APTD-1558
TRANSMISSION STUDY
FOR TURBINE AND RANKINE
CYCLE ENGINES
ENVIRONMENTAL PROTECTION AGENCY
Office of Air and Water Programs
Office of Mobile Source Air Pollution Control
Advanced Automotive Power Systems Development Division
Ann Arbor, Michigan 48105
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APTD-1558
TRANSMISSION STUDY
FOR TURBINE AND RANKINE
CYCLE ENGINES
Prepared By
M. A. Cordner and D. H. Grimm
Sunstrand Aviation
Division of Sunstrand Corporation
Rockford, Illinois 61101
Contract No. 68-04-0034
EPA Project Officer:
J. C. Wood
(NASA Lewis Research Center)
Prepared For
U.S. ENVIRONMENTAL PROTECTION AGENCY
Office of Air and Water Programs
Office of Mobile Source Air Pollution Control
Advanced Automotive Power Systems Development Division
Ann Arbor, Michigan 48105
December 1972
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The APTD (Air Pollution Technical Data) series of reports is issued by
the Office of Air Quality Planning and Standards, Office of Air and
Water Programs, Environmental Protection Agency, to report technical
data of interest to a limited number of readers. Copies of APTD reports
are available free of charge to Federal employees, current contractors
and grantees, and non-profit organizations - as supplies permit - from
the Air Pollution Technical Information Center, Environmental Protection
Agency, Research Triangle Park, North Carolina 27711 or may be obtained,
for a nominal cost, from the National Technical Information Service,
5285 Port Royal Road, Springfield, Virginia 22151.
This report was furnished to the U.S. Environmental Protection Agency
by Sunstrand Aviation in fulfillment of Contract No. 68-04-0034
and has been reviewed and approved for publication by the Environmental
Protection Agency. Approval does not signify that the contents necessarily
reflect the views and policies of the agency. The material presented in
this report may be based on an extrapolation of the "State-of-the-art."
Each assumption must be carefully analyzed by the reader to assure that it
is acceptable for his purpose. Results and conclusions should be viewed
correspondingly. Mention of trade names or commercial products does not
constitute endorsement or recommendation for use.
Publication No. APTD-1558
n
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TABLE OF CONTENTS
Section Section Page
No. Title No.
INTRODUCTION - ABSTRACT 1
A. Introduction 1
B. Abstract 2
II. RESULTS AND CONCLUSIONS 3
III. RECOMMENDATIONS 5
IV. FEASIBILITY STUDY 7
A. Introduction 7
1. Methodology 7
2. Requirements 7
B. Analysis of Transmission Types 8
1. Discussion of Types 8
2. Conclusions 10
C. Optimization of Selected Candidates 10
1. Hydromechanical 10
2. Traction 13
3. Comparison of Final Hydronechanical
and Traction Design Choices 16
V. TRANSMISSION DESCRIPTION 19
A. Hydromechanical 19
B. Traction Drive 30
C. Noise 40
D. Maintainability 41
VI. PERFORMANCE 43
A. Introduction 43
B. Ground Rules and Transmission
Parameter Summary 43
i i i
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Section Section
No. Title
C. Transmission Efficiency 46
D. Grade and Acceleration Performance 56
E. Fuel Consumption at Constant Speed
and Full and Part Loads 63
F. Fuel Consumption Summary 63
G. Tractive Effort Limits 63
^ai. CONTROL SYSTEM ANALYSIS 79
A. Control System Approach 79
B. Description of Operation 81
C. Stability Analysis 81
D. Safety Analysis 82
VIII. ESTIMATED TOTAL MANUFACTURING COST 85
A. Definition of the Cost Analysis 85
B. Cost Procedure 85
C. Results of Cost Analysis 86
IX. REFERENCES 89
The appendices in this report are identified by the number of the
section where first referenced. There are no appendices for
Sections II, III, IV, VIII and IX.
SECTION I APPENDICES
Appendix 1-1; Attachment 1, Scope of Work,
Contract 68-04-0034 91
Appendix 1-2; Prototype Vehicle Performance
Specification, January 3, 1972 95
Appendix 1-3; Federal Driving Cycle 109
Appendix 1-4; Rankine Engine Data 113
Appendix 1-5; Brayton Engine Data 117
Appendix 1-6; Idle Fuel Consumption 121
Appendix 1-7; Maximum Total Engine Power
vs. Vehicle Speed 123
Appendix 1-8; Vehicle Accessory Power Requirements 125
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Section Section
No. Title
SECTION V APPENDICES
Appendix V-l; Hydronechanical Transmission Outline
Drawing No. 2724A-E5 127
Appendix V-2; Hydromechanical Transmission Assembly
Drawing No. 2724A-L5 128
Appendix V-3; Traction Drive Transmission Outline
Drawing No. 2724A-E4 129
Appendix V-4; Traction Drive Transmission Assembly
Drawing No. 2724A-L4 130
Appendix V-5; Hydromechanical Transmission Component
Sizing 131
Appendix V-6; Traction Drive Transmission Component
Sizing 134
SECTION VT APPENDICES
Appendix VI-1; Hydromechanical Transmission/Computer
Performance Program 137
Appendix VI-2; ZB32 - Vehicle Performance Program,
Torque Converter and Traction Drive
Transmissions 147
Appendix VI-3; Vehicle Performance with a Typical
3 Speed Automatic Transmission 157
SECTION VTI APPENDICES
Appendix VII-1; Control System Parameters 165
Appendix VI1-2; Equations 167
Appendix VTI-3; Turbine Torque vs. Turbine Speed -
Rankine Engine 169
Appendix VII-4; Digital Program for Function Generation 170
Appendix VTI-5; Analog Computer Wiring Diagram 171
Appendix VII-6(1); Computer Readout for a 0.2 Per
Unit Throttle Acceleration and 50 Percent Load 173
Appendix VII-6(2); Computer Readout for a 0.2 Per
Unit Throttle Acceleration ana Zero Load 174
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LIST OF FIGURES
Figure Figure
No. Title
SECTION I. INTRODUCTION - ABSTRACT
SECTION II. RESULTS AND CONCLUSIONS
SECTION III. RECOMMENDATIONS
SECTION IV. FEASIBILITY STUDY
IV-1 Dual Mode Transmission 12
IV-2 Tri-Mode Transmission 12
IV-3 Quadri-Mode Transmission 12
IV-4 Traction Drive 15
IV-5 Effect of Torque Converter Power Absorption
Characteristics on Traction Drive Unit
Torque 17
SECTION V. TRANSMISSION DESCRIPTION
V-l Simplified Schematic 20
V-2 Geartrain Schematic 21
V-3 Hydraulic Unit Speeds 22
V-4 Speed of Various Links of Compound Summer 22
V-5 Axial Piston, Slipper Type Hydraulic Unit 26
V-6 Traction Drive-Torque Converter
Transmission Schematic 31
V-7 Traction Roller Steering Mechanism 32
SECTION VI. PERFORMANCE
VI-1 Tri-Mode Hydromechanical Transmission
Efficiency at Constant Speed -
Rankine Engine 43
VI-2 Traction Drive Transmission Efficiency at
Constant Speed - Rankine Engine 49
VI-3 Tri-Mode Hydromechanical Transmission
Efficiency at Constant Speed -
Brayton Engine 50
VI-4 Traction Drive Transmission Efficiency at
Constant Speed - Brayton Engine 51
VI-5 Hydromechanical Transmission Efficiency at
Full and Part Loads - Rankine Engine 52
VI
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Figure Figure
No. Title
VI-6 Traction Drive Transmission Efficiency at
Full and Part Loads - Rankine Engine 53
VI-7 Hydromechanical Transmission Efficiency at
Full and Part Loads - Brayton Engine 54
VI-8 Traction Drive Transmission Efficiency at
Full and Part Loads - Brayton Engine 55
VI-9 Hydromechanical Transmission Acceleration -
Rankine Engine 59
VI-10 Traction Drive Transmission Acceleration -
Rankine Engine 60
VI-11 Hydromechanical Transmission Acceleration -
Brayton Engine 61
VI-12 Traction Drive Transmission Acceleration -
Brayton Engine 62
VT-13 Hydromechanical Transmission Fuel
Consumption at Constant Speed -
Rankine Engine 64
VI-14 Traction Drive Transmission Fuel
Consumption at Constant Speed -
Rankine Engine 65
VI-15 Hydromechanical Transmission Fuel
Consumption at Constant Speed -
Brayton Engine 66
VI-16 Traction Drive Transmission Fuel
Consumption at Constant Speed -
Brayton Engine 67
VI-17 Hydromechanical Transmission Fuel
Consumption at Full and Part
Load - Rankine Engine 68
VT-18 Traction Drive Transmission Fuel
Consumption at Full and Part
Load - Rankine Engine 69
VI-19 Hydromechanical Transmission Fuel
Consumption at Full and Part
Load - Brayton Engine 70
VI-20 Traction Drive Transmission Fuel
Consumption at Full and Part
Load - Brayton Engine 71
VI-21 Maximum Tractive Effort - Hydromechanical
Transmission 77
VI-22 Maximum Tractive Effort - Traction Drive
Transmission 73
SECTION VII. CONTROL SYSTEM ANALYSIS
VTI-1 Control System Block Diagram 80
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Figure Figure Page
No. Title No.
SECTION VIII. ESTIMATED TOTAL MANUFACTURING COSTS
SECTION IX. REFERENCES
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LIST OF TABLES
Table Page
Title No.
SECTION I. INTRODUCTION - ABSTRACT
SECTION II. RESULTS AND CONCLUSIONS
SECTION III. RECOMMENDATIONS
SECTION IV. FEASIBILITY STUDY
SECTION V. TRANSMISSION DESCRIPTION
V-l Hydromechanical Transmission Weight
Breakdown 29
V-2 Traction Drive Transmission Weight
Breakdown 36
V-3 Clutch/Torque Converter Parameter
Trade-off Summary 39
SECTION VT. PERFORMANCE
VI-1 Combined Driving Cycle Transmission
Efficiency 47
VT-2 Idle, Acceleration and Grade Performance 58
VI-3 Federal Driving Cycle Fuel Consumption
with and without Air Conditioning 72
VI-4 Combined Driving Cycle Fuel Consumption 73
VI-5 Combined Driving Cycle Energy Consumption 74
VI-6 Vehicle Range at Federal Driving Cycle
and Cruise 75
SECTION VII. CONTROL SYSTEM ANALYSIS
SECTION VIII. ESTIMATED TOTAL MANUFACTURING COSTS
VIII-1 Transmission Cost Comparison 86
VTII-2 Transmission Manufacturing Cost
Breakdown Comparison 87
SECTION IX. REFERENCES
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ACKNOWLEDGEMENT
Th« EPA Project Officer was James C. Wood of the NASA-Lewis Research Center. Mr. Wood
worked for EPA under a special technical assistance agreement between NASA and EPA.
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!. INTRODUCTION - ABSTRACT
A. Introduction
During the study effort, "Hybrid Prpulsion System Transmission Evaluation — Phase I. "sponsored
by EPA under Contract 68-04-0034 and performed by Sundstrand Aviation, a transmission
configuration without a flywheel was investigated. Although the effort was minimal, it did indicate
that a transmission with an infinitely variable ratio offered many advantages for a conventional
automotive propulsion system, particularly if the engine had a limited speed range. These
conclusions, coupled with the efforts and interest by the Environmental Protection Agency Division
of Advanced Automotive Power Systems (EPA/AAPS) in the gas turbine and Rankine cycle engines,
made a more detailed evaluation of such a transmission desirable.
A study was, therefore, initiated by EPA/AAPS to quantitatively assess the technical and economic
feasibility of existing and potential types of transmissions most suitable for the gas turbine and
Rankine cycle engines. Such a study was to be aimed specifically at the AiResearch single shaft gas
turbine and the Aerojet Rankine cycle engine with a turbine expander. Both of these engines
operate over limited speed ranges, although the approach was applicable and advantageous to other
engines including the conventional spark ignition type. The Aerojet engine is now in the
development stage while the AiResearch engine is a conceptual design.
The study consisted of analytically evaluating the performance, physical characteristics and cost of
candidate transmissions. Although the study was analytical, design criteria, test data, and experience
from over 25 years involvement in the design, test and production of transmissions was used in the
analysis by Sundstrand.
Sundstrand's Aviation Division provided the program management, design, and analysis effort.
Detailed cost estimates of the transmissions were aided by personnel from Sundstrand's other
operating groups.
Requirements, scope of work, and other data utilized in and pertinent to the study are included in
Appendices 1-1 through 1-8.
B. Abstract
The study was carried out under contract to the Environmental Protection Agency, Office of Air
Programs. The object of the study was to determine the technical and economic feasibility of a
transmission to be utilized with gas turbine or Rankine cycle engines. Application of the
engine/transmission was to a full size "family car." Since the Rankine cycle prototype engine
hardware will be available before single shaft gas turbine hardware, priority was given to the
Rankine cycle engine.
The study was accomplished through a two-phase, multi-task program which included:
1) Evaluation of transmission types through a feasibility study and ultimate selection of a
transmission type.
2) Evaluation of the selected transmission type through design calculations and layouts,
performance analysis, control system analysis, and cost analysis. A number of different types of
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were initially evaluated including conventional multi-shift, hydrostatic, hydrokinetic,
electric, belt/chain, hydromechanical, and traction types. They were assessed in terms of efficiency,
technology and production status, cost, controllability, size, weight, and noise. The
hydromechanical and traction transmissions were eventually selected for more detailed evaluation.
Sketch layouts were made of both types and extensive system efficiency data was generated r,y tl.e
Sundstrand vehicle system computer model. System efficiency was the actual fuel consumption of
the vehicle, transmission, and engine system over the Federal Driving Cycle, Simplified U'ban
Driving Cycle, and Simplified Country Driving Cycle.
At the conclusion of this more "in-depth" evaluation, it was not evident that aither candidate
should be discarded. It was, therefore, decided to continue through the "Selected Transmission
Evaluation" phase with both the hydromechanical and traction transmissions. The hydromechanics!
transmission selected was a multi-mode type. It has three modes of operation - one hydrostatic and
tvr> hydromechanical (Tri-Mode). The traction transmission configuration selected was a toroidal
type with 2 rows of toroids, and output torque converter, and a forward/reverse gearbox.
detailed sizing of the two transmissions was done eventually resulting in detailed layouts of both.
in parallel with the mechanical layout effort, tne controls were defined and evaluated. An analog
study of the control system and a failure analysis were also completed.
Emmissions were a major evaluation criteria. However since little data was available relative to
various engine operating conditions, it was assumed in conjunction with EPA that minimum
emmissions occurred at maximum fuel economy. System efficiency, therefore became a major
evaluation criteria with considerable effort being expended in this area. Detailed system fuel
consumption values were determined utilizing a vehicle system computer simulation in which
vehicle, engine, transmission, and duty cycle characteristics were programmed. Both transmissions
were evaluated with the two specified engines — Aerojet Rankine cycle with a turbine expander and
the AiResearch regenerated single shaft gas turbine. The engine fuel consumption data for the
engine was supplied by EPA/AAPS at the initiation of the study and is reflected in the values shown
in the report. Modifications to this data have occurred particularly relative to the Aerojet engine,
which would change the absolute values in the report.
Cost estimates were made for both transmissions using comparative data, vendor quotations, and
in-house estimates. The resulting costs were also compared with a conventional three speed
automatic transmission. The study resulted in the determination that both the hydromechanical and
traction transmissions were technically and economically feasible for application to limited speed
range Rankine or gas turbine engines. Each had certain features better than its competitor and, in
some instances, they were equivalent. The hydromechanical transmission has slightly better
efficiency, is shorter, and represents less technical and program risk to develop a pre-prototype unit.
The traction transmission is estimated to be lower in cost, is lighter, and is inherently quieter.
A specific selection of the best type will have to be made by weighing the various characteristics, in
terms of their overall importance; a task not undertaken in this study.
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II. RESULTS AND CONCLUSIONS
1) Infinitely variable ration transmissions are the most flexible and best type for application to
engines required to run over a narrow speed range. They are also applicable to all other types of
engines, including the conventional spark ignition type, providing the ability with all engines to
optimize a selected characteristics such as fuel consumption, emissions or performance.
2) The hydromechanical and traction transmissions are the most feasible and desirable types to
provide an infinitely variable ratio.
3) Based on the given engine data, the hydromechanical or traction infinitely variable ratio
transmission becomes more attractive when utilized with the Ai Research single shaft gas turbine
engine than with the Aerojet Rankine cycle engine due to the shape of the specific fuel
consumption curves.
4) A multi-mode (3 mode) hydromechanical transmission and a two-element traction ratio
changer combined with an output torque converter and forward / reverse gearbox are the two
configurations selected as the best candidates for use with the AiResearch gas turbine and the
Aerojet Rankine cycle engine. The use of a slipping clutch rather than the torque converter with the
traction ratio changer may be acceptable although further study is required.
5) Overall feasibility assessment of the hydromechanical and traction / torque converter
transmissions indicates that their specific features are similar with the overall rating dependent upon
the relative importance of the various criteria. Transmission comparison utilizing two important
criteria, efficiency and technology status, indicates that the hydromechanical transmission has a
higher overall efficiency and presents a lower overall risk to produce a pre-prototype by early 1974.
Efficiency of the traction/slipping clutch transmission is equal to that of the hydromechanical
transmission.
6) Both the hydromechanical and traction transmission compare favorable with a typical three
speed automatic transmission in terms of performance, weight, and size. The cost increase must be
viewed in terms of the increased capability and flexibility of this type transmission. A comparative
summary is as follows:
Actual
Weight, Relative Actual Relative Volume
Type Pounds Weight Cost, $ Cost In^
Hydromechanical 92 0.80 122 1.37 1390
Traction 77 0.67 105 1.18 1275
3 Speed
Automatic 150 1.00 89 1.00 2500
7) The hydromechanical or traction transmission can be utilized with either the single shaft gas
turbine or the Rankine cycle engine with minor modification to the control system and a change in
torque converter diameter.
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8) The computer simulated performance of the full size automobile, Rankine cycle engine, and
either the hydromechanical or traction transmission met or exceec-J all start-up, acceleration,
gradeability, and maximum speed requirements of "Prototype Vehicle Performance Specification"
dated January 3, 1972.
9) The computer simulated performance of the full size automobile, Brayton cycle engine, and
either the hydromechanical or traction transmission met or exceeded all start-up, acceleration,
gradeability, and maximum speed requirements of "Prototype Vehicle Performance Specification"
dated January 3, 1972. However, to meet the 0-10 second acceleration requirement, the Brayton
cycle engine power must be increased to 155 HP (12% above specified) when coupled to the
traction transmission and to 145 HP (5% above specified) when coupled to the hydromechanical
transmission.
10) The hydromechanical transmission is inherently noiser than the traction / torque converter
transmission. However, utilizing presently known and demonstrated noise reduction techniques, it is
anticipated that the vehicle noise requirements of "Prototype Vehicle Performance Specification"
dated January 3, 1972, can be met.
11) The control of either the hydromechanical or traction transmission is essentially the same. It is
compatible with the specified engines, is stable, is simple, and provides a "driver-feel" comparable
to existing automative transmissions.
12) Installation of the hydromechanical or traction transmission is compatible with the specified
engines and when installed in a full size automobile as specified, requires no modification to the
structure.
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III. RECOMMENDATIONS
Initiate a hardware development program for an infinitely variable ratio hydromechanical or
traction transmission. Such a program should include the design, manufacture, dynamometer test
and vehicle test of the selected transmission for a specific engine. Analysis of the transmissions with
the two specified engines would seem to indicate that the single shaft gas turbine provides a better
application.
While selection of the specific transmission type is dependent upon the relative importance given to
the various transmission parameters, the hydromechanical type is the better candidate based on
higher efficiency and lower development risk.
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IV. FEASIBILITY STUDY
A. INTRODUCTION
1) Methodology
Identification of a large number of possible concepts indicated that within the short study time
available, a successive screening process would have to be utilized. Inital screening was a "gross"
process in that basic, known characteristics were evaluated and those concepts which were
promising were considered further.
As each screening process was undertaken more detailed evaluation was made. System
characteristics of efficiency cost, controllability, size, weight, noise, and technical and production
status were considered in varying degrees depending on the particular screening level. For the most
promising candidates sketch layouts were made to determine hardware feasibility and complexity.
Also a vehicle system computer model was utilized to determine vehicle eff icency and performance.
In most instances, information for evaluation was from "in-house" sources including test data,
studies, patent files, and literature searchs.
2) Requirements
Comparison of the technical and economic feasibility of the various transmission types for use with
the two specified engines — the Aerojet Rankine Cycle and the AiResearch Brayton cycle engines
required consideration of the following parameters.
a) Vehicle Parameters
Detailed vehicle description and performance requirements are given in Appendix I-2
"Prototype Vehicle Performance Specification." A summary of the more important
parameters that affect the transmission are:
• Test Weight — Wt. = 4600 pounds. This weight to be used for all fuel economy and
acceleration calculations.
• Gross Weight - Wg. = 5300 pounds. This weight to be used for sustained velocity
calculations at 5 and 30% grades.
• Maximum Vehicle Speed — 85 mph.
b) Engine Parameters
Brayton Cycle Rankine Cycle Rankine Cycle
47,460 Minimum Turbine Output Speed - RPM 16,000
83,055 Maximum Turbine Output Speed - RPM 22,000
145 (Tri-Mode) Maximum Power - HP 163
155 (Traction)
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Engine speed to be controlled by the transmission to give minimum specific fuel
consumption over the operating load range.
c) Other
Assumed maximum tractive effort — 2500 Ib.
B. ANALYSIS OF TRANSMISSION TYPES
1) Discussion of Types
a) Multi-Shift Fixed Ratio - Power Shifting Gearbox with Start Clutch for Fluid Coupling/
Converter
This type of transmission would need an energy dissipating device for starting the vehicle from
rest (friction clutch or fluid coupling) and would need 6 to 8 fixed ratios to keep the engine
speed within limits over the vehicle speed range. The primary advantages are good efficiency
and existing high production technology. The disadvantages are: inability to control engine
speed to any required operating curve; discontinuous power to the wheels; and power shifting
life limiting clutch packs. The cost of this type of transmission would be comparable with a
traction or hydromechanical type infinitely variable ratio transmission but without their
advantage of smooth, precise engine speed control.
b) Electric - Engine Driven Generator, D.C. or A.C. Motor Driven Wheels and Electronic
Solid State Controls
Where good efficiency, light weight, and low cost are requirements, electric transmissions are
invariably rejected. Efficiency of this transmission system is normally in the 60 to 70% range.
A study report (Reference 1, Section IX) shows in Table 4, page 25, that a family car system
weight exluding the A.C. generator power source is 348 pounds. Adding 100 pounds for the
generator, rectifier, speed increaser, and voltage regulator totals 448 pounds. The components
considered were all lightweight aircraft types. This total of 448 pounds is at least twice that of
the conventional 3 speed automatic transmission plus drive shaft and rear axle.
The cost would be at least twice that of a conventional torque converter automatic
transmission. Cost and weight, therefore, eliminated the electric transmission from further
consideration.
c) Hydrostatic — Engine Drive Variable Displacement Hydraulic Pump with Close Coupled
Fixed or Variable Displacement Hydraulic Motor
Pure hydrostatic transmission are invariably rejected for high speed vehicles where high
efficiency, light weight, and small size are important. Hydraulic unit displacement required to
transmit full rated HP at top vehicle speed can be as much as 18 times that required in a 3
mode hydromechanical transmission for the same vehicle. It is readily apparent that a pure
hydrostatic transmission is unacceptable for a high speed vehicle. Overall efficiency of the pure
hydrostatic transmission is considerably lower than the hydromechanical transmission type.
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d) Hydrokinetic — Torque Converter of Fluid Coupling
Hydrokinetic and aerodynamic torque converter / coupling devices can be used as the principle
means of varying the speed ratio across a transmission, such as a converter in series with a one,
two or three speed gearbox. (This category differs from the 6-8 speed multi-shift fixed ratio
devices discussed in (1) where the principle means of varying the speed ratio is by changing
gear ratios, and the clutch or converter is basically only a device for starting from rest.
This method of speed control is considered unsuitable because —
The converter is an energy dissipating device, and therefore inefficient unless operating
close to the coupling, or lock up point. Wide vehicle operating speed ratios require converter
operation considerably removed from the coupling point, in the very inefficient range.
The speed ratio for fixed blading type converters cannot be controlled at will, but is a
function of the instantaneous load and speed condition. Variable blading does allow some
measure of independent speed control, but at the expense of further operating efficiencies.
Converters, however, do work well and have their advantages when combined with some other
means of ratio control, such as the traction transmission.
e) Belt-Chain — Variable Sheaves either Belt or Chain Driven
This type transmission must also be provided with an energy dissipating clutch or hydraulic
coupling for vehicle start up as the output of the variable ratio belt cannot be brought down to
zero speed.
Many belt or chain variable speed transmissions have been developed and used successfully in
the machine tool and stationary construction or industrial machinery where a stepless variable
output speed is advantageous. For automobile transmission, several have been built and tested
in low power vehicles, but data is not available. Chains and belts are life limited items and are
also speed and power limited. Because belts are a high maintenance item in even normal
accessory drives, great efforts are being expended to replace them with more reliable drives
such as hydrostatic. Therefore, further investigation of this type of transmission was dropped.
f) Hydromechanical — Engine power is transmitted through both mechanical and hydraulic
paths to obtain infinetly variable ratios.
The hydromechanical transmission is an infinitely variable ratio transmissions and therefore
offers maximum flexibility. Considerable development work has been accomplished on this
type of transmission and hardware can be developed with a minimum amount of risk. Size,
weight, durability, controllability are proven. Efficiency, although probably lower than some
other alternatives, does provide an equivalent or better system efficiency due to its ability to
operate the engine at its optimum fuel consumption.
Noise represents a potential problem although noise reduction techniques developed over the
last few years should provide acceptable noise levels.
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g) Traction — Transfer of torque and speed through friction contact
The traction transmission is also an infinetly variable ratio device. The traction drive has
potential benefits of high efficiency, low vibration, and low noise. Development of a new
traction fluid by Monsanto has also improved force and stress levels such that it looks more
attractive.
2) Conclusions
Based on the analysis of the various types of transmissions, it was concluded that an infinetely
variable ratio transmission was the best type of transmission for the limited speed range engines
being considered. The hydromechanical and traction transmissions are the best infinetely variable
ratio candidates. It was therefore decided to continue with a more detailed optimization of both of
these candidates with one to be selected at the conclusion of that phase.
C. OPTIMIZATION OF SELECTED CANDIDATES
1) HYDROMECHANICAL
The typical hydromechanical circuit schematic is shown in the following sketch.
HYDRAULIC UNITS
^
TORQUE
SUMMING
POINT
(SINGLE GEAR
MESH)
zf,
SPEED SUMMING
POINT
(GEARED
DIFFERENTIAL)
The "V" unit is the variable displacement axial piston pump/motor type and the
"F" unit is the fixed displacement axial piston pump/motor type. The two units are hydraulically
ported to each other so that when one is a pump the other is a motor, and vice versa. The "F" unit
swashplate (or wobbler) is at a fixed angle of 14° and the "V" unit can be varied to any desired
angle from 0° to ± 15-1/2°. Both swashplates are non-rotating.
For best utilization of the hydraulic unit to cover the complete operating range of the vehicle, it is
desirable to run the "V" unit at its rated speed, and run the "F" unit through its full operating
speed range (plus to minus rated speed).
Hydraulic unit efficiency is a function of speed, pressure and wobbler angle.
Studies conducted in 1966 for TACOM R & E Directorate under contracts DA-11-022-AMC-695
(T) and 2269 (T) as well as in 1971 under Phase I of EPA Contract 68-04-0334 covered many
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different types of hydromechanical transmission schemes. Some of those eliminated from
consideration then, and are eliminated now for similar reasons, were the simple output differential,
input differential, multiple hydraulic units such as the David Brown, Stratos, or Ebert designs, and
the Borg-Warner multiple path arrangement which utilizes four variable displacement hydraulic units.
As stated previously, some hydromechanical schemes were eliminated through analysis of:
Hydraulic unit size, number of hydraulic units, quantity of power transmitted hydraulically, number
of clutches, amount of gearing, and overall complexity.
The remaining schemes for final analysis were the Dual Mode (DMT), theTri-Mode (TMT), and the
Quadri-Mode (QMT).
Hydromechanical transmissions can be designed to operate in more than one mode in order to
reduce the maximum HP that is transmitted hydraulically, which reduces the hydraulic unit size,
and increases efficiency. This may be achieved by having a straight hydrostatic mode for start-up
(whence all the engine power is transmistted hydraulically), and then one or more hydromechanical
modes of operation. In a single mode of hydromechanical operation, the fixed displacement
hydraulic unit is typically operated over its full rated speed range; that is, from plus its maximum
rated speed down through zero speed and up to minus its rated speed. Two modes of
hydromechanical operation then would typically consist of a system of gears and clutches that
would operate the fixed displacement unit over its full plus to minus rated speed, two times, and
three hydromechanical modes would do this three times, etc. The gears and clutches are arranged in
such a manner that there is no speed discontinuity at each clutch shift point when changing from
one mode of operation to another. Thus, all clutch shifting is done at synchronous speeds and no
power is absorbed by the clutches. The more modes that are added, the more efficient the
transmission can become, but at the cost of increased complexity of gears and clutches.
Sundstrand has considered the following hydromechanical transmission types:
a) The Dual Mode Transmission (DMT)
This transmission has one hydrostatic mode and one hydromechanical mode, schematically
represented in Figure IV-1. Sundstrand has developed a transmission of this type for the heavy
duty truck market and will be in production in 1973.
b) The Tri-Mode Transmission (TMT)
This transmission consists of one hydrostatic mode, and two hydromechanical modes. There
are many ways of achieving this schematically; one way is shown in Figure IV-2.
c) The Quadri-Mode Transmission (QMT). This transmission consists of one hydrostatic
mode and three hydromechanical modes. One way of achieving this is shown by the schematic
in Figure IV-3. It should be noted that this schematic is similar to the Orshansky Transmission
Corporation "Three Range Transmission" which was also evaluated. The Orshansky
transmission requires, however, one additional friction element.
In evaluating the optimum number of modes for this application, Sundstrand chose the Tri-Mode
Transmission as being the best compromise between efficiency and complexity. As more clutches
Sundstrand Aviation age
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Hh
Figure IV-1 Dual Mode Transmission
1 I 1
Figure IV-2 Tri Mode Transmission
INPUT
•
•
IB «
L T
~ I
V
c
A
\
k .
Figure IV-3 Quadri-Mode Transmission
Page 12
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are added to obtain increased number of modes, the spin losses from these clutches become
substantial, and reduce the gains in efficiency from this larger number of modes. As an example,
following is a tabulation including some relative cost and efficiency data for Dual Mode, Tri-Mode,
and Quadri-Mode Transmissions.
Sundstrand Hyd. Unit No. of Fuel Consumption
Transmission Displacement No. Clutches Gears Cost (over Veh. Life)
DMT 100% 2 11 100% 100%
TMT 42% 3 15 114% 94%
QMT 22% 4 21 125% 92%
2) TRACTION
The potential benefits of high efficiency, low vibration and low noise of traction transmission are
well known. With the availability of a suitable fluid and a design to reduce forces and stresses, the
traction drive becomes competitive enough to warrant a more thorough evaluation.
Variable ratio traction transmissions of various types have been studied over the years. The toroidal
type has emerged as the best design for highest power density, reasonable life, and good efficiency.
Sundstrand, Lycoming, Rotax, General Motors, English Electric, and others have built and tested
the toroidal types; Tractor also is developing a modified torodal type. Lycoming has been the only
company to market a toroidal traction transmission but other companies have successfully tested
prototype designs.
There are many options open in the design of a toroidal traction variable speed drive. One of the
basic design considerations is the use of a two row (dual toroid) design or a single row (single
toroid) design.
A single row toroidal drive derives its name from the fact that there is only one set of traction
rollers, typically three rollers per set, that transmit power between the input toroidal disk and the
output toroidal disk.
In a two row device, the power flow is split from the center, or input toroid, through two sets of
traction rollers. The two sets of rollers, with one set on either side of the input toroid, transmit
power to the output toroids which are located at each end of the traction drive.
Large thrust loads are required at the toroid/roller interface to provide the normal force necessary
to prevent gross slip between the rollers and the toroids. The benefit from the use of a two row
device is that since it is symmetrical, the thrust loads from the two halves of the unit cancel each
other and do not have to be taken thru thrust bearings. On the other hand, a single row traction
drive requires very large thrust bearings to react to the large axial thrust loads, and also, since there
is load sharing between the two halves of a two row device, the required toroid diameter for a two
row unit is considerably smaller than for a single row unit.
Sundstrand Aviation 4»^ age'
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One of the most important considerations throughout this study has been system energy efficiency.
The relatively large power loss that would be encountered in the large high speed thrust bearings of
a single row unit make the two row design much more attractive from an efficiency point of view.
Although a single row traction drive would cost less to produce, it is felt that the advantages to be
realized from using a two row unit more than offset the cost.
The basic scheme of the toroidal type is shown in Figure IV-4. Principal components are the input
toroid, rollers, and output toroid. The rollers are steerable and are "steered" to the necessary angle
to provide the input to output speed ratio desired. When the input toroid is rotated, the rollers turn
and exert a traction force on the output toroid. The toroids must be held together to insure
sufficient traction exists with the rollers to transmit the desired power. The rollers are steered or
tilted to the angle which produces the desired output speed. With the rollers angled as shown in the
figure, output speed is lower than input speed. At the opposite angle, output speed is higher than
input. Power capacity for a given size and given number of rollers is a function of the clamping
force between input and output toroids across the rollers, and the traction coefficient of the fluid
being used. The torque producing force at the point of roller contact is the product of the clamping
force and the traction coefficient. Life of the unit is a function of this force and devices have been
developed to vary this force in proportion to the load with a resultant increase in life.
The traction transmission design for this application must have a disengaging device to permit zero
output speed when the engine is running. Three ways of achieving this were considered, and they
are listed below along with the effects of each type on the traction drive unit.
a) Slipping Clutch
• The traction drive unit must be sized to take the maximum slipping torque of the clutch
which must be greater by some 10-30% than the maximum torque generated by the engine.
For narrow speed range engines, where minimum speed is close to maximum speed, the power
which must be dissipated in the clutch is high.
• The slipping clutch does not readily offer any load "shock absorbing" protection to
the traction unit. This is an important consideration with variable thrust types of traction
drive units in that thrust must always be maintained sufficient to prevent gross slip
between roller and toroid. Slips in excess of approximately 3% result in decreased torque
ability and damaging spin out to even greater slips until the load is removed. These
shock loads could be seen in the drive train under such conditions as accelerating on an
ice-patched surface or sudden wheel-lock when braking.
• Once locked up, the slipping clutch does have the advantage of virtually zero power loss,
giving the most efficient traction drive system of those being considered.
b) Split-Path with Planetary Summer
This consists of a power splitting circuit, similar to the hydromechanical concept, but replacing
the hydraulic unit with a traction unit.
• The traction drive unit would have to be upsized to take the high power which can
recirculate through this type of power circuit at start-up. The hydraulic ratio control system
Pa9e 14 Sundstrand Aviation
0..-W or S,"d*t'«"-*'-oi ^P J^ s
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ROLLERS
CLAMPING FORCE
OUTPUT
TOROID
CLAMPING FORCE
INPUT TOROID
CLAMPING FORCE
CLAMPING FORCE
Figure IV-4 Traction Drive
Sundstrand Aviation ™
Page 15
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could be used to act as a torque limiter. Without torque limiting, the torque would
theoretically go to infinity under conditions of holding the output stalled at wide open
throttle. The traction drive unit would have to be sized to take this limiting torque value,
which could be several times greater than maximum engine torque.
• This system does not offer any load "shock absorbing" protection for the traction drive
unit.
• The split-path system is non-dissipative and would be efficient. It could also be arranged
to give a "built-in" reverse ratio capability, obviating the need for a reverse gearbox.
c) Torque Converter, or Fluid Coupling
The torque converter can be placed on either the input or output of the traction drive unit.
• The speed versus horsepower absorption characteristics (torque at a given speed) of an
output torque converter can be utilized to down-size the traction drive unit by reducing the
maximum torque that can be seen by the traction drive output. This is illustrated graphically
in Figure IV-5. An input mounted torque converter will up-size the traction drive.
• The output mounted torque converter offers a high degree of load "shock absorbing" for
protection of the traction unit. The input mounted converter offers lesser protection.
• The torque converter system would be the simplest and most reliable although less
efficient than the other two schemes described previously.
From these considerations it was decided to use an output mounted torque converter as it is
the only device which gives any degree of shock load protection to the traction drive unit, and
it allows the use of the smallest possible traction unit. The torque converter has the additional
advantages of low cost, excellent reliability and virtually zero maintenance. The efficiency
penalty, in terms of MPG over the EPA Combined Driving Cycle, is about 5-8% relative to that
which a slipping clutch start-up system could achieve.
3) Comparison of Final Hydromechanical and Traction Design Choices
a) Size and Weight
When the two finalist transmiss;on designs were decided upon, preliminary hand sketched
layouts were made of each. The ? layouts indicated that both transmission schemes had the
potential to become practical automotive transmissions from a size and weight perspective.
b) Efficiency and Fuel Consumption
An exhaustive computer analysis was completed on each of the two finalist transmission
candidates using two full scale computer programs. Results of the computer analysis indicated
that both transmission types offer the capability to program engine speed to minimize fuel
consumption and / or emissions at a high level of energy efficiency. The computer efficiency
and consequently the fuel consumption of the two systems were very close.
16 *».,•....
Sundstrand Aviation
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ENGINE
TRACTION
DRIVE
UNIT
(A) TORQUE
T -.CONVERTER
/J*\ ^TO
\ ) WH
Speed~ torque curve
for torque converter with
stalled output at point®
Speed~ torque available at
traction drive output (Point@)
at max. Engine HP
4 _
Actual torque converter
input speed at stall
SPEED, N
Minimum Traction drive output speed
as limited by the ratio range limit.
The torque converter limits the maximum output torque of the traction drive unit to T^ rather
than T^ without a converter. The torque converter then, allows the traction drive unit to be
sized to torque T^.
Figure IV-5
Effect of Torque Converter Power
Absorption Characteristics on
Traction Drive Unit Torque
Sundstrand Aviation
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c) Cost
A preliminary total manufacturing cost comparison based upon a rate of 1 million per year
indicated that the hydromechanics! transmission would probably cost slightly more than the
traction drive transmission.
d) Noise
The traction drive torque converter type transmission is inherently quiet. Attention must be
given to the gearing to ensure that noise is reduced to a minimum. Sundstrand is committed to
meeting acceptable noise levels with the present production Dual Mode hydromechanical
transmission (DMT). Experience gained from testing the DMT has been applied in the
proposed Tri-Mode hydromechanical transmission design. It appears at this time that the noise
from either transmission can be brought within acceptable limits.
e) Producibility
There are no totally new unobtainable materials or processes involved in the manufacture of
either transmission type. Tooling requirements will be similar.
f) Technology Status
The hydromechanical transmission represents a more highly developed device than the traction
type. More companies are involved in actual testing and evaluation of hydromechanical
transmissions than with traction transmissions. Production hydromechanical vehicle
transmissions are being offered for sale, while production traction transmissions have been
produced only for aircraft constant speed drive applications. Design and development of a
pre-prototype transmission by late 1973 or early 1974 can be accomplished for either
transmission type, although the traction type represents a somewhat greater risk. Either type
could be ready for production by 1980.
The hydromechanical transmission primary development task will be the integration of its
controls with the engine. Although the basic control scheme has been mechanized and
demonstrated, operation with the specified engines will require additional effort.
The major development task for the traction type transmission is assurance and demonstration
of the required life. Since the life capability is highly dependent upon the vehicle load (toroid
and roller stress), vehicle and engine speed (traction ratio) and time at each condition,
determination of the actual vehicle duty cycle is very important. Since little experience has
been obtained with traction transmissions in vehicles this definition of the "real" operational
requirements and the mechanical design reflecting these parameters, becomes the major
development item.
g) Conclusion
After consideration of this comparative evaluation, Sundstrand felt they could not justify
dropping either transmission from further consideration. As a result, it was decided to
continue the detailed evaluation of both transmissions through to the completion of the study.
Pdoe 18
Sundstrand Aviation
C»'»>on jl Sur>«ttr«na Corporation
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V. TRANSMISSION DESCRIPTION
A. HYDROMECHANICAL
1. Mechanical Operation
The following is a discussion of the mechanical operation of the tri-mode hydromechanical
transmission with regard to the direction of power flow, component speed and torque relationships,
and variable unit displacement. The transmission is shown in simplified schematic form on Figure
V-1.
The transmission has three distinct modes of operation in forward. At 100% engine speed, the shift
between Mode 1 and Mode 2 occurs at 12.3 MPH, and the Mode 2 to Mode 3 shift occurs at 40.8
MPH. At lower engine speeds the shift points occur at proportionately lower vehicle speeds. During
Mode 1, the output from the fixed displacement hydraulic unit is geared directly to the output. In
Mode 2 and Mode 3 operations, the fixed unit is geared into the plantetary. Reverse is the same as
Mode 1, but in opposite direction and is obtained by stroking the variable displacement hydraulic
unit in the reverse direction.
Figure V-2 shows schematically the geartrain arrangement.
a) Component Speeds:
The variable displacement hydraulic unit is geared directly to the engine, therefore, its speed
will always be directly proportional to engine speed.
In Mode 1 the fixed displacement hydraulic unit is geared directly to the output planetary
link, so in Mode 1 its speed will be directly proportional to output speed, hence, vehicle speed.
When the fixed displacement hydraulic unit speed increases to the point where it is equal to
variable displacement hydraulic unit speed, a mode shift from Mode 1 to Mode 2 occurs.
In Mode 2, the fixed displacement hydraulic unit is geared to a leg of the planetary which
causes power to be transmitted both hydraulically through the hydraulic units and
mechanically through the planetary. The fixed unit speed decreases with increasing vehicle
speed until it passes through zero speed and then increases in the opposite direction. When the
fixed displacement hydraulic unit speed increases to minus one times the variable displacement
hydraulic unit speed, a second mode shift from Mode 2 to Mode 3 occurs. Both Mode 1 and
Mode 2 shifts are accomplished when the driving and driven clutch discs are at essentially equal
speeds.
In Mode 3, the fixed displacement hydraulic unit is geared to another leg of the planetary,
different from that of Mode 2, which again causes power to be transmitted both hydraulically
and mechanically. The characteristics of Mode 2 and Mode 3 are very closely related, the only
difference being the speed and torque ratios between the various elements. Increasing vehicle
speed further after the Mode 2 to Mode 3 shift results in decreasing fixed displacement
hydraulic unit speed (from its negative maximum) until it passes through zero, and then
increases to its positive maximum speed (one times the variable displacement hydraulic unit
speed) at maximum vehicle speed. Figure V-3 shows the hydraulic unit speeds schematically.
The speeds of the various links of the compound summer (in this case a four element
planetary) can also be represented on a nomograph, shown on Figure V-4. A straight line
Sundstrand Aviation * ^ a9e 19
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ENGINE
V
F
OUTPUT
V = Variable Displacement Hydraulic Unit
F = Fixed Displacement Hydraulic Unit
£4 = Four Element Differential
1 = Mode 1 Clutch
2 = Mode 2 Clutch
3 = Mode 3 Clutch
Figure V-l Simplified Schematic
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INPUT
V F
MODE 2 MODE 3
CLUTCH CLUTCH
T
~T
MODE 1
CLUTCH
rm
ill
OUTPUT
Figure V-2 Geartrain Schematic
Sundstrand Aviation
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PERCENT
SPEED
ENGINE & V-UNIT SPEED
+100%
0%
-100%
Figure V-3 Hydraulic Unit Speeds
% VEHICLE
SPEED
'F'UNIT MODE 2
OUTPUT
ENGINE
('F' UNIT, MODE 1) CV'-UNIT)
'F'UN IT, MODE 3
ZERO"?
SPEEDj
*
Figure V-4 Speeds of the Various Links
of the Compound Summer
Page 22
Sundstrand Aviation fc
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passing through any two link speeds defines the speed of the other two links. Thus, when
output speed and engine speed are known, fixed and variable displacement hydraulic unit link
speeds can be found. Since the fixed and variable displacement hydraulic units are related
directly to their respective planetary links by gear ratios, all the system speeds can be
calculated.
b) Hydraulic Unit Displacement:
The displacement of the variable displacement hydraulic unit can be calculated from the flow
continuity equation. This equation is shown below in its simplified form (neglecting
volumetric efficiencies):
Where:
Q= Flow (in3/min)
D = Displacement (in^/rev)
F = Fixed Unit
V = Variable Unit
N = Unit Speed (RPM)
Thus:
N
xD,
Therefore, for any given engine speed, the displacement of the variable displacement hydraulic
unit will vary directly proportionally to fixed displacement hydraulic unit speed (see the
following sketch).
PERCENT
V-UNIT
_ DISPLACEMENT
100%
% VEHICLE
SPEED
-100%
Sundstrand Aviation
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c) Torque:
The reaction torques in a compound summer may be represented as vectors acting on a beam
at positions that correspond to the link locations on the speed nomograph (Figure V-4).
Unknown torques may be found by applying the equations of statics to the torque
vector-beam analogy of the planetary (see the following sketch).
TYPICAL CASE: (MODE 2)
t
F-UNIT
| OUTPUT
1
i
V-UNIT
1
t
>
— l
i
ENGINE
Although there is more involved when efficiency is taken into account, the V-unit and F-unit
torques are related by the equations:
HPHYD = TVNV = TFNF
l!f.
TV = TFX Nv
Where:
HPj^YD = Hydraulic horsepower
T = Torque
N = Speed
V = Variable unit
F= Fixed Unit
d) Hydraulic Unit Pressure:
When the torque balance is solved for any given set of external load and speed conditions, the
working pressure can be calculated directly from the F-unit torque reaction. The basic formula
relating F-unit torque and the working pressure is:
Page 24
Sundstrand Aviation
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p , 2irxTp
~DF~
Where:
P = Working pressure (psi)
Tp * Fixed unit torque (in-lb)
O
Dp = Fixed unit displacement (in°/rev)
e) Horsepower:
Horsepower is the product of torque times speed. The basic methods of solving for torque and
speed in the transmission were defined previously. The magnitude of the horsepower in any
link is the torque in that link times the speed of that link divided by the appropriate
dimensional constant.
The direction of horsepower flow, on the other hand, must be determined from the direction
of link rotation and the direction of applied torque. Sign conventions were established for the
planetary speed nomograph (Figure V-4) such that any speed above the nomograph absissia is
positive, and any speed below is negative. In the planetary torque balance beam (sketch), any
vector pointing up is positive and any vector pointing down is negative.
2. Hardware Description
The following is a brief description of the various components which make up the tri-mode
hydromechanical transmission. Reference should be made to the cross section drawing, 2724A-L5,
shown in Appendix V-2 for indication of component arrangement and relative size.
a) Hydraulic Units:
The hydraulic units are the axial piston hydrostatically balanced configuration, typical of
Sundstrand's standard line of hydraulic units for the aircraft, agricultural, and construction
equipment market.
Figure V-5 shows a schematic cross section of a typical hydraulic unit of this configuration.
While a variety of hydraulic pump/motor units could have conceivably been evaluated for this
application, Sundstrand based hydraulic unit selection on our extensive experience in designing
hydrostatic and hydromechanical transmissions for a variety of applications over the last 30
years.
The hydraulic units are identical in construction to hydraulic units presently being
manufactured by Sundstrand for hydromechanical transmission applications where they have
proven their reliability, low cost, and good efficiency.
o
Both hydraulic units have a displacement of 1.5 inj/rev. One unit is variable displacement, the
other is fixed displacement. The units are designed for 3000 psi nominal, 7500 psi overloads,
and 9000 psi proof pressure.
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High Pressure
Oil Film
Figure V-5 Axial Piston, Slipper Type Hydraulic Unit
Sundstrand Aviation (
Page 26
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The units are mounted side by side with a common port plate manifold. Mounting the units in
this manner provides for a shorter transmission length and allows for better noise reduction
techniques to be utilized.
b) Clutches:
The clutches perform the shift function during the change from one mode to another.
These clutches are of the conventional mu I tip late disc type common to automotive
applications. They are simple to control, inexpensive, and have high torque capability. At the
shift, the shaft speeds are essentially synchronized, thereby allowing the use of light duty
clutches and are thus sized on torque capability and not energy dissipation.
Clutch design follows standard automotive practice. Steel separator plates are used with
organic linings and the drums are ductile cast iron. The piston and the back-up ring are
aluminum.
A centrifugal operated pressure sensitive check valve is incorporated within each clutch to
preclude centrifugal pressure from actuating the clutch.
c) Seals:
Standard lip seals are used on the transmission input and output shafts.
Rotating seals between concentric shafts are of the cast iron piston ring type common with
standard automotive practice.
d) Gears:
Helical gears have been assumed throughout the transmission, as in all automotive
transmissions, to minimize noise. The gears are all designed to permit use of economical mass
production techniques.
e) Charge Pump:
The charge pump is of the gerotor type common to automotive applications. It has been sized
to provide for main hydraulic unit charging, control operation, clutch application and cooling,
gear and bearing lubrication, and flow to the transmission cooler.
f) Bearings:
Extensive use has been made of radial and thrust load needle bearings. Bearings of this type are
widely used in automotive applications as they are inexpensive, reliable, and have minimum
lubrication requirements.
Tapered roller bearings are used in the hydraulic units as needle bearings are not suitable at
these locations.
. » Page 27
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g) Component mounting:
The clutch pack assemblies and the planetary gear set are mounted on coaxial shafting which is
supported by bearings at the input and output of the main housing and at the intermediate
support plate. The hydraulic unit assembly and its drive gears, along with the charge pump, are
mounted entirely on the intermediate plate which is mounted to the main housing. This type
of construction allows for very easy assembly, maintenance, and gives the best possible noise
isolation.
h) Controls:
The spool control valves are typical of those found in present automatic transmissions. The
valve bodies are cast iron, the spools are hardened steel and, where applicable, steel sleeves are
used.
The control linkages from the driver will be of similar type and construction to those presently
used in automotive applications.
Speed sensing governors are of the rotating flyweight type and act directly on a valve stem.
i) Transmission Cooler:
The transmission cooler is not an integral part of the transmission and is listed here only as a
reminder that it is required to dissipate the heat generated in the transmission. As this
transmission does not vary speed ratio by dissipating energy, such as the torque converter, the
cooling capacity would be less than required for a conventional automatic transmission while
the transmission fluid flow rate will be about the same.
3. Size and Weight
The tri-mode hydromechanical transmission is designed to fit within the requirements stated in
paragraph 6 of the "Prototype Vehicle Performance Specification" (see Appendix 1-2). In brief, the
transmission tunnel is not widened so as to decrease clearance between the accelerator pedal and the
tunnel; the tunnel height does not affect full fore and aft movement of the front seat; it does not
violate the ground clearance lines; it does not violate the space allocated for wheel jounce and
steering clearances; and it does not degrade the handling characteristics of the vehicle.
The input or mounting flange is not a standard to fit the conventional internal combustion engine.
However, as a reduction gearbox is required at the Rankine or Brayton cycle engine output, the
mounting flange and output shaft location may be located to suit the proposed transmission.
The weight of the tri-mode transmission is 92 pounds dry. A weight breakdown is shown in
Table V-1.
4. Design Analysis
By far, the majority of components in an automotive transmission are sized by considerations other
than material stress such as economy of manufacture, or requirements of fitting over or around
some other component. When weight is not a major consideration, components are often oversized
to "keep out of trouble," and no heed is taken or calculations made of the exact margin of safety.
Page 28 Sundstrand Aviation
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Obvious exceptions to this are gears, highly torqued small diameter or thin walled shafting, bearings
that see predictable loads, and clutches (or other forms of friction elements). Appendix V-5 gives
the summary of sizing this class of component along with a schematic which shows torques and
speeds. The hydraulic units are sized by proprietary Sundstrand methods to meet their rated speeds
and pressures.
In a study of this type where basic concept and feasibility are of prime importance, it is not
appropriate to go into extensive sizing detail analysis. This is especially true when the design is being
TABLE V-1
HYDROMECHANICAL TRANSMISSION WEIGHT BREAKDOWN
Planetary Gearset 5. 3
Transfer Gears 3. 8
Idler Gears and Bearings 1.2
Shafting and Bearings 9. 5
Clutches 13.7
Hydraulic Units (excluding shafts) 20. 0
Housing and Port Plate 24. 7
Control System and Charge Pump 6.2
Miscellaneous Hardware 7.5
Total (pounds) 91.9
m m Page 29
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made by personnel with many years of transmission experience. There are no areas in the
transmission that are so critical that any increase in component size, that may be required after a
detailed design study, would precipitate any significant cost performance or weight penalty.
B. TRACTION DRIVE
1. Mechanical Operation
This subsection is a discussion of the mechanical operation of the traction drive-torque converter
transmission. Figure V-6 shows the general schematic.
Transmission input speed is proportional to engine speed. Therefore, the speed of the input toric
disk is proportional to engine speed since it is driven by a gear on the transmission input shaft.
The speed of the output toric disks relative to the speed of the input toric disk is a function of the
inclination of the traction rollers. The speed ratio across the traction drive is the same as the ratio of
the radius of rolling contact on the output toric disk to the radius of rolling contact on the input
toric disk with respect to the axis of the traction drive.
Transmission ratio changes are effected by changing the "tilt angle" of the roller axis which varies
the radius of the two points of contact with the toroids. The "tilt angle" of the rollers can be
changed by either of 2 methods:
a) Application of an external force to the roller mounting yoke and physically forcing the axis
of rotation of the roller to the required angle.
b) "Steering forces" can be generated at the point of roller — toroid contact that will cause the
roller "tilt angle" to change by translating the axis of rotation of roller relation to the toroid
center. An explanation of how this is achieved is as follows:
Figure V-7a shows a cross-section of the traction unit with the roller in the 1:1 ratio position.
Figure V-7b shows the roller in an end view of the traction unit. The velocity vector of the
roller at point of contact with the toroid is shown by vector "V". With the roller positioned as
shown with zero slip between the roller and the toroid, vector "V" also represent the velocity
vector of the toroid.
Figure V-7c shows the axis of rotation of the roller displaced an amount "X" to the left of a
parallel center line ggoing through the axis of rotation of the toroid. The velocity vectors at
the point of contact between the roller (Vp) and the toroid (Bj) bonger coincide. This
difference causes a relative slip between the two members, represented by "Vg". This slip
vector will be "down" at the point of contact between the roller and the toroid shown, and
"up" as the other point of contact because the other toroid is rotating in the opposite
direction. These two equal and opposite speed vectors produce equal and opposite forces on
the roller which, if unrestrained at the roller bearing support will cause a turning movement on
the roller, at right angles to its axis of rotation, which will "steer" the roller to some new angle
to achieve equilibrium. The angle to which the roller axis will tilt to achieve equilibrium is a
function of the distance "X".
Page 30
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FORWARD/REVERSE
GEARSET
INPUT SHAFT
INPUT TOROIDAL DISK
OUTPUT TOROIDAL DISKS
TORQUE
CONVERTER
STATOR
TORQUE CONVERTER
IMPELLER
TORQUE CONVERTER
TURBINE
THRUSTER
Figure V-6 Traction Drive - Torque Converter
Transmission Schematic
Sundstrand Aviation £
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BEARING
ROLLER "YOKE"
ROLLER
fit. V 7.
fif. V 7S
V 7c
FIGURE V-7 TRACTION ROLLER STEERING MECHANISM
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Figure V-7d shows a three dimensional free-body diagram of the roller, illustrating the two
forces "Fs", which produce the couple that "steers" the roller. "Fj" is the tangential force on
the roller at the contact points with the toroids and "Fg" the bearing reaction forces.
The proposed design utilizes the "steering force" approach. Translation of the roller axis of rotation
is accomplished by applying a force to them through the hydraulic suspension and control
cylinders. When the desired change in ratio is achieved, the steering forces will be cancelled and the
unit will operate at the new ratio until the next ratio change is requested by the control system. The
traction rollers are hydraulically interconnected in, such a way that their tangential loads, rather
than their absolute positions, must correspond. Therefore, load sharing is positively assured.
Mechanization of this approach utilized in the design is defined by an English patent by McGill.
The required torque at the traction drive creates a tangential load at the traction roller/toric disk
interface. This tangential load is sensed by the hydraulic control-suspension system, and the
hydraulic pressure thus generated is applied to the hydraulic thruster which produces the axial
clamping force across the rollers. Thus, the normal force necessary to allow a torque producing
tangential force to develop at the roller contacting points is directly propcrtional to the torque
being transmitted. The normal force then is only as large as it must be to prevent traction roller
skidding, and unit life (which is inversely proportional to the cube of the normal force) is greatly
extended. Initial pre-load is provided by a belville washer which develops sufficient initial force to
allow charge pressure build-up. This force is negated when charge pressure is applied moving the
piston out of contact with the toroid (See Appendix V-4).
The one-way clutch is provided between the output of the traction drive unit and the transmission
housing. This clutch prevents the output of the traction drive unit from rotating backwards (such as
would happen if the vehicle were allowed to roll backwards while engaged in forward drive). This
reverse rotation of the traction drive could cause the traction rollers to "steer" themselves out of
position.
The output of the traction drive is connected directly to the torque converter input member, the
impeller. The speed of the torque converter output member, the turbine, is a function of vehicle
speed and the ratio of the transmission output gears.
The function of the forward/reverse gearset at the transmission output is twofold. It serves to bring
the relatively high torque converter speed down to a more favorable transmission output speed, and
it also provides the capability for reverse vehicle operation.
Sundstrand Aviation ^h Page
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2. Hardware Description
The following is a brief description of the principal components of the traction drive torque
converter transmission. Reference may be made to the cross section drawing, 2724A-L4, shown in
Appendix V-4 for indication of component arrangement and relative size.
a) Forward-Reverse Gearing
This gearing is all helical type constant mesh and is designed to conform to the conventional
automotive type manual synchromesh gearbox.
b) Torque Converter
The torque converter is a single stage three element converter with a 3.0 to 1 stall torque ratio.
The elements are typical automotive pressed steel construction scaled down in size from a
standard automotive converter. The maximum diameter of the oil path is 6.12 inches and the
maximum speed of the input impeller is 13,800 RPM at maximum engine speed.
c) Toroids and Traction Rollers:
Both the toroids and the traction rollers are form ground from M50 or M1 steel forgings. There
are three 2.5 inch diameter rollers in each of the two toroids. The rollers rotate at a radius of
1.66 inches from the axial centerline of the toroids. At the forward side of the first toroid disk
is the variable thrust device. A Belleville type spring imposes a 1700 pound thrust preload on
the toroids and rollers. This preload is held constant as the control pressure builds up
sufficiently to overcome the constant spring force. From then on the clamping force is directly
proportional to the control pressure. Maximum control pressure is 400 psi.
The traction roller steering and suspension mechanism is all hydraulic and is based on an
existing design which ensures the accurate load sharing described earlier.
d) Seals:
Standard lip seals are used on input and output shafts and rotating seals of the cast iron piston
type. All seals are typical of those found in standard automatic transmissions.
e) Bearings:
Standard anti-friction ball and roller bearings are used throughout the transmission.
f) Control and Lube Pump:
The control and lube pump is a proven single lobe vane unit driven at input speed. Maximum
flow is approximately 6 GPM at maximum engine speed.
g) Gears:
All gears except the input mesh are constant mesh of the helical type. The input mesh are the
spur type, 26 diametral pitch, 20° pressure angle with a modified involute profile.
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h) Controls:
The control valve block is cast iron with hardened steel spool valves or sleeves where
applicable. The speed sensing governor is the rotating flyweight type acting directly on a valve
stem.
Control linkages may be of similar type as presently used with automatic transmissions.
i) Transmission Cooler:
The transmission cooler is a separate item and is noted here as a reminder. The capacity
required is equivalent to the present automotive automatic type transmission cooler.
j) Lubricating Oil:
The design of the traction transmission is based on the use of Monsanta Santotrak as the
lubricating and cooling fluid.
3. Size and Weight
The traction transmission is designed to fit within the requirements stated in paragraph 6 of the
"Prototype Vehicle Performance Specification" (see Appendix 1-2). In brief, the transmission
tunnel is not widened so as to decrease clearance between the accelerator pedal and the tunnel; the
tunnel height does not affect full fore and aft movement of the front seat; it does not violate the
ground clearance lines; it does not violate the space allocated for wheel jounce and steering
clearances; and it does not degrade the handling characteristics of the vehicle.
The input or mounting flange is not a standard to fit the conventional internal combustion engine.
However, as a reduction gearbox is required at the Rankine or Brayton cycle engine output, the
mounting flange and output shaft location may be made suitable for the proposed transmission.
The weight of the traction transmission is 77 pounds dry. A weight breakdown is shown in
Table V-2.
By far, the majority of components in an automotive transmission are sized by considerations other
than material stress such as economy of manufacture, or requirements of fitting over or around
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TABLE V-2
TRACTION DRIVE TRANSMISSION WEIGHT BREAKDOWN
Transfer Gears
Idler Gear and Gearing
Shafting and Bearings
Converter
Traction Drive
Housings
Control System and Charge Pump
Miscellaneous Hardware
Total (pounds)
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some other component. When weight is not a major consideration, components are often oversized
to "keep out of trouble," and no heed is taken or calculations made of the exact margin of safety.
4. Design Analysis
The double row traction transmission is basically designed from a stress-cycle curve and previous
experience in designing and testing traction drives. The mean transmission input power requirement
was calculated for the Federal Driving Cycle, the Simplified Suburban Route, and the Simplified
Country Route. The maximum input power requirement was calculated from the acceleration and
grade velocity requirements. These input powers, the specified life of 3500 hours, a transmission
speed range of five to one, and appropriate toroid geometry ratios with a particular traction
coefficient provide the basis for the traction transmission design.
The toroid geometry ratios involved in the design are (1) the toroid pitch diameter to roller
diameter ratio, and (2) the conformity ratio, which is defined as the ratio of roller crown radius to
roller pitch radius. The first ratio (1) defines the size of the machine and the amount of rolling to
twisting contact that the rollers experience with the toroids. Since large toroid pitch radius to roller
radius ratios approach more nearly pure rolling, the traction coefficient increases and the speed
range decreases with this ratio. Therefore, in order to accommodate a 5 to 1 speed range and stay
within package size limits, a ratio of 1.33 was chosen.
The second ratio (2) affects the shape and size of the footprint as well as the normal stresses. Higher
conformity ratios for the same load result in higher stresses. Experience dictates a circular footprint
or one that has its major axis in the direction of rolling. As a result a conformity ratio of 50% was
chosen.
The traction coefficient also affects the overall transmission size and decreases as rolling contact
velocity increases. For this design, at an input speed of 8000 RPM, the rolling contact velocity is
1400 inches per second. A traction coefficient of 0.04 is reasonably attainable for traction fluids at
this velocity.
The maximum stresses calculated for this design are 448 ksi at maximum input power of 140 HP
and 234 ksi at the mean input power. Mean input power is weighted average power over the EPA
combined driving cycle as defined by the duty cycle. Using an assumed stress cycle curve and scaling
from 1,000,000 cycles at 700 ksi for M50 tool steel, it was determined that the 3500 hour life
requirement was satisfied.
Appendix V-6 shows a schematic of the transmission with individual component speeds and
torques. The actual component sizing summary is the same as for the hydromechanical sizing shown
in Appendix V-5.
In a study of this type where basic concept and feasibility are of prime importance, it is not
appropriate to go into extensive sizing detail analysis. This is especially true when the design is being
made by personnel with many years of transmission experience. There are no areas in the
transmission that are so critical that any increase in component size, that may be required after a
detailed design study, would precipitate any significant cost, performance, or weight penalty.'
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5. Traction ratio and torque converter optimization
An optimization study was carried out to determine the required system gear ratios, traction drive
ratio range, and torque converter type and size. These parameters had to be determined within the
system requirements of:
— Maximum vehicle 'creep' speed
— Maximum vehicle operating speed
— Engine speed range
— Reasonable torque converter idle HP absorption
These studies were made using the Sundstrand vehicle performance computer program for a traction
drive transmission. Fuel consumption transmission energy efficiency, and other important criteria,
were recorded for simulated runs over the Combined Driving Cycle while varying the transmission
parameters — some, one at a time, and some in combination.
Many different torque converters were simulated and studied to gain a better understanding of the
effects of converter characteristics on vehicle performance. For example, the study showed that a
high stall torque ratio converter gave better fuel consumption than a low torque ratio converter, due
to its more favorable efficiency curve at lower converter output/input speed ratios. Another
important factor was torque converter diameter. Making the converter diameter larger makes it
"tighter;" that is, it slips less, and is therefore more efficient. However the power absorbed at engine
idle by a torque converter also increases with diameter, and must be considered.
A study was also made replacing the torque converter with a friction clutch. Total energy efficiency
increased from 74% to 80%, and fuel consumption improved from 10.0 MPG to 10.8 MPG for the
Aerojet Rankine engine, and from 15.0 to 15.7 MPG for the AiResearch Brayton engine. It should
be noted that in realizing these gains, the advantages of having a torque converter as discussed in
Section IV are lost. It would appear that these advantages outweigh the efficiency disadvantage.
However, it is not completely evident that a clutch could not be used. A more detailed study of this
would be made prior to a hardware design commitment.
Studies were also made using a torque converter lock-up clutch, and an input clutch in the system.
The result of these studies, and some of the other optimization studies, are summarized in
Table V-3. Figures are for the Aerojet Rankine engine.
The parameters chosen for the final transmission design were a torque converter ratio of 3 to 1 and
a transmission speed ratio of 5 to 1. These result in only 3 HP absorbed at idle speed and 10.01
miles per gallon fuel consumption for the Combined Driving Cycle. A trade-off study of all of the
studies and computer runs involving complexity, cost, and overall economics resulted in the choice
of the above parameters.
C. Maintainability
It is expected that either the tri-mode or the traction transmission should provide no greater
maintainability problems than present automotive automatic transmissions.
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TABLE V-3
CLUTCH/TORQUE CONVERTER PARAMETER TRADE-OFF SUMMARY
Single Parameter
Being Varied
Converter Lock-Up Clutch
Without
With: (Locks up at
0.9 Spd. Ratio)
Converter Stall Torq.
Ratio
Converter Idle HP
Absorption Rate
(an input clutch req'd
to give 0 HP loss)
Traction Drive Ratio
Range
Torq. Converter
Stall Torq.
Ratio
2.23:1
2.23:1
1.82:1
3.00:1
2.23:1
2.23:1
2.23:1
2.23:1
2.23:1
2.23:1
2.23:1
2.23:1
HP Absorbed
at Idle
7
7
7
7
7
5
3
0
7
7
7
7
Traction Drive
Ratio Range
6:1
6:1
6:1
6:1
6:1
6:1
6:1
6:1
4:1
5:1
6:1
8:1
MPG
Combined
Driving Cycle
9.57
9.65
9.34
9.70
9.57
9.88
10.04
10.39
9.49
9.55
9.57
9.58
Finally Chosen
Combination
3.0:1
3
5:1
10.01
Sundstrand Aviation
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C.) NOISE
The transmission noise whether air-borne or structure borne is an important consideration for any
automotive transmission. It is of particuliar concern because the vehicle levels required are relatively
low. The hydromechanical transmission is inherently a higher noise generation source than the
traction transmission.
1. Hydromechanical
The primary potential noise source is the hydraulic units and the secondary source is the gears.
Solution to the latter is represented by fairly well known techniques utilized and demonstrated in
millions of automotive type transmission. Such techniques will be utilized in the recommended
configuration to minimize noise. Of primary importance will be the gear tooth profile and speeds
which will be similiar to present automotive transmissions.
Considerable effort has been expended in the last few years to understand and reduce hydraulic unit
noise. The cause is fairly well known and techniques have been developed to minimize it. However
it must be recognized that because of the large number of variables involved, the only positive
assurance of meeting the required noise levels comes through actual hardware demonstrations.
The basic approach to the hydraulic noise reduction is to minimize the noise or energy level at the
source, isolate or attenuate the conduction of the noise energy to the housing, and if necessary,
attenuate the energy at the housing through isolation blankets before it can be conducted or
radiated to the air and/or surrounding surfaces.
The hydraulic units represent the major noise source. This source is primarily related to the rate of
generation of high pressure from low pressure and vice versa, the level of maximum pressure and the
porting rate rotational speed. This process is accomplished within the hydraulic unit itself — pistons,
cylinder block, and port plate. Considerable experience has been gained in the last few years in
minimizing porting noise. This is accomplished by modifying the ports between the cylinder block
and port plate to prevent large, abrupt pressure transients. Another means of minimizing the noise is
to limit the maximum working pressure within the unit. In the recommended configuration, the
working pressure is limited to 4500 psi, which would only occur with "floored accelerator" below
about 20 MPH. Hydraulic unit operational speeds can be selected to insure the best noise
characteristics. Therefore the variables involved are pressure level, rate of pressure increase or
decrease in the individual pistons during parting, porting modifications, hydraulic unit speed and to
some degree the stroke or displacement of the hydraulic units. Optimization of these parameters
without degradation of hydraulic uni* efficiency can only be accomplished through extensive
testing.
Attenuation of the generated noise to the outside of the transmission is very important. The
attenuation itself is very important but it is also important to insure that component natural
frequencies are such that no resonants occur. Minimizing resonances will simplify energy
attenuation techniques. Also noise frequencies should be kept as high as possible as attenuation is
much easier at higher frequencies.
Air-borne noise within the transmission to the main housing has been suppressed by a deep-drawn
sound shield made from a special laminated sandwich around the hydraulic unit rotating
components. The oil pan is formed from the same material to prevent the air to fluid-borne noise
from being transferred outside.
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Structural-borne noise isolation is achieved by using a similar special laminated sandwich between
the hydraulic porting plate, the intermediate support plate, and between the support plate and the
main housing. The laminated sandwich is a composite of two metal plates with 2 layers of
viscoelastic material between them seperated by a steel screen. This isolation material has a high
crush force and good attenuation above 100 Hz. This double barrier should be very effective in
minimizing noise propogation. In addition to this, a similar type of isolation is provided between
the main input and output transmission bearings and the main housing, thus eliminating any "hard
path" between the noise producing dynamic components and the main housing.
As indicated previously noise tends to be in the category of "black art". Extensive testing and
evaluation has defined design techniques which will minimizing noise. Although it is impossible to
know at this time what the final noise level will be, it is anticipated that the noise requirements will
be met.
2. Traction
The primary source of noise in the traction transmission are the gears.
The output gears are constant mesh and of the helical type similar to present automotive practice.
The input gear mesh to the input toroid is constant mesh and is shown as a spur gear. The input
toroid cannot tolerate any external thrust. To ensure lowest possible noise generation, these input
gears will be fine pitch, low pressure angle, and with a modified involute profile.
Should it become necessary to further reduce the noise from the input mesh, helical gears with
thrust runners directly between the gears to cancel the resultant thrust will be used.
In addition to reducing gear noise to a minimum at its source, laminated sound insulating bearing
sleeves are used to isolate the noise from the- main housing.
D. MAINTAINABILITY
It is expected that either the tri-mode or the traction transmission should provide no greater
maintainability problems than present automotive automatic transmissions. The only normal
maintenance required will be to check the transmission oil level as is now done. Repair or overhaul
of the transmission should not require any additional complication. The only "new to the business"
component would be the hydraulic units or the toroids and rollers. It would be expected that these
assemblies would be provided to the garage or overhaul shop as reworked assemblies similar to
present torque converter assemblies.
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(SPACER PAGE - INTENTIONALLY BLANK)
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VI. PERFORMANCE
A. Introduction
The basic performance relating to the tri-mode hydromechanical and traction/torque converter
transmission is given in this section. Transmission efficiency is shown as a function of road
load at various vehicle speeds and also as a function of variable output loads from maximum
to 10% of maximum.
It is important to recognize that although transmission efficiency is interesting the real
measure of efficiency is that of the complete vehicle system — vehicle, engine, transmission.
This efficiency is reflected as fuel consumption over the defined driving cycle.
Acceleration data is also presented in several forms. This also represents total vehicle system
performance.
B. Ground Rules and Transmission Parameter Summary
The following ground rules for all performance calculations were either specified or mutually agreed
to by the Environmental Protection Agency.
1. Test vehicle weight = 4600 pounds (Prototype Vehicle Performance Specification).
2. Gross vehicle weight = 5300 pounds (Prototype Vehicle Performance Specification).
3. Vehicle road drag and air resistance (Prototype Vehicle Performance Specification) Frontal
area = 20 sq. ft.; Coefficient of drag = .6
4. Rolling radius of wheels =1.10 feet (assumed by Sundstrand).
5. Axle efficiency = .95 (assumed by Sundstrand).
6. Total rotating inertia of tires, wheels, and brakes for all four wheels = 11.2 ft-lb-sec^ (assumed
by Sundstrand).
7. Ambient air temperature was assumed by mutual agreement with EPA to be 85°F throughout
the study. Although differences in air temperature do make a difference in air drag forces, their
inclusion is somewhat meaningless without corresponding data on variation in engine performance
with temperature, which was unavailable.
8. Accessory power requirements (Prototype Vehicle Performance Specification).
NOTE: Performance specification accessory losses representative of 6:1 engine speed range. For
narrower speed range engines, the idle accessory power requirements were assumed unchanged, but
the accessory power requirements at maximum engine speed were reduced proportionately.
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9. Engine hosed ~ Power - Fuel Consumption data supplied by EPA. (See Appendices 1-4 and
1-5.)
10. Density of fuel used in fuel consumption calculations = 6.28 Ib/gallon (Prototype Vehicle
Performance Specification).
11. The driving cycle used to calculate fuel consumption was the Combined Duty Cycle. The
Combined Duty Cycle consists of: a) the Federal Driving Cycle (see Appendix I-3); b) Simplified
Suburban Route; and c) Simplified Country Route — (Prototype Vehicle Performance
Specification).
12. Acceleration and fuel economy performance for the referenced typical 3-speed automatic
transmission !as specified by EPA) is summarized in Appendix VI-3.
PARAMETER SUMMARY -TRI-MODE HYDROMECHANICAL TRANSMISSION
Transmission Input Speed (at 100% Engine Speed) 2336 RPM
Direction of Input Rotation (looking at Pad) Clockwise
Maximum Input Torque 315ft-lb
Transmission Output Speed (89.2 MPH and 100% Engine Speed) 5114 RPM
Direction of Output Rotation (looking at Pad) Counterclockwise
Maximum Output Torque 592 ft-lb
Assumed Drive Axle Ratio 4.50:1
Maximum Vehicle Creep Speed at Engine Idle 0 MPH
Maximum Vehicle Reverse Speed 12.5 MPH
Hydraulic Unii:
Displacement 1.50 in^/rev.
Rated Speed 5172 RPM
Maximum Pressure 4500 psi
Clutch Type Multi-Plate, Flat Disk
Axial Piston, Hydraulic
Planetary Type Four Element Ravigneaux
Lubricating Fluid Type F Automatic
Transmission Fluid
Cooler Size and Flovv Typical of Existing
Automatic Transmission Coolers
Tranvrission Weight (Dry) 92 pounds
44
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PARAMETER SUMMARY - TRACTION DRIVE TRANSMISSION
Transmission Input Speed (at 100% Engine Speed) 14,000 RPM
Direction of Input Rotation (looking at Pad) Clockwise
Maximum Input Torque 52.5 ft-lb
Transmission Output Speed (89.2 MPH and 100% Engine Speed) 5411 RPM
Direction of Output Rotation (looking at Pad) Counterclockwise
Maximum Output Torque 550 ft-lb
Assumed Drive Axle Ratio 5.00:1
Maximum Vehicle Creep Speed at Engine Idle 13.0 MPH
Maximum Vehicle Reverse Speed 20.0 MPH
(Arbitrary Limit)
Torque Converters:
Diameter 6.12 in. (Rankine)
7.06 in. (Brayton)
Speed at 85 MPH 14,140 RPM
Power Absorbed at Engine Idle 3.0 HP
Stall Torque Ratio 3.00:1
Traction Drive:
Maximum Input Speed 8000 RPM
Ratio Range 5.00:1
Lubricating Fluid Sanotrack 40
Cooler Size and Flow Typical of Existing
Automatic Transmission Coolers
Transmission Weight (Dry) 77 pounds
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C. Transmission Efficiency
Transmission efficiency has been calculated for both the tri-mode hydromechanical transmission
and the traction drive transmission for both the Aerojet Rankine engine and the AiResearch
Brayton engine. Conditions of output speed and load for which transmission efficiency tabulations
and graphs have been calculated include: 1) The Federal Driving Cycle, 2) The Simplified
Suburban Route, 3) The Simplified Country Route, 4) The Combined Driving Cycle (a
combination of 1, 2, and 3), 5) Constant Vehicle Speed (cruise), and 6) Part Load (tractive effort
at 100, 75, 50, 25, and 10 percent of maximum acceleration tractive effort).
The instantaneous transmission efficiency for each point in the driving cycle was calculated.
Also, an accumulative efficiency, that is, an average efficiency, for the driving cycles was
calculated and is presented as part of this report (see Table VI-1 and Figure VI-1 through
VI-8). This average efficiency represents the quiotent of the accumulative power utilized over
the given driving cycle and the accumulative power supplied.
Two computer programs, one for systems using hydromechanical transmissions and the other for
systems using traction drive transmissions were used to simulate the vehicle, the engine, the
transmissions, and the required duty cycles to generate the efficiency data. In the two programs,
every effort was made to simulate the system's realistically. Therefore, the absolute values of
efficiency presented in this report should be representative of actual hardware. It should also be
emphasized that since the two programs were developed together, the relative efficiencies of the
systems considered are also quite meaningful.
Transmission efficiency as used in this report is defined as the total power out of the transmission
output divided by the total power into the transmission input. The primary or engine gear reduction
has been assumed by Sundstrand to be part of the engine gearbox and is therefore not reflected in
the transmission efficiency data presented here.
Calculations for the power losses contributed by gears and bearings, planetaries, open clutch
spinning, charge pumps, and torque converters are well known and accepted.
The following paragraphs describe the background used in calculating hydraulic unit and traction
unit efficiencies (or losses).
1. Hydraulic Unit Efficiency — Hydromechanical Transmission
The efficiency of each of the hydraulic units for any given working fluid viscosity is a function of
the hydraulic working pressure, speed, and, in the case of the variable unit, actual displacement (or
wobbler angle). This efficiency is markedly reduced below certain levels of working pressure and
displacement (or wobbler angle). For example, at full stroke and 3000 psi working pressure,
hydraulic pump efficiency is in the 88-94% range depending on speed, while at 500 psi and 1/4
stroke, the corresponding efficiency range is 25-45%.
The hydraulic unit design, and the predicted operating efficiencies used in this study are based on
the testing and field experience of the past 30 years. The present axial piston-hydrostatic bearing
design has evolved from past experience with many hydraulic unit configurations including radial
piston units and anti-friction thrust bearino units, and has proved to be the best design in terms of
cost, size, efficiency, and reliability.
Page 46
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TABLE VI-1
COMBINED DRIVING CYCLE TRANSMISSION EFFICIENCY
Cycle
Federal Driving Cycle
Simplified Suburban Route
Simplified Country Route
Combined Driving Cycle
Average Transmission Efficiency Over
The specified Driving Cycles
Rankine Engine
HMT
78%
72%
90%
81%
TDT
67%
68%
85%
74%
Brayton Engine
HMT
81%
80%
87%
83%
TDT
71%
71%
86%
76%
HMT — Hydromechanical Transmission (Tri-Mode)
TDT — Traction Drive Transmission (with Torque Converter)
Vehicle Weight - 4600 Ib. (Test Vehicle Weight)
Accessory Power — Air Conditioner On. Total Vehicle Accessory Power; 4 HP at
Engine Idle, Linear to 4.83 HP (Rankine Engine) or 5.65 HP
(Brayton Engine) at maximum Engine Speed. See Appendix I.
Atmospheric Conditions - 85°F, 14.7 PSIA
Page 47
Sundstrand Aviation
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•8
&
3
a
-
•3
a
'>
5'
z
z
<
oc
100
40
TRI-MODE HYDROMECHANICAL TRANSMISSION
AEROJET RANKINE ENGINE
VEHICLE MEIQHT
40OO LB. (TEST VCHICLE WEIQNT)
ATMOSPHERIC CONDITION!
•B°F. 14 7 PSIA
10
20
30 40 50
VEHICLE SPEED (MPH)
60
70
BO
90
Figure VI-1
Tri-Mode Hydromechanical Transmission Efficiency at Constant
Speed - Rankine Engine
-------
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c
3
Q.
V)
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^f
o'
. 3
100
90.
80
70
2
UJ
4
o:
60
Jt 5(H
UJ
° 40
C/l
30
20
10-
TRACTION DRIVE/TORQUE CONVERTER TRANSMISSION
AEROJET RANKINE ENGINE
VJEHtCLE WEIGHT
44OO L8 1TCST VEHICLE WflGMTI
A1MOSPHEBIC CONDITIONS
BVf 14 7 PSIA
10
20
30 40 50
VEHCILE SPEED (MPHI
60
70
80
90
Figure VI-2 Traction Drive Transmission Efficiency at Constant Speed - Rankine Engine
-------
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3
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:
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16
10O
go-
_ 70.
o
ui
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40.
00
2 30J
2O
1O
TRI MODE HYDROMECHANICAL TRANSMISSION
AIRESEARCH 8RAYTON ENGINE
I
I
ViHictf WflOHT
•MO I*. Htlt WtMICLI WEIOHT)
ATMOSPMf NIC CONDITIONS
•ft°f. 14 7 Ml*
30
40 50
VEHICLE SPEED (MPH)
60
70
80
90
Figure VI-3
Tri-Mode Hydromechanical Transmission Efficiency at Constant Speed
Brayton Engine
-------
z
to
z
30-
20
10-
10
20
VEHICLE WEIGHT:
MM LB (TOT VIMICLl WtlQMTI
ATMOSTHEPtlC CONDITION*
§6°f. 14 7 '91*
TRACTION DRIVE/TORQUt COHV€RTER TRANSMISSION
AIRESEARCH BRAYTON ENGINt
30 40 50
VEHICLE SPEED (MPH)
60
70
80
90
Figure VI-4 Traction Drive Transmission Efficiency at Constant Speed - Brayton Engine
-------
100
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Q.
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Q.
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<
5'
•— t-
5'
3
20
10-
0,
0
TRI MODE HYDROMECHANICAL TRANSMISSION
AEROJET RANKINE ENGINE
VEHICLE WEIGHT:
«600 LB (TtST VEHICLE WEIGHT!
ATMOSPHERIC CONDITION*
BB°F \t J PSIA
10
20
30 40 50
VEHICLE SPEED(MPH)
60
70
25
PERCENT OF MAX
ACCELERATION
LOAD
.10
80
"1
90
Figure VI-5
Hydromechanical Transmission Eff5
Rankine Engine
mcy at Full and Part Loads -
-------
a
0)
- *+•
:O
*
at
(A)
30 40 50
VEHICLE SPEED (MPH)
PERCENT OF MAX.
ACCELERATION
LOAD
TRACTION DRIVE/TORQUE CONVERTER TRANSMISSION
AEROJET RANKINE ENGINE
Vf HICLt WilGHT
MOO i-8 ITtST VEHICLf Wf IOMTI
OSPHEHIC CONOITIONt
e»°F. 14 irSIA
Figure VI-6 Traction Drive Transmission Efficiency at Full and Part Loads -
Rankine Engine
-------
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0.
Cfl
u
UJ
o
z
o
I/)
CO
CO
50
40
30
20
10
0-
10
TRI-MODE HYDROMECHANICAL TRANSMISSION
AIRESEARCH 8RAYTON ENGINE
VEHICLE WEIGHT
480O LB (TEST VEHICLE WEIGHT)
ATMOSPHERIC CONDITIONS
65° F 14 7 f SI A
20
30 40 50
VEHICLE SPEED (MPH)
60
70
PERCENT OF MAX.
ACCELERATION
LOAD
10
80
90
Figure VI-7
Hydromechanical Transmission Efficiency at Pull and Part Load -
Brayton Engine
-------
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V)
•-+
E3
. 3
f*.
|o
01
01
o
QJ
o
UJ
Z
CO
Z
40 50
VEHICLE SPEED (MPH)
PERCENT OF MAX.
ACCELERATION
LOAD
TRACTION DRIVE/TORQUE QONVERTER TRANSMISSION
AIRESEARCH BRAYTON ENGINE
VEHICLC WEIGHT
4600 LB (TEST VEHICLE WEIGHT)
ATMOSPHERIC CONDITIONS
•B°F. 14 7 PSIA
Figure VI-8 Traction Drive Transmission Efficiency at Pull and Part Load -
Brayton Engine
-------
2. Traction Unit Efficiency — Traction Drive Transmission
There is always some degree of slip or creep at the roller/toroid interface. Since the contact area
between the roller and the toroid has a finite area, the drive rollers tend to spin with respect to their
toroidal races, so the motion of the rollers is rolling/spinning rather than pure rolling. Consequently,
there is a power loss due to roller spin. Also rolling resistance is encountered between the toroids
and the rollers, as well as in the rolling element bearings.
The overall efficiency of the traction unit is the product of the speed efficiency and the torque
efficiency. The speed efficiency is a measure of slip. The torque efficiency, in general, is a measure
of spin loss, rolling resistance, and windage.
The efficiency of the traction unit was calculated using experience gained from the development
and testing of a Sundstrand toroidal type variable input speed constant output speed traction drive
for aircraft applications. The Sundstrand efficiency data correlates well with data published by
General Motors and Tracor on the efficiency of rolling contacts.
D. Grade and Acceleration Performance
Grade and acceleration performance was calculated for both transmissions and engines. The grade
performance is a function of engine power and transmission efficiency. Acceleration performance is
a function of engine power, transmission characteristics, drive line efficiency, tire adhesion, the
engine time lag in going from idle to the maximum power condition, and the ratio of engine power
going into accelerating the engine, to that which is accelerating the vehicle during the time lag
period. This time lag itself is a function of the shape of the engine speed— torque curve, and the
combined engine— transmission inertia. Because of the many variables involved several assumptions
were made:
1) Maximum acceleration can be achieved by allowing the engine to accelerate to the
maximum power condition unloaded and then applying maximum power to the wheels. For
the required 0-60 mph acceleration time and the distance traveled in 10 seconds, an engine
acceleration time of 0.5 seconds for the Rankine cycle engine and 1.0 seconds for the Brayton
cycle engine was assumed based on discussion with the engine suppliers. For the 25-70 mph
and 50-80 mph acceleration times, an engine acceleration time of 0.25 seconds for the Rankine
cycle engine and 0.7 seconds for the Brayton engine was assumed. In practice, these time lags
would probably be unacceptable from the "driver feel" point of view, and to overcome this,
the engine power during this engine acceleration period would be split, some going to
accelerate the engine, and some to accelerate the vehicle. The exact ratio of this power split
would depend very much on "driver feel" and would be determined experimentally.
Regardless of the split, it has been assumed that the 0-60 mph and 0-10 sec. acceleration
performance would not be significantly different.
2) During the maximum acceleration from start conditions, the assumed vehicle weight and
weight distribution shift combined with a reasonable tractive coefficient will allow a maximum
tractive effort of 2500 Ib. at the wheels without wheel slippage.
3) The reflected inertia of the transmission to the engine is very small and for the purpose of
this study is ignored. The actual inertia is not only small but is reduced by the square of the
gear ratio between the two. For example, this factor is 1/1295 for the hydromechanical
transmission and 1/314 for the tractron drive transmission, when mated with the Brayton
Cycle engine.
Sundstrand Aviation •
-------
Calculation of the acceleration from start requirements when utilizing the Airesearch Brayton
cycle engine indicated that the maximum power as defined in Appendix 1-5 was not sufficient.
It was therefore necessary to assume a higher power to drive the vehicle and accessories — 145
HP for the hydromechanical transmission and 155 HP for the traction drive transmission. The
power available from the Aerojet Rankine cycle engine, 148 HP at zero vehicle speed
increasing to 163 HP at 85 mph, appears to be adequate to meet all performance requirement
limits.
No problem was encountered in meeting the gradeability requirements. The maximum
achievable vehicle speed along with the corresponding engine power requirements are tabulated
in Table VI-2.
Table VI-2 also lists the actual acceleration performance of the various systems, taking into account
engine lag. Also tabulated are the performance requirement limits.
Plots of vehicle speed and distance as a function of time during a maximum acceleration run are
presented in Figure VI-9, VI-10, VI-11 and VI-12. The plots are based on a start from maximum
power condition as can be achieved by locking the brakes.
Page 57
^^.
Sundstrand Aviation
-------
Table VI-2 Idle Acceleration and Grade Performance
Performance Requirements
Idle
Creep Speed
Accel.
Time to 60 MPH
Dist. to 10 sec.
Time 25* 70 MPH
Time 50* 80 MPH
Dist. 50 * 80 MPH
Grade Velocity
30%
5%
0%
18 MPH
(max)
13.5 sec
(max)
440ft.
(min)
15 sec
(max)
15 sec
(max)
1400ft.
(max)
5 MPH
(min)
65 MPH
(min)
85 MPH
(min)
Weight
1
1
1
1
1
1
2
2
1
Rankine Engine
HMT
0
11.1
490
12.3
11.7
1125
29
85
85
(91 HP)
TOT
13
13.2
445
14.4
12.3
1175
20
84
85
(86 HP)
Brayton Engine
HMT
0
12.6
440
15.0
12.9
1235
26
80
85
(93 HP)
TDT
13
12.6
440
14.6
12.7
1230
23
81
85
(89 HP)
HMT — Hydromechanical Transmission (Tri-Mode)
TDT — Traction Drive Transmission (with torque converter)
1 at 4600 Ib (test vehicle weight)
2 at 5300 Ib (gross vehicle weight)
Accessory Power — Air Conditioning On. Total Vehicle Accessory Power; 4.0 HP at Engine
Idle, Linear to 4.83 HP (Rankine Engine) or 5.65 HP (Brayton Engine)
at Max Eng. SpeeiJ. See Appendix I
Atmospheric Conditions - 85°F, 14.7 PSIA
Engine Power: In accordance with Figure 1-7 except as noted.
Page 58
Sundstrand Aviation
-------
3
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2500T
2250|
20004
17504.
t i 1500
UJ
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1250+
1000+
750+
5004-
2504-
VEHICLE WEIGHT-.
46OO LB. (TEST VEHICLE WEIGHT)
ACCESSORY POWER:
AIR CONDITIONER ON: TOTAL VEHICLE
ACCESSORY POWER: 4.00 HP AT ENGINE IDLE,
LINEAR TO 4.B3 HP (RANKINE ENGINE) OR
5.65 HP IBRAYTON ENGINE) AT MAXIMUM
ENGINE SPEED. SEE APPENDIX I.
ATMOSPHERIC CONDITIONS:
85°F. 14.7 PSIA
Figure VI-9
15 20
TIME (SECONDS)
Hydromechanical Transmission Acceleration
Rankine Engine
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VEHICLE WEIGHT:
4600 LB. (TEST VEHICLE WEIGHT)
ACCESSORY POWER:
AIR CONDITIONER ON; TOTAL VEHICLE
ACCESSORY POWER; 4.00 HP AT ENGINE IDLE,
LINEAR TO 4.83 HP (RANKINE ENGINE) OR
5.60 HP (ORAYTON ENGINE) AT MAXIMUM
ENGINE SPEED. SEE APPENDIX I.
ATMOSPHERIC CONDITIONS:
8S°F, 14.7 PSIA
10 15
20
25
30
TIME (SECONDS)
Figure VI-11
Hydromechanical Transmission Acceleration
Brayton Engine
-------
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E. Fuel Consumption at Constant Speed and Full and Part Loads
Constant speed fuel consumption and fuel consumption at full and part load in miles per gallon
were calculated for systems using the tri-mode hydromechanical transmission and the traction drive
transmission for both the Aerojet Rankine engine and the AiResearch Brayton engine.
Constant speed fuel consumption is defined as fuel consumption at zero vehicle acceleration.
Plots of constant speed fuel consumption based upon the simulated vehicle with its various
configurations of transmissions and engines are presented in this section in Figures VI-13, VI-14,
VI-15 and VI-16. Also presented here are plots of instantaneous fuel consumption in miles per
gallon vs. vehicle speed at maximum tractive effort, as well as at 75%, 50%, 25%, and 10% of
maximum tractive effort. This full and part load fuel consumption data is presented in Figures
VI-17, VI-18, VI-19 and VI-20.
F. Fuel Consumption Summary
Average fuel consumption in terms of miles per gallon and BTU/mile has been calculated for both
the tri-mode hydromechanical and the traction drive-torque converter transmission for both the
Aerojet Rankine engine systems and the AiResearch Brayton engine systems.
A summary of the average fuel consumption over the Federal Driving Cycle both with and without
the air conditioner operating is given in Table VI-3.
A breakdown of the fuel consumption over the Combined Driving Cycle is presented in Table VI-4.
The Combined Driving Cycle consists of the Federal Driving Cycle, the Simplified Suburban Route,
and the Simplified Country Route. Also presented in Table VI-4 is the fuel consumption that could
be expected from an ideal transmission. The ideal transmission is infinitely variable, 100% efficient,
and has no spin loss to load the engine at idle.
A fuel consumption breakdown of the Combined Driving Cycle in terms of BTU/mile is presented
in Table VI-5.
Vehicle range, which is a function of fuel consumption, was calculated and is presented in Table
VI-6. It was assumed that there was 25 gallons of fuel available initially. Vehicle range has been
calculated for the Federal Driving Cycle at a constant 70 mph cruise, both with and without the air
conditioner operating.
G. Tractive Effort Limits
Tractive effort limits for the simulated vehicle are established by a number of parameters. Among
these are road adhesion of the tires, vehicle weight and configuration, maximum available
transmission input power, maximum available transmission output torque, and transmission
efficiency.
Page 63
Sundstrand Aviation
-------
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Af-^OJET RANKINE ENGINE
HICLE WC K>NT
4600 LB (T£ST VtMlClt WkiCHT)
ATMOSPMtfltC CONDITIONS
8&°F 14 ? PSIA
10
20
30 40 50
VEHICLE SPEED - MPH
60
70
80
90
Figure VI-13 Hydromechanical Transmission Fuel Consumption at Constant Speed
Rankine Engine
-------
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AEROJET RANKINE ENGINE
£2*o/r/OA,f«
VEHICLt WfclGMT
««UO LB (TEST VIMICLt WilOHTI
ATMOSfHlBIC CONDITIONS
ai°f 14 7 PSIA
10
20
30 40 50
VEHICLE SPEED (MPH)
60
70
ao
90
Figure VI-14 Traction Drive Transmission Fuel Consumption at Constant Speed
Rankine Engine
-------
3"
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25
20
- TRI MODE HM. TRANSMISSION
- AlRESEARCH BRAYTON ENGINE
VEHICLE WEIQMT
40OO LI. (TEST VEHICLE WEIOHTI
30 40 SO
VEHICLE SPEED (MPH)
Figure VI-15 Hydromechanical Transmission Fuel Consumption at Constant Speed
Brayton Engine
-------
25
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TRACTION DRIVE/TORQUE CONVERTER TRANSMISSION
AIRESEARCH BRAYTON ENGINE
VtHICLE WtlGNT
46OOLB (TI»T VIMlCLl WClOHTl
ATMOSPHCAIC COMOITIONS
B5°F I47PSI*
10
20
30 40 50
VEHICLE SPEED IMPH)
60
70
80
90
Figure VI-16 Traction Drive Transmission Fuel Consumption at Constant Speed
Brayton Engine
-------
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20
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5
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- TRI MODE HM. TRANSMISSION
- AEROJET RANKINE ENGINE (PROTOTYPE)
VEHICLE WEIGHT
40OOLO (TEST VEHICLE WEIGHT)
ACCESSORY POWER
AIR CONDITIONER ON. TOTAL VEHICLE
ACCESSORY POWER. 4 OO HP AT ENGINE IDLE.
LINEAR TO 483 HP IRANKINE ENGINE) OR
565 HP IBRAYTON ENGINE) AT MAXIMUM"
ENGINE SPEED SEE APPENDIX I
ATMOSPHERIC CONDITIONS
85°> 14 I PSIA
PERCENT OF MAX.
ACCELERATION
LOAD
10-1
10
20
30 40 50
VEHICLE SPEED (MPH)
RUN ) 01
Figure VI-17 Hydromechanical Transmission Fuel Consumption at Full and Part Load
Rankine Engine
-------
25
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TRACTION DRIVE/TORQUE CONVERTER TRANSMISSION
AEROJET RANKINE ENGINE
VEHICLE WEIGHT
4600 LB (TEST VEHICLE WEIGHT)
ACCESSORY POWER
AIR CONOITIONEH ON TOTAL VEHICLE
ACCESSORY POWER. 4 00 HP AT ENGINE IDLE.
LINEAR TO 483 HP IRANKINE ENGINE) OR
5 h5 HP [BRAYTON ENGINE! AT MAXIMUM
INGINE SPIED SEC APPENDIX I
ATMOiPMI HIC CONDITIONS
»'juF 14 ; PbIA
PERCENT OF MAX.
ACCELERATION
LOAD
30 40 50
VEHICLE SPEED (MPHI
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Figure VT-18
Traction Drive Transmission Fuel Consumption at Full and Part Load -
Rankine Engine
-------
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VEHICLC Wf IOHT
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ACCfSSOXV POWEH
AIR CONOITIONtH ON: TOTAL VIMICLI
ACCESSORY rOWCR. 40O Hf AT ENOINC IDLE
LINEAR TO 4(3 HP IRANKINE ENGINE) OR
b 66 HP (BflAVTON ENGINE) AT MAXIMUM
CNGlNi SPtEO SEE APPENDIX I
TRI MODE HM. TRANSMISSION
AIRESEARCH BRAYTON ENGINE
PERCENT OF MAX
ACCELERATION
LOAD
10
30 40 50
VEHICLE SPEED (MPH)
Figure VT-19 Hydromechanical Transmission Fuel Consumption at Full and Part Load
Brayton Engine
-------
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3
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PERCENT OF MAX.
ACCELERATION
LOAD
O
a.
2
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a.
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to
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O
O
TRACTION DRIVE/TORQUE CONVERTER
TRANSMISSION
AIRESEARCH BRAYTON ENGINE
VEHICLE Wl IGMT
46OO LB ITtST VEHICLE WEIGHT!
ACCESSORY POWt H
AIR CONDITIONER ON
TOTAL VEHICLE
ACCESSORY POWER. 4 00 HP AT ENGINE
t INF AH TO 483 HP IRANKINE ENGINE) OR
560 HP IBHAYTON ENGINE! AT MAXIMUM-
tNGINC SPEIO SEE APPENDIX S
ATMOSPHERIC CONDITIONS
Bb0f 14 7 PSIA
40 50
VEHICLE SPEED (MPH)
Figure VI-20
Traction Drive Transmission Fuel Consumption at Full and Part Load
Brayton Engine
-------
TABLE VI-3
FEDERAL DRIVING CYCLE FUEL CONSUMPTION WITH AND WITHOUT AIR CONDITIONING
Accessory Power
Air Conditioner On
Air Conditioner Off
Average Fuel Consumption Over The
Federal Driving Cycle With and Without
Air Conditioner On
Rankine Engine
HMT
8.33
9.43
TDT
7.65
8.59
Brayton Enqine
HMT
12.91
14.29
TDT
12.03
13.31
HMT — Hydromechanical Transmission (Tri-Mode)
TDT — Traction Drive Transmission (with Torque Converter)
Vehicle Weight - 4600 Ib (Test Vehicle Weight)
Accessory Power — Air Conditioner On: Total Vehicle Accessory Power; 4.00 HP at
Engine Idle, Linear to 4.83 HP (Rankine Engine) or 5.65 HP
I (Brayton Engine) at maximum engine speed.
Air Conditioner Off: Total Vehicle Accessory Power; 2.00 HP at
Engine Idle, Linear to 2.23 HP (Rankine Engine)or 2.45 HP
(Brayton Engine) at maximum engine speed. See Appendix I.
Atmospheric Conditions - 85°F, 14.7 PSIA
Page 72
Sundstrand Aviation £™£
'• 5. :»!'s-s Cvo
-------
TABLE VI-4
COMBINED DRIVING CYCLE FUEL CONSUMPTION
Cycle
Federal
Driving Cycle
Simplified
Suburban Route
Simplified
Country Route
Combined
Driving Cycle
Fuel Consumption In Miles Per Gallon Over the Individual And
Combined Driving Cycles
Rankine Engine
HMT
8.33
12.23
12.30
10.62
TDT
7.65
11.77
11.78
10.01
Ideal
10.19
14.68
13.30
12.43
Brayton Engine
HMT
12.91
18.33
16.91
15.71
TDT
12.03
17.30
16.62
14.95
Ideal
15.08
19.84
19.28
17.79
HMT — Hydromechanical Transmission (Tri-Mode)
TDT — Traction Drive Transmission (with Torque Converter)
IDEAL - Hypothetical 100% Efficient Transmission
Vehicle Weight - 4600 Ib. (Test Vehicle Weight)
Accessory Power - Air Conditioner On. Total Vehicle Accessory Power; 4.00 HP at Engine
Idle, Linear to 4.83 HP (Rankine Engine) or 5.65 HP (Brayton Engine)
at maximum engine speed. See Appendix I.
Atmospheric Conditions - 85°F, 14.7 PSIA
Page 73
Sundstrand Aviation
-------
TABLE VI-5
COMBINED DRIVING CYCLE ENERGY CONSUMPTION
Cycle
Federal Driving Cycle
Simplified Suburban Route
Simplified Country Route
Combined Driving Cycle
Average Energy Consumption In BTU's
Per Mile Over The Individual And Combined
Driving Cycles
Rankine Engine
HMT
13947.
9500.
9446.
10940.
TDT
15187.
9871.
9862.
11606.
Brayton Engine
HMT
8999.
6338.
6870.
7395.
TDT
9658.
6716.
6990.
7771.
HMT — Hydromechanical Transmission (Tri-Mode)
TDT — Traction Drive Transmission (with Torque Converter)
Fuel Heating Value - 18500 BTU/LB
Fuel Density - 6.28 LB/GAL.
BTU
Ml
BTU
L8
LB '
GAL
Ml
GAL
Vehicle Weight - 4600 LB (Test Vehicle Weight)
Accessory Power — Air Conditioner On. Total Vehicle Accessory Power; 4.00 HP at
Engine Idle, Linear to 4.83 HP (Rankine Engine) or 5.65 HP
(Brayton Engine) at maximum engine speed. See Appendix I.
Atmospheric Conditions - 85°F, 14.7 PSIA
Page 74
Sundstrand Aviation
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TABLE VI-6
VEHICLE RANGE AT FEDERAL DRIVING CYCLE AND AT CRUISE
Cycle
Federal
Driving
Cycle
Federal
Driving
Cycle
70MPH
Cruise
70MPH
Cruise
Accessory
Power
Air Conditioner
On
Air Conditioner
Off
Air Conditioner
On
Air Conditioner
Off
Vehicle Range In Miles with and without
Air Conditioning (25 gallons of fuel)
Rankine Engine
HMT
208
236
285
324
TDT
191
215
280
291
Brayton Engine
HMT
323
357
376
393
TDT
301
333
378
398
HMT — Hydromechanical Transmission (Tri-Mode)
TDT — Traction Drive Transmission (with Torque Converter)
Fuel Tank Capacity — 25 gallons
The EPA "Prototype Vehicle Performance Specification" (Appendix I) Requires 200
Miles Minimum Range
Miles = MPG x Gallons
Vehicle Weight - 4600 Ib (Test Vehicle Weight)
Accessory Power - Air Conditioner On: Total Vehicle Accessory Power; 4.00 HP at
Engine Idle, Linear to 4.83 HP (Rankine Engine) or 5.65 HP
(Brayton Engine) at Maximum Engine Speed
Air Conditioner Off: Total Vehicle Accessory Power; 2.00 HP at
Engine Idle, Linear to 2.23 HP (Rankine Engine or 2.45 HP
(Brayton Engine) at Maximum Engine speed. See Appendix I
Atmospheric Conditions - 85°F, 14.7 PSIA
Page 75
Sundstrand Aviation
-------
It was assumed that the absolute maximum tractive effort that the wheels could produce without
slipping was 2500 pounds.
In the case of the systems with hydromechanical transmissions, a tractive effort of about 2400
pounds corresponds to a hydraulic system pressure of 4500 psi. For life, noise, and reliability
reasons, it is recommended that system pressure be limited to 4500 psi at maximum tractive effort.
Systems with traction drive transmissions are limited by road adhesion to 2500 pounds, with the
assumed 155 HP Brayton engine, or to 2450 pounds at start-up by engine power limitations with
the Rankine engine.
At higher vehicle speeds, the tractive effort available from either transmission type is limited by
engine power limitations and transmission efficiency. See graphs, Figures VI-21 and VI-22.
Page 76
Sundstrand Aviation
V W ,
-------
2500
2000
- 1500
CD
QC
O
1000
>
o
oc
500
VEHICLI WEIGHT
4000 LB. (TCST VEHICLE WEIGHT)
ACCESSORY POWf R
AIM CONDmONI R ON. TOTAL
ACCESSORY HOWIH 4 00 HP AT f NC
LINI AM TO 4NJ HP lHANKINE CNCINO Oft
6 0!» HP IHHAYTON [ NGINI I AT MAXIMUM
INGINl SI'l ID OCLAPP(NOIX)
ATMOSHMl niC CONDITIONS
8!)°t I47PJIA
10
20
30
40 50
VEHICLE SPEED (MPH)
Figure VI-21 Maximum Tractive Effort - Hydromechanical Transmission
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2500 T-
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VEHICLE WEIGHT:
4600 LB. (TEST Vf HICLE WElOHTI
ACCESSORY POWER.
AIM CONOITIONEH ON. TOTAL VEHICLE
ACCESSORY POWER. 4.0O HP AT ENOINE IDLE.
LINEAR TO »«3 HP IRANKINE ENGINE) OR
ttt HP IBRAYTON ENGINE) AT MAXIMUM
ENGINE SPEED SEE APPENDIX I
ATMOSPHERIC CONDITIONS
86°F 14 7 PSIA
40 50
VEHICLE SPEED (MPH)
Figure VI-22 Maximum Tractive Effort - Traction Drive Transmission
-------
VII. CONTROL SYSTEM ANALYSIS
The two engines under consideration, the Brayton and Rankine cycle, have similar ratings and
narrow speed ranges from idle to maximum speed. The tri-mode and traction transmission are both
infinitely variable ratio which provides the capability of operating the engine at speeds completely
independent of vehicle road speed. For these reasons, the control system for either transmission
when combined with either engine will be basically the same.
A. Control System Approach
Two fundamental methods of control available for use with infinitely variable ratio transmissions
are speed control or torque control. With speed control the accelerator pedal position is directly
related to transmission output speed and the output torque will attempt to reach the value which is
proportional to the difference between the output speed being called for and the actual output
speed. If this difference is great, the torque will be high and shock loading will occur with the
possible stalling of the engine and excessive wear or ultimate failure of driveline parts. The prime
advantage of speed control is that for a particular pedal position, the output speed will remain
constant regardless of torque required up to the engine capability. If this feature is desirable or
necessary, additional control devices must be incorporated to preventengine stall or shock loading.
With torque control, the accelerator pedal position is directly related to transmission output torque
and the output speed will change until the output torque equals the torque being called for. This
control scheme prevents shock loading and engine stall unless the engine is at a speed where it
cannot produce the torque called for. This type of control is very similar to that now used in the
standard passenger car with a 3-speed automatic transmission. A governor within the transmission
which controls the shift point prevents the engine from stalling; the shift to a higher gear does not
occur until the engine is up to a speed where it can produce the required torque.
Protection from excessive torque (or pressure in the hydromechanical transmission) is provided
inherently in the design of either type of transmission. For the traction drive, the torque converter
is the "relief valve" and for the hydromechanical the control system prevents the pressure from
exceeding 4500 psi by a pressure feedback which causes the variable wobbler to destroke at that
pressure.
The control system must also consider and provide for lowest fuel consumption, emissions, and
noise and for maximum acceleration and engine (dynamic) braking.
The block diagram of the control system selected to optimize the foregoing requirements is shown
on Figure VII-1. The diagram is similar for both transmission and both engine combinations with
only the addition of the shift device for the tri-mode transmission. The system is basically torque
control with the torque at the output directly proportional to the output of the governor valve. The
governor valve output is a function of accelerator pedal position and engine speed. These are
combined in such a manner that each pedal position calls for a particular horsepower and for an
engine speed which is optimum (lowest SFC) for that horsepower.
The speed sensor not only prevents the engine from being overloaded at any time but also allows
the engine to accelerate to the desired speed and horsepower so that maximum vthicle acceleration
can be obtained.
Page 79
Sundstrand Aviation £,-£.
-------
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ACCELERATOR _
INPUT '
BRAKE
INPUT
ENGINE
SPEED
SENSOR
-. fc^
»-
k-
VARIABLE RATIO
MECHANISM
t ! t
r- -i U n
r- -»- -|
SHIFT DEVICE
(TRI-MODEONLY)
- _,
1
GOVERNOR
VALVE
OUTPUT
Figure VII-1 Control System Block Diagram
-------
Dynamic engine braking is possible with either transmission type although the hydromechanical will
transmit power in the reverse direction more efficiently than the torque converter traction type.
With the accelerator pedal released, the brake input signal will place the control system in a state
where vehicle inertia will be capable of driving the engine up to its maximum speed for dynamic
engine braking.
Either transmission will have a manual control lever very similar to the present automotive control
for park, reverse, neutral, and drive (forward) operation. The tri-mode hydromechanical
transmission has two automatic mode changes (shifts) in forward which do not affect system
performance or stability analysis.
B. Description of Operation
The engine will be started with the control lever only in park or neutral position. In either position
the control system is designed so that the existence of any torque at the output will cause the
transmission ratio to change in the direction to reduce that torque. This is accomplished by
porting working pressure directly to the hydraulic unit stroking pistons.
When the operator depresses the accelerator pedal, two functions are accomplished. The engine fuel
control produces an engine torque which is a function of engine speed. Also, the transmission
governor valve controls the transmission output torque until the engine reaches the speed for that
particular accelerator pedal position, and maintains that speed. The combination of controlling
engine torque and speed provides a constant horsepower out of the engine which will accelerate the
vehicle until the road load equals this horsepower. A change in accelerator pedal position or load
will also change the transmission ratio and engine speed.
For the hydromechanical transmission, the controls automatically change to the second and third
mode when the transmission reaches the shift ratio. The transition from one mode to another is
completed very smoothly as the load is shared by both hydraulic and mechanical load paths.
When the driver removes his foot from the accelerator pedal, the engine will run at the minimum
speed compatible with the output speed and the vehicle will slow down until the engine reaches idle
speed. If the driver desires to reduce speed more rapidly, he will apply pressure to the brake pedal.
This will first bias the governor valve so that the vehicle will drive the engine to its maximum rated
speed for dynamic braking. Further pressure on the brake pedal will actuate the vehicle service
brakes.
C. Stability Analysis
A simulation of the control system was run on a hybrid computer. The parameters, equations,
and engine performance used are included in Appendix VII of this report. Also included are
representative traces of the computer readout for a 20% of maximum throttle (0:2 power
unit) acceleration at 50% load and no load.
The simulation was of the hydromechanical transmission in the start-up or hydrostatic mode. The
results indicate that the response is good and that the system is stable even when subjected to a
maximum transient. Experience dictates that similar results will exist in either hydromechanical
mode and also with the traction transmission.
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D. Safety Analysis
A failure and effect analysis has been carried out for each component associated with the control
system. The results of this study follow:
1. Accelerator Input Failure
If the input to both engine and transmission is lost, there will be no response when the accelerator is
depressed.
If only the input to the engine is lost, the engine will not be able to support any load and will not
accelerate away from idle.
If only the input to the transmission is lost, the transmission will try to keep the engine at idle by
assuming a minimum transmission rate and, therefore, the output power will be quite low.
2. Brake Input Failure
If the input linkage from the brake pedal fails, the system will perform normally except there will
be no added assist from the engine while braking. The engine will run at the minimum speed
possible for the vehicle speed. Normal friction brakes will not be affected.
3. Speed Sensor Failure
If the signal is lost to the speed sensor, or if the speed sensor sticks in the underspeed position, the
transmission will not be able to load the engine and when the accelerator is depressed, the engine
will accelerate as it does when in neutral.
If the speed sensor is stuck in the overspeed position, the transmission will load, stall the engine, as
the accelerator is depressed.
4. Governor Valve Failure
If the governor valve sticks in the high speed position, the engine will be loaded down. At high
speeds, the transmission will go to minimum ratio (minimum engine speed) and exhibit low power
output. At low speeds, or when standing still, the engine speed will be driven down until it stalls.
If the governor valve sticks in the low speed position, the engine speed will be driven up. If this
happens at high vehicle speeds, the engine speed could be driven beyond a safe speed. This is the
only potentially dangerous failure in the system. It is unlikely that this will happen while the vehicle
is running, and if it does, the operator can correct the overspeed by shifting into neutral. If this
failure occurs while shutdown, the vehicle will appear to remain in neutral after the engine is
started.
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5. Shift Device Failure (Tri-Mode Only)
If the shift device sticks in the hydrostatic mode, everything will be normal except the vehicle will
not be able to accelerate to any speed above the shift point.
If the shift device sticks in either other position, the transmission will remain in that range and the
engine will stall as the vehicle slows down or will not start if this occurs when stopped.
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VIII. ESTIMATED TOTAL MANUFACTURING COSTS
A. Definition of the Cost Analysis
The EPA Contract Specification paragraph 6.5 requires an original equipment manufacturer (OEM)
cost estimate for the transmission in quantities of 100,000 and 1,000,000 per annum and a cost
comparison made with a "conventional" (unspecified) multi-speed automatic transmission with
torque converter.
The figures shown in the following cost analysis are for the "total manufacturing cost" which can
be broadly defined as the cost of labor and materials, along with the operation and maintenance of
existing plant and tooling.
The price includes — cost of materials and purchased subcomponents, direct and indirect labor
(such as administration, supervision, production control, quality control, plant maintenance,
production engineering, etc.), and supplies and utilities for plant operation. Tooling and plant
amortization, and taxes for existing plant and equipment are also included. This price does not
include engineering and development, advertising, sales, distribution, interest or profit.
B. Costing Procedure
Although Sundstrand is not a supplier of transmissions to the automobile industry, large quantities
of transmissions for the trucking, farm equipment, construction and garden equipment industry
are produced. Personnel with cost estimating experience in the automotive automatic
transmission industry are available also.
This experience, coupled with cost data available from related product lines, is the basis for the
estimate of production rates of 1,000,000 per annum. Additionally, a "judgment factor" was
applied to arrive at figures for 100,000 per annum production rates. This "judgment factor"
acounted for the degree of complexity, type of processing, and the degree of process simplification
possible with high volume production for each type of component within the transmission.
In the area of the hydraulic units, Sundstrand produces approximately 30,000 units per annum of a
similar size and type as used in this study, and again "judgment factors" were applied to this cost
data to arrive at figures for the production rates required in this study.
Traction drive components were estimated by similarity to automotive parts as much as possible
with due consideration for the accurate form grinding required on the toroids and rollers.
All of the above cost estimating assumed the use of highly automated machine tools and material
handling equipment used in very high volume production.
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C. Results of Cost Analysis
The cost for a typical three speed automatic transmission with torque converter was estimated on a
major subassembly basis, and is included for reference along with the hydromechanical and traction
transmission costs in Table VI11-1.
For a major component cost breakdown and comparison for the hydromechanical, traction, and a
"typical" 3 speed automatic transmission, see Table VIII-2.
Table VIII-1 Transmission Cost Comparison
Yearly Production Rate
100,000 1,000,000
Hydromechanical Tri-Mode $182 $122
Transmission
Traction Drive-Torque $149 $105
Converter Transmission
'Typical" 3 Speed Automatic — $ 89
Transmission
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Table VI11-2 Transmission Manufacturing Cost Breakdown Comparison
'Typical"
3 Speed
Automatic
$
PLANETARY GEAR SET 12.00
SHAFTING 6.00
TRANSFER GEARS
(Includes synchronizer assembly for
traction transmission)
CLUTCHES 21.00
HYDRAULIC UNIT —
(Excludes Bearings, Shafts)
CONTROLS SYSTEM 7.00
(Valve Body, Charge Pump, Linkage)
HOUSINGS, BULKHEADS, COVERS 14.00
(Includes sound isolators)
TORQUE CONVERTER 14.00
TRACTION DRIVE UNIT
(Excludes shaft, includes steering assembly)
ANTI-FRICTION BEARINGS 1.00
(Excludes planet bearings)
MISCELLANEOUS 2.00
(Bolts, seals, gaskets, filter, etc.)
TRANSMISSION ASSEMBLY AND TEST 12.00
TOTAL (To Nearest Dollar) $89.00
(COSTS BASED ON 1 MILLION UNITS PER YEAR)
Tri-Mode
Hydro-
Mechanical
$
12.40
8.50
11.20
15.60
22.80
7.00
16.50
11.00
2.00
15.00
Traction
Drive/
Converter
$
5.30
9.50
6.50
16.20
13.00
30.20
6.80
2.00
15.00
$122.00 $105.00
Sundstrand Aviation
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IX. REFERENCES
1. Prospects for Electric Vehicles - A Study of Low Pollution Potential Vehicles - Electric -
National Air Pollution Control Administration and Arthur D. Little, Inc. 1969.
2. Final Report - U. S. Army Contract DA-11-022-AMC-2269(T) U. S. Army Tank Automotive
Center-April - 1966.
3. Final Report - U. S. Army Contract DA-11-022-AMC-6950") U. S. Army Tank Automotive
Center - June 30, 1966.
4. Phase 1 - Final Report - Contract 68-04-0034 EPA, Office of Air Programs, Advanced
Automotive Power Systems Division — February, 1972.
5. "Design Practice — Passenger Car Automatic Transmissions" Part 1 and 2, issued by S.A.E.
6. 'Tractive Capacity and Efficiency of Rolling Contacts", Hewko, L. 0., Rounds, F. G., Scott, R.
L., Proceedings of the Symposium on Rolling Contact Phenomena, Joseph B. Bidwell, Ed., Elsevier,
N. Y., 1962.
7. "Design and Test of the First Aerodynamic Torque Converter", C.C. Hill, R.A. Mercure, and C.D.
Cole, ASME Paper 69-GT-109; March 1969.
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APPENDICES
APPENDIX 1-1, 1-2, 1-3, 1-4, 1-5, 1-6, 1-7 and 1-8;
referenced in Section I.
APPENDIX V-1, V-2, V-3, V-4, V-5 and V-6;
referenced in Section V.
APPENDIX VI-1, VI-2 and VI-3;
referenced in Section VI.
APPENDIX VII-1, VII-2, VII-3, VII-4, VII-5, VII-6(1)
and VII-6(2); referenced in Section VII.
There are no appendices for Sections II, III, IV,
VIII and IX.
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APPENDIX 1-1 Attachment 1. Scope of Work, Contract 68-04-0034
Task 6 - Transmission Study for Turbine and Rankine Cycle Engines
This task is generally similar to the effort in tasks 1, 2, 3, 4, 5 and
is an extension of the study to cover transmissions, for the gas turbine
and rankine cycle engines. For each engine the contractor shall assess
quantitatively the technical and economic feasibility of existing and
potential types of transmissions most suitable for the particular engine.
Based on this study an optimum transmission for each engine shall be
recoraacnded and thoroughly evaluated as outlined below:
6.1 - Requirement
6.1.1 - The transmission systems considered shall be suitable for
application in a full size family car. The specifications
of this vehicle are given in an attachment to the original
statement of work entitled "vehicle design goals". Vehicle
weight for performance calculation shall be the same as for
previous tasks.
6.1.2 - The transmission study for each of the two heat engines,
that is 1) gas turbine 2) rankine cycle, shall be based
on the engine characteristics and accessories requirements
to be supplied by the project officer within one week froa
the initiation of this study.
6.2 - Technical Feasibility Studv
The contractor shall conduct technical and economic feasibility
analysis of the various types of transmissions for the two engir.as
specified in section 6.1.2. The transmission types for each engine
considered shall include existing and potential transmission such
as:
1. Mechanical
2. Hydrostatic
3. Combination of mechanical and hydrostatic
A. Kydrokinetic
5. Electrical
6. Traction
7. Belt Drive
EPA may propose specific transmissions for consideration in this study.
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-2-
7'.-.o contractor shall provide, when requested by the project
officer, buc not earlier the.-; 30 days from the effective date of
the contract layout drawings of the transmission or parts of the
transmission, in order that independent checks of stress analysis,
thcrr.al analysis and safety analysis can be made.
6.3 Perromance Analysis
6.3.1. - Steady State Efficiency and Fuel Consumption
The contractor shall calculate and provide graphical plots
of steady state part load and full load efficiency or the
two selected transmissions for vehicle speed ranging fron
0-80 aph. (1/10, 1/4, 1/2 and 3/4 full load) . The
corresponding plots of fuel consumptions shall also be
provided. This shall include plot of transmission
efficiency and fuel consumptions for cruise speed, on
level road, ranging from 0-80 raph with and without air
conditioning load.
6.3.2. - Driving Cycle Efficiency and Fuel Consurr.r/tion
The contractor shall calculate the average efficiency and
the corresponding average fuel consumption for the two
selected transmissions over the Federal driving cycle,
with and without air conditioning load. The detailed
procedure and methods for calculating above efficiencies
and fuel consumptions shall be included, in. a separate
appendix attached to the final report.
6.& Control System Definition and Analysis
The contractor shall conduct control systems analysis on the entire
transriission/engine/vehicle system. Control system analysis shall
include:
a) a cursory stability analysis
b) safety analysis
c) analysis of possible "pathological case" operator
induced instability.
The transmission for each er.gir.c shall be readily adaptr.ble to the
corresponding hear engine power system presently investigated by
I?A. The contractor is responsible for iicscr. with EPA Rankine system
Contractor suca t.'.at the recor.r.er.cled transmission and its accessories
as a whole unit is adaptable for integration in the vehicle. EPA will
provide necessary information on Srayton Cycle system . All cooling
systems, control system, etc for the transmission shall natch the overall
vehicle system thus avoiding unnecessary complications and duplications of
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- 3 -
sub-systems. The contractor shall provide sufficiently detailed drawings
of the transmission control inputs, cooling system and accessories and
shall indicate how the total transmission with all its control inputs
and accessories fit the overall vehicle system.
6.5 Cost Analysis
The contractor shall perform cost analysis of the various transmission
concepts. Tha quantity of transmissions in units per year to be
considered are 100,000 and 1,000,000. This shall be original
equipment manufacturer (OEM) cost. The reference transmission,
against x»hich all cost and performance comparisons shall be made, is
the conventional multi-speed torque converter ("Automatic")
transmissions.
The detailed procedure and method for cost estimate shall be included
in a. separate appendix attached to the report.
6.6 Transmission Recommendation
A recommendation of an optimum transmission based on the system
cost and efficiency shall be made. This recommendation shall include
designs of the optimum transmission in such detail that accurate cost
estimates required in 6.5 above can be made. The recommendation shall
include heat engine optimum operational mode and control.
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APPENDIX 1-2
ENVIRON MENTAL PROTECTION
AGENCY
PROTOTYPE VEHICLE PERFORMANCE SPECIFICATION
January 3, 1972
Division of Advanced Automotive
Power Systems Development
2929 Plymouth Road
Ann Arbor, Michigan 48105
Approved:
^Mr". George M. Thur'
Chief, Power Systems Branch
Approved:
' / / r, ^
C L (-, ^ ^/' J .^-^ ''
John J. Brogan
Director, Div. of Advanced Automotive Powe- Systems Development
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ADVANCED AUTOMOTIVE POWER SYSTEMS (AAPSj
PROTOTYPE VEHICLE PERFORMANCE SPECIFICATION *
January 3, 1972
The AAPS Vehicle performance design specification presented below
is intended to provide:
A common objective for prospective contractors.
Criteria for evaluating proposals and selecting a contractor.
Criteria for evaluating competitive power systems for
entering first generation system hardware.
Advisory criteria to assist the contractor in such areas
as rolling resistance, vehicle air drag etc.
The derived criteria are based on typical characteristics of the
class of passenger automobiles with the largest market volume produced
in the U.S. during the model years 1969 and 1970. It is noted that
emissions, volume and most weight characteristics presented are maximum
values while the performance characteristics are intended as minimum
values. Contractors and prospective contractors who take exceptions
must justify these exceptions and relate these exceptions to the
technical goals presented herein.
*Supersedes "Vehicle Design Goals - Six Passenger Automobile"
(Revision C - May 28, 1971)
Page 1 of 11
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CONTENTS
Introduction
Vehicle Weight Without Propulsion System
Propulsion System Weight
Vehicle Curb Weight
Vehicle Test Weight
Gross Vehicle Weight
Propulsion System Volume
Air Drag
Rolling Resistance
Propulsion System Emissions
Fuel
Start Up, Acceleration and Grade Velocity
Performance
Hinimum Vehicle Range
Fuel Consumption
Accessory Power Requirements
Propulsion System Operating Temperature and
Pressure Range
Passenger Comfort Requirements
Noise Standards
Operational Life
Reliability and Maintainability
Cost of Ownership
Safety Standards
SECTION
1
2
3
4
5
6
7
8
9
10
11
12
13
14
15
16
17
18
19
20
21
PAGE
3
4
4
4
4
5
5
5
5
6
6
7,8
9
9
10
10
10
10
11
11
11
11
Test Conditions
Table I
Page 2 of 11
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INTRODUCTION
The design of an automobile from a total systems
standpoint could be expected to result in major benefits
in cost, safety, and performance. This specification
is intended as a step along that path, describing a
propulsion system that can be installed into engine
compartments as they now exist. Integration of the
vehicle accessories within the propulsion system are
highly desirable. Following the successful demonstration
of this development further optimization of the propulsion
system with the power train, suspension, and vehicle
styling would be possible.
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1. VEHICLE WEIGHT WITHOUT PROPULSION SYSTEM - Wo
W0 is the weight of the vehicle excluding the propulsion system.
This weight includes, but is not limited to the frame, body, glass,
trim, suspension, wheels (rims and tires), service brakes, seats,
upholstery, sound absorbing materials, insulation, dashboard
instruments, accessory ducting and wiring, accessories, and all other
components not included as part of the propulsion system.
Accessories are defined as driver assistance and passenger
convenience components and subsystems not essential to propulsion
system operation. Included are power steering systems, power
brake systems and passenger compartment heating and air conditioning
systems.
W0 is fixed at 2700 Ibs.
2. PROPULSION SYSTEM WEIGHT - Wp
Wp includes the energy storage subsystem (including fuel, containment,
and supply and deliver ducting) , power conversion subsystem (including
auxiliaries and control) and power transmitting subsystem (including
transmission and drive train to the driven wheels).
Auxiliaries are defined as components and subsystems essential to the
operation of the power conversion system. Included are electric power
generating subsystems, starting subsystems, exhaust subsystems, motors
fans, blowers, pumps, and fluids.
Lightweight propulsion system are highly desirable, however,
the maximum allowable propulsion system weight, Wpm, is 1500 Ibs.
3. VEHICLE CURB WEIGHT - Wc
Wc = W0 + Wp.
The maximum allowable vehicle curb weight, Wcm, is 4200 Ibs.
(2700 + 1500 max. = 4200).
4. VEHICLE TEST WEIGHT - Wt
Wt = Wc + 400 Ibs. Wt is the vehicle weight at which all accelerative maneuvers,
fuel economy and emissions are to be calculated. (Items 9c, lie, lid,lie,
llf, 12,13).
The maximum allowable test weight, Wtm, is 4600 Ibs. (2700 + 1500 max.
+ 400 = 4600).
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5. GROSS VEHICLE WEIGHT - Wg
W. - Wc •»• 1100 Ibs. We is the gross vehicle weight at which sustained
velocity capability at 5 percent and 30 % grades is to be calculated. (Item llf).
The 1100 Ibs. load simulates a full load of passengers and baggage.
The maximum allowable gross vehicle weight, Wgn, is 5300 Ibs. (2700 +
1500 max. + 1100 = 5300).
6. PROPULSION SYSTEM VOLUME - Vp
Vp is the volume allotted for all items identified under item 2. The
propulsion system shall be packagable in such a way that the volume encroachment
on either the passenger or luggage compartment does not exceed the following:
a) The transmission tunnel may not be widened so as to decrease
the selected production vehicle clearance between the
accelerator pedal and the tunnel. The accelerator pedal may
not be relocated.
b) Intrusion of the tunnel into the passenger side of the vehicle
may be increased by a maximum of 1.5 inches.
c) The tunnel height may be increased by a maximum of 2 inches
but without affecting the full fore and aft adjustment of the
front seat of the vehicle. The front seat may not be
raised.
The propulsion system shall not violate the vehicle ground clearance
lines as established by the manufacturer of the vehicle used for propulsion
system/vehicle packaging. Additionally, the propulsion system shall
not violate the space allocated for wheel jounce motions and vehicle
steering clearances. Necessary external appearance (styling) changes will be
minor in nature. The propulsion system shall also be packagable in such a
way that the handling characteristics of the vehicle are not degraded.
7. AIR DRAG
The product of the drag coefficient, Cd, and the frontal area, Af,
is to be used in air drag calculations. The product of C^Af has a
value of 12 ft . The air density used in computations shall correspond
to the applicable ambient air temperature.
8. ROLLING RlSISTANCE
Rolling resistance, R, is expressed in the equation
R - (W/65) [ 1 + (1.4 x 10-3V) + (1.2 x 10~5v2)] ibs. V is the
vehicle velocity in ft/sec. W is the vehicle weight in Ibs.
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9. PROPULSION SYSTEM EMISSIONS
The vehicle is to be tested for emissions in accordance with the procedure
jLn_the_ July._2, 1971 Federal Register and as further described in the Code of
Federal Regulations Title 40 Part 85 for model year 1976 light duty vehicles
(CFR 40-85). The vehicle test weight shall be Wt and the accessory load as
defined in Section 14a. Ambient conditions are 14.7 psia and 85°F.
Emission tests will be run with fuel specified in Section 10.
The Federal emissions standards are:
Hydrocarbons - 0.41 grams/mile maximum
Carbon Monoxide - 3.40 grams/mile maximum
Oxides of Nitrogen* - 0.40 grams/mile maximum
Prototype vehicles are to meet the following emissions goals to allow
for production tolerances and life degradation. Measurement of emissions
is to be taken after the system has operated for 100 hours.
Hydrocarbons - 0.20 grams/mile maximum
Carbon Monoxide - 1.70 grams/mile maximum
Oxides of Nitrogen* _ 0.20 grams/mile maximum
*0xides of nitrogen are to be measured or computed as N02
Production of smoke, odors, aldehydes, ammonia, particulates or other undesirable
emissions not now specified in the July 2, 1971 Federal Register are undesirable.
10. FUEL
Emission tests will use the fuel specified below, however, the power svstem
shall have the capability of meeting emission levels using commercialIv
available unleaded fuels.
Item ASTM Designation Specification
Octane, Research, min. D1656 91-93
Pb. (Organic), gm/U.S. gal D 525 < -02
Distillation range D 86
I.B.P., °F - 100-115
10 percent point,°F - 140-150
50 percent point,°F - 240-250
90 percent point,°F - 330-340
E.P. °F (max) - 425
Sulfur, Wt. percent max. D1266 0.10
Phosphorous, theory - 0.0
R.V.P. Ib. D 323 5.5-7.5
Washed gum (max) mgm/gal D 381 4.0
Corrosion (not lower than) D 130 IB
Oxidation stability (not less than) D 525 240+
Hydrocarbon composition D1319
Olefins, percent, max. - 30
Aromatics, percent, max. - 40
Saturates - Remainder
For computation purposes the lower heating values of this fuel is to be assumed as
18500 Btu/lh. The cost to he assured for svstem cost analvsis is $0.31/gallon.
An A.P.I, gravity of 56.0 is to be assumed in all calculation.
Page 6 of 11 Page 101
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11. START UP, ACCELERATION. AND GRADE VELOCITY PERFORMANCE
a. Start Up:
The vehicle must be capable of being tested in accordance with
the procedure outlined in the July 2, 1971 Federal Register without
special driver startup/warmup procedures. The accessory load shall
be as defined in Section 14b.
The maximum time from "key on" to reach 65 percent of full power
level is 45 sec. Ambient conditions are 14.7 psia 60°F. The vehicle
is to be soaked at this temperature for a minimum of 12 hours prior
to initiation of start test.
Powerplant starting procedures in low ambient temperatures shall
be equivalent to or better than the typical automobile spark-ignition
engine. After a 24 hour soak at -20°F and 14.7 psia the engine shall
achieve a self-sustaining idle condition without further driver input
within 25 seconds. No starting aids external to the normal
vehicle system shall be needed at or above - 20°F.
b. Idle operation conditions:
The idle creep torque shall not result in level road operation of
the vehicle at a speed in excess of 18 mph in high gear, with the
entire propulsion system at steady state operating temperature and
ambient conditions of 14.7 psia and 85°F. The accessory load shall be
as defined in Section 14a.
c. Acceleration from a standing start;
The minimum distance to be covered in 10.0 sec. is 440 ft. The maximum
time to reach a velocity of 60 mph is 13.5 sec. Ambient conditions
are 14.7 psia, 85°F. Vehicle weight is Wt and accessory load as
defined in Section 14a. Acceleration is on zero grade and initiated
with the engine at the normal idle condition.
d. Acceleration in merging traffic:
The maximum time to accelerate from a constant velocity of 25 mph
to a velocity of 70 mph is 15.0 sec. Ambient conditions are 14.7
psia, 85°F. Vehicle weight is Wt and accessory load as defined in
Section 14a. Acceleration is on zero grade and time starts when
the accelerator pedal is depressed.
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e. Acceleration, DOT High Speed Pass Maneuver:
The maximum time and distance to go from an initial velocity
of 50 mph with the front of the automobile (18 foot length assumed)
100 feet behind the back of a 55 foot truck traveling at a constant
50 mphjto a position where the back of the automobile is 100 feet
in front of the front of the 55 foot truck ,is 15 sec. and 1400 ft.
The entire maneuver takes place in a traffic lane adjacent to the
lane in which the truck is operated. Vehicle is accelerated
until the maneuver is completed or until a maximum speed of 80 mph
is attained, whichever occurs first. Vehicle acceleration ceases
when a speed of 80 mph is attained, the maneuver then being completed
at a constant 80 mph. (This does not imply a design requirement
limiting the maximum vehicle speed to 80 mph). Time starts when
the accelerator pedal is depressed. Ambient conditions are 14.7 psia,
85°F. Vehicle weight is Wt and accessory load as defined in Section 14a.
Acceleration is on zero grade.
f. Grade velocity:
The vehicle must be capable of starting from rest on a thirty
percent (30%) grade and ascending the grade at a minimum speed of
5 mph. The vehicle must be capable of this maneuver in both the
forward and reverse directions at a vehicle weight of Wg, with
the accessory load as defined in Section 14a.
The minimum cruise velocity that can be continuously maintained on
a five percent (5%) grade shall be not less than 65 mph with a vehicle
weight of Wg and accessory load as defined in Section 14a.
The minimum cruise velocity that can be continuously maintained
on a zero percent (0%) grade shall be not less than 85 mph with a
vehicle weight of Wt and with the accessory load as defined in Item 14a.
Ambient conditions for all grade specifications are 14.7 psia and
85°F.
Performance degradation attributable to loss of powerplant efficiency
at extreme temperatures shall not exceed ten percent (10%) relative
to the performance values specified at 85°F. This limitation applies
to ambient temperatures from -20°F to 105°F.
The wind velocity is to be less than 10 mph for all acceleration and
grade tests.
Page 8 of 11
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12. MINIMUM VEHICLE RANGE
Minimum vehicle range without refueling will be 200 miles (maximum fuel
capacity is 25 U.S. gallons). The minimum range shall be calculated for,
and applied to each of the following modes:
1. Cyclic mode is: The Federal driving cycle which is in accordance
with the July 2, 1971 Federal Register. The range
may be calculated for one cycle and ratioed to
200 miles.
2. Cruise mode is: A constant 70 mph cruise on a zero grade for
200 miles.
The vehicle weight for both modes shall be Wt initially and with accessory
power levels as specified in Section 14 . The ambient conditions shall be
a pressure of 1A.7 psia, and a temperature of -20°F (air conditioner off)
and 105°F (air conditioner on).
13. FUEL CONSUMPTION
Using the fuel specified in Section 10,a "fuel economy" figure shall be
calculated based on 1) miles per gallon and 2) the number of Btu per mile
required to drive the vehicle through the following modes of operation:
Avg. Speed Hours % of Time
1) Federal Driving Cycle 19.84 1750 50
2) Simplified Surburban Route 30.00 1150 33
(equal times at constant
20,30 and 40 mph speeds).
3) Simplified Country Route 60.00 600 17
(equal times at constant
50,60 and 70 mph speeds).
Totals 30 3500 100
In all cases the system fuel consumption shall be calculated for a vehicle
weight of Wt initially, and power levels as specified in item 14a. Ambient
conditions are 14.7 psia and 85°F.
It is desirable that the fuel consumption rate at idle operating
condition not exceed 7 Ibs/hour.
Page 9 of 11
Page 104
-------
14. ACCESSORY POWER REQUIREMENTS
a. Accessory power requirements with the air conditioning in operation
are defined as 15 hp at maximum engine speed and 4 hp at engine idle speed,
with a linear relationship between these two points.
b. Accessory power requirements without the air conditioning in operation
are defined as 5 hp at maximum engine speed and 2 hp at engine idle speed,
with a linear relationship between these two points.
15. PROPULSION SYSTEM OPERATING TEMPERATURE AND PRESSURE RANGE
The propulsion system shall be operable within an expected ambient
temperature range of -40° to 125°F.
The propulsion system shall be operable within an expected environmental
pressure range of 9 psia to 15 psia.
16. PASSENGER COMFORT REQUIREMENTS
Heating and air conditioning of the passenger compartment shall be at a rate
equivalent to that provided in the present (1970) standard full size family car.
Present practice for maximum passenger compartment heating rate is
approximately 30,000 Btu/hr. For an air conditioning system at 110°F
ambient, 80°F and 40% relative humidity air to the evaporator, the
rate is approximately 13,000 Btu/hr.
17. NOISE STANDARDS
a. Maximum noise test:*
The maximum noise generated by the vehicle shall not exceed 77 dbA
when measured in accordance with SAE J986a. Note that the noise level
is 77 dbA whereas in the SAE J986 the level is 86 dbA.
b. Low speed noise test:*
The maximum noise generated by the vehicle shall not exceed 63 dbA
when measured in accordance with SAE J986a except that a constant
vehicle velocity of 30 mph is used on the pass-by.
c. Idle noise test:*
The maximum noise generated by the vehicle shall not exceed 62 dbA
when measured in accordance with SAE J986a except that the engine is
idling (clutch disengaged or in neutral gear) and the vehicle is
stationary. A 360° survey shall be made, the microphone being 10 feet
from the vehicle perimeter.
* The air conditioner will not be in operation during noise tests.
Page 10 of 11
Page 105
-------
18. OPERATIONAL LIFE
The design lifetime of the propulsion system In normal operation will
be 3500 hours minimum.
Termination of the operational life of an engine shall be determined
by structural or functional failure. Functional failure is defined as
power degradation exceeding 25 percent of maximum output of the rear wheels.
19. RELIABILITY AND MAINTAINABILITY
The reliability and maintainability of the vehicle shall equal or
exceed that of the spark-ignition automobile. The mean-time-between-
failure should be maximized to reduce the number of unscheduled service
trips. No failure modes shall present a serious safety hazard
during vehicle operation and servicing. Failure propagation should be
minimized. The power plant should be designed for ease of maintenance
and repairs to minimize costs, maintenance personnel education, and
downtime.
20. COST OF OWNERSHIP
The initial cost and net cost of ownership of the vehicle shall be
minimized for ten years and 105,000 miles of operation.
21. SAFETY STANDARDS
The vehicle shall comply with all Department of Transportation Federal
Motor Vehicle Safety Standards in force when the selected test vehicle
was manufactured.
Page 11 of 11
Page 106
-------
TABLE I
Section
9. EMISSIONS
11.PERFORMANCE
a. Start Up
b. Idle
c. Accel ( 0-60)
d. Accel (25-70)
e. Accel (50-80)
f. Grade (30%)
(5%)
(0%)
12. VEHICLE RANGE
13.FUEL CONSUM.
TEST CONDITIONS
(ENGINE DYNAMOMETER, CHASSIS DYNAMOMETER, ROAD)
Performance Requirements Accessory Power Weight Temperature
HC 0.20 grams per mile** 14a
CO 1.70 grams per mile** 14a
N02 0.20 grams per mile** 14a
65%* power in 45 sec** 14b
Driver Assistance 25 sec** 14b
Creep 18 mph** 14a
13.5 sec** to 60 MPH 14a
440 ft* in 10.0 sec. 14a
15.0 sec** from 25 to 70 MPH 14a
15.0 sec** and 1400 Ft** 14a
From 50 to 80 MPH**
0 to 5 MPH* 14a
65 MPH* 14a
85 MPH* I4a
Wt 85°F
Wt 85°F
Wt 85°F
Wt 60°F
Wt -20°F
Wt 85°F
Wt 85°F
Wt 85°F
Wt 85°F
Wt 85°F
W. 85°F
W* 85°F
200 MI* /"""^
1. During )«£(TDCj 14b and 14a W -20°F and 105°F
2. At 70 MPH 14b and 14a Wt -20°F and 105°F
MPG During FDC 14a A
MPG at 20,30, and 40 MPH 14a •
MPG at 50,60, and 70 MPH 14a •
'- Wt 85°F
Wt 85°F
Wt 85°F
Pressure
14.7
14.7
14.7
14.7
14.7
14.7
14.7
14.7
14.7
14.7
14.7
14.7
14.7
14.7
14.7
14.7
14.7
14.7
NOTES: Emission tests will be run with fuel specified in Section 10.
Road test wind conditions shall not exceed 10 MPH in any direction.
*Minimum values
**Maximum values
-------
{SPACER PAGE - INTENTIONALLY BLANK)
Page 108
Sundstrand Aviation
-------
APPENDIX 1-3 FEDERAL DRIVING CYCLE
RULES AND REGULATIONS
17311
Ami»ii A
^ MMM VTN AMOMCTVI MIVXNO BCIIKDVU
(0p**d T«r.u» Tim* Stqucno*)
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IT 17.3
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It 10.7
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M 32.1
M 21.8
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43 155
43 180
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10 326
11 31.3
11 190
11 17.1
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11 346
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17 30.1
88 304
89 307
80 30.7
tl 304
82 30.4
93 30 J
94 304
ti 30.1
99 30.4
97 299
91 294
99 29.8
100 303
101 30.7
102 30.9
103 31.0
104 30.9
10S 30.4
108 398
107 299
108 30.2
109 30.7
110 31.3
111 31.8
113 32.3
113 32.4
114 32.2
115 31.7
116 28.6
117 25.3
118 220
lit 18.7
120 15.4
121 12.1
122 88
123 55
12« 22
125 00
126 00
127 00
12A 00
120 00
130 00
131 00
132 00
133 00
134. 0.0
135 0.0
136 00
137 0.0
138 00
139 0 0
140 00
141 00
142 00
143 00
144 00
145 00
148 00
147 00
148 00
149 00
150 00
151 00
152 0.0
151 00
154 00
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101 00
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164 33
165 68
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Tim, Spc'4
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1&8 16.5
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170 223
171 243
173 25.8
173 264
174 25.7
17S 25.1
178 24.7
177 250
178 25.2
179 254
180 25.8
181 27.2
182 26 S
183 24.0
184 22.7
185 19.4
186 17.7
187 173
188 181
189 188
190 200
191 22.2
192 244
193 273
194 304
195 334
196 33.2
197 37.3
198 333
199 404
200 42.1
201 43.5
202 45.1
203 46 0
204 4G 8
205 47.5
206 47 5
207 47.3
203 47.2
209 47.0
210 470
211 470
212 470
213 47.0
211 472
215 474
210 470
217 485
218 43.1
219 435
220 50 0
221 50.6
222 51 0
223 615
224 52 2
225 53 2
226 54 1
227 54 8
228 54 3
229 55 0
230 54 9
231 546
232 548
233 54 8
234 55.1
235 55 5
238 557
217 00 1
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240 007
2»1 007
242 50 5
243 50 5
244 $0 5
245 50 5
248 565
347 56 5
248 664
243 66 1
250 658
251 65.1
AmNDU A— Continued
Time Sfttt
(«re.) («».*.>
253 54.8
253 64.3
354 64.0
355 53.7
256 536
257 534
258 54.0
259 54.1
250 54.1
261 63.8
2G2 53.4
363 53.0
364 52.6
265 52.1
266 52.4
267 52.0
268 614
269 SI. 7
270 S14
371 51.8
373 51.8
273 52.1
274 524
27S 53.0
378 534
277 540
278 54.9
279 55.4
280 55 6
281 56.0
282 560
283 55.8
284 55.2
285 544
286 53.6
287 524
288 514
239 51.5
230 514
291 81.1
292 50.1
233 50.0
234 50.1
230 50.0
236 49.6
237 495
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3(11 4110
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303 47 2
304 40.1
305 45 0
306 438
307 42 6
303 41 5
303 40 3
310 305
311 370
312 352
313 338
314 325
315 31.5
316 306
317 305
313 300
319 290
320 275
321 24.8
322 214
323 20.1
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125 18 5
325 170
327 15 5
321) 125
3M 108
330 80
331 4.7
332 1.4
33.1 0 0
334 00
335 0 0
336 00
337 0.0
338 0.0
Tim* Speed
(«re.) (m p ».>
339 0.0
340 00
341 0.0
343 0.0
343 0.0
344 0.0
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346 0.0
347 1.0
348 43
349 7.8
350 10.0
3S1 142
352 17.3
353 200
354 224
355 23.7
358 25.2
357 26 8
358 28.1
359 30.0
360 30.8
361 31.8
362 32.1
363 32.8
364 33.6
365 344
366 34.6
367 34.9
368 34.8
369 344
370 34.7
371 35.5
372 36.0
373 36.0
374 36.0
375 36.0
376 30.0
377 36.0
378 38.1
379 364
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382 3G.O
383 35.1
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330 203
331 17.5
332 145
333 120
334 87
335 5.4
336 2.1
337 0.0
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399 00
400 00
401 00
402 00
403 26
404 59
405 92
406 125
407 158
403 19.1
403 224
410 250
411 206
412 275
413 2')0
414 300
410 30 1
410 300
417 297
418 293
419 288
420 2dO
421 250
422 21.7
423 184
424 15.1
425 11.8
Tim* rfpfrJ
(•re.) (n ».A.)
428 84
427 1.2
423 1.9
429 00
430 OO
431 0.0
433 0.0
433 0.0
434 0.0
435 0.0
438 0.0
437 0.0
438 0.0
439 0.0
440 0.0
441 00
442 00
443 00
444 0.0
445 0.0
448 0.0
447 00
448 33
449 6.8
450 9.9
451 13.2
452 164
453 19 8
454 23.1
455 26.4
456 27.8
457 29.1
458 314
459 33.0
460 33.8
461 348
462 35.1
463 356
461 36.1
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4G« 36.1
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408 33 0
409 35.7
470 300
471 360
473 35 G
473 354
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470 30 3
478 30 2
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478 352
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481 350
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432 33 5
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001 103
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Ammix A— Continued
rim* trttt
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614 65
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617 104
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521 17.7
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529 24.9
530 25.0
531 250
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517 25.8
538 25.0
539 25.6
540 25.3
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545 23.1
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547 16.5
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570 66
571 99
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574 1GO
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576 170
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579 177
580 177
581 175
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534 10 6
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593 176
504 185
535 192
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537 210
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599 21.2
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601 32.0
602 224
603 22 S
604 224
60S 224
606 22.7
607 23.7
608 25.1
609 26.0
610 264
611 37.0
612 26.1
613 22.3
614 194
61S 16-2
618 12.9
617 96
618 63
619 3.0
620 0.0
621 0.0
622 0.0
623 0.0
624 0.0
62$ 0.0
626 0.0
527 0.0
6? 3 0.0
62J 0.0
630 0.0
631 0.0
632 0.0
633 00
634 00
635 0.0
636 0.0
637 00
638 0.0
639 0.0
640 0.0
611 0.0
G12 00
643 0.0
644 0.0
615 00
CI6 20
047 44
018 78
GO 103
G',0 125
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603 174
654 196
605 210
6S1 222
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GOO 25 6
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G02 291
GC3 20.2
004 262
655 264
666 265
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608 20 0
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670 23 8
571 214
072 185
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G70 110
070 87
G77 58
678 35
673 20
380 0.0
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633 0 0
633 0.0
634 0.0
685 0.0
Time Stn4
(«re.) (m p ».|
634 00
687 0.0
683 0.0
68) 0.0
630 0.0
691 0.0
833 0.0
693 0.0
694 1.4
695 33
696 4.4
697 65
698 9.3
639 11 J
700 134
701 148
703 1C 4
703 16.7
704 164
705 164
706 183
707 19.2
708 30.1
709 314
710 224
711 22 S
713 22.1
713 22.7
714 23.3
71S 234
716 224
717 31.6
718 204
719 18.0
720 150
721 12.0
722 9.0
723 62
724 4S
725 3.0
728 2.1
727 OS
723 OS
729 33
730 64
731 9.0
732 12 S
733 140
734 160
735 180
736 190
737 215
733 231
739 245
740 255
741 265
742 27.1
743 276
744 27 9
745 283
740 280
747 23 6
748 28 3
7O 23 2
750 23 0
751 275
752 26 8
753 15 5
754 235
755 2! S
750 190
757 16 S
753 143
753 l.'S
701) 9 4
701 83
7C3 30
7G3 1 S
7C» IS
765 05
7G6 00
767 30
7SS 63
7-3 96
770 139
771 158
773 174
•EDEIAL «£CISIE«, VOL 35, NO. 21?—TUESDAY, NOVEMIEI 10, 1»70
Table APP-I-3 DHEW Urban Dynamometer Driving Cycle
Sundstrand Aviation
Page 109
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11011
RULES AND REGULATIONS
Am
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109 143
• 10 14.3
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113 140
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• 14 138
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• 19 319
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•33 100
•34 29.9
135 199
•26 299
•37 29 9
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•31 393
•31 289
•33 38.3
•34 27.7
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148 340
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6C3 28.1
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•70 25.1
•71 25 6
•73 257
•73 263
•74 269
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•76 278
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•78 29 0
879 29 3
880 29.1
881 29.0
883 289
883 28.5
884 28.1
885 28.0
886 28.0
887 27.6
888 27.2
889 26.6
890 27 0
891 27.5
832 27.8
893 28.0
894 27.8
895 38.0
896 23.0
897 280
898 27.7
899 27.4
600 269
901 26.6
902 26 S
903 26 5
904 265
905 26.3
906 262
907 262
908 25.9
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910 25.6
911 259
912 258
913 255
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915 235
918 222
917 21.6
918 216
919 21.7
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921 234
922 24 0
923 24 2
924 24 4
925 249
926 25 1
927 25 2
928 253
923 255
930 253
931 25 0
932 25 0
933 25 0
934 24 7
935 245
936 243
937 24 3
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949 22.0
950 30.1
951 16.9
952 13.6
953 104
954 7.0
955 3.7
950 0.4
957 0.0
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959 0.0
900 2.0
961 6.3
962 86
963 11.9
964 15.3
905 17.5
9C6 18.6
967 20.0
968 31.1
969 32.0
970 33.0
971 345
973 26.3
973 374
974 28.1
975 28.4
976 28.5
977 28.5
978 28.5
979 27.7
980 274
981 272
982 26.8
983 26.5
984 26.0
985 25.7
986 252
987 240
988 230
933 21.5
930 21.5
991 21.8
992 22.5
993 23.0
994 22.8
905 228
996 23 0
997 227
998 22.7
999 22.7
1.000 23 5
1.001 240
1.002 24.6
1.003 24.8
1.004 25.1
1.005 25 S
1.008 25.6
1,007 25.5
1.008 250
1.009 24 1
1.010 23.7
1.011 232
1.012 22.9
1.013 22.5
1.014 220
1,015 216
1.016 205
t.017 175
1.018 142
1.019 109
1.020 76
.021 43
.022 1 0
.023 0.0
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.025 0 0
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.028 00
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.030 00
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1.033 00
1.034 0.0
Am
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1.036 0.0
1.037 00
1.038 0.0
1.039 0.0
1.040 0.0
1.041 0.0
1.012 00
1.043 00
1.014 00
1.045 0.0
1.046 0.0
1.047 0.0
1.048 0.0
1.049 0.0
1.050 0.0
1.051 0.0
1.053 0.0
1.053 1.2
1.054 4.0
1.055 73
1.056 10.6
1.057 13.9
1.058 17.0
1.059 18.5
1.060 20.0
1.061 31.8
1.062 23.0
1.063 240
1.064 34.8
1.065 256
1.066 264
1.067 26 8
1.068 27.4
1.069 27.9
1.070 28.3
1.071 28.0
1.072 27.5
1.073 270
1.074 27.0
1.075 26.3
1.076 245
1.077 12.5
1.078 21.5
1.079 206
1.080 180
1.081 15.0
1.082 12.3
1.083 11. 1
1.031 10.6
1.035 10.O
I.OSo 8.5
1.087 9.1
1.088 8.7
1.039 8.8
1.000 88
1.091 90
1.002 8.7
1.093 8 9
1.0D4 8.0
1.095 7.0
1.099 50
1.097 4.2
1.008 26
.009 1.0
.100 00
.101 O.I
.103 08
.103 1.8
.104 36
.105 69
.106 100
.107 12 8
.108 140
.109 145
.110 160
.111 181
.112 20.0
.113 210
.114 312
.115 2!3
.US 21.4
.117 21.7
.118 225
.119 230
.120 23 8
,121 244
inora A— ConU
Tim* Bete*
Ucr.l (mp.h.)
1.122 25.0
1.123 24.9
1,124 24.8
1.125 35.0
1.126 25.4
1,127 258
1,128 26.0
1.129 26.4
1.130 29.4
1.131 20.9
1.133 27.0
1,133 27.0
1.134 27.0
1.135 26.9
1.136 298
1.137 368
1.138 395
1.139 264
1.140 26.0
1.141 255
1.142 24.4
1.143 33.5
1.144 314
1.145 200
1.148 17.5
1,147 160
1.148 140
1.149 10.7
1.150 7.4
1.151 4.1
1.153 08
1.153 0.0
1.154 0.0
1.155 00
1.154 00
1.157 0.0
1.158 0.0
1.159 0.0
1.160 0.0
1.161 0.0
1.162 0.0
1.163 0.0
1.164 O.O
1,165 0.0
1.169 0.0
1,167 0.0
1.163 0.0
1.163 2.1
1.170 5.4
1.171 8.7
1.173 120
1.173 153
1.174 186
1.175 21.1
1.178 23.0
1.177 23.5
1.173 230
1.179 22.5
1.180 200
1.131 16.7
1.183 13.4
1.133 10.1
1.184 68
1.185 35
1,186 02
1.187 00
1,188 0.0
1.189 00
1.100 00
1.131 0.0
1.192 00
1.103 00
1.194 00
1.135 00
1.198 00
1.107 0.2
1.108 15
1,199 35
1.200 65
1.201 98
1.202 120
4.203 129
1.204 130
1.205 126
1.209 128
1.207 13.1
1.208 13.1
nued
Tim* S?rt4
(IK.) <».».*.)
1.209 140
1.210 155
1.211 17.O
1.213 18.6
1.213 19.7
1.314 21.0
1.215 21.5
1.216 21.8
1.217 31.8
1.218 315
1.319 31.3
1.220 31.5
1.221 21.8
1.223 220
1.223 21.9
1.224 21.7
1.225 314
1.226 314
1.227 21.4
1.228 30.1
1.223 19 5
1.230 193
1.231 19.6
1.232 19.8
1,233 200
1.234 19.5
1.235 175
1.238 15.5
1.23T 13.0
1.238 100
1.239 8.0
1.240 6.0
1.241 4.0
1.242 2.5
1.243 0.7
1.244 0.0
1.245 0.0
1.246 00
1.247 0.0
1.248 0.0
1.219 0.0
1.2JO 0.0
1.251 O.O
1.252 1.0
1,253 l.O
1.254 1.0
1J2SS 1.0
1.250 1.0
1.257 1.6
1.258 3.0
1.259 4.0
1.260 5.0
1.261 8.3
1.262 8.0
1.263 100
1.264 105
1.205 9.5
1.2C6 8.5
1.267 7.6
1.268 8.8
1.269 11.0
1.270 14.0
1.271 170
1.272 195
1.273 21.0
1.274 218
1,275 22 2
1.273 230
\.2~n 236
1.278 24.1
1.279 245
1.200 245
1 281 2 \ 0
1 212 23 5
1.233 235
1.284 235
1.283 23 S
1.210 235
1.207 235
1230 240
1.280 24 1
1.200 24 5
1.231 247
1.202 25.0
1.203 25 4
1.204 35 8
1.235 25.7
Am
Ttm* Sect£
l^M 20.0
1.237 2G.3
1438 37.0
14*9 274
1400 383
1401 29.0
1403 29.1
1403 39.0
1404 28.0
1405 34.7
1406 31.4
1407 18.1
1404 14.8
1409 115
1410 8.3
1411 4.9
1413 1.6
1413 0.0
1414 0.0
1415 0.0
1414 0.0
1417 0.0
1411 0.0
141* 0.0
1430 0.0
1421 0.0
CMBOt A—ConU
Tim* Bftt4
1423 00
1.323 00
1.334 0.0
1.325 0.0
1426 0.0
1,327 0.0
1.328 0.0
1,329 0.0
1.330 0.0
1,331 0.0
1.333 0.0
1.333 0.0
1.334 0.0
1435 0.0
1.334 00
1437 0.0
1.338 1.8
1.339 48
1.340 (.1
1.341 11.4
1443 133
1.343 1S.1
1444 168
1443 183
1444 194
nued
Tim* SHI
(fix:.)
-------
*J
H-
C
T3
(0 i
C M
3 I
CL w
(A
O
rt
vfl
5
O
M
(D
80
70 J
60 .
50
40
(/>
il
20
10
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t
' A
r y A
;
! '
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r -r ; • r~r
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'. i_ .4 -I-1-;- ---i
; r' i ' i ' ' i
i ••: -i : !•- !
300
GOO 900
TIME SEC.
1200
1350
-------
(SPACER PAGE - INTENTIONALLY BLANK)
Page 112
Sundstrand Aviation
-------
APPENDIX 1-4. RANKINE ENGINE DATA
To evaluate the performance characteristics of the EPA car with the Aerojet Rankine engine and an
infinitely variable transmission, it was necessary to define in detail the characteristics of the engine.
The optimum engine operating speed as a function of required engine power was determined, as
well as the specific fuel consumption characteristics.
From the engine performance map supplied by Aerojet and a knowledge of transmission
performance characteristics, it was possible to determine an engine operating speed curve as a
function of required engine power. The curve selected offers a good compromise in balancing the
engine efficiency and the transmission efficiency in the effort to maximize system efficiency (see
graph. Figure 1-4).
Specific fuel consumption data in pounds of fuel per horsepower hour as a function of engine
output power and vehicle velocity was supplied by Aerojet (see graph. Figure I-4A).
Engine output power as used here means total engine output power, that is, input power to the
transmission plus vehicle accessory power.
The engine reduction gear mesh has already been accounted for in the specific fuel consumption
data.
Sundstrand Aviation r^s
-------
2
•8
3
Q.
5<
° 55"
••* f*
5'
^,UUU -
i
GC
UJ
UJ
a.
^> in nnn
U-
/rt 10 «v\n
111
z
CD
1-
111
a
2
1
'
[
/
/
/
/
f
/
/
,x"
10 20 30 40 50 60 70 80 90 100 110
OUTPUT POWER (HP2)
120 130 140 150 160 170
Figure 1-4 Rankine Engine Performance - Speed vs. Output Horsepower
-------
CO
10 20 30
40 50 60 70 80 §0 100 110 120 130 140 150 160 170
ENGINE OUTPUT POWER (HP)
Figure I-4A Rankine Engine Performance - Specific Fuel Consumption vs,
Output Horsepower
-------
(SPACER PAGE - INTENTIONALLY BLANK)
Pa96116 Sundstrand Aviation
-------
APPENDIX 1-5. BRAYTON ENGINE DATA
To evaluate the performance characteristics of the EPA car with the AiResearch Brayton engine and
an infinitely variable transmission, it was necessary to define in detail the characteristics of the
engine. The optimum engine operating speed as a function of required engine power was
determined, as well as the specific fuel consumption characteristics.
Assumed engine data is based upon information supplied by AiResearch. Engine speed ranges from
60% speed at idle to 105% speed for short duration bursts at maximum power.
However, within the normal operating speed range of 60% to 100% speed, the highest continuous
permissible operating conditions of temperature and inlet guide vane position are 1700°F and 1.00
respectively.
From 0 to 15 HP, the minimum specific fuel consumption (min. SFC) is obtained at 60% speed.
Above 15 HP the min. SFC engine speed curve was assumed to coincide with the curve that defines
the 1700°F/1.00 curve mentioned above, since operating above that curve for any length of time is
detrimental to the life of the engine (see graph. Figure I-5).
When the power requirements exceed approximately 76 HP, engine speed is allowed to increase
above 100% speed for short periods of acceleration (see graph. Figure 1-5B).
Specific fuel consumption in pounds per horsepower hour was calculated from data supplied by
AiResearch (see graphs. Figure 1-5 A).
Engine power as used by Sundstrand implies total engine output power, that is, input power to the
transmission plus vehicle accessory power. Since AiResearch had assumed a constant 4 HP vehicle
accessory load, it was necessary to add 4 HP to all the data to find total engine power.
No engine reduction gear mesh had been assumed by AiResearch, therefore, it was necessary to
include this power loss in the fuel consumption calculations.
Page 117
Sundstrand Aviation
-------
3
Q.
(A
T
CO
a.
O
IGV"T4°F
1.025,1900
1.025,1900 F
WATER INJECTION
1.025, 1BOOF
WATER INJECTION
NOTE: INCLUDES 4 HP CONSTANT VEHICLE
ACC. IN ADDITION TO 6 HP ENGINE AC
55
60
65
70
75 80 85
PERCENT ENGINE SPEED
90 95 100 106
Figure 1-5 Brayton Engine Output Shaft HP vs. Percent Engine Speed
-------
W
•3
•3
•S
(0
1.40
1.30
.30
10 20 30 40
50 60 70 80 90 100 110 120 130 140 150 160
ENGINE POWER
Figure I-5A Brayton Engine Specific Fuel Consumption vs. Engine Output HP
-------
•8
M
O
3
a
-a
.a.
;<
55'
.6"
3
85.000
§ 75.000
a.
oc
2 65.000
tu
85
t 55.000
I
ui
45.000
oa
-------
APPENDIX 1-6. IDLE FUEL CONSUMPTION
Idle fuel consumption has proven to be an important parameter in this study, since a good deal of
the duty cycle is at idle. It was calculated from the specific fuel consumption data presented in
Appendicies 1-4 and 1-5. The formula for fuel consumption rate is: specific fuel consumption +
engine power.
_LB_
HP-HR
LB
HR
HP,
Engine power is total output engine power, that is, it includes the power absorbed by the
transmission at engine idle, and vehicle accessory power.
10 15 20
TOTAL ENGINE OUTPUT POWER (HP)
Sundstrand Aviation
Page 121
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(SPACER PAGE - INTENTIONALLY BLANK)
Page 122
Sundstrand Aviation
-------
APPENDIX 1-7 MAXIMUM TOTAL ENGINE POWER VS. VEHICLE SPEED
The engine data received from EPA indicated that the Aerojet Rankine cycle engine had a
maximum output power that increased with vehicle speed. This increase was due to the ram
air effect on condenser cooling. The Airesearch gas turbine has a maximum output power
which is not affected by vehicle speed. The power characteristics of the engine are shown in
Figure 1-7.
Page 123
Sundstrand Aviation
-------
a
V)
f+
: 3
•3
•0)
• >-••_
.5'
170
160
150 -
a.
I
O 140
130
i
BRAVION (155 HP) - FOR TRACTION TRANS.
BRAYTON (145 HP) - COR TRIMODF. TRANSMISSION
T ' 1 ~
10 20 30
i
40 50
VEHICLE SPEED (MPH)
60 70
80
90
Figure 1-7 Maximum Total Engine Power vs. Vehicle Speed
-------
APPENDIX 1-8. VEHICLE ACCESSORY POWER REQUIREMENTS
The vehicle accessory power requirements are defined in the Prototype Vehicle Performance
Specification, Section 14, page 10, as follows:
Accessory Power Requirements
a. Accessory power requirements with the air conditioning in operation are defined as 15 HP at
maximum engine speed and 4 HP at engine idle speed, with a linear relationship between these two
points.
b. Accessory power requirements without the air conditioning in operation are defined as 5 HP at
maximum engine speed and 2 HP at engine idle speed, with a linear relationship between these two
points.
These accessory loads were based upon the operating speed range of a typical internal combustion
(I.C.) engine. (Operating speed range is the maximum engine speed divided by the engine idle
speed.) The operating speed range of a typical I.C. engine is 6:1.
However, the operating speed range of the Rankine and Brayton engines used in this study are
considerably narrower. (1.375:1 for the Rankine engine, and 1.750:1 for the Brayton engine.)
It was assumed that vehicle accessory loads, and therefore vehicle accessory speeds, would be the
same at engine idle regardless of the type of engine that was used.
Then obviously at maximum engine speed, the vehicle accessory speeds and loads should be less
with an engine with a narrower speed range. Therefore, it seemed unreasonable to apply the same
maximum accessory power requirement to different engines with different speed ranges.
Consequently, with the agreement of the EPA, the accessory power requirement at maximum
engine speed was factored down linearly as a function of maximum engine speed ratio (see graph,
Figure
Page 125
Sundstrand Aviation *.*&
-------
1-8 (continued)
Page 126
Sundstrand Aviation
-------
I APPENDIX V-l j
Page 127
MOCWTlNO FLANGE TO MATE
WfTM ENGINE GEARBOX
•SELECTOR. LEVER (P,R,N,F>NO
CONNECTION TO EN3INE CONTROL
IN TH!S AREA. TO BE CO-ORDINATED
VAKA8LE HYDRAULIC INT
GEAR CENTER
(ALTERNATE INPUT CENTER)
DIRECTION OF
ROTATION CCW
LOOKING WTO
OUTRJT END
i- MOUNTING TO BE COORDINATED
2-OtL PORTS TO AND FROM COOLER
AND OH. LEVEL CHECK LOCATIONS
TO GE CO-OROINATEO
-------
-------
I APPBHBDC V-S
Page 129
DIRECTION OF
WPUT ROTATION-
,JNTING FLANGE
3 MATE WITH
ENGINE GEARBOX
OUTPUT SHAFT
SELECTOR LEVEL
(P.R.N.F.) AND
CONNECTION TO
ENGINE CONTROL
IN THIS AREA
NOTES
I-ENGINE MOUNTING TO BE
CO-ORDINATED
2-OIL PORTS TO AND fROM
COOLER AND OIL LEVEL
CHECK LOCATIONS TO BE
CO-ORDINATED
-------
APPENDIX V-4 !
Page 130
VARIABLE HVDRAUUC
THRUSTER
iOUTPUT SHAFT
STEER CONTROL
STOP
CRATO LIMITER)
LOUTPUT TOROO
-TORQUE CONVERTER
-------
APPENDIX V-5 HYDROMECHANICAL TRANSMISSION COMPONENT SIZING
SHEET I OF 3
TR1- MODE HNDROMECHANMCAL TRANSMISSION/ DATE
- AEROTET RANJVONJE £>OO'ME APPLICATION! BY
SCHEMATIC SUMMARY
(AA*. ToSO
80.5 fT.LS.
SPEEO 376o -
RPM
NVAy.
8O.S PT- i-S-
SPEED ± SlSO RPM
Page 131
I
Form G 7716
-------
APPENDIX V-5 (continued)
SHEET_2._OF.
DATE
FoR AE(^O^£T RAtOvCirJE
.xJfcMT SiZiM<5
Page 132
I
CLUTCHES
• S-/NJCU&ONJOOS si4(PTi/0£ - sizj~£> rsv MAX:
STATIC TO&QUE, ^°T E^EJ^CV
» MAXMt'M CLUTCH Liuiisjc OK»«T
«- SSO PS>
CAPAC-iTV AT MAX.
— 36 7o MODE: | CuUTC_H a
80 7o HODE. 2 AJvJb 3
t MAxrljv\uM S^i-EjCi.R. STfl£.SS "Dots M<3T"
MAX-
3.
• PACE U«I2T>4 S\z.Ejj R,-/ -rue. CREATOR OF
TOO CRtTE-RtA —
(0 MAXIMUM AUUOOA.G,I_£. AC,MA.
C,£>J^IM^ STRESS OP 120,000
Form G 7756
-------
APPENDIX V-5 (continued)
SHEET 3_OF.
DATE
BY
4.
• MEAJM LOA4D OS£D ^R
ASSUMED To BS '/S OP
O^D . ('N/OTC./ TH~E. MEAM
To TUe TRAMS . J)2.£S 0££Ji TO
A 3SOO HOO/IS
Page 133
|
Form G 7756
-------
APPENDIX V-6 TRACTION DRIVE TRANSMISSION COMPONENT SIZING
TRA<1TIONJ
ONI
SHEET__L_OF.
DATE
BY
SCVLEMATK
MAX TORQ
71 FT-LB.
SPEED
214-0 -14140
Page 134
I
rm G T1 .
-------
APPENDIX V-6 (continued)
SNFFT 2- OF
DATE
BY
TO
Page 135
I
Form G 7756
-------
(SPACER PAGE - INTENTIONALLY BLANK)
Page 136
Sundstrand Aviation
-------
APPENDIX VI-1. HYDROMECHANICAL TRANSMISSION/COMPUTER PERFORMANCE
PROGRAM
The purpose of this program was to determine transmission and vehicle performance of a simulated
vehicle with a hydromechanical transmission for a given duty cycle. The program also computes
speeds, torques, horsepower, working pressure in the hydraulic units, power losses, and efficiency.
The required input parameters are listed below:
Vehicle parameters
Engine specification
Fuel consumption data
Accessory specification
Duty cycle definition
Planetary dimensions
Gear ratios and efficiency
Hydraulic unit displacement
Charge pump pressure range
The program is a discrete simulation which calculates the conditions necessary within the system to
obtain the desired response. For each duty cycle condition, the program calculates the required
data. The following paragraphs explain in detail how the program works.
For each incremental time interval the vehicle speed, acceleration, drag, and tractive effort are
defined by given input data or predetermined duty cycles available as subroutines. An initial
estimate of required engine speed is made, and a check is made to be sure that the estimate is
greater than the minimum possible engine speed for the vehicle speed in question.
Next, the speeds of the various components of the system are calculated. Since the transmission is a
multi-mode transmission, a check is made to determine which mode is correct for the calculated
speed conditions.
After the speeds have been determined, the torques and horsepower in the various transmission
elements are calculated. Horsepower loss in the hydraulic units is calculated by a seoarate
subroutine. Torque and horsepower are found by a trial and error procedure. A working pressure
(which controls system torque) is assumed and hydraulic unit losses are calculated. Then the
equations of dynamic equilibrium are solved to find the unknown torques in the system. The
working pressure is then recalculated. If the recalculated working pressure differs by more than 10
psi from the assumed working pressure, the assumed working pressure is modified, and the whole
process is repeated until it iterates to a solution.
The horsepower flow in the elements of the transmission are found from the torques and speeds
determined above.
Then the other power losses that occur in the transmission are calculated. These losses include:
charge pump power requirements, planetary gearset losses, transfer gear losses, and open clutch
power loss.
Sundstrand Aviation
-------
The total power loss, the total power required from the engine to the transmission, and the
transmission efficiency are calculated. The total engine power includes the vehicle accessory power
requirements.
The calculated engine speed is compared with the engine speed that would yield the minimum
specific fuel consumption for the required power level. If the engine speeds do not correlate within
one percent, a new engine speed is assumed, and the process is repeated until it iterates to a
solution.
When a solution is achieved at minimum specific fuel consumption, engine speed, fuel consumption
and energy efficiency are calculated. Then the calculated output is printed, and the program goes on
to the next condition of the duty cycle.
When the driving cycle is completed, a summary of fuel consumption, energy efficiency, average
power requirements, distance, and time is printed.
A typical readout is included as part of this appendix. The example presented here is of the EPA car
at test weight. The engine is the Aerojet prototype configuration Rankine engine. The assumed
vehicle accessory load includes the air conditioner, but the accessory load at maximum engine speed
was scaled down as a function of maximum engine speed ratio because of the narrow operating
speed range of the engine (see Appendix 1-8).
The transmission featured in this computer run was the Sundstrand Tri-Mode hydromechanical
transmission.
The duty cycle assumed was the complete Combined Driving Cycle, which includes the Federal
Driving Cycle, the Simplified Suburban Route, and the Simplified Country Route.
The readout includes the input parameters for each driving cycle, the output parameters for each
driving cycle condition, and a summary of the results of each of the driving cycles. (Only the first
127 seconds and the last 135 seconds of the Federal Driving Cycle are shown here.) Also included at
the end of the readout is a summary of the combined duty cycle.
Page 138
Sundstrand Aviation
-------
APPENDIX VI -1 COMPUTER RUN - INPUT PARAMETERS
'UO. ECONOMY / T»AN«l5MnN PEPFPRCAKCE ANALYSIS
FOt 4 VEHICLE UMH * HYOP1»FCHANICAL TRANSH 55ICN
•-IT-7Z BUN 1.03 EP4 TUP9INE CAP FEDPAL DRIVING CYCLE
AEROJET RANKINE ENGINE TIT k/1.5 IN. LOG
INPUT PARAMETERS
VFHICl E Wf tr.HT ILP) ............................... hT- 4640.000
UKR I*E«TIA (FT-i.,_s6C2) ........................ i,,, U.2CO
FPDNTAL 4»FA (FT<| ................................ tf, 20.000
CnfFFICIF'lT 1C r,c«f, ............................... CP= 0.600
F-UFL DENSITY iLp/csLi ............................. at* *.2*o
ENGINE Clff ....................................... **'.' '.DOO
OUTY CVCIF Cn^S .................. . ................ CUTY* J.OOO
OUTPUT CODF ....................................... CUT
?.000
NO"OGR»PH
AT C
AT 85
0.4110
O.^'TO
1.0000
7.0000
1*.2000
ENGINE SPEFT / C>"MMF LINK SOFF" «l-
V-IINIT SO11 FT / V-UNIT II 'K "EEH '? =
MODE * F-ll'lT SPr^D / »nn= 1 F-(JMT LlK« S Pr t r,,...«?«
MOnf 2 F-PMT ^c>tcn / «pic p F_IJ.|T LIN' SPF=C....°4"
TRANS CUTOUT LPIK «»CE" /"'"'INS CUTOUT SPEE D. .'.'.'. ', Bf. UCOOO
TS AN S OUTPUT SPFFD / A >LE SPFFO FA* 4.5009
2.2000 O."FSH|
."E«"I
.950EFFI
HYD UNIT P I ^^lar F»ENT ......CISP* 1.^00
WHFFL 0 MM F TC 3 t .... ; ............................. «CW* 2(>.400
CEAP "ESH fFFICIE'lCY .....FG' 0.>»*
MODE 2 PL A>.FT«av c«'5 !C1« LINK .....KAR2- 4.000
"nns 3 PIANCTARY c»JO'c' LINK ......7CA»i* i.ono
MAXI»U" ClUTCM lf«t .T'SFPn.ER HCL<* 1.^00
M|NI»UM CHARGE Pl'"P P°FS>E PCN« >>0.000
MAXIMUM CHARGE Pu"P PRESSURE FC»- 250.000
Page 139
Sundstrand Aviation
-------
APPENDIX VI -1 COMPUTER RUN (first 127 seconds)
Uf\.
» A
/ TTAMS"I«M1N •* •FORI'lNCf: ANALYSIS
VEKtClF WfW A M»f>oO»fCM»MC»L T»A«»1SSK>>
g-|7.7j BUN 1.0? FPA 1UP» IN? CAR FEORAL DRIVING CYCLE
AEROJET HtSKINf ENGINE Itl k/1.5 IN. ICC
OUTPUT PAPAXETC9S
ENGINE
TR AKSf I S S [TN
OOAO
FUEL
SPEfO ACCEl
fft- F/S/S
0.0 0.0
0.0 0.0
0.0 0.0
0.0 0.0
0.0 n.n
o.n o.i
0.0 0.0
0.0 0.0
o.o o.n
0.0 0.3
0.0 O.C
0.0 0.0
0.0 0.0
0.0 0.0
o.o o.n
0.0 0.0
O.C 0.0
O.C 0.0
0.0 0.0
0.4 2. If
2.0 4.33
5.<: 4.72
8.7 4.H
11.5 4.Q1
14.2 3. "if
16.6 ?.4?
17." 0."
IP. 2 t.7«.
)5>.« 2.64
22.1 0.5<=
22.4 -O.C7
22. C -0.71
Jl .5 -0.77
20.9 -O.«l
20.1 -!.»•;
19.7 -2.40
17.1 -7.C?
15." -1.^4
15.1 -0.55
15.2 0.44
15.6 O.»l
16.1 1.17
17.2 2.05
I*?*! 2 « *• "*
21.0 2.1i
22.0 1.32
22. P 0."
22.7 -0.27
22.4 -1.4'
20. • -2.f4
10.0 -2.49
17.4 -2.35
16.1 -0.48
16.7 I. 10
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Page 140
Sundstrand Aviation
-------
APPENDIX VI -1 COMPUTER RUN (first 127 seconds)
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Sundstrand Aviation
-------
APPENDIX VI -1 COMPUTER RUN (last 135 seconds)
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19-105. .47 *0.* .3 0.191 9*0. 77. 0 If
l',240. .45 ?o. •, .1 0.215 779. »1.O «
131. 0.92* I*. 27 11. «1 e
:. 11. 1.067 5.65 1 .S3 «
. -T. f 4". C.7f 7 24.70 2. <1 o
•4'. 0.767 76.5° '.10 f
'. 492. C.»C« 2«.7« 3.» f
>. 43:. 0.311 24.01 .1 «
:""-H. .37 25.9 .1 0.257 556. K5.2 •*. ^3>. 0.936 21.6'' .* c
Ia2ll. .11 21.4 ,7 0.27S 4l». =4.* 37. ;«7. C.fit 19.99 .4 »
177«.4. .24 20.0 5.9 0.?95 105. »».0 »
1702% .14 17,1 5.4 0.114 569. »2.0 1
l6<"Cn. .01 14.4 4.7 C. 340 394. 7<>.2 ?c
14COO. .01 11.2 3.7 0.140 291. 67.1 r
IfOCO. .00 6.3 2.7 C.333 202. *T.A *c
C. O.C 0.0 0.0 0.0 0. 0.0 •
0. 0.0 0.0 0.0 0.3 0. n.O »'
(1. 0.0 0.0 ,1.0 0.0 0. 0.0 5
C. 0.0 0.0 0.0 0.0 0. O.O 9(
?32. C.«ft 17.74 .5 o
'. 170. 0.51* 15. »4 .« »
12". 0.9't 13. 76 1 . r P
79. 1.017 1 1 . 3 » 1 .17 P
'. 32. 1.CJ4 S.C7 1 .2" »
. -i;. c.o ».« » i . 33 f
. -24. C.O 9.79 1 ,r7 e
'. -1*. 0.0 8.74 1 .76 '
. -59. 0.0 S.69 1 .43 •
\- _at ri alii n A
C. O.n 0.0 3.n 0.0 0. 0.0 "4. -111. C.O 9.56 13."4 «
c. n.o o.o o.o o.o o. o.o 9
o. o.c o.n n.n n.n 3. n.o 9
C. 0.0 0.0 0.0 0.1 0. O.O 9
C. O.C 0.0 0.0 0.3 0. 0.0 7
o. o.o o.o n.n n.n o. o.o 7
o. o.n o.o o.n o.o o. o.o ^
0. 0.0 0.0 0.0 0.3 0. O.O 7
'. -14C. 0.3 ".48 12. C6 t
f. -16=. 0.0 0.29 IJ. 12 •
). -199. o.o ».29 11. «5 f
!. -255. C.3 P. 17 1C. 6?- •
^. -•9-. 0.3 7.CJ 9. 7J a
•. -567. 0.3 '.69 6. 57 f
!. -541. 0.0 7.44 4.43 •
r. o.o n.o n.o o.o o. o.n 7i. -v-. c.o 7.23 2.34 e
C. 0.0 0.0 0.0 0.3 0. 0.0 71. -238. 0.0 7.11 1.10 8
41
42
42
43
43
43
44
41
41
43
4}
42
42
42
41
41
41
4 C
4C
40
IS
39
19
39
3 9
39
36
» 7
31
37
36
36
36
36
25
25
34
33
22
32
31
31
2 I
3C
20
31
1 I
21
M
22
32
13
> -.)
33
33
3 3
33
33
33
Sundstrand Aviation
Page 143
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APPENDIX VI -1 COMPUTER RUN (analysis)
FUEL ECONOMY / TPAN5-M SSI^ pF°Fr°'M>>rE «N»I
FOR * VEHICLE KITH l rtYOon"FrHANIC4l_T''»NS»
H-17-7? RUN 1.03 EP» TUP.9INF CAP FEn3»L
AEROJET P4NKINF ENf.INf 1X1 k/1.5 I N. I rr-
, CYCLE
me AH ""'P KNr i>;r D»lvl'".i VMS.IIV'N* 2'}^
Fna ALL -ill>n ivFMitLE cnis'iw.l CONOITICNS a.2l>
MPG nVEBJll FOP ALL CONDITIONS
ENGINE OB IV INC MODE
TnTAL WhEELS-TP-»OAO ENFOr.Y.HP. 5EC
TOTAL ENGINE ENfBf.Y < I NCL. ACCE S < .) . HP. SEC
4VC. WHEELS TP BOAf) HP ij'5?
»VG. ENGINF t-P I INCL.SCCE SSI 22.55
TRANSMISSION ENERGY EFFICIENCY 78.205
I T.'.VSHl. 1367.SECI
ENGINE COASTING "HOE
TDTM. «0*0-TO-h«EElS EN£»C T,H». S£C
»VC. PlO»C-TO-tiHEELS HP
-JTT*.9»
-T.JS
FUEL ECPN1«Y / TOA*?*! SSI "" »FPFn<>»AM"F ANALYSIS
tr\* A VEHICLE WITH * H>oac»ECHANiciL T»AKS»ISS ICS
9-17-72 PUN 1.03 EP4 TUfJlNE CAB SI'FUFIEO SUSUKAN (CUTE
AEOOJET BANK IKE FNGINE T*I wi.5 IN. LOC
INPUT PARAMETERS
VEHICLE MEir.MT UCI MT. *60n.i5f>0
Mt-FEL |fE"TIA (F T-il-SEC2I M» 11 . 200
CTFFICIF'JT OF r»»r.II "m.T.'" Ill lit I""I '.'. '.'.'.'.'. Tt"' ollSfl!)
ENGINE c'"aE.......... I" I! I! i; II11II I!'. I'. li.'I I'l^ik-,* 'I ico
OIITY cYri f cnnc ,...:JTY- i.ooo
OUTPUT CnOE CUT- 2.000
NO»dftPAPH niui'i 'IHN!! I'.'.'.'.".'.'.'.'.'.'.'.'. ',...'.'.'.'.'.'.'.'.'. TJ= oTf-'oo
NO"OCDAPH CI'ENSIl)*' t* 1.0030
UB/t-B AT C*"f"Il".'.I limit II II ! ! II'.I I'.! .'. .'. ! ?3H\ i 7loOOO
LB/HR AT 85""*' FP"2 « 1
ENGINE ^"EECl / ENGIV LIN» <;=ctn 01,
"OOE 3 F-OviT SffFn / -'-IF > f-'_','\ '-' l_] i«'?cFF?I^ loi,
MPQF 2 F-UMT cPfFr / Mr - ? F-UMT LlK< Ssfc"^....^^»
MPDF 1 F-UMT SPffn / "ODE 1 F-L'IT lf^K Sc = cc....csa '*^''6^7
TRANS PUT»UT I !'.< S'EF" / T3-.NS TUTPUT ^i>EE3 '*= l.COOO
TRANS OUTPUT SPEED / AXLS ?PCEO SA- 4.5000 (6.950EFFI
f-YO UNIT Of«>l4CE"ENT CI5P. 1.500
MME?L OIA"ETEP ;«» 2(>.'.ir)
GFAR MfSH f EF I', It ',C" c"= 0.4°5
HA*|>'U« CLUTCH LTS5 HIO<;F ^"l.F5" ""!!'! II"I*.III ' FLCLX» l!500
MINIMU" C.ff'if Pt"» P5CS£(.5F cr\, 6O.OOO
MAXIMUM CHARGE PU«P PRESSURE FCX» 250.010
FUEL fCPN1"Y / TO/IN '«| |SS |TN
S-17-72 RUN 1.03 EpA TO»8INE CAR SI'HIFIED SJ8URBAN ROUTE
AEROJET RANKINE ENOINE T»T »/i.s IN. ICG
OUTPUT PARAMETER'
VEHICLE ENOI^'F TSAMMfSSrON «P40 FUEL
?8:8 8:8 |2S;J i?S;S ?f^S: l??^c: t:?8 19:? ,»:: 8:2« 5«: ZS:T ,55: igS; 4-?iS H-i, H-3? JJ-2?
4.27 22.3 6.5 0.5*8 339. 7«,5 126. 126. O^K 1<>'.19 li'.OI 12°.2J
Page 144
Sundstrand Aviation B
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APPENDIX VI-1 COMPUTER RUN (analysis)
FUF\. ECONOMY / TRANSMISSION PEPFPPHNrF ANALYSIS
fO» » VEHICLE WITH A Hincr.MFcmNiCAL Ta JN
8-17-72 »UN 1.03 EPA TURBINE CAB SfHIFIEO SUBURBAN ROUTE
AEROJET RANKINE ENGINE TIT w/1.5 IN. LOG
I«HF< DEa r.mLriN , »IL!=<; . SECCM1S
F0» ALL »HOn l?Nr,INE DRIVlNf.) C?NO! "I CSS 12.228 I 7.486*1, 898.SFC)
I«PG F0« MA -HPO (VEHICLE COASTING) CONDI TICNS 0.0 I olo MU
»M>G OVERALL FOR ALL CTNOITIONS 12.228 t T.486MI, 898.SEO
ENGINE DRIVING "ODE
' TOTAL Wt-FFLS-Tn-snAO FNFar.Y.HP.SEC 7875.40
TOTAL ENGINE ENERGY I I NCL .ACCE S5. I ,HP. SEC
AVC. WKEIS TP POAO HP 8.77
AVC. ENGINE hP CINCL.ACCESS) 16.88
TRANSMISSION ENERGY EFFICIENCY 72.42*
ENGINE COASTING MODE
TOTAl RPAC-TO-wt-fELS ENERGY.HP.SEC 0.0
AVC. ROAC-TO-WHEELS HP 0.0
FUEL ECTNO-Y / TBSNSHI R»»KrE ANALYSIS
FOR * VEHICLE WITH A HYOROKfCHANICAl. TPANS»ISS1CN
8-17-72 RUN l.OJ EPA TURBINE CAP SI»FLIFIED CCUNTRY RCUTE
46ROJET RANKINE ENf.INE TMT K/1.5 F N. ICG
'INPUT PARAfFTFRS
VEHICLE wflCHT (IB) .............................. .WT« 4600.000
wkEEL INERTIA IFT-LI-SECJ ......................... wi> u.zoo
fBfMTAL AREA (FTJl ................................ if, 21. .100
fnFFFKIENT OF T»«G ............................... CD- O.tOO
FUFL DENSITY (LB/r-AL) ............................. :EN« "-.z^o
FNr.lNE ClfF ....................................... E^".- J.)00
DUTY CYCLE COCE ................................... C )TY« 1.000
OUTPUT CODE ....................................... CUT» 2.300
NCWR'PH n^EN^ION ............................... t- 0.^100
NOror.RAPH ri^EN'in-i ............................... p« o.fjno
Nn»ocRA»H ni-ENSir.N ............................... c= i.ccoo
NFMAX/N^xtx ................................... ^^iTF= ».'0?l
L»/HR AT c "P" ............................... P»M1 7.CCOO
LB/HR AT 85 "PH ............................... fPHZ » 14.2000
ENCINE S»cfO / F'-IINF LINK ?»Fcn .................. M* o.41»4 IO."ESH)
V-UNfT SPfE} / v-fJIT lt.'.< rpejo .................. C2* 2.^157 (?.••>
"OPE 3 F-UMT SP'Cj / Mnr,c •) F-uMT |_tN« S fc £"...." 3 = 2.20T1 (1.«'S»-)
»n?E 2 F-UMT SPCFH / "C':E ? c-UMT LISK <. ft -.~i . ...»'•' 2.20JO I3.« = SH1
>«nnE 1 F-UNIT ^PfEO / "OOF 1 F-uMT LIN< S Fr c 1. .. . «5» 7.3^"-T l?.«E^f»
TRANS PUTPUT Ll'i" SPF.ED / TS'.N^ PUT'UT S»EET ...... =4= l.f";00 IO."ESH|
THANS OUTPUT SPfFD /AXLE JPFEn ................... SA« 4.5000 (0.950EFFI
HYD UNIT rUSPLACf "E'lT .................. . .......... C I S P- 1 . ">0 0
WHF.EL ClAMEtrc .................................... C'*» 26.400
GEAR «ESH Ecc!rir"C>' ........................ . ..... £'">* O.oq^
"OOE 2 PLAf:cTi5Y C'^0 15° L I •;« ..................... 7Ca=2= 4..1TO
"OOF 3 PLAN?T;«Y rflsojcp \_\\n ..................... :c»"1= 1.1"0
CLUTCH LT'S MF.O^E POv.CI< .................... HCL«= l.SOO
CfiA^r.f PI,"" PUCS'LB? ...................... FCN» >O.OCO
CHANGE PLUP PRESSURE ............... .......fcx« 250.000
FUFl FCPN1MY / TC4NS«I'MTN PEPFm>»AM;F ANALYSIJ
FOR * VEHICLE WITH » HYnRn»ECHANICAt TPSNSUSS I C 1
8-17-72 RUN 1.03 EPA TURBINE CAR SI^FIIFIEO CCUIVTRY RCUTE
AEROJET RANKISE ENGINE tMT k/1.5 IN. ICG
OUTPUT PARAMETERS
VEHICLE F.W5IM6 TH»SS»ISSICN _0^C _
" """ CRAG TRACT L«/ L / -P*-. «r
"_
21
cCfn»FT| nsT sPFCcsTTL 5 FoLTATrH A A L« L .
F/S/S INC CU" FT RP* HP »P FT-LS ("AT 1C P'Fs t'f^ _Lf __ L^ _ ..." ___ "* __ ':..-.-
Mn n n ICA - t * t. i IIACT 1C7CJ 4 4S 7Q.-S P.I 0«^3q 44Q. S'J.S 1 ' *> • !'*!• 0.6CJ ?3»77 13.21 13.
»,n*n n'S UtM 5f?'? U5nZ" jns%7 4*3 3<» 7 10^2 O.?20 I**. "1.7 1
-------
APPENDIX VI -1 COMPUTER RUN (analysis)
FUEL SCANTY / T6AN?«I5MCN l>fpcnp»«KCf *?*VTI!5i,
FOR * VEHICLE WITH A H>Ofn"ECH4NICAL TRASS»ISSie*
8-17-72 RUN l.OJ EPA TUPRINE CAP SIMPLIFIED COUNTRY ROUTE
AEROJET RANKINE ENGINE 1MT k/1. 5 IN. IOC
PERFORMANCE SUHMADY I»HF5 PE« GALL"N , "I 1= 5 , fECTKOJ ,,,„, 4A
H»c FP" »U «HT ifsr.iNf osivlNCi r.nvniTiCNS 12-P* J I-!11!!!1 *n fri
f-fG FOR ALL -HPO (VEHICLE C045TINGI CCA1I TICKS 0.0 I 0.0 tit O.SECI
WPG OVERALL FP« ALL cnNOinoNS iz.w* i T.etiMc. 4««.sfC)
ENGINE DRIVING "OCE
TOTAL WHFElS-TO-OD«n ENEROY.HP.5EC }S!i?'5i
TOTAL ENGINE ENHGY I INCL . AdCE 55. | tHP. SEC 195*6.58
»vc. WMFFLS TO on^n HP H*S1
AVC. ENGINE t-f I INCL.ACCES5I 41.TO
TRANSMISSION ENERGY EFFICIENCY 49.T35
ENCINE COASTING MODE
TOTAL ROAO-TO-Kt-EELS ENERGY.HP. SEC 0.0
AVC. ROAC-TO-WMEELS HP 0.0
FUEL ECONOMY / nuNSxij^tcN otcfrwif.re ANALYSIS
FOR A VEHICLE klTf A MYDOOCFCM A»:i C U TB1NS»I SS ICN
CC«efNEO FEQRAL OPIVING C YCLE , ?l »»LIF I EO SU9UPBAN CYCLE, 4kC SIMPLIFIEC COUNTRY CYCLE
PERFOPMANCE SU»M«OY IMHES pf « GALLON . »IL?S . SECCKCS
I>PG FTR AIL «MPO (ENGINE OBIVINGI C^NTt TICKS in. 341 ( 21.274MI. 2219. <PC FOR ALL -MPP IVEHJCLE COASTING! CCNDITICNS a.213 I 1. 47311, 51*. SEC)
KPC OVERALL FPR ALL CONOITICN; 10.621 I 22.7*6X1, 273*. SEC)
ENGINE PR IV ING NODE
TPT*. WKFELS-Tn-RnAO ENFKGY.MP. SEC 3*271. <»8
TOTAL ENGINE ENERGY I INCL .ACCE S 5. I ,MP. SEC 5392*. *7
AVC. t^EELS Tp PnAO HP 15.**
AVC. ENGINE H> ( INCL .ACC E 5 SI 2*. 30
TRANSMISSION ENERGY EFFICIENCY «1.227
ENCINE COASTING "GTE
TOTAL HOAO-TO-WHEElS ENERGY.HP. SEC -377*. ««
«VC. ROAC-TO-WHEEL S HP -7.33
Page 146
Sundstrand Aviation
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APPENDIX VI -2
ZB32 - VEHICLE PERFORMANCE PROGRAM,
TORQUE CONVERTER AND TRACTION
DRIVE TRANSMISSIONS
Program Language: Fortran IV
Purpose: To determine vehicle performance (Including fuel
consumption) for any given vehicle with a traction drive-torque
converter, or automatic shifting gearbox-torque converter type
transmission and with any given engine for (I) any given vehicle
output specified driving cyle, or (II) a standing start acceleration
run under the application of the given engine output.
Required Input
Environment parameters: Air temperature, road grade
Vehicle parameters: Weight, drive wheel radius, and total
wheel inertia, frontal area and aerodynamic drag factor.
(Rolling resistance factors built into program)
Engine parameters: Engine HP versus specific fuel consump-
tion map, desired engine speed versus engine HP operating
curve, maximum engine speed, torque and power, "closed
throttle" fuel consumption at idle and some other engine
speed, fuel density, engine speed versus vehicle accessory
HP curve, and engine Inertia.
Transmission parameters: Transmission and drive line gear
ratios and efficiencies, traction drive unit ratio range,
reference torque converter characteristics, and required
torque converter diameter. (Traction drive efficiency
computed in a separate sub-routine)
Driving cycle: Federal driving cycle, simplified suburban
and country routes specified in separate sub-routines.
Output conditions can also be specified point by point on
punched cards in terms of vehicle speed, acceleration and
and time increment, or in terms of required tractive effort,
speed, and time Increment.
Page 147
Sundstrand Aviation
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Page -2-
Drivlng cycle: (continued)
If the output performance conditions are not specified,
then they will be computed for the case of a vehicle starting
from rest under the application of (I) the given engine
speed-torque curve in the case of a fixed ratio-torque converter
transmission or (II) the given engine at its maximum power
point for a traction drive-converter transmission.
Computed Output
The program initially prints out some transmission limiting
parameters. Such as conditions at the maximum power stall
point, maximum creep speed, and some maximum system speeds.
The program then takes the vehicle through each point in the
requested driving cycle, printing out various speed, torques,
ratios, powers etc. in the engine-drive line system, as well
as speed, time, distance and instantaneous and cumulative fuel
consumption. At completion, a summary for the course, and
the sections within the course are printed out.
Operation
Following is a brief general description of how the program
operates.
Prom the given vehicle speed and acceleration (or tractive
effort) for each given driving cycle point, the program
calculates the required driveline speeds and torques from the
road wheels to the torque converter output shaft. Knowing
the torque converter size and characteristics, the torque
converter input speed and torque can be calculated,and thus,
the output conditons at the traction drive unit.
The engine speed is determined by the power required to satisfy
the particular duty cycle point and the given engine power
versus engine speed operating curve. Thus, the speed to the
traction drive input can be determined. The required
instantaneous traction drive ratio is then determined from
the required input and output speeds.
Page 148
Sundstrand Aviation
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Page -3-
Operation: (continued)
This ratio must be within the ratio ranee specified as input
data. If this range is exceeded, the unit will go to its
maximum ratio and the engine will be run at some speed off
its required operating line. (Part of the system optimization
is to choose a ratio range that coincides very closely with
that required by all the vehicle performance limits.)
The traction drive efficiency is calculated as a function of
the traction drive input torque, which means the input speed
must be determined by an iterative process of first assuming
an efficiency, calculating the required engine power, and
traction drive input conditions and then actual efficiency.
This is repeated until the actual and assumed efficiencies
are sufficiently close to each other.
Instantaneous fuel consumption is computed for each point
from the calculated engine power and speed and the given
specific fuel consumption map.
Time, distance, fuel consumption and energy are accumulated
through the duty cycle for the completed driving cycle
performance summary printout.
If a duty cycle is not given, the problem is worked in reverse,
that is,instead of calculating the required engine and
transmission conditions to give the specified vehicle speed
and acceleration, the speed and acceleration will be computed
with the specified engine output.
Sample Output
Following is a sample printout of the program, giving the
first 69 seconds, and the last 33 seconds of the Federal
Driving Cycle, the Simplified Suburban and Country Routes,
and the resultant combined course summary for the traction
drive-torque converter transmission and Aerojet Rankine cycle
engine.
Page 149
>"v.
Sundstrand Aviation
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APPENDIX VI -2 FIRST 69 SECONDS COMPUTER RUN
F
VEHICLE ....... E"a TiiKiINc CArf C"«Hl:i 10'J. . TiJirT | r, iv 1 /.- / r IR.'JE CQ
'
.
CnNvtklf.R ..... Fij30 j.r.l S.l'.H.
INPUT DATA
VEHICLE TAKE, OVER EpA COMPOUND ORIVINC CYCLE
DIAMETER OF CONVERTER/COUPLING T3 BE SCALED ............. 6.12 INS.
VEHICLE WEIGHT ........................................... 4500.
DRIVE KHEEL «AOIUS,FT .................................... 1.100
fQTAL HMEEL INERTIA, SLUG-FT-M ........................... 11.20
FRONTAL AREA, SO. FT ....................................... 2J.OO
MR TErP.,OEG.F .......................................... 85.0
AERODYNAMIC DRAG FACTOR ......................... . ........ 0.600
»0»0 GKApe»PtRCENT ....................................... O.O
H.I.T. ROLLING RESISTANCE FORMULA?. USED
ll30-B32/T(ia'JUi: CHNvfxTfR-FLU I 0 COUPLING SIZIMf. AI.O PERFORMANCE ANALYSIS
VEHICLE PFRI-JRMANCE vtiilU'i REVISION F
VEHICLE F.PA Tij-IKI-lf Ci^ C'.wdr.EC DUTY CYCLE
ENGl'l? ..AJH-MrT ^.V.nlr.5 E-jCl"?
CONVERTER..I!.FOSO J.-M S.
.22000.
ACCESSARY HP
*.00 4.23 ».55 *.rt2
IP Fps I
, ., - I n L e >' - -
101 E FL'F.l R.1. I E»l;
..... . .
{Ncifif smo NJ,RPH ...................................... 22 010. r
FUCL «ATc AT N2.LD/H-1 .................................... 12.0JO
FUEL OE:.SITY,LB/GAL ...................................... S.^8
TRtCTITi. OSIVF P.ATm 4i"4i,i 5.000
HAX.ENI.I-'F T rfOuE, FT . I F- > 8 .«.
HAi.t-M. !•.£ H 170.1
HAX.fcNO |r,6 i<.'-l 22000.C
CEAR Rftrig^.-T.URfVf >;<';3»'-G4 2.750 l.CC.'' l.fOO 1J!.7»0
tf F1C IE.'.ClFS.eol,EG2,E( 3,cG4 1.000 l.OOu 0.9S5 0.950
DESMEl' ENGI. E HP-SPLET ..il'ff'.iTIJN
ENGI'ifc HP
li. 14. 17. 20. 25. 30. 40. 50. 55. 60. 65. 70. 71.
ENGI'.C RP •
160JO. 16000. 16">30. 17670. 18570. 19400. 2f)74(>. 21530. 21790. 21930. 21990. 22000. 22000.
12.00 I1'. RfFFKE'lCF CruVERTtR/CTUPLI'iG «1S GIVF'. As II.PJT DATA
0.0 O.'OO 0.3CO U.J'.'O 0.400 0.4SO n.5ir> C'."i50 0.600 O.HC O.t'JO 0.'5O
O.BOO 'l.rtSi) 0.8tO 0.''i.0 0.9213 O.T.O P.15" 0.'60 0.970 0.9HO 0.9VO 0.'^92
TURC'jf RS'KI
J.CIU 2.6-10 ?.lc'0 2.010 1.300 |.6-»0 l.Sbr 1.5UO 1.370 1.3^) 1 .'00 1.120
1.030 I.''"" i.i-n LOU-) I.ooo 1.0)0 l.'jco l.ooo i.ooo t.c"> '.oor t.doo
CAPACITY r,'.CT:n.»
196. loJ. I".;. 1V2. 167. 165. 164. 1*5. 167. 16S. 172. l~6.
183. 196. iji,. 21-.. 225. 25^. 27ti. 3^0. 400. dO.l. 2000. 4000.
6.12 INCH COrr.ERTF«/COuPU-iG SCALEP FROH 12.00 INCH *EFE»E')CE UNIT
SPEED RATIO"
O.C O.ICO !>.."••? 0.3 O 0.410 0.41>n P.'f O.551 O.f,c0 0.650 0.70D 0.'5C
0.8OO 0.^50 0.^*'0 u.^ 'J 0.920 0.^4C 0.95f' 0.^*60 0.970 0.960 0.9-*o O.ct*2
TU»*0 l.5^" 1.5UO 1.370 1.3"*" :.?.00 1.120
l.Oij 1.000 l.O'u i.HUO I.OOO l.ooo !.0(.'0 l.">.'0 1.UOO 1.&.10 I.OOO l.tOO
CAPAC IT Y t dCTT*.*
1055. 1009. (66. 12J. J?7. BRr,. drtj. tB6. 897. 90*. 973. 9sB.
4*5. IOS5: 1-J77. 1152. 1211. 1340. M-»V. I7;j. 2153. 43?7.107*7:21534.
Page 150
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APPENDIX VI -2 FIRST 69 SECONDS OF COMPUTER RUN (continued)
iuo-f»32. ruHuuE ca-K'iaTE" - FLUID
VCHClE
F.OA Ti'-i.U'.. Cir
DUTY CtClE
..... .»,
51'. 'I..TH .-.en -: . )- lv--/ • •» ,'JE CG!*.vt»T?»
C» ..... FGUD >.vO i.T.H.
C»lCUL»T£0 C'EEP SPfeO/MPH
"VHP.? SI
I'.ul «E l-1>
"
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Sundstrand Aviation
Page 15,
-------
APPENDIX VI -2 FIRST 69 SECONDS OF COMPUTER RUN (continued)
VEHICLE [
SPEED
ACCEL
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Page 152
Sundstrand Aviation
-------
APPENDIX VI -2 FIRST 69 SECONDS OF COMPUTER RUN (continued)
VEHICLE
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Sundstrand Aviation
^» W ,
Page 153
-------
APPENDIX VI -2 LAST 33 SECONDS COMPUTER RUN
VEHICLE
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Page 154
Sundstrand Aviation
-------
APPENDIX VI -2 LAST 33 SECONDS OF COMPUTER RUN (continued)
VEHICLE
f-Prt |i,^
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Sundstrand Aviation ^'
Page 155
-------
APPENDIX VI-2 SIMPLIFIED SUBURBAN ROUTE,
SIMPLIFIED COUNTRY ROUTE,
SUMMARY OF COMBINED DUTY CYCLE
VEHICLE
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1130-B32» T'jRiJUE CO IVfPTfH - HLIO CDUPLII-0 SIZl'.G fij Pf-FHRMiVMCf ANALYSIS
MtVISi
VEHICLE f->* u'-'i-c ci r •«tit.eo ouiv CVCLE
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Page 156
Sundstrand Aviation
-------
APPENDIX Vi-3. VEICLE PERFORMANCE WITH A TYPICAL 3 SPEED AUTOMATIC
TRANSMISSION
Vehicle performance was computed for the vehicle specified in the "Prototype Vehicle Performance
Specification" (Apendix I) using data for a typical 350 cubic inch displacement internal
combustion engine and a typical 3 speed automatic transmission supplied by EPA under Phase I of
this contract.
A. Applicable Data
1) Engine power and SFC curve (Figure VI-3A).
2) Vehicle drag and resistance forces (see "Prototype Vehicle Performance Specification,"
Appendix I.
3) Remaining transmission and vehicle data, see Figures VI-3B and Figure VI-3C.
4) "Typical" 3-speed automatic transmission, vehicle vs. transmission efficiency curves, see
Figures VI-3D and VI-3E.
Page 157
Sundstrand Aviation
-------
y i
H-
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I >
(A
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; ; i .: j i I ' ! i
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3600
-------
Appendix VI-3 Figure VI-3B
f. PWr- Tcr-C1 UI^ CriUPLINr; SmiV, a\0 peocnsw^CF
VEHICLE Pfc^P°=»-'i\CL: V-r. efm ----- • ----------------------------------- ^EVISIPN
l/FHCLr ....... Fill SI7ri: f A' < P~ " FP'J
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fFi»?M<:c If! ..2 SPF=" -'I'lr"1 MC f'.5 1ST,
..... 11. 7e !'SfU ,2.0 STu- (f>F5
ENGINE
INF cor
°oc. !cnr. ?'-rc. •'pnc
arccc jF3M ............. . ...................... o-)f).0
ICI F FM-L PATr ,La /K< ..................................... I'
FUR "if? AT N?,IR/^ .................................... °.000
FUEL D6^«!ITV,L3/r,AL ...................................... 'S.za
SUFT POINT TAT 3
"~~~ -»niVT< ................. 2
h° AT OFr""»:r
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SHIFT ar-iMT «PE ^P S , '-'PH ...... ^^.o r^.O
M". x SHIFT nnr-it ^-n-^ns ............. cn.o 75.0
t75 IN. »FFCpfr;C= r.T'i Vf- P TC h /C CU D I I N - V-AS GIVrK) iS INPUT T
SPEED
-------
Appendix VI -3 Figure VI -3C
H ->0- ?•*.?, TfUfMlf c
VEHCLE PEC -^c r«."
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WHFEI IMFFTI/1 ,rl(jr.-FT-F T 11.20
SCA,?O.FT 20.00
1C CRA'*, FACTPQ O.f-00
POAC G0Ar)E»PcCCENT..... 0.0
P
-------
a
(A
I?
ti
Appendix VI-3D •'Typical*• 3 Speed Automatic Transmission
(Vehicle Speed versus Transmission Efficiency)
-------
— MAXIMUM ACCELERATION
CONSTANT SPEED C UlSE
20
30 40 50
VEHICLE SPEED IMPH)
60
70
80
Appendix VI-3E Transmission Efficiency vs. Vehicle Speed
••Typical" 3 Speed Automatic Transmission
Page 162
Sundstrand Aviation
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Appendix VI-3
B. Performance Summary for the •'Typical 3-Speed Automatic Transmission*•
Idle Creep Speed
Accel.
Time to 60 MPH
Dist. in 10 Sec.
Time 25 * 70 MPH
Time 50 + 80 MPH
(D.O.T. HI.-SPD. PASS)
Dist. 50 * 80 MPH
(D.O.T. HI -SPD PASS)
Grade Velocity
30%
5%
0%
Fuel Consumption
Fed. Driv. Cycle
Simplified Suburban Route
Simplified Country Route
Combined EPA Driv. Cycle
Weight
©
0
©
©
Q
©
©
©
©
©
©
©
©
Performance
Air Cond.
On
18 @
11. 74 sec.
460 ft.
12.69 sec.
12.0 sec.
1150ft.
19mph
84 mph
115 mph
11.94mpg
17.86 mpg
15.99mpg
14.85 mpg
Air Cond,
Off
—
—
••w
••f
—
—
—
i"»
—
12,60 mpg
19.38 mpg
17.14 mpg
15.87 mpg
Vehicle Weight 4600 Ib.
Vehicle Weight 5300 Ib.
In Third Gear
Total Vehicles Accessory Power with Air Conditioner on - 4.0 HP at min eng. speed (800
RPM), Linear to 15.0 HP at max. eng. speed (4800 RPM) - with Air Conditioner Off
2.0 HP at min. eng. speed, Linear to 5.0 HP at max. eng. speed.
Atmospheric Conditions, 85°F, 14.7 PSIA
Ptge163
Sundstrand Aviation
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(SPACER PAGE - INTENTIONALLY BLANK)
Page 164
Sundstrand Aviation
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VIM Control System Parameters
a — Governor Valve Porting Area Coefficient = 2 x 0.078 In.
A- - Governor Valve Spool Area = 0.1963 In2
Ay — Variable Unit Control Piston Area = 6.52 In2
B^Bn-Bulk Modulus of Type A Hydraulic Fluid
kQ — Porting Area Flow Coefficient
Temperature
°F
0
100
150
200
250
0
100
150
200
250
200
Pressure PSI
0
0
0
0
0
3000
3000
3000
3000
3000
6000
B-PSI
3.0 x 105
2.2 x 105
1.8 x 105
1.5 x 105
1.23 x105
3.6 x 105
2.7 x 105
2.3 x105
2.0 x 105
1.7 x 105
2.4 x 105
ko
91
93
94
95
96
91
93
94
95
96
95
B9
cv
DC
— Governor Valve Damping Coefficient = 0.0449 Ib/in/sec
— Variable Unit Control Damping Coefficient = 0.006 Ib/in/sec
- Fixed Unit Displacement (10 in3/rev) = 1.59 in3/rad.
— Variable Unit Displacement per Unit Stroke = 1.275 in2/rad.
o
- Polar Moment of Inertia of Engine = 4.0 in-lb-sec*
- Polar Moment of Inertia of Vehicle Reflected to Sun Gear of Transmission = 381 in-lb-sec2
— Governor Valve Spring Coefficient = 47.8 Ib/in
- Variable Unit Control Piston Spring Coefficient = 60 Ib/in
Page 165
Sundstrand Aviation
d.ntion at Su"Q*(r«"0 Ccrporj
-------
VI1-1 Control System Parameters (continued)
Ke - Engine Governor Pressure Coefficient = 0.000741 PSI/(rad/sec)2
KR - Pressure Regulator Valve Coefficient = 2.04 PSI/PSI
L - Vari./Fixed Unit Leakage Coefficient = 0.00585 c.i.s./p.s.i.
Ly - Control Piston Leakage Coefficient = 0.0001 c.i.s./p.s.i.
M_ - Mass of Governor Valve Spool = 4.46 x 10"4 Ib/in/sec2
Mx - Effect Mass @ Control Piston = 0.02435 Ib/in/sec2
Rj - Input Gear Ratio (Nv/Ne) = 0.8242
R2 -Gear Ratio (NF/NS) = 1.975
V = V - Volume of Fluid In Control Circuit = 12 in3
Page 166
Sundstrand Aviation fOk
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VI1-2 Equations
1. Vehicle Velocity, MPH
MPH = 0.04996 W$ m.p.h.
2. Angular Velocity of Sun Gear, Ws
Ws = 1/JSJ(R2TF - TL - TAR - TRR) dt rad/sec
3. Fixed Unit Torque, TF
TF = DF Pw in-lb
4. System Working Pressure, PW
vXyWv-DFR2Ws-LPw)dt p.s.i.
5. Variable Unit Angular Velocity. Wy.
W = R« W0 rad/sec
V 16
6. Variable Unit Control Stroke, Xy
\ = 1/mM(-CV Xv - kvXv + AVPC - Fw) dt dt INCHES
7. Control Pressure, PC-
Pc = Pl-P2 p.s.i.
where:
1 IjiOr* 1 oO 1 Q VV VC
P2 = B/VjA - a2k^P2- Pd + a4 k^Pp, - P2 + AVXV + LVPC) dt f
8. Governor Valve porting areas, a^, a2, a-r & a^
a1 =a2& a3 = a4 = a Xg IN.2
9. Regulated Pressure, PR
P — i^ p n c i
where:
Page 167
Sundstrand Aviation £o».£.
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VI1-2 Equations (continued)
10. Governor Valve Spool Position, Xg.
Xg = 1/mgjj(- Bg Xg - kg Xg + Ag PN£ - FpL - kg Xjn) dt dt INCHES
where:
Xjn = 0 to 0.5 in., Throttle input.
11. Engine Angular Velocity, We
we = 1/Je f (Te ~ R1 V dt irvlb-
12. Variable Unit Torque, Tv.
TV = DVXVPW In-Lb.
13. Engine Torque, Te
Te = f (We/ MPH. Xjn) In-Lb
14. Variable Unit Wobbler Force, FW
Fw = 0.045 Av Pw Lb.
Page 168
Sundstrand Aviation
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VII-3 TUBBINE TOBQUE VS. TURBINE SPEED - RANKINE ENGINE
f _i ., .i-i. * I ..,
rlltt.tr-;:
t ! I ! !--!—«-
-4—t- t t . .—t • t-
.-l fc . -i_l..._i.
. .ij-i.^i i.--1
--jTl.l-l.4-Ij
w
u-4-;.i"4-P
>
G
»• V -
n:/-l\
P-U • i: , .
•i MU:,
-i 4
tft-mitfi
•fi ' ! i
---—--
i i i--i--t I I I
Tt-j-4-l-H
.f > ! • !-»-••
i I , • ; ! •
/ooo
J'0<>
Soeo
Page 169
Sundstrand Aviation £»J-
dxiMO" at 5jfJiy*'4 wi-s--«t^- ^B Hf ^
-------
VII-4 DIGITAL PROGRAM FOR FUNCTION GENERATION
SCALED FRACTION NXC2),NY(2),X,Y,FXY
CALL OS> . ...
CALL QSC( l.IFRFt)'
10 CALL O^HAP'SfX, i . 1 ,IrRR)
CALL QRnADS(Y,2,l.IFRR)
CALL XN(X,f!X,IRFR)
IF(IPFR.MF.l) HO TO 50
CALL YN(Y,NY,IR"R)
CALL FXvR(NX,flY,FXY)
CALL QWJHASCFXY,1, IFRR)
GO TO 11
50 TYPF 200,X,riXtIRFR
PAUSE in
GO TO IP
60 TYPE 201fY,NY,IRFG
200 FORMATUPH ARR NORM FRR X=,S7,5X,S7,5X,IS)
201 FORfATdSH ARR NORfl ERR Y=,S7,5X,S7,5X, 13)
PAUSE 20
GO TO 113
END
SUBROUTINE XN(IX,I MX,IFRFR)
SCALED FRACTION XT(9)
DATA XT(l)/.1779S/fXT(?.)/.?.093f5/,XT(3)/.5>4n7.~/,
1 XTC4)/.2721S/,XT(5)/.293S/,XT<6)/.3?445/,
CALL V3MS(XT(1),XT(9),XT(5))
RETURN
END
SUBROUTINE YM(IY,IMY,IFRFR)
SCALED FRACTION YT(4)
DATA YT(l)/.PS/,YT(2)/.3S/,YT(7>)/.fiS/tYT(4)/.9S/
CALL VnNS(YT(l),YT(4),YT(?>))
RETURN
EMD
SIJBROIJTP'F FXYR(INX,INV,
SCALED FRACTION FT(3«")
DATA F
1 F
2
DATA FT(in)/./^"S/,FT(ll)/.
2 FT(l^)/!
DATA FTU?)/.
1 FT(f??)/..7l75(:;?/tFT(?..'.)/..:
2 FT(?5)/.
DATA FT(2P)/.
I FTC3D/.3S
2
CALL FNRN?.(9,FT(l))'
RETURN
END
Page 170
Sundstrand Aviation ^1%
-------
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to
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w
50
a
M
o
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s
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-------
FIGURE VII-6 (1)
COMPUTER READOUT FOR A 0.2 UNIT THROTTLE
ACCELERATION AND 50% LOAD
F/T;'•//•'•''/' '-illll!ilHHI^''"":!1'
&^i/fi||ffljfffliiFaiz|
tfc
t.
-f-
_
-"
_
1-
/..
• 1"
•
[
./, .. /. .1.. ......
i-i ///,'•/•-
-..!-• ' J j ) ! J | -
T
V
i i
..4
\ ,,^4
r|
...1
,v
till!
f~~ — "•
!
1
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11 \
./,. ./. : /. u ... Lj-
.| / .u../ ..
+
-i- Li_! _! .j... :..j....L ; _.;_;_! .IJ J.._LJ
]. ,| ;. f'fA"c£S..i/<:ttcfri'it,*>rfl \ '
!, ., u\-v-\-v ,-i
;.., i \.. , (..> .. ,...,
^U!..:. L.l.,4,.l
Page 173
Sundstrand Aviation £»«
fl...i,0n cl fturdt|,tno CorpOrM.on ^V W ,
-------
FIGURE VII-6 (2) COMPUTER READOUT FOR A 0.2 UNIT THROTTLE
ACCELERATION AND ZERO LOAD
/// /.- / / /./.../ ,• i-i-i i-ui- /-
^
_!..I.....-.!.-.4....l...........-1-.i :
..i.',... i .1.1 l.: U1...I...4-1...1...! ..!..;.! ..I !...! .'.
/-w^/ /-/-/-/;
Page 174
Sundstrand Aviation &J
-------
BIBLIOGRAPHIC DATA
SHEET
APTD-1558
3. Recipient's Acer-.si"M \>t.
Transmission Study for Turbine and Rankine Cycle
Eng i nes
5- Kepun U.uc
December 15, 1972
6.
A uthurf s )
M. A. Cordner and 0. H. Grimm
8- Performing Or,- .im/.it urn Kept.
No- AER 657
9. Performing Or^auizat ton Name and Address
Sundstrand Aviation
division of Sunstrand Corporation
Rockford, M1i nois 61101
10. Project Task 'iork Unit Ni
1 1. C'orurai t Grant No.
68-0*4-003^
12. Spon^orin^ Oiyani/at ion N,*me and A.iJtvs^
ENVIRONMENTAL PROTECTION AGENCY
Office of Air Programs
Division of Advanced Automotive Power Systems
Ann Arbor, Michigan ^8105
13. I >'pc of Kepurt .S: Period
Covered
14.
15. Supplementary Notes
16. Abstracts
A study was initiated to quantitatively assess the technical and economic feasibility
of existing and potential types of transmissions most suitable for the gas turbine and
Rankine cycle engines. Application of the engine/transmission was to a full size family
car. The study was accomplished through a two-phase, multi-task program which included:
(l) evaluation of transmission types through a feasibility study and ultimate selection
of a transmission type; (2) evaluation of the selected transmission type through design
calculations and layouts, performance analysis, control system analysis, and cost analy
sis. A number of different types of transmission were initially evaluated including
conventional multi-shift, hydrostatic, hydrokinetic, electric, belt/chain, hydromechani
cal, and traction types. Requirements, scope of work, and other data utilized in and
pertinent to the study are included in the appendices.
17. Key U'ords and Document Analysis. 17a. Descriptors
Ai r pol1ution
Automotive transmissions
Eng ines
Turb i nes
Rankine Cycle
Feas ib i1i ty
Economic analysis
Performance tests
Cost analysis
Des ign
17b. Idcntifiers/Opcn-Knded Terms
Control system analysis
17c. COSATI Field/Group
18. Availability Statement
Unlimi ted
19. Security ( la
He port i
20. Sei urity ( l.is- ( I his
Pape
i:xc i. \
-------
INSTRUCTIONS FOR COMPLETING FORM NTIS-35 (10-70) (Bibliographic Data Sheet based on COSATI
Guidelines to Format Standards for Scientific and Technical Reports Prepared by or for the Federal Government,
Pb-180 600).
1. Report Number. Each individually bound report shall carry a unique alphanumeric designation selected by the performing
organization or provided by the sponsoring organization. Use uppercase letters and Arabic numerals only. Examples
FASEB-NS-87 and FAA-RD-68-09.
2. Leave blank.
& Recipient'> Accession Number. Reserved for use by each report recipient.
4. Title and Subtitle. Title should indicate clearly and briefly the subject coverage of the report, and be displayed promi-
nently. Set subtitle, if used, in smaller type or otherwise subordinate it to main title. When a report is prepared in more
than one volume, repeat the primary title, add volume number and include subtitle foi the specific volume.
5. Report Dote. Each report shall carry a date indicating at least month and year. Indicate the basis on which it was selected
(e.g., date of issue, date of approval, date of preparation.
6. Performing Organization Code. Leave blank.
7. AuthoKs). Give name(s) in conventional order (e.g., John R. Doe, or J.Robert Doe). List author's affiliation if it differs
from the performing organization.
8. Performing Organization Report Number. Insert if performing organization wishes to assign this number.
9. Performing Organization Name and Address. Give name, street, city, state, and zip code. List no more than two levels of
an organizational hierarchy. Display the name of the organization exactly as it should appear in Government indexes such
as USGRDR-I.
10. Project/Tosk/Work Unit Number. Use the project, task and work unit numbers under which the report was prepared.
11. Controct/Gront Number. Insert contract or grant number under which report was prepared.
12- Sponsoring Agency Nome and Address. Include zip code.
13. Type of Report and Period Covered. Indicate interim, final, etc., and, if applicable, dates covered.
14. Sponsoring Agency Code. Leave blank.
15. Supplementary Notes. Enter information not included elsewhere but useful, such as: Prepared in cooperation with . . .
Translation of ... Presented at conference of ... To be published in ... Supersedes . . . Supplements
16. Abstract. Include a brief (200 words or less) factual summary of the most significant information contained in the report.
If the report contains a significant bibliography or literature survey, mention it here.
17. Key Words and Document Analysis, (a). Descriptors. Select from the Thesaurus of Engineering and Scientific Terms the
proper authorized terms that identify the major concept of the research and are sufficiently specific and precise to be used
as index entries for cataloging.
(b). Identifiers and Open-Ended Terms. Use identifiers for project names, code names, equipment designators, etc. Use
open-ended terms written in descriptor form for those subjects for which no descriptor exists.
(c). COSATI Field/Group. Field and Group assignments are to be taken from the 1969 COSATI Subject Category List.
Since the majority of documents are multidisciplinary in nature, the primary Field/Group assignment(s) will be the specific
discipline, area of human endeavor, or type of physical object. The application(s) will be cross-referenced with secondary
Field/Group assignments that will follow the primary posting(s).
18. Distribution Statement. Denote releasability to the public or limitation for reasons other than security for example "Re-
lease unlimited". Cite any availability to the public, with address and price.
19 & 20. Security Classification. Do not submit classified reports to the National Technical
21. Number of Pages. Insert the total number of pages, including this one and unnumbered pages, but excluding distribution
list, if any.
22, Price. Insert the price set by the National Technical Information Service or the Government Printing Office, if known.
FORM NTIS-38 (REV. 3-72' USCOMM-DC 14»S2-P72
*U.S. Government Printing Office: 1974--747-787/320 Region Nn 4
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