APTD-1517
TRANSMISSION FOR ADVANCED
AUTOMOTIVE SINGLE-SHAFT
GAS TURBINE AND TURBO-RANKINE
ENGINE
U.S. ENVIRONMENTAL PROTECTION AGENCY
Office of Air and Water Programs
Mobile Source Pollution Control Program
ed Automotive Power Systems Development Division
Ann Arbor, Michigan 48105
-------
APTD-1517
TRANSMISSION FOR ADVANCED
AUTOMOTIVE SINGLE-SHAFT
GAS TURBINE AND TURBO-RANKINE
ENGINE
Prepared by
R. C. Bowlin
Mechanical Technology Incorporated
968 Albany-Shaker Road
Latham, New York 12110
Contract No. 68-04-0033
EPA Project Officer: James C. Wood
NASA Lewis Research Center
Prepared for
U.S. ENVIRONMENTAL PROTECTION AGENCY
Office of Air and Water Programs
Mobile Source Pollution Control Program
Advanced Automotive Power Systems Development Division
Ann Arbor, Michigan 48105
May 1973
-------
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 Mechanical Technology Incorporated in fulfillment of Contract No.
68-04-0033 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 recom-
mendation for use.
Publication No. APTD-1517
11
-------
TABLE OF CONTENTS
I. SUMMARY
II. INTRODUCTION 3
III. DISCUSSION OF RESULTS 5
A. Selection of Candidate Transmission 5
B. Description of Selected Transmission 9
C. Performance of Selected Transmission 13
D. Cost and Physical Comparisons 31
IV. CONCLUSIONS & RECOMMENDATIONS 36
V. SUPPORTING INFORMATION
A. Selection of Transmission A-l
B. Description of Transmission and Controls B-l
C. Description of Methods for Determining Performance C-l
D. Performance Analysis D-l
E. Cost Analysis E-l
F. References F-l
APPENDIX A Al
iii
-------
FOREWORD
This report presents a summary of a study project performed by Mechanical
Technology Incorporated with cooperation from the Ford Motor Company.
The work was performed for the Environmental Protection Agency, Office of
Air Programs, Division of Advanced Automotive Power Systems EPA/AAPS under
Contract 68-04-0033.
A key consultant for the project was Mr. Edwin Charles in regard to the
costing analyses. Acknowledgement is given to Mr. George DeLalio who
submitted a proprietary design of a hydromechanical transmission; to Tracor
for supplying information on their traction drive; and to the Rohr Corpo-
ration for supplying proprietary information on their transmission design.
Mr. R. C. Bowlin was responsible for overall project direction. Other
major contributors from Mechanical Technology Incorporated were P. Lewis,
H. Jones, Dr. A. Smalley and Ms. Linda Almstead.
The 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.
-------
SECTION I
SUMMARY
The purpose of this study was to assess, on both a performance and cost basis,
the transmission most suitable for use with two types of advanced automotive
engines designed to power a medium sized family car. The advanced automotive
engines were a conceptual single-shaft gas turbine and a turbo Rankine engine
based upon design characteristics supplied by AiResearch and Aerojet, respectively.
The scope of the study consisted of a feasibility study to select a candidate
transmission, a preliminary design of the selected transmission, a detailed
performance analysis, control systems definition, cost analysis, and a specific
recommendation concerning the desirability of the selected transmission for the
engine types investigated.
As a result of considering eight different types of transmissions, a continuously
variable ratio, power-splitting, hydromechanical transmission was selected as
the best near-term transmission for application with the single-shaft gas-turbine
and the turbo-Rankine engines. This type of transmission combines hydraulic
elements with mechanical-gear elements to achieve a variable, stepless ratio that
achieves torque multiplication and control by means of the hydraulic elements. The
best long-term candidate was determined to be the Traction-type transmission. This
transmission holds great promise for the future, but requires substantial develop-
ment effort.
The selected hydromechanical transmission was designed in sufficient detail to
ascertain performance, cost and physical characteristics. The design philosophy
followed was to minimize the number of mechanical elements so as to achieve
simplicity and reliability with minimum size, weight, and cost while still meet-
ing specified vehicle performance goals. As a result, in comparison with a
current automatic transmission for a medium-sized car, the selected transmission
was only 3 percent heavier, 28 percent smaller in volume, and required 6 percent
more parts.
Production cost is a key factor in considering a transmission for automotive
applications. A detailed cost analysis, based upon costing procedures of the
Ford Motor Company, showed that the "original equipment manufacture" (O.E.M.)
cost of the selected transmission was 30 to 40 percent higher than a present day
-------
automatic transmission in production quantities of 1,000,000 units/year. On a
variable cost basis (more meaningful to the automotive companies on comparing
designs) the increase in cost was 44 percent. Higher material cost was the
major factor in causing the increased cost. However, since the weight of the
selected transmission was approximately the same as an automatic transmission,
this suggests that, with development of production techniques, future production
costs for such a transmission could approach that of a present day automatic
transmission.
From a performance viewpoint, the selected transmission was compatible with both
engine types and provided smooth acceleration characteristics that met specified
vehicle performance goals. Average transmission efficiency over the Federal
Driving Cycle ranged from 71 to 74 percent with the turbo-Rankine and single-shaft
gas turbine engines, respectively. Comparable efficiency of a conventional auto-
matic transmission powered by an 1C engine has been estimated at 78 percent
The slightly lower efficiency of the selected transmission was due to speed
dependent losses which cause the efficiency to be lower than that of a conven-
tional automatic at cruise-power levels below 25 mpho With respect to fuel
economy (miles/gallon), the selected transmission resulted in significant improve-
ment compared with the conventional automatic transmission for the single-shaft
gas turbine, but little improvement was shown for the turbo-Rankine engine. This
was because the minimum specific fuel consumption characteristics of the.single-
shaft gas turbine changed much more with engine speed variations than did the
turbo-Rankine engine used for this study. Thus, there was little advantage gained
in maintaining engine speed of the turbo-Rankine with a variable-ratio transmission
to achieve maximum miles/galIon.
For this reason, the most significant conclusion of this study was that further
development of the selected variable ratio hydromechanical transmission is re-
commended for application to the single-shaft gas turbine and not for the turbo-
Rankine engine, unless future design developments for the latter engine result in
showing a significant change in SFC with engine speed.
-------
SECTION II
INTRODUCTION
Recently various engine types have been considered as possible alternatives
to present automotive engines in order to improve exhaust emissions. Two
of the engine types are a single-shaft gas turbine and a Rankine-cycle engine
with a turbine expander. Both of these engines require a transmission in
order to achieve torque multiplication for adequate vehicle power. In addition,
based upon fuel economy, the transmission should allow these engines to operate
at conditions which minimize fuel consumption.
The purpose of this study was to select the most promising candidate transmission
for these types of turbine engines and to determine the resultant performance
and production cost of the selected transmission in comparison with a conven-
tional automatic transmission,,
Selection of the most promising transmission must take into account, the perfor-
mance requirements of the vehicle as well as the constraints imposed by development
time for new components, reliability, size and cost. On this basis an overall
assessment pointed to the desirability of the powersplitting type of trans-
mission. Even with this concept there are many possible variations which could
increase the capabilities of the transmission and achieve higher efficiencies,
etc. For this study, the selection and design philosophy concentrated on minimiz-
ing the number of mechanical elements to achieve simplicity and high reliability
with state-of-the-art components within the constraints of size, weight, and cost.
As a consequence, more complicated versions were not considered for this study.
A preliminary design of the selected transmission was made in sufficient detail
to provide the basis for a reasonable estimate of transmission weight, volume,
number of parts, and cost. In addition, the transmission control system was
defined by a control schematic; control logic details were investigated in con-
junction with determining power-train performance
-------
Engine characteristics used in this study were supplied by AiResearch based on
their conceptual design for a single-shaft gas turbine and Aerojet based on
their "prototype" design for a turbo-Rankine engine with an organic working
fluid. The respective engine maps were incorporated into digital computer models
in order to determine steady-state (cruise), maximum acceleration, and Federal
Driving Cycle performance with the selected transmission.
The following sections of this report discuss a summary of the results and then
present conclusions and recommendations. Detailed supporting information is
contained in the last section of the report.
-------
SECTION III
DISCUSSION OF RESULTS
This section summarizes and discusses the results of the study by considering
in the following order:
A. Selection of candidate transmission
B. Description of selected transmission
C. Performance of selected transmission
D. Cost and physical comparisons
Additional supporting details are given by topical headings in Section V -
Supporting Information,,
A. Selection of Candidate Transmission
A number of different types of transmissions were specified by EPA/AAPS for
for consideration as a candidate transmission. These types were:
Mechanical
Hydrostatic
Combination of Mechanical and Hydrostatic
Hydrokinetic
Electrical
Traction
Belt Drive
As discussed in more detail in Section V-A, all of the above types were considered
in order to select a candidate transmission for the single-shaft gas turbine and
the turbo-Rankine engine. Several of the above were eliminated from serious consi-
deration because of inherent limitations. For example, a purely mechanical
gear-type transmission was eliminated on the basis that it would not provide the
continuously variable ratio with stepless changes, which is required for smooth
operation. Similarly the hydrostatic and electrical were eliminated because of
large volume, high cost and relatively low efficiency for this application.
5
-------
As a result, the possible candidate transmissions were narrowed down to eight
specific types of transmission, which were:
1. Three-Speed Automatic with Variable Torque Converter Element
2. Advanced Hydromechanical
3. Conventional Hydromechanical
4. Traction - TRACOR
5. Traction - Power-Splitting
6. Friction - Composition Belt
7. Traction - Metal Belt
8. Three-Speed Automatic with Aerodynamic Torque Converter
Each of the above transmission types was reviewed and evaluated. Descriptive
details for each of the transmission types and the details of the evaluation are
given in Section V-A. Table 1 presents a summary of the descriptive evaluation
for each transmission type and Table 2 presents an evaluation summary.
As a result, on an overall basis, the power-splitting hydromechanical trans-
mission (subsequently described) was selected as the most promising candidate
transmission on a near-term development (1974) basis. Key attributes of this
type of transmission were relative simplicity, comparable size and weight to a
three-speed automatic transmission which is currently used, weight (which would
imply future cost comparable to the standard automatic), proven components well
within the current state-of-the-art, and the ability to provide optimum engine
speed throughout the desired operating range. Disadvantages of this type of
transmission include excessive noise particularly when operated at high
hydraulic pressures and reduced efficiency at low power levels if input speeds
are high.
From a long-range viewpoint, with additional development work, the traction
type of transmission was found to offer considerable promise as an alternate
candidate transmission.
-------
TABLE 1
DESCRIPTIVE EVALUATION OF CANDIDATE TRANSMISSIONS
1 . 3 Speed Auto-
matic w/var iable
(.-lem'.-nt
2 . Advanced Hydro-
mech .
3. Conventional
Hydromech.
4. Traction -
TRACOR
5. Tract ion -Power-
Splitting
6 . Traction -
Composition
Belt
7. Traction -
Metal Belt
8. 3-Speed Auto-
matic w/Aero
torque con-
verter - ROHR
Life &
Reliability
iimilar to existing
[except for variablt
.lenient)
iimilar to existing
\T. lie ne fi t ci siir-
)licity
iimilar to AT. Some
>cnalty over (2)
Jot defined at this
: ime
tot defined at this
time
Not defined at this
Lime
tot defined at this
time
tot defined at this
t imo
Noise &
Smoothness
Good - similar to
existing
Smoother Oper - Re-
quires attention to
noise
Smoo the r Oper - Re-
quires attention to
noise
Smooth Oper - w/low
Smooth Oper - w/low
noise potential
Smooth Oper - w/low
Smooth Oper - w/low
noise potential
Smooth Oper - w/low
noise potential
Cost
Additional
cost for extra
element
Slightly
ligher than
existing
ligher than
(2)
ligher than
(2)
Higher than
(2)
Potential for
Potential for
low cost
Higher than
(2)
Development
Status
Developed - not in
production
Developed - not in
auto production
Developed - not in
auto production
Not completely de-
veloped for auto
hp range
Concept Only
Not developed for
auto hp range
Not developed for
auto hp range
Not developed as a
transmission
Efficiency
Low - Penalty be-
cause of variable
element
Higher than exis-
ting AT, except at
low power levels
Higher than exis-
ting AT, except at
low power levels
Not as high as (2)
Potentially simi-
lar to (2)
Potentially simi-
lar to (2)
Potentially simi-
lar to (2)
Not as high- as (1) ;
significant penal tj
for Aero element
Note: AT designates current mass produced three-speed automatic transmission
Size &
Weight
toraina 1 increa se
over existing AT
lomparable to
existing AT
ligher than (2)
Higher than
existing AT
Higher than
existing AT
Not defined for
auto applic.
Not defined for
auto applic.
Not defined,
should be higher
than existing AT
Restriction
On Turbine
Engine
Slight penalty due
to steps
linimal
linimal
linimal
linimal
Possible speed
range limitations
Possible speed
range limitations
linimal
Control
Complexity
Minimal
Minimal
Minimal
Minimal
Minimal
Minimal
Minimal
More complex
Driver
Acceptability
Simlar to
existing AT
Similar to
existing AT
Similar to
existing AT
Similar to
existing AT
Similar to
existing AT
Similar to
existing AT
Similar to
existing AT
Similar to
existing AT
Environmenta 1
Restrict ions
Similar to
existing AT
Similar to AT, Re-
qires attention
to shock design
Similar to AT, Re-
quires attention
to shock design
Similar to AT,
Requires atten-
t ion to shock
de ign
Low shock absorp-
tion capability
Similar to
existing AT
Similar to
existing AT
Similar to
existing AT
KT1-U59
-------
TABLE 2
EVALUATION SUMMARY
TRANSMISSION TYPE
1. Three-Speed Automatic
with Variable Torque
converter element
MAIN ADVANTAGE
Similar to existing
automatic transmission
MAIN DISADVANTAGE
Lower efficiency than
existing automatic
2. Advanced Hydromechani-
cal
Simple mechanical
construction
Manufacturing techni-
ques need to be developed
and noise minimized
3c Conventional Hydro-
mechanical
Development experi-
ence existing
Manufacturing techniques
need to be developed and
noise minimized
4<, Traction - TRACOR
Future potential for
high efficiency with
low noise
o Traction-Power Split-
ting
Future potential for
high efficiency with
low noise
Not to design stage
6. Friction-Composition
Belt
7o Traction-Metal Belt
8. Three-Speed Automatic
with Aerodynamic Tor-
que Converter-Rohr
Manufacturing techni-
ques developed
Future potential for
high efficiency
Manufacturing techni-
que developed for bulk
of the transmission
Life and efficiency
undetermined
Not to development stage
Not to complete
transmission design
stage.
MTI-14596
-------
B. Description of Selected Transmission
The selected hydromechanical transmission is an infinitely variable, stepless
unit that obtains torque multiplication and control by means of hydraulic ele-
ments (pump-motor-combination). The transmission is based upon a proprietary
design of Mr. George DeLalio. Consequently, design of the transmission has
not been included in this report. The unit differs from the conventional torque
converter or fluid-coupling-type transmission, in that the hydraulic power is
transferred by fluid static pressure at low flow as contrasted to the high dynamic
action of fluid as utilized in hydrodynamic units. It also differs from a
purely hydrostatic transmission in that the much more efficient mechanical ele-
ments transmit a significant portion of the power. It is a "hard" type of drive,
in that slip is less than 2 percent under full load.
The design of the transmission was based upon minimizing the number of mechanical
parts in order to: a) make the transmission as simple as possible, b) keep the
size, weight and, particularly, the cost low, c) require the minimum amount of
development, and d) achieve high reliability while retaining a reasonably high
transmission efficiency. Additional gear trains and hydraulic functions could
have been employed to increase the capabilities of the transmission and reduce
the amount of power in the hydraulic elements. Clearly, however, this approach
would have involved more parts, and a more costly transmission. It was more con-
sistent with the stated design philosophy to concentrate on a straight forward,
simple, design.
Figure 1 is a functional schematic of the selected transmission. As shown on
this figure the transmission consists of an engine input shaft, a variable-dis-
placement hydraulic element and a fixed-displacement hydraulic element. Additional
components which are essential to the transmission operation are: a brake, which
locks one element of the planetaries for low-speed operation; a clutch, which closes
the mechanical power both for high-speed operation, and a control system.
As shown in Figure 1, the power path between input and output splits between hydro-
static and straight mechanical. The transmission has two operating ranges desig-
nated as low-range and high-range. At low power and output speed levels (up to
approximately 0.4 of maximum output speed) the transmission operates in the low
range and all power flows through the hydraulic elements. At higher power and
speed conditions the transmission operates in high range and the power path is
-------
SHIFT SELECTOR LEVER
THROTTLE POSITION
TURBINE
SHAFT
MEASURED
ENGINE
SPEED
I I
CONTROLS
VARIABLE
DISPLACEMENT
ELEMENT
FIXED
DISPLACEMENT
ELEMENT
SIMPLE
PLANETARIES
OUTPUT TO VEHICLE
Fig. 1 Functional Schematic Selected Hydromechanical Variable-Ratio Transmission
KTI-K628
-------
split between the hydraulic elements and the mechanical path. At approximately
0.7 of maximum output speed, all the power flow is through the mechanical path.
An important feature of this design is the synchronous shift which operates as
follows: The transmission has two operating ranges, low speed and high speed.
Transfer between these two ranges is effected by the concurrent opening of a .
brake and closing of a clutch. The shift is ideally synchronous if it occurs
under the following conditions:
1. In either low or high range the displacement of the variable hydraulic
element is at the same maximum point.
2. There is no change in relative velocity between the two sides of the
clutch.
3. The braked element is stationary whether the brake is applied or not.
In practice such effects as slip between hydraulic elements and control imper-
fections cause slight deviations from the ideally synchronous shift. Even so,
there is never a significant change in momentum demanded of the rotating parts
and the brake and clutch elements suffer little slippage or wear.
While theoretically unlimited , the practical low-speed torque ratio range lies
between 5:1 and 2.5:1 and the high-speed (mechanical path) range lies between
2.5:1 and 1:1. Thus the overall ratio range for the selected transmission
lies between 5:1 and 1:1. With the 3.08 rear-end ratio, the maximum torque ratio
between transmission input and the wheels is 15.4.
From a driver viewpoint, operation of the selected transmission is similar to
a present automatic transmission, except no "kick down" is required to downshift.
Gear shift lever functions are identical to a present three speed automatic trans-
mission and are described in detail in Section V-B.
The controls required for the hydromechanical transmission will be more sophis-
ticated than those required to operate the present automobile automatic trans-
mission. This results from the fact that to exploit full benefit from the con-
tinuously variable transmission it must be controlled to operate the turbo-rankine
11
-------
or gas turbine engine within the most economical fuel flow range. To accomplish
this objective an engine speed regulator governor and an extra flow path through
the transmission control valve are added.
The raw signals used to achieve optimum fuel economy are accelerator pedal posi-
tion and engine speed. The accelerator pedal position signifies commanded
vehicle speed and is translated via a cam into a signal, representing the corres-
ponding optimum engine speed (desired speed) (determined by design analysis).
The engine speed is translated, via the engine regulator governor, into a speed
signal, and compared to the desired speed. The error signal (hydraulic) causes
the piston actuator to move the swash plate cam, so establishing a new position
of the variable hydraulic element and a different transmission ratio. Subject
to limitations of the vehicle engine and control system time constants, optimum
fuel consumption at the commanded vehicle speed is achieved (on a level road).
Of course frequent acceleration and braking will tend to limit fuel economy as
with any vehicle power train.
One of the key features of the transmission control system is that at engine idle
speed the transmission is automatically disengaged from the engine, again mini-
mizing fuel flow. The operation is accomplished by sending an idle engine speed
signal from the engine engage governor through the additional flow path of the
control valve and on to the engaging spool valve. The engaging valve then
disengages the hydraulic element and the transmission from the engine.
When the engine speed is above idle the engage governor, through similar hydraulic
flow paths, causes the transmission to engage.
The shift from low range to high range, or vice-versa, is designed to occur
when the displacement of the variable element is at its extreme negative value.
A particular valve (the clutch valve) is used for this purpose. A cam, controlled
by the actuator piston has a profile which, over a short part of its travel,
slews the clutch valve from its low range position to its high range position.
The effect of this valve is to relieve pressure from the low-grange brake and apply
it to the high-range clutch, or vice-versa.
The controls hardware is described in more detail in Section V-B.
12
-------
C. Performance of Selected Transmission
The performance of the selected hydromechanical transmission was determined over
a wide range of operating conditions. Throe categories of performance were
investigated: 1) steady-state (cruise and constant power output), 2) Federal
Driving Cycle, and 3) full-power accelerations. In each analysis the perfor-
mance was determined for two different types of advanced automotive engines
powering a medium-sized family car with .power-train loading specified by EPA/
AAPS vehicle design goals (1). Comparisons, where possible, were made with the
corresponding performance of a similar power-train employing a conventional auto-
matic transmission.
The engine characteristics were supplied by Aerojet and AiResearch as specified
by EPA/AAPS as a result of work under separate EPA/AAPS contracts. The charac-
teristics of the Aerojet engine were based upon their "prototype" design for a
turbo-Rankine-cycle engine, which employs an organic working fluid. The AiResearch
engine characteristics were based upon their conceptual design for a single-shaft
gas turbine. Details of the engine maps and power-train loading are given in
Section V-D.
It should be pointed out that: (1) the respective engine characteristics were
based upon different design constraints not necessarily optimized for operation
with the selected transmission and (2) the engine performance predictions have
not been demonstrated experimentally. For these reasons comparisons between
engines, based upon these study results, are not warranted.
1. Transmission Performance Characteristics
*
The resultant transmission efficiency of the selected hydromechanical transmission
is dependent upon the mode of operation. In low range, at low vehicle speeds,
all the power passes through the hydraulic elements which limit the efficiency.
At high speeds, where the majority of the power passes through the mechanical
power path, the efficiency is higher. Peak efficiency occurs close to the
"straight-through" point, where power flow through the hydraulic path is essen-
tially zero.
^Transmission efficiency, as used throughout this report, relates the power output
from the transmission (delivered to the differential) and the input power to the
transmission input shaft.
13
-------
The above characteristics can be seen by referring to Figure 2, which presents
the variation of transmission efficiency with vehicle velocity under cruise-power
conditions for both engines. At 20 mph the efficiency is between 57 to 60 percent,
but increases to 90 percent at speeds of 50 mph or higher. The slight discontin-
uity in slope at 20 mph with the AiResearch engine, and at 25 mph with the Aerojet
engine, indicates the cruise-power shift point between low and high speed ranges.
The "straight-through" point occurs at 40 mph and 50 mph for the AiResearch and
Aerojet engines, respectively.
The marked difference between the transmission efficiencies with the two engines
at speeds below 50 mph reflects the differences in the manufacturer's specified
operating line. Up to approximately 24 hp output from the engine, the trans-
mission input speed with the AiResearch engine is 2300 rpm (minimum engaged speed).
However, with the Aerojet engine, the transmission input speed lies between
2700 and 3000 rpm. This difference in speed means that speed dependent trans-
mission losses are more significant with the Aerojet engine than with the AiResearch
engine and, at the low power levels, was 13 percent at 30 mph and 3 percent at
20 mph.
The variation in efficiency at constant power levels as shown in Figures 3 and
4 reveals some differences from the cruise efficiency curves. This is particu-
larly noticeable at low speeds. For example, at 10 mph, the transmission cruise
efficiency is 51 percent; whereas, with an output at the road of 10 percent of
rated power, the efficiency is 78 percent. This difference results from the fact
that the cruise power demand at 10 mph is only 2 hp at the road, exaggerating
the importance of speed dependent losses, which are of similar magnitude to the
road load. A further observation from the constant-power efficiency curves is
the reduced separation between the two engines at high power levels, indicating
the reduced significance of speed-dependent losses at high power levels. For a
cross-plot of transmission efficiency versus output power at constant speeds
refer to Section V-D.
The peak transmission efficiency is close to 95 percent, under conditions of
high power at the "straight through" speed condition. This high efficiency value
is a result of the minimal losses which occur when all power passes through the
14
-------
s
w
a
a
a
PK
fH
100
80
60
With AiResearch
Engine
With Aerojet Engine
20
Vehicle Wt.
Level Road
Rear End n
Rear End Ratio
4600 Ibs
0.96
3.08
10
20
30 40 50
VECHICLE VELOCITY, MPH
60
70
80
Fig. 2 Cruise-Power Efficiency of Selected Transmission
-------
100
1
w
I
X
§
H
w
Percent
Rated Power
20
30 40 50
VECHICLE VELOCITY, MPH
60
70
80
Fig, 3 Constant Power Efficiency of Selected Transmission
AiResearch Engine
MTI-1463
-------
100
a
a
I
a
Percent
Rated Power
40
20
10
30 40 50
VEHICLE VELOCITY, MPH
60
70
80
Fig. 4 Constant Power Efficiency of Selected Transmission
Aerojet Engine
MTI-U594
-------
mechanical path. The 5 percent losses which are incurred are the sum of the
cumulative losses in the output planetary, the residual power necessary to ro-
tate the variable hydraulic element, even when no power is passing through it,
and the "parasitic" losses such as the charge pump.
Figure 5 presents a comparison of the selected hydromechanical transmission
cruise power efficiency with that of the conventional automatic. The conventional
automatic used for this comparison was that currently selected by Aerojet for use
with their turbo-Rankine engine. The shift points are therefore designed to pro-
duce optimum system performance for this engine. The wide spread of the shift
points, even under the cruise conditions of Figure 5, is necessitated when a
vehicle speed range for zero to 85 mph is to be provided by an engine whose
ratio of maximum to idle speed is well below 2. In fact, the complete transmis-
sion incorporated by Aerojet includes a separate idle gear for use between 10 and
22 mph and a slipping clutch for speeds below 10 mph. Thus, in the range 10-85 mph,
the so-called "conventional automatic" actually behaves as a 4-speed, rather than
a 3-speed.
Below 30 mph the conventional automatic is significantly more efficient than the
selected transmission - the difference reaching 24 efficiency points at 10 mph
for the Aerojet engine. As shown in Section V-D (Figures D-10 and D-ll), the
efficiency of the selected hydromechanical transmission decreases rapidly below
20 percent load due to speed dependent losses. By contrast, the part load effi-
ciencies for a torque converter transmission increase with decreasing load within
the converter range. At increased load corresponding to vehicle speeds above
35 mph, it can be seen from Figure 5, that the selected transmission efficiency
was equal to, and in some instances better than, the automatic transmission.
Additional insight into the performance of the selected transmission is provided
by Table 3, which gives a detailed breakdown of the power flow and indicates
the contribution of mechanical and hydraulic losses to performance. Vehicle
speeds at 20 mph and 60 mph with the Aerojet engine were selected as typical
operating conditions. The first of these speeds is slightly below the shift
point, and the second speed is somewhat above the "straight-through" point. In
both cases the amount of power flowing through the hydraulic path is very similar.
18
-------
100
M
W
04
a
o
PK
W
80
60
40
20
<
V
Wi
AiResear<
-- K>_^
^
ehicle Wt.
:h
:h Engine
//
-"~/£.J
With Aero
s
Level Road
Rear End n =0.96
Rear End Ratio = 3.08
1 1
lf^>~- <
jet Engine
L.. O (.
Std. Automa
mission (Re
1
:ic Trans-
ar End Ratio
= 2 . 93 *
> Shift po
Automati
by Aeroj
Lnts for Std
: Selected
s t
0 10 20 30 40 50 60 70 80
VEHICLE VELOCITY MPH
Fig. 5 Comparison of Transmission Efficiency Cruise Power
Different Rear End Ratio for Standard Automatics selected by Aerojet to give optimum performance
with remainder of power train (engine + transmission)
-------
TABLE 3
TYPICAL POWER FLOW BREAKDOWN ~ AEROJET ENGINE
WITH AIR CONDITIONER - CRUISE POWER
r
ENGINE HP
ACCESSORY HP
TRANSMISSION
INPUT HP
MECHANICAL
PATH LOSSES
HYDRAULIC
PATH LOSSES
TRANSMISSION
OUTPUT HP
DIFFERENTIAL
LOSSES
ROAD HP
20 MPH
AVAILABLE
HP
13.114
8.374
4.776
4.592
HP
USED &
LOSSES
4.740
1.126
2.472
0.184
% OF
TRANSMISSION
INPUT HP
56.60
13.45
29.52
2.20
60 MPH
AVAILABLE
HP
39.908
34.996
31.826
30.602
HP
USED &
LOSSES
4.911
1.356
1.814
1.224
% OF
TRANSMISSION
INPUT HP
10.03
3.87
5.18
3.50
20
-------
However, at 20 mph the only power flow to the mechanical path is that necessary
to overcome friction and to drive the charge pump - no output power is delivered
by the mechanical path. At 60 mph most of the delivered power passes through the
mechanical path. The influence of this difference in power split is reflected in
the percentage contribution of the hydraulic and mechanical losses. At 20 mph
hydraulic losses account for 29.5 percent of the transmission input power, and
mechanical losses account for 13.5 percent. At 60 mph the hydraulic losses fall
to 5.2 percent and the mechanical losses to 3.9 percent.
The design operating pressure in the hydraulic elements of a hydromechanical
transmission is important with respect to reliability (life) and noise. Maximum
operating pressures above 3500 psi not only reduce the life of the elements but
also cause unwanted excessive noise. Thus, low operating pressures are highly
desirable to minimize transmission noise.
Typical design operating pressures for the selected transmission are shown by
Figure 6. At cruise-power, operating pressures were low, reaching a maximum of
330 psi at 85 mph. Under full-power demands the pressure remains close to its
limiting value of 3500 psi at speeds of 10 and 15 mph, reflecting the design limit
corresponding to wheel slip. At higher speeds, even in low range, the full power
can be transmitted through the hydrostatic path without exceeding the pressure
limits, and the pressure begins to fall with speed. At a speed of 85 mph, the
full-power pressure has decreased to about 530 psi. These results indicate that
the selected transmission design was conservative with respect to operating pres-
sure and therefore noise levels should be at a minimum, since the noise is most
strongly influenced by pressure and speed. Even though the low operating pres-
sures minimize noise, it should be recognized that they do not eliminate the
noise problem. Isolation and noise insulation techniques will probably have to
be developed to achieve driver aceptability.
2. Power-Train Performance - Aerojet Engine
The fuel economy of a vehicle power train, in mpg, is, of course, a meaningful
measure of the overall efficiency with which fuel is being converted to useful
work. The variable ratio selected transmission was controlled in a manner so as
to maintain engine speed for minimum fuel consumption. As pointed out earlier,
21
-------
4000
ro
3000
M
W
Oi
w
to
a
i
2000
1000
Vehicle Wt.
Level Road
Rear End T|
Rear End Ratio
- 4600 Ibs
- 0.96
= 3.08
VEHICLE VELOCITY, MPH
Fig. 6 Hydrostatic Pressures for Selected Transmission
MTl-14635
-------
the transmission changes ratio (hence engine load) so that engine operation was
maintained on a desired operating line except at large acceleration power levels.
Under maximum power accelerations, the transmission automatically provides the
largest necessary torque ratio.
Figure 7 presents the cruise-power fuel economy of the Aerojet engine with the
selected transmission. Optimum fuel economy occured at 35 to 40 mph; 15.4 mpg
with air conditioner and 17.4 without air conditioner.
A comparison in fuel economy (with air conditioner) to a torque-converter auto-
matic transmission, with the Aerojet engine, is given by Figure 8 (details of the
automatic transmission are given in Section V-D). The discontinuities in perform-
ance, associated with each shaft point for the automatic transmission are clearly
shown. However, on an average basis it is apparent that, at low speeds, the fuel
economy is slightly better with the automatic transmission, and at higher speeds
the selected transmission produced slightly better fuel economy. The predominant
reason for this similarity in power-train performance is the flat, symmetrical
nature of the Aerojet engine performance map. Thus, typically, a 500 rpm deviation
from the optimum engine speed, either up or down, causes only about 1 out of 17
deviation in engine efficiency. At low vehicle speeds the higher efficiency of
the automatic transmission actually results in higher fuel economy when the engine
is operating near its minimum SFC point.
Also shown by Figure 8, is the resultant fuel economy for an idealized situation
where the selected transmission has 100 percent efficiency. By comparison, even
with this perfect transmission and the Aerojet engine operating at maximum
efficiency, 17.9 mpg is the best fuel economy - only 3 mpg or 20 percent better
than the automatic transmission.
The Federal Driving Cycle provides an alternative operating condition to compare
performance. These results for the Aerojet engine are given by Table 4, and
show that the conventional automatic provides a small (4 percent) improvement in
average fuel economy over the hydromechanical transmission. Thus, again the two
transmissions are of similar benefit to the Aerojet engine. In this case a
perfectly efficient hydromechanical could provide 21 percent better average fuel
economy than the conventional automatic.
23
-------
ro
P-
O
8
Without A/C
Vehicle Wt.
Level Road
Fuel Density
Rear End T)
Rear End Ratio
6.30 Ibs/gal
0.96
3.08
10
VEHICLE VELOCITY, MPH
Fig. 7 Cruise Fuel Economy with Selected Transmission
Aerojet Engine
NTl-14626
-------
8
w
1-1
18
16
14
12
10
8
6
4
2
0
c
71
/
M
/
= 1002
/
~J>
~\
\
s
s
r-<
1
/\
u
"*
K. (
x^
Vehicle Wt. = 4600 Ibs
Level Road
Fuel Density = 6.30 Ibs/gal
Rear End T] =0.96
Rear End Ratio = 3.08
Cruise Power
i i i i i
-r2
**i
/\
^
"--
*=--,
-^
^^
^^
^*X
STD
Automatic
^-^
n
<
i
Selected
^ Transmission
^v
N
>
**"^
*^
^V
)
1 Shift
Points For
STD Auto-
matic
mis sic
ted b-s
Trans-
n Sel<
Aeroj
^^
\
c-
et
^^
^^
s.
V,
^
0 10 20 30 40 50 60 70 80
VEHICLE VELOCITY, MPH
Fig. 8 Comparison of Fuel Economy Aerojet Engine with
Air Conditioner Cruise Power
-------
TABLE 4
AEROJET ENGINE - DRIVING-CYCLE PERFORMANCE
Quantity
Average MPG
Average
Transmission
Average Engine
Power
Average Road
Power
Average
Velocity
Selected
Transmission
With A/C
19.6 mph
Selected
Transmission
Without A/C
19.6 mph
1007o
Transmission
With A/C
9.55
71%
16.76 hp
8.38 hp
10.85
717.
14.43 hp
8.38 hp
12.02
1007,
13.16 hp
8.38 hp
19.6 mph
Automatic*
Transmission
With A/G
9.9
* Date provided by Aerojet (transmission efficiency engine power data
not available)
Consider now the full-power acceleration performance of the selected transmission
with the Aerojet engine. As shown by Table 5, the power train exceeds all of the
EPA/AAPS maneuver specifications. The time history of vehicle velocity (Section
V-D) shows smooth, stepless acceleration for the power train consisting of the
Aerojet engine and the selected transmission.
TABLE 5
AEROJET MANEUVER PERFORMANCE
1. Distance traveled in 10 seconds
2. Time to reach 60*' mph from standing
start
3. High speed merge (25-70 mph)
4. DOT passing maneuver (time and
distance to overtake 50 mph truck)
TIME
DISTANCE
EPA
Specifications
440 ft.
13.5 sec
15.0 sec.
15.0 sec.
1400 ft.
Aerojet Engine
with Selected
Transmission
505 ft.
11.7 sec.
13.5 sec.
12.2 sec.
1166 ft.
26
-------
3. Power-Train Performance - AiResearch Engine
Figure 9 presents the fuel economy of the AiResearch engine coupled with the
selected transmission. Peak fuel economy was 28.3 mpg at 40 mph without air
conditioner and 26 mpg with air conditioner.
The power-train performance of a vehicle incorporating the same automatic trans-
mission as discussed for the Aerojet engine was computed. It is emphasized that
the conventional automatic could never provide satisfactory kinematic performance wit
the single shaft gas turbine but the comparison does provide an exagerated demon-
stration of the advantages of a continuously variable transmission to this type
of engine. A comparison of fuel economy is shown by Figure 10. These results
clearly show a considerable advantage in fuel economy with the selected trans-
mission. The reason for the poor performance shown by the automatic trans-
mission is the extreme sensitivity of the single-shaft, gas-turbine SFC to
engine speed for a given power demand.
Performance over the Federal Driving Cycle for the AiResear^h engine with the
selected transmission is presented in Table 6. No data is avaiable for com-
parison with the conventional automatic. The average fuel economy values ob-
tained with the selected transmission were 14.53 mpg with air conditioning and
15.76 without.
TABLE 6
DRIVING CYCLE PERFORMANCE - AIRESEARCH ENGINE
Selected Selected
Transmission Transmission
Quantity With A/C Without A/C
Average MPG 14.53 15.76
Average Transmission 74.4 73.5
Average Engine Power 15.94 hp 13.92
Average Road Power 8.38 hp 8.38 hp
Average Velocity 19.6 mph 19.6 mph
27
-------
ro
oo
28
26
24
22
20
18
16
14
12
10
10
X
Vehicle Wt.
Level Road
Fuel Density
Rear End T)
Rear End Ratio
I
1
4600 Ibs
6.30 Ibs/gal
0.96
3.08
I I I
With A/C
Without A/C
20
30
40
50
60
70
VEHICLE VELOCITY, MPH
Figo 9 Cruise Fuel Economy with Selected Transmission
AiResearch Engine
80
MT1-K627
-------
VO
o
o
o
W
24
20
16
12
Vehicle Wt.
Level Road
Fuel Density
Rear End T]
Rear End Ratio
4600 Ibs
6.30 Ibs/gal
0.96
3.08
10
\
MTI Transmission
\
STD Automatic
i
Shift
Points
20
30 40 50
VEHICLE VELOCITY, MPH
60
70
80
Fig. 10 Comparison of Fuel Economy AiResearch Engine with Air Conditioner
Cruise Power
-------
Full-power performance calculations showed that the AiResearch engine had to be
scaled upward by 15, percent more power, to meet all of the EPA/AAPS specifications.
The results are given in Table 7 below:
TABLE 7
AIRESEARCH MANEUVER PERFORMANCE
EPA
Maneuver Specifications
AiResearch
Engine
1. Distance traveled in 10 seconds
2. Time to read 60 mph for
standing effort
3. High speed merge (25-70 mph)
4. DOT passing maneuver (time and
distance to overtake 50 mph truck)
TIME
DISTANCE
440 ft.
13.5 sec,
15.0 sec.
15.0 sec.
1400 ft.
447 ft.
11.1 sec.
11.6 sec.
11.8 sec,
1139 ft.
30
-------
As discussed in Section V-D, the primary reason for scaling up the engine power
was in order to meet the requirement of 440 feet traveled in 10 seconds. This
was caused by the inherently low starting torque characteristics of a single-
shaft gas turbine.
The resulting time history of vehicle velocity (see Section V-D) showed smooth,
stepless acceleration characteristics for the AiResearch engine when coupled to
the selected transmission.
D. Cost and Physical Comparisons
One of the most important aspects of any transmission being considered for auto-
motive applications is cost. Consequently, a detailed cost analysis was performed
(refer to Section V-E for detail), using procedures currently practiced by the
Ford Motor Company, in order to determine the production cost of the selected
transmission. This analysis consisted of determining detailed cost estimates for
approximately 200 separate parts. The resulting costs were then compared to the
current cost of an automatic transmission for a medium-sized family car. Since
the cost of the automatic transmission was based upon proprietary information of
the Ford Motor Company, the results of the cost analysis are presented as ratios.
Detailed costs were determined on a "variable"cost" basis rather than an
"original equipment manufacturer" (O.E.M.) basis; since this approach is more
meaningful to the automotive industry in comparing designs. Variable cost in-
cludes: 1) purchased cost of a part, 2) direct labor required to get the part
to a desired condition, 3) indirect labor associated with the manufacturing
process, 4) variable overhead items which specifically relate to the manufactur-
ing process, and 5) specific (programmed) overhead expenses, such as specific
required testing.
The O.E.M. costs (specified by EPA/AAPS) were estimated from the aggregate
variable cost. O.E.M. cost includes transfer costs such as capital investment,
engineering development, facilities, etc. which are dependent upon management
strategy decisions. Consequently, O.E.M. cost estimates were given a range to
account for probable variations.
31
-------
A summary of the cost analysis is presented in Table 8. These results show
that, in production quantities of 1,000,000 units/year, the variable cost of
the selected hydromechanical transmission will be 1.44 times the cost of a
conventional three-speed automatic transmission - an increase in cost of 44
percent. On an O.E.M. cost basis, the increase ranged from 30 to 40 percent.
For smaller production quantities (100,000 units/year) the increase in cost
could be as high as 70-80 percent on an O.E.M. basis, when compared to larger
production quantities of the automatic transmission.
Also shown in Table 8, is a breakdown in costs attributed to controls, labor,
and material at production levels of 1,000,000 units/year. The selected
transmission control cost increase was 28 percent, additional labor content
34 percent, and material content 53 percent. Thus, the major factor causing
the increased costs was the material content of the transmission.
As subsequently discussed, the selected hydromechanical transmission weighs
approximately the same as a conventional automatic transmission. Therefore,
it is reasonable to believe that when the design and manufacturing skills of the
automobile industry, which have been applied over many years to existing trans-
missions, are applied to the selected transmission, the cost will approach the
cost/pound ratio of existing transmissions. This suggests that in the future the
production cost of the selected transmission would approach the present cost of
anautomatic transmission.
Consider now a comparison of pertinent physical characteristics. As summarized
in Table 9, the selected transmission design presented herein was 3 percent
heavier, had 6 percent more parts, and required 28 percent less volume than a
comparable three-speed automatic transmission. Figure 11 presents a comparison
of volume envelopes which shows that the selected transmission is even smaller
than a two-speed automatic transmission. Thus, it was concluded that the
selected transmission was smaller but was slightly heavier than a comparable
automatic transmission.
32
-------
TABLE 8
TRANSMISSION COST ANALYSIS -
COST RATIOS
PRODUCTION LEVEL
1,000,000 UNITS PER YEAR
STANDARD AUTOMATIC
TRANSMISSION WITH
TORQUE CONVERTER*
POWER SPLITTING
HYDROMECHANICAL
TRANSMISSION
PRODUCTION LEVEL
100,000 UNITS PER YEAR
STANDARD AUTOMATIC
TRANSMISSION WITH
TORQUE CONVERTER*
POWER SPLITTING
HYDROMECHANICAL
TRANSMISSION
1. VARIABLE COST RATIO (TOTAL)
a. CONTROL VARIABLE COST
RATIO
b. LABOR CONTENT RATIO
c. MATERIAL CONTENT RATIO
2. OEM COST RATIO
1.00*
1.44
1.29
1.00
1.00
1.00
1.28
1.34
1.53
1.29
1.50
1.20
1.00
1.30-1.40
1.25-1.35
1.65
2.02
1.84
1.70-1.80
USED AS REFERENCE, PRODUCTION LEVEL
OF 1,000,000 UNITS PER YEAR.
-------
TABLE 9
TRANSMISSION PHYSICAL COMPARISONS
TOTAL TRANSMISSION
AND CONTROL VOLUME
TRANSMISSION
VOLUME - FT3
CONTROL VOLUME
FT3
TOTAL TRANSMISSION
AND CONTROL WEIGHT
TRANSMISSION
WEIGHT - LBS.
CONTROL
WEIGHT - LBS.
TOTAL TRANSMISSION
AND CONTROL PARTS
TRANSMISSION -
NUMBER OF PARTS
CONTROL -
NUMBER OF PARTS
POWER-SPLITTING
TRANSMISSION
1.10 FT3
1.00 FT3
.10 FT3
146 LBS
130 LBS
16 LBS
250
175
75
AUTOMATIC THREE SPEED
WITH TORQUE CONVERTER
TRANSMISSION
1.52 FT3
1.40 FT3
.12 FT3
142 LBS
123 LBS
19 LBS
236
166
70
RATIO OF POWER-SPLITTING
TO STANDARD AUTOMATIC
THREE SPEED TRANSMISSION
.72
.714
.833
1.03
1.057
.842
1.06
1.054
1.071
34
-------
TWO SPEED AUTOMATIC TRANSMISSION
HYDROMECHANICAL TRANSMISSION
Fig. 11 Comparison of Transmission Envelopes
35
-------
SECTION IV
CONCLUSIONS AND RECOMMENDATIONS
! As a result of considering eight different types of transmissions, the
variable ratio, power-splitting, hydromechanical transmission was selected
as the most promising near-term (1974) transmission for application to both
the single-shaft gas turbine and the turbo-Rankine engine. With future de-
velopment, the traction type of transmission was considered a promising
alternate on a long-term basis.
2. The selected transmission design was 28 percent smaller, 3 percent heavier
and required only 6 percent more parts than a comparable conventional auto-
matic transmission.
3. The selected transmission was compatible and feasible for both the single-
shaft gas turbine and turbo-Rankine engines. Performance analysis demon-
strated that the transmission maintained engine operation (SFC) .along a
prescribed operating line (for maximum miles/gallon) and the resulting smooth
acceleration characteristics met all EPA/AAPS specified vehicle design goals.
4. Production cost is a key factor in considering a transmission for automotive
applications. A detailed cost analysis, based upon costing procedures of
the Ford Motor Company, showed that the "original equipment manufacture"
(O.E.M.) cost of the selected transmission was 30 to 40 percent higher than
a present day automatic transmission in production quantities of 1,000,000
units/year. On a "variable cost" basis (more meaningful to the automotive
companies in comparing designs) the increase in cost was 44 percent. Higher
material cost was the major factor in causing the increased cost. However,
.since the weight of the selected transmission was only slightly greater
than a comparable automatic transmission, this suggests that, with development
of production techniques, future production costs for such a transmission
could approach that of a present day automatic transmission.
36
-------
Detailed performance analysis showed that the resulting efficiency of the
selected transmission at cruise power was 57 to 60 percent at 20 mph, 66 to
74 at 30 mph and 90 percent at 50 mph or higher speeds. Average transmission
efficiency over the Federal Driving Cycle was 71 and 74 percent with the
turbo-Rankine and single-shaft gas turbine, respectively. A comparable con-
ventional automatic transmission (powered by a present-day automotive engine)
has been estimated to have an average efficiency of 77.6 percent over the same
driving cycle. Thus, it was concluded that the selected transmission had
somewhat lower efficiency than the automatic transmission at low cruise-power
levels (below 25 mph) and this was primarily caused by speed dependent losses.
The fuel economy calculated for the selected transmission with the single
shaft gas turbine showed clear advantages relative to a conventional auto-
matic. The automatic transmission is, of course,, incapable of (a) following
the critical minimum SFC line for the gas turbine and (b) providing satis-
factory power to the wheels over the vehicle speed range required. The con-
tinuously variable transmission provides both capabilities very well.
The fuel economy calculated for the selected transmission with the Turbo-
Rankine engine showed no advantage relative to a conventional automatic.
Although the automatic cannot closely follow the minimum SFC line for this
engine, there is no requirement that it should. The turbo-Rankine engine has
relatively flat engine characteristics and neither fuel economy nor. kinematic
performance is sensitive to speed.
Although the selected transmission provided smooth acceleration and optimum
engine operation for the turbo-Rankine engine, it was concluded that, since
there was little improvement in fuel economy compared to a conventional
automatic transmission, further development of the transmission for that
engine application is not advantageous. However, if future turbo-Rankine
engine design developments result in an engine characteristic where there is
a significant change in SFC with engine speed, then development of the selected
transmission for that engine would be beneficial.
37
-------
9. Further development of the selected variable ratio, hydromechanical trans-
mission is recommended for application with automotive single-shaft gas-
turbine engines (and other similar types of engines) since, in addition to
smooth vehicle operation, maximum fuel economy will be obtained.
38
-------
SECTION V
SUPPORTING INFORMATION
39
-------
A. SELECTION OF TRANSMISSION
The initial step in this study was to select, from a wide range of transmission
types, the most suitable candidate transmission type for use with the specified
single-shaft gas turbine and turbo-Rankine engine in a power train applicable
to a medium-sized family car. One of the more important criteria, specified by
EPA/AAPS, in the selection process was that the selected transmission should be
available for engine testing early in 1974.
Included for consideration were the following seven basic types of transmissions:
Mechanical
Hydrostatic
Combination of mechanical and hydrostatic
Hydrokinetic
Electrical
Traction
Belt Drive
As will be discussed, these and other variations were considered in reaching a
point of establishing the techical and economic feasibility.
Some of the basic transmission types can be ruled out on a qualitative basis
because of the inherent limitations which they have in this application. The
mechanical gear type was ruled out, since it does not provide the infinitely
variable ratio with stepless changes that are required for smooth, efficient
operation. The hydrostatic and the electric transmission overcome this limi-
tation. They can be considered to be similar in that they can provide an
infinitely variable ratio with smooth operating characteristics. Both suffer
from the limitations of high weight, relatively high cost and efficiency lower
than that attainable with other transmission types. In the case of the
electrical transmission, one often omitted consideration is the cost and weight
of associated controls and power-conditioning equipment.
Once having eliminated these transmissions from active consideration, the above
list of basic types was expanded to include the following:
A-l
-------
Three-speed Automatic with Variable Element
Conventional Hydromechanical
Advanced Hydromechanical
Traction - TRACOR
Traction - Power Splitting
Traction - Composition Belt
Traction - Metal Belt
Three-speed Automatic with Aerodynamic Torque Converter - ROHR
The standard three-speed automatic transmission was used in the latter
part of the evaluation as the datum for performance comparisons. There
is also merit in its consideration from the standpoint that it is widely
used currently with the standard 1C engine.
The candidate transmissions are described in the following paragraphs.
Transmission Descriptions
In this section each of the candidate transmissions is shown followed by a
brief description.
1. Three-Speed Automatic Transmission with Variable Element
Single -
Shaft
Turbine
Engine
Gear
Ratio
Variable
Torque
Converter
Secondary
Planetary
Primary
Planetary
Rear
Axle
A-2
-------
The standard three-speed automatic transmission is superior to a straight
mechanical gear-type transmission. However, the steps or shift requirements
still impose a penalty. The addition of a variable element would provide an
infinitely variable ratio without steps. This basic arrangement is shown
above.
A gear ratio is required ahead of the transmission to reduce engine output
speed to approximately 4000 rpm which is a practical input speed for the
torque converter.
A variable hydrodynamic torque converter makes use of a reactor element which
is varied in position to extend the normal torque multiplication range of the
standard hydrodynamic torque converter. The additional reactor element is
simple and inexpensive. The impeller and turbine members are similar to those
used in existing torque converters and thus adaptable to the same relatively
simple manufacturing and assembly techniques.
The remainder of the construction is similar to the standard three-speed
automatic transmission.
This configuration has the obvious advantage of maintaining the basic configu-
ration of and the majority of the parts of the standard automatic, although an
additional sub-assembly is required. Most particularly the poor efficiency
of the variable torque converter element was considered to rule out this
candidate.
2. Conventional Hydromechanical
Single-
Shaft
Turbine
Engine
Gear
Ratio
Gear
Ratio
Variable
Hydraulic
Element
Variable
Hydraulic
Element
Gear
Ratio
Low-
Range
Planetary
High Range Mechanical Path.
High-
Range
Planetary
Rear
Axle
This transmission utilizes a simple hydrostatic pump-motor circuit in combin-
ation with a planetary gear" train to provide a wide continuously variable
operating range.
A-3
-------
The design has two operating ranges. In the low-output-speed range, the
hydraulic circuit operates as a straight hydrostatic system driving through
the planetary gear set. This provides high output torque and variable opera-
tion in both the forward and reverse directions. In high range, the planetary
gear set and hydrostatic circuit function as a split torque or hydromechanical
system. This extends the range of output speed, increases the efficiency, and
also provides positive control. The low range is achieved by braking one of
the planetary elements. The high range is achieved by disengaging the brake and
engaging a clutch which closes the high range mechanical path.
By utilizing a combination hydrostatic-hydromechanical construction, the
range and pressure level over which the hydrostatic circuit must operate are
minimized. This substantially reduces the displacement and size of the hydro-
static elements, which increases the efficiency and at the same time provides
a smaller and more compact design.
In the design, the planetary gear train is constructed so that the clutch and
brake elements are in synchronization during transition from one stage to the
other. This eliminates high inertial loads on the clutch and brake, minimizes
slippage and wear, and provides positive and even drive over the entire operating
range.
3. Advanced Hydromechanical Transmissions
Single-
Shaft
Turbine
Gear
Ratio
Variable
Hydraulic
Element
Fixed
Hydraulic
Element
Low-
Range
Simple
High Range Mechanical Path * *-""*
High-
Range
Simple
Planetary
Rear
Axle
The advanced hydromechanical operates on basically the same principles described
above for the conventional hydromechanical. However, differences to be observed
are greater simplicity in design and controls, fewer parts, greater reliability
and lower cost. Note in particular the elimination of the gear ratios at input
to and output from the hydraulic elements.
The hydraulic units are operated at the reduced pressure of 3600 psi maximum and
speed of 3500 rpm maximum to achieve several objectives.
Minimize noise levels which must be carefully considered
in this type of design.
Achieve reliability by minimizing piston loading.
This design approach, however, does result in larger hydraulic units which
impose some weight and cost penalties, and some reduction in efficiency.
The noise problem, while minimized by the reduced pressure, may require
additional insulation or isolation to achieve driver acceptability.
A-4
-------
4. Traction Drive - TRACOR
Single-
Shaft
Turbine
Engine
Gear
Ratio
Toroid
Drive
Primary
Planetary
Torque
Converter
Secondary
Planetary
Rear
Axle
The TRACOR traction drive, which has progressed beyond the model stage, was con-
sidered as typical of all traction devices. A transmission schematic, shown
above, was made and evaluated using the TRACOR traction drive.
The essential element of the traction transmission is a toroid drive which pro-
vides a continuously variable ratio by changing the relative radii at which
power is delivered to and taken from a set of rotating disks. Fig. A-4 presents
the important details of the mechanism.
The theoretically available range of torque ratio from input to output varies from
3:1 to 1:3 - a factor of 9 variation. TRACOR has proposed an automotive gas
turbine transmission with a factor of 6 variation, which from the present studies
would appear very adequate. Continuous variability over this ratio with the toroic
drive results in a very high maximum output speed from the drive, and the im-
plications of this high speed are discussed further below.
Since the toroid drive itself is more effective when operating at high speeds,
the gear ratio shown ahead of the drive is a very simple low reduction gear
set. It is very possible that this gear reduction could be eliminated entirely.
The TRACOR traction drive must be disconnected to change ratio when the output
is at zero speed, and does not have internal provisions for reverse operation.
A torque converter was added to allow disconnection of the traction drive from
the rear end when at zero output speed. The torque converter provides an
additional benefit of further increasing the available torque ratio range.
Since the TRACOR traction drive operates at a high speed, a primary planetary
gear train to reduce the speed to that practical for the torque converter input
was required. A second planetary gear set was added to provide transmission
output at the appropriate speed and torque and to provide the reverse operation.
A-5
-------
The advantages of the TRACOR transmission are the wide ratio range and
relatively quiet operation. The disadvantages are the need for an additional
planetary set and the torque converter, and the associated efficiency reduction.
In addition, since a complete traction transmission has not been built for this
power range, significant development effort is to be anticipated before this
transmission type could be available for high production automotive applications.
5. Traction Drive - Power Splitting
Single-
Shaft
Turbine
Engine
Gear
Ratio
Toroid
Drive
Clutch
Low-
Range
Simple
Planetary
High-
Range
Simple
Planetary
Rear
Axle
During the transmission study, several purely conceptual transmissions were
reviewed. One was the power-splitting traction transmission.
Since the existing larger size traction elements do not directly provide idle
and reversing, an alternative scheme was generated which added this capability
and also reduced the amount of power going through the traction drive. This
reduction of power makes possible the use of a smaller traction element. The
block diagram of such a transmission is shown above.
The operation of the transmission is conceived as very similar to that of the
hydromechanical power-splitting transmission; that is, the traction units per-
form the same function as the hydraulic units.
In the low output horsepower and low-speed range, the traction circuit oper-
ates as a straight traction system driving through the planetary gear set.
This provides high output torque and variable operation in both the forward
and reverse directions. The clutch, which is shown in the diagram, is used
during the idl.e periods and during the switching from forward to reverse.
This clutch may not be necessary since several of the manufacturers of smaller
traction units claim that their units have built-in features that permit
reversing and operation with zero output speed.
A-6
-------
In the high range the planetary gear set and traction drive function as a
split torque system. By utilizing a combination of traction drives and
planetaries, the range of power over which the traction drive must operate
is minimized. With the proper combination of planetaries and traction
drives, the power through the traction drive can be reduced to as low as
15 percent of the output power. Thus, smaller and much more versatile
traction units can be used.
The remaining elements of the transmission are similar in operational and
constructional features to the hydromechanical power-splitting transmission.
Since this is a conceptual transmission only, it should not be considered
for near-future application. One disadvantage in common with the power
split hydromechanical is the likelihood of more parts than a conventional
automatic. However, the high-efficiency, high-speed operation and low-noise-
level potential of the traction element do encourage its exploration for a
transmission of the future.
6. Traction Drive - Composition Belt
Single-
Shaft
Turbine
Engine
Gear
Ratio
Clutch
Comp.
Variable
Belt
Drive
Secondary
Planetary
Primary
Planetary
Rear
Axle
Transmissions using composition belt drives as the variable elements have been
used in many low-horsepower, off-the-road vehicles. Most of these drives have
operated on vehicles requiring less than 50 horsepower. Several companies are
now considering the belt transmission for automotive application. These studies
are proprietary and were not made available for this review.
A possible approach is presented in the above diagram. Here the variable-
speed belt drive is used as the hydraulic units are used in the hydro-
mechanical transmission.
A-7
-------
A gear train is required to reduce the gas-turbine output speed to 6,000
rpm, which is a practical speed for variable speed belt drives.
Existing variable-speed belt drives do not operate effectively at zero
output speed (belts cannot be "slipped"); therefore, a clutch is provided
to protect the belt drive during idle operations. As the variable-speed
belt drives are developed for higher horsepower, the clutch may not be
required or may be integral with the belt drive.
The scheme presented is simply a concept and has not been reduced to a
design. Several of the areas that require exploration are:
1. The belt material and construction required to provide a
life of 3500 hours. (The load, heat, speed and rate of
speed change all tend to limit the life of existing belt
transmissions.)
2. The packaging of the belt drives into an envelope suitable
for automotive application.
3. The ability to achieve high efficiencies predicted by manu-
facturers of variable-speed belt drives, especially when
operating at the extremes of the speed and torque range.
4. The probability of the composition belt acting as a traction
drive and not a friction drive. (The wear associated with
friction would severely limit the load capacity and life.)
5. Design approaches consistent with maintenance requirements.
A-8
-------
One other approach would be to replace the torque converter of an existing
three-speed automatic transmission with a variable-speed composition belt
drive. Such a scheme would require that the belt drive handle more power
than the power-splitting approach. The advantages of using most of the
existing automotive transmission would make such an arrangement a worth-
while investigation.
Although there are many areas of the belt transmission that require explor-
ation, the possibility of achieving transmission efficiencies of around
92% over a broad operating range suggests that a preliminary design study
should be started.
7. Traction Drive - Metal Belt
Single -
Shaft
Turbine
Engine
Gear
Ratio
Metal
Belt
Variable
Drive
Clutch
Secondary
Planetary
Primary
Planetary
Rear
Axle
The composition-belt and metal-belt transmission are very similar in design.
However, the metal belt can be considered a traction-type element and, in
fact, several companies have built prototype transmissions using the belt
as the traction element.
The diagram is an arrangement which proposes that the torque converter of
an existing, three-speed, automatic transmission be replaced with a metal-
belt variable drive. A clutch is provided for the idling mode when the
output of the variable drive is at zero speed.
Many of the factors requiring exploration for the composition belt are not
a problem with the metal belt. However, the packaging of a metal-belt drive
into an envelope suitable for an automobile and the possibility of achieving
the high efficiencies predicted by the manufacturers are two areas that have
not been resolved.
A-9
-------
The arrangement presented above is a concept and has not been reduced to a
design. Therefore, additional study is required before a meaningful evalua-
tion can be accomplished.
8. Three-Speed Automatic Transmission with Aerodynamic Torque Converter
ROHR INDUSTRIES, INCORPORATED
Single-
Shaft
Turbine
Engine
Aerodynamic
Torque
Converter
Modulated
Clutch
Gear
Ratio
Secondary
Planetary
Primary
Planetary
Rear
Axle
An automotive transmission for use with a gas-turbine engine has been proposed
by Rohr Industries, Incorporated. A block diagram of the proposed transmission
is shown above. The normal hydraulic torque converter used with existing auto-
motive automatic transmissions has been replaced with an aerodynamic torque
converter (ATC) as developed by Rohr Industries, Incorporated. Further, the
ATC is shunted by a modulated clutch.
The ATC resembles the hydraulic torque converter in operation except it operates
directly at the high speeds associated with the gas-turbine engine and does not
require a geared reduction between the engine and the torque converter.
Since the ATC operates most effectively at high speeds, a gear reduction between
the ATC output and the planetary gears is required. An additional reduction
would be used to drive the required accessories at their normal operating speeds.
To perform adequately throughout all vehicle modes at good transmission effi-
ciency, the ATC is shunted by a modulated clutch. By judicious pressurization
of the ATC in combination with engagement of the modulated clutch, satisfactory
performance is envisioned to be achieved. The secondary and primary planetaries
are similar to those used in automotive power-shift transmissions. They provide
the additional torque ratio required and the reversing mode of operation.
A-10
-------
In considering a gas-turbine engine, there are several advantages when
using such a transmission:
1. The ATC can operate quite efficiently and effectively at the
high speeds of a gas turbine.
2. The ATC allows the gas turbine to operate at the most
efficient speeds required to satisfy the road load.
3. The ability of the ATC to regenerate its losses to the gas
turbine by return of coolant flow poses the possibility of
recovering some of the transmission losses.
4. The components all use existing mass production techniques.
Also, there are several questions that must be considered:
1. When considering system overall efficiency throughout the
driving cycle, would a free power turbine and a three-speed
shift transmission be more efficient than the ATC transmission?
2. Can an ATC transmission be arranged which would not require
an evacuation pump, modulated clutch, additional reduction gears,
and high speed seals; thereby, eliminating the increases in
weight and loss in reliability and efficiency associated with
such components?
3. Can the control of the ATC and return of coolant flow be
accomplished without additional complexity?
A-ll
-------
Although the initial development of the ATC is complete, the development
of a transmission using the ATC has not started. Therefore, such a trans-
mission should not be considered as a possibility for application during
the 1974 to 1975 period.
Selection Procedure
The first step in the selection was to define the evaluation factors.
These are given in Table A-l below:
TABLE A-l
EVALUATION FACTORS
1. Life and Reliability.
a. Number of parts
b. Application of parts
c. Design Simplicity
2. Noise and Smoothness
3. Cost.
a. Number of parts
b. Size
c. Weight
d. Manufacturing pre-
cision required
4. Development Status a. Paper Study
b. Complete Transmission
Built & Tested
c. In production
5. Efficiency
6. Size and Weight
7. Restriction on Turbine Engine a. Performance capability
A-12
-------
(Table A-l continued)
8. Control Complexity
9. Drive Acceptability
10. Environmental Restrictions a. Shock
b. Temperature
In applying these factors, it was important to bear in mind the short-
term implementation requirement (1974), which makes factor 4, Development
Status, one of major importance.
Although many of the transmissions must necessarily be treated somewhat
qualitatively and subjectively, it was important that some objective pro-
cedure be adopted for making the final comparisons and selection using
the factors shown in Table A-l . Table A-2 shows an unweighted overall
evaluation summary of these transmissions in which a score of 1-10 has been
assigned for each factor. It will be noted from Table A-2 that the three-
speed automatic with a variable element, the advanced hydromechanical and
the conventional hydromechanical rated very close to each other.
These transmissions were reviewed to determine their major strengths and
deficiencies. Table A-3 gives a short synopsis of the main advantage and
disadvantage of each of the transmissions.
Discussion
The evaluation of transmission types previously discussed shows that the
hydromechanical and traction-type transmissions were most worthy of con-
sideration. It is appropriate, therefore, to present some additional com-
parisons between these two basic approaches.
A-13
-------
TABLE A-2
EVALUATION SUMMARY
TRANSMISSION
TYPE
OVERALL RATING
(Unweighted)
1. Three-speed Automatic
with Variable Element
2. Advanced Hydromechanical
60
63
3. Conventional Hydromechanical
58
4. Traction-TRACOR
5. Traction-Power Splitting
48
35
6. Traction-Composition Belt
43
7. Traction-Metal Belt
32
8. Three-speed Automatic with
Aerodynamic Torque Converter
Rohr
28
9. Electric - Alternator and
Motor
21
A-14
-------
TABLE A-3
EVALUATION SUMMARY
TRANSMISSION TYPE
MAIN ADVANTAGE
MAIN DISADVANTAGE
1. THREE SPEED AUTOMATIC
WITH VARIABLE ELEMENT
SIMILAR COMPONENTS TO EXISTING
AUTOMATIC TRANSMISSION
LOWER EFFICIENCY THEN
EXISTING AUTOMATIC
2. ADVANCED HYDROMECHANICAL
SIMPLE MECHANICAL CONSTRUCTION
PRODUCTION TECHNIQUES WILL
HAVE TO BE DEVELOPED
3 CONVENTIONAL HYDROMECHANICAL DEVELOPMENT EXPERIENCE EXISTING
PRODUCTION TECHNIQUES WILL
HAVE TO BE DEVELOPED
4. TRACTION - TRACOR
FUTURE POTENTIAL FOR HIGH
EFFICIENCY
NOT TO DEVELOPMENT STAGE
5 TRACTION - POWER SPLITTING
FUTURE POTENTIAL FOR HIGH
EFFICIENCY
NOT TO DESIGN STAGE
6. FRICTION - COMPOSITION BELT
BELT MANUFACTURING TECHNIQUES
DEVELOPED
LIFE AND EFFICIENCY NOT
FULLY DETERMINED
7. TRACTION - METAL BELT
FUTURE POTENTIAL FOR HIGH
EFFICIENCY
NOT TO DEVELOPMENT STAGE
8 THREE-SPEED AUTOMATIC
WITH AERODYNAMIC TORQUE
CONVERTER-ROHR
MANUFACTURING TECHNIQUES
DEVELOPED FOR BULK OF THE
TRANSMISSION
NOT TO COMPLETE TRANSMISSION
DESIGN STAGE
-------
Several different kinds of traction devices were studied. Some of these
claimed:
1. Infinitely variable ratio.
2. Ratio change with the output member at zero speed (eliminating need
for clutch or torque converter).
3. Continuous operation through idle conditions (eliminating need for
special reversing element).
These drives were, at most, only to the very preliminary small-size-model
stage. As a result, they could not be evaluated for the present study. There-
fore, the TRACOR traction drive, which has progressed beyond the model stage,
was considered as typical of all traction devices.
Performance data for the traction drive was provided by the TRACOR Company
and is presented in Figure A-l. Using the traction drive data, MTI estimated
the performance of a traction transmission such as that shown in the diagram
on page A-5. The performance data calculated in this manner is shown in
Figure A-2.
Figure A-3 compares the performance with a comparable hydromechanical trans-
mission. The results indicate that the efficiency of the hydromechanical
is superior, particularly at lower power levels. It should be noted that
this is largely due to the high-speed ratios and the use of a torque con-
verter with this traction transmission concept.
Selected Transmission
Based upon these reviews, the hydromechanical (power-splitting) transmission
was selected. The key features of the selected transmission are:
Simple Construction
Proven Components
Low Cost Potential
Minimal Development
Size and Weight Compatible
A-16
-------
I
H-*
~^J
100
80
60
40
20
2 ROLLER - 4.5 INCH DIAMETER
7000 RPM - INPOT SPEED
100 HP
25 HP
10 HP
.5 1.0 1.5 2.0 2.5 3.0
TRANSMISSION OUTPUT/INPUT SPEED RATIO (Nln/Nou{.)
Figc A-l TRACOR Traction Drive Performance Data
-------
I"
00
16
RATIO OF TRANSMISSION INPUT TO WHEEL SPEED (Ne/NQ)
JL
' i
_L
90 66 40 30
20
10
Vehicle Velocity, MPH
igc A-2 Estimate of Traction Transmission Performance
MTI-14661
-------
fl
PL,
£
§
M
o
l-l
£
w
100
80
60
40
20
^^~ Hydromechanics 1
Traction
VEHICLE WT = 4600 LBS
FUEL DENSITY =6.30 LB/GAL
REAR END T\ = 0.96
REAR END RATIO =3.08
M
'8
S
10
20
30 40 50
VEHICLE VELOCITY - MPH
60
90
80
90
Fig. A-3 Comparison of Hydromechanical and Traction Transmissions
MTI-14606
-------
fo
o
. I
U1 N>
CD
SPRING
BIAS
CONTROL LINKAGE
OIL FOR
DRIVE &
SERVO
SYSTEM
PRECESS CAMS
ROLLER LOAD CAM
FOUR WAY
SPOOL VALVE
COMMAND
CYLINDER
SERVO
SUPPLY
PRESSURE
A- INPUT
SHAFT
OUTPUT
SHAFT
HYDRAULIC
THRUST
BEARINGS
TOROIDAL
DISCS
ROLLER
CONTROL
PISTONS
Fig. A-4 Traction Drive Schematic
XTI-U601
-------
B. DESCRIPTION OF TRANSMISSION & CONTROLS
This section contains a description of the selected hydromechanical transmission
and its associated controls. Design details have been omitted, since the
transmission is based upon a proprietary design of Mr. George DeLalio. In the
following discussion, some of the features of the transmission are described
with a subsequent discussion of the controls.
The selected hydromechanical transmission is an infinitely variable, stepless
unit that obtains torque multiplication and control by means of hydraulic
elements (pump-motor combination). The unit differs from the conventional
torque converter or fluid-coupling-type transmission in that the hydraulic
power is transferred by fluid static pressure at low flow as contrasted to
the high dynamic action of fluid as utilized in hydrodynamic units. It also
differs from a purely hydrostatic transmission in that the much more efficient
mechanical elements transmit a significant portion of the power. It.is a
"hard" type of drive in that slip is less than 2 percent under full load.
1. Description of Transmission
As pointed out earlier in this report, the design of the selected hydro-
mechanical transmission was based upon minimizing the number of mechanical
parts in order to: a) make the transmission as simple as possible, b) keep
the size, weight and, particularly, the cost low, c) require the minimum
amount of development, and d) achieve high reliability while retaining a rea-
sonably high transmission efficiency. Additional gear trains and hydraulic
functions could have been employed to increase the capabilities of the trans-
mission and reduce the amount of power in the hydraulic elements; however, it
was considered more realistic to concentrate on a unique, straight-forward
simple design in order to minimize development time.
Figure B-l is a functional schematic of the selected transmission. As shown on
this figure the transmission consists of an engine input shaft, and ah output
carrier and shaft. Other essential components are a brake for locking one
element of the planetary to achieve low range operation; a clutch for closing
the high speed mechanical power path; and a control system for varying the
displacements of the elements, for engaging, and for shifting range.
B-l
-------
W
SHIFT SELECTOR LEVER
THROTTLE POSITION »>
TURBINE
SHAFT
MEASURED
ENGINE
SPEED
I I
I
CONTROLS
VARIABLE
DISPLACEMENT
ELEMENT
FIXED
DISPLACEMENT
ELEMENT
SIMPLE
PLANETARIES
OUTPUT TO VEHICLE
Fig, B-1 Functional Schematic Selected Hydromechanical Variable-Ratio Transmission
KTI-14628
-------
As shown by Figure B-l, the power path between input and output splits be-
tween hydrostatic and straight mechanical. The transmission has two operat-
ing ranges designed as low-range and high-range. At low power and output
speed levels (up to approximately 0.4 of maximum output speed) the transmission
operates in the low range and all power flows through the hydraulic elements.
At higher power and speed conditions the transmission operates in high range
and the power path is split between the hydraulic elements and the mechanical
path. At approximately 0.7 of maximum output speed, all the power flow is
through the mechanical path.
An important feature of this design is the synchronous shift which operates
as follows: The transmission has two operating ranges, low speed and high
speed. Transfer between these two ranges is effected by the concurrent open-
ing of a brake and closing of a clutch. The shift is ideally synchronous if it
occurs under the following conditions:
1. In either low or high range the displacement of the variable
hydraulic element is at the same maximum point.
2. There is no change in relative velocity between the two sides
of the clutch.
3. The braked element is stationary whether the brake is applied or
not.
In practice such effects as slip between hydraulic elements and control
imperfections cause slight deviations from the ideally synchronous shift.
Even so there is never a significant change in momentum demanded of the
rotating parts and the motor and clutch elemnts suffer little slippage or
wear.
The introduction of the Mechanical connection reduces the range over which
the hydraulic elements must operate at 100 percent power level to approximate-
ly 40 percent of the total speed range. Since the mechanical path elements
are more efficient, compact, and less costly to produce,.this approach pro-
vides an optimum wherein the transmission is completely variable over its
range, the size of the hydraulic elements is minimized, the mechanical con-
struction is kept simple, and the range change is synchronous without any
steps, slippage or wear of brake and clutch elements.
B-3
-------
The maximum torque ratio available when the power path is through the
hydraulic branch is 5.0 from engine input to transmission output and 15.4
from engine input to rear-axle output when a standard rear-axle ratio of
3.08:1 is used.
The maximum torque ratio available when the power path is through the
mechanical branch is 2.5 from engine input to transmission output and is
7.7 from engine input to rear-axle output.
In the hydraulic units, the movement of the swashplate, which controls the
pump displacement, is a continuous function; therefore, the transmission
ratio control is completely stepless and infinitely variable within the
operating range.
A preliminary design layout of the transmission was made in sufficient detail
in order to determine size, weight, number of parts, and cost. Figure B-2
presents the design layout without the proprietary design details.
In the design, all the gears are made of automotive gear materials such as
forged and surface hardened AISI 8620 steel and are machined to an AGMA gear
quality level of 8 with a final polish. The manufacturing is accomplished
using automotive practices.
The two hydraulic elements are based upon a design originated by Dr. H.
Ebert* and recommended by Mr. G. DeLalio. They are considered to be of a
high power-density construction which produces the maximum capacity using
the minimum volume and weight. Typical power densities are approximately
20 horsepower per cubic inch. The construction of the unit is shown by Figures
B-3 and B-4. The compact size of these hydraulic elements is achieved by
using a rolling-element bearing for the swashplate bearing** and close coupling
it to the drum.
* Independent consultant. Dr. Ebert formally supervised the development of
hydromechanical transmissions at Daimler-Benz, Austin, NSU and Allgaier in
Germany.
** The design of the swashplate bearing was established by many hours of test-
ing in Germany and at the Stratos Division of the Fairchild Corporation.
B-4
-------
HYDRAULIC ELEMENTS
INPUT
SHAFT
OUTPUT
SHAFT
CONTROL HOUSING
0 I
8.75
SCALE (INCHES)
Pig. B-2 Preliminary Design Layout of the Hydromechanical Transmission
B-5
-------
Fig. B-3 Typical Axial Piston Hydraulic Element
B-7
-------
(.V\\H)
W
00
SCALE
Fig. B-4 Drawing of High-Density Hydraulic Element (SK-C-4266)
KTI-12520
-------
This type of hydraulic element has been successfully used in selected
applications for aircraft constant-speed drives, tractor transmissions,
truck transmissions and in postal vehicle transmissions. For example,
these types of hydraulic elements were specifically used in the following
programs conducted by the Stratos Division of the Fairchild Corporation:
1. 25-horsepower, constant-speed drive electric supply and hydraulic
pump for Fairchild Goose Missile.
2. 200-horsepower hydromechanical transmission for trucks evaluated
by Detroit Arsenal.
3. 150-horsepower hydromechanical transmission for M-34 trucks eval-
uated by Detroit Arsenal under Contract DA-30-069-ORD-2340.
4. 50-horsepower hydrostatic transmission for off-road vehicles
evaluated by Detroit Arsenal.
Tests conducted in various applications of the hydraulic elements have
established the requirements that must be met to achieve long life. These
3
results have shown that, for the?.5 in /rev size element, the system should
operate at pressures of 2500 psi or lower and speeds of 3000 rpm or lower
for 90% of the load schedule if a life greater than 3500 hours is expected.
Pressures to 3500 psi and speeds of 350 rpm for 10% of the load schedule
will not reduce the life of the elements below the 3500 hours.
With regard to structural aspects of the transmission, the main housings,
control housings and mounting plates will all be aluminum pressure die
castings. Automotive practices of thin-wall design, intricate sections
for less machining, high strength, maximum heat disipation and favorable
economics will be followed.
The high-range clutch and low-range brake are of conventional automotive
construction used in existing transmissions.
The automotive practice of using sleeve bearings or needle bearings to
support the radial and thrust loads has been followed.
B-9
-------
As in the case of the gears, the shafts were constructed using a forged
steel similar to AISI 8620. The bearing raceways and splines are hardened.
The displacement of the hydraulic unit is varied using a piston actuator
which moves a cam plate linked to the trunnion and swashplate of the ele-
ment. The actuator is similar in construction to that of the automobile
transmission actuators used to engage clutches and brakes. The cam plate
is steel with hardened cam tracks. A cam roller bearing is used to link
the cam to the trunnion.
2. Description of Transmission Controls
A control hardware implementation schematic is shown in Figure B-5.
The basic components are:
Shift lever
Control cams
Engine regulator governor
Engage governor
Control valve
Spool valve
Shuttle valve
Clutch valve
Element I swashplate cam
Swashplate piston actuator
Engaging valve
The manual shift lever is used to select the operating mode. The shift lever,
through control cams, acts directly on the control valve, spool valve and
regulator governor. A bias is applied to the regulator governor spring
while the control valve and spool valve, which are positioned by the shift
lever, direct pressures to the control elements as required for each mode.
B-10
-------
Fig. B-5 Hydromechanical Transmission Control Schematic
B-ll
-------
The engaging valve has several functions. It is a charge valve for the
pump and motor, a bypass valve for the pump, and is also a fast-acting
relief valve for the pump and motor.
The bypass function is controlled by the engage governor. The governor
is geared to the transmission input shaft to sense engine RPM and give
a positive neutral. When the engine RPM is too low, the governor causes
the valve to remain in the position to bypass. The bias pins and spring
react against the housing to open the engaging spool valve. As the speed
of the engine and the engage governor pressure increase the engaging
spool valve is moved to engage the hydraulic elements of the transmission.
As the hydraulic system working pressure builds up, it further reacts
against the bias pins which tend to open the engaging spool valve and
effect a smooth modulated engage action as a function of speed and work-
ing pressure.
The hydraulic pomp and motor relief valves, which are also located in the
engaging valve housing, are set at approximately 3600 psi. Theecontrol
system should hold the pressures below 2500 psi during normal operation
for maximum efficiency and minimum wear. Pressure surges are held below
the relief valve settings, during normal operation, by limiting the rate
at which the actuator varies postion. Shock loadings, accidental shift-
ing into reverse, or overloading greater than the full torque-speed ratio,
will not harm the- transmission, since the pressure relief valve is of a
low inertia construction which prevents over pressurization of the hydraulic
system.
The charge pump is geared to the transmission input shaft and supplies
charge pressure to the low side port of the pump and motor through the
charge valves.
The discussion has been limited to the controls required to accomplish
the primary functions of the transmission.
B-13
-------
Although not discussed, additional hydraulic elements, and controls, for
secondary functions such as bypass valves, flow restrictions, pressure
regulators, filters, and interlocks are required.
Operation
The shift pattern schematic as shown below follows the present day auto-
mobile shift pattern as closely as practical. The control sequences for
the various modes are discussed below.
P R N D 2 1
P - Park
R - Reverse
N - Neutral
D - Drive
2 - Second
1 - Low
Neutral
In neutral, pressure is directed by the control valve to the piston actua-
tor to move the sleeve to the position which limits piston stroke. Pres-
sure is transmitted through the shuttle valve to engage the;low-range brake.
The engaging valve is in the open position and the swashplate of Element I
is near zero displacement, which allows this element to rotate freely to
effect zero output torque. A cam, driven by the shift lever, sets the
;
engine regulator governor bias load to overcome the flyweight force, thus
moving the governor valve. The resultant position of the governor valve
provides flow to the actuator which moves the piston to the maximum torque
position. Accordingly, in neutral, the control system presets the primary
swashplate and low-range brake for low-ratio output for initial vehicle
acceleration from a stopped position.
B-14
-------
Drive
When the control valve is shifted to the drive position, the fluid pres-
sure connections to the low-range brake and high-range clutch remain the
same as in the neutral position. Pressure from the engage governor is
transmitted through the spool valve to the engage valve. The output
pressure from the engage governor increases as engine speed increases.
This pressure acts on the engage valve to ..engage the transmission at
2100 to 2300 rpm transmission input speed.
During drive operation the engine regulator governor controls hydraulic
flow to the opposite sides of the actuator piston to effect movement of
the Element I swashplate. The movement of the governor valve is regu-
lated in part by the position of the cam which adjusts the speed setting
as a function of the accelerator pedal.position. Accordingly, for every
throttle position, the governor valve mechanism continuously controls the
position of the actuator piston to vary the operating ratio to maintain
a set engine speed. This provides a means for ideally matching the engine
and vehicle speeds as a function of throttle position to provide optimum
engine performance.
It will be noted that, in initially accelerating, the clutch valve rides
upon the upper surface of the control cam, thereby providing fluid pres-
sure through the shuttle valve to engage the low-range brake. When the
transmission output speed increases, movement of the cam causes the
clutch valve to move down to the lower surface of the cam. In the extended
position of this valve, the fluid pressure to the low-range brake is vented,
thereby releasing the low-range brake. Also, when this valve is extended,
fluid pressure is provided to the high-range clutch, thereby engaging this
clutch. This transition is made while both the high-range clutch and low-
range brake are in synchronization to effect very smooth operation. This
provides wear-free operation.
B-15
-------
In Mode 2 the control cam introduces a bias to the engine regulator governor
which varies the engine speed setpoint at which the transmission switches
from low to high range and, in effect, holds the transmission in the low
range. In Mode 1 (Low) a further bias is introduced, enforcing an even
higher minimum torque ratio.
Reverse
In reverse, fluid pressure is transmitted from the control valve through the
shuttle valve, to engage the low-range brake. The control cam is positioned
by the mode selector to provide additional force on the engine regulator
governor spring thus moving the engine regulator governor valve downward and
providing fluid pressure to the bottom of the piston actuator. Pressure in
the top chamber of the actuator is vented through the control valve. Thus,
the piston and sleeve of the actuator both move to the top position. The
additional stroke of the cam plate due to movement of the sleeve causes the
pin connected with the swashplate housing to move into the negative angle
portion of the cam slot, thereby positioning Element 1 swashplate at a
negative angle for reverse output.
It will be noted that, when the control valve is moved to the reverse
position, fluid pressure is directed to the engage, valve so as to engage
the hydraulic elements in the transmission.
Park
Park, which opens the engage valve to bypass fluid around the hydraulic
motor and pump, permits the engine to idle without transmitting any torque
to the transmission output shaft. The engine idles at 2100 rpm. In
addition, the parking lock is positioned to, lock the transmission output
shaft to the rear wheels.
B-16
-------
C. DESCRIPTION OF METHODS FOR DETERMINING PERFORMANCE
This section presents: 1) a description of the method employed to determine
transmission performance, 2) a description of the computer program used for
determining steady-state performance, 3) a description of the computer pro-
gram used for driving-cycle analysis, and 4) a description of the computer
program for simulation of full-power acceleration of the vehicle (used for
calculating maneuver performance).
1. Method for Determining Transmission Performance
The selected hydromechanical transmission was considered to have the follow-
ing components:
Spur gears
Planetary gears
' Hydraulic pumps
Hydraulic motor
In transmitting power either under steady or transient conditions, each of
these components is a source of power loss. The losses considered may be
grouped into three types:
a. Mechanical losses
Mechanical losses always act to oppose rotation of a shaft, and
arise from such sources as friction in the bearings, friction at
gear teeth, and windage. All of the components listed above are
subject to mechanical losses. Part of the mechanical loss acts as
a function of its transmitted power (load dependent) and part is a
function of speed and independent of transmitted power (speed depend-
ent loss).
b. Flow losses
Flow losses are a loss in pressure head in a pump-motor combina-
tion (due*to entrance), exit fluid inertial losses, and viscous
pipe losses. The direction in which flow losses act is determined
by the direction of power flow. Only hydraulic units are subject
to flow losses. Flow losses are predominantly speed-dependent.
C-l
-------
c. Compressibility and leakage losses
Compressibility and leakage losses represent deviations from
ideal performance in the transfer of flow from one hydraulic
unit to the other. Thus, while nominally the flow transferred
to the motor equals the flow generated by the pump, the actual
or effective flows differ by a small amount due to compressibility
of the fluid and leakage through seals and past the pistons. The
direction in which compressibility and leakage losses act is also
determined by the direction of power flow. Only the hydraulic
units are subject to compressibility and leakage losses.
The treatment of the various components including losses is as
follows:
Spur gears
^ A
»irTK
Wheel 1
Normal Direction of Power Flow
Wheel 2
Speed equation:
N,,
RN,
(C-l)
Mechanical loss equation:
T2 =
where
(1 - n)
sign
] /R
(C-2)
R is the gear ratio
T-, T. are the input and output torques acting in the
normal direction of rotation
NI , N_ are the input and output speed
C-2
-------
(1 - n)T-
Sign (N. )
is the efficiency with which power
is transmitted through the gear pair
represents the absolute value of the quantity
(1 -
represents the algebraic sign of the quantity
Nj; i.e., if NX is positive. Sign (Hj) » + 1;
if N, is negative, Sign (NI) « -1.
Planetary gears
VNR
Cage
1
o»"c
S S
C=T
/Tl
/ \
L .
~i V
£ Sun \_
Ring
Assumed Normal
Direction of
Power Flow
Speed equation:
r,, + r
R
Mechanical loss equations:
s> T T
S LS
a T T
R LR
T > T + T
1C C LC
Torque relationships:
rr 1 '
2 I T
21 rj TS
NR
r
R
rk
(C-3)
(C-4)
(C-5)
' C-3
-------
Mechanical loss definition:
TLS = I ^V TSI B±* (V
TLR = I (1-V TR! 8i& (V
TLC = I ^V Tc' ^
-------
Flow loss equation:
P ' V/DP -
"
where a is an exponent taken to 1.5. (1.5 Is a value often used where both
viscous (1st power) and inertial (2nd power) losses must be simply accounted
for) -..'.. - . . . -
Note that all flow loss for the pump -motor combination Is taken In the pump
(This was simply a computational convenience contributing negligible error
in the calculation) .
Motors
Mechanical loss equation:
M
where
V
LM
V - TLM
PDM
I V (K2M + K3M V
Normal Power
>. Fiow Direction
Transfer of flow from pump to motor
M
where
LC
Qp - QLC
[KIL
In calculating hydraulic transmission ..performance, the pump and motor
are considered in combination. Equations C-6 through C-10 relate the
six values of torque, speed and displacement for the pump and motor in
such a way that, given four of the six values, the remaining unknown
two may be determined.
C-5
-------
In the above treatment of hydraulic units, the following definitions
apply:
T ,TM are the input torque to pump and output torque
from the motor, respectively.
VDM
VNM
max
Kf
v ir
2M' 3M
are the pump and motor displacements.
is the ratio of pump displacement to its maximum
value for the unit.
are pump and motor speeds, rad/sec.
are ratios of pump and motor speeds to the maximum
values for the unit.
2
is pressure, Ib/in .
is maximum pressure value for the unit.
is a flow loss coefficient (see below for values
imposed for all loss coefficients).
are mechanical loss coefficients.
K.. ,K_ ,K are coefficients relating to compressibility and
leakage losses.
VQM
are the effective pump and motor flows, in /sec.
is a leakage flow.
d. Speed dependent losses
In addition to the losses in gears and hydraulic units defined by
equations C-l through C-10 the following speed dependent losses
are applied:
Charge pump ,windage,and control losses (M)
Valve plate friction losses (H)
Clutch and brake losses (M)
Main thrust bearing losses (H)
Journal bearing losses (M)
C-6
-------
These are broken into two sets of lumped losses» one associated with
the hydraulic power path (items H) and one associated with the mech-
vanical p^wer fiaeh (W)= these ate located ae showa in Figure? C-l and
C62.
Numerical values for efficiencies and loss coefficients
The loss coefficients and mechanical efficiency numbers for the con-.
sidered components were generated from previous experimental work.
Values for mechanical losses were typical for the type of gearing
specified in the transmission design. The numerical values of gear
efficiency used for performance calculations are shown in Figures C-l
and C-2 for the low-speed and high-speed ranges, respectively.
The aquations defining losses in all hydraulic elements (presented
earlier) were correlated to experimental data obtained for similar
hydraulic units operating in various different transmissions that
ranged in power rating from 25 to 200 horsepower. The resulta&ti
values of hydraulic element loss coefficients employed in all per-
foraamea asalyeis calculations were as fallows;
^ - 0.018 Kj^ i 0.0075
K <* 0.025 K,T - 0.025
<&W Ait
** * 0.005 K,_ - 0.005
jPi 3 Li
The speed dependent losses described abov$ are as specified in
Table C-l for two speeds.
TABLE C-l
SPEED DEPENDS!! LOSSES
Speed
$PM '
2100
3500
Mechanical
HP Loss
1.00
1.25
Hydraulic
HP Loss
0.5
1.0
Total
HP Loss
1.5
2.25
Between these speeds linear variation of losses with speed is assumed.
VC-7
-------
.
in
HP.
n
i
oo
VEHICLE
ACCESSORIES
DEPENDANT
LOSSES
MECHANICAL
SPEED-
DEPENDANT
LOSSES
WHEELS
HP
W
Fig. C-l Power-Splitting Transmission Low-Speed
Range Diagram
-------
NE
N
.
in
HP,
E
o
HP.
VEHICLE
ACCESSORIES
HYDRAULIC
SPEED-
DEPENDANT
LOSSES
(H)
HP
M2CHANICAL
SPEED-
DEPENDANT
LOSSES
= .99
il=.9925
CAGE
n*i..
N
o
HP
o
DIFFERENTIAL
RATIO
= .96
Fig. C-2 Power-Splitting Transmission High-Speed Range Diagram
-------
2. Description of Steady-State Performance Computer Program
This computer program was used to compute constant-speed performance of
the transmission and power train. Two types of load can be applied to
the power train: 1) cruise road load as specified by EPA (1); or
2) constant power, limited at low speed by slip of the wheels.
The program operates by starting with a required power at the wheels and
moving back towards the engine. At each point in the transmission in turn
the program calculates the power required to overcome all losses between
that point and the wheels and to supply the power demanded at the wheels.
In determining the required power output from the engine speed reducer,
the vehicle accessory load as specified by EPA (1) is added. With
the continuously variable transmission under consideration, the trans-
mission input speed (directly related to engine speed) for a given power
requirement, is determined by an engine operating line, specified by the
engine manufacturer. In general this line represents the condition of
minimum SFC (maximum engine efficiency). At low and high speeds the
limits of the transmission ratio range and engine speed range may force
deviation from the operating line. The mechanics of this deviation will
be discussed below.
An additional result of the analysis is the setting of the variable dis-
placement unit necessary to achieve the particular operating condition.
Thus, in addition to calculating performance, this computer program provides
a means of defining the .swashplate displacement schedules as a function of
speed.
Detailed program procedure
The following sequence of operations describes the actual procedures
followed by' the steady-state computer program. The terms upstream and
downstream describe relative locations which respectively follow or oppose
the natural flow of power. Operation of the transmission in low range
(Figure C-l) is described first (items 1-12) followed by modifications
(items 13-15) to handle the high-range operation (Figure C-2).
,010
-------
1. For each vehicle speed of interest the program initially calculates
the resistance torque to be overcome at the wheels, which must, there-
fore, be supplied to the wheels by the transmission. This torque is
based on EPA specifications.
2. That transmission input speed is calculated which will provide the
required power with the engine on the desired operating line. This
first calculation of speed is made assuming no intermediate losses
apart from wheel resistance.
3. Using equations C-l and C-2, the speed and torque acting immediately
upstream of the rear axle differential ace computed*
4. Using equations C-l and C-2, the speed and torque acting immediately
upstream of the output gear ratio (R ) are computed.
5. The mechanical power path speed-dependent losses are calculated as a
function of the current value of transmission input speed. For the
low-speed range thesa losses are the only input required to the mech-
anical power path.
6. Using equations C-6, C-7, C-8, C-9,and C-10 with pump and motor speed,
motor torque and motor displacement specified, the pump displacement
and pump torque are determined accounting for all losses in the hydrau-
lic pump-motor combination.
7. Hydraulic speed-dependent torque losses based on Table C-l are added to
the pump torque to determine the torque input to the hydraulic power
path.
8. The torque inputs to the mechanical and hydraulic power paths are
added and multiplied by speed to give the transmission input power.
9. The accessory HP, which is a function of engine speed based on EPA
specifications, is added to the transmission input power, giving the
power required as output from the engine speed reducer.
" C-ll
-------
10. From the engine performance map the engine speed to provide this power
while running on the desired operating line is determined.
11. The above procedure, starting with item 3, is repeated until engine speed
and displacement values are repeatable between successive iterations
within 1 part in 10,000.
12. Using the manufacturers engine performance map, the SFC, fuel flow and
MPG corresponding to the calculated power-speed condition of the engine
are determined.
For operation of the transmission in the high range, the following steps
replace steps 4 and 5.
13. With ring (transmission input) speed and cage (upstream of R ) speed
specified, equation C-3 is used to calculate the sun speed - which is
equal to the speed of element II.
14. Using equations C-4 and C-5, the sun and ring torques for the output planet-
ary are calculated. The sun torque is the torque of element II. The ring
torque (T ) is the torque downstream of the mechanical speed dependent
Ix
losses. The cage torque is equal to the torque upstream of the rear end
differential and known (from step 3).
15. The mechanical power path loss torque is calculated as a function of the
current value of transmission input speed and added to T to give the
torque input to the mechanical power path.
Apart from the above modifications to handle the output planetary, the treatment
for high-range operating parallels that for low-range operation.
The following additional constraints apply:
If any of the engine speed requirements call for a displacement
of unit I which lies outside t-he upper or lower limits, then the
value of displacement is set at the limiting value and the equa-
tions solved for engine speed. However, if the resultant engine
C-12
-------
speed falls below the minimum engaged speed, then this condition
will cause the transmission to partially disengage and is not
directly calculable by the procedures described. The latter con-
dition only occurs at speeds below 10 mph. To obtain efficiencies
at speeds below 10 mph, linear interpolation between zero at zero
mph and the value calculated for 10 mph is used.
If the engine under consideration has been scaled up in power by
some factor, F, it is assumed that for a given engine speed the
SFC for a given power demand, P, is that corresponding to the
power P/F at the same speed on the unsealed (original map).
Output on steady-state program
The following quantities are calculated and printed by the computer program
as a function of vehicle velocity:
Wheel Speed
*
Engine Speed
Transmission Output Speed
Wheel Torque
Road Torque
Engine Input Torque
Engine Output Torque
Transmission Output Torque
Engine HP
Road HP
Transmission .Input HP
Accessory HP
Overall Efficiency
Transmission Efficiency
Fuel Flow
Specific Fuel Consumption
MPG
Speed, Torque, HP at Hydraulic Units
C-13
-------
Hydraulic Pressure
Displacements of Hydraulic Units
Ideal Hydraulic Flows
Input for steady-state program
Input to the computer program consists of:
Wheel Radius
Car Frontal Area
Wind Resistance Coefficient
Ambient Pressure Temperature
Vehicle Weight
Accessory HP Tables as a Function of Engine Speed
Engine Performance Tables
Gear Ratios
Sun, Gear, Cage Radii for Output Planetary
Mechanical Efficiencies
Compressibility, Leakage and Flow Coefficients
Limits of Displacements and Speed Pressure
3. Description of Driving Cycle Performance Computer Program
This computer program is used to calculate the time-varying and cumulative per-
formance of the vehicle power train over any driving cycle specified in terms of
velocity values at a sequence of discrete points in time. In this case the load
applied to the vehicle is calculated on the basis of cruise road load at the
appropriate velocity added to the power at the wheels necessary to accelerate*
the vehicle according to the driving cycle. The driving cycle used for all per-
formance calculations is that specified in the Federal Register dated, July 2,
1971.
The program processes each interval of the driving cycle in turn, starting off by
calculating the power required at the wheels, then moving back towards the engine.
The procedure followed is very similar to that previously defined for the steady-
state program except that the transmission is treated as a
* The required acceleration over each time increment of the driving cycle is ob-
tained by numerical differentiation (a=(V -V )/AT) where V , V , are
velocities at beginning and end of Ith increment.
C-14
-------
single component with an efficiency defined as a function of speed and power
by the steady-state program. Thus the detailed calculational procedure is
exactly as defined for the steady-state program except that steps 4, 5, 6,.
7, 8, 13, 14 and 15 are replaced by:
16- Using tables of efficiency data generated by the steady-state
program the transmission input torque is calculated and multi-
plied by speed to give the transmission input power.
Thus, as for the steady-state program, the engine is operated on the opera-
ting line specified by the engine manufacturer. The implicit assumption
when applying this approach on the driving cycle is that the control system
is perfect. Thus the results of this driving cycle analysis may be slightly
optimistic in relation to the performance of a real power train - control
system combination.
Certain special conditions peculiar to the driving cycle analysis are handled
as follows:
1. Deceleration
If the deceleration is so mild that the negative vehicle inertia
force remains less than the steady-state road load, the net out-
put power from the transmission must remain.positive. In this
case the treatment is the same as for any condition in which the
power flow is positive.
If the deceleration is sufficient for the negative vehicle inertia
force to exceed the steady-state road load, then it is assumed that
positive braking is required; that no power output is required from
the transmission; that the transmission disengages; and that the
engine speed falls to idle. The only power output from the engine
under idle condition is the accessory load (2.00 hp without air
conditioning, 4.00 hp with air conditioning).
C-15
-------
2. Low-speed operation of the vehicle
At speeds below approximately 10 mph, the transmission is par-
tially disengaged. It is assumed that transmission efficiency
can be interpolated linearly between zero at zero vehicle speed
and the value obtained at the minimum engaged speed (subject to
an arbitrary minimum of 2 1/2 percent efficiency). This treat-
ment is regarded as conservative.
3. Transmission out of range
If the operating line requirement calls for a transmission ratio
above the maximum value (-5.5:1) then the engine speed is calcu-
lated to satisfy this ratio. However, if the resultant engine speed
falls below the minimum engaged speed (implying a vehicle speed
below 10 mph), then the engine speed is held at the minimum en-
gaged value and the efficiency is interpolated as described
above.
t
Having iterated to establish the power and speed from the engine
speed reducer, the fuel flow is calculated by interpolating from
the manufacturer's engine map as for the steady-state program.
The above description applies to each individual time interval.
Additional calculations involve the calculation of cumulative
values of work done at various points in the transmission, dis-
tance travelled, and fuel used.
Output
.The following quantities are printed by the program as a function of time:
Vehicle velocity
Wheel speed
Transmission output speed
Transmission input speed
Acceleration power
Steady-state power
C-16
-------
Road power
Transmission efficiency
Overall efficiency
Fuel flow
SFC
MPG
Together with cumulative values of
Time of trip
Distance travelled
Road work done
Engine work done
Total fuel consumed
And average values of
Velocity
if'
Road power
Engine power
SFC
MPG
Transmission efficiency
Input
Inputs to the Driving Cycle Program are:
Weight of car
Ambient pressure
Ambient temperature
A x C,
d
Minimum engaged speed
Idle speed
Wheel radius
Differential efficiency
Differential gear ratio
,,C-17
-------
Accessory power
Fuel density
Engine operating line (speed vs. hp)
Engine map (fuel flow vs. speed, power)
Transmission efficiency map (efficiency vs. output, speed, power)
4. Description of Full-Power Acceleration Performance Computer Program
This computer program calculates the histories of acceleration, velocity and
distance travelled for a vehicle accelerating between two specified velocities.
This program is used to calculate the performance of the vehicle relative to
the EPA maneuver specifications.
The program solves the combined equations of motion for the engine and vehicle,
subject to the assumption of perfect controls. The maximum engine power as a
function of speed is obtained from the engine map, and the engine is operated
for as much of the maneuver time as possible at the speed providing maximum
power.
The combined equation of motion for the engine-inertia system is as follows:
(more detailed explanation)(follows nomenclature).
T = T + J N +RRJN+RRT + RR Tn +RT,, (C-ll)
e a ee ooo o res o lo 11
N = R RN (C-12)
o o e
N = R RN + R RN (C-13)
o o e o e
where:
T is the output torque from the engine speed reducer.
6
T is the accessory torque.
cl *
J is the engine inertia as reflected at the transmission input shaft.
N is the engine speed (rad/sec).
R is the transmission ratio.
C-18
-------
R is the differential ratio.
o
.T is the vehicle inertia as reflected at the rear wheels.
N is the wheel speed.
T is the wheel torque required to overcome rolling and air
1T6S
resistance (as specified by EPA).
T.. 0 is the torque loss in the transmission as reflected to the
transmission output shaft.
T.... is the torque loss in the differential as reflected to the
rear axle.
*
N , N are wheel and engine accelerations, respectively.
Equation C-ll provides a means of determining the distance travelled, velocity
and acceleration history of the engine and vehicle provided that some relation-
ship defining the variable ratio R in equations C-ll, 12, 13 as a function of
time is available. Equation C-ll expresses the fact that the available engine
torque (which we know, as a function of engine speed) must, in general, provide
torque to accelerate engine and vehicle, and overcome internal losses and the
road load. According to the relationship for R (discussed below) certain terms
drop out.and the equation is rearranged to provide values either for acceleration
of the engine alone the vehicle above or some combined acceleration of the two.
So as to simplify the problem relationships for R are considered only for the
following 3 special cases
1. Engine Speed = Constant
Under this condition N =0 and equation C-ll becomes a one-
e «
inertia equation for vehicle speed. The identity N =0 is
substituted into equations C-12, 13; these are, in turn,
substituted
its subject,
substituted into C-ll, and C-ll is rearranged to make N
o
C-19
-------
2. Vehicle Speed = Constant
Under this condition N =0, and equation C-ll becomes a one-
inertia equation for engine speed. The identity N =0 is
substituted into equations C-12, 13; these are in turn
substituted into C-ll, and C-ll is rearranged to give N
6
as its subject.
3. Transmission Ratio = Constant
Under this condition R=0, and equation C-13 allows either
N or N to be eliminated from equation C-ll, so that a
e o
single effective inertia problem again exists. C-ll
provides a relal
as the subject.
provides a relationship with either N or N (arbitrary)
Adequate performance results for the EPA maneuvers can
be generated using combinations of these three
conditions. Figure C-3 shows the relationship between
engine and wheel speed which has been imposed to cover
all acceleration maneuvers. Five regions (I, II, III,
IV, V) are identified on this figure, and are defined
below.
C-20
-------
V
Engine
Speed
NDES
NL
N.
ENGAGE
N.
IDLE
IV
II
N.,, Wheel Speed
Fig. C-3 Relationship Between Engine Speed
and Wheel Speed for Acceleration
from Standing Start
Region I ,
The engine is accelerated from idle to some speed, N which is intended
engage
to correspond to the point in a real system where the transmission engagement
valve is closed. During this initial acceleration the vehicle is maintained
stationary. The delay involved before the vehicle starts to move is, in all
cases, less than 1/4 sec., which is considered an acceptable short time
(N is 200 rpm above idle speed for the Airesearch engine, 100 rpm above
cngagts
idle speed for the AeroJ et engine) .
Region II
During region II the engine and vehicle are considered to follow a linear
relationship until the ratio of engine speed to wheel speed reaches a value
corresponding to the maximum steady-state transmission ratio. The purpose
of region II 'is to bridge the gap between the condition when the engagement
valve is closed but there is 100 percent slip and the point where the slip
falls to a steady-state value (-10-12 percent slip for full power at 10 mph) .
C-21
-------
In region II equations C-12 and C-13 are replaced by the condition
N = AN + B
.e
N = AN
e o
A and B are constants during a particular acceleration and are determined by
the values imposed for N , N. ,. and the maximum steady-state trans-
GtigdgC 1QX6
mission ratio. Equation C-14 is substituted into equation C-ll to eliminate
one of N or N and the problem is solved as a single-inertia problem.
Region III
During region III the ratio R remains constant at the maximum value for the
transmission. Equation C-ll is solved as described above for a constant
ratio. Region III ends when the engine speed reaches the desired operating
speed for full-power acceleration.
Region IV
During region IV the engine speed remains constant at the desired operating
speed for full-power acceleration. Equation C-ll is solved as described
above for constant engine speed. Region IV normally continues to the comple-
tion of the maneuver.
Region V
Only in the case of the Aerojet engine, towards the end of the high-speed
passing maneuver (50-80 mph) , does the engine get forced, by the lowerlimit
of the transmission, to go above the desired operating speed (-2970 rpm) .
The system then operates in region V which is a constant-ratio region,
solved exactly as for region III.
The above idealization is intended as a simplified representation of the
more gradual, smooth changes in condition occurring in a real power-train-
control system. For example, the actual engagement process will start below
C-22
-------
the speed Neneaee» causing a build-up in hydraulic pressure, torque output
from the transmission, and acceleration of the vehicle, thus eliminating
the first discontinuity in slope of Figure C-3 between region I and II.
Similar smoothing of the curve of Figure C-3 will occur over its entirety.
During the early stages of acceleration from standing start, if all available
engine power were transmitted to the rear wheels they would skid. This was
controlled in the analysis in a manner analagous to the operation of.the
transmission (in which relief valves limit pressure to the value correspond-
ing to wheel slip). Thus in the analysis torques corresponding both to
limiting hydraulic pressure and to wheel slip were calculated and the trans-
mitted torque was not allowed to exceed either value.
The numerical calculations were performed using the Runge-Kutta 4th order
integration method.
C-23
-------
D. PERFORMANCE ANALYSIS
This section presents details of the performance which was determined for
the selected hydroraechanical transmission and the resultant power-train
performance of two selected advanced automatic engines with the trans-
mission. Three categories of performance analysis are considered:
Steady-state analysis (cruise and constant output power)
Driving-cycle analysis
' Full-power acceleration (standing start and DOT maneuvers)
In each analysis the performance was calculated for the selected engines
(specified by AiResearch and Aerojet) with the selected transmission
powering or medium-sized family car, as specified by EPA/AAPS vehicle
design goals ( 1)* Comparisons, where possible, are made with the corres-
ponding performance of a similar power-train employing a conventional
automatic transmission.
1. Specified Engine Characteristics
The characteristics of the two advanced automotive engines used for this
study were based upon engines selected by EPA/AAPS in order to ascertain
the performance gain offered by an advanced transmission. The engine
characteristics were supplied by Aerojet and AiResearch as a result of
their work under seperate EPA/AAPS contracts. These characteristics
are based upon differing design constraints, not necessarily optimized for
the selected variable-ratio transmisson, and have not been demonstrated
experimentally.
The characteristics of the Aerojet engine are based upon their "prototype"
design for a turbo Rankine-cycle engine which employs an organie working
fluid. Aerojet is currently testing and developing the "pre-prototype"
version of this engine under an EPA/AAPS contract. Figure D-l shows the
*Denotes references listed in Section F.
D-l
-------
Fig. D-l Aerojet Engine Performance Map
(Ram Air Velocity = 0)
HII-14607
-------
performance map for the Aerojet Rankine-cycle engine. Engine power level
(defined as the power output from the engine speed reducer) is plotted
against overall engine efficiency and speed reducer output speed. Figure
D-2 shows the percentage increase in power and efficiency as a function of
vehicle speed due to ram air velocity
The AiResearch engine characteristics are for a single-shaft, gas turbine
as described in their report to EPA/AAPS ( 2 ). Appendix A defines, in
tabular form, the performance of the AiResearch single-shaft gas turbine.
However, to illustrate the nature of the power variation with speed and
to define the operating line, a partial map (without fuel flow data) is
plotted in Figure D-3 The engine, as analyzed, was sealed up in power
output by 15 percent. The scaling law used was that, for given speed
and SFC, the engine could put out 15 percent more power than specified
by the map. The purpose in scaling up the engine power is to allow
achievement of all EPA maneuver specifications. This is discussed sub-
sequently in more detail. It is to be noted that the AiResearch Engine
data already has 4 hp subtracted to account for accessory load. Since
a variable accessory power was imposed in the present study, this 4 hp
was re-included in the engine power, before scaling.
Additional engine and transmission interface data are given in Table D-l.
TABLE D-l
ENGINE AND TRANSMISSION INTERFACE DATA
AiResearch
SS Gas Turbine
3500 rpm
Manufacturer
Type
1007» Transmission input speed
Engine inertia (reflected at
transmission input)
Transmission input speed at idle
Max. power
Idle fuel flow
0.766 lb.ft.sec
2100 rpm
147 hp at 100% speed
(scaled up 15%)
3.23 Ib/hr without A/C
3.81 Ib/hr with A/C
Aerojet
Rankine Cycle
3500 rpm
0.306 lb.ft.sec2
2100 rpm
125 hp at 0 mph
133 hp at 60 mph
2.42 Ib/hr without A/i
3.55 Ib/hr with A/C
D-3
-------
c
V
o
14
0)
^
o
28
24
20
16
12
O
n
CD
(U
10
20
30
40 50 60
Vehicle Velocity ( mph)
70
80
90
100
Fig. D-2 Effect of Ram Air Velocity Aerojet Engine
MTI-U663
-------
9
Ul
a
e
ss
CO
S
60
70
80 90
PERCENT ENGINE SPEED
100
INLET
TEMPERATURE
°F
1900
1900
1800
1700
1700
1700
1700
1700
Fig. D-3 AiResearch Engine Performance Map
(Includes Constant 4-hp Accessories)
MTI-14662
-------
Superimposed on each of the aforementioned engine maps (Figures D-l and
D-3) is an operating line which uniquely defines desired engine speed as
a function of power level. The operating line was specified by the engine
manufacturer as the desired steady-state operating condition of the engine.
The operating line tends to coincide with the maximum engine efficiency
(minimum SFC) under steady-state conditions. As described elsewhere in
this report, in the computation of steady-state and driving-cycle perform-
ance, the continuously variable ratio transmission is adjusted (controlled)
to maintain engine speed (by varying engine load) on the selected operating
line, within the extreme limits of the transmission.
2. Power-Train Loading ""
Under steady-state cruise conditions the road load is a combination of
rolling resistance and air resistance which is specified as a function of
velocity by EPA/AAPS ( 1 ). The variation of this load with vehicle vel-
ocity is plotted in Figure D-4. In addition the engine is subjected to
Accessory loads, also based on EPA specifications as a function of engine
speed, for the cases of with and without air conditioner. The accessory
load variation with engine speed is shown in Figure D-5.
For constant-power loading, a range of power-level demands, each related
by some fraction to the full-power load (taken to be 100 hp at the road),
is applied in turn. These loads are limited at low vehicle velocity by
the constraint that the implied tractive force at the wheels should be
less than 2150 Ib. The resultant road power level variation with vehicle
velocity is shown in Figure D-6.
Under driving-^ycle and full-power acceleration conditions the same road
load as described for steady-state cruise conditions was applied. In
addition the power to accelerate a 4600 Ib car and appropriate engine
inertia were applied. The power available to accelerate the vehicle is, in
both cases, limited firstly by the maximum engine power under wide-open
throttle conditions and secondly, by the wheel slip condition (maximum
tractive force of wheels = 2150 Ib).
D-6
-------
80
Vehicle Wei
Level Road
C A = 12 Sq
;ht = 4600 L
Ft
>s
60
w
to
I
40
20
20
30 40 50
VEHICLE VELOCITY, MPH
60
70
80
Fig. D-4 Road Horsepower Steady-State Cruise Conditions
-------
00
I
CO
I
2000
2200
2400 2600 2800 3000
ENGINE SPEED, RPM
3200
3400
Fig. D-5 Accessory Load - EPA/AAPS Specification
3600
MTI-14658
-------
100
100 Percent
80
2
I
60
40
50 Percent
20
25 Percent
10 Percent
10
20
30 40 50
VEHICLE VELOCITY - MPH
60
70
80
Fig,, D-6 Constant Power Low Speed Limits
-------
3. Transmission Performance Character! ti
The efficiency of the selected split-power path hydromechanical transmission
is highly dependent upon the mode of operation. At high speeds, where
the majority of power passes through the mechanical power path, the
efficiency is high. The shift point between low- and high-speed ranges
occurs at between 20 and 30 mph. Peak efficiency tendsto occur .close ..to
the "straight-through" point, where power flow through the hydraulic path
is essentially zero. The "straight-through" point occurs at between 35
and 50 mph.
The above cruise efficiency characteristics are shown very clearly in
Figure D-7, which presents the variation of transmission efficiency with
vehicle velocity under cruise conditions for the two engines. At 20 mph
the efficiency is between 57 and 60 percent, but rises to 90 percent at
50 mph. The slight discontinuity in slope at 20 mph with the AiResearch
engine, and at 25 mph with the Aerojet engine indicates the shift point.
The "straight-through" point occurs at 40 mph and 50 mph for the AiResearch
and Aerojet engines, respectively.
The marked difference between the transmission efficiencies with the two
engines at speeds below 50 mph reflects the differences in the manufacturer's
specified operating line. Up to approximately 24 hp output from the engine
the transmission input speed with the AiResearch engine is 2300 rpm (mini-
mum engaged speed). However, with the Aerojet engine the transmission
input speed lies between 2700 and 3000 rpm. This difference in speed means
that speed dependent transmission losses are more significant with the
Aerojet engine than with the AiResearch engine and, at the low power levels
demanded by the oruise condition, the effect of these losses on efficiency
was 13 percent at 30 mph. From a systems point of view this suggests the
possibility of improved overall performance resulting from operating the
Aerojet engine at a speed below optimum. However, this possibility has
not been explored under the present study.
D-10
-------
100
S
a
w
80
60
40
With AiResearch
Engine
With Aerojet Engine
20
Vehicle Wt.
Level Road
Rear End n
Rear End Ratio
4600 Ibs
0.96
3.08
10
20
30 40 50
VECHICLE VELOCITY, MPH
60
70
80
Fig. D-7 Cruise-Power Efficiency of Selected Transmission
MTl-li629
-------
The maximum transmission efficiency under cruise conditions is close to 90
percent. It occurs near to the "straight through" point (40 mph and 50 mph
for AiResearch and Aerojet respectively) and remains almost constant at speeds
above this.
The variation in efficiency at constant power levels as shown by Figures D-8
and D-9 reveals some differences from the cruise efficiency curves. This is
particularly noticeable at low speeds. For example, at 10 mph, the transmission
cruise efficiency is 51 percent; whereas, with an output at the road of 10 per-
cent of rated power, the efficiency is 78 percent. This difference results
from the fact that the cruise power demand at 10 mph is only 2 hp at the road,
exaggerating the importance of speed dependent losses, which are of similar
magnitude to the road load. A further observation from the constant-power
efficiency curves is the reduced separation between the two engines at high
power levels, indicating the reduced significance of speed-dependent losses
at high power levels.
The peak transmission efficiency is close to 95 percent, under conditions of
high power at the "straight through" speed condition. This high efficiency
value is a result of the minimal losses which occur when all power passes
through the mechanical path. The 5 percent loss which is incurred represents
the sum of all losses in the output planetary, the residual power necessary to
turn the variable hydraulic element even when no power is passing through it,
and the "parasitic" losses such as the charge pump.
Figures D-10 and D-ll are cross-plots of transmission efficiency vs output
power at 3 constant speeds, which reinforce the significance of speed-dependent
losses at low power levels. At 85 mph the speed-dependent losses are highest,
and at 10 hp output, the 85-mph efficiency is clearly the lowest. The fact
that the 50-mph efficiency is consistently highest over the output power
range reflects the closeness of this speed to the "straight through" point.
Figure D-12 presents a comparison of the selected hydromechanical transmission
cruise power efficiency with that of a conventional automatic. The conventional
automatic used for this comparison was that currently selected by Aerojet for
use with their Turbo-Rankine engine. The shift points are therefore designed
to produce optimum system performance for this engine. The wide spread of the
'D-12
-------
100
s
1
04
I
>4
Pw
W
Percent
Rated Power
20
10
20
30 40 50
VECHICLE VELOCITY, MPH
Fig0 D-8 Constant Power Efficiency of Selected Transmission
AiResearch Engine
MTI-14634
-------
100
o
w
w
a,
En
Percent
Rated Power
40
20
10
20
30 40 50
VEHICLE VELOCITY, MPH
80
Fig. D-9 Constant Power Efficiency of Selected Transmission
Aerojet Engine
MTl-14594
-------
a
CJ
w
20
40 60
OUTPUT POWER HP
50 MPH
85 MPH
20 MPH
10CT
Fig. D-10 Constant-Speed Efficiency of Selected Transmission
Aerojet Engine
D-15
MTl-14665
-------
100
g
W
w
Oi
50 MPH
85 MPH
2U MPH
20 40 60 80 100
OUTPUT POWER HP
Fig. D-H Constant-Speed Efficiency of Selected Transmission
AiResearch Engine
D-16
-------
w
Ou
a
o
P-H
w
100
80
60
With
AiResearch Engine
Std. Automatic Trans-
mission (Rear En^ Ratio
' = 2.93 *
40
20
H B- Shift
U Automati
by Aero;
Vehicle Wt.
Level Road
Rear End n
Rear End Ratio
= 0.96
= 3.08
s for Std
: Selected
point
10
20
30 40 50
VEHICLE VELOCITY MPH
60
70
80
Fig, D-12 Comparison of Transmission Efficiency Cruise Power
Different Rear End Ratio for Standard Automatics selected by Aerojet to give optimum performance
with remainder of power train (engine + transmission)
XTI-14632
-------
shift points, even under the cruise conditions of Fig. 5, is necessitated when
a vehicle speed range from zero to 85 mph is to be provided by an engine whose
ratio of maximum to idle speed is well below 2. In fact the complete" .'trans-
mission incorporated by Aerojet includes a separate idle gear for use between
10 and 22 mph and a slipping clutch for speeds below 10 mph. Thus, in the
range 10-85 mph, the so-called "conventional automatic" actually behaves as
a 4-speed rather than a 3-speed. '
Below 30 mph the conventional automatic is significantly more efficient than
the selected transmission - the difference reaching.24 efficiency.points..a_t.. _.
10 mph for the Aerojet engine. The implications of this difference in efficiency7
are discussed subsequently.
Additional measures of behavior, or performance which reflect most clearly the" ".
operation of the selected hydromechanical transmission are the fraction of trans-
mission input power which goes through the hydrostatic power path and the oper-
ating pressure in the hydraulic units.
Figures D-13 and D-14 show the variation of hydrostatic power with vehicle
velocity for the two engines, respectively. In each case, up to the shift
point, the majority of the power goes to the hydrostatic path-- 99 percent
under full power conditions and about 80 percent under cruise.conditions. The '_
small amount of power going to the mechanical power path under low-range con-
ditions is that associated with mechanical path speed-dependent losses such as
clutch friction.
At speeds above the shift point the fraction of power passing through the hydro-
static path falls rapidly to a minimum at the "straight-through" point. At speeds
above the "straight-through" point the hydrostatic power increases again to about
35 percent at 85 mph.
The main differences in power split characteristics for the two engines can be
associated with the difference in transmission shift behaviors. For the Aerojet
engine the sharp reduction in hydraulic power corresponding to the shift point
occurs at similar speeds, both for cruise and for 100 percent operation. How-
ever, for the AiResearch engine, the sharp reduction at the shift point occurs'
at 20 mph under cruise conditions and at 30 mph under 100 percent power. This
difference is a result of the insensitivity to power level of the desired Aerojet
D-18
-------
I
0*
em
100
Vehicle Wt,
Level Road
Fuel Density
Rear End T)
Rear End Ratio
6.30 Ibs/gal
0.96
3,. 08
Full Power
20
Fig. D-13
30 40 50
VEHICLE VELOCITY, MPH
Power Through Hydrostatic Path of Selected Transmission
with Aerojet Engine (with Air Conditioner)
MTI-14664
-------
w
s
H
ss
CO
§
v g
N>
o
2
100
Vehicle Wt.
Level Road
Fuel Density
Rear End 7)
Rear End Ratio
= 6.30 Ibs/gal
= 0.96
= 3.08
Full Power
20
30 40 50
VEHICLE VELOCITY, MPH
80
Fig. D-14 Power Through Hydrostatic Path of Selected Transmission
with AiResearch Engine (with Air Conditioner)
MTI-1&6C4
-------
engine speed, and the corresponding sensitivity to power level of the
AiResearch engine speed. Power level influences the shift point much less
for the Aerojet engine than for the AiResearch engine.
Table D-2 provides a more detailed breakdown of the power flow and indicates
the contribution of mechanical and hydraulic losses to performance,, This has
been done for the two vehicle speeds of 20 mph and.60 mph with the Aerojet
engine. The first of these speeds is slightly below the shift point, as the
second speed is somewhat above the "straight-through" point. In both cases
the amount of power flowing through the hydraulic path is very similar. How-
ever, at 20 mph the only power flowing to the mechanical path is that necessary
to overcome friction and to drive the charge pump - no output power is delivered
by the mechanical path. The influence of this difference in power split is
reflected in the percentage contribution of the hydraulic and mechanical losses.
At 20 mph hydraulic losses account for 29.5 percent of the transmission input
power, and mechanical losses account for 13.5 percent. At 60 mph the hydraulic
losses fall to 5.2 percent and the mechanical losses to 3.9 percent.
Consider the operating pressure in the hydraulic elements. It is important to
point out that maximum operating pressures above 3500 psi not only reduce the
life (and reliability) of the hydraulic elements, but also cause unwanted
excessive noise in the elements. Thus, low operating pressures are highly
desirable to reduce transmission noise to a minimum.
Figure D-15 shows the variation in hydrostatic pressure as a function of vehicle
speed for the selected hydromechanical transmission,, It can be seen that, for
cruise conditions, the operating pressure was below 400 psi. Note that the cruise
pressure drops slightly following the shift point, then remains nearly constant
up to the straight-through" point, and steadily increases in pressure up to a
maximum of 330 psi at 85 mph. Under full-power demands the pressure remains close
to its limiting value of 3500 psi at speeds of 10 and 15 mph, reflecting the design
limit corresponding to wheel slip. At higher speeds, even in low range (17-30 mph)
the full power can be transmitted through the hydrostatic path without exceeding
the pressure limits, and the pressure begins to fall with speed. At a speed of
85 mph the full-power pressure has decreased to about 530 psi.
D-21
-------
TABLE D-2
TYPICAL POWER FLOW BREAKDOWN - AEROJET ENGINE
WITH AIR CONDITIONER - CRUISE POWER
ENGINE HP
ACCESSORY HP
TRANSMISSION
INPUT HP
MECHANICAL
PATH LOSSES
HYDRAULIC
PATH LOSSES
TRANSMISSION
OUTPUT HP
DIFFERENTIAL
LOSSES
ROAD HP
20 MPH
AVAILABLE
HP
13.114
8.374
4.776
4.592
HP
USED &
LOSSES
4.740
1.126
2.472
0.184
% OF
TRANSMISSION
INPUT HP
56.60
13.45
29.52
2.20
60 MPH
AVAILABLE
HP
39.908
34.996
31.826
30.602
HP
USED &
LOSSES
49.11
1.356
1.814
1.224
% OF
TRANSMISSION
INPUT HP
14.03
3.87
5.18
3.50
D-22
-------
4000
o
NJ
OJ
3000
CO
CO
CO
CM
a
i
2000
1000
Vehicle
Level Road
Rear End
Rear End Ratio
VEHICLE VELOCITY, MPH
Figo D-15 Hydrostatic Pressures for Selected Transmission
HT1-U635
-------
The only significant difference in pressure level between the two engines
occurs at the shift point as a result of the differences in the shift
velocity for the two engines. Apart from this point, at which the trans-
mission is actually in a different mode for the two engines, the pressure
is dictated almost directly by the power demand at the wheels.
4. Power-Train Performance ~ Aerojet Engine
The fuel economy of the vehicle power-train, in MPG, is the most meaning-
ful measure of the efficiency with which fuel is being converted into
useful work. The cruise fuel economy for the Aerojet engine is shown in
Figure D-16. The two different lines reflect the influence of the air
conditioner on fuel economy, which can reach almost 3 mpg at low vehicle
speeds, but falls to less than 0.5 mpg at high speeds.
The optimum fuel economy occurs at 35-40 mph. Here the Aerojet engine
gives 15.4 mpg with air conditioner, and 17.4 mpg without air conditioner.
The drop-off in fuel economy at low and high speeds is pronounced. At
low speeds the main reason for the fall in fuel economy is the increase
in significance of transmission losses and vehicle accessories. At
high speeds the reason for the fall in fuel economy is the square-law
dependence of air resistance and rolling resistance with vehicle velocity.
Aerojet has designated an automatic transmission for use with their engine.
This transmission includes a 3-speed automatic gear box, a torque converter,
and an idler gear which provides additional speed reduction between 10 mph
and 22 mph. Using data for the efficiency of the 3-speed automatic gear
box and torque converter as supplied by EPA/AAPS ( 3 ) and assuming 97
percent efficiency for the idler gear, the fuel economy of the Aerojet
power-train employing this automatic transmission has been calculated and
compared with that for the selected transmission.
The comparison is shown in Figure D-17, and demonstrates the discontinuities
in performance associated with each shift point of the automatic. However,
D-24
-------
0
O
8
w
18
16
14
12
10
8
Vehicle Wt.
Level Road
Fuel Density
Rear End 7)
Rear End Ratio
- 4600 Ibs
= 6.30 Ibs/gal
= 0.96
= 3.08
Without A/C
With A/C
10
20
30
40
50
60
70
80
VEHICLE VELOCITY, MPH
Fig. D-16 Cruise Fuel Economy with Selected Transmission
Aerojet Engine
MT1-U626
-------
o
w
-
g
18
16
14
12
10
8
6
4
2
0
t
/
1- *
j-<
1
1
^
"
>- <
x^
Vehicle Wt. = 4600 Ibs
Level Road
Fuel Density = 6.30 Ibs/gal
Rear End T) = 0.96
Rear End Ratio = 3.08
Cruise Power
Tl
i**i
^
"
^-^
^r;
s>-
><
STD
Automatic
^^
M
<
\
Selected
Transmission
j^s
N
>^
^^
t
I Shift
Points For
STD Auto-
matic
missic
ted b\
TransJ
n Sele
Aero
\
c-
et
V.
^v
\
0 10 20 30 40 50 60 70 80
VEHICLE VELOCITY, MPH
Fig. D-17 Comparison of Fuel Economy Aerojet Engine with Air Conditioner
Cruise Power
MTI-U630
-------
by comparing average levels, it is apparent that, at low speeds, the fact
economy is slightly better with the automatic and, at high speeds, the
fuel economy is slightly better with the selected transmission. The pre-
dominant reason for this similarity in power-train performance is the flat,
symmetrical nature of the Aerojet engine performance map. Thus, typically,
a 500 rpm deviation from the optimum engine speed, either up or dom, causes
only about a 1 out of 17 deviation in engine efficiency. At low vehicle
speeds the higher efficiency of the automatic transmission actually results
in higher fuel economy when the engine is operating near its minimum SFC
point.
As an extreme illustration the power-train performance with a 100 percent
efficient hydromechanical transmission was calculated. The comparison of
power-train performance with the automatic transmission and the idealized
hydromechanical transmission is presented in Figure D-18.
The conclusion from this comparison is that, even with a perfectly efficient
transmission and the engine always operating at maximum efficiency, 17.9
mpg is the best cruise fuel economy which can be achieved - that is only
3.0 mpg or 20 percent better than with the standard automatic.
The ability of the selected hydromechanical transmission to keep the engine
operating at minimum SFC is demonstrated in Figure D-19. The solid lines
on this plot represent the engine SFC with and without air conditioner.
The engine SFC with the conventional automatic is superimposed on this
plot and only below 15 nph does the SFC with conventional automatic fall
below that with the selected transmission. The exception occurs because,
at 10 raph, the standard transmission actually allows the engine to operate
closer to the minimum SFC line then the selected transmission.
The Federal Driving cycle provides an alternative operating condition under
which to measure power-train performance. On the basis of the cruise per-
formance, it is not to be expected that the driving cycle will reveal any
significant performance advantages for the selected hydromechanical trans-
mission. Table D-3, which summarizes the Aerojet driving-cycle performance,
and compares it with available information for the automatic transmission
confirms this expectation.
D-27 /
-------
18
16
14
1 12
1 10
8
w
j 8
g
6
4
2
0
(
H =
/
/
100% -
/
L/
Vehicle Wt.
Level Road
Fuel Density
Rear End T]
Rear End Ratio
A
x>'
^<
^
^
/
^
^~
^~~~
^^.
MM^^^B
Efficiency as
Calculated
= 4600 Ibs
= 6.30 Ibs/
= 0.96
= 3.08
gal
<^
'
"^
K^^
^^
.^
^>»,
^^
^
'^
^^
) 10 20 30 40 50 60 70 80
VEHICLE VELOCITY, MPH
Fig. D-18 Effect of Transmission Efficiency Aerojet Engine
with Air Conditioner
MTI-U605
-------
0.9
0.8
i
I
W
0.7
0.6
With A/C
Vehicle Wt.
Level Road
Fuel Density
Rear End T)
Rear End Ratio
= 4600 Ibs
= 6.30 Ibs/gal
= 0.96
= 3.08
SID Automatic
Values (Without A/C)
20 40 60
VEHICLE VELOCITY, MPH
80
100
Fig. D-19 Specific Fuel Consumption with Selected Transmission
Aerojet Engine
D-29
MTI-U603
-------
TABLE D-3
AEROJET ENGINE - DRIVING-CYCLE PERFORMANC
Quantity
Average MPG
Average Transmission
Average Engine Power
Average Road Powe
Average Velocity
Selected
Transmission
With A/C
Selected
Transmission
Without A/C
100% 1]
Transmission
With A/C
9.55
717.
16.76 hp
8.38 hp
19.6 mph
10.85
71%
14.43 hp
8.38 hp
19.6 mph
12.02
100%
13.16 hp
8.38 hp
19 . 6 mph
Automatic
Transmission
.. With A/C
9.9
Note* Data Provided by Aerojet. (Transmission efficiency and engine power data not
available)
On the basis of this data, the Aerojet engine provides marginally better
performance with the automatic transmission than with the selected trans-
mission. It is of benefit to rationalize this conclusion as follows.
The average driving cycle vehicle velocity is 19.6 mph. On the basis of
cruise performance at 20 mph the automatic transmission might be expected
to provide significantly better fuel economy (see Figure D-17). However,
it should be noted that the average road power over the driving cycle
(8.38 hp) is 83 percent higher than the cruise road power at 20 mph (4.59
hp). This higher average power level results in a higher average trans-
mission efficiency (71 percent vs 62 percent) for the selected transmission,
and is the reason why the selected transmission gives very similar driving-
cycle fuel economy in comparison with the conventional automatic transmission.
D-30
-------
The effect of a perfectly efficient hydromechanics! transmission is to
give an average fuel economy of 12.02 mpg with air conditioner - a gain
of 2.47 mpg (25 percent) relative to performance with the actual trans-
mission efficiencies.
Consider now the full-power acceleration performance of the Aerojet engine
with the selected hydromechanical transmission. As shown by Table D-4,.'
the power-train exceeds all of the EPA/AAPS maneuver specifications.
TABLE D-4
AEROJET MANEUVER PERFORMANCE
Maneuver
>
1. Distance travelled in 10 seconds
2. Time to reach 60 mph from
standing start
3. High speed merge (25-70 mph)
4. DOT passing maneuver (time and
distance to overtake 50 mph truck)
TIME
DISTANCE
EPA
Specifications
440 ft.
13.5 sec.
15.0 sec.
15.0 sec.
1400 ft.
Aerojet Engind
with Selected
Transmission
505 ft
11.7 sec.
13.5 sec.
12.2 sec.
1166 ft.
Thus, no special optimization of the transmission, differential or overall
-power-train is necessary to satisfy these full-power vehicle performance
requirements. They are achieved without a requirement to exceed the max-
imum temperature or 100 percent speed.
D-31
-------
Typical time domain plots of engine velocity (rpm), vehicle velocity
(ft/sec), and distance traveled are shown in Figures D-20, D-21, and
D-22, respectively. The engine speed plot, in particular, reflects the
imposed control law described in Section C - Description of Methods for
Determining Performance. During region I, in which the engine alone is
accelerated to 100 rpm above idle, the high power available and relatively
low engine inertia make this an almost instantaneous acceleration. During
region II, in which the vehicle is accelerated from a standing condition
to a speed relative to the engine corresponding to the maximum transmission
ratio, the increase in engine speed is greatly slowed, taking 1.1 seconds
to provide a 200 rpm increase in engine speed. The reason for this slowing
is that the vehicle inertia, as seen by the engine, is increasing and that
there is an energy increase associated with the increase in inertia itself.
During region III the engine and vehicle are accelerated together at a
constant transmission ratio. This again results in a very rapid (0.3 sec)
increase in engine velocity from 2300 to 2970 rpm. Finally, in region
IV the constant engine speed is seen.
It is to be noted that, in spite of the sharp discontinuities in slope shown
by the engine speed variation, the vehicle acceleration of the vehicle is
limited to a value corresponding to wheel slip.
The distance vs time curve again closely confirms the maneuver results of
Table D-4. The relatively flat power-speed curve of the Aerojet engine
allows very rapid engine acceleration as discussed above. As a result,
the vehicle responds early and the resultant distance covered, \ V dt, in
10 seconds was 505 ft or 15 percent further than the EPA/AAPS specification.
5.. Power-Train Performance AiResearch Engine
The resultant fuel economy of the AiResearch engine with the selected hydro-
mechanical transmission is shown by Figure D-23. Peak performance occurs
at 40 mph, at which speed the fuel economy is 26 mpg with air conditioning,
D-32
-------
ENGINE SPEED - FULL POWER ACCELERATION
Vehicle Wt.
Level Road
Fuel Density
Rear End T]
Rear End Ratio
= 6.30 Ibs/gal
= 0.96
= 3.08
Fig. D-20. Engine Speed Under Full-Power Acceleration With
Selected Transmission (0-60 MPH) - Aerojet Engine
D-33
MTI-11600
-------
VEHICLE SPEED - FULL POWER ACCELERATION
TOT
.J.-.J--I--
Vehicle Wt.
Level Road
Fuel Density
Rear End T)
Rear End Ratio
= 6.30 Ibs/gal
= 0.96
= 3.08
Fig. D-21. Vehicle Velocity Under Full-Power Acceleration With
Selected Transmission (0-60 MPH) Aerojet Engine
D-34
MTI-H599
-------
Vehicle Wt.
Level Road
Fuel Density
Rear End T)
Rear End Ratio
6.30 Ibs/gal
0.96
3.08
Fig. D-22 Distance Travelled Under Full-Power Acceleration with
Selected Transmission (0-60 MPH) Aerojet Engine
D-35
MTI-14666
-------
28
26
24
22
20
o 18
OJ
16
14
12
10
8
(
/
/t
//
/
/
/
/
/ /
/
/
' /
/
/
/
Vehicle Wt.
Level Road
Fuel Density =
Rear End 7] =
Rear End Ratio =
^^*
^^
^^
^
lu*^*.
4600 Ibs
\
\
6.30 Ibs/gal
0.96
3.08
1
,\
\
\
\
^x
\
With A/C
Without A/C
^s
S
N^s
V
N^
\s
) 10 20 30 40 50 60 70 80
VEHICLE VELOCITY, MPH
Fig. D-23 Cruise Fuel Economy with Selected Transmission
AiResearch Engine
MTI-14627
-------
and 28.3 mpg without. It is noticeable that air conditioning has less of an
effect than with the Aerojet engine. This difference is attributable to the
lower operating speed of the AiResearch engine, and the correspondingly lower
accessory power level.
The power-train performance of a vehicle incoprorating the same automatic trans-
mission as discussed for the Aerojet engine has been calculated. It is recognized
that this transmission has not been optimized in any way for the single-shaft}
gas turbine engine and, indeed that a conventional automatic could never provide
JV
satisfactory kinematic performance with this engine. However, the calculation
does provide an exaggerated demonstration of the benefits of the selected trans-
mission for single shaft gas turbine application. The comparison of the fuel
economy for the two transmissions with the AiResearch engine is shown in Figure
D-24. Clearly the selected transmission offers considerable advantages in this
application. The reason for the poor power-train performance with the conven-
tional automatic transmission is the extreme sensitivity of the single-shaft,
gas-turbine SFC to engine speed for a given power demand.
This is confirmed in Figure D-25 in which specific fuel consumption is plotted
as a function of vehicle velocity. In most cases the selected transmission
produces substantially lower specific fuel consumption than the standard auto-
matic.
The performance over the Federal driving cycle of the AiResearch engine train
with the selected transmission is presented in Table D-5.
TABLE D-5
DRIVING'CYCLE PERFORMANCE - AIRESEARCH ENGINE
Selected Selected
Transmission Transmission
Quantity With A/C Without A/C
Average MPG ' 14.53 15.76
Average Transmission n 74.4 73.5
Average Engine Power 15.94 hp 13.92 hp
Average Road Power 8.38 hp 8.38 hp
Average Velocity 19.6 mph 19.6 mph
. D-37
-------
U)
00
24
o
I!
s
a
Vehicle Wt.
Level Road
Fuel Density
Rear End T)
Rear End Ratio
Mil Transmission
6.30 Ibs/gal
0.96
3.08
STD Automatic
Shift
Points
12
30 40 50
VEHICLE VELOCITY, MPH
80
Fig. D-24 Comparison of Fuel Economy AiResearch Engine with Air Conditioner
Cruise Power
MTI-14631
-------
.7
Z
o
en
z
o
o
o
w
CU
1.4
1.2
i.o
0.8
Without
A/C
0.6
0.4
o'
/r
Standard Automatic
Data
Vehicle Wt.
Level Road
Fuel Density
Rear End J\
Rear End Ratio
= 4600 Ibs
= 6.30 Ibs/gal
= 0.96
= 3.08
20
30 40 50
VEHICLE VELOCITY, MPH
60
70
80
Fig. D-25. Specific Fuel Consumption With Selected Transmission AiResearch Engine
-------
For operation both with and without air conditioner, the average driving-
cycle fuel economy is close to 5 mpg less than the corresponding cruise fuel
consumption. The air conditioner changes the average fuel economy by 1.4 mpg
or 9 percent. No data are available for performance of the AiResearch engine
over the driving cycle with a standard automatic transmission.
Full-power performance predictions with the AiResearch engine are given in
Table D-6 below.
TABLE D-6
AIRESEARCH MANEUVER PERFORMANCE .
Maneuver
1. Distance traveled in 10 seconds
2. Time to read 60 mph from standing
start
3. High speed merge (25-70 mph)
4. DOT passing maneuver (time and
distance to overtake 50 mph truck)
Time
Distance
EPA
Specifications
440 ft.
13.5 sec.
15.0 sec.
15.0 sec.
1400 ft.
AiResearch
Engine
447 ft.
11.1 sec.
11.6 sec.
11.8 sec.
1139 ft.
As discussed in relation to the engine data, the AiResearch engine power level
has been scaled up in order to meet these maneuver specifications. The most
critical maneuver requirement was found to be the distance traveled in 10
seconds; it may be seen that all other requirements are very comfortably met
by the power-train. The reason for this distance problem is the nature of the
variation with speed of available power with the AiResearch engine. As shown
in Figure D-3 the available power falls off very sharply at engine speeds below
100 percent. -Thus, the velocities reached in the first few seconds of the
acceleration are low, as demonstrated in Figure D-26, which gives vehicle velocity
as a function of time, in a 0-60 mph acceleration. During the latter part of the
acceleration maneuver the higher maximum power of the scaled-up AiResearch engine
produces higher accelerations so enabling it to meet the 0-60 mph requirements.
D-40J
-------
It is noted that Rosbach (5) shows the same AiResearch engine, unsealed,
as being able to meet the distance requirement. However, close examination of
this reference reveals that an idle speed equal to 70 percent of design speed
was used. The present studies used a value of 60 percent, and showed that
acceleration of the engine up to 70 percent speed took almost 0.8 seconds,
and that the vehicle moved only 4 feet in this period (see Figures D-26 and D-27),
In terms of distance traveled in 10 seconds the increased idle speed used by
Reference (5) will result in approximately 65 additional feet. However, the
penalties for using this increased idle speed are substantial. As an example,
a 300 rpm increase in the minimum engaged speed was found to cause over 4 mpg
decrease (20 percent) in fuel economy at 30 mph.
Figure D-28, which plots engine speed vs time for a 0-60 mph acceleration, shows
that with the selected variable ratio transmission, it took 3.3 seconds for the
engine to reach 100 percent speed (3675 rpm at transmission input), due to the
low initial engine torque characteristic typical of single-shaft gas turbines.
D-41
-------
Vehicle Wt.
Level Road
Fuel Density
Rear End T]
Rear End Ratio
6.30 iWgal
0.96
3.08
D-26 Vehicle Speed Under Full-Power Acceleration with Selected
Transmission (0-60 MPH) - AiResearch Engine
D-42
MTI-U667
-------
igj D-27 Distance Travelled Under Full-Power Acceleration with
Selected Transmission (0-50 MPH) AiResearch Engine
D-43
MTI-1«668
-------
Vehicle Wt.
Level Road
Fuel Density
Rear End T|
Rear End Ratio
Figa D-28 Engine Speed Under Full-Power Acceleration with Selected
Transmission (0-60 MPH) AiResearch Engine
D-44
MTI-M669
-------
E. COST ANALYSIS
Included in this section are a description of the methods employed to
determine transmission costs and a discussion of the results. The cost
data presented were generated using the experience of automotive cost
consultants based upon cost information and practices of the Ford Motor
Company. Therefore, this procedure provided a sound approach to com-
paring the cost of the selected hydromechanical transmission with that
of the standard-multispeed torque converter (automatic) currently in
mass production.
The objectives of the cost analysis were to determine the original equip-
ment manufacturer cost (OEM) for production quantities of 100,000 and
1,000,000 units per year of the hydromechanical transmission and then
compare that cost to similar costs for a multispeed (automatic) trans-
mission.
The procedure used to determine costs was to estimate the detail manu-
facturing cost of all components of the selected hydromechanical trans-
mission on a variable-cost basis rather than an OEM basis. Then, as sub-
sequently discussed, the total cost of the transmission on a variable-
cost basis was converted to an estimated-cost range on an OEM basis.
A variable-cost estimating approach is commonly used in the automotive
industry when making decisions on implementation of a new design or replace-
ment system. The major value of this approach is that it eliminated certain
transfer costs which may be affected in various ways. The OEM or transfer
costs would include cost allocations for fixed burden, scrap, factory cost
a.dju|5ents, general and administrative costs, profit and capital invest-
investment would include costs for facilities, tooling and
e. Since many of these transfer costs would vary with
different ftiijitoiBOtive companies, the OEM data were not as basic as the
variable cost data and therefores were not considered as reliable when
comparing information from different sources.
E-l
-------
The items which are included in a variable-cost comparison are the
purchased cost of the part, direct labor required to get to the desired
condition, the indirect labor associated with the manufacturing process,
variable overhead items which specifically relate to the manufacturing
process and programmed overhead expenses such as specific testing re-
quired.
Aside from the need to generate cost data on a basis which was consistent
and meaningful in the automotive industry, there was the requirement that
proprietary information be protected. For this reason, the cost analysis
is presented in the form of ratios using the standard automatic trans-
mission as the datum.
In addition, transfer costs for facilities would not reflect the same in
the transmission cost ratios. For instance, the cost of the facilities
for the automatic and hydromechanical transmissions could be the same.
As an example, let:
1.0 = Variable costs of automatic transmission
1.44 = Variable costs of power-splitting transmission
0.1 = Facilities costs for either transmission.
1 44
Variable cost ratio without facilities include = ' ', = 1.44
1 54
OEM or cost ratio with facilities included = -' . = 1.40
J. J.
Therefore, when dealing with ratios the increased cost of the basic trans-
mission is not correctly identified if only the OEM cost ratios are presen-
ted. OEM cost ratios are estimated. However, these are given a range to
account for possible variances.
E-2
-------
1. Typical Details of Costing Procedure
The technique for obtaining the costs for the selected transmission is
outlined and a sample sheet is included as Table E-l. The cost of 175
components, some of which were assembles of more than one item (an
example would be the park gear lock assembly) was reviewed in order to
obtain a valid cost comparison of the transmission. Approximately 25
percent of these were components now used in the standard automatic
transmission. The few examples presented on the sample sheet were, in
most instances, selected to present the costs of the hydraulic components
unique to the power-splitting transmission.
An overall design layout of the selected transmission was made in suffi-
cient detail in order to establish costs. Components, such as the hydro-
static pumps and motors not normally found in an automobile transmission,
were detailed with sufficient dimensional and material information for an
accurate cost estimate. The approach is similar to that used by high-
volume car manufacturers and is discussed in the following paragraphs.
The initial column of Table E-l describes the part or function to be
costed.
Columns 2 and 3 are the part number and number of such parts called out
on the transmission parts list as given by the transmission design layouts.
Column 4 presents the method of manufacturing the part consistent with
current automotive practices for mass production of transmissions.
In column 5 the material costs were established, and they include all
costs to bring the part to the "as-purchased" condition. For example, a
die cast component would have rough weight established to develop material
cost. The material cost was the actual purchase price in the "as-purchased"
condition.
E-3
-------
TABLE E-l
COST SAMPLE SHEET
TOTAL PER ASSEMBLY REMARKS
COLUMN 1 234567 8 9
LABOR
ITEM MAKE MAT'L COST LABOR COST TOTAL COST
DESCRIPTION NO QTY BUY (DOLLARS) MIN. (VAR) (DOLLARS) (DOLLARS)
Transmission Ass'y.
Shaft Engine Input
Valve Plate
Cylinder Block
Piston - Motor-Pump
Trunnion - Swashplate
Snaahplate - Motor-Pump
Support - Swaahplate
Gear - Motor - Planetary
Carrier - Planetary "B"
Planetary Carrier - Sub Ass'y.
Ring Gear - Planetary "B"
Planet Gear - Planetary "B"
Governor - Engage
Control Body Asa'y - Including
Valves, Sleeves, Springs, and
Linkage
NOMENCLATURE
LA
3
7
8
9
10
11
13
14
21
20 & 21
48
50
1
1
1
2
18
1
2
1
1
1
1
1
3
1
1
P.F.
P.R. "
P.S./F
M
A
A
PR
PR
PR
PS/F
PR
PF
PR
PR
PR
A
PR
M
PF
P.F. /P.R.
M/A
Purcha t
Purcha t
Purcha t
Manufac
Assembl
.992
.990
2.440
2.592
1.510
4.800
.600
1.550
1.540
1.484
.360
1.611
5.620
!d as finished item
id In rough condltic
id on a seml-f Inlshe
Cured In house
48.05
8.50
7.60
18.00
4.50
12.00
7.02
14.20
6.80
4.42
9.40
5.30
10.25
37.00
i, such as
1 item
8.385
1.483
1.326
3.140
.785
2.094
1.225
2.478
1.187
.772
1.640
.925
1.789
6.480
casting, and forging:
8.385
2.475
2.316
5.580
3.377
3.604
4.800
2.125
4.028
2.727
.772
3.124
1.285
3.400
12.100
Incl. Sub. Ass'y. not Specified (Cost Sht 1)
Forging - AISI 8620 Steel (Cost Sht 1)
Meehanlte Casting (Shell Mold) (Cost Sht 1)
Meehanite Casting (Shell Mold) (Cost Sht 1)
Cold Extrusion (H.T.) (Cost Sht 1)
Nodular Iron (H.T.) (Pearl. Mall.) (Cost Sht 1)
Heavy Coined Stpg. for Blank Torr. (Coat Sht 1)
Cast Iron (Cost Sht 1)
Forging - AISI 8620 Steel (Cost Sht 1)
Malleable Iron (Cost Sht 4)
Sub-Assembly (Cost Sht 4)
Heat Treat Nodular Iron (Cost Sht 4)
Stl. Bar (Cost Sht 4)
(Cost Sht 2)
Typical of Automatic Transmission (Cost Sht 2)
Control
TI-H150
-------
The "in-house" manufacturing costs to finish a specific part are developed
on extension of variable-minute costs times labor-minute content. Variable-
minute costs include direct labor, indirect labor and non-variable burden.
In Column 6, the actual "in-house" number of labor minutes to complete the
manufacturing task were listed and in Column 7 the variable-minute costs
were listed.
Column 8 is the total cost in dollars for each item or task labeled in
Column 1, and is the sum of Columns 5 and 7.
Tabulation of costing sheets similar to that shown by Table E-l provided
the basis for the variable cost of the selected hydromechanical trans-
mission. It should be noted that all cost-saving design improvements
suggested by the automotive cost consultants were faetored into the
results.
2. Resultant Cost Ratios
As a result of the detailed costing procedure outlined above, cost ratios
were established as shown by Table E-2. All ratios presented are the cost
of the selected hydromechanical transmission divided by the cost of a con-
ventional multispeed torque converter (automatic transmission) for medium-
size family car as currently mass produced. The range in OEM cost ratios is
shown in Table E-2 account for the estimated variation associated with
these costs.
Detailed cost estimates were made for production levels of 100,000 and
1,000,000 units/year with tooling and facilities appropriate for each of
these production rates.
The results given by Table E-2 show that for 1,000,000 units/year the ratio
of the variable cost of the selected hydromechanical transmission to that
of a typical presently produced, automatic transmission was 1.44 - a 44 per-
cent increase in cost. On an OEM basis the increase in cost ranged between
30 to 40 percent.
E-5
-------
TABLE E-2
TRANSMISSION COST ANALYSIS - COST RATIOS
PRODUCTION LEVEL
1,000,000 UNITS PER YEAR
STANDARD AUTOMATIC
TRANSMISSION WITH
TORQUE CONVERTER*
POWER SPLITTING
HYDROMECHANICAL
TRANSMISSION
PRODUCTION LEVEL
100,000 UNITS PER YEAR
STANDARD AUTOMATIC
TRANSMISSION WITH
TORQUE CONVERTER*
POWER SPLITTING
HYDROMECHANICAL
TRANSMISSION
1. VARIABLE COST RATIO (TOTAL)
a. CONTROL VARIABLE COST
RATIO
b. LABOR CONTENT RATIO
c. MATERIAL CONTENT RATIO
2. OEM COST RATIO
1.00*
1.44
1.29
1.00
1.00
1.00
1.28
1.34
1.53
1.29
1.50
1.20
1.00
1.30-1.40
1.25-1.35
1.65
2.02
1.84
1.86
1.70-1.80
USED AS REFERENCE, PRODUCTION LEVEL
OF 1,000,000 UNITS PER YEAR.
-------
At lower production rates, 100,000 units/year, it can be seen from Table
E-2 that the variable cost of the automatic transmission would be 1.29
times the cost for 1,000,000 units/year. Thus it follows that the variable
cost of the selected hydromechanical in quantities of 100,000 units/year
was 1.86 times (1.44 x 1.29 = 1.86) the cost of the automatic transmission
produced at the rate of 1,000,000 units/year.
Also shown by Table E-2 is a breakdown in costs attributed to controls, labor,
and material. At production levels of 1,000,000 units/year, the control cost
increase was 28 percent, additional labor content 34 percent, and material
content cost increased 53 percent.
Several features of the hydromechanical transmission can be cited as
contributors to the cost increase:
1. The additional governor required to control engine speed.
2. The infinitely variable ratio of the transmission is
achieved by using the planetary gearing. Thus, the
gears of a power splitting transmission are always loaded,
even under 1:1 conditions. A present day automatic locks
up in high gear and therefore the gears are unloaded for a
good proportion of driving time. This difference in
operation requires that the gears of the selected trans-
mission be heavier in construction.
3. The design and manufacturing techniques of a power-splitting
hydromechanical transmission have not been developed to the
degree of those of the automatic transmission. For example,
highly developed, economical, stamping and brazing techniques
such as used in current torque converter manufacture could
not be considered at this time.
E-7
-------
When an article is manufactured in high production, it is common practice
to assign a cost to the item as so many dollars per pound. The power-
splitting transmission weights 146 pounds versus 140 pounds for the standard
automatic transmission. Therefore, it is reasonable to believe that when
the design and manufacturing skills developed by the automobile industry,
over a period of several years of manufacturing the existing transmission
are applied to the selected hydromechanical transmissions, the cost will
decrease. Assuming an equivalent cost per pound for both transmissions,
the variable cost percent increase would be 4 percent.
In summary, it was concluded that the selected hydromechanical transmission
would initially cost 44 percent more than the standard automatic trans-
mission and finally, after several years of production and refinement,
would approach 4 percent nore.
E-8
-------
F. REFERENCES
1. "Advanced Automotive Power Systems (AAPS) Prototype Vehicle Performance
Specification," Environmental Protection Agency, Division of Advanced
Automotive Power Systems Development, January 3, 1972.
2. "Automobile Gas Turbine Optimization Study," AiResearch Manufacturing
Company of Arizona, Report No. AT-6100-R6, December 23, 1971.
3. Efficiency Curves for 3-speed Gearbox and Torque-Converter. Provided
by EPA. Also reprinted in ref. 4.
4. "Flywheel Drive Systems Study," R. R. Gilbert, G. E. Heuer, E. H.
Jacobsen, E. B. Kuhns, L. J. Lawson and W. T. Wada, Ground Vehicles
Systems, Lockheed Missiles and Space Company, Inc., California, Report
No. LMSC-D246393, July 31, 1972.
5. E. J. Rosbach (editor) Final Report - Automobile Gas Turbine - Optimum
Cycle Selection. EPA Contract No. 68-01-0406. G. E. Space Division.
6. "TRACOR Toroidal Traction Drive" TRACOR promotional literature 1972.
7. Rohr, "Proprietary Information Package"
8. J. J. Edwards and C. C. Hill "Application of the Aerodynamic Torque
Converter to Closed Cycle Systems." SAE paper 719164
9. C. C. Hill "Aerodynamic Torque Converter For Gas Turbines" ASME paper
69-GT-108
10. C. C. Hill, R. A. Mercure, C. D. Cole, "Design and Test of the First
Aerodynamic Torque Converter for Gas Turbines. ASME paper 69-GT-107.
F-l
-------
APPENDIX A
PERFORMANCE OF THE AIRESEARCH SINGLE-SHAFT. GAS-TURBINE ENGINE
Table A-l presents the performance data for the single-shaft, gas-turbine
engine as received from the engine manufacturer. This data accounts for
losses in the engine speed reducer and for a 4 hp constant accessory load.
All torque and speed data refer to the engine shaft.
Al
-------
TABLE A-l
PERFORMANCE DATA FOR THE SINGLE-SHAFT GAS-TURBINE ENGINE
Engine Scaled for 4600 Lb Vehicle 105 F Day, Sea Level
IGV Setting
Water Injection
Torque
(Ft-Lb)
Fuel Flow
(Lb/Hr)
Engine
RPM
Turbine Inlet
Temp (R)
1.02500
1.02500
1.02500
1.02500
1.02500
1.02500
1.02500
1.02500
1.02500
1.02500
1.02500
Water Injection
1.02500
1.02500
1.02500
1.02500
1.02500
1.02500
1.02500
1.02500
1.02500
1.02500
1.02500
1.02500
1.02500
1.02500
1.02500
1.02500
1.02500
1.02500
1.02500
1.02500
1.02500
1.02500
8.476
80236
7.605
6.869
5.89
5.001
4.103
3.258
2.676
2.081
1.587
7.322
7.328
6.844
6.248
5.395
4.568
3.763
2.969
2.428
1.864
1.411
6.52048
6.33683
5.85136
50 28404
4.53156
3.84759
3.15644
2.50573
2 o 05890
1.60082
1.22117
65.17
57.03
48.53
40.83
33.47
27.21
21.50
16.14
13.38
10.40
7.96
61.6
54.47
46.59
39.28
32.30
26.19
20.77
16.17
12.92
10.03
7.68
50.13635
43.87011
37.33375
31.40953
25.75446
20.92541
16.54436
12.87627
10.29384
8.00597
6.12664
83064.70000
79109.20000
75153.70000
71198.20000
67242.70000
63287.20000
59331.70000
55376.20000
51420.70000
47465.20000
43509.70000
83064.70000
79109.20000
75153.70000
71198.20000
67242.70000
63287.20000
59331.70000
55376.20000
51420.70000
47465.20000
43509.70000
83064.70000
79109.20000
75153.70000
71198.20000
67242.70000
63287.20000
59331.70000
55376.20000
51420.70000
47465.20000
43509.70000
2359.70000
2359.70000
2359.70000
2359.70000
2359.70000
2359.70000
2359.70000
2359.70000
2359.70000
2359.70000
2359.70000
2259.70000
2259.70000
2259.70000
2259.70000
2259.70000
2259.70000
2259.70000
2259.70000
2259.70000
2259.70000
2259.70000
2359.70000
2359.70000
2359.70000
2359.70000
2359.70000
2359.70000
2359.70000
2359.70000
2359.70000
2359.70000
2359.70000
A2
-------
(TABLE A-l CONT'D)
N III 2V ENGINE SCALED FOR 4600 LB VEHICLE 105 F DAY, SEA LEVEL
IGV Setting
1,
1,
1,
1.
1,
1.
1.
1.
1.
1.
02500
02500
02500
02500
02500
02500
02500
02500
02500
02500
1.25000
1.00000
1.00000
1,
1.
1,
1.
1,
1.
1.
1.
00000
00000
00000
00000
00000
00000
00000
00000
1.00000
.95000
.95000
.95000
.95000
.95000
.95000
.95000
.95000
.95000
.95000
.95000
.90000
.90000
.90000
.90000
.90000
.90000
.90000
.90000
.90000
.90000
.90000
Torque
(Ft-Lb)
5.63231
5.63714
5.26518
4.80603
4.15052
3.51416
2.89535
2.28427
1.86830
1.43460
1.08504
4.50329
4.74686
4.54973
4.21023
3.66930
3.10374
2.58053
2.00324
1.62480
1.23159
.88790
4.15621
4.36602
4.20550
3.84409
3.33133
2.80127
2.30842
1.81989
1.40877
1.06671
.74329
3.69192
3.93939
3.84652
3.46176
3.00534
2.53520
2.05547
1.51023
1.20657
.89830
.59617
Fuel Flow
(Lb/Hr)
47.39872
41.90301
35.83623
30.22312
24.85373.
20.15443
15.97729
12.43825
9.93734
7.71882
5.90734
43.45813
39.16062
33.55177
28.38931
23.42959
19.01482
15.15423
11.78439
9.37452
7.27073
5.53629
40.04014
36.05083
31.04099
26.13879
21.47865
17.35890
13.86534
10.82391
8.51587
6.62678
4.99698
36.97856
33.37912
28.79323
24.14606
19.88584
16.06917
12.76527
10.03611
7.76594
6.05864
4.57479
Engine
RPM
83064.70000
79109.20000
75153.70000
71198.20000
67242.70000
63287.20000
59331.70000
55376.20000
51420.70000
47465.20000
43509.70000
83064.70000
79109.20000
75153.70000
71198.20000
67242.70000
63287.20000
59331.70000
55376.20000
51420.70000
47465.20000
43509.70000
83064.70000
79109.20000
75153.70000
71198.20000
67242.70000
63287.20000
59331.70000
55376.20000
51420.70000
47465.20000
43509.70000
83064.70000
79109.20000
75153.70000
71198.20000
67242.70000
63287.20000
59331.70000
55376.20000
51420.70000
47465.20000
43509.70000
Turbine Inlet
Temp (R)
2259.70000
2259.70000
2259.70000
2259.70000
2259.70000
2259.70000
2259.70000
2259.70000
2259.70000
2259.70000
2259.70000
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
A3
-------
(TABLE A-l CONT'D)
IGV Setting
.85000
.85000
.85000
.85000
.85000
.85000
.85000
.85000
.85000
.85000
.85000
.80000
.80000
.80000
.80000
.80000
.80000
.80000
.80000
.80000
.80000
.80000
.75000
.75000
.75000
.75000
.75000
.75000
.75000
.75000
.75000
.75000
.75000
.70000
.70000
.70000
.70000
.70000
,70000
.70000
.70000
.70000
.70000
.70000
Torque
(Ft-Lb)
3.21515
3.54201
3.44061
3.04786
2.63594
2.23652
1.77420
1.37177
1.00330
.72302
.45745
Fuel Flow
(lb/Hr)
33.88733
30.83968
26.47738
22.13567
18.22650
14.79780
11.68450
9.08968
7.08779
5.49914
4.16674
Engine
RPM
83064.70000
79109.20000
75153.70000
71198.20000
67242.70000
63287.20000
59331.70000
55376.20000
51420.70000
47465.20000
43509.70000
Turbine Inlet
Temp (R)
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
2.70715
3.09817
2.95335
2.60773
2.24579
1.90405
1.52317
1.09418
.78042
.53901
.31242
2.27605
2.56293
2.42355
2.15166
1.88119
1.54983
1.22536
.82010
.55712
.34899
.14381
1.70357
1.99172
1.85513
1.65484
1.46790
1.17852
.90137
.51862
.31933
.13693
.01083
30.80587
28.22065
24.09902
20.20578
16.56778
13.42425
10.69869
8.21204
6.39494
4.91990
3.73569
27.88643
25.47337
21.70740
18.16663
14.99190
12.05540
9.61813
7.29660
5.67738
4.36702
3.33508
24.80789
22.63508
19.30451
16.11233
13.31192
10.78126
8.46226
6.36979
4.96326
3.82038
2.90044
83064.70000
79109.20000
75153.70000
71198.20000
67242.70000
63287.20000
59331.70000
55376.20000
51420. 70000
47465.20000
43509.70000
83064.70000
79109.20000
75153.70000
71198.20000
67242.70000
63287.20000
59331.70000
55376.20000
51420.70000
47465.20000
43509.70000
83064.70000
79109.20000
75153.70000
71198.20000
67242.70000
63287.20000
59331.70000
55376.20000
51420.70000
47465.20000
43509.70000
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
A4
-------
(TABLE A-l CONT'D)
IGV Setting
.65000
.65000
.65000
.65000
.65000
.65000
.65000
.65000
.65000
.65000
.65000
.60000
.60000
.60000
.60000
.60000
.60000
.60000
.60000
.'60000
.60000
.60000
Torque
(Fl-lb)
1.12997
1,32501
1.28382
1.21992
1.02301
.80923
.55255
.23529
.08813
-.06539
-.18341
.33612
.70237
.71189
.67152
.51866
.39334
.16514
-.05371
-.16955
-.26051
-.33485
Fuel Flow
(Ib/hr)
21.82657
19.79218
16.89241
14.25839
11.66346
9.43478
7.31575
5.48314
4.26976
3.29390
2.49250
18.45764
17.08063
14.61040
12.22382
10.02985
8.05270
6.15717
4.60849
3.60177
2.78586
2.12841
Engine
RPM
83064.70000
79109.20000
75153.70000
71198.20000
67242.70000
63287.20000
59331.70000
55376.20000
51420.70000
47465.20000
43509.70000
83064.70000
79109.20000
75153.70000
71198.20000
67242.70000
63287.20000
59331.70000
55376.20000
51420.70000
47465.20000
43509.70000
Turbine Inlet
Temp (R)
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000 .
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
2159.70000
A5
-------
TECHNICAL REPORT DATA
(Please read Instructions on the reverse before completing)
1. RC°ORT NO.
APTD-1517
2.
------- |