalog: 

http://www.mobility.siemens.com/mobility/global/SiteCollectionDocuments/en/railsolutions/components-and-systems/bogies-catalog-en.pdf 

11. Siemens Taurus Quill Drive: http://tams-www.informatik.unihamburg.de/applets/jfig/gallery/taurus-antrieb.pdf

12. Siemens Vectron bogie: http://www.mobility.siemens.com/mobility/global/en/interurbanmobility/rail-solutions/locomotives/vectron/technology/modular-locomotiveconcept/flexible-drive/Pages/flexible-drive.aspx#Vectron the drive system

13. Talamini, Jeong, Gordon (2007). Estimation of the Fatigue Life of Railroad Joint Bars, from

http://www.volpe.dot.gov/coi/pis/docs/integrity/jrcice2007-40065.pdf

14. Technology Vehicle Report issued in August 2011 by the Next Generation Equipment 

Committee.

15. Trains Magazine http://trn.trains.com/en/Railroad Reference/Railroad History/2009/09/The 

Stages of US Railroad Electrification.aspx
107

Appendix A. 

Truck Design Fundamentals

A.1 Technical background – high-speed bogie design requirements

This section reviews, in some detail, the basic design requirements required to produce a 

successful high-speed bogie. The following information provides the technical background to 

understand in some depth the analyses that follow in Tasks 1–6. 

Truck Design Requirements and Performance Characteristics

The truck is the interface between the locomotive and the track. As illustrated in Figure A.1.1a 

below, the interactions between wheel and track geometries provide the dynamic inputs to the 

truck and locomotive. A complete understanding of the correct and relevant inputs and system 

responses is a fundamental basic to designing a safe and reliable passenger locomotive truck.

Figure A.1.1a: Wheel-Rail Contact and Track Quality

Truck Functional Requirements

The truck subsystem provides the following basic functions:

• support the car body weight and distribute the locomotive weight to the axles

• guide and steer the locomotive on the tracks safely and with stability under all operating 

conditions 

• produce traction and braking effort between the wheels and rails

• transmit traction and braking force to the car body

• reduce and isolate vibrations and impact loads from wheel-track interactions

• mount and support auxiliary components and equipment
108

The basic functional requirements stated above must be met by a system design that will provide 

an acceptable level of the following six key truck performance measures in a cost effective, 

reliable, and safe package under the specified operation conditions: 

1) Running stability 

2) Derailment safety 

3) Wheel-rail dynamic forces

4) Braking capability

5) Ride quality and passenger comfort

6) Component and structural reliability and life

Each of these important truck characteristics is described in more detail in the following sections.

A.1.1.1 Running stability

The locomotive truck must be designed to be dynamically stable under all specified track and 

operating conditions. Theoretically, “stable” means that none of the Eigen values of the 

locomotive dynamics system would have a “positive” real part. In practice, stability is achieved 

by limiting the vibration of the locomotive system to a set of specified acceleration values that 

will ensure the system does not go into hunting mode.

Wheel and rail rolling contact is a unique mechanical element in the locomotive-track interaction 

system. The characteristics of its creepage-force (relative speed-force) relations are essential for 

the running stability of the locomotive truck. As shown below in Figure A.1.1-b, the forces 

developed between the wheel and the rail, in the contact patch, are determined by the following 

relationships: 

Figure A.1.1b: Wheel-rail rolling contact
109

�

Tξ

Tη

M

� = �

fξ(νξ, νη, ω)

fη(νξ, νη, ω)

fζ(νξ, νη, ω)

�

where: 

Tξ – Longitudinal tangential creep force,

Tη – Lateral tangential creep force,

M − Moment around the normal of the contact patch,

fξ(νξ, νη, ω) – function of longitudinal tangential creep force; determined by the contact 

geometry between the wheel and the rail as well as by the normal contact force,

fη(νξ, νη, ω) – function of lateral tangential creep force; determined by the contact 

geometry between the wheel and the rail as well as by the normal contact force,

fζ(νξ, νη, ω) – function of moment around the normal of the contact patch; determined by 

the contact geometry between the wheel and the rail as well as by the normal contact 

force,

νξ – Longitudinal creepage, νξ = Vξ_w−Vξ_r

V ,

νη – Lateral creepage, νη = Vη_w−Vη_r

V ,

ω – Spin creepage, = ωw−ωr

V ,

Vξ_w, Vη_w, ωw – Longitudinal, lateral, and spin speeds of the wheel in the coordinate 

system (ξ, η, ζ) at the contact patch (see Figure A.1.1b),

Vξ_r, Vη_r, ωr – Longitudinal, lateral, and spin speeds of the rail in the coordinate system 

(ξ, η, ζ) at the contact patch (see Figure A.1.1b),

V – Locomotive running speed

The total tangential force at the contact patch can be determined by:

T = �Tξ

2 + Tη2

The creep forces Tξ, Tη and M are nonlinear functions of the wheel and locomotive speeds, as 

illustrated in Figure A.1.1c. For a given locomotive speed, the creep forces increase when the 

creepages (relative speed between wheel and rail) increase. But once the creepages reach a large 

value, the creep forces will decrease with further increase of the creepages, and this introduces a 

negative damping in the wheel-rail contact. On the other hand, if the locomotive speed 

increases, the slope of the creep-force creepage curve will become smaller. This means that the 

damping in the wheel-rail contact will decrease as the locomotive speed increases.
110

Figure A.1.1c: Normalized total tangential force of T

μN = �Tξ

2 + Tη2/μN as a function of the total 

tangential creepage

The creep forces in the contact patch are saturated when the entire contact patch becomes a 

“sliding” area due to large creepages. The relation between the three creep forces Tξ, Tη and M is 

described by a “saturation surface,” as shown in Figure A.1.1d. Due to extreme creepage in any 

of the three directions—longitudinal, lateral or spin—the creep force can be saturated in that 

direction. For example, if lateral or spin creepage is too large in the contact patch, there will not 

be much tractive effort produced in longitudinal direction. 

Figure A.1.1d: Saturation surface of creep forces
111

The fact that the wheel-rail contact forces are locomotive speed dependent and the damping in 

the contact patch decreases with the increase of locomotive speed can lead to an unstable 

phenomenon known as hunting. 

The speed at which the wheelset becomes unstable is the “wheelset hunting speed,” also called 

“critical speed.” Hunting begins when the critical speed for a specific wheel-rail profile 

combination is reached. At speeds below the critical speed (Vc), the lateral and yaw motions 

(Figure A.1.1-e) of the wheelset are damped down when a track induced disturbance occurs 

(Figure A.1.1- f (a)) due to sufficient damping in the system. At speeds above the critical speed, 

any disturbance on the track can cause the wheelset motions to grow (Figure A.1.1-f(c)) due to 

lack of damping in the system. The unstable motions will grow until wheel flange contact occurs 

to create a forced limit cycle motion. 

Figure A.1.1-e: Hunting motion of a wheelset: y- lateral displacement, φ- yaw angle
112

Figure A.1.1-f: Lateral displacement of wheelset on ideal straight track

In addition to the wheelset hunting mode described above, there are two other hunting modes for 

a locomotive with primary and secondary suspension systems. The three basic hunting modes 

are listed below:

• Wheelset hunting – combination of lateral and yaw motions of the wheelset

• Truck hunting – combination of lateral and yaw motions of the truck frame

• Car body hunting – combination of lateral and yaw motions of the car body

Any of these instable modes can cause severe vibration and conditions that can lead to 

derailment. Selection of the proper suspension mechanisms and parameters is vital to ensure that 

the critical speeds for all the instable hunting modes will be well above the maximum locomotive 

operating speeds for all new and worn locomotive and track conditions. 

Design strategies which generally improve locomotive running stability and increase critical 

speed include: (1) increasing axle guiding stiffness in lateral and yaw direction, (2) higher lateral 

and yaw stiffness in secondary suspension, and (3) longer wheelbase and proper selection of the 

distance between the two trucks.

A.1.1.2 Derailment Safety

It is an obvious fundamental requirement that the locomotive must run safely on both straight 

and curved tracks at all operating speeds and all track and equipment conditions. The optimum 

configuration for the most stable operation on straight track may be at odds with the best 

configuration for operation on curved track. As with many areas of truck design, it is necessary 

to understand and manage the tradeoffs to create a design that will meet the safety requirements 

for all conditions. 

Derailment potential is measured by wheel and axle derailment coefficients as well as wheel load 

reduction:

• Wheel derailment coefficient L

V

, where L is the lateral force acting on the wheel, and V is 

the vertical force, as shown in Figure A.1.1g below
113

• Axle sum L

V |axle, where:

L

V |axle = �

L

V |left wheel� + �

L

V |right wheel�

• Wheel load reduction dV

V0

, where V0 is the static wheel load and dV = V − V0

Figure A.1.1g: Lateral and vertical forces acting on the wheel

Wheel and axle derailment coefficients and wheel load reduction must be within specified limits 

to ensure derailment does not occur.

Curve Negotiation

Derailment safety is of special significance on curved track. Curved track is typically superelevated in order to provide better force balance while negotiating the curve (see Figure A.1.1-h). 

During a curve, the wheel on the high rail (outside rail of the curve) has to travel more distance 

than the wheel on the inner rail in order for the wheels to purely roll (no slip) on the rails. This 

travel distance can be achieved by providing tapered wheel surfaces (conicity), which provide a 

rolling radius difference between the outer and inner wheels, as shown in Figure A.1.1-i below. 

For negotiating curves with large radius, the difference of travelling distances between the outer 

and inner rails is small. The conicity of the wheel profiles can provide sufficient rolling radius 

difference between the outer and inner wheels to meet the necessary requirement for pure rolling. 

But for curves with small radius, the rolling radius difference between the wheels will be limited 

by the wheel profile design. The radius difference will not be large enough to allow pure rolling. 

This leads to a difference between the angle of the wheelset axle and the radial direction of the 

curve called the attack angle (see Figure A.1.1-j below). The existence of the attack angle can 

cause wheel slip and result in a “forced negotiation.” A large attack angle increases the L/V ratio 

and therefore the risk of derailment.
114

Figure A.1.1-h: Configuration of curved track—superelevation provides centric force to 

help curve negotiation

Figure A.1.1-i: Attack angle during curve for a 2 axle truck
115

Figure A.1.1-j: Pure rolling and self-steering

When the wheelset negotiates a curve, the longitudinal creepage at the outer wheel (on the high 

rail) is larger than that at the inner wheel, due to its larger rolling radius. Therefore, the 

longitudinal creep force on the outer wheel will be larger than that on the inner wheel, as shown 

in Figure A.1.1-j above. The resultant moment due to this force difference will help the wheelset 

try to position itself in a more radial orientation. The capability of the wheelset to purely roll (as 

stated in the above section) and position itself in radial orientation due to its rolling radius 

difference is known as “self-steering.” Depending on the curve radius, especially for sharp 

curves, this self-steering capability of the wheelset with tapered wheel profiles may not be good 

enough for the wheelset to steer through the curve completely in the radial direction.

Radial Steering

For best curving performance, each of the truck wheelsets should be aligned in the radial 

direction, as shown above in Figure A.1.1-k below. The truck design, especially the design of the 

axle guiding devices, should provide enough compliance for the wheelset to steer when it enters 

a curve. The secondary suspension should be designed in such a way that it will allow the truck 

to rotate smoothly when the locomotive enters a curve and also provide enough restoring 

stiffness for the truck to rotate back to its normal position when it leaves the curve. Shorter 

wheelbase and softer primary and secondary yaw stiffness will help to improve curving stability.

Wear on both wheel and track is generally higher on curved track. In addition to increasing 

maintenance costs, this can also lead to poor performance and stability. Reducing attack angle 

and avoiding flange and multipoint contact by properly selecting the suspension parameters will 

help to slow wheel-rail wear.
116

Figure A.1.1-k: Radial steering of 2 axle truck

As shown, there is a conflict between running stability and curving performance in suspension 

system design. A tradeoff analysis must be properly conducted so that both performance modes 

can meet the safety requirements.

A.1.1.3 P2 and other wheel-rail dynamic forces

Extreme dynamic forces exerted by the wheels on the rails may damage both the track and 

locomotive components, as well as increase wear. The following performance index/variables 

are used to evaluate the wheel-rail dynamic forces: 

• Lateral wheel force, L

• Track shift force (lateral axle force), ∑ Laxle

• Truck side, L/V

L

V |truck side = ∑ Ltruck side

∑ Vtruck side

• P2 force – a vertical force of wheel on rail, determined as 

P2 = P0 + 2αV � Mu

Mu + Mt

�

1/2

�1 − πCt

4[Kt(Mu + Mt)]1/2� (KtMu)1/2

Where (see Figure A.1.1-l below)

P0 – Static wheel load (N)

2α – Total angle of vertical ramp discontinuity (rad)

V– Locomotive speed (km/h)

Mu – Unsprung mass (kg/wheel)
117

Mt– Effective track/rail mass per wheel (kg)

Kt – Effective track/rail stiffness per wheel (N/m)

Ct– Effective track/rail damping per wheel (Ns/m)

Figure A.1.1l: P2 force

The lateral wheel-rail contact forces, which include lateral wheel force, track shift force, and 

truck side L/V, heavily depend on the locomotive stability and curving performance. As 

indicated previously, proper design of the axle guiding mechanism and selection of primary and 

secondary suspension parameters, as well as the wheel base and loco length, will help improve 

the lateral wheel-rail forces.

P2 force is a measure of the total vertical force of a wheel acting on the rail when the wheel runs 

through a ramp, or dip angle, on the rail. It includes the static wheel load and the dynamic force 

due to the unsprung mass. The amplitude of the P2 force is determined by both the track 

characteristics (track mass, stiffness, damping, and discontinuity angle) and the locomotive 

characteristics (unsprung mass and wheel load), as well as the operating condition (locomotive 

speed). From a truck design perspective, the P2 force will be primarily determined by the 

locomotive wheel load (locomotive weight) and the unsprung mass, for a given track 

infrastructure and locomotive speed. The smaller the wheel load and unsprung mass are, the 

lower the P2 force would be. 

The unsprung mass, also called dead weight, includes the weight of the wheelset and the weight 

of components mounted on the axle. Any design which reduces the weight of the wheel and 

axle, and any reductions to the weight of components mounted on the axle and wheel, will 

reduce the unsprung mass and therefore reduce the P2 force. Frame or body hung motors, and 

semi- or fully suspended drivetrains, are the most common methods to significantly reduce the 

unsprung mass. 

A.1.1.4 Braking capability

Brake capability is essential for high-speed locomotive. The brake system design should meet the 

specified stopping distance or deceleration requirement for the specified locomotive speed range 

and braking frequency. Depending on the customer specification, the brake performance can be 

achieved by air brake only, blended brake (DB + air brake), or a combination of air brake and 

other brake types, such as eddy current brake or aero DB. However, “air brake only” is the 

common minimum requirement for emergency brake, due to safety regulations. 

High-speed passenger locomotives have a more demanding braking application than slower 

freight locomotives, in terms of energy that must be dissipated by the brake system. The 
118

traditional tread brakes alone, which are normally used for low-speed application and are limited 

by the capacity of the locomotive wheels to absorb heat without exceeding transformation 

temperature and damaging the wheel and brakes, will no longer be capable of meeting the highspeed locomotive application standards. Therefore, disc brakes are commonly used for higher 

speed application because of their higher energy absorption capability and the lack of wheel 

damage. But, disc brakes also have a limit on the amount of energy that can be absorbed. To 

meet the braking requirement, proper numbers of discs must be determined.

Disc brakes can be wheel mounted discs, axle mounted discs, or separate axle mounted discs. As 

indicated, adding discs either on the wheel or on the axle will increase the unsprung mass, and 

therefore the P2 forces. This adds significantly to the challenge of designing a high-speed diesel 

locomotive which is already very heavy because of its engine and its support equipment. In 

addition, adding a disc brake system to a location on the truck other than the wheel and axle will 

take up space in the already crowded bogie system. 

A.1.1.5 Ride quality or comfort

In addition to controlling and reducing the vibrations on the locomotive introduced by track 

irregularities which may cause damage to locomotive components, a suitable working 

environment for the crew in the operator cab should be created by proper design of the 

locomotive suspension system. This working environment will be defined and evaluated as ride 

quality or comfort. 

There are different standards for ride quality or comfort. Ride index according to Sperling (see 

References) is a popularly used evaluation method because of its simplicity. ISO 2631 is an 

internationally recognized standard for evaluation of human exposure to whole-body vibration. 

Depending on the standards, either accelerations or velocities measured in the operator cab or on 

the operator seat will be used for the ride quality or comfort evaluation. Therefore, reducing the 

vibrations (lateral, vertical, and longitudinal accelerations or velocities) in the operator cab and 

operator seat by proper design of the primary and secondary suspension systems will help 

improve the ride quality and operator comfort. Once again, design tradeoffs will have to be 

thoroughly understood and optimized, as characteristics that provide the most comfortable ride 

may not be the most desirable for stability.

A.1.1.6 Life and reliability 

As indicated in the previous sections, the truck, while operating in an environment with extreme 

dynamic loading, will support the locomotive car-body weight, guide the locomotive on the 

straight track and negotiate curved track, generate and transmit tractive and braking efforts, as 

well as reduce and isolate vibrations. All of the truck components have to withstand not only the 

static loads, but also the dynamic alternating loads. The design of the components must meet 

both proof and fatigue strength requirements with sufficient safety margin for the specified 

design life. Clear understanding and prediction of the loads and duty cycles under specified U.S.

track and operation conditions, proper application of suitable design methods, as well as proper 

selection and application of materials and manufacturing technologies will help increase the 

component/structure reliability for safe operation of the high-speed truck. 

Bogie Design Tradeoffs

The six key performance traits described above must all be designed to meet the specifications 

required to ensure safe reliable operation of a high-speed passenger truck. Though already 
119

noted, it is worth repeating that the best design to meet each performance trait may conflict with 

the optimal designs for other traits. Examples include designing for stability in straight versus

curved track, P2 force versus unsprung mass increases due to additional brake systems, and 

meeting component life and reliability requirements versus the need to reduce weight, and many 

others. Providing a final optimized design that meets all the performance requirements will 

require detailed and complex tradeoff analyses. This is particularly true in the case of the highspeed truck design that will be required to meet the new PRIIA specifications. The need to 

reduce P2 forces significantly below the current levels, while dealing with weight increases 

resulting from crashworthiness and T4 emissions requirements, will add difficulties to a design 

that was already challenging.
120

Appendix B. 

Survey results tables—Truck design parameters

Truck type (axle 

arrangement) SF1 (Bo-Bo) for E-loco

SF2 (Bo-Bo) for ElocoSF3 (Bo-Bo) for Eloco SF4 (F4)

Builder Siemens Siemens Siemens Siemens

Speed range 230km/h (357 kph) 140 km/h 160 km/h 200 km/h

Axle load 21.5t 22t 22t 80~90/4=20~22.5t

Motor Suspension 

design Frame-hung Axle-hung Semi frame-hung Frame-hung

Drivetrain design

Quill Gear box Gear box

Hollow shaft drive 

(option: fully suspended 

quill drive)

Brake design and

arrangement 2 discs on separate shaft Wheel disc brake Wheel brake discs Wheel discs

Parking brake 

design and

arrangement Spring loaded brake Spring loaded system Spring loaded brake Spring loaded brake

Axle design Forged Hollow axle Hollow-bored axle shaft

Wheel diameter 

(new and worn) 1150/1070 mm 1250/1170 mm 1100/1020 mm 1250/1170 mm

Wheelbase 3000 mm 2900 mm 2700 mm 2900 mm (3000 mm)

Max TE/Axle 300KN/4 300KN/4 62.5KN 300KN/4=75KN

Power per axle 1.6 MW/4 1.6 MW/4 750 KW

Bogie frame design Fabricated Fabricated Fabricated Fabricated

Primary suspension

(vertical) 2 coil springs per JB, 1 

vertical damper per JB

2 coil springs per JB, 

1 vertical damper per 

JB

2 coil springs per JB, 

1 vertical damper per 

JB coil springs

Axle guiding design 3 points, bushings 

(vertical arrangement)

3 points, bushings 

(vertical arrangement)

3 points, bushings 

(vertical)

Bushings (vertical 

arrangement)

Secondary 

suspension 2 coil springs per side 

(lateral arrangement), 

vertical and yaw dampers

2 coil springs per side 

(lateral arrangement), 

vertical, yaw dampers

2 coil springs per side 

(lateral arrangement), 

lateral & vertical and 

yaw dampers

Flexcoil spring with low 

rotational stiffness

Traction link design 

between truck and 

car body Pivot Pivot / sliding plates Pivot / sliding plates Pivot

Gauge 1435 mm 1435 mm 1435 mm 1435 mm

Minimum curve 120 m 80 m 120/100 m 90 m (E)/100 m (D)

Bogie weight 18t 17t 14t

Application Taurus at Austria Federal 

Railway (OBB); OBB / 

Austria DB AG / Germany

MAV / Hungary

GYSEV / Hungary

Dispolok / Europe

RTS / Austria

SZ / Slovenia

SNCB / Belgium

DB AG / Germany

Dispolok / Europe

SBB / Switzerland

Mitsui / Europe

CP / Portugal

OBB / Austria

KCRC Hong Kong / 

China

Dispolok / Europe

Several European 

operators

Vectron (European 

application) 2010; 

Belgium State Railways 

(SNCB)

Northeast Corridor / 

Keystone corridor –

ACS-64
121

Truck type (axle 

arrangement) Flex 120 (Co-Co)

Flex Power 140/200 

(Bo-Bo)

Flex Power 250 (BoBo) Flex Power 350 (Bo-Bo)

Builder Bombardier Bombardier Bombardier Bombardier

Speed range 80–120 km/h 140/200km/h 250km/h 330km/h

Axle load 22–30t 22t 21t (APL 45-32.66t) 17t

Motor suspension 

design

Axle-hung

Flexx 140:Axle-hung, 

Flexx 200: Framehung (Fully 

suspended drive)

Car body, Bogie 

frame-hung (Fully 

suspended drive)

Hung motor (Axle-hung 

gearbox; motor fully 

suspended in bogie 

frame = partly

Drivetrain design

Brake design and

arrangement Double side tread brakes Wheel disc brakes

Hollow shaft mounted 

axle disc brakes

Mounted disc brakes 

and 2 axle mounted disc

Parking brake 

design and

arrangement

Axle design

Wheel diameter 

(new and worn) 1067mm 1250mm

1118 (44”)/1046 

(41.2”)(APL45-44”) 1040mm

Wheelbase 2 x 1850mm 2600mm 2650mm 2650mm

Max TE/Axle 259-350KN/3 160KN/2 150KN/2 100/2 KN

Power per axle

Bogie frame design Fabricated, flat Fabricated, flat Fabricated Fabricated, center low

Primary suspension

(vertical) 2 coil springs per JB

2 coil springs per JB, 

1 vertical damper per 

JB

2 coil springs per JB, 

1 vertical damper per 

JB

2 coil springs per JB, 1 

vertical damper per JB

Axle guiding design Bushing (lateral 

arrangement), end axle

2 point links, bushing 

in lateral

2 point links, bushing 

in lateral

2 point link, bushing in 

lateral

Secondary 

suspension 2 coil springs per side 

(longitudinal 

arrangement), 1 vertical 

damper per side, 2 lateral 

dampers

(longitudinal 

arrangement), 1 

vertical damper per 

side, (2 lateral 

dampers), 1 yaw 

damper per side

(longitudinal 

arrangement), 1 

vertical damper per 

side, (2 lateral 

dampers), 1 yaw 

damper per side

(longitudinal 

arrangement), 1 vertical 

damper per side, (2 

lateral dampers), 2 yaw 

damper per side

Traction link design 

between truck and 

car body Traction rod

Traction rod at the 

end transom (low 

traction point)

Traction rod at the 

end transom (center traction rod)

Gauge 1000 / 1067 / 1435mm 1435 1435 1435mm

Minimum curve

Bogie weight 23.8t 16t 16t 11.7t

Application

Blue Tiger locomotives; 

IORE locomotives

TRAXX locomotives, 

D B A G,

SBB, BLS,

SNCB, Angel Trains, 

Mitsui and others

North American ALP 

46 locomotives *260); 

ALP 45 DP AVE**S-102
122

Truck type (axle 

arrangement) Prima II locomotive (Bo-Bo) DF11G (Co-Co) Maxima (Co-C0)

Builder Alstom China Voith

Speed range 200 km/h (Freight 140kph) 170km/h 120~160km/h

Axle load 22t (21.5~22.5t) 23t 126~135/6t (30 C C/40C C)

Motor suspension 

design

Frame-hung (gear box on 

axle) Frame-hung, Quill

Drivetrain design Quill Hydrodynamic

Brake design and

arrangement Wheel disc brakes

Electro-pneumatic and Hydrodynamic

Parking brake 

design and

arrangement

Axle design

Wheel diameter 

(new and worn) 1160mm (1150 freight) 1050

Wheelbase 3000mm (2600 freight) 2*2000

Max TE/Axle 320/4 KN 245/6 KN 519/6 KN

Power per axle

Bogie Frame Design Fabricated

Primary 

suspension(vertical)

2 coils/JB, 1 vertical 

damper/JB Coil spring

Axle guiding design Diagonally arranged 

traction links / JN 

(bushings) “Z” link

Secondary 

suspension

2 coils in side (longitudinal 

arrangement) Coil spring

Traction link design 

between truck and 

car body Traction rods at both ends 4-roads parallel

Gauge 1435mm 1435

Minimum curve 145m 80m

Bogie weight 18.6t (16.95t freight)

Application

Prima II

Germany, Poland, Benelux 

and others
123

Truck type (axle 

arrangement)

Vossloh Wuro 3000 AC 

(Euro Light) Bo-Bo

Vossloh Euro 4000 

AC (Co-Co) PL 42 AC (Bo-Bo) Bo-Bo

Builder

Vossloh Bossloh

Bossloh, EMD, 

Alstom Talgo

Speed range Up to 200km/h 160km/h 160km/h 200km/h

Axle load 88/4 21t 130/4t 17?

Motor suspension 

design

Axle-hung (Nose 

suspension) Nose suspension Frame-hung

Drivetrain design Universal shaft Cardan 

shaft)

Brake design and

arrangement Wheel discs Disk/wheel Wheel discs Wheel discs

Parking brake 

design and

arrangement

Axle design Hollow

Wheel diameter 

(new and worn) 1117mm 1067 1010

Wheelbase 2600mm 2800mm

Max TE/Axle 305/4 311KN/4

Power per axle

Bogie frame design Fabricated Fabricated Fabricated Fabricated

Primary suspension

(vertical)

2 coil springs + JB, 

vertical dampers

Axle guiding design Link

Secondary 

suspension

2 coils /side (longitudinal), 

vertical / lateral dampers, 

Anti-roll bar

Traction link design 

between truck and 

car body Center pin (pivot) Rubber-metal

Gauge 1435mm 1435

Minimum curve

Bogie weight

Application
124

Truck type (axle 

arrangement) Genesis: Bo-Bo MBTA Bo-Bo Ansaldo Bo-Bo

Builder GE/Vossloh GE/MPI Amsaldo Breda

Speed range 117km/h (110mph) 165.8km/h (103mph) 220kph

Axle load 30.38t (67000lbs) 33t (72750lbs) 22t

Motor suspension 

design Axle-hung Axle-hung Frame-hung

Drivetrain design Gear Gear Quill

Brake design and

arrangement Tread Tread Wheel discs

Parking brake 

design and

arrangement

Axle design

Wheel diameter 

(new and worn) 1016 (40”) 1016 (40”) 1100 (43.3”) and 40.94”

Wheelbase 2750 2845mm (112”) 2750

Max TE/Axle 315KN/4

Power per axle 1500KW (2000HP)

Bogie frame design Fabricated Cast Fabricated

Primary suspension

(vertical) 2 coil springs + Journal box Coil springs Coil springs (on top of JB)

Axle guiding design

3 points link, vertical 

bushings Pedestal leg

Integrated rotational arm (JB 

and link integrated into one 

part)

Secondary 

suspension

Rubber + coil springs (in 

series), 2 sets per side 

(longitudinal arrangement) Coil spring

Coil springs (2/side, 

longitudinal)

Traction link design 

between truck and 

car body Center pin Traction rod

Gauge 1435

Minimum curve 90m (yard), 149m (line), max 

cand deficiency 6”

Bogie weight 15.37t (3900 lbs)

Application E403 for Trenitalia
125

Appendix C. 

Basic configurations of surveyed trucks

Figure C.1a: Siemens SF1 High-Speed Truck with Frame-Hung Motor

Figure C.1b: Siemens SF2 Truck with Axle-Hung Motor
126

Figure C.1c: Bombardier Flexx 140 Truck with Frame-Hung Motor

Figure C.1d: Bombardier Flexx 250 Truck with Frame-Hung Motor
127

Figure C.1e: Alstom Truck with Frame-Hung Motor and Double Traction Rods

Figure C.1f: Talgo Truck with Body-Hung Engine and Cardan Shaft Drive
128

Figure C.1g: Vossloh Truck with Axle-Hung Motor
129

Appendix D. 

Task Modeling Results

Figure D.1a: Pareto plot

Figure D.1b: Sensitivity plot

Figure D.1b—Vertical car body acceleration in m/s^2 (Frame-hung motor, new wheel and worn 

rail profiles, straight track, Class 7, 125 mph)
130

Figure D.1c: Pareto plot

Figure D.1d: Sensitivity Plot

Figure D.1d—Lateral car body acceleration in m/s^2 (Frame-hung motor, new wheel and worn 

rail profiles, straight track, Class 7, 125 mph)
131

Figure D.1e: Pareto plot

Figure D.1f: Sensitivity Plot

Figure D.1f—Lateral truck acceleration in m/s^2 (Frame-hung motor, new wheel and worn rail 

profiles, straight track, Class 7, 125 mph)
132

Figure D.1g: Pareto plot

Figure D.1h: Sensitivity Plot

Figure D.1h—Track shift force in N (Frame-hung motor, new wheel and worn rail profiles, 

straight track, Class 7, 125 mph)
133

Figure D.1i: Pareto plot

Figure D.1j: Sensitivity Plot

Figure D.1j—Wheel L/V (Frame-hung motor, new wheel and worn rail profiles, straight track, 

Class 7, 125 mph)
134

Figure D.1k: Pareto plot

Figure D.1l: Sensitivity Plot

Figure D.1l—Vertical car body acceleration in m/s^2 (Frame-hung motor, new wheel and worn 

rail profiles, 1.1 degree curve, 6” superelevation, Class 7 track, 125 mph)
135

Figure D.1m: Pareto plot

Figure D.1n: Sensitivity plot

Figure D.1n—Lateral truck acceleration in m/s^2 (Frame-hung motor, new wheel and worn rail 

profiles, 1.1 degree curve, 6” superelevation, Class 7 track, 125 mph)
136

Figure D.1o: Pareto plot

Figure D.1p: Sensitivity plot

Figure D.1p—Track shift force in N (Frame-hung motor, new wheel and worn rail profiles, 1.1 

degree curve, 6” superelevation, Class 7 track, 125 mph)
137

Figure D.1q: Pareto plot

Figure D.1r: Sensitivity plot

Figure D.1r—Wheel L/V (Frame-hung motor, new wheel and worn rail profiles, 1.1 degree 

curve, 6” superelevation, Class 7 track, 125 mph)
138

Figure D.1s: Pareto plot

Figure D.1t: Sensitivity plot

Figure D.1t—Wheel load reduction rate (Frame-hung motor, new wheel and worn rail profiles, 

1.1 degree curve, 6” superelevation, Class 7 track, 125 mph)
139

Figure D.1s: Pareto plot

Figure D.1t: Sensitivity plot

Figure D.1t—Pareto and sensitivity of vertical car body acceleration in m/s^2 (Frame-hung 

motor, worn wheel and worn rail profiles, straight track, Class 7, 125 mph)
140

Figure D.1u: Pareto plot

Figure D.1v: Sensitivity plot

Figure D.1v—Lateral car body acceleration in m/s^2 (Frame-hung motor, worn wheel and worn 

rail profiles, straight track, Class 7, 125 mph)
141

Figure D.1w: Pareto plot

Figure D.1x: Sensitivity plot

Figure D.1x—Lateral truck acceleration in m/s^2 (Frame-hung motor, worn wheel and worn rail 

profiles, straight track, Class 7, 125 mph)
142

Figure D.1y: Pareto plot

Figure D.1z: Sensitivity plot

Figure D.1z—Track shift force in N (Frame-hung motor, worn wheel and worn rail profiles, 

straight track, Class 7, 125 mph)
143

Figure D.1aa: Pareto plot

Figure D.1ab: Sensitivity plot

Figure D.1ab—Wheel L/V (Frame-hung motor, worn wheel and worn rail profiles, straight track, 

Class 7, 125 mph)
144

Figure D.1ac: Pareto plot

Figure D.1ad: Sensitivity plot

Figure D.1ad—Vertical car body acceleration in m/s^2 (Frame-hung motor, worn wheel and 

worn rail profiles, 1.1 degree curve, 6” superelevation, Class 7 track, 125 mph)
145

Figure D.1ae: Pareto plot

Figure D.1af: Sensitivity plot

Figure D.1af—Lateral car body acceleration in m/s^2 (Frame-hung motor, worn wheel and worn 

rail profiles, 1.1 degree curve, 6” superelevation, Class 7 track, 125 mph)
146

Figure D.1ag: Pareto plot

Figure D.1ah: Sensitivity plot

Figure D.1h—Lateral truck acceleration in m/s^2 (Frame-hung motor, worn wheel and worn rail 

profiles, 1.1 degree curve, 6” superelevation, Class 7 track, 125 mph)
147

Figure D.1ai: Pareto plot

Figure D.1aj: Sensitivity plot

Figure D.1aj—Track shift force in N (Frame-hung motor, worn wheel and worn rail profiles, 1.1 

degree curve, 6” superelevation, Class 7 track, 125 mph)
148

Figure D.1ak: Pareto plot

Figure D.1al: Sensitivity plot

Figure D.1al—Wheel L/V ( Frame-hung motor, worn wheel and worn rail profiles, 1.1 degree 

curve, 6” superelevation, Class 7 track, 125 mph)
149

Figure D.1am: Pareto plot

Figure D.1an: Sensitivity plot

Figure D.1an—Wheel load reduction rate (Frame-hung motor, worn wheel and worn rail 

profiles, 1.1 degree curve, 6” superelevation, Class 7 track, 125 mph)
150

Abbreviations and Acronyms

AAR Association of American Railroads

Bo-Bo A locomotive with two bogies, each bogie having two powered axles

CCD Central Composite Design

CFR Code of Federal Regulations

DB Dynamic Brake

DFSS Design for Six Sigma

DMU

DOE

Diesel Multiple Unit

Design of Experiments

EMU Electric Multiple Unit

EN European

FMEA Failure Modes and Effects Analysis

FRA Federal Railroad Administration

GET GE Transportation

HSR High-Speed Rail

kph Kilometers per Hour

lbf Pounds Force

MBTA Massachusetts Bay Transit Authority

MCAT Minimally Compliant Analytical Track

mph Miles per Hour

MPI Motive Power Inc.

MT Metric Tons (1 metric ton = 2,204.6 lbf)

NJT New Jersey Transit

PRIIA Passenger Rail Investment and Improvement Act

RCF Rolling Contact Fatigue