MTI-71TR75
APTD-118)
                         FEASIBILITY ANALYSIS
                         OF THE TRANSMISSION
                                              FOR
                     A FLYWHEEL/HEAT ENGINE
                  HYBRID PROPULSION SYSTEM
                                          NOVEMBER 1971
                                            prepared for
                              ENVIRONMENTAL PROTECTION AGENCY
                                     OFFICE OF AIR PROGRAMS
                       DIVISION OF ADVANCED AUTOMOTIVE POWER SYSTEMS

                                     CONTRACT NO. 68-04-0033
                             MECHANICAL TECHNOLOGY INCORPORATED

-------
                                  FOREWORD

This report presents a summary of the work performed by Mechanical  Technology
Incorporated and Bendix Research Laboratories,  Bendix Corporation with  coopera-
tion from the Ford Motor Company.  The work was performed for  the Environmental
Protection Agency (EPA), Office of Air Programs, Division of Advanced Automotive
Power Systems under Contract 68-04-0033.

Key consultants to the project were Mr. George  DeLalio, with respect to overall
transmission design, and Mr. Edwin Charles in regard to costing  analysis,

Mr. R. C. Bowlin was responsible for overall project direction.   Other  major con-
tributors from MTI were P. Lewis, H. Jones, Dr. A.  Smalley and Mrs. Linda
Almstead.  Also participating were John Broderick and Paul Dean  of  MTI.

Mr. W. Datwyler was responsible for coordinating the Bendix work on controls.
Other major contributors from Bendix include R. Presley and N. Sikora.

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.  The contribution of Dr. Karl Hellman  of EPA is also  acknowledged.
                                    -iii-

-------
                                TABLE OF CONTENTS
                                                                            Page
      FOREWORD	  iii
  I.  SUMMARY	    I
 II.  INTRODUCTION 	    3
III.  DISCUSSION OF RESULTS 	    9
      A.  Selection of Candidate Transmission 	    9
      B.  Description of Transmission Design 	   10
      C-.  Controls and Operational Aspects	   15
      D.  Comparative Performance 	   20
      E.  Safety Aspects	:   34
      F.  Regenerative Braking Analysis 	   35
      G.  Cost and Physical Comparisons	   36
 IV.  CONCLUSIONS AND RECOMMENDATIONS 	   A3
  V.  APPLICATION TO OTHER PROPULSION SYSTEMS	   47
 VI.  SUPPORTING INFORMATION
      A.  SELECTION OF CANDIDATE TRANSMISSION 	  A-l
      B.  DETAILED DESCRIPTION OF TRANSMISSION 	  B-l
      C.  PERFORMANCE ANALYSIS 	  C-l
      D.  CONTROLS DESIGN AND ANALYSIS 	  D-l
      E.  COST ANALYSIS	  E-l
      F,  SAFETY ANALYSIS	  F-l
      Gc  REGENERATIVE BRAKING ANALYSIS 	  G-l
      H.  REFERENCES	  H-l
APPENDICES
  I.  PROCEDURE FOR OBTAINING THE LOSS COEFFICIENTS AND EFFICIENCIES USED
      FOR PERFORMANCE CALCULATIONS 	  1-1
 II.  STABILITY AND ANALOG COMPUTER SIMULATION ANALYSIS 	 II-l
                                     -iv-

-------
                                    3i::CTION  I
                                    SUMMARY

The overall purpose of this study was to quantitatively assess the practicality,
from both a performance and cost viewpoint, of a transmission that will meet the
requirements of a flywheel/heat engine hybrid propulsion system for a family car.

The scope of the study included:  consideration of various types of feasible trans-
missions; selection of a candidate  transmission; detailed performance and controls
analysis of the candidate transmission; and lastly, a detailed cost analysis of
the transmission.  The heat engine  considered in the study was a standard internal
combustion engine typically used in present-day, medium sized, family cars.

In theory, the flywheel/hybrid propulsion system is quite simple.  However, from a
practical viewpoint the system interrelationships of the control and transmission
are much more complex than those that exist in a present family car with an auto-
matic transmission.  Thus, even though the transmission designed in this study is
compact and highly efficient, it would be 75 percent heavier and have 65 percent
more parts than a comparable conventional automatic transmission.  Based upon cost
information of the Ford Motor Company, this resulted in the "original equipment
manufacturer" (O.EnM.) cost of the  transmission being at least 125 percent higher
than a present day automatic transmission for production quantities of 1,000,000
units per year.  On a "variable cost" basis (more meaningful to the automotive
industry in comparing designs), the increase in cost was 140 percent.

Consideration of propulsion system costs showed that the flywheel/hybrid propul-
sion system (130 HP conventional .1C engine, flywheel and transmission) can be ex-
pected to have an O.E.M.  cost at least 60 to 70 percent more than a comparable
propulsion system (177 HP engine and automatic transmission) currently used in
automobiles.

Although the main objective for considering a flywheel/hybrid propulsion system
is to reduce emissions,  the resultant effect on emission levels was not evaluated
due to lack of emission data for the engine.  Consequently,  study efforts were
directed at minimizing engine fuel consumption.
                                     -1-

-------
From a performance viewpoint, the efficiency of the transmission was higher than

a conventional automatic transmission at cruise power and slightly lower at maxi-

mum power conditions.  Due to flywheel losses, the fuel consumption of the fly-

wheel/hybrid propulsion system was higher than a conventional system at idle and

low-speed cruise power conditions often encountered in urban driving situations.


A simplified control approach was selected and shown to be generally feasible.

However, additional studies are required in order to investigate system inter^

actions and optimize the control system.  The selected control approach provides

perfect "Total Kinetic Energy" (TKK) control (flywheel plus vehicle kinetic energy

remains constant) for a given schedule of vehicle loads and approximates it tit

other Loads.


in summary, the most significant conclusions and recommendations of this study

are as follows:

     1.  The design concept using the power-splitting transmission can effec-
         tively couple a heat engine with a flywheel power storage system.

     2.  Significant deviations from the desired total energy concept will be
         encountered during rapid or emergency vehicle stops.  This is because
         the rear drive wheels are the only wheels which can regenerate energy
         to the flywheel and only a fraction of the required energy can be re-
         covered.  This deficiency in flywheel energy must be provided by the
         heat engine to return the system to normal operation.

     !l.  System complexity and developments which are required for safety as
         well as further control development, suggest that the practical  im-
         plementation of such a system would not accrue by .1975.

     4.  The cost of the flywheel/hybrid propulsion system and associated
         transmission is significantly higher than a comparable present-day
         propulsion system and automatic transmission,
     5.  Based upon the above conclusions, hardware development of the trans-
         mission for the flywheel/hybrid propulsion system is not recommended
         without further detailed studieso  These studies should include:  (a)
         a realistic appraisal.of emission reductions (when emission data be-
         comes available), and (b) a comparison on a systems trade-off basis
         against other propulsion systems which are currently being investiga-
         ted by the Environmental Protection Agency, Division of Advanced Auto-
         motive Power Systems.
     6.  Other implications of the study are that the power-splitting trans-
         mission without power storage would have applicability to several
         engine types including the conventional 1C engine, the gas turbine
         engine, the turbo-Rankine engine, and the diesel derivative engines.
                                     -2-

-------
                                   Section  II
                                 INTRODUCTION

The concept of a hybrid propulsion system  for automobiles has recently received
considerable attention because it offers the possibility of reducing air pollu-
tion.   J.n this concept, an engine and energy storage device are combined so
that acceleration power is supplied by the energy storage device.  Then ideally,
during  decelerations, the energy storage device is recharged by absorbing energy
from the vehicle.  Compared to present automobile propulsion systems, this
would result in a smaller engine which could be operated at optimum running
conditions to reduce emissions.

However, in order for a hybrid propulsion system to be a viable alternative to
present or future automotive propulsion systems it should:
     1.  Meet the Federal emission standards
     2.  Meet vehicle requirements as established by the Environmental
                              *
         Protection Agency (1)
     3.  Require minimum increase in cost compared to present engine/
         transmission systems
     A.  Be operational and ready for mass production by automotive model
         year 1975 or soon after.

In addition,  in our opinion,  it is desirable for the hybrid propulsion system
to require no significant changes in driving habits by the operator and to require
minimum changes to the vehicle.

In a previous study, Lockheed Missiles and Space Company (LMSC)  (2) investigated
the feasibility of flywheels  for various types of vehicles.   As  a result of
that study,  LMSC recommended  that a hybrid propulsion system for a family car
consisting of a flywheel and  a heat engine offered considerable  promise for
reducing emissions.  In addition, LMSC recommended that in order to minimize
* Numbers in parenthesis refer to references listed at the end of this report.
                                     -3-

-------
history  dependent  vehicle performance and  flywheel .size, the best control.
approach for the  flywheel was a Total Kinetic Energy (TKE) concept.  In that
concept the sum of the vehicle plus flywheel kinetic energies is maintained
constant by transmission controls.

Since the transmission which links the flywheel, heat engine, and vehicle drive
train together in an optimum manner is the critical link in a flywheel/hybrid
system, this study was initiated by the Environmental Protection Agency,
Division of Advanced Automotive Power Systems (EPA/AAPS).

As part of a parallel effort,  LMSC was the systems contractor for EPA/AAPS.
As the system contractor LMSC supplied information for this study pertaining
to 1) desired vehicle tractive efforts, 2) flywheel characteristics (including
costs), and 3)  preliminary characteristics of flywheel containment housings.

In order to quantitatively assess the practicality, from both a performance and
cost viewpoint, of a transmission that will meet the requirements of a family
car flywheel/heat  engine hybrid propulsion system, the detailed objectives of
this study were as follows:

     1.   Establish feasible  design concepts for  linking the power transfer
         between an 1C heat  engine, flywheel, and load  (vehicle)  by considering
         mechanical,  hydrostatic, and hydromechanical (power-splitting) types
         of transmissions,

     2.   Establish practical control  system conceptual  designs for the entire
         integrated system consisting of  the heat engine, flywheel,  selected
         transmission concept  and the vehicle including analysis  of stability,
         safety,  and  possible  operator induced instability.

     3   Determine the comparative performance efficiencies of the various trans-
         mission design concepts  as a basis for  screening and  selection of a
         final  concept and design approach.

     4.   Determine how cost  and complexity of the transmission design  concepts
         are affected by using all or a fraction of the regenerative power from
         the rear  wheels of  the vehicle.
                                     -4-

-------
      5.   Estimate,  from sufficiently  Detailed  preliminary  designs,  the  original
          equipment  manufacturer  (OEM)  cost:  of  the  transmission and  control  system
          design concepts  for  selected  production quantities.

      6.   Determine  the  performance efficiencies of the selected transmission
          design concept on both  a steady-state and dynamic  (efficiency  versus
          time) basis for  selected vehicle operating conditions with the transmis-
          sions as  an integral part of  the overall propulsion/vehicle system with
          controls.

      7.   Comparatively  evaluate  the cost and performance of the selected trans-
          mission and control system concept to a conventional multispeed torque
          converter  transmission.

      8.   Recommend  an optimum transmission based upon system cost and efficiency.
          The recommendation shall include all pertinent physical configuration
          and operation characteristics  (optimum flywheel speed ratio, heat
          engine operational mode, etc.).  If, in the opinion of MTI the optimum
          transmission/control system cannot actually be found that fulfills
          the requirements this conclusion should be made.

In developing an approach to accomplish the aforementioned objectives consistent
with  the necessary  and desirable goals for a hybrid propulsion system mentioned
earlier, emphasis was placed on selecting transmission concepts and components
which were based upon current technology.  In addition, EPA/AAPS directed that
the basic approach  to the transmission  (system) controls was to follow the TKE
control concept originally suggested by LMSC..

The methodology of  the study approach is shown diagrammatically by Figure  II-l.
Initially, several  transmission concepts were investigated.  These included
mechanical, hydrostatic, and power splitting.  Then,  by a screening process, a
"candidate" transmission concept was selected for more detailed investigation.
In a similar manner, two control concepts were investigated for stability and
other considerations.   As part of the control studies the propulsion system
was simulated on an analog computer.   Final assessments of the candidate trans-
mission and a preferred control system included system comparisons of cost,
                                     -5-

-------
                                PHASE I APPROACH
                                           System
                                       Characteristics
                Steady-State
                Performance
 Production
 Costing of
Transmission
 & Controls
                                    Screen & Select Best
                                    Transmission Concept
                                        Transmission
                                     Preliminary Design
                        Control
                       Parameters
Transmission
  Dynamics
   Control
   Design
                                     Dynamic Analysis of
                                     System on Analog &
                                      Digi tal Computers
                                         Performance
                                         Comparisons
                                  4Cost Comparisons  1

i
Recommendation
& Report
                            II-l    Phase  T Study Approach
                                      -6-

-------
package-size, weight - as well as steady-state and dynamic performance based
upon the results obtained from digital computer models of the complete propulsion
system.

The following sections of the report:  1) summarize the study results; 2) present
conclusions and recommendations; and 3)  discuss applications to other propulsion
systems.  Detailed supporting information is contained at the end of the report.
                                      -7-

-------
                                           TTT
                              DISCUSSION OF RESULTS

Tlu.s section summarizes the results of the study.  Additional supporting de-
tails -ire given by topical headings in Section VI - Supporting Information,

A_, __ Selection of Candidate Transmission
There are various transmission concepts that can transfer power from the
engine and flywheel to accommodate the required road loads.  In order to
determine which concept would be the best suited as a candidate transmission
for the flywheel/hybrid propulsion system, various applicable power transmit-
ting devices were separated into two categories:  hydraulic and mechanical.

The following devices were investigated in each category:
   Hydraul ic
      1.  Hydrostatic or positive displacement hydraulic pumps and motors
      2.  Hydrokinetic fluid drive or fluid coupling
      3 ,.  Hydrodynamic coupling or torque converter
      4.  Hydroviscous shear drive or slipping cluat'cb

   Mec_ha_ni£a_l_
      1.  Traction drives
      2.  Cam or gear drives
      3.  Belt drives

As discussed in more detail in Section VI-A,  the most promising of these
devices were used as the basis for seven different transmission design concepts
Schematics and sketch layouts were made for each concept, in sufficient detail
so that selected evaluation factors could be  estimated.

-------
 The  resulting  evaluation  of  transmission concepts  is  summarized by Table III-l.
 The  comparative  ratings shown  have  been normalized in each category such that.
 the  lowest  numbers  indicate  the  best, rating.   As  shown by Table III-l,  the Two
 Element  Hydrodynamic  concept, had the best overall  rating on a  numerical basis.
 [n I hot  concept,  l he  variable  ratio could be  achieved by using two fluid coup L-
 ing.s  or  i wo lorqut;  converters.   Since fluid coupling  slip significantly lowers
 transmission efficiency,  the two converter approach was  selected.   However,  wiili
 that  approach, a  valve was  required to schedule  flow  of  fluid  and  power for  the
 convenors,  Has fed  upon our available information,  this  concept  (including
 valving)  h;is not  been  reduced  to practice or  developed..   Truis ,  (his concept  jnd
 other concepts which  required  component development were excluded  from  consid-
 eration  en  the basis  of lacking  near-term availability.

 Consequently,  the power splitter with the flywheel axis  in line with the engine
 shaft (to reduce  complexity) was selected as  the most promising transmission
 concept   The  basis for the selection was because:  1)  it  did  not.  require
 component development;  2) it  was considered  the best, concept  with regard to
 the  factors of cost,  specific  fuel  consumption (SFC) ,  and  efficiency;   and 3)
 with all  evaluation factors considered,  it was only exceeded by one other con-
 cept and  that  conc.ept  required component development.
jj^ __ DescjMf j._i on of Transmits ion Design
The bijsic design approach  followed  in determining  the  best arrangement of  the
power splitting transmission was  i.o:
     I.  Minimize package  size and  weight,  to  reduce cost
     2.  Have the design as  compatible as  possible with present auto-
         mobiles so as t.c  minimize  vehicle body or suspension system
         changes
     3   To transmit power at high  efficiencies —  from the engine to the
         road and engine to  the flywheel in order  to minimize emissions under
         steady- s tate driving conditions

Based on size, power path  efficiency and vehicle compatibility, the design was
oriented to place the flywheel in the region currently occupied by the torque
converter.   The  required hydraulic units and associated planetary gearing were
integrated  by design to form a single, compact transmission package.   This was

                                     -10-

-------
TABLE III-L    NORMALIZED SCREENING FOR TRANSMISSION CONCEPTS

SK-J-4258
3 Element
Hydrodynamic
SK-E-4260
2 Element
Hydrodynamic
SK-J-4257
Hydrostatic
SK-E-4256
Power Split
SK-E-4261
Power Split
(Offset)
SK-D-4259
Hydrodynamic
Hydrostatic
SK-A-4262
Hydrodynamic
2 Coupling
Cost
1.266
1.0
1.4
1.266
1.333
1.266
1.133
Weight
1.146
1.0
1.265
1.142
1.173
1.064
1.059
SFC
1.3
1.4
1.0
1.0
1.0
1.3
1.4
Size
1.21
1.00
1.46
1.35
1.46
1.13
1.02
No.
Parts
1.44
1.23
1.00
1.27
1.44
1.41
1.30
Eff .*
1.39
1.07
1.26
1.00
1.00
1.13
1.38
Overall
Rating
7.757
6.700
7.385
7.028
7.406
7.300
7.292
Rating
Cost-SFC
Eff .
3.961
3.470
3.66
3.266
3.333
3.696
3.913
Notes
(a)
(a)
i

(b)
(b)

(a)
        * The normalized  efficiency  is obtained using the weighted efficiency  data.
      (a) Component development  is required.
      (b) Optimum availability of engine power.

-------
 considered  a  more  efficient  approach than,  for example, using two functionally
 separate  power  splitting- transmissions.

 Figure  1II-1  presents  a  schematic  of the recommended transmission design.  As
 discussed in  more  detail in  Section VI-B,  the primary transmission utilizes
 a simple  pump- motor  circuit  (Elements  I  and II) in combination with a planetary
 gear train  to provide  a  wide stepless  variable speed and torque ratio.  The
 r.-if LO  LS  changed by  varying  t.he  displacements of Elements I and II.  With this
 type of cony t. rue I- ion,  the  hydrostatic  units are not subjected to more than 33
 percent ol  t.he  power over  a  great,  percentage of t.he operating range.
The- p ropo<;<'d 'lesion  has  I wo  modes  of  ope rat', ion.   In the low-output-speed
the liydranlir circuit  functions  as  d  straight,  hydrostatic system driving
t.hroMgh  (!ie Planetary  "C"  gear set.   This  provides high output torque and vari-
able  operation  in  both  the forward  and  reverse directions.   In the high-ojt.put
speed stage, the Planetary "B" gear set and  hydrostatic circuit, act as a split
torque or hyd romechanical  system.   This extends  the range of output speed and
increases the efficiency with positive  control.   By utilizing a combination
hydros tatic-hydromechanical  construction,,  the  range and pressure level over
which the hydrostatic circuit must  operate,  is minimized.  This substantially
reduces  the displacement and size  of  the hydrostatic elements which increases
the efficiency  by  f ransmi 1 1. ing most  of  t.he power through the more efficient
mfchoii].ca I e I uinent •-;  and ;ji the same  time provides  a smaller and more compact!
de '-i i >',ri .

In the p ropob'.'d design, the  planetary gear train was constructed •so tluit  the
cl'il.ch and brake elements are in synchronization during transition from one
stage LO  the other.  This eliminates  high  inertial loads on the clutch and brake
which reduces wear and provides  positive and even  drive over the entire opera-
ting  range.

The secondary transmission provides  a power path from the prime mover through
the carrier of Planetary "A" to  the  flywheel or  from the flywheel to the plane-
tary carrier and on  to the rear  axle.   The path  to the  flywheel is direct, and
efficient.  The displacements of the hydrostatic pump-motor circuit (Elements
III and  IV) are varied to est.ablish whether  the  flywheel is accelerated or
decelerated with respect to  the  input.  The hydrostatic drive functions on
                                    -12-

-------
                                             Planetary  "A"
I
I—'
U)
      Engine
      Input
      1200 RPM
        to
      3750 RPM
                                                                         Primary Transmission
                                                                               I     I
                                                                                                                       Suashplnre
                              6.5
                               2.75
                            -Li.ol"


~ J
=


"•
Flywheel
24,000 RPM
10,000 RFM
.
                                            1.00
                                            1.00
                                                 1.27
                                                          Variable Displacement
                                                          + 3200 RPM
                                                            3200 PSI
                                                            7.5 In3/rev.
                                                              ELEMENT 1
                         Variable Displacement
                         + 3200 RPM
                           3,200 PSI
                           7.5 InVrev.
                                                                                          ELEMENT II
                                                  1. 00
                   I     I

                   L_J
                   I     I
Variable Displacement
+ 3400 RPM
  3400 PSI
  6.0 tn3/rev.
                                                                ELEMENT HI
                                                                       \
Variable Displacement
  3400 RPM
  3400 PSI
  6.0 InVrev.
                                                                                         ELEMENT IV
                                                                       Secondary Transmission
                                                                                                                                        Planetary "B"
                                                                         Planetary  "C"
                                                                             Tranamloa :.oi
                                                                             Output
                                                                             0 RPM
                                                                              to
                                                                             4470 RPM
                                                                                            f*-- Lov Range Brake

                                                                                      I/}/)/}/
                                                                          High Range Clutch


                                                 Notes:  1.  Rear Axle Differential Ratio = 3.55:1
                                                        2.  Rear Wheel Diameter = 2.1  ft.
                                                        3.  Vehicle Speed Equals 94.4  raph
                                                           at Transmission Output Speec"
                                                           of 4470 rpm.
                                              Fig.  III-l     Schematic  of  Recommended  Transmission  Design
                                                                                                                                                               MTI-12340

-------
both sides of  the mechanical  drive  of Planetary  "A"  only  as  a  means  to  add  or
subtract,  speed  to effect positive control.  This  reduces  the hydrostatic  por-
tion of the  flywheel drive  to a minimum.  The  flow of  power  from  the  flywheel
to  the output  is carried through the primary transmission.   Since  the primary
transmission is already sized to provide maximum  torque multiplication  to the
output, the  flywheel acceleration power does not  change the  size  of  the hy-
draulic or mechanical elements.

An  important feature of the transmission design was  the hydraulic  units.  These
type of hydraulic uni-ts are fully developed and have been successfully  used  for
aircraft,  constant: speed drives, tractor transmissions, truck transmissions,  and
in  postal vehicle transmissions.  The key features of  the unit, are that they
have hi;;h efficiency and high power dens i ties (2 horsepower/cubic  inch)  at
reasonable operating pressures in order to reduce wear and achieve reliability.
Additional details on the hydraulic units are  given  in Section VI-B Supporting
Information.

Although  the foregoing design utilizes an inline  flywheel to achieve  the
efficiency, compactness and compatibility discussed  previously., offsetting  the
flywheel  for systems reasons (such as gyroscopic effect) could  be  readily
accommodated.

O'her flywheel   locations were also explored.    For example., location ot"  t.he
flywhc:"!  fi.irt.her downstream (nearer to the rear wheels) was evaluated.  How-
ever., an  even greater ratio would be required  in  the planetary gearing  result-
ing in larger sizes and less efficient operation.

It should be noted that the design presented can  be  further enhanced by the
use of compound planetaries.  The basic benefit of this approach would be to
provide additional operational points where transmission  of power  through the
hydraulic elements would be even further reduced with a resulting  performance
increase.   Because this approach will require  additional  components with at-
tendant increase in complexity and cost, it was not  further evaluated for the
current program.
                                   -14-

-------
C.  Controls and Operational
The controls required for the flywheel/hybrid propulsion system, of necessity,
will be more complicated than those presently employed in an automobile with
an automatic transmission.  This results not only from the fact that there are
two sources of power to the rear wheels but also because the system is com-
pounded by other characteristics which include:
     1.  Initial charging of the flywheel with no power flow to rear wheels
     2.  Flywheel charging by engine while vehicle is moving or at a stop
     3.  Engine operation geared to produce minimum emissions
     4.  Design and functioning compatible with minimum need for changing
         the habits of motorists

Furthermore, as recommended by Lockheed Missiles and Space Company (2) in a
preliminary study of the flywheel/hybrid propulsion system, it is  desira'^e to
have the control system of the transmission follow a "Total Kinetic Energy"
(TKE) control law.   This tends to minimize flywheel weight and offers the possi-
bility of not having the propulsion system "history dependent".  The basic idea
of the TKE approach is to control the heat-engine output so as to  have the
kinetic energy of the vehicle plus flywheel remain constant.

Maintaining ideal TKE is complicated by factors such as:
     1.  Hills — which can introduce significant, unscheduled, system
         kinetic energy variations due t.o potential energy variations;
     2.  Dissipative Braking — which can introduce rapid and significant
         unscheduled changes in system TKE level which must be made up
         by the engine;
     3.  Variable Loading— the number of passengers carried, trunk loading,
         or pulling a trailing vehicle can change the vehicle mass from the
         value used in computing and implementing the TKE control scheme
     4.  Wheel Slip — which can introduce unrealistic indications  of
         actual vehicle kinetic energy status to the control system,
         especially if the rear drive wheel speed is used to measure
         vehicle speed in the control.
                                   -15-

-------
As described in more detail in Section VI-D, two types of control approaches
were investigated on a non-linear basis by simulating the complete propulsion
system as an analog computer (Appendix II).

Many operational features were common to both control approahces.  These in-
cluded:
     1.  Control of the primary transmission by a schedule cam from
         driver pedal position and measured engine speed.  The cam
         schedule can be contoured so that engine speed is varied to
         follow as closely as possible a minimum emissions curve for
         the engine.  Since no emissions data were available for the
         engine, the schedule cam was set for minimum specific fuel
         consumption (SFC),
     2.  A position on the driver shift lever for initial flywheel
         charging.  Other positions were similar to a present car
         with an automatic transmission.
     3.  Complete disengagement of the transmission at engine speeds
         belov 900 RPM to permit engine startup with the shift lever
         in either Neutral or Park.
     4.  Engagement of only the secondary transmission when the shift
         lever is in the charge position to permit rapid initial fly-
         wheel charging with the control automatically regulating engine
         power.
     5.  When the shift lever is in  Drive,  the primary transmission auto-
         matically disengages at zero vehicle speed while the secondary
         transmission remains engaged..  In addition to minimizing ideal
         fuel flow requirements, this permits flywheel charging at stops.
         The amount of engine power available for charging under these
         conditions is limited by engine capability at low speeds (about
         40 horsepower At 1200 RPM).

The major difference between the two approaches was the method of deriving
vehicle velocity for TKE when operating with the gear shift lever in the Drive
position.  The closed-loop control (Type A), shown by Figure III-2, measured
                                   -16-

-------
                                            FLYWHEEL
            RATE
            VALVF,
     I IRAKI'
Accelerator
   Pedal
                           SKCONDARY
                            CONTROL
  SKCONDARY
TRANSMISSION
                   RATE
                   VALVR
                              ENGINE
                                     ENGINE
                                    GOVERNOR
                                                            OUTPUT
                                                           GOVERNOR
                  PRIMARY
                TRANSMISSION
                                                            PRIMARY
                                                            CONTROL
                 Fig.  III-2   Closed-Loop  TKE  Control (Type A)
    BRAKE
    PEDAL
                               FLYWHEEL
                               GOVERNOR
  FLYWHEEL
  SECONDARY
TRANSMISSION
  ACCELERATOR
     PEDAL
                                    ENGINE
                                   GOVERNOR
                                                            PRIMARY
                                                         TRANSMISSION
                                                            PRIMARY
                                                            CONTROL
             Fig.  III-3   Open-Loop Scheduled TKE Control (Type  B)
                                   -17-
                                                                              MTI-12527

-------
 vehicle  speed  (transmission  output  speed)  and -flywheel  speed  to maintain TKE
 constant  by  modulating  the  secondary (flywheel)  ratio.   The open-loop  Type  B
 control  (Figure  III-3)  approximated vehicle  speed based upon  a  predetermined
 scheduled  with driver pedal  position and compared that  to measured  flywheel
 speed  in  order to  approximate corvsitant.TKE.  .

 UIKJ  to (.lie naLure  of the  TKE computation with  the Type  A cont.rol.  tliu  vehic. U<
 speed  feedback is  positive  (regenerative)  in that an increase: in vehicle spivd
 commands  a decrease  in  flywheel  speed which  in  turn results in  a hirther in-
 crease in  vehicle  speed.  This could result  in  a  difficult driver  control situ-
 ation  whenever wheel-slip occurs  (e.g., icy  roads).   Also,  as the  vehicle pro-
 ceeds  down a hill, the  resultant  increase  in speed commands the flywheel to
 accelerate the vehicle  even  more —  thus, a destabilizing feeling for the driver.

 Although control systems  can be  implemented  with  positive feedback  loops, the
 gain of  these loops must  be  less  than unity  for stable  operation.   As  shown by
 the analog computer analysis this restricts  the dynamic operation  of the system,
 especially with  the large variation in gains inherent to the  Type A control.

 With the open-loop Type 15 control approach (Figure HI- 3),  changes  in  vehicle
 load torque  from that torque used to establish  schedule settings will  cmise
 deviations from  desired TKE  as shown by Figure  III-4..   For  example, with a
 schedule set for perfect  TKE control on the  level  when  proceeding  up a  five
 percent grade, the flywheel  would not be able to  provide power  at speeds
 greater than 60 MPH.  However, this  situation is  somewhat analogous to  that
which  exists in a present car where  increased load torque decreases accelera-
 tion capability.

 In addition,  with the Type B control,  the driver  pedal  tends  to  act like a
 torque command for the propulsion system.  This is  a  desirable  feature  since
 it is  analogous  to present automotive  systems.  Thus, hills do  not  result in
 the flywheel influencing  vehicle speed unless the  driver commands it to.  Also,
 the driver can use the flywheel  (departing from TKE)  to aid in  climbing  hills
and for dynamic braking when descending hills.  However,  there  is a disadvantage
when going down a long steep hill when the driver  has his foot  off  the  pedal;
                                     -18-

-------
   1.2
   1.0
   0.8
H

TD
CO
01
Q
   0.6
   0.4
   0.2
          10% Increase in Vehicle Mass at 07. Grade
•Minimum Flywheel
      Speed
                                 Note:   Changes in wind load, fric-
                                        tion, and engine conditions
                                        have effect similar to
                                        change in grade.
                  20           40          60          80

                       Vehicle Speed,  Miles Per Hour
                   100
        Fig.  III-4   Effect on TKE of Varying Vehicle Mass and
                     Grade Conditions — Type B Open-Loop Control
                              -19-
                                                                        OTI-12499

-------
 eventually  the  flywheel reaches  its maximum  speed.  This  results  in  loss of
 dynamic  braking  from the  flywheel which could cause an undesirable feeling for
 the driver.

 On a comparative basis, although the Type A  control is initially  attractive
 because  it  measures vehicle speed to control TKE,  the Type B approach was
 selected for final design evaluation and costing at this  time  for the following
 reasons:
     1)  reduced complexity
     2)  lower cost
     3)  inherently better stability
     4)  better  operational performance and  driver feel characteristics

 From an  implementation viewpoint, a hydraulic-mechanical  type  of  control was
 considered  preferable on a near  term basis when compared  to other types (such
 as an electronic type), since it is compatible with the hydraulics in the trans-
 mission, power is readily available and maximum use can be made of control
 components  presently available,  in a conventional automatic transmission in
 order to reduce cost.  A detailed discussion of the implementation and opera-
 tion of  the control system is presented in Section VI-D.

 In summary, it is important to point out that (even with  the control approach
 selected) there are the previously discussed problems that remain to be solved,.
 These and others which, in general, related  to overall systems operation ,>;id
 interactions during the multitude of conditions under which a  family car must
 be capable  of operating must, be explored,.   In theory, the flywheel/hybrid pro-
 pulaion system is quite simple; however, from a practical viewpoint,  the control
 and system  interrelationships are much more complex than  those that exist in
 a present family car having an automatic transmission.

 D.  Comparative Performance
 Performance characteristics were determined for the transmission as an integral
 part of the flywheel/hybrid propulsion system.  A detailed discussion of these
 results is contained in Section VI-C.   Power levels selected included idle,
steady-state cruise, and a full power transient.   In addition, the time re-
quired  initially to charge the flywheel was evaluated.
                                    -20-

-------
The vehicle characteristics were defi.;:^U by the EPA/AAPS Design Goals (1).
The heat engine for the flywheel/hybrid Dropulsion system was a conventional
internal combustion engine with characteristics specified by EPA/AAPS.  Maxi-
mum engine power at 3800 RPM was approximately 177 horsepower.  Accessory
and/or auxiliary power losses assumed were typical for this size engine with
air conditioning included.

Comparable performance for a present day car were determined by using the same
vehicle characteristics with the aforementioned 1C engine and a conventional
automatic transmission whose characteristics were specified by EPA/AAPS.

1.  Idle Conditions
An important operating condition for a family car is when the vehicle is idling.
This condition occurs more frequently than any other on the DREW driving cycle
and, of course, is quite common to urban driving.  At idle conditions, the fuel
flow rate for the flywheel/hybrid system would be approximately 18 percent more
than the conventional system.  This increase was a direct result of the higher
engine idling speed required in order to have the capability to maintain flywheel
speed due to flywheel losses and possible deficiencies in TKE.  Recjlling that.
the primary transmission is disengaged at idle to minimize losses, these results
indicate comparatively poor performance for dhe flywheel/hybrid system at idle
conditions.

2.  Steady State Cruise 'Power Conditions
As described in Section VI-C, performance characteristics at cruise power were
calculated based upon a steady-state model of the complete propulsion system
which included transmission losses accounted for on a component basis as a
function of component operating levels (see Appendix I).

Under steady-state level road conditions for the selected vehicle characteris-
tics, the transmission was in the high-speed range for all vehicle speeds above
11 mph.   Based upon the detailed breakdown of transmission power of flow.'paths
given in s'ection VI-C, it was shown that for the aforementioned conditions  the
hydraulic elements were not an  important  factor in determining transmission
efficiency — particularly at vehicle speeds arouml 20 mph characteristic  of
                                   -21-

-------
 urban  driving.   For  example,  at  20 mph,  94 percent of the power in the primary
 flows  through  the  mechanical  path  to the rear wheels  with the remaining 6  per-
 cent flowing through the  hydraulic units.

 A  comparison between the  transmission efficiency  of the  flywheel/hybrid and  con-
 ventional  automatic  transmissions  is shown by Figure  III-5.   The transmission
 efficiency includes  the engine power to  the flywheel.  For the particular  fly-
 wheel  configuration  assumed,  the efficiency of the hybrid transmission ranged
 from 85  to 90  percent.  On  a  comparative basis, the hybrid transmission had  a
 higher efficiency  than the  conventional  transmission,  particularly at  cruise
 speeds from 20  to  40 mph.   For example,  at a  vehicle  speed of 25 mph,  the  effi-
 ciency was 20 percent higher  than  the conventional transmission.   However, de-
 creasing the flywheel losses  will  require  less power  flow in  the transmission
 and the effect  of  residual  (or parasitic)  power losses in the transmission will
 result in  a lower  transmission efficiency  at  low  vehicle  speeds.   As pointed out
 in Section VI-C, a substantial reduction (57)  percent) in flywheel losses  resul-
 ted in about a  13  percent decrease  in efficiency  from  that shown by Figure III-5
 at 10  mph.

 Figure II1-6 presents a comparison  of power train efficiencies  at  cruise power.
 Note that  the power  train efficiency (defined  on  Figure III-6)  does not  include
 flywheel power.  The  effect of flywheel  losses is  clearly evident  in the lower
 power  train efficiency of the flywheel/hybrid  transmission at  low  vehicle speeds.

 A more important comparison is on  the basis of fuel economy.   Figures  III-7  and
 III-8  present this comparison at cruise  power  conditions.  These  results indicate
 that the flywheel/hybrid propulsion  system will decrease  in cruise fuel economy
 at speeds  typically  encountered  in urban driving.  This loss  in  fuel economy is
 primarily  due to flywheel losses as  shown  by Figure III-8.  For  the higher loss
 flywheel (Pc =  2.94  psi) the  decrease  in miles per gallon  at  20 mph was 18 per-
 cent.  Reducing the  flywheel  power losses  by a total of 57 percent resulted  in
 improved fuel economy but it was still 10  percent  less than the  conventional
 system at  20 mph.  Thus, even with the improved performance capabilities of  the
variable ratio  transmission (compared  to the conventional  automatic), a decrease
 in cruise  performance at low  speeds  can be  expected.  At higher speeds, as shown
by Figure  III-7, an  improvement up to nine  percent in fuel economy was obtained
since  flywheel losses are minimal.

                                     -22-

-------
         100
NJ
U)
                                                   Flywheel/Hybrid  Transmission
                                                    (With  Inline  Pierced Flywheel, P  = 2.94 psi)
                                                               Standard Automatic  Transmission (Without Flywheel)
IVehicle Weight = 4600  Ibs
}07, Grade
-Engine Accessories  Include Air Conditioner—:
                                                              Transmission Efficiency (Cruise) =
          20
                                                            40          50
                                                          Vehicle Speed, MPH
                              Fig. III-5    Comparison  of Transmission  Efficiencies  at Cruise Power
                                                                                                                       MTI-12492

-------
„.
                          F        ~T  T~Y
                                           -y-Standard Automatic Transmission
                                           *  (Without Flywheel)
  - . _4	 ..  i —

r • ' ~
 80 .	
                                                        h     i     !•    \            I
                                                     Flywheel/Hybrid Transmission Power Train
                                                     (With  Inline Pierced Flywheel, P  = 2.94 psi)
                                                      Cruise Power Train Efficiency
                                                 40          50

                                               Vehicle Speed,  MPH
                       Fig. III-6    Power Train Efficiency Comparison  at  Cruise Power
                                                                                                          MTI-12493

-------
NJ
Ul
I
          20
          18
::».!~F4f;!--!  i-- i  '^fc;;U  1 ,  •   i
   Vehicle Weight = 4600  Ibs
 j_ 0% Grade
 * Fuel Density = 6.152  Ib/gal

  ~Engine Accessories Include Air  Conditioner

                                                                          ^Standard Automatic  Transmission
                                                                          ^(Without Flywheel)
                                                        Flywheel/Hybrid Transmission
                                                        (With Inline Pierced Flywheel,±
                                                           = 2.94 psi)
                                                                  ~  '
                                                    .:— 3.-.:. ....- -;.: i t±:
                                                30           40           50

                                                         Vehicle  Speed,  MPH
                                     Fig.  III-7   Comparison of Fuel Economy at Cruise Power
                                                                                                                       MTI-12494

-------
   20
      Steady-State Cruise Power
     !Vehicle Weight = 4600 Ibs
      0% Grade
     .Fuel Density = 6.152 Ib/gal.
      Engine Accessories Include Air Conditioner
u
&
K
                                                 Pierced  Flywheel,  P
                                                            Pierced Flywheel, P  =2.94 psi
                                                                               c
                                                   40          50
                                                 Vehicle Speed, MPH
                      Fig. III-8   Percentage Change in Cruise MPG Compared to Conventional
                                   Automatic  Power Train
                                                                                                            MTI-12495

-------
Engine operation at minimum specific fuel consumption (SFC) conditions is, of
course, desirable in order to achieve good fuel economy.  However, since no
emission data were available for the engine during the course of this study,
it cannot: be ascertained at this time whether this type of engine operation
would indeed result in minimizing engine emissions.

Consequently, the transmission and control systems were designed to control the
engine so that. it. operated at minimum SFC as much as possible based upon the
assumption that, if this did not result in minimizing emissions, some ot.her
operating schedule which did so could be achieved without extensive redesign.
In order to permit random flywheel charging and for other reasons (see Section
VI-C) a minimum engine speed of 1400 RPM was selected.  Whenever the engine
power demanded is less than the minimum SFC horsepower at 1400 RPM. operation at
a higher SFC will result.

A comparison of the engine SFC at cruise power is shown by Figure III-9.   In
general, the flywheel/hybrid propulsion system tended to result in operation
closer to minimum SFC than did the conventional system.  At 10 mph. a decrease
of 20 to 32 percent (dependent upon flywheel losses) in engine SFC was obtained.
Operation at improved engine SFC was due to flywheel losses (more engine power
required) and the design features of the transmission.

Since ('he size of the heat engine required for the flywheel/hybrid propulsion
system is dependent upon the selection of maximum steady-state cruise power.
EPA/AAPS concurred that, a 5300-pound vehicle should be capable of climbing a
five percent grade at 70 mph without flywheel acceleration power for a period
of 100 seconds.

After accounting for all losses (including transmission losses), an engine
capable of developing at least 130 horsepower will be required:   26 percent
smaller than the 177 horsepower engine specified for this study.  This is an
indication of the reduction in engine size possible for the flywheel/hybrid
system compared  to an engine commonly used in present day propulsion systems
for medium-sized family cars.
                                   -27-

-------
         50
         40
               Steady-State  Cruise  Power
               Vehicle Weight = 4600  Ibs
               0% Grade
               Engine Accessories Include Air  Conditioner
00
I
                                    Pierced Flywheel, P  = 2.94 psi
                                            Pierced Flywheel, P  = 0.581 psi
                       10
   40          50
Vehicle Speed, MPH
                                                                                                                      90
                                 Fig. Ill-9    Percent Decrease  in Cruise SFC Compared To
                                               Conventional Automatic Power Train
                                                                                                                    MTI-12549

-------
3.  Full-Power Transient
In order to obtain dynamic performance results a nonlinear model of the complete
propulsion system was set up on the digital computer.  The model simulated all
significant dynamics (torque and speed relationships) of the flywheel, trans-
mission witli controls, engine, and family car in the time domain.  Compress ibi l-
ily dynamics in both hydraulic pump-motor circuits were  included and transmission
losses were calculated as described in Appendix I.  Engagement and d^engagement
of: (.he secondary and primary transmission were not included.  Input to The model
to initiate a transient was driver command pedal as a function of time.

The control system was the open-loop Type B control (discussed earlier) setjp
for TKE with flywheel speed varying between 24,000 rpm and 10.000 rpm for
vehicle speeds of 0 to 85 mph.  Control gains and rate limits were adjusted to
maintain stability, but were not optimized for maximum performance capability.
Additional tuning would have resulted in obtaining further improvement, in vehi-
cle acceleration characteristics.

The results of the full power transient demonstrated the performance capabili-
ties of the transmission under a "worst case" condition with respect re the
efficiency of power transfer from the engine and flywheel.  In addition, it
demonstrated the capabilities of the design under simulated operation condi-
tions .

A detailed discussion of the results obtained is contained in Section VI-C.  A
summary of those results is presented in the following discussion.

Based upon an energy balance of the full-power transient, the following results
were obtained:

     1.   The effective efficiency of the transmission was 89.5 percent
         and the power train efficiency was 86 percent.
     2.   The engine supplied approximately 13 percent less energy than Lt
         should have for ideal control;  consequently,  the flywheel had to
         provide the additional energy.
                                  -29-

-------
     3.  The engine supplied 41 percent of  the  input energy and the  flywheel
         the remainder.  This  indicated that the engine could have been re-
         duced about 50 percent in size if  it were not for steady-state require-
         ments .

Those  results  show that, although further tuning of the conlrol system  Is re-
quired., the transmission operated at. a relatively high efficiency and that, on
an energy basis, the system operated reasonably close t:o desired condif. ions.

Figure III-10  presents the transmission efficiency obtained during a full power
transient as a function of vehicle speed.   The  lowest efficiency (78 percent)
occurred at 25 mph, where the  primary shifts range under full power conditions.

Figure III-11  presents a comparison with a conventional automatic transmission
on the basis of transmission efficiencies.  These results indicate that the
conventional transmission has  better full-power efficiency than the flywheel/
hybrid transmission at speeds  between 15 and 30 mph and at very high speeds.
For example, at 25 mph the efficiency was 13 percent lower than a conventional
transmission.   The lower efficiency occurred at the range switching point.
At. lower power levels, range switching occurs at lower vehicle speeds — II mph
;jt cruise power.  Thus, under  normal driving conditions (power levels less
than maximum), the efficiency  of the flywheel/hybrid transmission should be
comparable to  a conventional automatic transmission.

A comparison on the basis of fuel economy (Figure III-12) clearly shows the
advantage of the flywheel even though engine operation was not optimized.   For
about  three seconds during the initial portions of the full-power transient
(when  the engine operates at low power levels), the engine SFC was about 50
percent higher than with a conventional system.

These  results   indicate that further improvements in the design will be
necessary in order to obtain a better engine SFC during maximum acceleration.
At lower acceleration rates (gradual changes in the driver pedal)  more common
to the DHEW driving cycle, the engine would tend to operate close to the
minimum SFC operating conditions.   As an example —• with output power equiva-
lent to a five percent grade — engine operation at minimum SFC was obtained
for all vehicle speeds above 25 mph.
                                   -30-

-------
  100
      Vehicle Weight =  5300  Ibs
      0% Grade
      Engine Accessories Include  Air  Conditioner
                                                  Full.  Power Transient
                             i                     TUJ
                       '" ~	I"      i    H   H~
I 60
                                                                                                       ,
                                        Transmission Efficiency (Acceleration)

•


j


. ,v
, 4 . : •
i






rfc




.
-.
"' : ' rt
; -
-;. .. . J

!
!
i


- - .

:


                                         30
   40         50

Vehicle Speed, MPH
60
70
80
90
                   Fig.  111-10   Flywheel/Hybrid  Transmission Efficiency at Maximum  Power
                                                                                                                MTI-1254*

-------
         100
LO
ro
i
              t I  —i	»—«^ I—r-
             -U-1 4-4-;
                                                        Conventional Automatic  Transmission (Based  Upon

                                                        Steady State Operation  at Maximum Power)
                                   4_i- .J.J.J.M-- • .-• <	-i	-4	i ----...i U44-
                                  FljwheeI/Hybrid Transmission  (Based Upon Full

                                 LPower Transient with Pierced  Flywheel, P  = 2.94 psi
                                                                            ;Vehicle Weight =  5300  Ibs

                                                                             0% Grade
                                                                            .Engine Accessories  Include Air Conditioner
                                                                           Transmission Efficiency (Acceleration)
                                                             40          50

                                                           Vehicle Speed, MPH
                                Fig.  III-H    Comparison  of Transmission Efficiencies at  Full Power
                                                                                                                         MTI-12541

-------
          20


          18


          16
Vehicle Weight = 5300 Ibs
0% Grade
Engine Accessories Include Air Conditioner
Fuel Density = 6.152 Ib/gal.
i
UJ
LJ
          14
                                                        4-
                                                                               	I	,	
                                                        Flywheel/Hybrid Transmission  (Based Upon  Full
                                                        Power Transient with Pierced  Flywheel,  P   =2.94  psi)
                                                   .   I
                                                         Conventional Automatic Transmission (Based Upon
                                                         Steady State Operation at Maximum Power)

                                                                                               70
                                            40          50
                                         Vehicle Speed, MPH
60
80
90
                                    Fig.  111-12   Comparison  of Full Power Fuel  Economy
                                                                                                                     HTI-12543


-------
4.  Flywheel Charging Times
As part of  Che subject investigation  it was of significance  to determine what
magnitude of charging times could be  expected for  the system.  In order to
accomplish  this, it was assumed that  there is a fixed ratio  between the engine
speed and the flywheel speed.  This implies that the hydraulic elements were
not. taking  part in the charging cycle and that the charging  was directed
through the gearing with the ring gear held stationary.  The  inertia of t.he
flywheel was referred to engine speed and the torque available for acceleration
uva determined from the engine performance map which related  horsepower JIH!
speed..

Knowing torque and reflected inertia., the average  rat.e of acceleration could
be determined.  The acceleration and  the speed increment were then used to
establish the time required between speed steps.

The subject calculation was made for  charging at minimum SFC  as well as
charging at maximum engine horsepower.  Two end points of engine speed were
also considered, 3200 and 3800 rpm.  The times required for  full charging was
greatest at 44.9 seconds at minimum SFC hp and an engine speed of 3200 rpm;
and least,  25.1 seconds at maximum engine hp and an engine speed of 3800 rpm.
This was well-within the goal that 65 percent, of full power  be available
within 45 seconds..

E.  Safety Aspects
Since the transmission controls the flywheel power flow, a preliminary analysis
was made with regard to safety aspects.   Based upon t.he details given in
Section VI--F, it. was concluded that failures of certain transmission elements
and control failures leading to primary and secondary transmission ratio
errors can  result in potentially dangerous vehicle acceleration due to the
energy storage capability of the flywheel.

In analyzing these  types of failures,  it: was assumed that, they would require
the same type of driver reactions as conventional vehicles require.   The mal-
functions  noted above,  which fall into the pathological category,  would not be
immediately corrected by the normal reactions of the driver.   For example,
                                   -34-

-------
under abnormal conditions, t.he drive', would  lift, his foot, from the accelerator
and apply  the brakes expecting to slow.  He  would still note an increase  in
vehicle speed or no indication of deceleration for a period of time.  Similar
comments apply to the condition where unexpected decelerations take place.

In summary, based upon the preliminary analysis conducted in this study (see
Section VI-F), it was concluded that, a more  detailed investigation of the cri-
tical safety aspects identified herein and t.he remifications of their solution
is required..

Although not included in our considerations, it is very probable that, should
the flywheel/hybrid propulsion system come into common usage, certain correct.ive
actions could be built in and integrated with what normally would be expected
from the driver.  This would require added control complexity.

F   Regenerative Braking Analysis
In reviewing the requirements for dynamic (regenerative) braking, it was  impor-
tant, to assess the feasibility of such braking and to determine if the use of
flywheel dynamic braking would unduly penalize the overall transmission design.
In the latter case, this would require a trade-off between any added complexity
or design compromise and the extent of their benefit.   In terms of assessing
feasibility, this was considered on the basis of a total energy concept, prac-
tical, braking limits for vehicles, peak torque and horsepower requirements, and
any other influences brought: about by the flywheel.

The first consideration with respect to braking was  based on the premise that
the driver must be capable of making an emergency stop which for purposes of
                                                           2
this  study was taken to be a deceleration rate of 20 ft/sec .   This indicated
that  with rear wheel drive,  a maximum of 35 percent  of the required braking
force could be realized through the rear wheels.   This results from the fact
that,  at  this deceleration rate, the vehicle weight load on the rear wheels
is 35 percent.   Thus,  it was concluded that four-wheel mechanical braking was
definitely required, that in an emergency situation  there is a severe
limitation on regeneration and that additional engine  energy sometime during
the cycle would be required  to bring the flywheel back to the  desired energy
level.
                                   -35-

-------
                                                                      2
 Further  analyses  indicated  that,  at  a  deceleration  rate  of  8.2  ft/sec  ,  it was
 possible  to have  regeneration.  However,  the sizing of the  flywheel  transmission
 would have to be  changed.   In  fact,  the capacity would have  to  be doubled.  As
 shown in  Section  VI-G, design  modifications would be required to achieve  the
 above.  The estimate of cost change  would be an additional  12 to 15  percent of
 overall  transmission cost.
With  the transmission design unaltered,  the flywheel system could be utilized
                                                               2
as a  retarder  to provide deceleration capability of 4.41 ft/sec  with full re-
generation.
 It was concluded tihat some benefit as a retarder could be realized, but this
 would not negate the requirement for four-wheel mechanical brakes.  Provisions
 would have to be made for different levels of braking command such that this
 retarder action could be realized.

 Further details regarding regenerative braking are given in Section VI-G.

 G.  Cost and Physical Comparisons
 One of the most important aspects of any system being considered for automotive
 use is cost.  Consequently, a detailed cost analysis (Section VI-E) was per-
 formed using procedures currently practiced by the Ford Motor Company, in order
 to determine the cost of the transmission and control system design presented
 herein.  This analysis consisted of determining detailed costs for approxi-
mately 390 separate parts by procedures which are currently practiced in the
 automotive industry.  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" (OEM) basis, since this approach is common
 to the automotive industry when comparing designs.  Briefly, variable costs
 include the cost of material (in "as purchased" condition)  and in-house costs
 to complete the necessary manufacturing and assembly.  The latter costs include
 direct labor, indirect labor and nonvariable burden.  The OEM costs (specified
 by EPS/AAPS) were estimated from..'the aggregate-variable' oest.
                                   -36-

-------
 A summary of the cost analysis results is presented by Table III-2.  These re-
 sults show that in production quantities of 1,000,000 units per year, the vari-
 able cost of this transmission will be at least 2.40 times a conventional auto-
 matic transmission — an increase of 140 percent.  On an OEM cost basis, the
 minimum increase ranged from 125 to 135 percent.  For smaller production quanti-
 ties (100,000 units per year), the minimum OEM cost was 124 to 135 percent higher
 than a conventional automatic transmission dependent upon the type of production
 tooling assumed.  This increase in cost was primarily due to:
    1.  Two power sources - this required two transmissions to he packaged Into a
        single unit and the costs reflect the torque transfer elements of two
        transmissions.
    2.  Added control complexity - the control system had more than twice as mary
        parts compared to the controls for a conventional automatic transmission.
 The results shown by Table III-2 also indicate that in order to minimize the
 cost of the transmission, it is desirable to have the axis of the flywheel in
 line with the engine shaft.

 Another important comparison is on the basis of propulsion system costs.  In or-
 der to determine this comparison, the total OEM cost of 1) a smaller engine,
 2) the flywheel*, and 3) associated transmission were ratioed to the equivalent
 cost of a larger engine and an automatic transmission.   These results showed that
 in production quantities of 1,000,000 units/year the ratio in OEM cost was 1.63.
 Thus, the flywheel/hybrid propulsion system can be expected to cost at least 60
 to 70 percent more than a comparable propulsion system currently used in auto-
 mobiles.

 Consider now a comparison of pertinent physical characteristics.   As summarized
 by Table III-3, the transmission design presented herein was five percent larger
 in volume, 75 percent heavier and had 65 percent more parts than a comparable
 conventional automatic transmission.  These results clearly indicate the in-
 creased complexity of the transmission compared to a conventional automatic
 transmission.
•-•The flywheel assumed was the in-line pierced configuration.   Cost data for the
flywheel and its associated parts were supplied by LMSC.
                                     -37-

-------
                                                           TABLE III-2

                                            Final Transr-.i ssian. Cc5t Analvsjs~Ratios
                                        Standard
                                        Multi-Speed Torque
                                        Converter  ("Automatic")
                                        Transmission  for
                                        Medium  Size Vehicle
Power
Splitting Transmission
for Flywheel/Heat
Engine Medium Size
Vehicle (Inline Flywheel)
Power Splitting
Transmission  for
Flywheel/Heat Engine Medium
Size Vehicle  (Offset Flywheel)

1) Variable Cost Ratio
2) O.E.M. Cost Ratio
3) Control Variable Cost Ratio
4) Labor Content Ratio
5) Material Content Ratio
A
1.00
1.00
1.00
1.00
1.00
.B
1'.19
1.15-1.25
1.19
1.00
1.20
C
1.29
1.25-1.35
1.29
1.50
1.20
A
2.40
2.25-2.35
2.87
2.13
2.65
B
2.85
2.70-2.80
3.41
2.13
3.18
C
3.10
2.95-3.05
3.70
3.19
3.18
A
2.53
2.35-2.U5
2.87
2.25
2.79
B
3.01
2.85-2.95
3.41
2.25
3.35
C
3.26
3.10-3.20
3.70
3.37
3.35
I
UJ
00
I
              A)   Ratios based on manufacture of 1,000,000 units per year

              3)   Ratios based on manufacture of 100,000 units per year with  tooling as  for  1,000,000 units

              c)   Ratios based on manufacture of 100,000 units per year with  tooling suitable  for  maximum yearly
                   manufacture of 100,000 units.
                                                                                                                            MTI-12540

-------
TABLE III-3   TRANSMISSION  PHYSICAL COMPARISONS

Transmission
Volume - Ft3
Control Volume
Ft3
Transmission
Weight - Lbs.
Control
Weight - Lbs.
Transmission -
Number of Parts
Control -
Number of Parts

Total Transmission
and Control Volume
Total Transmission
and Control Weight
Total Transmission
and Control Parts
Flywheel/Hybrid
Transmission
1.70 Ft3
.40 Ft3
218 Lbs
31 Lbs
235
155

2.1 Ft3
249 Lbs
390
Automatic Three Speed
With Torque Converter
Transmission
1.85 Ft3
.15 Ft3
123 Lbs
19 Lbs
166
70

2.0 Ft3
142 Lbs
236
Percent Increase Compared
to Conventional Automatic
3-Speed Transmission


5%
75%
65%
  * Does not include  the volume of the flywheel.
                    -39-

-------
On a propulsion system basis, the weight of the flywheel/hybrid propulsion
(Table III-4) was just barely within the limits specified by EPA/AAPS Design
Goals.  Table III-5 shows that the volume of the complete propulsion system
was well within EPA/AAPS allowable limits.
                                    -40-

-------
                                 TABLE III-4
                               Weight Summary
Weight of 1970 Propulsion Systems

  "    "   "   Transmissions

  "    "   "   Propulsion System Minus
 Transmission

Weight of Hybrid Propulsion System Flywheel

Weight of Hybrid Propulsion System Power
 Splitting Transmission
Total Weight of Hybrid Propulsion System

Total Allowable System Weight
                                                      1300 Pounds

                                                       140 Pounds

                                                      1160 Pounds


                                                       185 Pounds


                                                       249 Pounds
                                                      1594 Pounds

                                                      1600 Pounds
                                TABLE III-5

                               Volume Summary
Volume I.C. Engine and Accessories
Volume I.C. Engine Cooling and Heating System
Volume Drive Axle
Volume Rear Axle
Volume Fuel Tank
Volume Exhaust System
Volume Flywheel
Volume Power Splitting Transmission
    Total Volume of Hybrid Propulsion Syste
                                           m
    Total Allowable System Volume
                                                  =  20.2 FtJ


                                                      3.0 Ft3


                                                       .5 Ft3


                                                      1.1 Ft3


                                                  =   2.7 Ft3


                                                       .7 Ft3


                                                      1.3 Ft3


                                                      2.1 Ft3
                                                  =   31.6  Ff
                                                  =   35    Ft'
                                    -41-

-------
                                 SECTION IV
                       CONCLUSIONS AND RECOMMENDATIONS

1.  As a result of evaluating various transmission concepts, the power-splitting
    transmission was considered to be the best for the flywheel/hybrid propul-
    sion system.  Thus, a preliminary de°ign of this type of transmission was
    made and evaluated in this study.

2.  The transmission design presented herein has sufficient flexibility to
    accommodate flywheel speeds ranging from 8,000 to 24,000 rpm without major
    design changes.  Therefore, the operating speed ratio>'of the flywheel can
    be based upon considerations other than transmission size.

3.  After accounting for all losses (including transmission losses), it was
    determined that an engine capable of developing at least 130 horsepower is
    needed to meet EPA/AAPS requirements:   26 percent smaller than the engine
    specified for this study (177 horsepower).  This is an indication of the
    reduction in engine size possible for the flywheel/hybrid system compared
    to an engine typically used for present day medium sized family cars.

4.  Significant deviations from the desired total energy concept will be
                                                                  2
    encountered during rapid or emergency decelerations (20 ft/sec ).  Under
    these conditions, the rear wheels can regenerate 35 percent or less of the
    required flywheel charge.   Since full benefit of regenerative braking could
    not be realized:   a)  a conventional braking system was recommended for the
    vehicle and  b) for lower cost, the transmission was sized  compatible with
    acceleration requirements.

                                                           2
5.  With the selected transmission si::e,  during a 20 ft/sec  deceleration,  only
    17 percent of the required  flywheel charge can be regenerated through the
    rear wheels.   The engine supplies the  additional energy to  return the
    system to normal  energy levels.  However,  at lower deceleration rates,
    characteristic of driving cycle operation, the flywheel (as a retarder)
    can be fully regenerated except for losses (transmission, dissipative
    braking, flywheel losses,  etc.),
                                    -43-

-------
 6.  Based upon stability, system, and cost considerations, an open-loop
     scheduling-type of control was selected for the transmission.  However,
     additional studies are required in order to more thoroughly investigate
     system interactions, operational aspects, and to optimize the control sys-
     tem.  The selected control approach provides perfect IKE control for a given
     schedule of vehicle load and approximated it at other loads.  In addition,
     it provides more flexible operation of t.he flywheel on hills and under
     wheel-slip conditions.

 7.  At. idle conditions., the fuel flow rate for flywheel/hybrid system would lie
     approximately 18 percent more than a conventional system.

 8.  At steady-state cruise power conditions, the efficiency of the transmission
     (after accounting for flywheel loss power) is better (20 to 25 percent
     higher) than a conventional automatic transmission.  However, the fuel
     economy of the propulsion system would be 10 to 18 percent lower at 20 mph
     than a conventional system due to flywheel losses.

 9.  At maximum power conditions, the efficiency of the transmission was 13 per-
     cent, lower than a conventional transmission at 25 mph.  At. other vehicle
     speeds, the efficiency was essentially the same as a conventional trans-
     mission .

10.  The engine would operate at better SFC than a conventional system at cruise
     power conditions; at 10 mph the decrease (improvement) in SFC was 20 to 30
     percent.  Under maximum power accelerations, the engine operates for 3
     seconds at 50% poorer SFC than a conventional system;  thus,  further design
     improvements are required.  At lower, acceleration rates,  t.he engine tends  to
     operate close to minimum SFC.  With output power equivalent,  to that re-
     quired for a 5 percent grade, the engine would operate at minimum SFC for
     all vehicle speeds above 25 mph.

11.  Failures leading to primary and secondary ratio errors can result in poten-
     tially dangerous vehicle accelerations or decelerations,   A  more detailed
     investigation of critical safety aspects and the ramification of their so-
     lution is required.
                                     -44-

-------
12.  Even though Che transmission was designed Co be compact, lightweight, and
     highly efficient, in comparison to a conventional automatic transmission, it
     was 75 percent heavier (249 versus 142 pounds) and would have 65 percent more
     parts (390 versus 236).

13.  The OEM cost of the transmission is at least 125 percent higher than a
     standard automatic transmission for production quantities of 1,000,000
     units per year.  On a variable cost basis (more meaningful to the automotive
     industry), the increase in cost would be 140 percent.  Decreasing the quan-
     tity of production had no major effect on cost comparisons.

14.  Consideration of propulsion system costs showed that the flywheel/hybrid
     propulsion system can be expected to have an OEM cost at least 60 to 70 per-
     cent more than a comparable propulsion system currently used in automobiles,

15.  On a propulsion system basis, the weight of the flywheel/hybrid propulsion
     is just within the limits specified by EPA/AAPS Design Goals, while the total
     volume <*f the complete propulsion system is well within EPA/AAPS allowable
     limits.

16.  System complexity and developments which are required for safety (contain-
     ment, overriding mechanisms, etc.,) as well as further control development;
     suggest that the practical production implementation of the transmission
     would not be possible by 1975.

17.  Based upon the above, development of this transmission is not recommended
     until further evaluations are conducted.   As part of these evaluations,  the
     flywheel/hybrid propulsion should be compared on a trade-off basis  to other
     alternative propulsion systems for the family car.

18.  As pointed out in the following section of this report, the basic power-
     splitting transmission (without provisions for the flywheel)  would  have  ap-
     plicability to several engine types including the conventional  1C engine,
     the gas  turbine engine, the turbo-Rankine engine, and the diesel derivative
     engines.
                                     -45-

-------
                                    SECTION V
                     APPLICATION TO OTHER PROPULSION SYSTEMS

The desirability of an infinitely variable transmission for vehicular applica-
tions is generally accepted.  In terms of the technology, the ability to do
this is available in both mechanical and hydraulic components.  To date, the
mechanical configurations have been limited by virtue of their stage of devel-
opment or by virtue of limitations in power level where they can operate effi-
ciently.  Similarly, hydraulic elements such as a pure hydrostatic transmission
accomplish 'the function; however; they are'limited in terms of the hydraulic
efficiencies.  Thus the split transmission combining mechanical and hydraulic
elements represents a logical path to pursue.  Its major advantage is the
ability to utilize the hydraulic elements for smoothly changing ratio and
range while utilizing highly efficient mechanical gearing to transmit the
bulk, of the power flow.

Although many of the engine requirements particularly those which show promise
for reducing vehicular contributions to pollution can effectively utilize an
infinitely variable transmission, it is well to review a few of the important
benefits which would accrue.  In the Brayton cycle, or gas turbine area, it is
desirable to maintain the engine at a given speed, that is, the high efficiency
point at the desired pressure ratio.  Where it is not possible to do this
because of limiting ratios available in the transmission or because of the
excessive losses during the shifting process; two spool type gas turbines have
been-extensively utilized.  This configuration tends to approach the problem
by keeping the gas generator at essentially constant speed and varying the
output of the power turbine.  Not only is the mechanical configuration more
complex and expensive,  but the controls are  equally affected.  An infinitely
variable transmission which did not incur severe penalties over a wide oper-
ating speed range would permit the use of the more simple single shaft machine.

Similarly,  the Turbo-Rankine cycle engines have an optimum speed where maximum
efficiency can be realized.  Such engines as the Wanke1 have characteristics
similar to the diesel and the internal combustion heat engine; ?s such,

                                     -47-

-------
matching of engine characteristics for desired SFC performance and subsequently
emissions and pollutants is well-known.

As a result of the current program,  the design of a transmission which can
provide a complete variable ratio was evolved.  Although this transmission was
designed for combined flywheel/heat engine hybrid propulsion requirements, the
portion between heat engine and the vehicle represents a design which, with
adaption and optimization, can provide performance benefits t:o the engines
listed above.  The current program has demonstrated advantages of this trans-
mission in terms of many of the features required by passenger cars.

In order to demonstrate the application of this type of transmission  to other
prime movers, estimates were made of the characteristics of the transmission
for constant-speed inputs at two different power levels: maximum and  cruise.
The road horsepower levels selected are shown by Figure V-l.  The maximum power
level selected corresponds to 140 horsepower at the road up to the skid limits
of the rear wheels.  This should be quite sufficient to meet: EPA/AAPS acceler-
ation power specifications.

The resultant efficiency of the transmission for the selected power levels is
shown by Figure V-2.  It is important to note that this transmission was not
optimized for the Turbo-Rankine application and further Improvements  in per-
formance would result from the program proposed herein.

The power train (or overall) efficiency from the turbine output shaft to road
horsepower is shown by Figure V-3.

A comparison of efficiency between this transmission and a variable-blade-angle,
torque-converter, multi-speed transmission is shown by Figure V-4.  The trans-
mitted power level of the variable torque converter for which data was available
was approximately 100 road horsepower at. 80 mph.  These results indicate a
significant improvement in power train efficiency.

Thus the improvement in efficiency coupled with the completely variable ratio
capability of this type of transmission should result, in optimum performance
(maximum MPG) capability for engines such as the Turbo-Rankine over the
                                    -48-

-------
   160
   140
  100
I
e.

-------
  100
   80
S
g
*J
U-l
w
   60
   40

   20

                     :
                             •

                                                                   See  Figure V-l  for  Output  Power  Levels

i i : c±
..,._!.
. !


±up:
..
' : :
[- •
f "•
• ,
;


11.
. . . -

                        -:   I

                                                    J

                         j .
                         •
                                                            !"
                                                V
                                                                              I

                10
20
30
  40          50

Vehicle Speed, MPH
60
70
80
•     -;
L ..._._!
     90
                     Fig. V-2   Transmission Efficiency of MTI Variable Ratio Transmission at
                                Constant Input Speed

-------
100
                                                 Cruise  Power
r, Percent
5 5
•
Efficiencj
P- c
D C
t
i ' :
i
20
0
,...._. _ ,| _ 	

: ' !
... ...„,. ,.
....

L !
* i

. i. '..




.. . . ..T
-*-*ii
1 -
._.:._..
• -
	

: I T" . ; ; .'
j • : , ;



;




~ . i . i
.. rihj .
-:--t -<•<"-
t- -- • 4-

:
; rrf
TTTI

:i ri±
:. . ..; U
i T _T
h»»
•»,

• »


-

- *
- 1 .-.






. .1. -.

. _p
IT
...i_


:j' ..
_L
— i — i. —
:rf
_, ^.
_ *—

,
•
i ' ;
. : - - •
. ,_.k. ..._ .




"1 ' "


.)-4 	 i_!~
. ,
• r rtt

Jttt
, . .„, ,
i — L .i-i 	 i...
i*
.
1
'

. '.
' ". '
ttj
r— r-- • -
~ ~ r "~'
: •• !


7 - ; ! i




|
\
Maximum Power
- ...
: 	 , I
:


St±t±--: .;;-' •
. r - *• J 'i


Constant „
Spec
ft±'
_;it£
1 1 "7~r~p

d Turbine
Hp
ifr T- • ' i
*• } 	 •' •
- 	 t -;
, : T

...'


.: ..I. ... . .. j .. .
HPR
fficiency = — ~
HPIN
• •{- r-| i- • 	 f- f
••-;:• 	 h 	 i 	



: — 	 ,.._.
. . .U-L-4—
i
H||

.ssion
iH;.
t
. | !
j-
:
j
4^..

*
_

—




t
•
t
•
~H

5:



N


J , j. ,j ,
L .* ... -i-.

1
.. !.;,. <

Road

ifitt:
: : : •
•4-\--rr~
i f E 	 •;
	 	 	 :
' j
I 1 - : •
...
	 -t ~ - 1 t . .
Horsepower
o«
PR
.:LL." .;.'!
i • I ' . . ':
-. ' ; : : : r
t . - . i ; *
                          20
30          40          50
         Vehicle Speed, MPH
                                                                          60
70
80
90
                   Fig. V-3   Power Train Efficiency of MTI 150A Transmission for Constant
                              Speed Turbine Engine

-------
100
 80
-u
c
0)

JJ  60

2
>»

|
-H
U

£  40
Cd
 20
                                                  MTI 150A Transmission

                          .

               _
                                 -
                                 i
              10
                            20
                                        30
                                               Typical Variable Angle Torque  Converter

                                               with 3-Speed Transmission

                                                         .
                                                      	i	
                                                               i
                                                          1	'..>.
                                                                                                   r
                                                                                                  - -r	i-	


  40          50


Vehicle Speed, MPH
60
70
 '
       L

80
90
              Fig.  V-4   Power Train Efficiency Comparison at Maximum Power and Constant  Input  Speed

-------
complete vehicle operating range.  :: is also important  to note  that,  at  idle
conditions (zero output speed), the tr_.v_':i.ission  is automatically disengaged
which minimizes the idle fuel  flow required  for  turbine-type engines.

In conclusion, it potential exists for improving  the performance  and  reducing
emissions of various engine types currently  under consideration  by EPA/AAPS
with this type of transmission design.
                                     -53-

-------
                                                                           CO
                                                                           T)
                                                                           TJ
                                                                           o
                                                                           O
                                                                           3D
VI.  SUPPORTING INFORMATION

A.   SELECTION OF CANDIDATE TRANSMISSION
B.   DETAILED DESCRIPTION OF TRANSMISSION
C.   PERFORMANCE ANALYSIS
D.   CONTROLS DESIGN AND ANALYSIS
E.   COST ANALYSIS
F.   SAFETY ANALYSIS
G.   REGENERATIVE BRAKING
H.   REFERENCES

-------
                                                                  o
                                                                  o

                                                                  o
                                                                  n>
                                                                  T3
             SECTION A



SELECTION OF CANDIDATE TRANSMISSION

-------
                     A.  SELECTION OF CANDIDATE TRANSMISSION

This section deals with the initial technical task of the heat engine Elywheel
propulsion system transmission study.  It consisted of searching for power
transmitting devices and screening these devices to select those most promising.
The most promising candidates were subsequently combined into transmission
concepts to transmit torque and speed from both the heat engine and the fly-
wheel storage system to the rear wheel of a medium-size vehicle.

The basic methodology employed was to go through the following procedure:

   1.  Establish criteria for both selection and evaluation.
   2.  Gather data on available candidates.
   3.  Screen all but the most promising.
   4.  Combine these into overall transmission concepts.
   5.  Evaluate and review these concepts.
   6.  Select the most promising candidate for analysis.
In terms of the program,  shown below is the flow chart  of the  process  discussed
above.
   ESTABLISH CRITERIA
  (SCREENING FACTORS)
   CONSIDER VARIOUS
   TYPES OF POWER
   TRANSMITTING
   DEVICES
COMBINE THE MOST
PROMISING INTO
TRANSMISSION
CONCEPTS
1 EVALUATE AND SCREEN
! CONCEPTS
                                                       CANDIDATE
                                                       TRANSMISSION

-------
Evaluation Factors

The application required that the two sources of energy input be joined by
the transmission to provide one source of output.  Such torque transmitting
components are many and varied; therefore, to expeditiously arrive at a promis-
ing candidate evaluation, screening factors were established using the com-
bined engineering experience of MTI, Bendix and Consulting personnel.  The
factors were as follows:

   1.  Availability of components
   2.  Control and stability aspects
   3.  Compatibility with vehicular requirements
   4.  Safety
   5.  Performance
        • efficiency
        • emissions level of system
        • fuel economy
        • dynamic response
        • operational aspects (startup, braking, idle)
   6.  Packaging
   7.  Weight
   8.  Complexity
   9.  Cost
  10.  Required development
  11.  Regenerative power aspects.

The same combination of engineering experience was called upon to select
for evaluation only those components considered practical for such a trans-
mission.  It should be recognized that there are probably areas of research
or development of torque transmitting components unknown to the group which
were not considered.   Unpublished research accomplished by private industry
would be the most likely area of omission.  The specification that the trans-
mission was to be built within the year established the first evaluation
factor — availability of components.   Therefore, it was decided that torque
                                     A-2

-------
transmitting components in the research or initial development phase would not
be used and any such omission would not adversely affect the conclusions reached
for the evaluation and selection task.

For the initial evaluations, preliminary comparisons of these factors for the
transmission under consideration to that of an existing transmission were made.
A more detailed treatment was reserved for the selected candidate design con-
cept.

Power Transmitting Devices

By contract definition the transmission must transmit engine power to the road
for all steady-state requirements and must also feed in power from a variable
high-speed flywheel to a variable low-speed output shaft and provide vehicle
acceleration.

The definition of possible transmissions which then could be evaluated using
the factors previously listed required the initial selection of torque trans-
mitting apparatus for each function.

These devices were separated into two categories — hydraulic and mechanical.

Under the hydraulic category the following elements were studied:

   1.  Hydrostatic or positive displacement hydraulic pumps and motors.
   2.  Hydrokinetic fluid drive or fluid coupling.
   3.  Hydrodynamic coupling or torque converter.
   4.  Hydroviscous shear drive or slipping clutch.

All of these did offer the following desired characteristics:

   1.  Variable speed ratio
   2.  Controllable output torque
   3.  Developed technology
   4.  Existing applications
   5.  Ample horsepower capacity
                                     A-3

-------
Many drive applications using hydraulic devices were reviewed.  Some of the more
relevant are summarized in Table A-l.

Under the mechanical category the following elements were studied:

   1.  Traction drives
   2.  Cam or gear drives
   3.  Belt drives

Several of the mechanical devices reviewed are summarized in Table A-2.

Candidate Transmissions

In addition to the concepts generated by combining the most promising components,
several concepts from other studies were reviewed.  Although not specifically
included in our screening chart, they merit discussion.  One of these is a
flywheel/transmission for automobiles as proposed by R. C. Clerk.  His paper,
The Utilization of Flywheel Energy, dated June 1963, discussed the concept.
The flywheel was controlled to follow the speed of the engine resulting in a
low speed ratio and a low percentage use (approximately 20 percent) of the fly-
wheel energy.  Thus, the flywheel had to be quite large and heavy to provide
the desired acceleration torque.  In addition, a variable fluid coupling was
used as the flywheel connecting link which would have a low efficiency at high
speed ratios.  Although heavy, the arrangement did prove the feasibility of
such a hybrid/flywheel propulsion system.

Another concept locating the transmission as part of the rear axle could make
use of the rear axle differential to lead the power from the engine and flywheel
to the rear wheels.  Such transmission arrangements were discussed in detail
in Mr. G. M. DeLalio's patent, Number 3,212,358, dated October 1965.

The heavy rear axle and transmission assembly would mandate that the assembly be
fixed and that the rear wheels be independently suspended.  This is a major
change to American car design.
                                      A-4

-------
                                TABLE  A-l
         Description
            Comments
1.  Variable Hydraulic Drive

The device uses a planetary gear
train as positive displacement
hydraulic pump and a valve to
control the amount of oil or air
in the gear train and the flow
of oil out of the gear train.
Torque and speed are controllable,
 2.  Internal Gear Torque Motor
An internal gear pump and motor
and a bypass control valve are
used to provide torque multi-
plication.  The element is used
in a small tractor transmission.
The feature of bypassing oil as
a control means results in a low
efficiency at speeds other than
the straight through or lock-up
speed.  Thus the unit is limited
to low horsepower applications
(under 50 hp) which do non neces-
sarily require high efficiency.
The overall efficiency is low
and the unit is limited to below
50 horsepower.
3.  Vane, Axial Piston & Radial
	Piston Pump and Motor	

These drives use positive displace-
ment hydraulic pumps and motors
to provide a variable output speed
and torque.  The unit transmits
all of the power through the hy-
draulic pump and motor.

4.  Clutch/Brake Drive

These drives transmit torque and
speed by viscous shear of a thin
oil film between the discs.  The
slipping-clutch—type drive has
been developed and a long life can
be expected.

5.  Fluid Couplings as Manufactured
	for the Automotive Companies
The SAE Paper 359B, 1961, Fluid Coup-
lings, by J. Qualman and E. Egbert,
General Motors Corporation, was used
to investigate the capabilities of
these elements.
Has an lower overall efficiency
than the power splitting type
drive.  The unit manufactured
for industrial applications is
too bulky for automotive appli-
cation .
The disadvantage of a low effi-
ciency at all part-load conditions
has not been overcome.
These elements do provide a versa-
tile variable drive coupling neces-
sary for the transmission;  however,
the hydrokinetic action used to
transmit torque and speed results
in a very low efficiency at part
load conditions where the slip of
the input of the coupling relative
to the output of the coupling is
high.
                                  A-5

-------
                          TABLE A-l (Continued)
         Description
            Comments
6.  Torque Converters as Manu-
    factured for the Automotive
	Companies	

The SAE papers 359A, 1961, Appli-
cation of Hydrodynamic Drive Units
to Passenger Car Automatic Trans-
missions, by E. Upton, G.M. Corp.,
and 359C, 1961, Multiturbine
Torque Converters, by F. Walker,
G.M. Corp.
The torque converter offers the
advantage of a simple, automatic
and reliable method of changing
torque ratios or speed ratios.
The converter has been manufactured
in single, two and three stage
configurations which provide a
wide range of torque multiplica-
tion.  Development models of a
single-stage converter with a
variable reactor have been tested
successfully.  The variable re-
actor provides a way to match the
speed characteristics of a parti-
cular engine to the requirements
of a particular vehicle.

The efficiency remains high over
much of the driving requirements.
However, as the torque ratio
increases and the output speed
decreases the efficiency falls
off rapidly.
                                   A-6

-------
                               TABLE A-2
         Description
             Comments
1.  Several companies  have  been
    engaged in programs  that  use
    traction drives as transmis-
    sions.  These drives are  in
    the form of planetary friction
    drives which make  use of  a
    tilting traction roller or ad-
    justable disc to adjust ratio.
There appears to be no horsepower
or speed limit and the noise
level is low while the efficiency
is high.  The operational technology
and development of the drive is en-
couraging and indicates the possibil-
ities of many applications within the
next few years.  However, the develop-
ment status of the drive eliminated
its application for the system under
immediate consideration.
    Many versions of the cam,  rol-
    ler or gear type drive have
    been produced or proposed.
    These devices are in produc-
    tion as high ratio speed re-
    ducers .
Introducing a ratio controlling
element would place these drives
in  the development category only
to  be considered for transmissions
of  the future.
    The Maroth drive proposed by
    Tranco, Inc., is a high ratio,
    in-line drive that uses tapered
    rollers and a nutating motion
    to transmit torque.
4.  Many companies have built vari-
    able speed transmissions us-
    ing belts and chains.
A small hydraulic pump and motor
could be used to obtain varying
torque and speed ratio.  The
drive offered low weight, high
efficiency, low noise and low
vulnerability as possible advan-
tages.  However, the development
status of the drive again placed
it in the category of transmissions
in the future.

To date the package size needed to
transmit the horsepower required
for an automobile is larger than
the installation space if composi-
tion belts or chains are used.
When using metal belts or chains
the problems of preventing metal
fatigue and preventing or remov-
ing the heat has limited the
amount of power than can be hand-
led within a given package size.
Transmissions using these elements
require both technology and opera-
tional development.
                                   A-7

-------
In summary, the R. C. Clerk concept, transaxle concept and several others of a
similar nature were abandoned because of the attending design and structure
changes to the car.

Those power transmitting devices that were selected as applicable were arranged
as transmission configurations:

   I..  Figures A-l and A-2 (package drawing SK-J-4258) present a concept that
       uses three torque converter automotive-type transmissions.  Transmission 1
       runs the engine at the optimum point for low emissions; Transmission 2
       provides speed flexibility for the flywheel; and Transmission 3 provides
       the correct output speed and torque.

   2.  Figures A-3 and A-4 (package drawing SK-E-4260) show a concept that uses
       two torque converter automotive-type transmissions.  Similar functions
       to that described in the first transmission are performed by the trans-
       mission.  The engine will ope rate off the minimum specific fuel consump-
       tion line since the engine speed range is fixed by Transmission 1 which
       does not include a torque converter.

   3.  Figure A-5 is a concept that uses two variable fluid couplings to provide
       the variable speed ratio.  A valve is used to control the speed ratio.
       Development of a valve to control speed and torque without introducing
       a  large efficiency penalty is not,  to our knowledge, being explored by
       any of today's transmission manufacturers.

   4.  Figures A-6 and A-7 (package drawing SK-D-4259) are a concept that uses
       a  torque converter automotive-type transmission and a hydrostatic trans-
       mission to provide control for the flywheel.

   5.  Figures A-8 and A-9 (package drawing SK-J-4257) give a concept that uses
       three  hydrostatic units for primary torque and speed control  and three
       hydrostatic units for flywheel torque and speed control.   The arrange-
       ment is essentially two separate transmissions which would ease the
                                     A-8

-------
                          Scheme A
        Tran #1
        (Engine)
\
Trans #3
(Main)
Offset Gears
   Trans #2
   (Flywheel)
Wlieol s
Fig. A-l   Automotive 3-Element Hydrodynamic (Ref. SK-J-4258)
                              A-9

-------
INPUT DRIVE
(FROM ENGINE)
                                                                                         4'       8'       I'
                                                                                           SCALE
                            Fig. A-2    Three-Element Automotive  Transmission  (SK-J-4258)
                                                                                                                       MTI-12518

-------
                                    Scheme A'
•—Engine
                Trans #1
                .(Engine)
                                                Trans #2
                                                (Flywheel)
Offset Gears
                                                                   Wheels
                                                             Flywheel
            Fig. A-3   Automotive 2-Element Hydrodynamic (Ref. SK-E-4260)

-------
                                                           SCALE
Fig. A-4   Two-Element  Automotive Transmission (SK-E-4260)
                                                                                    MTI-12517

-------
                                        -FLWHEEL AND IDLER SHAFTS -  15  OFF
                                         VERTICAL CENTERLISE (ABOVE HORIZONTAL)
                                                                                      25'
75'
                                                                                                                           OUTPUT
                                                                                                                           SHAFT
INPUT SHAFT
(FROM HEAT ENGINE)
                                                                                         SCALE
                           Fig.  A-5    Transmission Automotive-2  Coupling Hydrodynamic
                                        Heat  Engine/Flywheel  (SK-A-4262)
                                                                                                                           MTI-12523

-------
                                      HYDRODYNAMIC TRANSMISSION
>
i
\
1
1




1
1
N







k
i'




ii

.

H


.j
•

1
_
Lq
"


•
r r
•
r


»



•




Output to
Rear Axle



                                                               FLYWHEEL

                                                               TRANSMISSION
                          Fig.  A-6   Automotive  Hydrodynamic/Hydrostatic  (Ref.  SK-D-4259)

-------
          HYDROSTATIC DRIVE
                                                       TOP VIEW
                                                                           2.5"
5"
7.5"
Fig. A-7   Hydrodynamic/Hydrostatic Automotive  Transmission
            (SK-D-4259)
                                                                              SCALE
                                                                                         MTI-12498

-------
                                              PRIMARY
                                            TRANSMISSION
Engine
Input
                      V.D.
                      7.5  in
                       V.D.
                       7.0 in
                                      V.D.   .
                                      7.0  in'
F.D.
7.0 in
                                                                                   Output  to  Rear  Axle
-T- tt/ff
V.D.
6.0 in

F.D.
6.0 in



P
\
-A.
•
«i

J
FLYWHEEL
TRANSMISSION
•i
•
•
•
•
n^
^
^
^
»
w
i^
\\x\\





                         Fig.  A-8   Hydrostatic  Transmission  CRef.  SK-J-4257)

-------
I
t—'
^J
                                                Fig.  A-9(a)    SK-J-4257 (Sheet  1)
                                                                                            SCALE
                                                                                                          9"
                                                                                                                             MTI-12525

-------
I*
I-*
00
                                                   SECTION THROUGH


                                                   HORIZONTAL CENTERLINE
                                                         Fig.  A-9(b)    SK-J-4257 (Sheet 2)
                                                                                                                             7J5T
                                                                                                              SCALE
                                                                                                                                                MTI-12526

-------
       problem of controlling the flywheel acceleration torque.   However, the
       two transmissions would result in low efficiency.

   6.  Figures A-10, A-ll and Figures A-12, A-13 and A-14 (package drawings
       SK-E-4283, SK-E-4261, and SK-J-4301) are concepts  that use hydrostatic
       units and planetary gear trains to control speed and torque from the
       engine and flywheel to the output.  The three concepts vary only in the
       location of the flywheel.

Screening of Transmission Candidates

Seven different transmissions were evaluated using the screening factors pre-
viously discussed.  The results are shown in Table A-3.  As preparation for
this final table several additional tables were generated.   In all of these
tables data for the standard-multi-speed automatic torque converter transmission
were; included to provide a present-day base for comparisons.  This was the
preliminary .screening process previously described.

Tables A-4 and A-5 present the transmission's physical characteristics of woight,
volume and number of major components.  These numbers were  obtained by using
the package drawings generated for the candidates.

As expected, the existing standard transmission is lower  in weight, volume and
number of components.

A comparison of cost in terms of a ratio of the estimated cost of the new trans-
mission versus the standard automatic transmission is given in Table A-6.  Cost
ratios as used and discussed in the final power-splitting transmission cost
analysis were used.  Mr. E. Charles, using the drawings generated, provided
the cost information.

An estimate of the efficiencies as determined by following  the power paths
for varying output load and speed, assigning an efficiency  for each path and
multiplying those numbers to obtain an overall efficiency are given in Table A-7.
                                     A-19

-------
ENGINE
INPUT
                                      PRIMARY TRANSMISSION







_L
f
1
Tl
T


JL





















^^

V




B








VD













VD













i
i
*





                                    VD
VD
                          FLYWHEEL TRANSMISSION
                                                           1
                                                                         ±    I
                                                                              f
T    T


    mf
                                          OUTPUT
                                          TO REAR
                                          AXLE
                                Fig. A-10  Power- Splitting Transmission

-------
                                                                       PRIMARY TRANSMISSION
i
K)
                                                                       FLYWHEEL TRANSMISSION
                                                                                                                   Output

                                                                                                                   to Rear

                                                                                                                   Axle
                               Fig.  A-ll    Power-Splitting-Offset  Flywheel  (Ref. SK-E-4261)

-------
                                                                     .^^W^g)(^l®®®l®®^l®@^@@@^^^^^



                         ;   '  . ^     •    i       \ •     \    '    ;     -; '/
                               '      /        .    ,•'.,,• i  ,•  i
1 FLYWHEEL  I TRANSMISSION
                                          HORIZONTAL SECTION
                                                                                 3"
                                                                                    SCALE
                                  Fig.  A-12(a)    SK-E-4283  (Sheet  1)
                                                                                                                MTI-12519

-------
                                                            GRAVITY
>
u;
                                                                                                             SCALE
                                                 Fig. A-12(b)   SK-E-4283  (Sheet 2)
                                                                                                                             MTI-12521


-------
                                                          -CASE, SHOWING TRUNNION BEARING SITPORT AM)
                                                           BREAK-AWAY SHOWING CASE INTERIOR MINI'S INTER-SAL PASTS
                                                ACTUATOR (2 REQ'D)
                                          CAM PLATE GUIDE
                                          BRONZE     fcj
GRAVITY
CAM PLATE - STEEL
CAM TRACKS HT. TR. TO Re 50 MIN.

 SECONDARY TRANSMISSION
 PRIMARY TRANSMISSION
                                          Fig.  A-12(c)     SK-E-4283  (Sheet  3)
                                                                                                           SCALE
                                                                                                                                         KII-12496

-------
          IXPIT SHAFT
          (.FRO- HEAT ENGINE)

>
ro
                                                                 -^^—
                                                   -FLWHEEL AW) IDLER SHAFTS - 15  OFF
                                                   VERTICAL CESTERLI-SE (ABOVE HORIZONTAL)
                                                                                                25'
75'
                                                                                                    SCALE
                                     Fig. A-13    Power-Splitting  - Offset  Flywheel Transmission
                                                   (SK-E-4261)
                                                                                                                                      KII-12530

-------
NJ
             ENGINE INPIT
                             Fig. A-14   Vertical Flywheel Axis Installation Concept (SK-J-4301)
                                                                                                                8'
                                                                                                          SCALE
                                                                                                                         KTI-12524

-------
                                TABLE  A-3   NORMALIZED SCREENING FOR TRANSMISSION CONCEPTS



SK-J-4258
3 Element
Hydrodynamic
SK-E-4260
2 Element
Hydrodynamic
SK-J-4257
Hydrostatic
SK-E-4256
Power Split
SK-E-4261
Power Split
(Offset)
SK-D-4259
Hydrodynamic
Hydrostatic
SK-A-4262
Hydrodynamic
2 Coupling
Std-Multi Speed
Automatic
Transmission
w/o Flywheel


Cost
1.266


1.0


1.4

1.266

1.333


1.266


1.133


.67





Weight
1.146


1.0


1.265

1.142

1.173


1.064


1.059


.46





SFC
1.3


1.4


1.0

1.0

1.0


1.3


1.4


1.3





Size
1.21


1.00


1.46

1.35

1.46


1.13


1.02


.96




No.
Parts
1.^4


1.23


1.00

1.27

1.44


1.41


1.30


.93





Eff .*
1.39


1.07


1.26

1.00

1.00


1.13


1.38


1.03




Overall
Rating
7.757


6.700


7.385

7.028

7.406


7.300


7.292


Rating
Cost-SFC
Eff.
3.961


3.470


3.66

3.266

3.333


3.696


3.913


Note: Comparative Information


Notes
(a)


(a)


(M

(b)

(b)





(a)


for
Standard Transmission is Shown
for Reference Purposes Only.

I
IV)
                                       *
                                     (a)
                                     (b)
The normalized  efficiency is obtained using the weighted  efficiency  data.
Component  development is required.
Optimum availability of engine power.

-------
                              TABLE A-4.

                              ESTIMATED

                  TRANSMISSION WEIGHTS AND VOLUMES

(Weights are dry weights and include flywheel, controls, and housing)
                                     Wt.                  Vol.
                                    fibs)               (ct.ft.)
SK-E-4283
Power Split                          337                  2.80

SK-E-4261
Power Split
Offset Flywheel                      347                  3.05

SK-J4257
Hydrostatic                          374                  3.04

SK-J-4258
Automotive
3-Element
Hydrodynamic                         339                  2.52

SK-E-4260
Automotive
2-Element                            295                  2.08

SK-D-4259
Automotive
Hydrodynamic/
Hydrostatic                          314                  2.37

SK-A-4262
Automotive
Hydrodynamic
2-Couplings                          313                  2.12

STD Transmission
Without Flywheel or
Flywheel Transmission                140                  2.0
                                   A-28

-------
Number
                                               n.-.-jone: . ts  in  Transmission*
1
: to
I VI CO
; : oo i c
C ! -i-i
•H i .— I •
i 0 —ID.
; -i-i i Q. D
i *-*
CU CO CO







Power Splitting
SK-E-4283
i
, Auto-3 Element Hydrodynamic
SK-J-4258
Auto-2 Element Hydrodynamic
SK-E 4260
Hydrostatic
SK-J-4257
i— 1 4-1 4-1
J3 CO C
co o cu
•H H E
I-i 13 CU
CO ^% »"H
> X W
4


-

_

4

Auto Hydrodynamic /Hydrostatic 1
SK-D-4259 ,
3 0
0 C_>
o
cu
T) . 3
•H a-
3
t— i
fc
_


_

.

_
Xj
O
H
.


3

2

mf

_

Auto Hydrodynamic 2-Coupling - 2
!

„
SK-A-4262 :
Power Split-Offset Flywheel i 4 ' •-
SK-E-426I
Standard Transmission

— — : 1

00 CO 13 in
C CU >-i O
•^ J= CO J=
Q. O ! T3 0
Q, iJ C 4-1
•1-13 co 3
i— 1 i— 1 4-1 f— 1
CO O CO O
2
i

4 3
i
4 2
I
— . '?

: 2


oo
c
•r-l
C w
C CU
3 J=
J-i O
I-I 4-1
4) 3
> •-*•
00
_


2

2

_

3

22 2

2 . -

2 , 2

2



^j
CO
cu
o

^
I-I
CD
4-1 CO
0) CU
c >
CO -i-l
^ »-l

3


4

3

1

3

4

3

2












CO
}-l
CO
cu
o
10


_

2

9

4

_

12

-











(O
4J
y-i
CO
.c
CO
8


8

7

8

9

9

10

5

CO
oo
c
•H
M
CO
cu
PQ

0)
^
CU
CU
,_(
CO
10


12

11

2

12

15

11

9


4-1
C
CU
E
ID
i— 1
W
co
00 OO
C C
•H -t-l
— i M
.-1 CO
O 
-------
                               TABLE A-6
              Preliminary Transmission Cost Analysis - Ratios
                       1,000,000 Units Per Year
                                 Cost
                 1,000,000 Units Per Year
                           Cost
SK-J-4258
3 Element
Hydrodynamic

SK-E-4260
2 Element
Hydrodynamic

SK-J-4257
Hydrostatic

SK-E-4283
Power Split

SK-E-4261
Power Split
(Offset)

SK-D-4259
Hydrodynamic
Hydrostatic

SK-A-4262
Hydrodynamic
2 Coupling

Standard
Multi -Speed
Torque Converter
("Automatic")
1.9
1.5
2.1
1.9
2.0
1.9
1.7
2.45



1.94



2.71


2.48


2.58



2.48



2.19
                           1.29
                                  A-30

-------
                                      TABLE A-7

                          ESTIMATED TRANSMISSION  EFFICIENCIES



Power Splitting
Hydrosta tic
SK-J-4257
Auto - 3 Element
Hydrodynamic
SK-J-4258
Auto - 2 Element
Hydrodynamic
SK-E-4260
Hydrodynamic/
Hydrostatic
Au to -Hydrodynamic
2 Couplings
Std-Multi Speed
Automatic Trans-
mission without
Flywheel


20 MPH
75-70%
60-55%
50-45%

70-65%

68-63%
51-46%
73-68%




50 MPH
92-87%
73-68%
80-75%

85-80%

78-73%
80-75%
89-84%




80 MPH
84-79%
71-66%
85-80%

90-85%

83-78%
82-77%
93-88%




WEIGHTED*
76.8%
60.8%
55%

71.3%

67.7%
55.8%
74 . 5%


50 MPH
PLUS FLYWHEEL
ACCELERATION
80-75%
73-78%
72-67%

75-70%

74-69%
69-64%
	


* The weighted efficiency combines  the  efficiencies  of  the  20  MPH  and  50  MPH cases
  for the amount  of time spent  at  each  of  these  speeds  during  the  driving cycle.

-------
 The  efficiency used for the screening process was the weighted efficiency.
 This  efficiency was normalized using the 20 MPH and 50 MPH estimates and was
 weighted  to account for the amount of time spent at each of these speeds for
 a driving cycle by using the Velocity Histogram of the DHEW Urban Dynamometer
 Driving Schedule dated 8-5-71.

 The  factors from these tables and several additional factors were combined
 into  rating Table A-3.  The specific fuel concumption for each candidate was
 an estimate made by Mr. G. DeLalio.  His estimates were based on the power
 paths through the transmissions and whether the engine could or could not run
 on the minimum SFC line.

 Table A-3 which summarizes the screening performed on the transmission concepts
 and their basic comparison to the standard automatic transmission would basic-
 ally  indicate that the two-element hydrodynamic transmission would be the primary
 overall selection.  (The overall rating of Table A-3 was simply the addition of
 all of the weighting factors).  The second choice and close contender would be
 the power-splitting transmission with the inline flywheel.  Accepting this over-
 all rating at face value and introducing one of the major criteria with respect
 to the urgency of implementation led us to reject the two-element hydrodynamic
 unit on the basis of the valve development required.  It should be recognized,
 however, that considerable optimism regarding the torque converter efficiency
 is included in the subject rating.  With this additional constraint, however,
 it was clear that the power-splitting transmission because of its overall
 comparative rating as well as the level of development of all the components
 made is the primary selection.  Referring again to Table A-3,  if one were to
 select three criteria which would cover the objective of reduced pollutants
 and acceptance by the public of a new concept, we concluded that the elements
 of cost, specific fuel consumption and efficiency fulfilled this requirement.
On this basis, the power-splitting transmission was clearly the favored selec-
 tion (Figures A-10 and A-12 — drawing SK-E-4283).

 The selected power-splitting transmission has four :hydraulic elements as the
major torque and;speed control elements.  These-represent^approximately 50 per-
cent of the transmission weight and 50 percent of the volume.   The successful
                                      A-32

-------
application of the transmission depends to a great extent on the proper selection
of the hydraulic elements.  Therefore, prior to finalizing this concept, a study
was made to select the best type of pump and motor.  The results of the study are
presenLed in Table A-8.  All of the different pumps and motors are currently
inanii l;acLurt.'cl by several companies.

The conclusion drawn from this study and summarized on Table A-8 is that the
axial piston pump and motor was the correct selection for the transmission.

The limited installation space and transmission weight required that the elements
be as small as possible.  Therefore, a high density hydraulic element was con-
sidered an advantage.  A drawing of such an element is shown on Figure A-15
(drawing SK-C-4266) and a picture on Figure A-16.  The unit is packaged very
compactly to achieve maximum power per cubic inch of volume.  The pump or motor
designated for the MTI power-splitting transmission produces approximately 2
horsepower per cubic inch of volume.

A relative comparison of the operation of a three-speed manual, a three-speed
torque converter, and a power-splitting transmission can be shown using a plot
of the amount of engine power available at the rear axle versus the vehicle
speed (see sketch on page A-37).

The manual transmission is capable of delivering the maximum amount of engine
input power to the rear wheels.  Therefore, it could be concluded that a smaller
engine, less fuel and lower emissions are possible when using a manual transmis-
sion.  This is true if there are a sufficient number of gear ratio steps, approx-
imately 6; but, this is not considered reasonable based upon present driving
habits of vehicle operators.

As is shown by the curve the percent of output power to the rear wheels falls off
drastically when the vehicle speed does not match the optimum point of operation
for the gear ratio being used.   In fact it falls below that of both the torque
converter and power-splitting transmission.

One additional comparison should be discussed which is how these transmissions
permit the engine to operate with respect to the minimum SFC line.

                                     A-33

-------
                                                      TABLE A-8
                                          Hydraulic Pump  and Motor Rating

Axial
Piston
Vane
Gear
Gerotor
-Radial.
j Piston
;
Previous Use
For Such An
Application

1
3
3
3

2
Reliability

1
3
2
2

1
Efficiency

1
3
2
2

1
High
Pressure
Capability

1
3
3
2

1
Modulation
Capability

1
1
2
1

1
Control
Adaptability

1
2
3
2

2
Ranee

1
2
2
2

1
I
OJ

-------
u>
Ln
                                                                                                  1.5"       3.0"      45"
                                                                                                    SCALE
                            Fig.  A-15   Drawing  of High-Density  Hydraulic  Element (SK-C-4266)
                                                                                                                       MTI-12520

-------
Fig. A-16   Typical Axial Piston Hydraulic  Element
                          A-36
                                                                MTI-10560

-------
             100
              75
        c
        -l
        
-------
The power-splitting transmission with the infinitely variable ratio i§ a
that will provide the maximum power to the wheeli while providing maximum §v@r§lv
engine efficiency by operating on the minimum 8FC line ler a greater p§re@nea§§ of
the time.
                                     A-38

-------
              SECTION B

DETAILED DESCRIPTION OF TRANSMISSION
                                                                     CD
                                                                     K
                                                                     o

                                                                     O
                                                                     S
                                                                     n
                                                                     ^
                                                                     •o'
                                                                     «-»
                                                                     5'

-------
                     B.   DETAILED DESCRIPTION OF.TRANSMISSION

Tlit operation and mechanical description of the power-splitting transmission as
proposed by MTI is covered in this section.

Figure A-12, shown previously, depicts the parts of the transmission in suffi-
cient detail to perform an accurate cost analysis.  Figure B-l is a simplified
schematic of the transmission which was used for the performance analysis.
Shown in Figure B-2 is a preliminary plot of the operational parameters of the
transmission when the engine is operated at maximum speed and power.  A plot of
the displacements of the hydraulic elements which control the flywheel accelera-
tion and deceleration is presented in Figure B-3.  Figures B-4, B-5, B-6 and B-7
are schematics of the power path through the transmission when operating in low
range, high range, flywheel accelerating the vehicle in low range, and flywheel
accelerating the vehicle in high range.

The MTI power-splitting transmission is an infinitely variable, stepless unit
that obtains torque multiplication and control by means of hydraulic principles.
It is intended for use in a medium-size automobile.

The unit differs from the torque converter or fluid coupling hydrodynamic-type
transmissions in that the power in the hydraulic circuit is transferred by fluid
static pressure at low flows, whereas the hydrodynamic unit utilizes high flows
and the dynamic or inertial motion of the fluid to transfer power.  Further, the
power-splitting unit is a "hard" type drive in that the transmission slip is less
than two percent under full load.

Basically the unit consists of engine input shaft, flywheel, flywheel planetary
gear train, hydraulic variable displacement elements, connecting drive gears,
output planetary gear train, output gear train and a control system.  The control
system is described in the control section.  The transmission description may be
understood by reference to Figures A-12 and B-2.
                                     B-l

-------
Planetary "A"
Primary Transsi sslcr.
 /         \
                                                            Svashplate
                                  \
                          \\\\\\\v
Engine 	
Tnput
1200 RPM
to
3750 RPM








b.5 _[
2.75
! n



=
=

Flywheel
24,000 RPM
10,000 RFM
-



\
/]



1.00
1.00


\
\
1.27
1.00
1 \
Variable Displacement
+ 3200 RPM
~ 3200 PSI
7.5 In3/rev.
ELEMENT 1

\
\

Variable Displacement
+ 3400 RPM
3400 PSI
6.0 InVrev.
ELEMENT III
1
L

\
Variable Displacement
+ 32OO RPM
X2OO PSI
7.5 tn3/rev.
ELEMENT II
,1


/ »


Variable Displacement
f 3400 RPM
340O PSI
6.0 in3/rev.
ELEMENT IV
\ . . /
Secondary Transmission


1.00
1.10
/
/
1.00









1.62
1.00
Notes: 1. Rear Axle Diffe
2. Rear Wheel Diaz
3. Vehicle Speed E
at Transmissior
of 4470 rpm.
3.48
1.24
1L
1.00

III
High
t
t
s_

Range
recer = 2.1 £
qua Is 9A,- •=-
Output Spee
	 5

r
Clu
)h
Planetary "C"
\\\\\ /
E /
2.50 /
.75 /
1 (QQ Transml aslon
0 RPM
to
= 4470 RPM
1 I I I* — I.OM Range Brake
l/l/l//
tch
3.55:1
Fig.  B-l    Schematic  of the  Recommended Transmission  Design
                                                                          MTI-12341

-------
     + 100
c
o
o
D.
CO
o
H
50
      -50
     -100
                                  8          12          16


                                       N Flywheel/N Engine
                                                              20
                              Fig.  B-2   Secondary Hydraulic System

-------
Fig. B-3  A Plot of Hydraulic Element Displacements

-------
CO

Ln
   ENGINE
   INPUT
                                       PRIMARY TRANSMISSION
                 PLANETARY "A"
                                                       SWASHPLATE
              FLYWHEEL
                                t
VD
in
VD
                           FLYWHEEL TRANSMISSION  HIGH RANGE CLUTCH
                                                                       PLANETARY "B"




















.


_L
T

1
T~l
1


r^ll
Tl 1
1U
/







r
i
•
i








» «







P
r

•
i
=i
i


\
i
ii



mm> *^










mm





\ir\
vu
I














L1





\in
v u
n



• «• ^MB M^ •

1

Ir
/
/
/!



i^ mi
1









1

ri
•


























j
r ]

.



1


i
j
L
[

-



L
LOV

^
•
!
i
j

*
i
j
:
;

^
jr
\





P
                                                                            LO\V RANGE BRAKE
                                          Tl_,
                                        OUTPUT
                                        TO REAR
                                        AXLE
                                                                                PLANETARY  C'
                                         TORQUE  PATH
                                         TORQUE REACTION
                            Fig. B-4   Power Path for Low-Range Steady State
                                                                                       HTl-12538

-------
                                      PRIMARY TRANSMISSION
ENGINE
INPUT
              PLANETARY "A"
                                                       SWASHPLATE
                Tl-l
                                    in
	
\yn
VU
n



h*™ ^^m ^^"*
1


^B





•••

1
1
/
/:



•~1
— •
•|.

a








••







™
1
™

IT j
*—* -i
•j
i
1 	 1
~i •
1 4
LOV
\l/2

t J|

"I
p -1

J
i n
                                                                             LOW RANGE BRAKE
12:
                                          OUTPUT
                                          TO REAR
                                          AXLE
                                                                                    ANETARY  C
                          FLYWHEEL TRANSMISSION^1   HIGH RANGE CLUTCH
                                                                           PLANETARY  "B"
                               1 TORQUE PATH
                                TORQUE REACTION
                            Fig. B-5   Power Path for High-Range Steady State
                                                                                           MTI-12511;

-------
CD
I
ENGINE
INPUT
                                        PRIMARY TRANSMISSION
                 PLANETARY "A"
                                                       SWASHPLATE
                                     VD
                                      I
                                              VD
                                              n
                                	,  "I	,      ',     ,—,  =f    £
                                L     LC     iM    P4  r*\
                                ___.•»___.!».1|±»     J_      -L    I
                                   VD        VD   _/.__	I      1Lj^§rP
                                                 VD
                                                 is:
                            FLYWHEEL TRANSMISSION^  HIGH RANGE CLUTCH
                                                                        PLANETARY "B"
                                                                            LOW RANGE BRAKE
I      I
                                                                                         OUTPUT
                                                                                         TO REA^
                                                                                         AXLE
                                                                                 PLANETARY  C
                                                                           TORQUE PATH
                                                                           TORQUE REACTION
                             Fig.  B-6  Power Path for Low-Range Acceleration
                                                                                         MTI-12546

-------
                                         PRIMARY TRANSMISSION
»  ENGINE
00  INPUT
                 PLANETARY "A"
                                                         SWASHPLATE
              FLYWHEEL




f

-T^
T
1
T


n
11




^
r
|
•
1




• ••
• «

r
i
i
1
rll
r|

i
•
rl J
1
T
/




»



• •


VD
I




^m



11


VD
n


» ^^ a^» ^BB

^



^B
1
"Y "
/i


^«
K
                                                                               LOW RANGE BRAKE
                                      nr
12:
                                          OUTPUT
                                          TO REAR
                                          AXLE
                                                                                      ANETARY "c"
                            FLYWHEEL TRANSMISSIO^Pl  HIGH RANGE CLUTCH
                                                                             PLANETARY  "B"
                               TORQUE PATH
                               TORQUE REACTION
                            Fig. B-7  Power Path for High-Range Acceleration

-------
Component Description

The flywheel is located in line with the transmission and in the cavity
normally occupied by the torque converter of a standard automatic transmission.
It was believed that the location would best meet the specification that the
flywheel and transmission be capable of Bitting into an existing medium-size
car with no major structural or design changes to the car.  However, thp.
location and the design of the flywheelare to be decided on a system basis.

Three planetary gear trains are used in the assembly:  1) to provide a power
path for the flywheel; 2) to direct the output power when the vehicle is in the
high ratio range; and 3) to provide a low-ratio range power path.  These gears
and the gears that connect the hydraulic elements to the power shafts are manu-
factured using automotive practices.

The gears are made of automotive gear materials such as forged and surface-hard-
ened AISI 8620 steel and are machibed to an AGMA gear quality level of 8 with a
final polish.

The four variable hydraulic elements are based upon a design originated by
Dr. H.  Ebert* and recommended by Mr. G. DeLalio.  They are considered a high
power density construction which produces the maximum capacity using the minimum
volume and weight.  Typical power densities are approximately 2 horsepower/
cubic inch.   The construction of the unit is described by Figure A-15 (earlier).
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 Palmer-Benz,  Austin,  NSV and Allgaier
     in Germany.
**   The design of the swashplate bearing was established by many hours  of
     testing in Germany and at the Stratos Division of the Fairchild
     Corporation.
                                   B-9

-------
 This  type  of hydraulic element has been  successfully used in selected applica-
 tions  for  aircraft constant speed drives,  tractor  transmissions,  truck  trans-
 missions 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
         evaluated 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 7.5 in /rev size element (primary trans-
mission), the system should operate at pressure 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 3200 psi and speeds of 3200 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, con-
trol 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.
                                    B-10

-------
The high-range clutch and low-range brake are of conventional automotive con-
struction used in existing transmission.-,.

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 element.  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 hard-
ened cam tracks.  A cam roller bearing is used to link the cam to the trunnion.

The automotive practice of using sleeve bearings or needle bearings to support
the radial and thrust loads has been followed.

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 mechanical description of the controls is covered in the control section.

The necessary control by the driver is to be established by the appropriate posi-
tioning of a single lever.  The lever is so located to be hand-operated from the
driver's position much the same as the control of present automobiles.

Operation

The general operation may be followed by reference to Figures B-l, B-2 and B-3.
The arrangement combines the primary and secondary transmission:  functions into
a single integrated unit.

Figure B-l is a schematic of the working elements of the transmission.  The input
and output speed, the data for the hydraulic elements, and the relative ratio of
the gears are shown.  Figures B-2 and B-3 are plots of the operational character-
istics of the transmission.  Figure B-2 is plotted for an engine speed of 3750
RPM.

The number of mechanical elements has been minimized to emphasize the simplicity
of such a transmission, maintain a high reliability and keep the cost as low as
possible.  Additional gear trains and sketches could have been added to increase
                                     B-ll

-------
 the  capabilities  of  the  transmission or  reduce  the  amount  of  power  in  the  hydro-•
 static  portion of  the transmission; however,  the  time  schedule  from design  to
 production mandated  the  straightforward  approach.

 Referring to the primary section of the  transmission,  the  hydrostatic  power
 circuit consists of  two  identical positive displacement units which  are referred
 to as Unit I and Unit II.  These are hydraulically-connected by porting in  the
 common valve plate member.  Both units are variable displacement and the dis-
 placement is regulated by controlling the swashplate position, which is
 generally indicated  by a heavy line on Figure B-l.  When the swashplate is
 parallel to the drum face, the respective unit  is at zero  displacement and when
 the swashplate is at 15° angle, the unit is at  full displacement.  The engine
 drives Unit I through a 1:1.27 gear reduction.  As shown on Figure B-2, Element
 I is varied to full  displacement to drive Element II through zero speed to full
 output speed, at which time Element II is varied  to achieve the overdrive
 speed ratio of 1:1.1.  Unit II through a step-up gear ratio of 1.57:1.0 drives
 the sun gears of planetary "B" and "C".  The low-range steady-state power path
 through the transmission is shown on Figure B-4.  The heavy solid lines
 represent the flow of torque through the transmission while the heavy dotted
 lines represent the  torque reaction when needed.

 The output torque is a function of the hydraulic pressure and the displacement
 of the hydrostatic units.  If the torque load increases the displacement or
 pressure of the units must increase.   The pressures shown on Figure B-l are con-
 sidered to be maximum for reliable operation when using high-density pump and
motor construction.  The hydraulic elements have been sized for a pressure level
of less than 2500 psi during most of their operating life.   At this pressure
 level, infinite life can be expected.   Typical output torques for a 7.5 in /rev
displacement element are 250 ft-lb at  2500 psi and 320 ft-lb at 3200 psi
 (maximum pressure).

 In low-range operation corresponding  to high output torques, the low-range brake
is applied.   Element II  drives the sun gear of Planetary C.  The sun gear drives
                                     B-12

-------
the planet gears which react on the locked ring gear to drive the rear planet
carrier and output shaft at a reduced speed.  Planetary "C" has a reduction
ratio of 1:3.50.  The power then proceeds through the rear axle with a reduction
ratio of 1:3.55 to the wheels.  In this mode of operation the vehicle can be
driven in reverse.  Unit 1 is positioned past its zero displacement to a negative
displacement.  Flow is therefore reversed, pumping fluid into Element II in the
reverse direction turning the output shalL in the reverse direction for reverse
operation of the vehicle.

The transmission stays in low range to 1/3 of maximum output speed.  Maximum out-
put speed is dependent upon the engine input speed.  Thus, the absolute speed
value for the maximum low-range operation depends on engine input which is
required to provide the engine power to satisfy the wheel torque load.

In order to provide higher output speed, a mechanical change in the planetaries
is required.

Referring to the forward gear set of Planetary B:  when Element II is driving
through the rear sun gear, it also drives the forward sun gear.  The forward sun
gear drives corresponding planet gears which are also driven by the forward ring
gear.  This ring is connected to the rear planetary carrier and output.  The
forward carrier is geared to the high-range clutch which is connected to the
transmission input.

This unique construction effects a synchronous mechanical range change.  In low
range as the Element II drives the output it also drives through the forward
planetary to drive the inner element of the high-range clutch in the same direc-
tion as the outer elements are driven by the input.  The gear ratios are selected
so that when Element II is driving the output at maximum speed in low range, the
forward planetary is driving the inner elements of the clutch at the same speed
as the input is driving the outer elements.   At this point of operation, the
low-range brake is disconnected and the high-range clutch is applied.

Since there is no difference of speed between the engaging and disengaging ele-
ments, there is no change in input-to-output ratio, direction of rotation of
                                     B-13

-------
 components, or control position of Elements I and II.  Accordingly,  this range
 transition is completely smooth and with substantially no  slippage or wear  of
 clutch and brake elements.  When changing from high range  to low range, this
 same operation occurs in the reverse direction.  At the point where  the low-
 range brake is applied the ring gear is stationary.

 With the high-range clutch engaged, a mechanical connection from the input  to
 Planetary B is established,  this changes the mode of operation of the primary
 transmission from pure hydrostatic in low range to hydromechanical in high
 range.

 In high range as the Element II speed is decreased, higher output speed is
 effected by the differential action of the forward gear set of Planetary B.  As
 the Element I displacement is decreased to zero, locking Element II  from
 rotating, the forward sun gear of Planetary B is stationary and all  the power
 flow is mechanical from the input through the gear set to  the output.  This
 corresponds to 67 percent of operating range.

 Since the Element II rotation is controlled in the opposite direction of rotation,
 its speed is additive to the mechanical drive of the Planetary "B", and at  the
maximum speed of Element II the output is also driven at maximum speed.

The introduction of the mechanical connection reduces the range over which  the
 hydrostatic drive must operate to approximately one third.  Since the mechanical
elements are much more efficient,  compact and less costly to produce on  a high-
production basis, this approach provides an optimum construction wherein the
 transmission is completely variable over its range,  the size of the hydrostatic
elements is minimized, the mechanical construction is kept simple,  and the  range
change is completely synchronous without any steps,  slippage,  or wear of brake
and clutch elements.

Figure B-5 shows the flow of power through the transmission* when operating in
*Refer to Section VI-C, Performance Analysis for values of power levels and
 other operating characteristics in the transmission at vheicle speeds of 20
 and 70 mph.

                                    B-14

-------
the high range, while Figure B-2 shows the speeds and displacements of the
units.  Although little or no power flows through the primary units pressure
is developed in these units to provide a torque reaction at the sun gear of
Planetary "B".

To achieve the overdrive or 1:1.1 speed ratio Element I is held at maximum
displacement while the displacement of Element II is decreased.  The increased
speed of Element II increases the output speed.

If the torque load remains constant for the overdrive condition, then the
pressure must increase to compensate for the drecrease in displacement.

In the overdrive condition, the power flowing through the hydrualic path may
be greater than the 33 percent previously discussed  (See Section VI-C).  How-
ever, if overdrive is desired, the straight-through point could be altered
to reduce this power.  With the addition of some mechanical components the
maximum hydraulic power could be reduced below 20 percent.

The torque ratio available when the power path is through the hydraulic branch
Ls 2.7 from engine input to transmission output and is 9.6 from engineiinput
to rear axle output when a standard rear axle ratio  ( 3 ) of 3.55:1 is used.

The torque ratio available when the power path is through the mechanical
branch is 1.38 from engine input to transmission output and is 4,9 from
engine input to rear axle output.

The secondary or flywheel drive section will now be examined.  Since the fly-
wheel operates over a relatively small speed ratio and only operates for short
bursts of power in and out, the drive is connected to the relatively constant
speed transmission input.  This reduces the operating range to approximately
3:1.   Since the primary transmission already is sized to provide maximum
torque multiplication to the output, approximately 10:1, there is no advantage
in building a  parallel power path in the secondary drive.
                                    B-ij

-------
 Referring  to  the schematic Figure B-l,  the  input  drives  the Planetary  "A"
 carrier and extends to the rear and drives  hydrostatic displacement  Element  IV.
 Hydrostatic Element III connects to the  ring gear.  Hydrostatic Elements III
 and  IV form an infinitely variable and  reversible drive  similar to Elements
 I and II.  Elements III and IV are not  hydraulically-connected to Elements
 I and II.

 The  operation of this" section is quite  straightforward.  The displacement of
 Units ill  and IV at any particular Llywheel to engine speed ratio is shown  in
 Figure B-3.   Primary control of power is effected by varying the displacement
 of Element IV.

 When Element  IV is at zero displacement, Element  III is  stationary 	 locking
 the  ring gear.  The input drives the carrier and  planet,  gears which react on
 the  locked ring gear to drive the sun gear  and flywheel  at a speed 7.55 times
 engine speed.  At this point of operation,  there  is no hydraulic flow  and all
 power to and  from the flywheel is through Planetary "A".  When the Element IV
 displacement  is controlled to drive Element III and the  ring gear in the same
 direction as  the input, the speed component is subtractive and the flywheel
 is driven at  a lesser speed.  Similarly, when Element III and the ring gear
 are  driven in the opposite direction to  the input, the speed component is
 additive, driving the flywheel at a higher speed.  The operating speed range
 of the flywheel required the displacements shown  in Figure B-3.

 The  flow of power in and out of the flywheel is completely reversible  through
 the  secondary drive, and any small change in the displacement of Element IV
will determine whether the flywheel is accelerated or decelerated with
 respect to the input.

 The  advantages of this construction are apparent.  The major power to and
 from the flywheel is transmitted by the planetary.  The hydrostatic drive
 functions on both sides of the mechanical drive to serve only as a positive
speed control.  This reduces the size of Elements III and IV to less than
    3
6 in /rev.  Since the hydraulic portion of the drive is small, the movemer
of power to and from the flywheel is achieved at the maximum efficiency.
                                    B-16

-------
The flow of accelerating power from the flywheel to the output when operating
in the low range is shown on Figure B-6.  Again, the path for the required
torque reaction is shown using dotted lines.

The I low of accelerating power and the torque reaction for the flywheel to
the output when operating in the high -~nge is shown on Figure B-7.

In all of the hydraulic units, the movement of the swashplate which controls
the pump and motor displacement is a continuous function; therefore, the
transmission ratio control is completely stepless and infinitely variable
within the operating range.

The variable power-splitting transmission herein described affords maximum
efficiency while maintaining minimum complexity and improved operating
characteristics.  The general construction, weight, and size are within the
present state of development of conventional automobile transmissions.
                                   B-17

-------
      SECTION C



PERFORMANCE ANALYSIS
                                                          o
                                                          •o
                                                          a

                                                          -*

                                                          S

                                                          3

-------
                            C.  PERFORMANCE ANALYSIS

Presented in this section in the following order are:  1) a review of the system
considerations and requirements, 2) a definition of efficiencies, 3) a descrip-
tion of the method of analysis, 4) typical cruise power flows through the trans-
mission, 5) steady-state performance results and comparisons, 6) dynamic per-
formance results and comparisons, and, 7) a discussion of flywheel charging
characteristics.

1.  System Considerations and Requirements

The design of a transmission to transfer power between the flywheel, heat engine,
and vehicle wheels is dependent upon the overall propulsion system design goals.
These goals were defined by EPA/AAPS for a family car (1) and then were related
specially to the hybrid flywheel/heat engine propulsion system by LMSC as the
systems contractor for EPA/AAPS.

As a result LMSC defined (4 ) the maximum tractive effort in order to meet the
EPA/AAPS acceleration design goals.  These tractive effort requirements as a
function of vehicle speed are shown by Figure C-l.  Since the size of the heat
engine is dependent upon the selection of maximum steady-state cruise power,
EPA/AAPS concurred that a 5300 pound vehicle should be capable of climbing a
five percent grade at 70 MPH without flywheel acceleration power for a period of
100 seconds.  The resulting road horsepower required to meet those cruise power
conditions is shown by Figure C-2 to be 95 horsepower.  Then at zero grade con-
ditions with vehicle weight reduced to 4600 pounds (per EPA/AAPS design goals),
Figure C-2 shows that the vehicle will be capable of achieving a speed of 96 MPH.

As the systems contractor, LMSC was responsible for determining all characteris-
tics of the flywheel.  From a performance viewpoint, an important parameter of
the flywheel is its loss characteristics.  In order to minimize windage losses
and tip heating, it is necessary to operate the flywheel in a reduced chamber
pressure environment.  The flywheel losses, as determined by LMSC (5 ) ,  are
*
 Since the completion of this study, slightly lower values of flywheel losses
 have been determined by LMSC.
                                     C-l

-------
                                            Note;
                                            Data supplied by LMSC.
Reverse
Driving
Reverse
Braking
                                        30    40    50
                                        Velocity, MPH
       Fig. C-l   Tractive  Effort vs.  Velocity Requirements for

                  Heat Engine/Flywheel/Hybrid Passenger Car Drive  System
                                     C-2

-------

120
100
                          40          60
                        Vehicle Speed, MPH
100
             Fig. C-2   Required Cruise Road Power
                              C-3

-------
 shown  by  Figure  C-3.  Representative  flywheel  types applicable for the trans-
 mission configurations were pierced and non-pierced operating at the chamber
 pressures shown  by Figure C-3.  A  chamber pressure of 2.94 psi was considered by
 LMSC as a reasonable pressure which could be maintained by engine vacuum.  From
 a performance viewpoint, the pierced  flywheel  operating at chamber pressures of
 0.581  and 2.94 psi was considered  representative of a reasonable range of fly-
 wheel  losses and  therefore was used as a performance parameter for results dis-
 cussed subsequently.

 The characteristics of the heat engine, a conventional 1C engine, and the con-
 ventional automatic transmission used for this study were defined by EPA/AAPS.
 The observed horsepower as a function of engine speed for the engine is shown by
 Figure C-4.  Figures C-5 and C-6 give the pertinent characteristics which defined
 the conventional  transmission.

 The accessory and/or auxiliary horsepower requirements of the heat engine (except)
 for the water pump) were defined by EPA/AAPS.  The resultant horsepower required
 as a function of  engine speed is shown by Figure C-7.  For the performance re-
 sults presented herein, it is important to note that the air conditioner was in-
 cluded, since the propulsion system should realistically have the capability of
 handling the extra load.  The sizable additional horsepower required to handle
 the air conditioner is indicated on Figure C-7.

As suggested by LMSC in an initial study (2) of the flywheel/hybrid propulsion
 system, it is desirable to have the control system of the transmission follow a
 "Total Kinetic Energy" (TKE) control law.  This tends to result in a minimum
weight flywheel and theoretically offers the possibility of not having the pro-
pulsion system "history dependent."  The basic idea of the TKE approach is to
control the heat-engine output so as to hold at a constant value the sum of the
kinetic energies of the flywheel and vehicle.  This can be expressed by:
       2       2
     MV   +  Jw   =  constant                                               (C-l)

Then for a selected maximum vehicle speed and maximum flywheel speed knowing
 that maximum flywheel speed occurs at zero vehicle speed, it can be shown that:
                                     C-4

-------
          10
n
i
       CO
       0>
       13
       C
       cd
       oo
       c
       0)
       CQ
       
-------
n   £
I    O
                Data Supplied by  EPA/AAPS
                                                                                                                              Constant  Specific

                                                                                                                              Consumption Lines (Ib/hr/hp)


                                                                                                                                           I    I     I
              800     1000     1200      1400    1600     1890     2000     2200     2400     2600


                                                                            Engine   RPM
2800      3000
                  3200
                            3400
                                     3600      3800
                                                Fig. C-4    Medium Size  Engine  Fuel  Economy Map
                                                                                                                                                MTl-12"!38

-------
    30
    25
                     I               I               I
         Typical "B" Car Transmission Torque Efficiency
         (see Note 1) and Losses

           1st Gear Ratio 2.5:1, Eff. = 95.67.
           2nd Gear Ratio 1.5:1, Eff. = 94.0%
           3rd Gear Ratio 1.0:1, Eff. = 1007=
    20
O
H
to
CO
O
C
•H
C.
c
O
•.H
co
(0
•H
e
CO
c.
    15
10
                                1st Gear
                                          Note 1:
                                                Spin  loss  torque  from
                                                plotted  data  is  in addition
                                                to  torque  loss.

                                                Rear  axle  efficiency esti-  '
                                                mated  at 967,.

                                                Data  furnished by EPA.
                                                   Axle Spin Loss
                                                   (see Note 2)
                                                   I
                   1000
                               2000          3000
                              Propeller Shaft, RPM
4000
                                                                               5000
       Fig.  C-5    Typical "B" Car Transmission Torque Efficiency and Losses
                                       C-7

-------
o

00
                 4000
                 3600
                 3200
                 2800
          2.5
          2.0
        "* 1 5
        4J J. . ->

        as
        01
        
-------

/
/


/
/
f



'

















^^^ Without
Air Conditioner












             1000
                                                      4000
                    2000         3000


                    Engine Speed,  RPM



Fig. C-7   Engine Accessory  plus Auxiliary Power
                                                               5000
                              C-9

-------
     M    I  max [
     J  =  V
          L maxj
1  -
                                     21
      u
       max
(C-2)
Therefore, setting the flywheel speed ratio, maximum vehicle speed, and maximum
flywheel speed determines a fixed ratio of vehicle mass-to-flywheel inertia.
Figure C-8 presents typical flywheel inertias as a function of vehicle weight
for various flywheel speed ratios.  It can be seen that for a flywheel speed
                                                          *
ratio of 24,000/10,000 and a vehicle weight of 5300 pounds , the required fly-
                               2
wheel inertia is 0.49 Ib-ft-sec  in order to follow TKE at vehicle speeds up to
85 mph.  Consequently, the aforementioned value of flywheel inertias was used
for the transient performance results subsequently discussed.

In addition, it should be noted that, for the cruise-power results which follow,
the flywheel losses were computed by assuming the flywheel followed TKE between
24,000 RPM and 10,000 RPM corresponding to vehicle speeds from 0 to 85 MPH, re-
spectively.

2.  Definition of Efficiencies
It is worthwhile to clearly define various efficiencies which were used to des-
cribe the performance of the flywheel/hybrid transmission in this study.  The
approach used to define the efficiency of the flywheel/hybrid transmission was
to follow the customary definition of efficiency for a conventional automotive
transmission.

For a conventional automotive transmission, transmission efficiency is simply
defined as:
     ~     .   .    .......         Transmission Output Power    , _„           ,„ ^N
     Transmission Efficiency  =	—t——	  x 100           (C-3)
                                 Transmission Input Power

At steady-state (cruise-power) conditions with the flywhee/hybrid system, it is
necessary for the engine to supply power through the transmission to the flywheel
K
 Car weight specified by EPA/AAPS (1) for acceleration and grade requirement
 design goals.
                                     C-10

-------
         1.0
n
i
     o
     01
     in
     i
     O
     C
     0)
     PL,
         0.8
         0.6
        0.4
        0.2
                   1      ]      I      T      I

              Maximum  Vehicle  Speed  =  85  MPH

              Maximum  Flywheel Speed = 24,000

                                           RPM
                                                                                 Minimum Flywheel
                                                                                   Speed, RPM
                   Flywheel  Speed  Ratio
                                                                                  12,000
                                                                                       I
                                                                                  10,000
                                                                                       I
                                                                                   8,000
           2800
3200
3600
                                               4000         4400         4800        5200

                                                         Vehicle  Weight,  Ibs.

                                   Fig. C-8   Required Flywheel Inertia for TK.E  Control Approach
                                                                         5600
                                                                         6000
6400

-------
for flywheel bearing, seal, and windage losses in order to maintain  the  flywheel
at constant speed according to TKE control.  This results in transmission output
power which is in addition to that power going to the vehicle wheels.  Thus, it
follows from Equation C-3 that, at cruise-condition, the flywheel/hybrid trans-
mission efficiency was defined as:
                                          HP      HP
     Transmission Efficiency (Cruise)  -    ™   •	  ^"                     (C-4)
                                              Eng-net
where:
     HP ..  ..    «•  Power required at flywheel due to bearings, seals and
       F ly—.loss
                   windage losses (see Figure C-3).
     HP         =  Net power from engine at transmission input or observed
       LSilE~H6 L
                   brake horsepower of engine (Figure C-4).minus auxiliary
                   and accessory power (Figure C-7).
     HP         =  Drive train power at transmission output or the road
                   horsepower developed at the vehicle wheels divided by the
                   rear-end differential efficiency.
Any power required for flywheel vacuum and lube pumps (shown earlier by Figure
C-3) was assumed to be additional engine accessory power.

The overall efficiency at which steady-state cruise-power was transmitted to
the rear wheels was termed power train efficiency and defined as:
     Cruise Power Train Efficiency  =  Road Horsepower	
                                       Transmission Input Power
                                       HP
     Cruise Power Train Efficiency  =  -r^-	   x 100                    (C-5)
                                       HP_
                                         Eng-net

Where HP  was that road horsepower required at the vehicle rear wheels to main-
tain a given vehicle speed and HP_,       was the transmission input power as
                                 Eng-net
previously defined by Equation C-4.
                                     C-12

-------
During vehicle acceleration, when the flywheel supplies power to the rear wheels,
the transmission input power from the fb"-»heel is given by:
      HP,,,       =  Jr N, • a - HP,,.                                         (C-6)
        Fly-net      f  f        Fly-loss
where:
      J   =  Flywheel Inertia
      N   =  Flywheel Speed
       a  =  Flywheel Acceleration (dN /dt)
Thus, the efficiency of the flywheel/hybrid transmission during vehicle accelera-
tions was defined as:
                                             HP
      Transmission Efficiency (Ace)  =  —	°U + Hp	  x 100        (C-7)
                                          Eng-net     Fly-net

A more refined breakdown in efficiencies, for example, in order to account  for
the efficiency at which power is being supplied from or to the flywheel, was
not possible with the type of transmission design presented herein.  This was
because engine power and flywheel power flowed through common paths.

For comparative performance evaluations, the efficiency of a conventional auto-
matic transmission was calculated.  The characteristics of the conventional
transmission were given earlier by Figures C-5 and C-6.  Since no flywheel was
involved, the efficiency of the conventional transmission was simply defined
by Equations C-3 and C-5 presented earlier.

3.  Method of Analysis

In order to determine the performance of the flywheel hybrid transmission, non-
linear models of the complete propulsion system were developed and programmed
on a digital computer to calculate:  1) steady-state performance and 2) dynamic
performance.  Refer to Appendix I for details of the equations used to describe
the transmission (including losses) and a description of the computer programs.
In addition, a model of the complete propulsion system was setup on the Bendix
analog computer in order to investigate system stability (see Appendix II).
                                     C-13

-------
The steady-state computer model included the transmission  (with losses described
on a component basis), engine, flywheel and associated vehicle characteristics.
The basic input to the model was desired vehicle speed.

The dynamic performance digital computer model was a simulation of the complete
propulsion system.  The model simulated all significant dynamics (torque and
speed relationships) of the flywheel, transmission with controls, engine and
vehicle (characteristic of a family car) in the time domain.  Compressibility
dynamics in both hydraulic pump-motor circuits (primary and secondary) were
included in addition to other losses in the transmission on a component basis.
Engagement  and disengagement of the secondary and primary transmissions were
not included on the model.  The basic input to the dynamic model was driver
command pedal as a function of time.

In both the steady-state and dynamic models of the transmission, losses were
accounted for on a component basis as a function of component operating levels.
Component losses were based upon correlation with measured losses for similar
components.  For example, typical losses in the hydraulic elements which were
calculated based upon correlation to experimental data included:
     1.  Compressibility of oil
     2.  Leakage which includes leakage past the pistons and past the
         valve plate and drum seals.
     3.  Mechanical which includes piston friction, bearing friction, and
         valve plate friction.
     4.  Flow which includes valve inlet and outlet hydrodynamic losses and
         duct losses.

As shown in Appendix I, these losses vary with system pressure, element speed,
swashplate displacement, and flowrate.

It is important to realize that in the flywheel/hybrid transmission design
presented herein that it is possible for the hydraulic elements to have low
component efficiency and yet not have a major effect upon transmission effici-
ency.   This is because the hydraulic elements can transmit high torque while

                                    C-14

-------
transmitting low power.  Since pressure is proportional to torque, a high  torque
results in high pressure which in turn causes high compressibility and leakage
losses.  This, of course, results in a relatively low efficiency for the hydrau-
lic element on a component basis.  However, since little power is going through
the hydraulic element, the effect on transmission efficiency is negligible.

4.  Typical Cruise Power Flows in Transmissions

As discussed earlier in Section VI-B, one of the key features of the split--pov"r
transmission design presented herein was that, in the primary transmission, moj>.
of the power flowed through the mechanical rather than the hydraulic path  in
order to achieve high efficiency.  In order to give a typical example of th°
power-split and also to provide some insight to the performance calculations,
consider the results obtained for steady-state vehicle speeds of 20 and 70 MPH.
with a pierced-flywheel configuration operating at a chamber pressure of 2.94 nt-.
A complete table of computer results for these conditions is given in Appendix T

Figure C-9 shows all the power paths through the transmission in the high-speed
range.  Under steady-state, level-road conditions the transmission was in  the
high-speed range for vehicle speeds above 11 MPH.

Locations throughout the power paths are identified and values of power at each
of these locations are shown by Figure C-9.  Also shown are the resultant  effi-
ciencies for the transmission, primary hydraulics, and secondary hydraulics.
Note that negative values of power indicate power flow in a direction opposite
to the arrows shown on the diagram of Figure C-9.

At 20 MPH, 94 percent of the power in the primary flows through the mechanical
path to the rear wheels.  Under these conditions, the power through the primary
hydraulics (about 0.3 horsepower) flows in a reverse direction.  Since the pri-
mary hydraulic units are at very low pressure (121 psi), speed, and flow condi-
tions for their respective ratings, their losses are negligible.   For all  prac-
tical purposes, at 20 MPH steady-state speed, the transmission can be considered
similar to a mechanical geared drive.
                                     C-15

-------
 Kng I na
                                                                           	•  	 	 _
                                                                           l,oge
                                                                                                               Wheels
Flywheel

'R
rs
"6I

Ring
.'ETARY
1
1

R
K6





Element
III
•
Element
IV






R5

                                                                                                 PRIMARY PLANETARY
R = Gear Ratio
r = Gear Radius
                                             , M ,
                                   Power Path Locntlnn
A Trminnil ss 1 on input. Power
R 	
C 	
[) 	
K 	
F 	
C 	
H Transmission Output Power
I Road Horsepower
J 	
K 	
L 	
M 	
N 	
0 Flywheel Power
12.37 HP
. . • 11.37 IIP
-.32 IIP
-.32 HP
-.33 IIP
5.39 HP
. . 5.26 HP
4.77 HP
4.59 HP
. . 2.71 HP
. . 3.55 HP
. . 3.51 HP
. . 3.41 HP
. . 3.38 HP
6.00 HP
                                                                  Vehicle Spend, MPII
                                                                  20               70
                                                                               52.51  HP
                                                                               51.11  HP
                                                                               19.38  HP
                                                                               19.18  HP
                                                                               18.26  HP
                                                                               29.25  HP
                                                                               28.54  HP
                                                                               45.25  HP
                                                                               43.51  HP
                                                                                2.94  HP
                                                                                -.70  HP
                                                                                -.71  HP
                                                                                -.71  HP
                                                                                -.72  HP
                                                                                2.19  HP
                                  Performance  Summary

                         Power Train Efficiency                 37.11%           82.86%
                         Transmission Efficiency                 87.107.           90.37%
                         Primary Hydraulics Efficiency          100.32%           95.17%
                         Secondary Hydraulics Efficiency         97.13%          100.89%
                         Primary Pressure                      120.7 PSI        455.2 PSI
                         Secondary Pressure                    187.0 PSI         63.8 PSI
                         Engine Horsepower                      19.98 HP         67.17 HP
                         Auxiliary Horsepower                    7.61 HP         14.67 HP
                         Engine Speed                           1400 RPM          2817 RPM
                         Flywheel Speed                       23,444 RPM        15,911 RPM

                         Note:  I. Efficiency greater than 100% and negative powers mean
                                  power flow in direction opposite to arrows.
                               2. Refer to Appendix I for additional detailed  results.
             Fig.  C-9    Typical Power  Flow  in  Transmission  at  Steady-State
                            Speeds  of  20 and  70  MPH
                                                          C-16
                                                                                                            MTl-12342

-------
Although the effect of flywheel losses is subsequently discussed, it should t>3
observed here from Figure C-9 that at 20 MPH 12.4 horsepower was coming into the
transmission and 6 horsepower was required at the flywheel with only 4.8 horse-
power at the transmission output to produce 4.6 road horsepower.  Thus it is
clear why the power train efficiency was only 4.6/12.4 or 37.1 percent when the
transmission efficiency was  (4.8 + 6)/12.4 or 87.1 percent.

At 70 MPH, the results given by Figure C-9 can be interpreted to show thst 3°!
percent of the power in the primary flows through the hydraulic elements and fii
percent through the mechanical path.  Under these conditions, the pressure in
the primary  (455 psi) was only 15 percent of maximum rated for the hydraulic
units.  As a result of the relatively low pressure, the efficiency of the hyd1 •..;•;.••
lie units was 95 percent.

As pointed out earlier in this section, the efficiency of all hydraulic units
was correlated to measured experimental data.  The resultant loss coefficients
are presented in Appendix I with the equations used to describe the character-
istics of these hydraulic units in the performance models of the transmission.
It can be readily shown from the loss coefficients presented in Appendix I that,
under the worst combination  of rated conditions (pressure, speed, displacement,
etc.) these hydraulic units would have an efficiency of 85 percent.  Under com-
parable conditions other types of hydraulic units have slightly lower efficien-
cies.  The higher efficiency capability as well as the higher power density are
important features of the hydraulic units selected for the transmission design.

As shown by the results of Figure C-9, under steady-state level road conditions,
particularly at low vehicle speeds, characteristic of urban driving, the hydrau-
lic elements were not an important factor in determining transmission perform-
ance.  This was due to the design characteristics of the transmission resulting
in element operation at high efficiency and low-power flow.

5.  Steady-State Performance Results and Comparisons

Consider now the steady-state performance of the flywheel/hybrid transmission
over a range of vehicle speeds.  The calculated transmission efficiency is shown
by Figure C-10 for two different but representative flywheel losses.  Note that

                                     C-17

-------
        100
o
i
oo
         80
         60
     c
     o>
     u
     (-1
     0)
     CL,
u

£   40
•H
U
•r-<
M-l
14-1
U
         20
                                Pierced Flywheel,  P   = 2.94 psi
                                            Pierced  Flywheel, P  = 0.581 psi
                                                                          Vehicle Weight =  4600 Ibs

                                                                          0% Grade

                                                                          Engine Accessories  Include Air  Conditioner
                       Efficiencies < 10 MPH

                       are  Approximate
                                               Transmission Efficiency (Cruise)
                                                                    (HP   + HP_.  ,  ) x 100
                                                                      out	Fly-loss	
                                                                               HP
                                                                                 Eng-net
                       10           20           30           40           50          60            70


                                                           Vehicle Speed,  MPH


                         Fig.  C-10  Cruise Power Transmission Efficiency of Flywheel/Hybrid  Transmission
                                                                                                           80
90

-------
efficiencies below 10 MPH are at iest inly an approximation.  Referring to
Figure C-10, it can be seen that the efficiency varied between 77 and 90 percent
for vehicle speeds from 10 to 85 MPH, respectively, with the low-loss flywheel
(Pc = 0.581 psi).  With the higher-loss flywheel, efficiencies varied from 85
to 90 percent over a similar vehicle speed range.  This trend is as expected,
since increased flywheel losses require additional power flow in the transmission
and the effect of residual (or parasitic) power losses in the transmission are
reduced - particularly at low vehicle speeds.

In order to minimize emissions, it is important that the transmission be capable
of transferring engine power to the road and flywheel with high efficiencies 31
cruise power.  The relatively high efficiencies shown by Figure C-10 indicate
important design feature of the transmission presented here.

In a similar manner, Figure C-ll presents the calculated cruise-power train
efficiency.  The rapid decrease in power train efficiency as vehicle speed is
decreased is due to flywheel losses.  This occurred because the power loss of
the flywheel was, of course, not reflected in wheel horsepower.  Thus, reducing
the flywheel losses caused a substantial improvement in power train efficiency
as shown by Figure C-ll.

A most important performance parameter of the flywheel/hybrid propulsion system
at steady-state cruise power conditions is the level of engine emissions.   Since
emission data were not available for the engine from EPA/AAPS during the course
of this study, discussion must be limited to fuel consumed by the engine (or
miles/gallon) with the assumption that on a first approximation basis this is a
reasonable indication of emissions,

Figure C-12 presents the calculated fuel economy of the flywheel/hybrid propul-
sion system at steady-state cruise power as a function of vehicle speed.  These
results indicate that with a low-loss flywheel (Pc = 0.581 psi),  the best  fuel
economy obtained was 16.2 miles/gallon at 30 MPH.  Increasing the flywheel
losses (Pc - 2.94 psi) lowered the results to 15 miles/gallon at  30 MPH.
                                     C-19

-------
        100
n
i
         80
             Vehicle Weight = 4600  Ibs

             0% Grade

             Engine Accessories Include  Air Conditioner

                 (Figure C-5)

             Flywheel Losses per  Figure  C-3
                                                           ,Pierced  Flywheel, P  = 0.581 psi
     c
     
-------
o
I
NJ
20
    Vehicle Weight = 4600  Ibs
    0% Grade
18  Engine Accessories Include Air
        Conditioner (Figure  C-5)
    Flywheel  Losses per Figure C-3
16  Fuel Density = 6.152 Ib/gal
          14
                                                                 Pierced  Flywheel,  P   = .581 psi
                                                                                                    _ j:_
                                       kF-
                                       fi
          10
                                                                               .   .;
                                                                         .
                                                   i -     :]__:-

 4
           2


           0

                                      ;
                          -' ;""-i :, ! •'••":. . . ! "
                          trcrttttrrjtftti'' '•* ~
           gfnjf
           *T1 "^1            i
                 .  I I
                                            I'
                    i-;;::
                    •
                                               '.
                 .  -   ~
                         10
20
 j
 J

30
                                                                                   ,
                                                                                 L±tH       _..:. i.

                                                                                            :  --•-..::
                                                                                         •  -—-r—


                                                                                                                       ^p
                                                                                                      rrrrt
                                                                                                      tfJ.fnlr^i•  '•   •  :."" '.'
                                                                                                      fnTuiTT'"  '  • "~~-;~
                                                                   .-.-:
                                                    ^^-"—- :. ;:-~fc 4"
                                                                                 '   ,.
                                                                                1

                                                     40           50
                                                 Vehicle Speed,  MPH
                                                                                        60
    -
., .

    70
                                                                            80
.

  90.
                               Fig.  C-12  Fuel  Economy of Flywheel/Hybrid Propulsion System at Cruise  Power

-------
 Comparable  steady-state cruise  power performance calculations were made for a
 conventional  propulsion system  with a conventional automatic transmission and
 the  same  size 1C  engine (defined by Figure C-4).  It was assumed on a level
 road  at cruise power  that  the conventional automatic transmission was in first
 gear  at up  to 10  MPH,  second gear at 15 MPH and in third gear from 20 to 85 MPH.
 Depending upon the  particular characteristics of the transmission controls and
 their design  settings, Lt  is quite possible that the automatic transmission
 could be  In third gear from 10  Lo 85 MPII til: c-.ruf.se conditions.

 The resultant  transmission efficiency calculated for the conventional automatic
 transmission  is shown  by Figure C-13.  The affect on transmission efficiency
 of operating  in various gears is clearly indicated.  The corresponding fuel
 economy is  shown  by Figure C-14.

 A comparison  between  the transmission efficiency of the flywheel/hybrid and
 conventional  automatic transmissions is shown by Figure C-15.   These results
 indicate that  the hybrid transmission design presented herein should be more
 efficient than  the automatic transmission particularly at cruiso speeds from
 20 to 40 MPH.  Although flywheel losses Lend to improve the ei'lMc i.ency of t lie
 flywheel/hybrid transmission as noted earlier, the hydromechanics I (power-
 splitting)  transmission design used will sl.ill exhibit better efficiency at. low
 vehicle speeds  even without flywheel losses.

 Figure C-16 presents a comparison of power train efficiencies at cruise power.
 The affect  of  flywheel losses is clearly evident in the lower power train
 efficiency  of  the flywheel/hybrid transmission at low vehicle speeds.  These
 results can be misleading  since they do not indicate the relative effect upon
 engine performance and emissions.

A more important  comparison is on the basis of fuel economy.  Figure C-17  pre-
 sents this  comparison at cruise power in terms of miles/gallon.   The comparable
 percentage  change in  fuel  economy is then shown by Figure C-18 for two different
 flywheel losses.  These results indicate that, the flywheel/hybrid propulsion
 system will have  a decrease in cruise fuel economy at low vehicle speeds
                                    C-22

-------
       100
r>
to
            Vehicle Weight  =  4600  Ibs
            07, Grade
            Accessories  Include  Air  Conditioner
                (Figure  C-5)
        80
        60
        40
        20
/
/
3rd Gear

— .„ . .


t
J
i
•
•
                                Solid  Curve  (      )  Indicates  Standard
                                Automatic  Transmission Efficiency Selected
                                for Cruise Power
                                                i
                                    •      r      ]
                                Dashed Curves  (	) Indicate Effect of
                                Staying in Various Gears
                                        I
                                                                                       .
                                                t .  . \

                                                                                  (Transmission Output Power) x 100
                                                        Transmission Efficiency = 	-	:	f-	—	
                                                                                      Transmission Input Power
                     10
20
30
   40          50
Vehicle Speed, MPH
60
                                                                                             70
80
90
                  Fig.  C-13   Transmission Efficiency at  Cruise  Power —  Conventional  Automatic  Transmission

-------
n
t-o
   20
   18
   16
   L4
|
o
§  10
u
5J
        Vehicle Weight = 4600  Ibs
        0% Grade
        Fuel Density = 6.152 Ib/gal.
        Engine Accessories Include  Air
              Conditioner  (Figure  C-5)

                                                                     Solid Curve Indicates Fuel Economy
                                                                     With Standard Automatic Transmission
                                                                     at Cruise Power
                                                                                                            ±jr.!
                                                                                                    ..  •},.

                                                                                : _:
                                                                  3                            ^
                                                                   "* 2nd  Gear
                                                                          ;
                                                          Dashed Curves Indicate Effect
                                                          of Staying in  Various Gears
                                                       !

                                                                 .. ,i
                                                                  :•
                                                                 •
                                                                                   -__—i   --^..*rr- -,: r 4-'*"—-"•"
                                                                                            .}—•::  -
                                                                                            i_lL^r:
                                                                                		jrr::,
                                                                                   ,_ .^l-r-
                                                                                      • :—"I'. :
                                                                                               •


                                           30
                                                       40          50
                                                   Vehicle Speed, MPH
                                                                                     60
   .

70         80
                                                                                                             _
                                                                                                              "" -  ; " .  -
                                                                                                           • •
                                                                                                             '  I
                                                                                                            ;	•
90
                           Fig. C-14    Fuel  Economy at Cruise Power —  Conventional Automatic
                                        Transmission (No Flywheel)

-------
         100 ,
                                                   Flywheel/Hybrid Transmission
o
                                                    (With  Inline  Pierced Flywheel, P  = 2.94  psi)
                                                               Standard  Automatic Transmission  (Without Flywheel)
                                                                                                        1             •   '
±±±t--hrri--r- -
                                                                          : Vehicle  Weight = 4600 Ibs

                                                                           0% Grade

                                                                           Engine Accessories Include Air Conditioner --
                                                                                        (HP   + HP_, ,  ) x 100
                                                                                         out	Fly-loss	
                                                              Transmission Efficiency (Cruise)
                                                            40          50


                                                          Vehicle Speed, MPH
                             Fig.  C-15   Comparison of Transmission Efficiencies at Cruise Power
                                                                                                                        MTI-12492

-------
         100 ;	
ro
CT
                                                        Standard Automatic Transmission
                                                        (Without Flywheel)

                                                               Flywheel/Hybrid  Transmission Power Train
                                                                (With  Inline  Pierced  Flywheel, P  = 2.94 psi)

                                                                 Cruise Power Train Efficiency =
                                                                         I     ,'-   .  T::/
                        10
20
30
                                                            40          50

                                                          Vehicle Speed, MPH
                                  Fig.  C-16   Power  Train Efficiency Comparison  at  Cruise Power
                                                                                                                       MTI-12493

-------
o
20




18




16




14




12




10
                                               f-.. ::••:
         Vehicle Weight = 4600 Ibs

         0% Grade

        "Fuel Density = 6.152 Ib/gal
      tut
        "Engine Accessories Include Air Conditioner

                                                                  ^Standard Automatic Transmission

                                                                  :r (Without Flywheel)
                                                                                                    • • '   '
                                                Flywheel/Hybrid Transmission
                                                (With Inline Pierced Flywheel,£
                                                     40          50

                                                  Vehicle Speed,  MPH
                              Fig.  C-17   Comparison of Fuel Economy at Cruise Power
                                                                                                              MTI-1249A

-------
            20
n
ho
CO
         •H
               Steady-State Cruise Power
               Vehicle Weight = 4600  Ibs
               0% Grade
              .Fuel Density = 6.152 Ib/gal.
               Engine Accessories Include Air Conditioner
         u
         c
*
on
B
I
r«
                                                                       	   I	. _  t__
                                                         Pierced Flywheel. P
                                                                     Pierced Flywheel, P  = 2.94 psi
                                                                                        c

                                                            40          50
                                                          Vehicle Speed, MPH
                               Fig. C-18    Percentage Change  in Cruise MPG Compared to Conventional
                                            Automatic Power Train
                                                                                                                     MTI-12495

-------
(typically encountered in urban driving)  compared  to  a  conventional  propulsion
system with an automatic transmission.   The maximum decrease occurred  at 20 MPH.
This loss in fuel economy is due to flywheel losses.   For the higher loss fly-
wheel (Pc = 2.94 psi) the decrease in miles/gallon at 20 MPH was 18  percent.
Reducing the flywheel power losses by a total of 577» (see Figure C-3 earlier)
resulted in improved fuel economy, but it was still 10 percent less  than the
conventional system at 20 MPH.  Thus, even with the improved performance capa-
bilities of the variable ratio transmission (compared to the conventional auto-
matic), a decrease in cruise performance at low speeds can be expected.   At
higher speeds, as shown by Figure C-18, an improvement up to nine percent in
fuel economy was obtained since flywheel losses are minimal.

A comparison of fuel flow used at idle by the existing heat engine/conventional
automatic transmission system and the heat engine/flywheel power-splitting
transmission system was made.  The heat engine/flywheel system uses  approximate-
ly 18 percent more fuel.  The increased fuel flow is  a direct result of  the higher
idling speed for the flywheel system.  The standard engine idles at  800  RPM,
has an idling torque requirement of approximately 85  foot-pounds and has a fuel
flow of 7.7 pounds per hour.  The flywheel system idles at 1200 RPM, has a torque
requirement of 52 foot-pounds and a fuel flow of 9.2  pounds per hour.

The idling speed for the flywheel engine was selected to provide sufficient
engine horsepower to charge the flywheel within an acceptable time span.  At
the speed selected, approximately 40 engine horsepower is available  to charge
the flywheel.  If the flywheel required a 20 percent  makeup in energy  this would
take approximately 12 seconds.
                                                      )
Any trade-off that would lower the hybrid system idling speed would  reduce the
fuel flow and tend to equalize the idling performance of the two systems.

The flywheel/hybrid transmission was designed to allow the engine to operate at
minimum SFC whenever possible.  Operation at minimum  SFC was considered  desirable
in order to minimize emissions.  For the engine size  specified by EPA/AAPS (see
Figure C-4) it was difficult to have engine operation at minimum SFC at  low ••
                                    C-29

-------
 engine power requirements and:  1) transmit engine cruise power to the rear
 wheels (and flywheel) at maximum efficiency; 2) have a minimum weight transmis-
 sion; and 3) provide the capability of charging the flywheel at zero vehicle
 speed.

 A comparison of the engine SFC at cruise power is shown by Figure C-19.  In
 general, the flywheel/hybrid  propulsion system tended to result in operation
 closer to minimum SFC than the conventional system.  At 10 MPH, a decrease of
 20  to 32 percent in engine SFC was obtained dependent upon flywheel losses.
 Operation at improved engine  SFC was due to flywheel losses (more engine power
 required) and  the design features of the transmission.

 The resultant  engine horsepower required for the flywheel/hybrid propulsion
 system at cruise conditions is given by Figure C-2Q.  Also shown is the engine
 horsepower required by a conventional system with a conventional automatic
 transmission.  Recall that flywheel losses and accessory losses were given
 earlier, the maximum engine power at 3800 RPM is approximately 177 horsepower.
 Thus, the 1C engine could have been reduced approximately 26 percent in size
 for the flywheel/hybrid propulsion system and still meet the EPA/AAPS design
 goals.  Reducing flywheel losses will have little effect on the sizing of the
 1C  engine since at 70 MPH, the flywheel losses are relatively small.

 It  is important to point out  that the reduced engine size required for the
 f lywheeil/hybrid propulsion system could, perhaps, result in better cruise per-
 formance than  the results presented earlier.  Further analysis is required in
 order to determine whether such an improvement is possible.

 3.  Dynamic Performance

 Before discussing the dynamic performance of the flywheel/hybrid  propulsion
 system,  some fundamental differences between that system and  a  conventional
 automotive propulsion system should be clarified.  The basic  concept of the
flywheel/hybrid system is that the flywheel supplies the kinetic  energy to ac-
 celerate the vehicle, and the heat engine only supplies that  power  required to
                                    C-30

-------
         50
         40
               Steady-State Cruise Power
               Vehicle Weight = 4600 Ibs
               0% Grade
               Engine Accessories Include Air Conditioner
n

      01
      K
      03
      O
      o  20
                                    •Pierced Flywheel, P  = 2.94 psi
         10
                                            Pierced Flywheel, P  = 0.581 psi
                                                          40          50

                                                       Vehicle Speed,  MPH
                                 Fig.  C-19    Percent Decrease  in  Cruise SFC Compared  to
                                              Conventional Automatic  Power Train
                                                                                                                   KTI-12M9

-------
    160
         Engine Accessories Include Air Conditioner  (Figure C-5)
         Based on Inline Pierced Flywheel  P   =2.94  psi
    140
 a
 8
w
•a
OJ
1-J
                                                         4600 Ib Vehicle
                                                         0% Grade
                                       Standard Automatic Transmission
                                       Without Flywheel
                  20
40          60
    Vehicle Speed, MPH
100
120
            Fig. C-20   Required Engine Power  at Cruise Conditions
                                      C-32

-------
provide steady-state power and to overcome system energy losses.   These losses
include accessory, auxiliary, flywheel, transmission,  air drag,  and rolling
resistance losses.  Since the torque (or power) developed by the flywheel is
proportional to its angularj acceleration, the flywheel produces  no useful power
                           ' dn
at steady-state conditions
                            dt
sion system, it is not feasible to
   Unlike a conventional automotive propul-
obtain steady-state performance maps over the
complete output power spectrum of the flywheel/hybrid propulsion system.   This
means that maximum torque output can only be achieved when flywheel and vehicle
speed are changing as a function of time.

Furthermore, the dynamic response of a flywheel/hybrid propulsion system for  a
given engine, flywheel, and vehicle is primarily governed by the transmission
design characteristics and its associated controls.   Design characteristics such
as the speed ratio are particularly important.  With variable ratio transmissions,
the time rate of change of ratio can have a significant affect- ;ipon system rer
sponse during accelerations or decelerations.

Two types of vehicle operating modes can be considered in order to evaluate the
dynamic performance of the flywheel/hybrid propulsion system:

   1.  Maximum Performance 	 defined by EPA/AAPS Design Goals (1).

   2.  Driving Cycle Performance 	 defined by DHEW driving cycle.

Analytical determination of realistic performance characteristics for  either
of these modes requires solution of torque and speed relationships in  the time
domain.  Comparison to a conventional propulsion system requires a similar
approach to be followed for the conventional system  (&)•

Since the scope of this study was limited to determining  transmission  efficien-
cy, dynamic performance analysis was limited to the  maximum power mode.  This is
a  "worst case" condition for.the transmission with respect to efficiency of
power transfer from the engine and flywheel.  Also,  it demonstrates the maximum
capabilities of the design.
                                     C-33

-------
The dynamic results for the flywheel/hybrid propulsion system under
the maximum power mode were then cross-plotted versus vehicle speed so that
a comparison could be made to maximum power performance of a conventional
system with a standard automatic transmission.

As pointed out elsewhere in this report, the desired maximum power capability
of the transmission determines the size (capacity) of the hydraulic units in
the transmission.  Hydraulic units were sized to meet the tractive effort re-
quirements given earlier by Figure C-l.  Once a particular size is selected,
the maximum power throughout is limited by the pressure relief valve settings
in the hydraulic units.  Thus, control characteristics in the secondary trans-
mission (governing the flywheel) must be selected so as to call for control
power outputs that do not exceed, except for very short durations, the maximum
capability of the hydraulic units in the secondary transmission.  In addition,
the control characteristics must provide for stable operation of the complete
system with desired engine performance.

Pertinent characteristics of the full-range digital simulation of the propul-
sion system used to determine dynamic performance were as follows.  The full-
power transient was initiated by a step input in the driver pedal commanding
                                *
a speed change from 8 to 85 MPH.   Vehicle weight was 5300 pounds (per EPA/AAPS
Design Goals) operating on a level road.  Engine accessory horsepower was given
earlier by Figure C-7.  The flywheel was a pierced configuration operating at
                                                                2.
a chamber pressure of 2.94 psi with an inertia of 0.49 ft-lb-sec     The 1C
engine used was defined earlier by Figure C-4.  An engine inertia of
                2
0.0568 ft-lb-sec , typical for that size engine, was assumed.  Other inertias
*
 A lower speed of 8 MPH was selected to safely avoid operation at lesser
 speeds where the model was invalid.
                                    C-34

-------
included in the dynamic model were the vehicle rear axle and differential,
vehicle wheels, and major rotating components in the transmission.
                        t4

The: control, system was the open loop scheduling or Type B (see Section Vl-D)
setup for TKE with the flywheel speed varying between 24,000 RPM ami 10,000 RPM
for vehicle speeds of 0 to 85 MPH, respectively.  Control gains and rate limits
were adjusted in order to:

   1.  Provide a reasonable level of performance.
   2.  Avoid wheel slip (which was not simulated).
   3.  Stay within maximum power limitations of the transmission.
   4.  Provide engine power to make up for system losses.
   5.  Stay within control stability limits.

Additional tuning and optimization of the system controls would have easily
resulted in improved acceleration characteristics of the vehicle.

Fi r.urc- C-21 presents the resulting vehicle and flywheel speed transients Cor
36 seconds after the step change in driver pedal.  The vehicle reached 60 MPH
in 1.4 seconds and 80 MPH in 26 seconds.  This was slightly below the EPA/AAPS
goal of 0 to 60 MPH in 13.5 seconds.  As previously mentioned, further tuning
of the controls (closer to wheel slip) would have resulted in achieving the goal
without altering the propulsion system.  Since the average power developed at
the wheels from 25 to 80 MPH was 120 HP and within the limits specified by
LMSC (see Figure C-l) for that speed range, these data were considered represen-
tative of a full power transient.

As shown in Figure C-21, flywheel speed reached its minimum value (10,000 RPM)
prior to vehicle speed reaching 85 MPH.  This was caused by the control system
in order to minimize excessive "undershoot" in flywheel speed.  Lowest flywheel
speed was 9891 RPM at. 32 seconds.  Thus, the maximum "undershoot" in flywheel
speed was only 1.1 percent less than the final steady-state speed.  This demon-
strates quite acceptable control performance.
                                   C-35

-------
o
  110
100
   90
   80
   70
     ac
     O-.
   60
   50
                30
               28
                26
                24
                22
             £  20
                18
              a
              in
u
•H
j:
                16
             .c
   30
   20
   10
         0 L
                14
                12
                10
               Step Change in Driver  Pedal  at  t = 0
               Vehicle Weight =  5300  Ibs
               070 Grade
               Pierced Flywheel,  P  = 2.94  psi
                                          8     10     12     14     16     18     20     22     24     26     28    30     32   34    36
                                                                  Time, Seconds
                              Fig.  C-21   Engine and Flywheel Speed During  Full-Power  Transient
                                                                                                                               MTt-' 514

-------
Referring again to Figure C-21  it- -.m be seen that at 36 seconds  (end of com-
puter run) vehicle speed is within '.I'.ree percent and gradually approaching its
steady-state value of 85 mph.  Final steady-state conditions would have been
reached in approximately another 14 seconds.  This type of damped  response is
desirable since it shows that as the system transient settles out, there are
no undesirable oscillations in vehicle speed.

The change in energy developed by the flywheel (less losses) and engine during
the transient is presented by Figure C-22.  Table C- 1 presents an  energy balance
for the transient for a vehicle speed change from 8 to 80 mph.  The latter spp^d
was selected since the flywheel contribution was minimal at higher speeds as
shown by Figure C-22.  These results indicate that during the full power transient:

1.  Total energy transferred to the rear wheels was 2845 horsepower-seconds or
    74 percent of the input energy from the flywheel and engine (3852 horsepow°r-
    seconds).  If accessory power is subtracted, 84 percent of the input energy
    is transferred to the rear wheels.
2.  The transmission and rear axle (power train) losses were 12 percent (475
    horsepower-seconds) of the input energy.  This resulted in an  effective
    power train efficiency of 86 percent calculated as follows:
               .,.•=•<-«:   2845 x 100
         Power Train Eff = o    + 475 =
    Dividing by the rear axle efficiency (96 percent) results in a transmission
    efficiency of 89.5 percent on an energy transfer basis.  Thus, the trans-
    mission under full power transient conditions transferred energy to the
    vehicle wheels at a relatively high efficiency.
3.  Although control settings were not optimized, the engine supplied only 13
    percent less energy than it should have in order to follow ideal TKE control;
    the change in vehicle kinetic energy was 90 percent of the change in fly-
    wheel energy.  This can be seen by noting in Table C-l that the engine
    supplied 1576 horsepower-seconds or 41 percent of the input energy required.
    For perfect TKE control, the engine must supply the total energy input minus
    the vehicle change in kinetic energy or 3852-2047 = 1805 horsepower seconds.
    Similarly, the flywheel change in kinetic energy was 2276 horsepower-seconds,
    while the vehicle change in kinetic energy was only 2047 horsepower seconds.
                                    C-37

-------
Lo
OO
*ouu
2400
2200
2000
1800
1600
•o
c
o
0
% 1400
01
a 1200
n
i-i
o
x 1000
800
600
400
200
0
ill • T '
Step Change in Driver Pedal at t = 0
Vehicle Weight = 5300 Ibs
' 07. Grade
Pierced Flywheel, P = 2.94 psi










_-=:









/
' -








/
'
	







/


	 '






/



	 "





/












Net Energy Out
of Flywheel — ^


X




X


X




X


x




' /







x











X



/







^












^


.^^•k"
w










•
/
	 -











/




/


- Engine
Energy















/












/













0246      8     10    12    14    16     18    20   22

                                              Time,  Seconds
24    26     28    30    32    34
                                                                                                                       36
                                  Fig.  C-22   Change  in  Energy Produced by  Engine and Flywheel
                                              During  Full  Power Transient
                                                                                                                      HTI-12501

-------
                        TABLE C-I
Energy Balance for Full-Power Transient From 8 to 80 mph
Description
Change in Vehicle Kinetic Energy
Change in Engine Kinetic Energy
Flywheel Losses
Road Resistance & Air Drag Losses
Accessory Losses
Transmission & Rear End
(power train) Losses

Change in Flywheel Kinetic Energy
Change in Engine Energy

HP-Sec
2047
7
76
798
449

475
3852
2276
1576
3852
Percent
53
ft* 0
2
21
12

12
1007o
59
41
100%
                         C-39

-------
      These  results  indicate  the  system approximated TKE control fairly close.
      Further adjustments in  control should provide somewhat better performance.
 4.    The combined energy from  the  flywheel and engine was 3852 horsepower sec-
      onds (800 watt-hours) with  the flywheel supplying 59 percent of the energy
      required and the engine furnishing  the remaining 41 percent.  These results
      imply  that  for  level road operation, it would have been possible to reduce
      the engine  size roughly 50  percent, with some margin, if it were not neces-
      sary to meet EPA/AAPS steady-state  design goals for climbing a five percent
      grade  as discussed earlier.

 Consider now the resultant transmission  efficiency as a function of vehicle
 speed during the full-power  transient.   Figure C-23 presents these results from
 10 to 80 MPH.  The lowest efficiency, 78 percent, occurred at approximately
 25 MPH where the primary transmission shifts range.  Efficiencies above 90 per-
 cent were obtained at higher vehicle speeds.  As noted earlier, the average
 efficiency  during the full-power transient was 89.5 percent.

 Before presenting a comparison in performance to a conventional system employing
 a conventional automatic transmission, the method used to calculate the perfor-
 mance of the conventional system under full-power conditions is summarized as
 follows.  The engine, defined earlier by Figure C-4, was assumed to operate at
maximum horsepower.  The automatic transmission under full-power conditions was
 in 1st gear from 0 to 25 MPH, 2nd gear from 30 to 55 MPH, and 3rd gear from 60
 to 85 MPH.  Engine accessory losses were as defined earlier by Figure C-5 and
 included air conditioning.  The resulting wheel horsepower developed as a func-
 tion of steady-state vehicle speed was approximately equivalent to that obtained
 for the flywheel/hybrid system during the full-power transient.

The resultant comparison of transmission efficiencies as a function of vehicle
speed is presented by Figure C-24.  These results indicate that the conventional
 transmission had better full-power efficiency than the flywheel/hybrid trans-
mission at speeds between 15 and 30 MPH and at speeds over 65 MPH.   For example,
at 25 MPH the hybrid transmission efficiency was 13 percent lower than the effi-
ciency of a conventional transmission.  This lower efficiency,  which occurs at
the range switching point,  is characteristic of power-splitting transmission and
is similar to the gear-shift points of a conventional transmission.
                                    C-40

-------
100
     Vehicle Weight = 5300  Ibs
     07, Grade
     Engine Accessories Include Air Conditioner
 80
                                                  Full Power  Transient

                                                                                             -t-t-t- -f-f J—
                                                                                               •   TH r
                           •
                                       Transmission Efficiency (Acceleration)
                                          .... i  ,  -    I  j-H-f
                                                 f-f-^-J  - M
                                                     40         50
                                                  Vehicle Speed, MPH

                  Fig. C-23   Flywheel/Hybrid  Transmission Efficiency at   Maximum  Power
                                                                                                                 MTI-12544

-------
         100
n
i
                                                        Conventional  Automatic Transmission (Based  Upon

                                                        Steady State  Operation at Maximum Power)
                                  Flywheel/Hybrid Transmission (Based Upon  Full
                                 Power Transient with Pierced Flywheel, P   =  2.94 psi)
                                                                           |  Vehicle Weight =  5300 Ibs

                                                                         - . — 0% Grade

                                                                             Engine Accessories  Include Air Conditioner
                                                                           Transmission Efficiency (Acceleration)
                                                                      r-f-J-r-'l-1~f t-f-44-f- -r-f-4-r -T~j H-i—H-| ••••"! -,--•
                                                                      U_Li_L4_U f-14-f ---^4- f •  ! I I l-i-4 +4 - '- J-f-* L
                                                                      pm i t-u; rti [... n T. TillU-lt-I - . -. - -H-..
                                                                      I i I I T i  : 1 1 1 ! ' . TTTl  I ! I i   n I . . •  • I i !
                                                             40           50


                                                           Vehicle Speed,  MPH
                                 Fig. C-24   Comparison of  Transmission Efficiencies  at Full  Power
                                                                                                                           MTI-12541

-------
At  lower  pow e r  levels,  " an ge  s wi t c h In g  '6 c  c u r s   at
lower vehicle speeds.  For example, ^ader cruise power conditions range switch-
ing occurs at 11 mph.  Thus, under normal driving cycle conditions and reduced
power levels the lower efficiency at range switching should be comparable  to
the efficiency of SL conventional  transmission.

A more important performance comparison is fuel economy.  Figure C-25 presents
this comparison.  At full power,  the advantage of the flywheel/hybrid system is
maximized and this is reflected in the dramatic improvement in miles/gallon
shown by Figure C-25.  For example, at 25 mph vehicle speed the hybrid system
fuel economy was approximately 10 MPG compared to 2 MPG for the conventional
system — an improvement of 400 percent even with somewhat lower transmission
efficiency.

Before discussing a comparison of  engine operation, it is helpful to consider
some of the operational aspects of the flywheel/hybrid transmission in more
detail during a full-power acceleration from a low vehicle speed.  A step
change in the driver's command pedal, initiates the ratio control in the
secondary transmission to decrease,flywheel speed :and the ratio control in the
primary transmission moves to give the maximum possible ratio (torque multi-
plication).  Since engine fuel flow is scheduled to produce little engine
power at low speeds (in order to approximate TKE control), the engine fuel
flow is only increasing gradually as a result of the step change in driver pedal.
                                                          *
Recalling that the engine and flywheel are linked together  by the secondary
transmission ratio control (see Section VI-B), this results in a rapid increase
in engine speed until the error signal in engine speed sensed by the primary
control system starts the primary ratio to decrease.  This can be seen by
referring to Figure C-26 which presents the engine speed transient during  the
full power transient.

Prior to the transient, the engine was operating at a steady-state speed of
1400 rpm (selected idle during low speed cruise conditions in order to supply
flywheel charging capability).  As a result of the step change in pedal,  the
engine speed rapidly increased to 3500 rpm and was subsequently regulated by
changing the ratio in the primary transmission at approximately 3200 rpm.   Thus,
*This is done to allow flywheel charging as a result of deviations from TKE due
 to system losses during deceleration or application of mechanical braking.
                                    C-43

-------
20
18
Vehicle Weight = 5300 Ibs
07. Grade
Engine Accessories Include Air Conditioner
Fuel Density = 6.152 Ib/gal.
16                            -4-
                     ,

14
        T"
                                            r  Flywheel/Hybrid Transmission  (Based Upon  Full
                                            /Power Transient with Pierced  Flywheel, P   =2.94  psi)
                                              'Conventional Automatic Transmission (Based Upon
                                               Steady State Operation at Maximum Power)
                                                                              i I '• !    I      i   '!"'..
                                                                       I I  I
                                                                                                   m
              10
                    20
30          40          50
         Vehicle Speed, MPH
                                                                         60
                                                                                70
80
90
                           Fig. C-25    Comparison of Full  Power Fuel Economy
                                                                                                           tfTI-12543

-------
n

Ul
                                                                              Step Change  in Driver Pedal at  t
                                                                              Vehicle Weight = 5300 Ibs
                                                                              0% Grade
                                                                              Pierced Flywheel, P  =  2.94 psi
= 0
           024     6     8     10    12    14    16    18    20    22    24    26    28    30    32    34    36

                                                          Time, Seconds
                                 Fig. C-26   Engine Speed Transient During Full Powor Transient

-------
 during most of  the  full power  transient,  the engine was operating at approxi-
 mately a  constant speed of  3200 RPM.

 As a result, during the initial portions  of the full-power transient (at very
 low engine power) the engine operated  for about 3 seconds at a specific fuel
 consumption which was about 50 percent higher than a conventional system at full
 power.  These results indicate that further improvements in the transmission and
 control design will be necessary in order to obtain engine operation at a better
 SFC during maximum acceleration.  At lower acceleration rates (gradual changes
 in the driver pedal), more common to the  DREW driving cycle, the engine would
 tend to operate close to the  minimum SFC operating conditions.  As an example,
with output power equivalent to a five percent grade, engine operation at mini-
mum SFC was obtained for all vehicle speeds above 25 MPH.

As mentioned earlier, emission data were  not available for the engine specified
by EPA/AAPS.  Thus, at this time no conclusion can be made with respect to the
possible  reduction in emissions for the flywheel/hybrid propulsion system
during transients with this transmission.

4.  Flywheel Charging Analysis

As part of the subject investigation it was of significance to determine what
magnitude of charging times would be expected for the system.  In order to
accomplish this, it was assumed that there was a fixed ratio between the engine
speed and the flywheel speed.  This would imply that the hydraulic elements
were not  taking part in the charging cycle and that the charging was directed
through the gearing with the ring gear held stationary.  The inertia of the
flywheel was referred to engine speed and the torque available for acceleration
was determined from the engine performance map which related horsepower and
speed.  Knowing torque and reflected inertia, the average rate of accelerating
could be  determined.  The acceleration and the speed increment were then used
to establish the time required between speed steps.
                                    C-46

-------
The subject calculation was made for charging at minimum SFC as well as charging

at maximum engine horsepower.   Two end points of engine speed were also consid-
ered, 3200 and 380C rpm.  The  key results are summarized below.
                                   TABLE C-2
                            Flywheel Charging Time
    Condition

    Min.  SFC HP

    Min.  SFC HP

    Max.  HP

    Max.  HP
Engine
Speed
RPM

 3200

 3800

 3200

 3800
Time To
Reach Idle
Sec.	

   2.2

   2.2

   2.2

   2.2
Time-
Idle To
Full
Speed Sec,

   42.7

   37.7

   25.5

   22.9
Total
Charge
Time Sec.

  44.9

  39.9

  27.7

  25.1
                                    C-47

-------
                                                             o
                                                             o
                                                             3
         SECTION D



CONTROLS DESIGN AND ANALYSIS

-------
                        D.  CONTROLS DESIGN AND ANALYSIS

This section presents:  1) an overview of the basic control approach; 2) a review
of the constant total energy control constraint; 3) a discussion of the two
control approaches considered; and 4) a detailed description of the implementa-
tion and operation of the selected control approach.  Additional detailed analy-
sis  of-the1 control'-'system is  contained  in 'Appendix II ViiS't'ab^lity' andi Analog. •
.Computer  Simulation  Analyses.

1.  Basic Control Approach
The basic interrelationships of the engine, flywheel, transmission and vehicle
are indicated in the block diagram of Figure D-l.  The power paths of the engine,
flywheel and vehicle are  interconnected through the transmission.  Allocating
and directing of system power flow is established by the controls, which accept
acceleration, deceleration and modal commands from the operator; engine, fly-
wheel and output speed signals from the system; and provide throttle and ratio
commands to the engine and transmission,  respectively.

The overall transmission  consists of two  separately controlled split-path
hydrostatic links:   the Primary path, which establishes a given ratio between
the engine and vehicle for optimum torque-speed loading condition of the engine
in the steady-state; and  the Secondary path, which controls the direction and
magnitude of power flow to and from the flywheel during vehicle velocity
transients.

2.  Constant Total Energy Constraint
As pointed out by LMSC (2 ), it is advantageous to maintain the flywheel charge
by following the requirement that the total kinetic energy (TKE) of the system
remains constant.   In other words, the kinetic energy of the flywheel plus the
kinetic energy of the vehicle equals a constant.  This relationship is shown
graphically in Figure D-2.
                                     '! -I

-------
                                                      FLYWHEEL
ENGINE
                                    Engine
                                    Output
*
                    Throttle
                    Command
                                                               Flywheel
                                                                Output
                                                Transmission
                                                   Output
TRANSMISSION
}
                     Primary
                  Transmission
                     Conmaod
                                          Engine
                                         _Speed
N>
                      Accel •  Pedal
                      Selector T-*
                      Brake
                         1
                                          I

                              Secondary   '
                            Transmission  |
                               Command    I
                                   r—	'
                                   Output
                                I   Speed
                                                       CONTROLS
^
                                                                 Vehicle
                                                                 Velocity
                                                                                              VEHICLE
                                             Flywheel
                                              Speed
                                                     —— ^— -^ Link Dependent on Detailed  System Form
                                    Fig.  D-l    System  and Controls  Interrelationships

-------
      MAX
 u
 o
 r-l
 
-------
As  indicated  by  the  locus  curve  of Figure  D-2,  the maintenance  of  a  constant
TKE  level  implies  that  the  flywheel velocity ( hence,  the  "state  of charge")
will be maximum when  the vehicle  is stopped, and at its specified  minimum when
the  vehicle is at  its specified maximum velocity.  Considered as an  energy
storage element, the  flywheel has the highest charge  level when  the  vehicle is
in most need  of energy  for  acceleration, and conversely is in best condition
to accept energy from the  vehicle for deceleration when the vehicle  velocity
(hence, charge) is highest.

To the extent that a constant TKE value is maintained on  an instantaneous basis,
the energy utilization  and  storage action  of the system is constant  and indepen-
dent of load  conditions and immediate past history.

In an ideal case where  the  road is horizontal and all losses (including those
due  to the transmission and dissipative braking) are zero, total system kinetic
energy would  inherently be  preserved.  Energy would be transferred back and
forth between the  flywheel and vehicle, and only the rate of transfer  (hence,
accel-decel rate of the vehicle) would be  of concern.

In the real case,  observation of  the TKE constraint is complicated by  factors
such as:
   (1)  hills — which can  introduce significant unscheduled system kinetic
        energy variations due to potential energy variations,
   (2)  dissipative braking - which can introduce rapid and significant
        unscheduled changes in system TKE  level, and
   (3)  variable loading — as by the number of passengers and/or trunk load
        carried, or in pulling trailing vehicles, which can significantly
        alter the value of vehicle mass used in computing and implementing
        the TKE control scheme.

3.  Control Approaches Considered
Two basic  approaches to maintaining constant TKE were considered:  one involving
the measuring and utilization of vehicle velocity in an attempt to maintain
system TKE by closed-loop feedback control techniques; and the  second using an
approximated value  for vehicle velocity based upon driver pedal position in

                                    D-4

-------
an  open  loop  scheduled manner.

Functional block diagrams  for  the  two basic  types of TKE control are  shown  in
Figures  D-3 and D-4,  respectively.

The major basic difference between  the  two control configurations  is  the method
of  deriving a vehicle velocity  value for controlling system  total  kinetic energy.
The closed-loop (Type A) control of Figure D-3 uses a speed  governor  for measur-
ing output shaft: (hence, vehicle) speed.  In  the open-loop scheduled  (Type  B)
control, output speed is approximated from pedal position by a '-cam-generated
schedule.

The Type A configuration (Figure D-3) constitutes closed-loop control of TKE
since actual vehicle  velocity  feedback  is used, allowing a continuous comparison
of  flywheel and vehicle velocities  to be made.

                                                         2            2
Due to the nature of  the TKE computation (TKE = 1/2 Jp Nf  + 1/2 MV v ) , the
vehicle  speed feedback is  positive  (regenerative) in that an increase in
vehicle  speed commands a decrease in flywheel speed, which in turn results  in
a further increase in vehicle speed.

It  is important to point out that whenever wheel slip occurs, such as vehicle
operation on road ice, these regenerative control effects could result  in
difficult driver control situations.  Also, with this type of control approach,
a destabilizing feel can occur when climbing or descending hills.  In other
words, as the vehicle proceeds down a hill, the resultant increase in vehicle
velocity commands the flywheel to accelerate the vehicle.  This situation
would be disturbing to the driver.

Although control systems can be implemented with positive feedback loops, the
magnitude of gain must be carefully restricted to be less than unity  to prevent
regenerative behavior.  If gain is equal to or greater than  unity, the  system
will be unstable in that, once the output is started in motion,it will  continue
to move until system saturation occurs,  even without an input to the control
loop.   If the gain is too  low, system transient response will be too slow and
general performance poor.  Thus, in principle there is little range for adjusting

                                     D-5

-------
            RATE
            VALVE
     HRAKE
Accelerator
   Pedal
                           SECONDARY
                            CONTROL
                                        SECONDARY
                                      TRANSMISSION
                   KATE
                   VALVE
                                       v
                              ENGINE
                                     ENGINE
                                    GOVERNOR
                                                             OUTPUT
                                                           GOVERNOR
                                                        PRIMARY
                                                      TRANSMISSION
                                                            PRIMARY
                                                            CONTROL
                  Fig.  D-3   Closed-Loop  TKE  Control (Type A)
BRA£E
PEDAL
                               FLYWHEEL
                               GOVERNOR
                                           SECONDARY
                                         TRANSMISSION
  ACCELERATOR
     PEDAL
                                    ENGINE
                                   GOVERNOR
                                                            PRIMARY
                                                         TRANSMISSION
                                                            PRIMARY
                                                            CONTROL
              Fig.  D-4   Open-Loop  Scheduled TKE Control  (Type  B)
                                   D-6
                                                                              MTI-12527

-------
 the gain  for  performance  considerations,  from an  implementation  standpoint,
 gain variations  can  result  in  inadvertafit regeneration with a  resultant  loss  of
 control.

 In  the Type A TKE control,  the  sensitivity  to gain normally experienced  with
 positive  feedback is complicated by  the  fact that  the feedback velocity  function
 is  actually the  square  of velocity as required to  follow  the kinetic energy
 equation.  This  means  that  the  effective feedback  gain increases  with vehicle
 velocity  (is  proportional to it) and,for stability,  the gain must be restricted
 to  less than  unity at  the highest velocity  and allowed to degrade at lower
 velocities -  resulting  in significant variations in  response with velocity.

 As  indicated by  the  curve of Figure  D-2, the TKE equation as implemented by
 the Type A approach  merely  prescribes the locus of flywheel and vehicle  velocity
 pairs for a given constant  TKE  value.  The  rate valves shown in Figure D-3 are
 required  to provide  transient  signals to the secondary (flywheel) transmission
 to  establish and direct the rate of  energy  flow to and from the flywheel for
 vehicle acceleration and dynamic braking.

 With the Type B  open-loop control approach  (Figure D-4) the vehicle "kinetic
 energy" input for the TKE control function  has the form of a well-defined
 command rather than  an  instantaneous feedback signal — eliminating the positive
 feedback effect  inherent in the Type A approach.   System  gains can therefore
 be  varied over a wider  range to optimize performance and  stability; and  normal
 gain variations will not result in instability or  loss of control.  Also, the
 driver pedal  tends to act like  a torque command for  the propulsion system
 which is analogous to present automotive systems.

 As  previously mentioned, the regenerative effect of  the Type A approach  would
 cause a destabilizing feel  to hill climbing and descending; however, with the
 Type B control hills would not  result in the flywheel influencing vehicle .speed
 unless the driver commands it to.

As shown in Figure D-4, the Type B control  is programmed  by a^cariKdrdverayby the
 accelerator pedal to provide a signal proportional to vehicle kinetic energy
at specific values of grade, vehicle mass,  wind load, friction, and engine
                                     D-7

-------
condition.  A change in any of these pardiucters could cause the vehicle speed
and  thus  the vehicle kinetic energy to be different from the programmed value.
Thus, some compromise or selection of nominal condition is required in order
to establish the optimum open loop schedule.  This is illustrated by Figufe D-5
where the schedule was selected to provide constant TKE at zero grade.  Operation
at five percent grade conditions then results in lower flywheel speeds at the
same vehicle velocity under steady-state conditions.

Figure D-6 shows the possible variation of TKE at different vehicle speeds due
to the effect of changing vehicle mass and grade conditions.  These results
clearly indicate that as the load torque increases, the driver will have  less
available torque to accelerate the vehicle.  A situation similar to that
encountered with conventional propulsion systems.

Figure D-7 shows the change in flywheel speed which occurs in going up or down
a hill, at constant vehicle speed with the Type B control.  Flywheel speed
decreases when going up a hill, thus providing some power for climbing the hill.
Flywheel  speed increases when going down a hill, thus transferring part of the
vehicle potential to the flywheel.  One of the significant advantages of  the
Type B control configuration is the use of flywheel power to aid in climbing
short hills.  Of course, the total kinetic energy will, not remain constant
during hill climbing.  Upon reaching the top of a hill, the flywheel speed will
be low.  The accelerator pedal must be depressed farther, at a given vehicle
speed, to provide power for accelerating the flywheel.  After the flywheel is
charged,  the accelerator pedal can be returned to the normal position.

Upon reaching the bottom of a hill, the flywheel speed will be high.  Therefore,
dynamic braking cannot be used until the flywheel returns to the normal speed.
If a stop must be made at the bottom of a hill, the brake pedal must be depressed
far enough to engage the mechanical brakes.  There will be a difference in brake
feel under this condition.   It should be possible to maintain constant brake
pedal feeling  by adding compensation to the brake pedal.  However, this would
add to the control complexity.   Upon coming to a stop, flywheel speed will be
at the correct value.  An overspeed limiter prevents the flywheel from acceler-
ating to a higher speed.
                                     D-8

-------
   25,000
   20,000
TJ
01
0)
a
CO
01
01
   15,000
   10,000
()I Grade  (Constant
 Total  Kinetic Energy)
                                                                     100
                        Vehicle Velocity,  Miles Per Hour


               Fig. D-5   Flywheel Speed  Versua Vehicle  Velocity
                          Type  B Open Valve Control
                                     D-9

-------
 1.2
 1.0
0.8
"O
01

•H
m
01
O
0.6
0.4
0.2
  0
        10%  Increase  in Vehicle  Mass  at 0% Grade
                                               •Minimum Flywheel
                                                     Speed
                              Note:  Changes  in wind  load,  fric-

                                     tion, and engine  conditions
                                     have effect similar  to
                                     change in grade.
               20           40          60

                    Vehicle Speed, Miles Per Hour
                                                       80
100
      Fig.  D-6
                    Effect  on  TKE  of  Varying Vehicle Mass and

                    Grade Conditions  - Type  B Open-Loop Control
                          D-10
                                                                     MTI-12499

-------
ALTITUDE
           HILL
STEEPER
  HILL
ACCELERATOR
PEDAL
/V
                                                                  LJ
BRAKE
PECAL
                                                BRAKE
                                                APPLIED
VEHICLE
VELOCITY
                               CONSTANT VEHICLE SPEED
FLYWHEEL
SPEED
                                        TIME
        Fig. D-7   Effect of Hills on Flywheel Speed at Constant Vehicle Speed
                                         D-ll

-------
On a comparative basis, although the Type * TKE control is initially attractive
because it would monitor vehicle velocity and would therefore utilize an accurately
computed value of kinetic energy, the Type B approach was selected for final
design evaluation and costing on the following major bases:
   (1)  Control System Complexity and Cost
   The Type B approach is inherently less costly by virtue of eliminating the
   output speed sensor and the brake and accelerator rate valves.  Although the
   detailed dynamic signal shaping techniques were not finalized for either
   Type A or Type B approaches, within the scope of the study, the Type B
   requirements appear to be-.basically, simpler  because  of the inherent  better
   stability of that  approach.

   (2)  General Performance and Feel
  Assuming  the  Type A approach is compensated appropriately  to achieve acceptable
  stability,  the steady-state TKE performance of  the Type A  and  B  systems would
  be  similar  for horizontal driving conditions.  As indicated previously,  the
  Type A approach  increases driver control requirements  on hills and icy roads,
  and  it eliminates the  possibility of using the  flywheel  (assuming no TKE
  error)   either  to provide energy in climbing or to store  energy on descending
  a hill.  At  some departure from TKE, the Type B approach provides the capability
  for  the flywheel to provide energy during hill climbing and to store energy
  (as desired) when on descending hills.  Also, the Type B control tends to make
  the driver pedal a torque command to the system.
                                    D-12

-------
4.  Implementation and Operation of  the Type B Open-Loop Control System
General Description
A simplified system diagram is illustrated in Figure D-8.  The inputs  to  the
system are the same as those of the  typical family car; selector lever position
(plus a position for initially charging the flywheel), accelerator pedal  position,
and brake pedal position.  The control variables are the hear, engine throttle
position, and the continuously variable ratios of the  two transmissions.  Feed-
back signals are engine speed and flywheel speed.

Acceleration
A step change in accelerator pedal position in the direction to increase  vehicle
speed results in three inputs to the control system.   The first input acts
through the primary control system to command a new steady-state engine speed.
The second input acts through the secondary control system to effectively vary
the secondary transmission ratio between the flywheel  and engine speed.  Fuel
flow to the engine is modulated by the third input providing engine power
approximately equal to the steady-state losses at each speed.

Varying the secondary transmission ratio results in a  decelerating torque being
applied to the flywheel and an accelerating torque applied to the engine  shaft.
The engine speed increases to the commanded steady-state value and is held at
this speed by the primary transmission ratio being varied with changes in load
through action of the primary transmission controls.

Kinetic energy is transferred from the flywheel to the vehicle with vehicle
acceleration resulting.

Dynamic compensation is used to maintain the correct sequence and rates between
the control functions.

Deceleration
When the accelerator pedal is returned to the idle position at high vehicle
speed, the vehicle speed decreases slowly in the same manner as a conventional
vehicle.  Depressing the brake pedal part way varies the secondary transmission
ratio to transfer kinetic energy from the vehicle to the flywheel — thereby
using dynamic braking for decelerating the vehicle.   Depressing the brake pedal
farther engages  the mechanical brakes for faster deceleration.
                                     D-13

-------
                  Mechanical
                  Brake
     Flywheel
     Governor
 Flywheel
Brake Pedal
      Accelerator
      Pedal
  Selector Lever
   Secondary
   Control
Secondary
Transmission
                  j—»• CAM
v2  (scheduled)

—•*
                                  Engine
                                      Engine
                                      Governor
   Modal
   Control
                Primary
                Transmissior
                                                            -ft
                                     1
                                                               Primary
                                                               Control
                     Fig.  D-8   Basic  System Functional  Diagram
                                          D-14

-------
Fast Stop
The reduction of engine speed when the vehicle is brought to a stop results in
disengagement of the primary transmission.  The secondary transmission normally
remains engaged.

The flywheel speed will be approximately correct when the vehicle stops due to
dynamic braking only.  After a fast stop in which the mechanical brakes are
applied, the flywheel speed will be low.  The secondary control:system then
causes the flywheel to slowly accelerate to the correct speed at a rate within
the power capability of the engine at idle.

Total Kinetic Energy Behavior
The secondary control system is programmed to vary the flywheel speed to main-
tain the total system kinetic energy (flywheel plus vehicle) at a constant value
for zero grade nominal load conditions.  Variations of grade, load and engine
condition can cause variation of the total kinetic energy.

Implementation and Operation
A control component diagram and hardware implementation schematic are shown in
Figures D-9 and D-10.

The manual shift lever is used to select the operating mode.  The shift pattern
schematic as shown below follows the present day automobile shift pattern as
closely as practical.

5 R C I
1 i i
1 i 1
« D 2 1
P - Park
R - Reverse
C - Charge
N - Neutral
D - Drive
2 - Second
1 - Low
                                     D-15

-------
                           SECONDARY CONTROLS
O
i
                                                        Secondary
                                                        Cam Plates
                                                                                                                Mechanlca1
                                                                                                                Path
                                                                                                                Hydraulic
                                                                                                                Path
                                                                                        Primary
                                                                                        Transmission
1
i
»

Engine
Engage
Governor





























High
Range
Clutch
Low Range
Clutch
Accel.
Pedal




Engine
Dynamic
Compensator
-\
\
«

Flywheel
Charge
Piston














\
Control
Cam
*


Cover
Cam
1


Engage
Valve
i
h. f
fcj

, 	 1
i
»


Primary
Piston
Actuator
i
i

Regulator
Governor
i
*

1
Control
Cam


\
^



Governor
Cam
                                                                                                    PRIMARY CONTROLS
                                                                                                                           W
                              Fig.' D-9   Flywheel/Hybrid Transmission — Control  Component Diagram
                                                                                                                           MTI-12536

-------
o
I
                              Fig. D-10   Control System - Hardware Implementation Schematic
                                                                                                                    KII-12522

-------
 Control  Valves  1  and  2,  which  are  positioned by the  shift lever,  direct pressures
 to the  control  elements  as  required in'each  mode.

 In the  charge mode,  the  primary transmission and high-range clutch are disen-
 gaged.   The  kinetic  energy  summing valve provides  a  pressure proportional to
 kinetic  energy  error  which  acts on the  flywheel charge  piston to  move the
 throttle.  Thus,  the  engine speeds up to charge the  flywheel.   When the flywheel
 is fully charged,  kinetic energy error  is zero  and  the engine  slows down to
 idle  speed.  The  secondary  engage  valve is actuated  at  an engine  speed of 900
 to 1100  rpm, providing pressure to hydrostatic elements III and IV, thus en-
 gaging  the secondary  transmission.   The secondary  transmission  remains engaged
 for all  shift lever  positions  as long as engine speed  is above  approximately
 900 rpm.

 In neutral,  pressure  is  directed by Control  Valve  1  to  the  primary piston actua-
 tor to move  the sleeve to the  position  which limits  piston  stroke.  Pressure is
 transmitted  through  the  shuttle valve to engage the  low-range brake.  The engage
 valve is in  the open  position,  and the  swashplate  of Element I  is near zero dis-
 placement 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 re-
 sultant  position  of  the  governor valve  provides flow to the primary actuator
 which moves  the piston to the maximum torque position C.  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.

 When  the control valves  are  shifted to  the drive position,  the  fluid pressure
 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 Control Valve 2  to  the  primary  and secondary engage valves.   The  output
 pressure from the engage governor  increases  as  engine speed increases.   This
 pressure acts on the primary and secondary engage  valves  to engage  the  secondary
 transmission at 900 to 1100 rpm  engine  speed   and  to engage  the primary  trans-
mission at 1200 to 1400 rpm engine  speed.
                                     D-18

-------
During drive operation the engine xemulator governor controls hydraulic  flow
to the opposite sides of the primary actuator piston to effect movement  of
the Element I swashplate.  The movement of the governor valve is regulated  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 primary
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 for  mini-
mizing emissions.

It will be noted that in initially accelerating, the clutch and brake valve
rides upon the upper surface of the control cam, thereby providing fluid
pressure, through the shuttle valve, to engage the low-range brake.  When the
primary transmission output speed increases, movement of the cam causes  the
clutch and brake 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,
in addition to non-power shift of both the clutch and brake, provides wear-free
operation.

In Modes 2 and 1, the control cam introduces a bias to the engine regulator
governor which varies the engine speed setpoint at which the primary  transmission
switches from low to high range.

In reverse, fluid pressure is transmitted from Control Valve 1, 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 gover-
nor spring  thus moving the engine regulator governor valve downward,  and pro-
viding fluid pressure to the bottom of the primary piston actuator.  Pressure
in the top chamber of the primary actuator is vented through Control Valve  1.
Thus, the piston and sleeve of the primary actuator both move to the  top
position.  The additional stroke of the cam plate due to movement of  the
                                    D-19

-------
 sleeve  causes  the  pin  connected  with  the  swashplate  housing to move  into the
 negative  angle  portion of  the  cam slot,  thereby positioning Element  I swashplate
 at a  negative angle  for reverse  output.

 It will be noted that,  when Control Valve  I  is  moved to  the reverse  position,
 fluid pressure  is  directed to  the  primary  engage  valve so  as to engage  the
 primary hydrostatic  transmission.

 Park, which opens  the  engage valve  to  bypass  fluid around  the  hydraulic  motor
 and pump;(Elements I and II) permits  the engine to idle  without transmitting
 any torque to the  transmission output  shaft.  The engine idlds  below  1200 rpm.
 In addition the parking lock is  positioned to lock the transmission  output  shaft
 to the rear wheels.

 The secondary transmission and control system continuously vary the  ratio
 between the engine and  flywheel  to maintain a nominal constant  total  kinetic
 energy.   The kinetic energy summing valve  provides an output pressure  propor-
 tional to kinetic energy error,  which  acts on the secondary piston actuator to
 vary the  swashplate angles of hydrostatic Elements III and IV,  thus varying
 the ratio.

Each position of the accelerator pedal, at constant  grade  and  road load,
corresponds to a steady-state vehicle  velocity,  and  therefore  to a scheduled
vehicle kinetic energy.  A cam driven by the accelerator pedal  varies  the
 load on the kinetic energy summing valve spring in proportion  to the  commanded
value of vehicle kinetic energy.  The spring preload is proportional  to  the
desired value of total  kinetic energy.  The flywheel governor,  which  is  driven
by the flywheel through a 6 to 1 reduction ratio, provides  a force proportional
to flywheel kinetic energy.  This force is applied to the  kinetic energy
summing valve through a coupling spring.

At zero accelerator pedal position, the kinetic  energy summing  valve  spring
applies the maximum load to the  flywheel speed  governor valve.  A cam causes
this spring force to decrease as the accelerator pedal is  depressed.   At any
accelerator pedal position, if the flywheel speed is too low, the kinetic energy
summing valve spring forces the governor valve  downward.    High  pressure  is then
                                   D-20

-------
directed to the bottom of the secondary piston actuator, and the top is vented.
The secondary piston actuator then moves the swashplates in the direction  to
increase the ratio of flywheel to engine speed, thus increasing flywheel speed.
When the flywheel speed reaches the desired value, the  flywheel governor force
equals the kinetic energy summing valve spring force.   The governor valve  then
moves to the neutral position and the actuator motion ceases.  If flywheel
speed is too high, the governor flyweight force moves the governor valve
upward, thus providing flow to move the actuator as required to decrease
flywheel speed.

Pressure feedback from the secondary actuator to the kinetic energy summing
valve is used to reduce the phase lag of the actuator as required for stable
operation.  The pressure feedback also reduces the effect of engine speed  on
the secondary control system.  A further advantage of the pressure feedback
is that torque applied to the flywheel becomes proportional to kinetic energy
error.

The flywheel dynamic compensator is used to limit the secondary piston actuator
velocity as required to obtain fast acceleration without saturating the secondary
pressure.

The hydrostatic transmission relief valves, which are located in the primary
and secondary engage valve housings, are set at approximately 3600 psi.  The
control system should hold the pressures below 3000 psi during normal operation
for maximum efficiency and minimum wear.  Pressure surges are held below the
relief valve settings, during normal operation, by the primary actuator rate
limits, and by the engine dynamic compensator and flywheel dynamic compensator.
                                    D-21

-------
                                                      o
                                                      o
  SECTION E



COST ANALYSIS

-------
                              E.  COST ANALYSIS

Described in this section is the method employed to determine transmission costs
and a discussion of the results is included.   The cost data presented was based
on the experience of automotive cost consultants based upon cost information and
practices of the Ford Motor Company.  Therefore, this procedure provides  a sound
approach to comparing the cost of the flywheel-hybrid transmission with that of
the standard multispeed  torque converter (automatic) transmission.

The objectives  of the cost analysis were to determine the original equipment
manufacturercost (O.E.M)  for production quantities of 100,000 and 1,000,000
units per year  of the flywheel-hybrid transmission and then compare that  cost to
similar costs for a  multispeed (automatic) transmission.

Feasible Transmission Concept Cost
During the initial feasibility analysis, seven transmission concepts were iden-
tified as possible candidates and sketch layouts were drawn.  Estimates of the
O.E.M. costs for these transmission concepts  were made and referenced against
the automatic transmission used on current medium sized vehicles.   The results
are presented in Table E-l.

The power-splitting transmission was selected from the seven candidates as the
type of transmission required to accommodate  the dual power path propulsion
system.  Estimates of the final cost ratios were made and are presented in
Table E-2.

The method of establishing the information presented is discussed  in the  follow-
ing paragraphs.   Preliminary cost ratios did  not have the depth of analysis as
did the final ratios and  were, as a result, optimistic.  Since the same optimism
prevailed throughout the  analysis of the seven candidates,  the ratios of
Table E-l were  on a consistent basis and adequate for the purpose  	 that  of
assessing the cost category for the intial screening task.
                                    E-l

-------
                             TABLE  E-l

         Preliminary Transmission Cost Analysis - Ratios
                    1,000,000 Units Per Year
                           O.E.M.  Cost
                      100,000 Units Per  Year  '•
                            O.E.M.  Cost       !
SK-J-4258
3 Element
Hydrodynamic

SK-E-4260
2 Element
Hydrodynamic

SK-J-4257
Hydrostatic

SK-E-4256
Power Split

SK-E-4261
Power Split
(Offset)

SK-D-4259
Hydrodynamic
Hydrostatic

SK-A-4262
Hydrodynamic
2 Coupling

Standard
Multi-Speed
Torque Converter
("Automatic")
1.9
1.5
2.1
1.9
2.0
1.9
1.7
2.45
1.94
2.71
2.48
2.58
2.48
2.19
1.0
1.29
                                  K-2

-------
                        TABLE E-2    FINAL TRANSMISSION COST  ANALYSIS RATIOS
                        Standard
                        Multi-Speed Torque
                        Converter ("Automatic")
                        Transmission for
                        Medium Size Vehicle
Power
Splitting Transmission
for Flywheel/Heat
Engine Medium Size
Vehicle (Inline Flywheel)
Power Splitting
Transmission for
Flywheel/Heat Engine Medium
Size Vehicle (Offset Flywheel)

1) Variable Cost Ratio
2) O.E.M. Cost Ratio
3) Control Variable Cost Ratio
4) Labor Content Ratio
5) Material Content Ratio
A
1.00
1.00
1.00
1.00
1.00
B
1.19
1.15-1.25
1.19
1.00
1.20
C
1.29
1.25-1.35
1.29
1.50
1.20
A
2.40
2.25-2.35
2.87
2.13
2.65
B
2.85
2.70-2.80
3.41
2.13
3.18
C
3.10
2.95-3.05
3.70
3.19
3.18
A
2.53
2.35-2.45
2.87
2.25
2.79
B
3.01
2.85-2.95
3.41
2.25
3.35
C
3.26
3.10-3.20
3.70
3.37
3.35
A)  Ratios based  on  manufacture of 1,000,000 units per year

3)  Ratios based  on  manufacture of 100,000 units per year with tooling as for 1,000,000 units
c)  Ratios based on  manufacture of 100,000 units per year with tooling suitable for maximum yearly
    manufacture of  100,000 units.

-------
 Although not shown on the transmission concept  sketches,  cost  saving  design
 improvements suggested by the  automotive  consultants were factored  into  the
 finalized cost  ratios.

 Costing Procedures

 The  technique for  obtaining  the  costs  for the power-splitting  transmission is
 outlined and a  sample sheet  ia included as  Table E-3.   It was  necesary to re-
 view  a  total of  390  items  to obtain a  valid cost comparison.   The few examples
 presented on the sample  sheet  were, in most instances,  selected to  present the
 costs of the hydraulic components peculiar  to the power-splitting transmission.

 The package  drawing,  SK-E-4283 (Figure A-12, pp. A-22 to .A-2&), was;used to identify
 the items for costing.   Components, such  as the hydrostatic 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-3  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.

Column 4  presents  the method  of manufacturing the part  as established  by  the
 automotive company.

 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.

 The "in-house" manufacturing costs to  finish a specific part are  developed  on
 an extension  of variable minute costs  times labor minute content.   Variable
minute costs  include  direct labor, indirect labor and non-variable burden.

                                    E-4

-------
TABLE C-3   COST SAMPLE SHEET


DESCRIPTION

Shaft, Element I

Cylinder Block - Primary Transmission


Pl3t°" 	 — — ^ 	 — 	 anSm
Trunnion - Swashplate - Prl .

Support - Swashplate - Pri .

Swashplate - Motor- Pump - Prl .


Planet Gear - Fly, Planetary

RlnR Gear - Fly. Planetary


Gear - Element III - Fly. Planetary

Plnnptflry Aiiy 	 Flywheel 	
Governor - Eneaee - Flywheel

TKE Valve and Governor Assy







NOMENCLATURE











TOTAL PER ASS'Y REMARKS
ITEM
NO.

4

8



11

12

13


26

26


23


--

--




















QTY

i

2



1

1

2


3

1


1


1

1







P.P.
P.R.
P.S./F
M
A







MAKE
BUY

PR

PR



PR

PR

PF


M

PR


RR


PF

PF







Pur
Pur
Pur
Mar
As;







MAT'L
COST (D

1 .120

2.440



1 .510

.900

4.800


.810

1.860


.581


1.611

6.717







chased as
chased In
chased as
u fa c cured
emble








DLLARS)
































finished
rough cc
o semi-f
in house








LABOR
MIN.

9.70

18.00



12.00

7.02

	


5.70

8.30


6.30


10.25

22.50







(tern
ndltlon ,
Lnished 1












































such as c
tern









LABOR
COST
fVARUD

1 .693

3.140



2.094

1 .225

	


.995

1.448


I .099


1.789

3.933








istlng, o











1LLARS)

































d forRin










TOTAL
COST(DO

2.813

5.580



3.604

2. 125

4.800


1.805

3.308


1.679


3.400

10.650








s











.LARS)

FOFR

Mech



Nodu

Cast

Heav


Stl.

Heat


ForR



























nR - AIS

nite Cas



ar Iron

Iron

' Coined


Bar

Treat No


ng - AIS


























^ n0t SP
8620 St

Ing (She



H.T.) (P



itpR. for




lular Iro


[ 8620 St



























el

1 Mold)



a r 1. Ha 1 L



Blank pe







>el


























"
<

(



(

(

Torr. (


(

(


(


(

<






















ost Sht

ost Sht



ost Sht 1

ost Sht 1

ost Sht 1


ost Sht 2

;ost Sht


lost Sht


ost Sht7

xist Sht 1




















-------
In Column 6, the actual "in-house" number of labor minutes to complete the manu-
facturing task were listed and in Column 7 the variable minute costs were listed,

Column 8 is the total cost dollars for each item or task labeled in Column 2.

The study was based on variable costs and not O.E.M. costs.  The O.E.M.  or
transfer costs would include cost allocations for fixed burden, scrap, factory
cost adjustments, general and administrative costs, profit and capital invest-
ment.  Capital investment would include costs for facilities, tooling and engin-
eering expense.  Since many of these transfer costs would vary with different
automotive companies, the O.E.M. data were not as basic as the variable  cost
data, and; therefore, were not considered as reliable when comparing information
from different sources.

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 power-splitting transmissions could be the same.

As an example, let:

   1.0  =  Variable costs of automatic transmission
   2.0  =  Variable costs of power-splitting transmission
   0.1  =  Facilities costs for either transmission.

                                                       2.0
   Variable cost ratio without facilities included  =  -H-r  =  2.0

                                                       2.1
   O.E.M. or cost ratio with facilities included    =  T^T  =  1.9
                                                       J. • J.

Therefore, when dealing with ratios the increased cost of the basic transmission
is not correctly identified if only the O.E.M.  cost ratios were presented.

The cost analysis task dealt in ratios because the actual variable dollar cost
of an automatic transmission was automotive company proprietary information.
                                     E-7

-------
The range in the O.E.M. ratios provided in Table E-2 account for the possible
variances associated with the costs.

The cost estimates are presented in Tables E-l and E-2 as a ratio of hybrid
propulsion system transmission costs to conventional automatic transmission
costs for a medium-size car line.

All ratios represent the cost of the item being considered divided by the cost
of a multi-speed torque converter ("automatic") transmission for medium-size
vehicles as now produced in quantity by automotive manufacturers.

The detailed cost study was made for 1,000,000 units per year volume of both
standard automatic and power-splitting transmissions.  A second estimate for
the manufacture of 100,000 units per year volume was presented.  This estimate
assumes the same tooling and facilities as for the 1,000,000 units per year
volume.  Essentially it reflects the increase in purchase costs when the volume
is reduced.

A third estimate for the manufacture of 100,000 units per year volume was pre-
sented.  This estimate uses tooling and facilities suitable for the manufacture
of 100,000 units per year with no increase, in production predicted for the fu-
ture.  This number represents the increase in purchase material costs and labor
content costs.

As shown in Table E-2, the ratio of the variable costs for the power-splitting
to that of the multi-speed converter transmission was 2.40 — an increase of
140 per cent.   Primarily, two factors were responsible for the increased cost:

   1.  In reality, the power-splitting transmission was designed to accept
       power from two separate propulsion sources.  Such a feat required that
       the components of two transmissions be packaged as one.   Therefore, the
       costs reflect the major torque transfer elements of two transmissions.
                                     F.-8

-------
   2.  The control required to properly relate the heat engine,  flywheel,
       heat engine transmission and flywheel transmission has essentially
       two times the number of components as the standard transmission control.

Also in Table E-2 the control variable cost ratio has been presented.   The ratio
of 2.87 or an increase of 187 percent in costs reflects the complexity of  pro-
viding controls for such a propulsion system.

Finally, the labor content ratio and material  content ratio are  given  to es-
tablish the area of the major portion of the increased cost.   Since the power-
splitting type has the torque elements of two  transmissions,  the increase  in
material costs was the major contributor to the high cost ratio.

Included as part of Table E-2 are the cost ratios for a power-splitting trans-
mission with an offset flywheel.  The increased values reflect the additional
gear shafts needed to offset the flywheel.  Again the material content contrib-
uted the major portion of the cost increase.

In summary, the hybrid propulsion system transmission requires twice the num-
ber of torque transmitting components and the  control components resulting in
a cost of 2.40 times that of the standard automatic transmission.

One additional cost comparison was investigated.  This was a  determination of
the ratio of the costs of the smaller engine plus the inline  pierced flywheel
plus the power-splitting transmission divided  by the  costs  of the  large  engine plus
the standard automatic transmission.  Costs were based on the production of
1,000,000 units per year.  The ratio was 1.63.

If the non-pierced flywheel is used, the cost  ratio is 1.69.

Therefore, the hybrid propulsion system would  cost approximately 60 to 70
percent more than the system presently installed in automobiles.
                                     E-9

-------
   SECTION F
SAFETY ANALYSIS
                                                        9!.

-------
                              F.  SAFETY ANALYSIS

The approach for evaluating potential safety problems with the heat engine/fly-
wheel hybrid transmission was to look initially at the main assemblies and how
they would influence operation in the event of a malfunction or failure.  Using
this approach, we found that the basic operation would fall into two categories:
1) a loss of performance which would be similar to what would happen with a
standard transmission, that is, loss of power to the road and requirement that
the automobile would have to be driven to the shoulder and repair would be
required; 2) a behavior which might be startling to the driver which would be
classed as malfunctions which result in a pathological effect.

Transmission Elements

Rather than dwell on what we might term as standard failures, Table F-l sum-
marizes the malfunctions and the effects as determined by the first review.
These are identified in accordance with the transmission schematic in Figure B-l.
However, noted in Table F-l, are three specific failures in the transmission
which fall into the pathological category and which could result in significant
safety hazards unless some corrective means was incorporated into the design.
In general, these latter types of failures are associated with the fact that
the flywheel is a separate source of power which in the event of malfunction
can operate in an unexpected and uncontrolled fashion.  The three failures in
this category are as follows:

   1.  Lockup of the Output Planetary B
       The effect would be to rapidly change the gear ratio between the flywheel
       and output shaft such that a rapid change of output speed would be
       realized.   For example, during this type of failure, the vehicle may
       accelerate without driver command.  Generally speaking, this type of
       problem will be of short duration.  However,  it would present a signi-
       ficant surprise and probably a pathological reaction.
                                    F-l

-------
                                                TABLE F-l

                                   TRANSMISSION FAILURE ANALSIS SUMMARY
  TRANSMISSION
   COMPONENT
     TYPE
  OF FAILURE
   EXAMPLE
  OF FAILURE
         EFFECT
    CORRECTIVE ACTION
Primary
Hydrostatic
(Elements I &
ID
Clutch & Brake
Interlock Sys-
tem for Clutch
and Brake
Lockup
                  Free Wheeling
                  Fixed Ratio
Open or Closed
Open or Closed
Bearing Seizure
                Loss of Hydraulic
                Pressure.   Mode
                Switch
                Trunnion Bearing
                Failure
Actuator Failure
Valve Seizure
Lock rear wheels with
probable slip of brake
and clutch.

Gradual vehicle decelera-
tion.

Engine stall.

Flywheel overdrives Ele-
ment III with subsequent
opening of pressure relief
valve.
                     Loss of drive torque,
                                                      Vehicle  slows.
                     Loss of variable speed
                     control with subsequent
                     engine stall.
Abnormal high or low
range operation.
Abnormal high or  low
range operation.
Set slip level for clutch
and brake.
                                                                                   Size relief  for required
                                                                                   flow.
Use interlock system.

-------
                                             TABLE  F-l  (cont.)

                                    TRANSMISSION  FAILURE  ANALYSIS  SUMMARY
  TRANSMISSION
   COMPONENT
     TYPE
  OF FAILURE
   EXAMPLE
  OF FAILURE
         EFFECT
    CORRECTIVE ACTION
Secondary
Hydrostatic
  Element III*
  Element IV
  Element III.
  and/or IV
Output
Planetaries
  Planetary B*
  Planetary B
  Planetary C
Freeze Up or
Lockup
Freeze Up or
Lockup
Free Wheeling
Lockup
Free Wheeling
Lockup
                  Free Wheeling
Bearing Failure
Bearing Failure
Sheared Shaft or
Loss of Hydraulic
Pressure  Mode
Switch
Bearing or Gear
Failure (Freeze
Up)
Sheared or Frac-
tured Shaft.
Bearing Seizure
                Loss of Gear Teeth
Change of vehicle speed.
Possible flywheel over-
speed.

No output.  Engine stalls,
Same as Planetary "A"
lockup.
Loss of Planetary "A"
reaction and loss of fly-
wheel power.
Rapid change of vehicle
speed, + or —, depending
on range condition.
Engine stall.
Loss of output power,
Slip of low-range brake
with gradual loss of
vehicle speed.

No low range power.
Hydraulic dump output of
Element III.

Containment or growth ring.
Shear joint or override
mechanism.

-------
                                              TABLE F-l  (cont.)

                                    TRANSMISSION FAILURE ANALYSIS SUMMARY
  TRANSMISSION
   COMPONENT
     TYPE
  OF FAILURE
   EXAMPLE
  OF FAILURE
         EFFECT
    CORRECTIVE ACTION
Input Planetary
  Planetary A*
  Planetary A
Flywheel
  Vacuum
  Flywheel
  Fracture
Lockup
Free Wheeling
Lockup
Loss of Pump
Breakup
Bearing Seizure
Sheared Shaft or
Loss of Gear Teeth
Bearing Seizure
Tri-hub
Rapid increase in engine
speed and vehicle speed.
Blow relief valve.
Loss of flywheel power,
High torque to transmis-
sion housing and mounting.
                     Overheating
Lockup or free wheel
Planetary A.
Shear joint or override
mechanism.
Bearing outer race slip,
Shear section.
                            Warning system.
Containment^
*Pathological failures.

-------
    2.  Freezeup of Element Three in the Secondary Hydrostatic Transmission

       Again, this can result in a sudden unexpected change in vehicle speed
       without driver command.  Because the flywheel is tied to the engine, a
       second effect could be flywheel overspeed.

    3.  Lockup of Planetary A.

       Once again, the major effect will be the change in vehicle velocity with-
       out driver command.

 In  all probability, pressure buildup, relief valve venting, and slipping of
 the clutch and brake system will minimize this effect for all three of the
 above faults.  With respect to the planetary gearing, some form of shear joint
 or  override member could be introduced into the design to minimize or eliminate
 this condition.  Certainly the failure mode which could result in flywheel
 overspeed would ultimately rely on some form of containment or possibly a
 growthring-type of failure brake mechanism.

Control System

Regardless of the detailed source,  a.failure in a control system signal com-
ponent becomes important only in terms of how it ultimately affects functioning
of the transmission and the flow of system power.  For example,  a variety of
detailed failures can occur in the  components which control the  secondary
transmission.  The net result will  be a value of the command transmission
ratio, either higher or lower than the correct value, which would result in
undesired vehicle deceleration or acceleration,respectively.

Viewed in this manner,  failures in  the control system can be analyzed as com-
mand errors in:  1)  Primary Transmission Ratio, 2) Primary Transmission Range,
3) Secondary Transmission Ratio,  4) Mode Switching,  5)  Engage-Disengage,  and
6) Engine Throttle Position.
                                    F-5

-------
Of these failures, the Primary Transmission Ratio,  Primary Range,  Secondary
Transmission Ratio, Failure of Mode Switching from  Charge to  Drive fall  into
the pathological category as defined for the transmission. These, as  well
as the loss of performance variety, are discussed below.

   1.   Primary Transmission Ratio

       Errors in primary ratio result in improper  loading at  the  engine  shaft.
       If a higher ratio is commanded,  the vehicle  will  try to  speed up  the
       engine shaft;  a lower ratio will try to drag the  shaft speed down.
       Since the significant flywheel inertia will  tend  to hold the engine
       speed constant, a step increase  in primary ratio  will  also  tend to
       decelerate the vehicle as it accelerates the engine and  flywheel.  A
       step decrease in primary ratio will tend to  accelerate the  vehicle.
       Both deceleration and acceleration levels will be limited  by the  relief
       valves in the hydrostatic links, but a step  primary ratio  error can re-
       sult in a harsher and potentially more dangerous  undesired  velocity
       change with a flywheel system than with a conventional vehicle  where in-
       ertia storage is much less.

   2.   Primary RanKfc

       Results of errors in commands to the high-speed clutch and  low-speed
       brake are similar to malfunctions in those  components  as discussed
       previously.

   3.   Secondary Transmission Ratio

       Erroneous ratio commands to the  secondary transmission will result in
       undesired vehicle acceleration or deceleration due to  flywheel  action.
       Although not exactly like a full pedal or brake input, a fast full sec-
       ondary ratio error would call for essentially maximum  acceleration or
       deceleration capability of the flywheel augmented  system.
                                    F-6

-------
   A.  Mode Switching

       As in a conventional vehicle,  failures in tirade  selection  or  establish-
       ment logic mainly result in abnormal driving conditions with improper
       available power or engine loading conditions.   In  a  flywheel vehicle a
       failure into the charge mode disengages the primary  transmission and
       could initiate abnormal flywheel chargeup until limited by the  flywheel
       velocity limit.  Change without a command from  charge  to  drive  could
       cause a brief vehicle acceleration burst.

   5.  Engage-Disengage

       Failures in the primary and secondary engage-disengage systems  lead to
       inconvenient but not dangerous maloperation. Erroneous disengage  opens
       the hydrostatic links and causes loss of drive  or  flywheel capabilities.
       Erroneous engage can lead to stalling the engine or  prevent  a startup.

   6.  Engine Throttle Position

       As with a conventional vehicle, an erroneous throttle  "on" command leads
       to runaway vehicle acceleration.  The "normal"  driver  response  is  to
       apply the brake and/or cut the ignition.

Summary

As indicated, failures of certain transmission elements and control failures
leading to primary and secondary transmission ratio errors  can result  in  po-
tentially dangerous vehicle acceleration or deceleration  due  to  the special
energy storage capability of the flywheel.

In the preceding discussion, it was assumed that these types  of  malfunctions
would require the same type of driver reaction as conventional vehicles re-
quire.  The malfunctions noted above which  fall into the  pathological  category
would not be immediately corrected by the normal reactions  of the driver.  For
example, under abnormal conditions the driver would lift  his  foot from the
accelerator and apply the brakes expecting  to slow down.  He  would  still  note
                                    F-7

-------
an increase in vehicle speed or no indication of deceleration for a period of
time.  Similar comments apply to the condition where unexpected decelerations
take place.

Although not included in our considerations.it is very probable that,  should
the flywheel hybrid propulsion system come Into common usage, certain  correc-
tive action could be built in and integrated with what would normally  be ex-
pected from the driver.

Runaway accelerations can be simply guarded against by providing a solenoid
input to the secondary disengage valve such that the normal response of turn-
ing off the ignition will also open the secondary link.

Corrective action to abnormal abrupt deceleration is not as simple, since the
normal response to a quick surprise braking would not be to cut the ignition.
An automatic means for protecting against such a situation would be to sense
the vehicle gross acceleration or deceleration levels and disengage the secon-
dary l.f the respect I. vi.- levels are not commensurate wttli  tlu1 existing accelera-
tor or brake pedal positions.
                                   F-8

-------
      SECTION G


REGENERATIVE BRAKING
                                                           3)
                                                           n>
                                                           
-------
                       G.  REGENERATIVE BRAKING ANALYSIS

In reviewing the requirements for dynamic (regenerative) braking, it was impor-
tant to assess the feasibility of braking and to determine if the use of flywheel
dynamic braking would unduly penalize the overall transmission design.  In the
latter case, this would require a trade-off between any added complexity or
design compromise and the extent of their benefit.  In terms of assessing feasi-
bility, this was considered on the basis of a total energy concept, practical
braking limits for vehicles, peak torque and horsepower requirements, and any
other influences brought about by the flywheel.

In terms of the total energy concept, there is no real question that in theory,
sufficient energy can be absorbed in the flywheel to effectively absorb the
kinetic energy of the vehicle in going between two speeds.  The questions have
to do with the practical limitations of road coefficient of adhesion, flow of
power from the rear wheels only, and possible modifications if some combination
of flywheel and mechanical braking is required for overall vehicle passenger
safety.

One of the primary considerations with respect to braking must concern itself
with operator or passenger safety.  Therefore, the initial point for assessment
concerns itself with an emergency stop type situation, which for analysis pur-
                                                      2
poses was taken to be a deceleration rate of 20 ft/sec .  For purposes of estab-
lishing the outside limit, a vehicle weight of 5300 Ibs. was selected and a
factor to compensate for the inertia of wheels and transmission parts of 1.05 was
applied to this value giving a total equivalent vehicle weight of 5565 Ibs.  The
                              2
deceleration rate of 20 ft/sec  is equivalent to a coefficient of road adhesion
of 0.62 which gives a braking force of 3450 Ibs. with the above vehicle weight.
During a deceleration of this  magnitude,  the  distribution of  vehicle weight  is
such that approximately 35 percent will be applied to the rear wheels.  This
represents a rear wheel braking force of 1208 Ibs.  At this point, several
conclusions regarding the braking situation can be made:

   1.  For an emergency stop, braking must utilize an assist to the dynamic
                                     G-l

-------
        braking possible from the flywheel since only 35 percent can be
        dissipated through the rear wheels.

   2.   There is a limitation with regard to regeneration if this is to
        come purely from rear wheel braking energy alone.  (The engine
        must make up at least 65 percent of the total flywheel energy
        required.)

   3.   Energy dissipated from other elements such as front wheel brakes
        and the combined effects of air resistance and rolling resistance
        are not recovered as regeneration energy.

There are several implications of the above data with respect to operator safety.
Certainly the basic conclusion is that there can be modes of operation where the
flywheel would be unable to dissipate the required energy necessary for emergency
stopping.  Recent literature indicates that one-half g stops are certainly com-
                                   2
mon and that the use of a 20 ft/sec  stopping rate is not unreasonable.  (See
Reference  7.)  Thus, it is concluded that the four-wheel braking capability
currently utilized in vehicles could not be eliminated nor significantly com-
promised.  However, further consideration must be given to other implications of
the system and lesser rates of deceleration which are more consistent with the
typical driving cycle.  Before discussing these, it is appropriate to consider
some additional qualitative implications suggested by the above discussion.
These items are as follows:

   1.   If recuperation of the flywheel is achieved by dynamic braking and
       engine input, there is a dual power path which involves additional
       complexity in control.  Extreme care would be necessary to pre-
       clude the rotating inertia from coupling to the vehicle to reflect
       itself as an even larger equivalent vehicle weight.

   2.  During this engine-assisted flywheel charging process,  it would
       be an additional control requirement to match the conditions
       necessary to maintain a reasonable engine specific fuel consump-
       tion.
                                     G-2

-------
Based on the above it is logical to consider the lower rates of deceleration and
recovering energy as a retarder which might be set up with minimum modification
and provide added brake-shoe life on both front and rear wheels.  In order to
demonstrate this we considered the same road coefficient of adhesion (0.62), an
effective vehicle weight of 5565 Ibs., and took into account the revised weight
distribution resulting from the deceleration rate to determine that a decelera-
                  2
tion of 8.2 ft/sec  would just permit regeneration.  The braking force corres-
ponding to this rate of deceleration is 1,417 Ibs.  As in the previous calcula-
tions, the effects of air resistance and rolling resistance were neglected since
this is a conservative approach with respect to braking safety.  Figure G-l shows
the braking behavior for the maximum tractive force and constant rates of decel-
eration.
                                    2
Since the deceleration of 8.2 ft/sec  was determined to be the value where full
regeneration was possible, this was considered further in terms of the transmis-
sion system concept.  One of the important factors concerning the ability of the
system to provide recuperation is the instantaneous flow of power which must be
delivered from the rear wheels to the flywheel.  At 85 miles per hour, the in-
stantaneous power requirement is 319.7 lip.  The size of the hydrostatic compo-
nents permits a peak power transfer rate of 164.4 hp.  This power is referenced
to the wheels and includes the braking benefits due to system losses.  In terms
of a rear wheel braking force, the 164.4 hp corresponds to 725.6 Ibs.  The con-
clusion here is that to achieve regeneration up to a maximum deceleration rate
             2
of 8.2 ft/sec , the capacity of the flywheel transmission must be doubled.  The
attached Figure G-2 shows how the system could be altered to realize this addi-
tional capacity.  The added components and control complexity were estimated to
add between 12 and 15 percent to the overall transmission cost.

Finally, the flywheel was considered for a retarder.  In essence this considered
that the maximum braking force available would be the 725.6 Ibs.  If this is now
used in conjunction with normal brakes for an emergency deceleration, again
                     2
taken to be 20 ft/sec , the flywheel would be capable of absorbing approximately
16.8 percent of the vehicle kinetic energy when stopping from 85 miles per hour.
The above calculation is based on an assumption of 80 percent efficiency in
returning the regenerative energy to the flywheel.  Still another way to consider
this is to determine what deceleration or retarding capability could be provided

                                     G-3

-------
                                               	 .3 x 10  KE Flywheel at 10,000 RPM
        Max. KE of Vehicle  (1.4 x  IO6)
   10
                     5 ft/sec
                              v
                                              f
                                             7
1.7 x IO6 KE Flywheel
at 24,000 RPM
o
.—i
x
 1
4J
14-1
t>0
Cd
0)
  AKE
Required
   of
Flywheel
                                    .3 x  10  KE  Flywheel
                                    at 10,000  RPM
   io
                           Braking KE -  .62 Coefficient
                           Decel. =  10 ft/sec2
                                     Braking  KE  Available  at  Rear Wheels
                                        .62  Road  Coefficient
                                       Decel.  =  20  ft/sec2
                           Brake KE Matches Vehicle  KE
                           When:  Decel.  = 8.17  ft/sec2,
                                  Road  Coefficient = .62
                                      i  i
                                                  l   l  l  i    i
                  20      30   40  50 60 70 80   10 12 14 16 18 20  24
                       MPH
                                             Flywheel RPM x  10
                                                                              o
                    Fig. G-l    Flywheel Regenerative Braking  Behavior
                                             G-4
                                                                                         KII-12502

-------
O
Ul
Bngin_£
Input






\e








































J


1



]
J






r~


r



~
F—
t 	 L _-























1




]
P

I
b









~

s
1
1
^

















\T)
T C • 3
7 . 5 in




FD
7.5 in3

VD
7.5 In3












J





FD
3
7 .5 In







VD
7.5 in3

































F
i





i
4




                                                                                                                Output
to Rear
Axle
                             Fig. G-2   Power-Splitting Transmission Sized for  Regenerative Braking
                                                                                                                     MTI-12503

-------
by the flywheel transmission as shown in Figure B-l, page 9-2'-.  In this instance,
the flywheel i
regeneration.
                                                                   2
the flywheel could provide a deceleration capability of 4.41 ft/sec  with full
It was felt that the above represented reasonable guidelines for what could be
expected from regenerative braking.  There are obviously certain technical
problems which would have to be resolved to take advantage of the flywheel even
as a retarder.  For example, there would have to be some mechanism provided so
that the proper braking force at the rear wheels could be shared by the rear
wheel brakes and the flywheel circuit in agreement with the proportion discussed
above.  This would, in turn, have to be reconsidered in terms of the implication
on safety.   Automatic retarding action would be realized when there was no
throttle requirement (foot removed from the accelerator).
                                    G-6

-------
                               H.   REFERENCES
1.  Environmental Protection Agency, Advanced Automotive Power Systems,
    "Vehicle Design Goals - Six Passenger Automobile," Revision C,
    May 28, 1971.

2.  Lockheed Missiles and Space Company, "Flywheel Feasibility Study and
    Demonstration," Final Report LMSC-D007915, April 30, 1971.

3.  Automotive-Industries,  March, 1969, Page 131.

4.  Lockheed Missiles and Space Company, Letter LMSC-D179983
    0/50-33, B/528, August 12, 1971.

5.  Lockheed Missiles and Space Company, Letter LMSC-D244218,
    October 20, 1971.

6.  David N. Hwang, "Fundamental Parameters of Vehicle Fuel Economy and
    Acceleration," Paper //690541, Society of Automotive Engineers,
    October 30, 1968.

7.  Rudolf G. Mortimer, "Hard Braking is More Common Than You Might Think,"
    Automotive Engineering, August, 1971.
                                     H-l

-------
TJ
m
Z
g

x

-------
                                  APPENDIX I
                           DESCRIPTION OF METHODS FOR
           DETERMINING TRANSMISSION AND PROPULSION SYSTEM PERFORMANCE
Presented herein are (1) a discussion of the transmission performance analysis,
(2) a description of the computer program for steady-state performance, and (3)
a discussion of the digital dynamic simulation used for determining transient
performance.

A.  Transmission Performance Analysis
The transmission is considered to comprise the following components:
    • Spur Gears
    • Planetary Gears
    • Pumps
    • Motors

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:
    1.  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.

    2 .  Flow Losses
    Flow losses represent deviations from ideal performance of a hydraulic
    pump or motor.  Thus, while nominal performance assumes that pump flow
       T
    (in /sec) is equal to the product of speed (rad/sec) and displacement
    (in /rad), the actual flow differs from this product by a small amount
    — similarly for the motor.  The direction in which flow losses act is
    determined by the direction of power flow.  Only hydraulic units are
    subject to flow losses.
                                     1-1

-------
3.  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.  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
                           Wheel   1
                Wheel   2
Speed Equation:
Mechanical Loss Equation:
                                 Normal Direction of Power Flow
                                                                       (I-D
                                                                       (1-2)
where
     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

                                 1-2

-------
     T) is the efficiency of transmission


     I (1 — T])T1  represents the absolute value of the quality  (1 — Tj)T1


     Sign (N ) represents the algebraic sign of  the quality N  ; i.e.,


               if N  is positive, Sign  (N..) = +1; if N^ is negative,


               Sign (N) = -1.
Planetary Gears
 V \
                            Cage
                                          T    N
                                              '
                                                       Assumed Normal

                                                       Direction of

                                                       Power Flow
Speed Equation:
N,.  =  N.
                •S     *R



Mechanical Loss Equations
                                                                        (1-3)
     T    =  T  - T
      S       S    LS
          =  T  — T
              R    LR
                                                                  (1-4)
      T   =  T   + T
       C      C     LC
Torque Relationships:
                                                                       (1-5)
                  Ts'
                                 1-3

-------
Mechanical Loss Definition:
      LS
LR
                                
-------
Flow Loss Equation:
     Qp  -  NpDp
Pressure Calculation:
     P  =  Tp'/Dp
                                 sign (NpTp))
                             (1-6)
                                                                       (1-7)
Motors

 P»QM
                                                Normal Power Flow Direction
Mechanical Loss Equation:
     T   =  T   —  T
      M      M      LM
where

     LM
                  P   TT-
                  max
Flow Loss Equation:
                           (K2M + K3M °P
sign (NM)
                                                                       (1-8)
                                                                        (1-9)
Transfer of Flow From Pump to Motor
     QM  =  QP  ~  Q
where
                    LC
     QLC  =  LlP/Pmax
              K1L + K2L + K3L
                                        + [KIL] | QM
                                                                        (1-10)
In calculating hydraulic transmission performance, the pump and motor are
considered in combination.  Equations (1-5) through (1-9) relate the six
values of torque, speed and displacement for the pump and motor in such a
                                  1-5

-------
way that given four of the  six values,  the remaining unknown  two may
be determined.

In the above treatment of hydraulic units, the  following definitions apply:

     T ,31    are the input  torque to pump, and  output torque  from the
             motor respectively.
     Dt>»Dv<   are the pump and motor displacements
      r  el
        D    is the ratio of pump displacement  to  its maximum value
             for the unit
     N ,N    are pump and motor speeds, rad/sec
     Nn>N..   are ratios of  pump and motor speeds to the maximum
      f  M
             values for the unit
                               2
         P   is pressure, Ib/in
      P      is maximum pressure value  for the  unit
       max
       K     is a flow loss coefficient (see below for values imposed
     ...^,
    In  3M
             for all loss coefficients)
             are mechanical loss coefficient
   ,K  ,K    are coefficients relating to compressibility and leakage
 1L  2L  JL
      n>«
      P  M
             losses
                                                       o
             are tne effective pump and motor flows, in /sec
       Q     is a leakage flow
        LC

Additional Losses
In addition to losses in gears and hydraulic units, further losses are con-
sidered to occur at various points in the transmission to account for the
power needed to drive such components as the flywheel vacuum pump.   (See
Figures AI-1 and AI-2  —  Parasitic Losses.)  Such losses are treated in
a manner exactly analogous to the spur gear losses using a unit gear ratio.
Additional losses are also considered to overcome flywheel windage and
bearing friction (as specified by LMSC), to overcome road resistance (as
specified by EPA).

                                  1-6

-------
                                                  PRIMARY HYDRAULICS
   TpL=0.985
                         Actuator
                         Stroke
                                     Tl = 0.99
Te,N
e.e
 Parasitic
 Losses
                   T| =1.00
              Secondary Planetary
                                                   Element
Element
   II
                                                SECONDARY HYDRAULICS
                                            R  =  Gear  Ratio

                                            r  =  Gear  Radius
                                            Ne =  Engine  Speed
                                                                                                          0.96
I^U.b
rR
rs
V
»J
1
l| -u. -y

R6
1
?
N3


Element
III


Element
IV



N4

"; = U.VB.
R5

                                                                                 Actuator
                                                                                 Stroke
                    Fig.  AI-1    Power-Splitting Transmission - Low-Speed Range  Diagram
                                                                                                               Wh e e1s
                                                                                                                 HTI-12516

-------
i
oo
                                       Actuator
                                        Stroke
                                                           PRIMARY HYDRAULICS
                                                                                     1) = 0.99
                                                                               I              T       — _r\ nn
                                                                                   odge
          NFI
             I  T| = 0.99
                                                        SECONDARY HYDRAULICS
                  Sun
             i     sun                    i


                  SECONDARY PLANETARY

rR
rs
N6 "
" 1
1
D
R6
Ring 1
N3


Element
III


Element
IV



\

T]=0.985
R5

                       PRIMARY PLANETAR"
Actuator
 Stroke
                                                            R  =  Gear  Ratio

                                                            r  =  Gear  Radius
                                                           N'e  =  Engine  Speed



                               Fig. AI-2   Power-Splitting Transmission —  High-Speed  Range Diagram
                                                                                                                          MTI-12548

-------
    Numerical Values for Efficiencies and Loss Coefficients
    The loss coefficients and mechanical efficiency numbers for the considered
    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 efficiency used for
    performance calculations are shown in Figures AI-1 and AI-2 for the low-
    speed and high-speed ranges, respectively.  The loss associated with TL
    has a minimum value of 1.5 hp at engine speeds above 3200 rpm.  At lower
    speeds the minimum loss is linearly interpolated between 1 hp at 1400 rpm
    and 1.5 hp at 3200 rpm.

    The equations 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 resultant values of hydraulic element loss
    coefficients employed in all performance analysis calculations were as
    follows:

         K    =  0.0225                   K    =  0.0075
          1 r                               1L
         ^M  =  °'°25                    K2L  =  0>°25
         *™  =  °-005                    KQT   =  0.005
          jM                               JL

B.  Digital Computer Program for Steady-State Performance Analysis
In broad terms, the computer program operates by moving backwards, from the
wheels and the flywheel, towards the engine.  At each point in the transmission
the program calculates the torque required to overcome all hydraulic and mechan-
ical losses between that point and the wheels or flywheel, together with the
torque to overcome resistance to motion of the wheels (resistance specified by
EPA) or flywheel (windage and friction specified by LMSC).  The flywheel speed
is related to the wheel speed by the TKE relationship.  The primary and secon-
dary (flywheel) transmissions are linked together to satisfy the necessary speed
and torque relationships of a geared torsional system under steady-state operation.
                                     1-9

-------
The above statement of operation requires the addition of some constraint to
define engine speed.  The standard procedure followed by the program is to seek
the engine speed which provides the necessary power at a condition of minimum
SFC.   To do this and not overconstrain the problem it is necessary to leave one
displacement in each of the two hydraulic power converters undefined.  The
program then establishes the displacement value for each hydraulic power  con-
verter which provides the necessary transfer of speed and torque.   Thus,  in
addition to calculating performance,  the program provides a means  of defining
the primary and secondary displacement schedules as a function of  speed.

Under certain conditions of speed and load,  steady-state operation on the minimum
SFC line cannot be achieved, due to displacement limits of the hydraulic  units  or
limits on the engine speed.  These special conditions will be amplified subse-
quently .

Detailed Program Procedure
The following sequence of operations  describes the actual procedures followed by
the computer program.  Figures AI-landAI-2  provide a reference for this  descrip-
tion.  The terms upstream and downstream describe relative locations which,
respectively, follow or oppose the arrows of Figures AI-1 and AI-2 .  Operation  of
the transmission in low range (Fig. AI-1) is described first (items 1-20) followed
by modifications (items 21-25) to handle the high-range operation  (Fig. AI-2)

    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.  The flywheel speed corresponding to  the vehicle speed of interest is
        calculated via the TKE relationship.
    3.  The flywheel windage and friction losses corresponding to  this  flywheel
        speed are calculated according to LMSC specifications.
    4.  That engine speed is calculated which will provide the required power
        at minimum SFC.  This first calculation of speed is made assuming no
        intermediate losses apart from resistance and windage.
                                     1-10

-------
 5.  Using equations 1-1 and 1-2, the speed and torque acting immediately up--
     stream of the rear .axiel %. differential (RA) are computed.
 6.  Using equations 1-1 and 1-2, the speed and torque acting immediately up-
     stream of the output gear ratio (R3) are computed.
 7.  Using equations 1-1 and 1-2, the speed and torque acting immediately up-
     stream of the primary hydraulic output gear (R2) are computed.
 8.  Using equation 1-1, the speed downstream of the primary pump input gear
     (Rl) is calculated from the current value of engine speed.
 9.  Using equations 1-6, 7, 8, 9 and 10, with pump and motor speed, motor
     torque, and motor displacement specified, the pump displacement and pump
     torque are determined for the primary hydraulic power converter (this
     system of equations is mildyly non-linear in displacement, but is solved
     effectively by direct (Picard) iteration).
10.  Using equation 1-2, the torque upstream of the primary pump input gear
     (Rl) is calculated from the torque at input to the primary pump
     (element I).
11.  Using equations 1-4 and 1-5, the torques at the ring and cage of the fly-
     wheel planetary are calculated from the flywheel windage torque (sun
     torque).
12.  Using equation 1-3, the speed of the ring of the flywheel planetary is
     calculated from the flywheel (sun) and engine (cage) speeds.
13.  Using equation 1-2 for R,, the torque acting at output from the secondary
     motor (element III) is calculated.
1A.  Using equation 1-1 for R ,_the speed of the secondary pump (element IV)
     is calculated from the engine speed.
15.  Using equations 1-6, 7, 8, 9 and 10 with motor torque, motor speed, motor
     displacement and pump speed specified, the pump torque and displacement
     are calculated for the secondary hydraulic power converter.  (Direct
     iteration is again used to handle the non-linearities in this system
     of equations).
16.  Using equation 1-2 the torque upstream of the secondary pump gear (R,)  is
     calculated.
17.  All the torques acting on the engine output shaft downstream of the
     parasitic losses are now defined.   These torques are added to give the
     torque downstream of the parasitic losses.
                                   1-11

-------
     18.   Equation  1-2 with a  gear  ratio  uf  unity  is  imposed  to  calculate  the
          power upstream of the  parasitic losses,  which  is  also  the  engine output
          torque.
     19.   From the  engine performance  charts the engine  speed to define  this
          torque at minimum SFC  is  interpolated.
     20.   The above procedure, starting at item 5,  is  repeated until  engine speed
          and displacement values are  repeatable between successive  iterations
          within 1 part in 10,000.

For  operation of the transmission  in  the high range,  the following  steps  replace
Steps 6 and 7:
    21.  Using equation  1-1  for gears R   and  R  ~,  the  speed downstream  of R  n
         is calculated.  This speed is also the  cage speed for  the  output
         planetary.
    22.  With cage and ring  (upstream of  R.)  speeds specified,  equation 1-3  is
         used to calculate the sun speed  - which is the speed downstream of  R  .
    23.  Using equation  1-4  and 1-5, the  sun  and cage  torques for the output
         planetary are calculated.  The sun torque is  the torque downstream  of
         R .  The cage torque is the torque downstream of R-.
    24.  Using equation  1-2  for ratio R,n, the torque  upstream  of R n is cal-
         culated from the cage torque.
    25.  Using equation  1-2  for ratio R , the torque upstream of R  is  cal-
         culated.
Apart from the above modifications to handle the output planetary and ratios
Rq and R-,n» the treatment for high-range operation parallels that for low-
range operation.

The following additional constraints apply:

A.   If the speed which provides the required torque at minimum SFC is less than
     1400 RPM or greater than 3800 RPM, the desired speed is taken to be 1400 RPM
     or 3800 RPM respectively.
                                     1-12

-------
B.   If any of the engine speed requirements call for displacements of  either
     of the primary hydraulic units which exceed the allowable limits,  then  the
     value of displacement is set at  that limiting value and the equations
     solved for engine speed.  This constraint override!* Constraint A.

Sample Calculation

Table AI-1 shows sets of computed quantities at vehicle speeds of 20 and  70 MPH.
The tables include quantities at every major point in the transmission.
C.  DIGITAL DYNAMIC SIMULATION ANALYSIS AND COMPUTER PROGRAM
For the purposes of dynamic analysis the vehicle and transmission are modeled
as a three-inertia system consisting of the Engine (J ), Flywheel (Jf), and
Output (J ) inertias.  The output inertia represents the vehicle mass referred
         o
to the rear axle.  All secondary inertias are referred to the appropriate speed
and combined with one of these three model inertias.
The equations of motion for the three inertias are:
                 J N      =    T  - T                       (1-11)
                  o w           w    R
                 J N      =    T.  - T                      (1-12)
                  e e           me                     ^
                 JfNf
Where N , N ,  Nf are the angular speeds of the output, engine, and flywheel
inertias respectively in rad/sec.

T        is the torque transmitted to the wheels via the transmission

T        is the resistance torque acting on the wheels (rolling, wind and
         grade as specified by EPA)

                                     1-13

-------
    TABLE AI-1   TYPICAL  DETAILED  RESULTS  OF  STEADY-STATE
                     PERFORMANCE COMPUTER PROGRAM
INPUT DATA:
   Vehicle  Weight
   Cd x A
   Grade
   Ambient  Temperature
   Ambient  Pressure
   Fuel Density
 4600 LB
12.00 FT2
 0.00 Per Cent
85.00 °F
14.70 PSI
6.152 LB/CAL
   Gear ratios  as specified In Fig.  B-l
   Engine characteristics as specified by EPA
   Pierced Flywheel - chamber pressure = 2.94 PSI
   Resistance  losses as specified by EPA
VEHICLE VELOCITY

Power Train  Efficiency
Transmission Efficiency
Primary Efficiency
Secondary Efficiency

Wheel Speed
Output Speed
Flywheel Speed
Engine Speed
Speed At Element   I
Speed At Element  II
Speed At Element III
Speed At Element IV
Road Torque
Output Torque
Flywheel  Torque
Engine Torque
Torque At Element   I
Torque At Element  II
Torque At Element III
Torque At Element  IV

Road Horse Power
Output Horae Power
Flywheel  Horse Power
Engine Horse Power
HP At Element    I
HP At Element  II
HP At Element  III
HP At Element  IV
Auxiliary Horse  Power

Fuel Flow
SFC
MPO

Displacement Of Element   I
Displacement Of Element  II
Displacement Of Element III
Displacement Of Element  IV

Primary Pressure
Secondary Pressure

Ideal Flow At Element   I
Ideal Flow At Element  II
Ideal Flow At Element III
Ideal Flow At Element  IV

Effective Flow At Element   I
Effective Flow At Element  II
Effective Flow At Element III
Effective Flow At Element  IV
Primary Leakage Flow
Secondary Leakage Flow
                                           20  M.P.H.
                                                                    70   M.P.H.
37.114
87.100
100.317*
97.135
266.774
947.049
23444.553
1400.000
1102.360
142.699
-1991.470
1272.740
90.400
26.483
-1.344
74.974
-1.549
-12.003
-9.003
14.502
4.592
4.775
-6.001
19.985
-0.325
-0.326
3.414
3.514
7.613
12.276
0.614
10.023
0.968
7.500
3.633
-5.841
120.661
187.056
17.788
17.837
-120.584
-123.901
17.806
17.835
-122.723
-123.070
-0.029
-0.347
per cent
per cent
per cent
per cent
RPM
RPM
RPM
RPM
RPM
RPM
RPM
RPM
Ft. Lb.
Ft. Lb.
Ft. Lb.
Ft. Lb.
Ft. Lb.
Ft. Lb.
Ft. Lb.
Ft. Lb.
HP
HP
HP
HP
HP
HP
HP
HP
HP
Lb. Hr.
Lb. Hr. HP

In3/Rev
In-j/Rev
In /Rev
In /Rev
Psi
Psi
In^/Sec
In,/Sec
In^/Sec
In /Sec
In^/Sec
In,/Sec
In^/Sec
In /Sec
In^/Sec
In /Sec
82.864
90.355
95.172
100.886*
933.710
3314.671
15911.520
2817.460
2218.468
-2835.430
802.989
2561.353
244.733
71.696
-0.724
125.222
45.419
-33.821
-4.657
-1.447
43.508
45.248
-2.193
67. 174
19.185
18.259
-0.712
-0.706
14.669
35.309
0.526
12.196
-7.500
5.625
5.500
1.710
455.166
63.831
-277.308
-205.822
73.607
73.004
-272.983
-271.121
73.413
73.192
-1.862
-0.221
per cent
per cent
per cent
per cent
RPM
RPM
RPM
RPM
RPM
RPM
RPM
RPM
Ft. Lb.
Ft. Lb.
Ft. Lb.
Ft. Lb.
Ft. Lb.
Ft. Lb.
Ft. Lb.
Ft. Ub.
HP
HP
HP
HP
HP
HP
HP
HP
HP
Lb. Hr.
Lb. Hr. HP

Inil/Rev
In^/Rev
In,/Rev
In /Rev
Psi
Psi
In^/Sec
In^/Sec
In^/Sec
In /Sec
In^/Sec
In^/Sec
In^/Sec
In /Sec
In^/Sec
In /Sec
 Efficiency  >  1007. Implies that  flow of power is in reverse of arrows in Figs. Al-1 and
 AI-2 (see values for power).
                                            1-14
                                                      MTl-12335

-------
T       is the engine input torque
 in

T       is the engine output torque  (acting on  transmission)
 e

T  .  ,   is the windage and friction  torque acting on  the flywheel  (as  specified
 wind
        by Lockheed)
T       is the flywheel output  torque  (acting on  transmission)
The interrelationship between the system torques,  including mechanical  losses
follows the treatment of Section A of this Appendix.

The flow, compressibility and leakage losses are also calculated  according
to the procedure of Section A of this Appendix.
Thus, the torques T  , T_, T.  , T  , T  .  ,, T,, can be calculated  at  any  time,  t,
                   w   K   in   e   wind    t
as a function of the three velocities N , N   N  , and  the primary and secondary
                                       w    i   e
pressures.
In determining the pressure in the hydraulic units  the  treatment differs  from
the steady state procedure.  Under dynamic conditions,  the rate of change of
pressure, rather than the pressure itself is determined by the torques and
speeds of the pump and motor as follows:
                             =                              (1-14)
                         at       v                         v    '
where:      p is the pressure
            6 is the bulk modulus of the fluid
            Q is the net mismatch in pump and motor flows
            V is the volume of fluid in the unit.
                                   /t-15

-------
Equation 1-14 is applied both to the primary and to the secondary hydraulic

transmission units.  Thus, together with the three equations of motion, 1-11,

12, 13, the dynamics of the mechanical and hydraulic system are described by

5 dynamic equations, and the torque-speed pressure relationships of Section A

of this Appendix.  A further series of dynamic equations define the performance

of the control system, whose function is to adjust the displacements of the

hydraulic units in response to commands of the pedal and requirements of main-

taining constant IKE.  The control system is defined by the full range block

diagram, Figure AII-7.


The complete system of twelve equations defining the dynamic performance of the

vehicle and transmission in response to pedal commands is solved via a fourth

order Runge-Kutta integration algorithm.  Output quantities are printed at

regularly spaced time intervals (typically every 0.1 sec).


The following quantities are calculated and printed by the computer program as

a function of time.
        Vehicle Velocity
        Vehicle Acceleration
        Wheel Speed
        Flywheel Speed
        Engine Speed
        Transmission Output Speed
        Wheel Torque
        Road Torque
        Engine Input Torque
        Engine Output Torque
        Flywheel Loss Torque
        Flywheel Output Torque
        Transmission Output Torque
        Engine HP
        Road HP
        Flywheel HP
        Overall Efficiency
        Overall Transmission Efficiency
        Primary Efficiency
        Secondary Efficiency
        Fuel Flow
        Specific Fuel Consumption
        MPG
        Speed, Torque, HP at Hydraulic Units
                                    1-16

-------
        Primary Pressure
        Secondary Pressure
        Displacements of Hydraulic Units
        Ideal Hydraulic Flows
        Cumulative Fuel Consumption
        Cumulative Engine HP-Sec.
        Cumulative Flywheel HP-Sec.
Input to the computer program consists of:
        Initial Values of Spped,  Pressure,  Control Settings
        Control Gains, Time Constants
        Hydraulic Fluid Compressibility
        Primary,  Secondary Volumes
        TKE
        Wheel Radius
        Grade
        Car Frontal Area
        Wind Resistance Coefficient
        Ambient Pressure, Temperature
        Vehicle Weight
        Accessory HP Tables as  a  Function  of Engine Speed
        Engine Performance Tables
        Idle Fuel Flow
        Inertia Values
        Gear Ratios
        Sun, Gear,  Cage Radii  for Flywheel  and  Output  Planetary  Gears
        Mechanical  Efficiencies
        Compressibility, Leakage  and Flow  Coefficients
        Numerical Integration  Controls
        Range Switch Controls
        Limits of:   Pressure;  Fuel  Flow Rate of Change; Actuator Position;
                    Displacements;  Rate of  Change  of Actuator Position.
                                   1-17

-------
                                 APPENDIX II

               STABILITY AND ANALOG COMPUTER SIMULATION ANALYSIS

Presented herein is a discussion of the stability analysis performed for the
transmission as an integral part of the flywheel/hybrid propulsion system.  This
is followed by a discussion of the related analog computer simulation analysis.

1.  Stability Analysis
The objectives of the linearized stability analysis were:
     A.  To determine whether the feedback control loops are basically
         stable or unstable.

     B.  If unstable loops exist, add compensation as required to obtain
         stable operation.

     C.  To determine the optimum (maximum allowed) gain values.  If the
         gains are too low, the transient response will be too slow.  If
         the gains are too high, the system will be unstable.

From a general viewpoint, open loop gain and phase lag both vary with frequency.
If the phase lag between the input to a control loop and the feedback signal is
180 degrees and the gain is 1.0 (zero decibels) or greater at the same frequency,
the loop will be unstable.  The open loop gain is usually set to give zero
decibels at about 135 degrees, to provide adequate damping.

Stability analysis techniques apply only to linear systems.  Therefore,  the sys-
tem nonlinearities must be replaced with equivalent linear functions.  The gain
varies with the operating point for a nonlinear system.  Therefore, stability
must be investigated over the full range of the parameters.

The effects of nonlinearities and interaction between control loops is best
investigated on an analog computer.   Some  modification of the calculated gain
values is usually required because of these effects.
                                    II-l

-------
The full-range nonlinear block diagram for the complete system is shown in Figure
AII-7. The linearized block diagram of the secondary control system is shown in
Figure AII-1. The linearized models were derived from the nonlinear block diagram.

Secondary Transmission Control
The secondary transmission has two integrations:  one from the flywheel, and one
from the piston actuator.  Since two integrations provide 180 degrees phase lag
at all frequencies, the secondary transmission will be unstable for any reasonable
gain unless feedback compensation is used to reduce the phase lag.  Two types of
feedback around the piston actuator were considered:  a mechanical feedback
linkage and pressure feedback.

The steady-state displacement of hydrostatic Element IV, D,,* varies with the
values of engine speed and flywheel speed.  With a mechanical feedback linkage
around the secondary piston actuator, D,  is given by

                                    K13 K4
                             D4  -  -if—  AKE
where:
       Kf  =  feedback gain around primary piston actuator
      AKE  =  kinetic energy error

Thus, a kinetic energy error is required to maintain a steady-state value of D .
The kinetic energy error is proportional to D,.   The kinetic energy error inher-
ent with this type of feedback would not be desirable.

With pressure feedback, a kinetic energy error is not required to maintain
steady-state values of D,.   The pressure feedback is equivalent to flywheel
torque.  Thus, with pressure feedback, flywheel torque is proportional to kinetic
energy error.  At steady-state, the kinetic energy error is proportional to fly-
wheel friction.  Flywheel friction torque is small compared to the maximum fly-
wheel torque.  Therefore, the steady-state kinetic energy error will be small.
 *Nomenclature  is  given  at  the  end  of  this Appendix.
                                     11-2

-------
At a constant value of D,, an increase of engine speed results in an increase of
flywheel speed.  Engine speed is an undesired signal input to the secondary con-
trol system.  Pressure feedback will reduce the effect of this undesired engine
speed input.  Pressure feedback will also reduce any tendency of the secondary
transmission pressures to saturate.  Therefore, pressure feedback was selected
to reduce the phase lag of the piston actuator.

The linearized block diagram of the secondary control system is shown in Figure
AII-1. The speed of hydrostatic Element IV, N, , is considered constant in the
linearized model.  The square function in the flywheel speed feedback was
linearized as follows.  The square function is described by the equation
                                         Nf2                             (II-l)
Differentiating
                                AX  =  2 N  ANf                          (H-2)
In Equation II-2, ANf and AX are the linearized variables.  2Nf is the gain of
the square function.  Next, the open-loop gain and phase lag versus frequency
curves were plotted (Bode plots).   A Nichols chart was used to determine closed-
loop frequency response from the open loop Bode plot.

Figure AII-2 shows open- and closed-loop Bode plots of the inner control loop with
pressure feedback, K .   The gain of this loop varies with engine speed by 2.7
(9 db) over the operating range.  The gains must be selected for adequate damping
at the highest engine speed.  The  highest open-loop gain for stable operation at
3800 rpm engine speed is 18 (25 db) as shown in Figure AII-2.  The required com-
bined value for the gain parameters K  and K, for 25 db open-loop gain is K  K, =
4.1 x 10~5.

Figure AII-3shows the Bode plots for the secondary transmission outer loop with
linearized flywheel speed feedback.  The maximum allowable gain at maximum engine
speed and maximum flywheel speed is 8.5 (18.5 db).  The required value of the
                                     II-3

-------
                        *»
                                                1/L2
                                                  TPS
Nf ^
Fig. AII-1   Linear  Block Diagram — Secondary Control System
                                                                                     MTI-12545

-------
-20
                                         Frequency, (Rad/Sec)
                                                                                              100
                 Fig. All-2   Bode Plots  -  Secondary Transmission  Inner Loop
                                                                                                            MTl-12339

-------
 20
                             Open Loop Cain at
                               Ne -  3800 RPM
                               Nf -  24, .000 RPM
  Closed Loop FH«se Lag at
          3800 RPM
          24,000 RFM
Ne  -  3800 RPM
                                                                                                           200
                                                                                                           ISO
                                                       7\
Closed Loop Gala it
     -  3800 RFM
  \Ne  -  380
  Nf  -  2*.
      000 RPM
                                                                                                           100
-20
                                                          \\
 X
                            50
                                          Closed Loop Gain at
                                               -  1400 RPM
                                               -  10.000 RPM
                                                                   \
   Closed Loop Gain at
    Ne  - 1400 RPM
    Nf  - 24.000 RPM
                                                     10
                                                   Rod/Sec
                             100
                  Fig.  AII-3    Bode Plots -  Secondary Transmission Outer Loop
                                                                                                                   MTI-12337

-------
gain parameters K   and K  to give 18.5 db outer loop gain is K  /K  = 0.264.

Primary Control System  — High Range
The linearized block diagram for the primary control system in high range is
shown in Figure AII-4-  This block diagram includes the engine and vehicle
dynamics.  It is necessary to simplify this model before applying linear analysis
techniques due to the number of feedback loops.

There are two parallel feedback paths from P..  (the hydrostatic differential
pressure) to the gain block R2 D_.  The gain through the vehicle is less than
the gain through the engine by a factor of about 100.  Therefore, the signal
path through the vehicle has negligible effect on primary control stability,
and this path can be eliminated from the model.  The signal through gain block
RI P./12 is small for normal values of P.. ; therefore, this path can also be
omitted from the analysis.

There are two feedback loops from engine speed, N , to the compressibility block.
Closing these loops gives the transfer function

                           N
                            e
                                        2c s + _i^ S2
                                        (0      io/2
                                         n      n
The parameters of the transfer function vary with the value of D.. as shown below
in /rad
0
+1.19
-1.19
C
0.52
0.35
1.03
Ll
rad/sec
158
80
80
C
0.085
0.17
0.17
The set of parameters having the greatest effect on stability is on the lower
line since the gain is highest and the natural frequency is lowest.  Those
parameters were used for the inner loop transfer function.  The open- and closed-
loop Bode plots for the outer loop are shown, in Figure AII-5.  The inner-loop
dynamics actually have negligible effect on the outer loop dynamics.  The value
                                    II-7

-------
i
oo
                                                         R1D1
                                      Actua tor
                                                             Compressibility
                                                                                           Engine
             + 1 -
ec
K,
                                                         + f +
                                                                    1/Ll

                                                                   l+TpS
                                                         R2D2
                                                                       *2R12D2
                                                                          12    12
                                                                                  Ripio
                                                                                   12
                                                                                               m
                                                                                                     R12


                                                                                                   RaR10
                                                                                        Vehicle
                                                                                 11  2
                                                                                 12
                                                                                             11
                                                                                                            +1
                       Fig. AII-4.    Linear  Block  Diagram  -  Primary  Control  System -  High Range

-------
   30
   20
   10
.o
•H
u
01
O
  -10
  -20
                                      'Open Loop Gain
                                                Open Loop Phase Lag
                                                                Closed Loop  Gain
                                              10

                                      Frequency, (Rad/Sec)
                                                                                        200
       60
       «


       01
       00
       4
        03

        (II
        4)
        M
        00
                                                                                         100
100
                  Fig.  AII-5.    Primary Control System Frequency Response
                                            IT-9

-------
of gain term K. which was used for this ^ode plot is K_ = 3.68 x 10

Primary Control System  —  Low Range
The linear block diagram for the primary control system in low range is shown in
Figure AII-6.  The signal paths through the vehicle inertia and R. P,/12 gain
block can be omitted for reasons explained in the high-range analysis.  Closure
of the inner feedback loop gives the transfer function
                          N               C0
                           e               2
                                       _     +
                                        0         2
                                       n2      »
where
     C2  -  1.35
     C2  =  0.25
    u) „  =  52 rad/sec
     n/

The outer loop is not closed over most of the low range since Hydrostatic Element
I, Displacement D , does not change.  At the upper end of the low range, the
outer loop is closed.  The outer loop is then the same as in the high range.  The
primary control system is therefore stable in low range.  At the upper end of the
low range, damping is a little lower than in the high range due to the higher
gain, and lower natural frequency of the inner loop.

Results of Stability Analysis
The stability analysis above thus indicates that both the primary and secondary
control systems are basically stable with the gain values established.

2.  Analog Computer Simulation
The objectives of the analog computer study were to:

    A.   Verify validity and accuracy of mathematical model.

    B.   Modify the gain values determined from the linear analysis, if required,
        to compensate for the effect of nonlinearities.

                                    11-10

-------
'ec
                                                     RLD1
                                Actuator
                                   dl
                                                     + 1'-
1/L1
                                                             1-t-T   S
R1D1
 12
                                                                            R1P10
                                                                              12
                                                                    J0S
                                                                           Vehicle
                                Engine
5N/oT
                    Fig.  AII-6.   Linear Block  Diagram -  Primary Control System - Low Range

-------
    C.  Study operation of system and determine whether compensation is
        required to meet performance ^cais without saturating pressures.

    D.  Determine the effects of variation of parameters.

The block diagram of the mathematical model used for computer simulation is shown
in Figure AII-7. The significant nonlinearities of the engine, vehicle, and hydro-
static transmissions are included in this model.  Only those transmission lossao
affecting dynamics were included.

The control functions were partially linearized, since the control configuration
was not established at the time of the computer study.

The system was simulated on an Applied Dynamics AD/4 Analog Computer.

The secondary control system was initially observed to be unstable for all values
of gain K,.   This instability resulted from two integrations within the secondary
control system.   The first integration is performed by the secondary piston
actuator and the second by the flywheel.  These two integrations provide 180
degrees phase lag at all frequencies, and thus result in instability.  A
mechanical feedback linkage around the secondary piston actuator was then simu-
lated to reduce the phase lag of this integration.  This feedback corrected the
instability.  Performance was not satisfactory however.   When the secondary
piston actuator acts as an integrator, the flywheel speed will continuously
change as long as a kinetic energy error exists.  With the feedback linkage,
integration is no longer performed by the secondary piston actuator.  A large
kinetic energy error is then required to obtain a significant change in flywheel
speed.  It is necessary to have a significant kinetic energy error at all speeds
to obtain a steady-state value of D,.  Therefore, the feedback linkage was con-
sidered unsatisfactory.

Pressure feedback from the secondary hydrostatic transmission 'around the secon-
dary piston actuator was then simulated.  The use of pressure feedback with
appropriate values of the gain parameters eliminated the instability in the
secondary control system which occurred as a result of the double integration.
                                   11-12

-------
ACCELERATOR
  PEDAL
                                   Fig.  AII-7   Full-Range Block Diagram  -  Engine Flywheel
                                                 Propulsion System
                                                                                                                              KTI-12539

-------
 The  Typo  A  control,  system  was  then  inveni igaued.   The  Type  A  control  system
 utilizes  ,-in  output  speed governor for  computing vehicle  kinetic  energy  as
 i 11 us trau-.cl  in  Figure D-3,  page  D-6 .    The  output  speed  input  to  the  secondary
 rontrol system  forms a positive  feedback, or  regenerative,  type  of  control  loop.
 When  the  secondary  control  system gain,  K  , was too  high, the  system  was diffi-
 cult  to control at  high vehicle  speed.   Once  the vehicle speed started  to in-
 crease, it would continue  to increase  until the computer overloaded.  Returning
 the  accelerator pedal to zero  would not  stop  this  runaway condition.  It was
 possible  to  bring the system under  control  by depressing the  brake  pedal.

 Reducing  K   eliminated the  runaway  condition  at high vehicle  speed.   However,
 with reduced K, gain, the  transient response  at low  vehicle speed was too slow.
 Several methods of  eliminating the  regenerative condition at high speed without
 reducing  the transient response  at  low speed were  considered.

 One  possible method is to  schedule  gain  K   as a function of vehicle speed.  At
 low  speed, K^ would be high to obtain good  transient response.  At  high speed,
 K  would  he  reduced as required  for stable operation.

 The  square function in the  output speed  input to the secondary loop is  the  cause
 of the gain  variation.  The effective gain of this input is proportional to out-
 put  speed as a result of the square function.  Eliminating the square function
would eliminate the variable gain and a  constant value of K, could  be used  at all
 speeds.   However, with this control configuration, flywheel speed would decrease
 Linearly with increasing vehicle speed.  The total kinetic energy would then vary
with vehicle speed.

Several other approaches were  also considered.  However, with any of  these
approaches,  the positive feedback loop seemed objectional.   An unexpected gain
variation could possibly  lead to loss of control.  Therefore, the B-type control
system configuration shown  in  Figure D-4, page D-6 , was selected in preference
 to the A-type.

The B-type control system uses scheduled vehicle velocity as a function of
accelerator pedal position  for the input to the secondary control system.   The
                                    11-14

-------
positive feedback loop, which is the major problem with the A-type control, do is
not exist in the B-type control.

The primary control system gain determined in the linear stability analysis was
found to be the optimum value.  The secondary control gains used were a little
lower than the values determined from the stability analysis.  The system was
stable with these gain values.

The computer study indicated that some type of dynamic compensation was required
in order to achieve the desired acceleration rates without exceeding the pressure
limits.  When fuel flow increases too rapidly, the engine leads the flywheel.
Engine lead results in engine power being used to accelerate the vehicle, and
also can result in the flywheel accelerating during a vehicle acceleration.
When flywheel torque increases too rapidly, excessive vehicle acceleration and
pressure saturation can occur.

Several compensation techniques were investigated, including lagged accelerator
pedal input signal, and lagged fuel flow.  The best results were achieved with
dynamic compensation of fuel flow and the secondary piston actuator.  These
dynamic compensators effectively modulate the fuel flow and secondary actuator
rate limits to achieve the desired acceleration rates without pressure satura-
tion.  There are several possible sources for the input signals to the dynamic
compensators, including primary piston actuator displacement, primary piston
actuator differential pressure, secondary piston actuator displacement, and fuel
flow.  The signals from each of these sources appeared to be effective in modu-
lating the dynamic compensators.  A more detailed study is required to select the
dynamic compensator input function.

Conclusions
    1.   Pressure feedback from the secondary hydrostatic transmission around
        the secondary piston actuator  is required to stabilize the secondary
        transmission.

    2.   The use of an output governor  for computing vehicle kinetic energy
        is undesirable.  This input to the secondary control forms a positive
        feedback loop with resulting control problems.

                                    11-15

-------
3.  Dynamic compensation of the fuel flow and secondary piston actuator
    is required.

4.  The system is stable and is capable of meeting the performance goals
    with the gain values selected and with dynamic compensation of fuel
    flow and secondary piston actuator.
                                11-16

-------
NOMENCLATURE



    D     displacement of hydrostatic unit, in /rad

                                                  2
   F,     gain of hydrostatic unit A cam block, in

                                                  2
   F      gain of hydrostatic unit 3 cam block, in

                            2
    J     inertia, ft Ib sec


   K_     primary control system gain, in/sec/rpm


   K.     gain parameter


  K       gain parameter, in/sec/ft Ib

                                                  2
  K       gain of hydrostatic unit 1 cam block, in


   K      pressure feedback gain, in/sec/psi
    P
                                                                3
   L      primary hydrostatic transmission leakage parameter, in /sec/psi

                                                                  3
   L      secondary hydrostatic transmission leakage parameter, in /sec/psi


    N     speed, rad/sec


   P      primary hydrostatic transmission pressure, psi


   P      secondary hydrostatic transmission pressure, psi

                       3
    Q     flow rate, in /sec


    R     transmission ratio


    S     d/dt, I/sec


   T      engine time constant (J  3N/8T),.sec


   T      compressibility time constant (V/LB), sec


3N/3T     reciprocal of slope of engine torque-speed curve, I/ft Ib sec
Subscripts (except as noted in Nomenclature)


    e     engine


    f     flywheel






                                    11-17

-------
o     primary transmission output




1     hydrostatic unit 1




2     hydrostatic unit 2




3     hydrostatic unit 3




4     hydrostatic unit 4
                               11-18

-------