DETAILED DESIGN
RANKINE-CYCLE POWER SYSTEM
WITH ORGANIC-BASED WORKING FLUID
AND RECIPROCATING EXPANDER
FOR AUTOMOBILE PROPULSION
VOLUME I - TECHNICAL REPORT
Prepared for
Division of Advanced Automotive Power Systems Development
Environmental Protection Agency
Ann Arbor, Michigan
Prepared by
Thermo Electron Corporation
Research and Development Center
101 First Avenue
Waltham, Massachusetts
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Report No. 4134-71-72
DETAILED DESIGN
RANKINE-CYCLE POWER SYSTEM
WITH ORGANIC-BASED WORKING FLUID
AND RECIPROCATING EXPANDER FOR
AUTOMOBILE PROPULSION
Edited by:
Dean T. Morgan, Program Manager
Prepared by: Rankine Power Systems Department
Edward F. Doyle, Manager
Robert J. Raymond, Expander Development
Ravinder Sakhuja, Heat Exchanger Development
Herb Somi, System Integration and Packaging
William Noe, Controls Development
Chi Chung Wang, Performance Analysis
Andrew Vasilakis, Combustor Development
Lucb DiNanno, Feedpump and Rotary Shaft Seal Development
Thermo Electron Corporation
Research and Development Center
85 First Avenue
Waltham, Massachusetts 02154
Prepared tor:
Division of Advanced Automotive Power System Development
Environmental Protection Agency
Ann Arbor, Michigan
Contract No. EHS 70-102
Work Performed: May 6, J970 - November 5, 1971
Report Issued: May 5, 1972
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THERMO BIBCTROM
CORPORATION
TABLE OF CONTENTS
Chapter Page
1 ABSTRACT 1-1
2 SUMMARY 2-1
2. 1 INTRODUCTION , 2-1
2.2 SYSTEM DESCRIPTION AND CHARACTER-
ISTICS 2-1
2.3 COMPONENT DESCRIPTIONS 2-16
2. 4 MAJOR CONCLUSIONS 2-40
3 INTRODUCTION 3-1
3. 1 PROGRAM GOALS AND HISTORY 3-1
3.2 TECHNICAL BASIS OF SYSTEM 3-3
4 SYSTEM DESCRIPTION AND CHARACTERISTICS . . 4-1
4. 1 INTRODUCTION 4-1
4.2 WORKING FLUID-LUBRICANT, SYSTEM
SCHEMATIC, AND DESIGN POINT
CONDITIONS , 4-4
4. 3 SYSTEM INTEGRATION AND PACKAGING
IN 1972 FORD GALAXIE 4-21
4.4 ACCELERATION PERFORMANCE AND
FUEL CONSUMPTION CALCULATIONS 4-30
4. 5 EMISSION PROJECTIONS FROM RANKINE-
CYCLE SYSTEM 4-39
REFERENCES FOR CHAPTER 4 4-44
5 COMPONENT DESCRIPTIONS 5-1
5. 1 EXPANDER-FEEDPUMP-TRANSMISSION
SUBASSEMBLY -. 5-1
5.2 COMBUSTION SYSTEM-BOILER SUB-
ASSEMBLY 5-90
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THERMO ELECTRON
CORPORATION
TABLE OF CONTENTS (continued)
Chapter Page
5 5.3 REGENERATOR 5-120
5.4 CONDENSER SUBASSEMBLY 5-126
5.5 STARTUP SEQUENCING, SAFETY CONTROLS
AND ACCELERATOR PEDAL LINKAGE 5-140
5.6 BOOST PUMP-INDUCER-RESERVOIR
SUBASSEMBLY 5-151
5.7 ACCESSORY AND AUXILIARY COMPONENTS. . 5-160
IV
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THERMO ELECTRON
CORPORATION
1. ABSTRACT
The Division of Advanced Automotive Power System Development
of the Environmental Protection Agency (EPA) is sponsoring part of
the development of a low-emission Rankine-cycle propulsion system
for automobiles at Thermo Electron Corporation (TECO); the Ford
Motor Company is contributing both financially and technically to the
development effort.
The system under development at TECO is based on use of an
organic-based working fluid with reciprocating expander. The working
fluid used is FluorinolrSS, a mixture of 85 mole percent trifluorinol
and 15 mole percent water. In this report, a description is presented
of the detailed, optimized design of the system including packaging of
the complete system in the reference car, the 1972 Ford Galaxie.
The results of experimental development in several critical areas are
also presented. The system is designed to provide performance
approximately equivalent to use of a 351 cubic inch displacement
internal combustion engine in the reference car. Performance cal-
culations indicated that a system designed to provide 131 hp net shaft
horsepower (feedpump and condenser fan power subtracted) at an
expander speed of 1800 rpm (equivalent to 90 mph vehicle speed) pro-
vides acceptable performance for the 1972 Ford Galaxie. Some pre-
dicted performance characteristics are:
Wide-Open Throttle Performance:
0-60 MPH, 0% Grade 13.4 seconds
Grade for 70 MPH Constant Speed 6. 8%
Top Speed, 0% Grade 103 MPH
1-1
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CORroRATION
Fuel Consumption:
Steady Speed - 30 MPH 15.4 MPG
60 MPH 11.8 MPG
Federal Driving Cycle for Emissions 10.8 MPG
Emission measurements were made with a burner designed for
a 100 shp Rankine-cycle system. The measurements were made with
the burner operated transiently to simulate operation of a Rankine-
cycle system over the Federal driving cycle for emission measure-
ments. The Federal specified procedure for the measurements was
followed exactly including use of the constant volume sampling unit.
The measurements confirmed the low emission potential of the
Rankine-cycle system:
Emission Level, grams/mile
Pollutant
NOX
CO
UHC
The next phase of the program involves development of all
components and testing of the complete preprototype system in the
laboratory.
Measured
0.29
0.22
0. 14
Federal 1976 Standard
0.4
3.4
0.41
1-2
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THBRIK10 BIBCTROM
CORPORATION
:;2. SUMMARY
2. 1 INTRODUCTION
The Division of Advanced Automotive Power System Development
of the Environmental Protection Agency (EPA) is sponsoring develop-
ment of a low-emission Rankine-cycle propulsion system for automo-
biles at Thermo Electron Corporation (TECO). The system under
development at TECO is based on use of an organic -based working
fluid with reciprocating expander. In this report, a description is
presented of the detailed, optimized design of the system including
packaging of the complete system in the reference car, the 1972 Ford
Galaxie. The results of experimental development in several critical
areas are also presented. The system is designed to provide per-
formance approximately equivalent to use of a 351 cubic inch displace-
ment internal combustion engine in the reference car. Experience
gained in 1-1/2 years of testing of a complete 5-1/2 hp Rankine-cycle
power system at TECO provides a firm technical basis for the design.
The same working fluid and similar cycle conditions are used for the
automotive system design as in the 5-1/2 hp system.
The Ford Motor Company (FOMOCO) is contributing both financially
and technically to the development effort, particularly in the areas of:
(1) integration of the system into the 1972 Ford Qalaxie, (2) manufactur-
ing considerations, (3) expander design, and (4) transmission design.
2. 2 SYSTEM DESCRIPTION AND CHARACTERISTICS
Performance calculations indicated that a system designed to
provide 131 hp net shaft power output (feedpump and condenser fan
power subtracted) at an expander speed of 1800 rpm (equivalent to
2-1
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THBRMO KIBCTRON
CORPOBATION
90 mph vehicle speed), and with condenser and condenser fan charac-
teristics based on ram air resulting from 90 rpm vehicle speed,
provided acceptable performance for the 1972 Ford Galaxie relative
to the EPA specifications. All component sizes are, therefore, based
on this design point condition. The working fluid used is Fluorinol-85;
the state point diagram for the system at the design point is presented
in Figure 2. 1.
The system schematic including all control functions is presented
in Figure 2. 2. The driver interface to the system is limited to the
ignition switch, accelerator pedal, gear shift lever, and brakes as in
present automobiles. System startup and operation are completely
automatic.
2.Z.I System Integration and Packaging in 1972 Ford Galaxie
In the design of the components described in Section 2. 3, an
essential input has been packageability of the complete system in
the 1972 Ford Galaxie with only minor modifications to the vehicle.
The complete system layout is illustrated in Figure 2.3, a side view
looking from the driver's side of the engine compartment, and in
Figure 2.4, a view from the top of the engine compartment. Sectional
views are provided in Figures 2. 5, 2. 6 and 2. 7, as identified in
Figure 2. 3. As can be seen from these drawings, the expander-
feedpump-transmission subassembly is located to the rear of the
engine compartment, the condenser is located in the very front of
the engine compartment with the condenser fans mounted to the
condenser shroud on the rear of the condenser, and the combustion
system-boiler subassembly is located between the expander and
condenser and is placed as close as possible to the expander, leaving
&
Halocarbon Products, Inc. , Hackensack, N. J.
2-2
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IT-Z48-F
|P- 4O Mia 0- 4I400O Btu/hr
!•"• •'• /• "i"~"^
,' REGENERATOR ,'
I-' '•• '-' '.. I
T=248 °F
P=39 pile
,e(HHV)=8I.O%
T=2I4 "F
P-627 MM
0=188 « lO'BTWhr
AIR FLOW
-I730OCFH AT
85* AMBIENT
-75700 Ib/hr
FAN SHAFT POWER
-7.0 dp
CONDENSER
x.
X
3
T=ZIO'F
?- 35 Olio
DIRECT
DRIVE
FROM
H EXPANDER
COMBUSTION «IR
2168 IB/h.
9.0 l« WC,
2.31V SHIFT
^ % NET SHAFT HP = 131.1
(GROSS LESS FEEDPUMP
AND CONDENSER FANS)
cr*
WF85= 29.4 GPM
BOOST PUMP,
SHAFT POWER = O.5 hp
INDUCER
WFB5= 13.5 GPM
ATOMIZING AIR
COMPRESSOR
5.5 CFM
ISpsig
O.4 tip SHAFT POWER
wf 85 ' l5'9 GPM ° 9B6° lt/hr
CYCLE EFFICIENCY = 14.75%
OVERALL EFFICIENCY (HHV)= 12.0 %
Figure 2. 1 State Point Diagram.
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I
.*•
REGENERATOR
( SECTION B)
HIGH
PRESSURE
BYPASS
VALVE
LEGEND:
ORGANIC
FUEL, AIR OR
LUBRICANT
CONTROL
IGNITION
SWITCH
TRANSMISSION
AND
ACCESSORIES
IR
fllTER a
WfTRIum
^•••••••••••1
^
^••••••1 Xg
STARTER
AlUMI&INto
AIR
CONTROL
*
ATOM 1 7 IMft
AIR
BLOWER
STARTUP
SEQUENCING
CONTROLS
•
•
ro
BURNER
FUEL
CONTROL
>:
•
f
COMWSTION]
AIR 1
CONTROL
HYDRAULIC
SUPPLY
4
COMBUSTION
AIR
SERVO
CONTROL
IMBUSTWM I
, AIR |+MA
•LOWER r A
( SAFETY I
|CONTROLS!
SYMBOLS:
PT - PRESSURE, INDICATING ORGANIC TEMPERATURE
PBD - PRESSURE, BOILER DISCHARGE
Apc - PRESSURE DIFFERENTIAL, SIGNAL ORIFICE
Xa - DISPLACEMENT, ACCELERATOR
N - SPEED, EXPANDER CRANKSHAFT
0V - DISPLACEMENT, AIR CONTROL VALVE
Kg - GAIN FUNCTION, FEEDBACK
TAC - TEMPERATURE, CONDENSER DISCHARGE AIR
Figure 2.2 System Schematic.
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BOILER
CM
Ul
j— JCCELEK XTOff LINKAGE
/ AND TK^SDUCER
GNITION SYSTEM
AK CONDITIOHER
RECEIVER
'^ THREE SPEED
; AUTOMATIC TRANSMISSION
Figure 2. 3 Side View, Packaged System.
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STO. HEATER
AIR CONDITIONER CASE
srsrow
CONDENSER
POriER S JEfRING PUMP
'«
«
CSJ
CT^
•*»
00
~1
CHITION SYSTEM
CONDENSER FANS
AIR CONDITIONER
CONDENSER
Figvire 2. 4 Top View, Packaged System.
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S7Q HEATER /AIR CONG/TONER
to
I
-J
EXPANDER INTAKE
VALVE OPERATOR
I
IS)
STARTER MOTOR
Figure 2. 5 Section A-A from Figure 2. 3.
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ro
i
oo
POWER STEERING
PUMP
©
FUEL LINE to BURNER
Figure 2. 6 Section B-B from Figure 2. 3.
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ro
i
SO
CONDENSER FAN
IGNITION SYSTEM
HYDRAULIC VALVE
CONTROL
'• TEM CONTROL
Figure 2. 7 Section C-C from Figure 2. 3.
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THERMO ELECTRON
CORPORATION
sufficient flow area for exhaust of the condenser cooling air. The
regenerator is located directly above the expander and is mounted
to the expander. All auxiliary components for the Rankine-cycle
system as well as accessory components for the normal automotive
functions are also packaged in the system.
2.2.2 Acceleration Performance and Fuel Consumption Calculations
Computer models of the Rankine-cycle system and the vehicle
have been used in calculating the acceleration performance and fuel
consumption of the vehicle over typical driving conditions. In Figures
2. 8 and 2.9, performance maps of the Rankine-cycle power system
are presented in the form of net shaft horsepower vs. expander rpm
with lines of constant efficiency shown. The maps are based on use
of a special two-speed transmission designed by the Dana Corporation;
the first map applies to "L.o" gear with a high expander rpm relative
to vehicle speed, and the second to "Hi" gear with a low expander rpm
relative to vehicle speed.
The acceleration performance and gradabillty of the vehicle are
presented in Tables 2, 1 and 2.2, respectively. These estimates are
based on a vehicle curb weight, fully-fueled, of 4276 Ibs, with 300 Ibs
passenger load for acceleration performance and 1000 Ibs passenger
load for gradability. The vehicle with the Rankine-cycle system, which
weighs 210 Ibs more than the same vehicle with the 1972 Ford 351 CID
internal combustion engine, meets EPA specifications for acceleration
and gradability.
The fuel consumption is presented in Table 2. 3 both for steady
speed on 0% grade and for three driving cycles: the Federal driving
2-10
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140
120
100
(T
UJ
180
UJ
CO
a:
o
x
£60
^
(O
40
20
FIRST GEAR WITH DANA TRANSMISSION
IR MAX =0.325
FLUORINOL-85 WORKING FLUID
FULL THROTTLE
12%
3 %
N
-J
200 400 600 800 1000 1200
EXPANDER SPEED, RPM
1400
1600
1800
2000
Figure 2. 8 Performance Map with Transmission in First Gear
(High Expander Rpm Relative to Vehicle Speed).
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IN)
I
C\)
SECOND GEAR WITH DANA TRANSMISSION
IR MAX = 0.325
FLUORINOL-85 WORKING FLUID
200
400
600
800 1000 I2OO
EXPANDER SPEED, RPM
1400
1600
t\>
1800 2000
Figure 2.9
Performance Map with Transmission in Second Gear
(Low Expander Rpm Relative to Vehicle Speed) .
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THBHMO KJ.KCTMOH
CORPORATION
TABLE 2. 1
VEHICLE ACCELERATION PERFORMANCE
Vehicle Test Weight
Ambient Temperature
Dana Transmission
4576 Ibs
85°F
Two Speed
0-60 mph
0-10 seconds
25 - 70 mph
Passing, 50 - 80 mph
System Performance
13.36 sec
457. 9 ft
15.0 sec
15.4 sec
EPA Spec
« 13.5 sec
1 440 ft
£ 15.0 sec
S 15.0 sec
£ 1400 ft
2-13
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THBRIMO ELECTRON
CORPORATION
TABLE 2.2
GRADABILITY
Vehicle Test Weight 5276 Ibs
Ambient Temperature 85 °F
Dana Transmission Two Speed
Vehicle Speed, mph
0
10
15
20
30
40
50
60
70
103
Grade %
System Performance
35. 3% :
34. 6%
30. 7%
26. 8%
19. 8%
13. 8%
11.3%
8. 97%
6. 84%
0%
EPA Spec.
s 30%
i? 30%
a 30%
-
-
-
-
-
Grade for 70 mph
-
2-14
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THKRMO BLBCTRON
CORPORATION
TABLE 2 3
FUEL CONSUMPTION
Vehicle Test Weight 4576 Ibs
Ambient Temperature 85 °F
Dana Transmission
Constant Speed 0% Grade
MPH
30
40
50
60
70
80
85
Constant Speed, 5% Grade
60
. Two Speed
MPG
15. 38
15 15
13.52
11. 76
10. 33
8. 54
7 84
5.68
70
4.99
Driving Cycles
Federal Driving Cycle (or Emissions 10. 81 mpg
FOMOCO Suburban Cycle 11.41 mpg
FOMOCO City Cycle 8. 61 mpg
FOMOCO Customer Average 10.01 mpg
2-15
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THERMO ELECTRON
CORPORATION
cycle for emission measurements, and the Ford Motor Company
suburban and city driving cycles. The FOMOCO customer average
is the arithmetic average of the FOMOCO suburban and city driving
cycles.
2.2.3 Emission Levels Measured Over Federal Driving Cycle
To demonstrate the potential of the Rankine-cycle automotive
propulsion system for very low emission levels, Thermo Electron'
Corporation has made emission measurements on a full-size burner
for a 100 shp automotive system, operated transiently over burning
rates, corresponding to operation of the vehicle over the Federal
driving cycle for emission measurements. The fuel/air control
used in these transient tests was similar to that to be used in the
automotive system. The burner fired into a water-cooled "boiler"
with approximately the same configuration as in the system. The
procedure for making the emission measurements on the exhaust
from the boiler was identical to that of the Federal Register and
included use of the three-bag constant volume sampler.
The te'st results with a ceramic-lined burner are summarized
in Table 2.4. The gram/mile emission levels are below the 1976
Federal standards by a factor of 1.4 for NO , 15.4.Tfor CO, and 2.,9
for UHC. Steady-state'measurements indicate that use-of exhaust '
gas recirculation (EGR) will result in even lower NO emission rates;
j£
thus, EGR will be used on the system.
2. 3 COMPONENT DESCRIPTIONS
A description of the component designs and characteristics is
presented in this section. These components are identical to those used
in the system packaging.
2-16
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TABLE 2.4
EMISSION LEVELS MEASURED OVER
FEDERAL DRIVING CYCLE
Emissions
(grams/mile)
NO
X
CO
UHC
Transient
Test .
Result*
0.29
0.22
0 14
Federal 1976
Standard j
0. 4
3.4
0.41
Actual gas mileage used for tests was 12. 1 mph. The
latest performance calculation predicts 10. 8 mpg for the
Federal emission test driving cycle and the measured
emission levels have been increased by 12% to reflect
the change in fuel economy
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THBRMO BLBCTRON
CORPORATION
2.3.1 Expander- Feedpump-Transmission Subassembly
2,3.1.1 Expander Design'
The expander is a single-acting V-4 with 4.42 inch bore, 3. 00 inch
stroke, and total displacement of 184 in , The expander speed at a
vehicle speed of 90 mph is 1800 rpm. Variable cutoff intake valving
is used for power control from the expander to maximize: (1) wide-
open-throttle acceleration performance and (2) the part-load efficiency
of the system. Power is controlled completely by the expander variable
cuttoff intake valving; no throttle valve is used between the boiler and
expander. The automatic exhaust valve is similar to that used on
Thermo Electron's 5-1/2 hp expanders and permits exhaust over most
of the piston return stroke, thereby maximizing the power output per
unit of expander displacement.
The front and side layouts of the expander are illustrated in
Figures 2. 10 and 2. 11, respectively. With the exception of the
valving, the expander construction is similar to that of internal com-
bustion engines. The block and cylinder head are cast iron, the piston
and connecting rod are cast aluminum, and the crankshaft is cast
steel. Needle bearings are used throughout to reduce initial develop-
ment difficulties. Spray lubrication is used. The lubricating oil is
thermally stable and compatible with the working fluid at the peak
cycle temperature of 550 °F, so that oil blowing by the expander can
be allowed to pass through the boiler with no deleterious effects on
the system.
The variable cutoff intake valves for the preprototype expander,
which are hydraulically actuated, are being developed by the American
Bosch Corporation. The valve actuator was constructed and bench-
testing during the program by American Bosch. With a hydraulic
2-18
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DO
I
if'*/« Ol* HCAO KK.TS
§
Figure Z. 10 Expander Layout with American Bosch Valving
Cross Section Through Front Cylinders.
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Figure 2. 11 Expander Layout with American Bosch Valving -
Cross Section Through Rear of Expander Showing
Feedpump and Oil Pump Drive.
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THERMO ELBCTROM
CORPORATION
pressure of 1500 psi, the full-lift opening and closing times are 2. 0
and 3.0 milliseconds, respectively. The intake valves are 1.25 inches
in diameter with 0, 3 inch lift. In addition to this system, a detailed
design of a mechanically (cam-driven) actuated valve system was
developed and is presented in the report. A second hydraulic system,
financed by Thermo Electron Corporation, is under experimental
development.
The lubricating oil for the expander is used as the hydraulic fluid
for the valving system. The hydraulic pump and a separate oil reser-
voir for the valve system are integrated in the expander design.
2.3.1.2 Feedpump
The main system feedpump is a radial, 7 cylinder, reciprocating
piston pump. The pump is directly driven by the expander and is
integrated with the expander, as illustrated in Figure 2.10. Since, at
any expander speed, the required system pumping rate can vary from
zero to a maximum of about 16 gpm, depending on the intake valving
cutoff point (or system power output), the pump is variable displace-
ment to minimize the feedpump power requirement at part-load con-
ditions. An additional development goal was maintaining high pump
efficiency over a wide speed and pumping rate range.
The feedpump design is illustrated in the cross section of
Figure 2. 12. The seven cylinders are located radially around the
axis of the rotating shaft. Each cylinder houses seal-ring pistons
which can be varied from zero to full stroke by axially moving the
angled portion of the shaft through the center eccentric ring. The
cylinders receive fluid from a common inlet plenum and discharge
2-21
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THERMO ELECTRON
CORPORATION
into a common outlet plenum. Both the inlet and outlet valves are
simple spring-loaded washer types. The piston shoes ride on a
septagonal ring, with a circular ring used to return the pistons.
The cylinder bore is 1. 50 inches, stroke is 0.466 inch, and
maximum displacement is 5. 76 inches. The pump delivers 15 gpm
with full stroke at a speed of 600 rpm. The design discharge pressure
is 850 psia. The pump has an aluminum housing with steel pistons
and steel drive mechanism,
The pump has been constructed and tested; some of the test
results are presented in Figures 2, 13 and 2. 14. The volumetric
efficiency varies from ~ 98% at low speeds to 71% at the maximum
speed of 1800 rpm. The overall pump efficiency (hydraulic power/
shaft power) varies from about 82% at low speeds to 68% at high
speeds. The efficiency at a given speed is insensitive to discharge
pressure and relatively insensitive to the pump displacement. Pres-
sure pulsations are very small and acceptable over the entire range
of operating speed and displacement tested.
Boiler outlet pressure is controlled by varying the pumping rate
to the boiler through control of the pump displacement. The feedpump
design of Figure 2, 12 includes a spool control valve operated by a
spring-biased diaphragm to which boiler outlet pressure is directly
applied. This spool valve controls application of the feedpump
discharge fluid to a power piston which is connected to the stroke
control shaft of the pump and is used to vary the displacement.
2.3.1.3 Transmission
The system uses a conventional, three-speed automatic trans-
mission with 12 inch diameter torque converter coupling (FOMOCO
2-22
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1-2677
Figure 2. 12 Reciprocating Piston Pump with Variable
Displacement - Cross Section.
2-23
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1-2680
OPERATING CONDITIONS'
P|N = 9 to 18 PSIA
°-pouTs 4I5 PSIA
X~POUTS 5I5 PSIA
» 615 PSIA
s 7I5 PSIA
A~POUTS 5I5 PSIA
-N=500RPM
600RPM
100 -
90 -
3* 80 -
• 70 -
I 60
UJ
S£ 50
u.
u.
uj 40
30 -
20 -
1
0 -
0
1.0
2.0
3.0
4.0
DISPLACEMENT ( IN5)
1
5.0 5.57
Figure 2.13 Efficiency vs. Displacement for 7-Cylinder Feedpump.
2-24
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10
9
8
t-
LJ
UJ
O
<
110
100
>
i
"90
80
70
z
CM —
Q60
dD
a. 4
3
2
o
UJ
o
40
- iZ30 —
UJ
20
10
* - - DISPL
00
I
I
2OO
400
600 800 1000
SPEED-N ( RPM)
1200
I4OO
1600
I80C
Figure 2. 14 Efficiency, Displacement, and Shaft Power
Input vs. Speed for 7-Cylinder Feedpump.
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THERMO ELECTRON
CORPORATION
C-4 automatic transmission). A planetary speedup gear with 1:2. 25
speed ratio is used between the expander and transmission. Gear
ratios in the transmission are 2.40 low, 1.47 intermediate, and
1.00 high. The driveline is standard for the 1972 Ford Galaxie with
2.75:1 axle ratio.
2.3.1.4 Rotary Shaft Seal
The system has been designed so that only one dynamic shaft
seal, that on the 3 inch diameter expander shaft, is required. To
positively prohibit either air leakage into the system or working
fluid leakage out of the system, a double shaft seal is used with
prepressurized buffer fluid between the two seals. The system
lubricating oil is used as the buffer fluid so that any leakage into
the system through the seal is the system lubricant. The two seals
are of the face-seal type.
Bench-testing of two designs of the full-size seal has been
carried out under simulated system conditions with excellent results.
Over a test period of 3187 hours in one test, the average leakage
rates of the lubricating oil were 0. 183 pints/1000 hrs operation into
the crankcase, and 0.255 pints/1000 hrs operation through the out-
board seal. Leak rates in the shutdown mode are approximately
a factor of 10 lower than the operating leakage rates.
2. 3, 2 Combustion System-Boiler Subassembly
The assembly cross sections of the combustion system-boiler
subassembly are presented in Figures 2. 15 and 2. 16. These drawings
include the boiler, burners, combustion air blower and motor drive,
atomizing air compressor, and fuel/air controls.
2-26
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POWER
DIAPHRAGM
COMBUSTION AIR
CONTROL VANE
ATOMIZING
AIR COMPRESSOR
BOILER TUBES
O
00
COMBUSTION
CHAMBER
COMBUSTION
BLOWER
AIR ATOMIZING
NOZZLE
FUEL
SOLENOID VALVE
c
T
Figure 2. 15 Front View of Combustion System-Boiler Subassembly.
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1-2709
WORKING
FLUID
EXIT
DC MOTOR
Figure 2. 16 Side View of Combustion System-Boiler Subassembly.
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THERMO ELECTRON
CORPORATION
2.3.2.1 Boiler Design
The boiler tube bundle is designed for a maximum heat transfer
rate of 2.25 x 10 Btu/hr with 81% efficiency, based on the fuel higher
heating valve (HHV). At 10% load, the boiler efficiency rises to 88%
(HHV). The flow path of working fluid and combustion gases through
the boiler is illustrated in Figure 2. 17. At the design point, com-
bustion gases enter the bottom of the boiler at 2975°F and exit from
the top at 600°F. The working fluid enters the preheat state as
liquid at 287°F and 826. 5 psia and exits from the superheat stage as
superheated vapor at 550°F and 700 psia. The maximum tube wall
temperature on the working fluid side is 569°F. The design point
characteristics of the boiler stages are summarized in Table 2.5.
The boiling and superheat stages of the boiler are bare tube
bundles with a water jacket buffer to positively prevent either gross
or local overheating of the working fluid. Dual tube construction is
used for these stages, as illustrated in Figure 2.15, with organic
flowing through the inner tube. The annular space (~60 mil gap)
between the tube bundles is sealed and filled with water, with an
external thermal expansion tank to permit thermal expansion of
the water. Heat transfer from the combustion gases to the organic
occurs by boiling the water on the inner surface of the outer tube
and condensing the water vapor on the outside of the inner tube. The
organic tube wall temperature can therefore not exceed the saturation
steam temperature corresponding to the water jacket pressure; this
pressure provides a convenient and sensitive means of controlling
the maximum temperature to which the working fluid is exposed. The
boiling and superheat stages are brazed construction with the tubes
brazed into machined steel headers.
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THERMO ELECTRON
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The preheat stage is a conventional, finned-tube heat exchanger.
A water jacket is not used for this stage, since the combustion gas
temperature is relatively low and the tubes are filled with liquid
Fluorinol-85.
The fins, tubes, and headers for all stages are constructed of
AISI4130 steel.
2.3.2.2 Burner and Fuel/Air Supply Designs
Two burners firing in parallel are used to provide a burning rate
(HHV) of 2. 78 x 10 Btu/hr with JP-4 fuel. The design turndown ratio
is 20:1. With reference to Figure 2. 15, the cylindrical combustion
chamber is air-cooled with combustion air entering at the top of the
combustion chamber and flowing down the space between the combustion
chamber and outer wall of the burner. At the bottom of the burner,
the air flow reverses direction and flows through swirl vanes into the
combustion chamber. The fuel nozzle is an air-atomizing Sonicore
nozzle. Near the nozzle, the combustion chamber is lined with a
ceramic insert. The combustion chamber wall is flared outward above
the main combustion zone to diffuse the combustion gases and provide
a uniform gas velocity over the entire area of the rectangular boiler
tube bundle. To insure proper balance of fuel/air flow between the
two burners, the air and fuel flow paths from the common fuel control
and common air control are symmetrical.
The combustion air blower is a cross-flow or transverse type. The
air. blower provides a pressure head of 9 in. W. C. at the design air
flow rate of 770 CFM and a mean mix temperature of 165 °F. The
design is based on 20% excess air (2468 Ibs/hr at 60°F) and 20%
exhaust gas recirculation (521 Ibs/hr at 600°F). A gerotor-type fuel
2-30
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EXHAUST GAS FLOW
t
U)
PRP Up AT
STAGE
LIQUID >.
INLET
SUPERHEAT
STAGE
BOILING >
^^x *\ *\ S\ S\ ^\
I Y~~~^
-/ ^ ^
^ Ji 1 JL L L L JL JL JL JL 1 ,
Q) Q) G) (D O (D 0) (DOG) G) (D () ()
T T T^.T^ T T TTv« " T
o o^cb o (b cb ()"*c
' ' f £ ' '
^ or>r^r> ortoortr
FLOW
NO.
~ ^rx
>^ VAPOR
u) ( ) (b 3
O <)• 4
•> f* — : > s
t
COMBUSTION GAS FLOW
FROM BURNER
I
^>
OJ
ro
Figure 2. 17 Flow Path Schematic of Working Fluid Through Boiler.
-------
TABLE 2. 5
BOILER DESIGN POINT CHARACTERISTICS
Stage
Boiling
Superheat- 1
Super heat- II
Preheat
Total
Tube
Row
No.
5
4
3
1 and 2
Heat Transfer
Rate
Btu/hr
625,000
444, 800
224,400
958,000
2.25 x 106
Combustion Gas
Temperature
Inlet
°F
2975
2371
1939
1709
Outlet
°F
2371
1939
1709
607
FL-85
Temp.
In
°F
437
445
502
287
Out
°F
445
502
550
437
Pressure Drop
Gas Side
In w. c.
2.38
0.74
0. 11
0. 17
3.40
Organic
Side
psi
40
30
32.5
24
126.5
Water
Jacket
Pressure
psia
1470
1405
1431
-
Mass of
Core
Without
Water
Ibs.
66
40
40
80
226
Mass of
Water
Ibs
2.7
2. 7
2.7
-
8. 1
TUBE SPECIFICATIONS
Boiling Stage
Inner Tube - 1. 00" O. D., 0. 083" wall
Outer Tube - 1. 313" O. D. , 0. 093" wall
Superheater I and II Stages
Inner Tube - 5/8" O. D., 0.049" wall
Outer Tube - 7/8" O. D. , 0. 058" wall
Preheat Stage
5/8" tube expanded (0.577" O. D. , 0.035" wall)
18 fins/inlet (rippled)
Tube and Header Material = AISI 4130 Steel
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THERMO ELECTRO
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pump is used to provide constant pressure (25 psig) fuel to the burner
fuel metering valve. The atomizing air compressor is of the rotary-
vane type.
The functional schematic of the burner controls is presented in
Figure 2.2 and the location of the control components on the burner-
boiler are illustrated in Figures 2.1.5 and'2e 1'fc. Detailed designs of the
control elements are presented in Chapter 5 of this report. The
burner controls are designed to provide rapid response to large .
power transients of the system, regulating'the burning rate to a level
corresponding to the new power level in a time of about 200 milli-
seconds. A &P created across.an orifice in the inlet Mine to'the boiler is
applied to the air servo control; this control is diaphragm-actuated,
thus providing rapid response to organic flow rate changes to the
boiler and an approximate balance between the organic flow rate
entering the boiler and the burning rate. For fine tuning of the
burning rate, a thermal-expansion temperature sensor responding
to the vapor temperature is provided. This sensor generates a
pressure signal, proportional to the organic temperature, which is
also applied to a diaphragm in the combustion.air servo control. The
combustion air servo control modulates the air flow to the burner in
response to these input signals. For low emissions, reasonably
tight control on the fuel/air ratio must be provided. The fuel control
is thus directly linked to the combustion air control.
2.3.3 Regenerator
The regenerative heat transfer rate at the design point is 414,000
Btu/hr, with vapor entering and leaving at 375°F and 239°F, respectively,
and liquid entering and leaving at 208°F and 287°F. The regenerator
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THBRMO BJ.BCTRON
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design is illustrated in Figure 2. 18. The core is brazed aluminum
construction with flat tubes carrying the liquid and with fins on both
the liquid and vapor sides of the core. The liquid and vapor flow paths
are illustrated in Figure 2. 19.
The regenerator is also used as an oil separator to collect and
return to the crankcase a major portion of the oil droplets in the exhaust
vapor from the expander. Provision is made for gravity return to the
expander crankcase of the separated oil.
2.3.4 Condenser Subassembly
In the design of the condensing subassembly, a prime goal has been
minimizing the parasitic fan power. To meet this goal, (1) the maximum
condenser frontal area which can be packaged in the 1972 Ford Galaxie
without major modifications has been used; (2) the condenser fan drive
includes a control to optimize the fan speed under part-load and low
„•
vehicle speed operating conditions, and (3) an inducer is used to main-
tain the condenser free of condensed liquid so that the entire condenser
core is effective for condensation.
The condenser design is illustrated in Figure 2,20. The condenser
is T-shaped, as illustrated, in order to provide the maximum condenser
frontal area without modifying the frame at the front of the car. The
central condenser section sits between the two frame members and the
two side condenser sections sit on top of the frame members (see
Figure 2. 7). Total core frontal area is 8. 21 ft and the core thickness
is 4. 3 inches. At the design point rate of 1. 88 x 10 Btu/hr, the air flow
fate is 17; 300 CFM at 85 °F, the air side pressure is 3. 5 in. W. C. ,
the ideal air power with the fans on the downstream side is 11.2 hp,
2-34
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EXHAUST GAS FLOW
t
PREHEAT
STAGE
LIQUID
INLET
STAGE
CS)
Ul
STAGE
VT
_
^
(EAT
S >
1 FLOW
A: N°
N - t "
f > > > VAPOR
^(j)(f)n)(p()(t)O(pOd)(t)u)(b(b(b(p(p ^
£ £ •£.
<() o o d> o d> <)(>(> o o o c» o i 4 *
^ c\ f\ f\ r> o r>r>.r>r> r> r> •> s
-J
(JO
t
COMBUSTION GAS FLOW
FROM BURNER
Figure 2. 17 Flow Path Schematic of Working Fluid Through Boiler.
-------
1-2728
VAPOR OUT
c
LIQUID IN
VAPOR
IN
I
c>
LIQUID
OUT
Figure 2. 19 Liquid and Vapor Flow Paths
in Regenerator.
2-36
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OJ
-vl
L_
— — \
HI
r
Tl
Figure 2. 20 Condenser Design Layout.
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THERMO BLBCTRO
and the organic side pressure drop is 5 psi. The condenser sections
are constructed with top and bottom headers connected by flat tubes;
both internal and external finning is used. The condenser construction
is brazed aluminum and the design is suitable for one-step brazing
for ease in manufacture.
As seen in Figure 2.7, the three condenser fans, one 24 inches
in diameter and two 16 inches in diameter, are required to provide
uniform condenser air flow and the high cooling air flow rate. The
hub thickness of the fans is 2. 5 inches and the design, point rpm's
are 2400 and 4200 for the 24 inch and 16 inch diameter fans respectively.
The fans are designed to provide 1. 5 inches W. C. head at the design
point, with 2. 0 inches provided by ram air at the vehicle design point
speed of 90 mph.
For optimizing the fan speed, a Morse variable-speed belt drive
is used in conjunction with an Eaton Tempatrol viscous clutch. The
variable speed drive uses a centrifugally-controlled sheave to vary
the expander-fan drive speed ratio. This control provides a constant
fan/expander speed ratio of 4. 91:1 up to 550 rpm expander speed.
Above an expander speed of 550 rpm, the centrifugal sheave maintains
a constant fan drive speed of 2700 rpm irrespective of expander speed.
The viscous clutch provides control for part-load conditions as
well as ambient temperature variation. This clutch modulates fan
speed by sensing the air temperature leaving the condenser and,
if the air temperature is low (indicating excess air flow and fan
power),the clutch slips, thereby reducing the fan speed. One clutch,
constructed integrally with the central 24-inch fan, provides control
for all three fans, since the 16-inch fans are belt-driven from the
clutch rim.
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2.3.5 Startup Sequencing, Safety Controls, and Accelerator Pedal
Linkage
Startup sequencing controls are provided for completely automatic
system startup, requiring the driver only to turn on the ignition switch.
Safety controls are provided to protect the system in case of abnormal
operation. These include flame sensing to cut off fuel in case of
flameout and overpressure and over-temperature cut-off switches.
The American Bosch hydraulic valving system is electrically
controlled. An electrical interface through a LVDT is used between
the accelerator pedal and the intake valving control system.
2.3.6. Boost Pump - Inducer - Reservoir Subassembly
The flow schematic of this subassembly is indicated in Figure 2.21.
The components of this subassembly provide the following functions:
(1) produce the NPSH required by the feedpump during normal operation
and during startup; (2) eliminate required condenser subcooling, thereby
making more efficient use of the available frontal area for condensing;
(3) provide reservoir capacity for working fluid inventory transfer during
start condition and transient operation; and (4) prevent separation of
lubricant from working fluid in condenser and reservoir.
The centrifugal boost pump, illustrated in Figure 2. 22, is designed
for a flow rate of 29. 4 rpm with head rise of 26. 6 feet. The minimum
NPSH is 10 inches. The shaft power required is 0. 5 hp. The pump
is driven by the accessory drive shaft. To eliminate the requirement
for a dynamic shaft seal, a permanent magnet drive is used. The pump
is constructed primarily of aluminum.
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THERMO BIBCTROM
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The aluminum reservoir has a capacity of 1. 16 gallons and is
located at the very top of the engine compartment (see Figure 2.4
and 2.6).
The-inducer is of conventional construction; nozzle flow is taken
from the centrifugal boost pump.
2.3.7 Accessory and Auxiliary Components
The system includes power steering, power brakes, heating, and
air conditioning; these components are identical to those now used in
automobiles.
The battery-alternator supplies electrical power to both the
system and the normal automotive functions requiring electrical
power. The electrical system is 12 Vdc.
An analysis of battery-alternator requirements for both starting
and sustained operation at high powers was made to establish the
size of these components. For high-power system operation, electrical
power is drawn from both the battery and alternator. The battery is.
an AABM size 24C 84 amp hr battery; the 14 volt alternator is a
standard, heavy-duty type and provides 130 amps at 5000 rpm and
90 amps at 2000 rpm (idle condition).
2.4 MAJOR CONCLUSIONS
• A Rankine-cycle automotive propulsion system based on a
reciprocating expander and organic based working fluid and
competitive in performance to a 1972 351 CID internal com-
bustion engine can be completely packaged in the engine
compartment of a 1972 Ford Galaxie with only minor internal
sheet metal modifications in the engine compartment required.
2-40
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VENT
CONDENSER
IS)
INOUCER
RESERVOIR
FEED PUMP
I
r\>
-j
Figure 2.21 Boost Pump-Inducer-Reservoir Subassembly.
-------
1-2745
V
k \\\\ N\\ V
y ////////
ss
'////(^/^/^
^v\\\\ vvvv
/\
>
\\:
S.I
DIA
Figure 2. 22 Cross Section of Centrifugal Boost Pump.
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THBRMO ELECTRON
The customer-average fuel consumption with the 1972
Ford Galaxie is predicted to be 10 mpg.
Transient burner tests on an automotive size burner confirm
the potential of the Rankine-cycle automotive propulsion
system for very low emission levels of NO ,. CO, and un-
burned hydrocarbons.
Thermo Electron's system design 'is based on operation of
a complete 5-1/2 hp system with the .same boiler outlet
temperature as that used for the automotive?system. This
4
experience provides a firm technical basis fpr the system
design presented in this report.
Onlylow-cost materials are used in the system and the
system design is adaptable to high volume production
techniques.
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THERMO ELECTRON
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3. INTRODUCTION
3. 1 PROGRAM GOALS AND HISTORY
The Division of Advanced Automotive Power System Development
of the Environmental Protection Agency (EPA) is sponsoring develop-
ment of low-emission power systems for automotive propulsion systems.
The primary goal of these development programs is demonstration of
one or more power systems which: (1) produce very low gram/mile
emission levels of unburned hydrocarbons, carbon monoxide, nitrogen
oxides, and particulates, and (2) fulfill all other requirements for an
automotive power system. As part of this program, the development
of a Rankine-cycle power system for automobiles is underway at
Thermo Electron Corporation. The effort at Thermo Electron is
funded partly by EPA and partly by the Ford Motor Company through
funds made available under a Ford-TECO business agreement. In
addition to its financial support, the Ford Motor Company is providing
substantial technical support to the development effort.
The system under development at Thermo Electron Corporation
is based on use of an organic-based working fluid with reciprocating
expander. Work on this system started in June, 1969; this report
presents the results of work performed between June 1970 and November
1971. The program history is summarized below.
3. 1. 1 June 1969 - June 1970: Conceptual Design Study
The preliminary designs of components for a 100 shp power plant
and package drawings for installation of the powerplant in a 19^9 Ford
Torino, an intermediate-sized family car, were prepared. A mockup
of the powerplant was also fabricated and installed in the engine
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TMBRMO ELECTRON
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compartment of a 1969 Ford Torino. Computer models for the com-
ponents and system were prepared and used in performance and fuel
consumption estimates. Thiophene was used as the reference working
fluid.
Results of this study are described in the final report issued in
June 1970. l
3.1.2 June 1970 - November 1971: Detailed Design and Experimental
Development^in Selected Areas
This phase of the program involved conversion of the conceptual
design to a much more detailed, optimized design and experimental
development in several of the more crucial areas, specifically:
a. Analysis and bench-testing of full-size, variable cut-off
expander intake valving system.
b. Analysis and bench-testing of full-size expander exhaust
valve.
c. Simulated testing of full-size rotary shaft seal.
d. Heat transfer and pressure drop measurements on ball
matrix fin and evaluation of its use for very compact
heat exchangers.
e. Construction and testing of a variable-delivery feedpump
based on the conceptual design study.
f. Bearing-lubricant testing with thiophene working fluid.
The detailed design developed in this phase and the experimental
results obtained are described in this report. The main body of the
report is the final detailed design; the experimental results are
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THERMO ELECTRON
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described in the appendices, except for the expander test results,
which are covered in the main body of the report.
During this phase of the program, two significant changes were
made. The reference car size was increased by EPA from the inter-
mediate size (Ford Torino) to the full size (Ford Galaxie) car and
2
new EPA performance specifications were issued. These changes
required an increase in the system design point power level from
100 shp to 131 hp. At the beginning of this phase, thiophene was the
reference working fluid. In January, 1971, after approximately 4
months of system testing in a 5-1/2 hp system with Fluorinol-85, the
decision was made to convert from thiophene to Fluorinol-85 as the
system working fluid. This change was made primarily because the
safety characteristics of Fluorinol-85 are superior to those of thiophene
and the thermodynamic properties are almost as good.
3.1.3 Starting November 5, 1971: Experimental Development Testing
of Preprototype System
In this phase of the program, the complete system, as described
in this report, is to be experimentally developed and tested as a pre-
prototype system. The schedule calls for initial testing of the complete
system in December, 1972.
3. 2 TECHNICAL BASIS OF SYSTEM
While its potential for very low emission levels is the main
reason for consideration of a Rankine-cycle propulsion system, any
new powerplant for automotive application must fulfill many other
characteristics, if it is to be seriously considered for large-scale
use in automobiles. These other characteristics are:
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• Manufacturing cost.
• Reliability and maintenance requirements.
• Driver convenience.
• Acceleration performance
• Fuel economy.
• Packageability and weight.
• Safety.
• Startup and operation over ambient temperature range
of -40°F to 125°F.
The selection of the system under development at Thermo Electron
was based on attempting to meet all of these characteristics in order to
have a competitive and practical system. The most important system
considerations which influence meeting these characteristics are
working fluid and peak cycle temperature and expander type.
The automotive system design presented in this report is based
on experience gained in 1-1/2 years' testing of a 5-1/2 hp system
at Thermo Electron Corporation. The working fluid,lubricant,
peak cycle temperature, most materials of construction, expander
construction, automatic startup sequencing, etc. , are based on
this test experience. The emphasis behind the 5-1/2 hp system
development is commercialization of Rankine-cycle systems in the
5 - 20 hp range for applications such as fork lift trucks, engine
generator sets, and other applications where low noise and low pollution
are important. Since the prime competitor for these applications is
the internal combustion engine, the development goals for the small
horsepower Rankine-cycle systems are similar to those for the
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THERMO ELECTRON
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automotive system where the prime competitor is also the internal
combustion engine.
*
In selecting the working fluid, Fluorinol-85 , for these systems,
several requirements were set, based on the anticipated use in
commercial systems:
a. The maximum acceptable freezing point is -40°F.
b. With respect to flammability, the fluid must be self-
extinguishing. Also, vapor-air mixtures should either
be non-explosive or have a very mild reaction.
c. The vapor inhalation and dermal application toxicities
should be acceptable for use in the development labora-
tories at Thermo Electron Corporation with no special
requirements other than good ventilation.
d. The maximum working fluid cost when produced in large
volumes must be no greater than $1 - $2 per pound.
e. The working fluid must be compatible with low-cost
materials of construction.
f. A lubricant which is thermally stable and compatible
with the working fluid at peak cycle temperature must
be available.
Fluorinol-85 is the best available working fluid which meets all of
these requirements in a system with reciprocating expander while
*
Halocarbon Products, Inc. , Hackensack, New Jersey.
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THERMO ELECTRON
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still providing a useable cycle efficiency. This fluid is used in the
5-1/2 hp system being tested at Thermo Electron Corporation at a
boiler outlet temperature of 550°F, and can probably be used at boiler
outlet temperatures up to 600°F. It is expected that fluids with higher
thermal stability than Fluorinol-85 will become available in the future,
permitting higher cycle efficiencies while still retaining all of the
required characteristics for the working fluid. If the flammability
requirements were relaxed, some currently available fluids would
provide a higher cycle efficiency.
The expander used is of the reciprocating type with variable
cutoff intake valving. The primary advantages of this type of expander
for this application are its low shaft speed without gearing and main-
tenance of high expander efficiency over a wide speed and power range,
thus permitting use of a relatively simple and inexpensive transmission.
In addition, production technology for internal combustion engines is
directly applicable to the reciprocating expander, and the optimum
fluids for a Rankine-cycle system with reciprocating expander require
a small regenerator size.
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THBRMO ELECTRON
CHAPTER 3
REFERENCES
1. Morgan, D. T. and Raymond, R. J. , "Conceptual Design,
Rankine-Cycle Power System with Organic Working Fluid and
Reciprocating Engine for Passenger Vehicles, " Report No.
TE4121-133-70, June 1970, Thermo Electron Corporation,
Waltham, Massachusetts.
2. "Vehicle Design Goals - Six Passenger Automobile, " Revision
C, Division of Advanced Automotive Power Systems Development,
Environmental Protection Agency, issued May 28, 1971.
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4. SYSTEM DESCRIPTION AND CHARACTERISTICS
4. 1 INTRODUCTION
The propulsion system design is based on meeting the perform-
ance and vehicle specifications established for the Advanced Automo-
tive Power Systems Program by the Division of Advanced Automotive
Power Systems Development of the Environmental Protection Agency.
The EPA vehicle specifications are for a full-sized American auto-
mobile, such as the Ford Galaxie 500 sedan, and the performance
specifications are approximately equivalent to those for a 351 CID
i
internal combustion engine with 1970 emission controls and three-speed
automatic transmission. A summary of the primary EPA performance
specifications is given in Table 4. 1.
The vehicle selected as the reference car for system packaging
and for performance and fuel economy estimates is the 1972 Ford
Galaxie 500 four-door sedan. Performance calculations presented in
this chapter indicate that a system designed to provide 131 hp net shaft
powe r output (feedpump and condenser fan power subtracted) at 90 mph
vehicle speed provides acceptable performance for the reference car
relative to the EPA specifications. All component sizes are therefore
based on this design point condition.
The system design has been developed to provide driver con-
venience equal to that provided by current automobiles. The driver
interface is limited to the ignition switch, accelerator pedal and
gear selector; system startup and operation are completely automatic.
A very important consideration in the system design has been insurance
of startup and proper system operation, regardless of environmental
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TABLE 4. 1
PERFORMANCE SPECIFICATIONS FOR ADVANCED AUTOMOTIVE
POWER SYSTEMS PROGRAM
A. GENERAL REQUIREMENTS
Startup: 65% of full power in 45 seconds at 60 °F ambient
temperature.
Self-sustaining idle within 25 seconds at -20 °F
ambient temperature.
Idle Fuel Consumption: Not to exceed 14% of fuel consumption
rate at maximum power conditions.
Performance Degradation with Ambient Temperature:
All performance specifications are to be degraded by
no more than 5% at ambient temperature of 105 °F.
B. ACCELERATION REQUIREMENTS; VEHICLE WEIGHT = FULLY
FUELED VEHICLE PLUS 300 LBSJ 0% GRADE
Acceleration from Standing Start at 85 °F Ambient Temperature:
Distance in 10 seconds §440 ft
0-60 mph time £ 13. 5 seconds
Acceleration in Merging Traffic at 85 °F Ambient Temperature:
25 to 70 mph time 5 15. 0 seconds
Acceleration, DOT High Speed Pass Maneuver at 85 °F Ambient
Temperature:
Initial
Condition
50 mph
(truck ]Q-». so mph
Final v H8fth 100ft »Ptjnr»l
Condition^
[carp—*- speed §80 mph
[truck )D-»- 50 mph
100 ft
Time to Accomplish Maneuver ^15 secpnds
Distance Traveled by Automobile 5 1400 feet
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TABLE 4. 1 (continued)
C. GRADABILITY AND REQUIRED MAXIMUM SPEED REQUIRE-
MENTS AT 85°F AMBIENT TEMPERATURE; VEHICLE
WEIGHT = FULLY FUELED VEHICLE PLUS 1000 LBS
Grade
Vehicle Speed
30%
5%
0%
Start from rest and accelerate
to 15 mph
60 mph continuous
65 mph for at least 180 seconds
after acceleration from 60 mph
70 mph for at least 100 seconds
after acceleration from 60 mph
85 mph continuous
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THBRMO ELECTRON
conditions or orientation of the vehicle. Consideration has thus been
given to conditions such as startup with the car parked uphill or down-
hill on a steep grade and startup at low ambient temperatures when
the internal system pressure is very low (< 0. 1 psia).
In the system design and packaging the complete powerplant has
been packaged in the existing engine compartment of the 1972 Ford
Galaxie with only minor modifications to the base vehicle. In addition,
the standard accessory equipment for passenger convenience and
comfort has been retained, and packaged with the Rankine-cycle system.
These functions include power steering, power brakes, air conditioning,
and heating.
In the remainder of this chapter, a description is given of the
overall system integration and packaging in the 1972 Ford Galaxie, and
of the car performance in fuel economy projections. In Chapter 5, the
detailed designs of the system components are described.
4.2 WORKING: FLUID-LUBRICANT/ SYSTEM SCHEMATIC, AND
DESIGN POINT CONDITIONS
4. 2. 1 Working Fluid-Lubricant
The selected working fluid is Fluorinol-85, a mixture of 85 mole
percent trifluoroethanol and 15 mole percent water. The characteristics
of the working fluid are summarized in Table 4. 2. This working fluid
is currently produced in industrial quantity by Halocarbon Products Cor-
poration, Hackensack, New Jersey. Fluorinol-85 has been used at
Thermo Electron since September, 1970 in testing of small horsepower
(5-1/2 hp) Rankine-cycle power systems with satisfactory results.
A lubricating oil which is thermally stable and compatible with
low-cost materials of construction and the working fluid at the peak
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THERMO ELECTRON
TABLE 4.2
WORKING FLUID CHARACTERISTICS
Chemical Composition
Average Molecular Weight
Freezing Point
Boiling Point
Critical Temperature
Critical Pressure
Flammability Characteristics
Toxicity Characteristics
Liquid Density (ZOO°F)
Lubricant
Thermal Stability and
Material Compatibility
85 Mole Percent Trifluoroethanol
CF CH OH, 15 Mole Percent Water
(Fluorinol-85)
87. 74
-82°F
165°F
452°F
800 psia
Non-Supporting, Non-Explosive
Not classified as toxic via dermal
or inhalation pathways (MCA, 1970) '
1.25 gms/cm
Commercial Refrigeration Oil -
Immiscible with Fluorinol-85
Demonstrated at 550 °F (Boiler outlet
. temperature) by operation of complete
power system for 450 hours
Capsule tests indicate potential for
use at higher boiler outlet temperature
4-5
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THERMO ELECTRON
CORPORAT1O
cycle temperature of 550°F is used for lubrication of the expander and
feedpump bearings and sliding surfaces. Because of its thermal
stability, it is not necessary to absolutely prohibit the oil from passing
through the boiler and high-temperature side of the expander; it is
expected that a small fraction (< 1%) of the circulating flow through the
condenser and boiler is the lubricant. Satisfactory experience has
been obtained with use of this oil in complete Rankine-cycle systems
operating with the same boiler outlet temperature as the automotive
system. The vapor side of the regenerator is used as an oil separator
to return oil to the crankcase and maintain the circulating oil fraction
at an acceptable level to prevent serious degradation of the regenerator,
condenser, and boiler performances.
The lubricant is immiscible with the working fluid and has a
density less than that of the working fluid. Consideration in the system
design and integration has therefore been given to insure that no lubri-
cant traps occur in the system, resulting in an inadequate lubricant
level in the crankcase. The immiscibility of the working fluid-lubricant
facilitates adequate and positive lubrication of the expander-feedpump
bearings on startup and provides a supply of a working fluid - free oil
to the expander hydraulic valving pump.
4.2.2 System Description and Design Point Conditions
4. 2. 2. 1 Basic Cycle
The basic components and cycle conditions comprising the Rankine-
cycle powerplant are illustrated in terms of a flow schematic and a
T-S diagram in Figure 4. 1. The Fluorinol-85 vapor leaves the boiler
at a temperature of 550°F and a pressure of 700 psia (State Point 1) .
This vapor is expanded through the reciprocating expander, producing
shaft power applied to the load. The exhaust vapor leaves the expander
4-6
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I-Z653
FUEL
Working Fluid
Shaft Power
---- Fuel
600
960
520
480
44O -
400 -
360-
u.
".- 3ZO -
| 280-
I
240 -
200
160 -
120-
40-
P- 43 psio
P-39psia
-0.1
0.1
02 0.3 0.4
Entropy. Blu/lb *R
0.5
0.6
0.7
O.B
Figure 4. 1 Illustration of Basic Components and Cycle for Rankine-cycle
Power System with Fluorinol-85 Working Fluid.
4-7
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THBRIMO ELECTRO
CORPORATION
at a pressure of 43 psia and temperature of 376°F in a superheated
condition (State Point 2) . Because of the relatively high temperature
of the vapor, the overall cy.cle efficiency can be improved by using a
regenerative heat exchanger in which energy is transferred from the
expander exhaust vapor (State Point 2-* State Point 3) to the feed liquid
going to the boiler (State Point 5 -• State Point 6) , thereby reducing the
fuel requirement for a given system power output. The vapor leaves
the regenerator at a pressure of 39 psia and enters the air-cooled con-
denser, where the vapbr^i's.completely condensed at an average tem-
perature of 21 5°F (State Point 3- State Point 4) .
The condensed liquid, then enters the feedpump, where the fluid
pressure is raised from the condenser discharge pressure of 35 psia
.' • »* ." •
to 820 psia. (State Point 4.- State Point 5). The liquid then flows
through the regenerator, liquid side (State Point 5 -• State Point 6) and
enters the boiler at State Point 6. Energy from the combustion of fuel
is then transferred to the Fluorinol-85 flowing through the boiler,
producing the high pressure-high temperature vapor at State Point 1
and completing the basic cycle.
4. 2. 2. 2 System Description
In arriving at a complete system for automotive propulsion, many
alternative choices are available on the approach to be followed in
synthesis of the system. In this section, a listing is presented of
the various components and subassemblies making up the complete
system, and the logic underlying the selected approach is briefly out-
lined. In Chapter 5, a detailed description of the various components
comprising the complete system is presented.
4-8
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THERMO Ei-ECTROM
CORPORATION
The system block diagram is illustrated in Figure 4. 2 and includes
all components or subassemblies making up the system. For purposes
of discussion, here as well as in Chapter 5, the various components are
grouped in subassemblies based either on function (as the accessory and
auxiliary components) or on preassembled and integrated units to be
installed in the system (as the expander-feedpump-transmission sub-
assembly).
a. Expander-Feedpump-Transmission Subassembly; The expander
is a reciprocating piston type with hydraulically operated intake valving.
The method of controlling power output from the expander is to vary the
cut-off point (or intake ratio) of the expander intake valving with vapor
at full boiler pressure supplied to the intake valves. Thus, for high
power or torque outputs, a large intake ratio is used; for low power
outputs, a small intake ratio, is used. The intake valve control is
directly controlled by the accelerator pedal, as illustrated in Figure 4. 2.
The hydraulically-operated intake valves used in the system can be con-
trolled down to the point of zero lift with no vapor flow through the
expander. Thus, no throttle valve is required between the boiler and
expander.
An alternative control procedure for the expander power output is
use of fixed cut-off intake valve timing with a throttle valve to reduce
the working fluid pressure entering the expander. The reasons for
selection of the variable cut-off approach, discussed in the June 1970
4
report, are: (1) to maximize the wide-open-throttle acceleration
performance of the system with given component sizes, and (2) to maxi-
mize the system efficiency in the low-power and low-speed range where
an automobile operates on the average, thereby providing good fuel
economy.
4-9
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•kCfmmmmmmmi
V,
0
N
0
E
N
S
E
R
()
\
u
— (
.rf'
VISCOUS
CLUTCH
«1
REGENERATOR
( SECTION B)
OIL
SEPARATOR
LEGEND:
FUEL, AIR OR
LUBRICANT
!••••••• CONTROL
IGNITION
SWITCH
STARTUP
SEQUENCING
CONTROLS
COMBUSTION
AIR
SERVO
CONTROL
• •••••••I Kgbmmmmmmmmm
Ul
SYMBOLS:
PT - PRESSURE. INDICATING ORGANIC TEMPERATURE
pBD - PRESSURE, BOILER DISCHARGE
Apc- PRESSURE DIFFERENTIAL, SIGNAL ORIFICE
Xa - DISPLACEMENT. ACCELERATOR
N - SPEED. EXPANDER CRANKSHAFT
flv - DISPLACEMENT. AIR CONTROL VALVE
Kg - GAIN FUNCTION. FEEDBACK
TAG * TEMPERATURE. CONDENSER DISCHARGE AIR
Figure 4. 2 System Schematic.
-------
TMBRMO ElBCTROM
conroiiATioN
Included as part of the expander is an intake ratio Limit control
and a governor using expander speed as input. At any given expander
speed, as the cut-off point of the intake valving is increased, the
organic flow rate increases. Since the boiler heat transfer rate
required is approximately linear with organic flow rate, assuming
constant boiler outlet pressure and temperature, the intake ratio
limiter automatically prevents exceeding the intake ratio al. which
the boiler capacity would be exceeded with a resultant drop in boiler
outlet pressure and temperature. Since, as discussed below, the
expander is allowed to idle at zero vehicle speed in order to drive
the automotive accessories, a governor is required to maintain the
expander idle speed under varying power demands by automatic
adjustment of the intake ratio.
An automatic transmission is used between the expander and
vehicle drive shaft. The transmission serves two functions: it permits
the expander to idle at zero vehicle speed, to operate the vehicle
accessories, and it supplies either two or three gear ratios to provide
improved wide-open-throttle acceleration response from a.given sys-
tem. Its'function is therel-or-?. identical =1:o the transmission used in
I/C engine-powered cars.
The feedpump is directly driven by the expander at expander speed
and integrated with the expander. This procedure eliminates any
dynamic seal for the feedpump. With the variable intake expander
valving, the organic flow at a given expander speed can vary from
zero to a maximum corresponding to the boiler capacity. A variable
displacement piston pump is used to provide the proper organic flow
rate in an efficient manner over the complete power-speed range of
the system, thereby minimizing the feedpump power at any operating
4-11
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THBMMO BILBCTROM
CORPORATION
point. The boiler outlet pressure is used to vary the feedpump dis-
placement through a servo-control and force amplifier to maintain
constant boiler outlet pressure.
b. Combustion System-Boiler Subassembly; This subassembly
includes the burners, boiler, burner controls, combustion air blower,
fuel supply, and atomizing air compressor. The combustion air
blower and atomizing air compressor are driven by an electric motor
,f
operating from the battery on startup or from the alternator during
normal operation. This requirement results from the desire to
operate the burners at peak burning rate during startup to minimize
the cold startup time. In order to have reasonable motor, battery,
and alternator sizes, as well as to reduce the parasitic load on the
system, a prime consideration in the design of the burner-boiler
unit has been low combustion side pressure drop within the packaging
restrictions for the 1972 Ford Galaxie
The burner fuel/air control has been designed for rapid response
to transient power changes of the system.and regulation of the burning
rate to a level corresponding to the new power level in the fastest
practical time. The approach followed to provide rapid response to
power transients is use of an orifice in the liquid organic feed line to
the boiler; the AP created across this orifice by the organic flow is
then applied to the air servo control, providing rapid response to
organic flow rate changes to the boiler. Thus, for a power increase
resulting from depression of the accelerator pedal by the driver,
the vapor flow rate through the expander and from the boiler increases.
This results in a transient decrease in boiler outlet pressure; the
feedpump displacement is then automatically increased by the boiler
4-12
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THERMO ELECTRON
CORPORATION
outlet pressure control, bringing the organic flow rate to a higher
level. The increased organic flow rate provides an increased orifice
£P to the combustion air servo control, which responds in approxi-
mately 200 milliseconds to bring the burner rate to the level corres-
ponding approximately to that required for the organic flow rate.
Power decreases result in the inverse of the above.
The orifice control input provides an approximate balance between
the organic flow rate entering the boiler and the burning rate. For fine
tuning of the burning rate to insure constant boiler outlet temperature,
a temperature sensor responding to the vapor temperature leaving the
boiler is provided. This sensor generates a pressure signal, propor-
tional to' the organic temperature, which'.is also applied to the com-
bustion air servo control.
The combustion air servo control modulates the air flow to the
burner in response to the input signals. Feedback for stability is also
provided in the servo control. For low emissions, reasonably tight
control on the fuel/air ratio must be provided. The fuel control is
thus directly linked to the combustion air control.
The burner uses an air atomizing fuel nozzle. For low emissions,
it may be desirable to control the atomizing air pressure for different
burning rates. Current tests at Thermo Electron indicate that this
control may not be required with constant atomizing air pressure used
at all burning rates.
c. Regenerator^ The regenerator is located directly above the
expander and mounted to the expander exhaust flange. The regenerator
is also used as the system oil separator, returning most of the oil in
the expander exhaust vapor to the crankcase. A screen separator is
used in the vapor inlet header to remove a major fraction of the oil
4-13
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THERMO BQ.BCTROM
CORPORATION
before the flow enters the regenerator vapor-side fins in order to
minimize the effect of the oil film on the regenerator performance.
The vapor side fins of the regenerator act as an additional separator.
The separated oil drains by gravity back to the expander crankcase.
The oil leaving the regenerator in the vapor passes through the
condenser, pumps, boiler, and expander back to the regenerator.
d. Condenser-Condenser Fans-Condenser Fan Drive and
Controls Subassembly; The power required for the condenser fans
represents the largest parasitic load to the system; strong considera-
tion has thus been given in the condenser-condenser fan selection to
minimizing the fan power for a specific condensing rate. In addition,
a fan drive and control system has been devised which approximately
optimizes the fan speed for peak system performance under all
operating conditions. The drive and controls are made of commerciaily
available parts.
The condenser fans are directly driven by the expander through
a centrifugally-controlled variable speed belt drive. This drive
provides a high speed ratio (high relative fan speed) at low expander
speeds and low speed ratio at high expander speeds (low relative fan
speed) for maximum utilization of ram air. A thermostatically-
controlled viscous clutch operating on the air temperature leaving
the condenser provides additional control. This viscous clutch is
identical to those now used on some automotive I/C engines.
4-14
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THERMO ELECTRON
e. Startup Sequencing and Safety Controls; Startup sequencing
controls are provided for completely automatic system startup re-
quiring the driver only to turn on the ignition switch. The startup
sequencing first operates the burner at full firing rate for rapid
boiler heatup. When the boiler has been heated to a preset tempera-
ture, the starter or cranking motor is switched on, turning over the
expander and feedpump until the system becomes self-sustaining. The
startup can be safely and effectively accomplished, even if the boiler
is initially completely dry or initially filled with working fluid.
Sufficient safety controls are provided to prevent damage to the
system and vehicle if system malfunction occurs* These controls
include functions such as high pressure and high temperature shut-
down of the system and flame sensing to stop fuel flow if flame-out
occurs.
f. Inducer-Receiver-Boost Pump Subassembly: These com-
ponents, while not part of a basic Rankine-cycle system, are extremely
important for reliable startup and operation of the system under any
possible operating conditions of the vehicle and for proper handling of
the lubricant passing through the condenser, that is, for preventing
accumulation of oil in the condenser. The use and design of these
components is based on startup and operating experience with the
completely automatic 5-1/2 hp systems constructed and tested at
Thermo Electron. For application in a car, the condensate header
at the bottom of the condenser is the lowest part of the system, and,
if the vehicle is parked on a steep downhill grade, can be lower than
any part of the system.
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THBRMO ELBCTRO
CORPORATION
The feedpump is a piston pump with spring-loaded suction valving
and requires an NPSH of several feet for operation of the suction valving
and for pump operation without cavitation. On cold startup, the NPSH
to the pumping system, must be provided solely by liquid head; in ad-
dition, the internal pressure in the system on cold startup is normally
very low (0.08 psi at 0°F). During operation, the required NPSH can
be provided by subcooling of the liquid from the condenser. However,
this procedure detracts from the frontal area available for the con-
denser, requires an additional air-cooled heat exchanger, reduces the
system efficiency, and is not applicable for system startup. By utiliz-
ing the centrifugal boost pump to pressurize the feedpump suction to
a pressure — 5.0 psia,positive feedpump operation under all conditions
without cavitation can be guaranteed without subcooling. The centri-
fugal boost pump is designed for an NPSH of 10 inches, which the
receiver provides under all conditions. The boost pump also provides
flow for operation of the inducer. The boost pump is driven by the
condenser fan drive, as illustrated in Figure 4.2. To eliminate a
dynamic seal, a standard magnetic drive is used for the boost pump.
The inducer operates as a sump pump and maintains the con-
denser free of liquid, maximizing the condenser performances as well
as prohibiting accumulation of oil in the condenser. The inducer is
operated by flow from the boost pump and pumps directly into the
receiver.
The receiver provides allowance for working fluid inventory
changes in the system and guarantees liquid supply to the boost pump
and feedpump under all conditions. Under normal operation, the
receiver would be about three-fourths filled with liquid. Under ab-
normal conditions or on start-up, when liquid might accumulate in
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THBMMO BIBCTROH
CORPORATION
the condenser or some other location in the system, the receiver
provides sufficient liquid volume to insure liquid priming of the boost
pump. The receiver is located at the top of the engine compartment
and provides 10 inches head to the boost pump when completely empty.
g. Accessory and Auxiliary Components: .Auxiliaries are defined
as other components required for operation of the system, and acces-
sories are components required for the vehicle operation and for
passenger convenierfce and comfort. These components include the
alternator, battery, starter or cranking motor, air conditioning and
heating of passenger compartment, power steering, and power brakes.
In arriving at the alternator size for the system, consideration has
been given to the worst operating conditions for the vehicle to insure
sufficient alternator capacity for the electrically-operated system
components as well as normal vehicle requirements.
4.2.2.3 Design Point Conditions
Component sizes have been based on a net shaft horsepower output
of 131. 1 hp (gross shaft power less condenser fan power and feedpump
power) at a vehicle speed of 90 mph and an expander speed of 1800 rpm,
with ambient temperature of 85 °F. Performance calculations indicate
that this system power output and components sized on this basis satisfy
the EPA performance specifications, as outlined in Table 4. 1 when
installed in the 1972 Ford Galaxie.
In Figure 4.3, the state points at the design point are illustrated
on the system flow schematic; in Table 4. 3, the design point conditions
for the major components are summarized. The design point conditions
were calculated from the computer model of the complete system.
4-17
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oo
0= 1.88 i lO'BTLVhr
AIR FLOW
-17300 CFM AT
85° AMBIENT
-757OO Ib/hr
FAN SHAFT POWER
-7.0hp
IPs 4Q nit Q. 414000 Btu/hr
-.-. .1 ..-. .-I—\
! REGENERATOR ,'
= 2I4 °F
NET SHAFT HP - 131.1
(GROSS LESS FEEDPUMP
AND CONDENSER FANS)
INOUCER
WF85= 13.5 GPM
CYCLE EFFICIENCY = 14.75%
OVERALL EFFICIENCY (HHV)= 12.0 %
U)
Figure 4. 3 State Point Diagram.
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T H • R M O BIBCTROM
CORPORATION
TABLE 4.3
DESIGN POINT CONDITIONS
Vehicle
Boiler
Outlet Temperature
Outlet Pressure
Heat Transfer Rate
Efficiency (HHV)
Burning Rate (HHV)
Expander
Intake Ratio
Speed
Gross Shaft Power
Displacement
Bore (V-4)
Stroke (V-4)
Overall Efficiency
Regenerator
Heat .Transfer Rate
Effectiveness
Vapor Temperature
Liquid Temperature
Condenser
Heat Transfer Rate
Average Pressure
Average Condensing
Temperature
Ambient Air Temperature
Air Flow Rate .
Effectiveness
Fan Power (with Utilization
of Ram Air at 90 mph)
1972 Ford Galaxie
550°F
700 psia
2.26 x 106 Btu/hr
81.0%
2. 78 x 106 Btu/hr
0. 175
1800 rpm
146.6 hp
184 in3
4.41 in
3-00 in
75%
414..000 Btu/hr
81%
376°F -*248°F
214°F -283°F
1.88 x 10
37 psia
Btu/hi
214°F
85°F
75,700 Ib/hr
17,300 CFM Entering
80%
7 hp
4-19
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THBRMO BIBCTHO
CORPORATION
TABLE 4. 3 (continued)
Feedpump
Organic Flow Rate 9860 Ib/hr
15. 9 gpm
Efficiency 85%
Shaft Power 8. 5 hp
System
Net Shaft Power (Less Feedpump)
and Condenser Fan) 131. 1
Cycle Efficiency 14. 8%
Overall Efficiency (HHV) 12. 0%
4-20
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THERMO ELECTRO
CORPORATION
4. 3 SYSTEM INTEGRATION AND PACKAGING IN 1972 FORD
G A LAX IE
In designing the components described in Chapter 5, an essential
input has been packageability in the engine compartment of the 1972
Ford Galaxie 500 with only minor modifications. Considerable iteration
in the component designs has been required in fitting the system into
the car. The system packaging is described in this section, and in-
cludes every component of the system as well as power steering, power
brakes, and passenger compartment heating and air conditioning. It
has been possible to retain the standard air conditioning-heater case
assembly on the 1972 Ford Galaxie 500, which occupies considerable
volume in the engine compartment.
In packaging the system, the following modifications to the vehicle
were required to eliminate local interference points.
• Transmission Tunnel
A standard three-speed automatic transmission with torque
converter has been used in the packaging. The diameter of
the torque converter plus the necessity to locate the expander-
feedpump-transmission subassembly as far to the rear of the
engine compartment as possible without violation of the fire-
wall necessitated some increase in the transmission tunnel
size hear the. firewall. The transmission tunnel is thus
enlarged locally (adjacent to torque converter housing) by
1-1/2 inches vertically and 1-1/2 inches on the passenger
side. The position of the accelerator pedal and its operation
are not affected.
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THERMO ELECTRON
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• Steering Linkage
The drag link of the steering linkage must be depressed
locally by 3/4 inch to clear the expander oil pan.
• Number Two Cross Member
The rear upper edge of the number two cross member must
be depressed locally (at the center) by 3/4 inch to clear the
expander oil pan.
• Sway Bar
The center of the sway bar must be depressed locally by
2 inches to clear the combustion blower motor.
9 Rear Expander Mount
The rear mount for the expander must be moved 10 inches
rearward.
• Head Lamp Sockets
The headlamps must be moved 1-1/4 inches forward to clear
the condenser.
The following modifications are required to improve the flow of air to
the condenser and through the engine compartment:
• Redesign of grill.
• Replacement of four headlamps with two headlamps.
• Louver front surface of wheel aprons to increase flow area
for exhaust of cooling air.
• Removal of horizontal panels at front of apron assembly.
4-22
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N
The complete system layout is illustrated in Figure 4.4, a side
view looking from the driver's side of the engine compartment, and
Figure 4.5, a view from the top of the engine compartment. Overlays
are included with these figures for aid in identifying the various com-
partments of the system. Sectional views are provided in Figures 4.6,
4.7 and 4.8, as identified in Figure 4.4. Section A-A, given in
Figure 4.6, is a front view just in front of the expander looking .
toward the rear of the vehicle; this view includes the V-4 expander,
feedpump, starter motor, regenerator, and battery. Section B-B,
given in Figure 4.7, is a front view just in front of the combustion
system-boiler, looking toward the rear. Section C-C, given in
Figure 4.8, is a rear view of the condenser-condenser fan arrange-
ment.
As can be seen from these drawings, the expander-feedpump-
transmission subassembly is. located to the rear of the engine com-
partment. The cranking motor is integrated with this assembly. The
regenerator is located directly above the expander and is mounted
un the expander. The condenser is located in the very L-rmt of the
engine compartment, with the condenser fans mounted to the condenser
shroud on the rear of the condenser. The combustion system-boiler
is located between the expander and condenser and is placed as close
as possible to the expander to leave sufficient flow area for exhaust
of the condenser cooling air. The combustion blower and burner
controls are located between the two burners.
The accessory drive is taken from the rear housing of the ex-
pander so that only one dynamic shaft seal is required in the system.
A aplined shaft with universals is used to bring the accessory drive
4-23
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9OILER
STD. H ATER / A/P CONDI 'IOHER
CASE
S—SrOR LINKAGE
/f AND TR,\NSDUCER
CHITON SYSTEM
AIR CONDITIONER
RECEIVER
AUTOMATIC TRANSMISSION
BURUERS
Figure 4. 4 Side View, Packaged System.
-------
a
,„
STD. HEATER
AIR CONDITIONER CASE
— POWER STtERING PUMP
CONDENSER
IGNITION SYSTEM
CONDENSER FANS
AIR CONDITIONER
COHDENSER
CM
o-
00
Figure 4. 5 Top View, Packaged System.
-------
S7D. HEATER / AIR CONDITIONER
CASE:
EXPANDER INTAKE
VALVE OPERATOR
AUXILIARY AND
ACCESSORY DRIVE
V-4 EXPANDER
STARTER MOTOR
I
fO
Figure 4. 6 Section A-A from Figure 4. 4
-------
POWER STEEPING
PUMP
\ fUEL -
\ SOLENOID VALVE
_! 1
/ I
/
\ \
\\
\ J\ p
\ -i\
J
T
' VARIABLE
COHOENSB
— RECYCLE RE'.
BURHEF!
SPEED RATIO
R FAN DRIVE
\
\- AIP CONCH
COMPPES
'ERVOIR
TIONEP
SO*
4
(.//V£ K> BURNER
Ul
tCTiOM
Figure 4. 7 Section B-B from Figure 4. 4
-------
I
N
00
CONDENSER FAN
IGNITION SYSTEM
,— VARIAB.E SPEED RATIO
CONDEHSER PAN DRIVE
HYDRAULIC VALVE
CONTROL
SYSTEM CONTROL
e
01
Figure 4. 8 Section C-C from Figure 4. 4
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THERMO ELECTRON
CORPORATION
to the front of the expander. A belt drive is used at this position to
a second shaft to bring the accessory drive to the rear of the con-
denser. Belt drives from this point drive the condenser fans (through
the condenser fan variable speed belt drive), boost pump, air con-
ditioning compressor, power steering pump, and alternator. These
components are all located just to the rear of the condenser and
positioned so as not to restrict condenser cooling air flow significantly-
through the engine compartment.
The battery is located at the top rear of the engine compartment,
above the expander and to the rear of the regenerator. The receiver
for the Rankine-cycle system is located beside the boiler tube bundle,
at the top of the engine compartment on the passenger side. The in^
ducer and boost pump are located near the bottom of the engine com-
partment, as required by functional considerations.
Components located in front of the condenser are the ignition
system, air conditioning condenser, and air conditioning receiver.
The system control assembly is located at the front of the engine
compartment to provide a relatively cool environment for the electrical
components. The system control assembly includes inlet valve controls,
relays, startup sequencing, and safety cut-offs.. An electrical trans-
ducer in the passenger compartment and attached to the accelerator
pedal linkage provides the operator signal to the control assembly.
Because of relative motion between components mounted on the
vehicle frame and those mounted on the expander, flexible connections
are provided in the lines connecting the boiler outlet and expander,
boost pump discharge and feedpump suction, and regenerator outlet
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THERMO BLBCTHOM
CORPORATION
and condenser inlet. For clarity, the piping has not been shown on
the drawings presented here. Combustion gas exhaust ducts (not
shown on these drawings) are used to bring the exhaust gases to the
bottom of the engine compartment, and to the rear of the vehicle. These
ducts are taken from the top rear of the boiler and pass on either side of
the expander near the rear of the engine compartment.
4,4 ACCELERATION PERFORMANCE AND FUEL CONSUMPTION
CALCULATIONS
Calculations of acceleration performance and fuel consumption
over typical driving cycles have been made using steady-state com-
puter models of both the Rankine-cycle power system and the vehicle.
The calculational procedure is outlined in Figure 4. 9. Computer
models of the Rankine-cycle system components plus the Fluorinol-85
thermodynamic and physical properties are used in the system per-
formance prediction program. System performance characteristics
are generated by this program in the form of tables providing hp,
burning rate, and system efficiency (as well as any other desired
system characteristic) over the range of expander speeds and intake
ratios encountered in system operation.
In Figures 4. 10 and 4. 11, performance maps, cross plotted from
these tables, are presented for the 131.1 hp system in the form of
hp vs expander rpm, with lines of constant efficiency shown. The
maps are based on use of the Dana two-speed transmission, with
the first map applying to first gear with a high expander rpm relative
to vehicle speed and the second map to second gear with a low expander
rpm relative to vehicle speed. Characteristics of the Dana transmission
are discussed in Appendix VII.
4-30
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to
COMPUTER MODEL
OF RCS SYSTEM
COMPONENTS
IF-65 THERWODYNAMIC
AMD PHYSICAL
PfJOPERTIES
AUXILIARY AND REMAINING
ACCESSORY COMPONENT
MODELS
VEHICLE INfirTHA
AND ROAD TEST
MODELS
SYSTEM PERFORMANCE
PREDICTION PROGRAM
VEHICLE PERFORMANCE
AMD FUEL
CONSUMPTION PROGRAM
VEHICLE
AMD r'UEL
CONSUMPTION CHARACTERISTICS
I
ro
DRIVING CYCLE
CHARACTERISTICS
Figure 4. 9 Vehicle Perforrra nee and Fuel Consumption Calculation.
-------
UJ
N
140
120
100
leo
UJ
1800
20OO
Figure 4. 10 Performance Map with Transmission in First Gear
(High Expander Rpm Relative to Vehicle Speed) .
-------
140
120
100
or
UJ
80
UJ
en
§
»-60
u.
40
20
SECOND GEAR WITH DANA TRANSMISSION
IR MAX = 0.325
FLUORINOL-85 WORKING FLUID
10
FULL THROTTLE.^^ |2
"i — —" ""1^-""**
%
200 400 600 800 1000 1200
EXPANDER SPEED, RPM
1400
1600
13%
14 %
ro
-j
1800 2000
Figure 4. 11 Performance Map with Transmission in Second Gear
(Low Expander Rpm Relative to Vehicle Speed).
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THBKEMO BJ.BCTROM
CORPORATION
These performance maps are used as input to the vehicle per-
formance and fuel consumption program. This program includes all
inertia effects, road load, mechanical efficiencies of rear end and
transmission, vehicle accessories such as power steering and air
conditioning, and any accessory loads not included in the system
performance predictions. Any desired driving requirement or con-
dition can be provided as input to the program.
The estimated weight of the 131. 1 hp system is given in Table 4. 4.
In Table 4. 5, the vehicle weight breakdown is provided for the 1972
Ford Galaxie with a Rankine-cycle powerplant. The vehicle curb weight
with full fuel tank is 4276 Ibs; for comparison, the weight of the same
car with 351 CID I/C engine and three-speed automatic transmission is
4066 Ibs. For wide-open-throttle acceleration and fuel consumption
calculations, 300 Ibs weight was added to the curb weight to provide
a "test" weight of 4576 Ibs. For gradability, 1000 Ibs weight was added
to the curb weight, to provide a gradability "test" weight of 5276 Ibs.
The wide-open-throttle acceleration performance of the vehicle is
provided in Table 4. 6 and compared with the EPA specifications. The
system meets the specifications for 0-60 mph acceleration, distance
traveled in 10 seconds from standing start, and 25-70 mph acceleration.
The system does not quite meet the passing specification but is very
close.
In Table 4. 7, the gradability of the vehicle is presented. The vehicle
is able to pull off a 35% grade, and is able to maintain 15 mph on a 30. 7%
grade. The car can maintain 70 mph on an 8. 9% grade. The vehicle
gradability meets all EPA specifications. The vehicle top speed is approxi-
mately 103 mph on a level road.
4-34
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TNKRMO BLBCTRON
TABLE 4.4
ESTIMATED SYSTEM WEIGHT BREAKDOWN
Expander 345
Burner-Boiler 330
Combustion Blower
Atom. Air Compressor
Fv.el Pump
C^n-iburtion Blower Motor 15
Feedpump 42
Transmission 138
Step-up Gear 19
Regenerator 21
Condenser 72
Condenser Shroud 3
Condenser Fans, Pulley,
Brackets 22
Variable Speed Fan Drive 21
Boost Pump \
".nducer /
J Reservoir 2
i
I Starter 12
i
Alternator 19
Igniter 2
Buffer Fluid Reservoir Z
Exhaust Ducts 12
Electrical Box 4
Plumbing 16
Fluid and Lubricant 30
Battery 45
TOTAL 1213
4-35
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VHBRESIQ
CORPORATIO
TABLE 4, 5
VEHICLE WEIGHT BREAKDOWN
FOR PERFORMANCE CALCULATIONS
BASED ON 1972 FORD GALAXIE 500
Vehicle Less Propulsion Powerplant
Complete Rankine- cycle Powerplant
with Three- Speed Automatic Transmission
Vehicle Curb Weight (Fully Fueled)
"Passenger Weight for Performance and Fuel Consumption
Test Weight for Performance and Fuel Contraption
Passenger Weight for Grade Velocity
Test Weight for Grade Velocity
3063 Ibs
1213 Ibs
4276 Ibs
300 Ibs
4576 Ibs
-
1000 Ibs
5276 Ibs
4-36
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THBRMO ELECTRON
TABLE 4. 6
VEHICLE ACCELERATION PERFORMANCE
Vehicle Test Weight 4576 Ibs
Ambient Temperature 85 °F
Dana Transmission, Two Speed
0-60 mph
0-10 seconds
25-70 mph
Passing, 50-80 mph.
System Performance
13.36 sec,
457.9 ft.
15.0 sec.
15.4 sec.
1472 ft.
EPA Spec
* 13.5 sec.
a 440 ft.
£ 15. 0 sec.
3 15, 0 sec.
& 1400 ft.
4-37
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THBRMO ELECTION
COIPODATIOII
TABLE 4. 7
GRADABILITY
Vehicle Test Weight 5276 Ibs
Ambient Temperature 85 °F
Dana Transmission Two Speed
Vehicle Speed, mph
0
10
15
20
30
40
50
60
70
103
Grade %
System Performance
35.3%
34!6%
30, 7%
26. 8%
19.8%
13.8%
11.3%
8. 97%
6.84%
0%
EPA Spec.
Start from Rest
on 30% grade
30%
—
—
—
—
—
5%
85 mph
4-38
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THERMO ELECTRON!
CORPORATION
The vehicle fuel economy is presented in Table 4. 8 for steady speed
and for three drive cycles. One driving cycle is that specified in the
Federal test procedure for emissions measurement. The other two are
driving cycles used by the Ford Motor Company for evaluation of their
automobiles, one cycle being typical of suburban driving conditions and
the other typical of city driving conditions. The Ford Motor Company
customer average is the arithmetic average of these two driving cycles
and gives 10.0 mpg for the 1972 Ford Galaxie with Rankine-cycle
powerplant.
4. 5 EMISSION PROJECTIONS FROM RANKINE-CYCLE SYSTEM
The primary incentive for development of a Rankine-cycle auto-
motive propulsion system is its potential for very low emission levels
which not only meet the 1976 Federal objectives, but also are signifi-
cantly less than the Federal objectives. To demonstrate this potential,
Thermo Electron Corporation has made emission measurements on a
burner designed for a 100 hp automotive Rankine-cycle system for an
intermediate-size American ca'r such as the Ford Torino; the burner
was operated transiently over burning rates corresponding to operation
of the vehicle over the Federal emission test driving cycle. The fuel/air
control used in these transient tests was similar to that to be used in
the automotive system. The burner fired into a water-cooled "boiler"
with approximately the same configuration as in the system. The
procedure for measuring the exhaust emissions from the boiler was
identical to that of the Federal Register and included use of the three-
bag constant volume sampler. The details of the measurement are
described in Appendix VI.
Theresults of this test are presented in Table 4.9. The measured
emission levels in grams /mile are below the 1976 Federal standard by
4-39
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THBWMO BIBCTROM
COnrOftATlOX
TABLE 4. 8
FUEL CONSUMPTION
Vehicle Test Weight 4576 Ibs
Ambient Temperature 85 "F
Dana Transmission Two Speed
Constant Speed, 0% Grade
MPH MPG
30 x5.38
40 15. 15
50 13.52
60 11.76
70 10.33
80 8.54
85 7. 84
Constant Speed, 5% Grade
60 5.68
70. 4, 99
Driving Cycles
Federal Driving Cycle for Emissions 10, 81 mpg
FOMOCO Suburban Cycle 11.41 mpg
FOMOCO City Cycle 8.61 mpg
FOMOCO Customer Average 10. 01 mpg
4-40
-------
TABLE 4.9
EMISSION LEVELS MEASURED OVER
FEDERAL DRIVING CYCLE
Emissions
(grams/mile)
NO
X
CO
UHC
Transient
.Test .
Result*
0. 29
0.22.
0. 14
Federal 1976
Standard
0.4
1
3.4
0.41
Actual gas mileage used for tests was 12. 1 mph. The
latest performance calculation predicts 10. 8 mpg for the
Federal emission test driving cycle and the measured
emission levels have been increased by 12% to reflect
the change in fuel economy.
4-41
-------
>*»•
i
.&•
ro
DATA TRACK
BLOWER
TO
EMISSION
ANALYZERS
INPUT
CONTROL
SIGNAL
AIR-FUEL
CONTROL
VALVE
SAMPLING BAG
CVS UNIT
CONDENSER
BURNER
DRAIN
"^
f
\
^
f
- — •
f
J
V71
ro
DILUTION
AIR
Figure 4. 12 Burner Configuration Used in Emission Tests.
-------
THERMO ELECTRON
CORPORATION
a factor of 1.4 for NOX, 15.4 for CO, and 2.9 for UHC. These tests
conclusively demonstrate the low emission levels attainable with a
Rankine cycle power plant for automobiles.
Steady-state measurements indicate that use of exhaust gas re-
circulation (EGR) would result in even lower NO,, emission rates.
JL
Transient tests with EGR have not yet been made, but such tests
are planned in the future.
4-43
-------
CHAPTER 4
REFERENCES
1. "Vehicle Design Goals - Six Passenger Automobile, " Revision C,
Division of Advanced Automotive Power Systems Development,
Environmental Protection. Agency, issued May 28, 1971.
2. Manufacturing Chemists Association. (1970) Guide to Precautionary
Labeling of Hazardous Chemicals, 7th Ed. , Manual L-l, Washington,
D. C,
3. Blake, D. A., and Brown, D. R. , Evaluation of Trifluoroethanol
Toxicity and Hazard, April 1971, University of Maryland, Baltimore,
Maryland.
4. Morgan, D. T., and Raymond, R. J. , "Conceptual Design, Rankine-
Cycle Power System with Organic Working Fluid and Reciprocating
Engine for Passenger Vehicles, Report No, TE4121-133-70, June
1970, Thermo Electron Corporation, Waltham, Massachusetts.
5. "Exhaust Emission Standards and Test Procedures," Federal
Register, Vol. 36, No. 128, Friday, July 2, 1971, Part II.
Environmental Protection Agency.
4-44
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THERMO ELECTRON
CORPORATION
5. COMPONENT DESCRIPTIONS
In this section, a detailed description of the component designs
and characteristics is presented. These components are identical
to those used in the system packaging described in Chapter 4.
5. 1 EXPANDER-FEEDPUMP-TRANSMISSION SUBASSEMBLY
5.1.1 Expander Design
5.1.1.1 Variable Cut-Off Intake Valving Systems
System performance analyses have indicated that load control
by variable inlet valve timing, as opposed to throttling, is a very
desirable feature for a Rankine-cycle automotive powerplant because
of the higher part-load efficiency and higher wide-open-throttle
performance which can be achieved. During the initial conceptual
design study, both mechanical and hydraulic schemes were derived
conceptually to accomplish variable timing. During the second phase
of the program reported here, the most promising valve concepts
were analyzed in more detail and one approach was selected for
experimental bench testing. Work carried out under this program
was concentrated on (1) analysis of a mechanical or cam-driven
approach with two inlet valves in series, and (2) analysis and bench-
testing of a hydraulic approach by the American Bosch Company of
Springfield, Massachusetts. In addition, another hydraulic approach
is under experimental investigation at the British Internal Combustion
Engine Research Institute in England, and is financed entirely by
Thermo Electron Corporation.
For the mechanical and hydraulic schemes to be comparable, the
valve sizes and motions would have to be such that they both would
5-1
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THKRMO BLHCTWON
CORPORATION
give the same power and efficiency when installed in a given expander
operating at the design point conditions. A computer program was
developed to calculate an expander cycle, given operating characteristics
such as bore, stroke, speed, inlet pressure and temperature, etc.
Instantaneous inlet valve area as a function of expander crank angle is also
an input variable. The output of this program includes a pressure-
volume diagram, flow rate, power output and expander cycle efficiency.
Since the mechanical and hydraulic valves have different actuating
mechanisms and, therefore, different opening and closing rates, the
size and lift of the two valving approaches must be different to give
the same expander performance.
The maximum flow area of the hydraulically actuated valve had
initially been estimated at 1. 20 in , and this size was used to calculate
the valve mass and response characteristics in the studies on this
system performed by American Bosch under subcontract from Thermo
Electron. Once the size of the valve was chosen, the response for
various servo pressures was experimentally determined and the power
to drive the valve gear calculated. This information was then used to
conduct an analytical study of expander performance as a function of
inlet valve size, servo pressure, bore, stroke, and speed. The
results of this study determined the characteristics of the expander
for 147 gross shaft horsepower output as given in Table 5.1.
These characteristics represent a reasonable optimum in the
trade-off between expander size (as well as the size of other com-
ponents of the system) and expander efficiency at the design point
condition. Increasing valve size and/or servo pressure improves the
expander indicator diagram performance but the improvement is
5-2
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THERMO ELECTRON
CORPORATION
nullified by the additional power required to drive the valves. Attempts
to use a higher design point rotational speed in order to reduce the
expander size and the overall system size were not successful with
the current intake valving systems, since intake valve throttling losses
increase above 1800 RPM. A higher intake ratio is then required to
maintain a given power level and the expander efficiency is reduced
requiring a larger boiler and condenser as well as other components
of the system. TABLE 5. 1
EXPANDER CHARACTERISTICS
WITH
HYDRAULICALLY ACTUATED INTAKE VALVES
Number of cylinders 4
Bore 4. 42 inches
Stroke 3. 00 inches
Total Displacement 184 in3
Maximum Average Piston Speed 900 ft/min
Corresponding Maximum Speed 1800 rpm
Inlet Valve Size 1.25 inch diameter
x 0. 3 inch lift
Hydraulic Pressure 1500 psia
Intake Ratio 0. 175
The mechanical, two valves-in-series design had to provide the
same expander performance as the hydraulic design in order for the
comparison to be valid. The principle of the two valves-in-series
approach is shown in Figure 5. 1. The valve operated by cam No. 1
operates with fixed timing whereas the 'timing of the valve operated
by cam No. 2 is varied. The overlap of the two valve events determines
5-3
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THERMO ELECTRON
CORPORATION
the effective intake ratio of the expander. It is obvious from the
valve event diagram shown that each valve of the two valves in series
must have an appreciably larger flow area than the single hydraulic
valve, since the full lift of the cam is not utilized at the short intake
ratio (0. 175) at the design point. The final configuration taken by
the two valves in series is discussed in more detail in part b of this
section, and the detailed design illustrated in Figures 5. 15, 5. 16
and 5. 17. The size shown in Figures 5. 15 gives the same performance
as the hydraulically-actuated valve. Figure 5.2 shows valve area
as a function of crank angle for the two systems at the design point
conditions. Figure 5.3 shows the instantaneous flow rate through
the two valves, and Figure 5. 4 is an expander P-V diagram for the
two valving schemes at the design point condition. It can be seen
from these results that the two schemes give very nearly equal per-
formance at the design point. No off-design runs were made conn-
paring the two systems, but one would expect that the hydraulically-
actuated valve would show an advantage, since its lift vs. crank angle
diagram approaches a square wave as expander speed is decreased;
the two valves-in-series profile remains the same for a given valve
timing as speed decreases.
a. American Bosch Hydraulically Actuated Inlet Valve; The
valve actuating mechanism is illustrated in Figure 5. 5 and consists
of the following basic elements: A double-spool servo valve, a spool
check valve, the driving plunger and piston, and a solenoid or mag-
netic actuator (not shown on Figure 5.5).
Referring to Figure 5.5, the right-hand sketch shows the valve
in the closed position. High pressure fluid is applied to the top of
5-4
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1-2655
(5o)
VAPOR
INLET
CRANK ANGLE
Figure 5. 1 Two Inlet Valves in Series,
Variable Cut-Off Mechanism.
5-5
-------
1.2
CM .
c
LJ
5
0.8
^
006
_l
Ll_
LU
0.2
T. D.C.
I I
N = 1800 RPM
BOSCH VALVE, l.25in. DIA. x 0.3 LIFT
SERVO PRESSURE- 1500 PSI
TWO VALVES IN SERIES
2.186in. DIA. x0.5 LIFT
10
20°
CRANK ANGLE
40«
ro
H—
~J
50«
Figure 5. 2 Valve Flow Area aa a Function of Crank Angle at Design Point Conditions.
-------
Ul
I
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N=I800RPM
HYDRAULIC VALVE
1.25 DI A.
0.3 LIFT
TWO VALVES IN SERIES
2.186 OIA.
0.3 LIFT
ro
20°
CRANK ANGLE
Figure 5. 3 Flow Rate Through Intake Valve versus Crank Angle at Design Point Conditions.
-------
1-2707
700
600
500
400
UJ
tr
tr.
a.
300
I
BORE - 4. 42 INCHES
STROKE - 3. 0 INCHES
SPEED - 1800 RPM
INTAKE VALVE OPENS - 5° BTDC
INTAKE VALVE CLOSES - 45° ATDC —
EXHAUST PORT OPENS - 130" ATDC
EXHAUST PORT CLOSES - 34° BTDC
INLET PRESSURE - 692 PSIA
EXHAUST PRESSURE - 44 PSIA
INLET TEMPERATURE - 550°F —
FULL OPEN FLOW COEFFICIENT ASSUMED
0. 6 FOR BOTH VALVES
HYDRAULICALLY ACTUATED VALVE
ACTUATING PRESSURE - 1500PSI
VALVE DIAMETER - 1. 25 INCHES
MAXIMUM VALVE LIFT - 0. 3 INCH
IMEP - 189 PSI
INDICATED EFFICIENCY - . 805
• TWO VALVES IN SERIES —
MEAN VALVE DIAMETER - 2. 186 INCHES
MAXIMUM VALVE LIFT - 0. 5 INCH
IMEP - 191 PSI
INDICATED EFFICIENCY - . 804
200
100
I 2
CYLINDER VOLUME (ft'x|02)
Figure 5. 4 Expander P-V Diagram at Design Point Conditions.
5-8
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171
I
vO
OUTWARD MOTION
STOPPED BY
HYDRAULIC
SNUBBING
ACTION
SERVO SPILL
ANNULUS
HIGH PRESSURE
INLET
DRIVE PLUNGER
SERVO VALVE
SPOOLS
SOLENOID ON
VALVE ACTUATING
PISTON
TANK DRAIN
(CONDENSER PRESSURE)
SOLENOID OFF
Figure 5. 5 Servo Actuated Inlet Valve (American Bosch).
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THKRMO BLBCTRON
CORPORATION
the small spool of the servo valve. A bleed line equipped with an
orifice and a spool-type check valve is provided across the two spools
of the servo valve. This line produces pressure force equalization
across the spools when the solenoid valve is closed. The servo spools
are thus maintained in their closed position by the force differential
generated by the difference in spool diameters. The high pressure
fluid is ducted through the servo valve passages and applied to the
underside of the valve actuating piston. This fluid pressure holds
the engine valve in its closed position.
Refer now to the left-hand sketch. When the solenoid actuated
valve is opened, the fluid pressure under the large spool valve is
sharply reduced and a large pressure drop occurs across the two
spools which causes them to move downward and the spool check
valve to close by spring action. When the servo spools have moved
to the point where the spill annulus is closed off (position shown in
sketch), the pressure under the large spool increases sufficiently to
stop the motion of the spools. Since the bleed line is open, the
pressure under the spools will increase, and they will start to move
upward. Motion in this direction, however, will cause the spill
annulus to be reopened slightly, which re-establishes flow through
the solenoid control valve. When this flow exactly matches the flow
through the bleed orifice, the spools become stabilized in their "open"
position.
With the servo valve spools in their "open" position, the area under
the valve actuating pistion is switched from high pressure to "tank"
pressure (=» condenser pressure) and the area above the drive plunger
is switched from "tank" pressure to high pressure. The resultant high
5-10
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THERMO ELECTRON
CORPORATION
pressure drop across the drive plunger and valve actuating piston
drives the intake valve open. When the valve actuating piston moves
to the point where its spill annulus is closed off (position shown in
sketch), the pressure under the piston increases sufficiently to stop
the motion of the valve. This "snubbing" pressure is indicated by the
dot-dash area. The engine valve will remain open as long as the
solenoid valve is energized.
When the solenoid valve is closed, flow through the bleed line
causes the pressure under the large servo spool to increase. This
forces the spool check valve upward, thus opening a large parallel
feed passage under the servo spools for fast response. The servo
spools are driven upward to the point where the top end of the small
spool closes off the bleed line annulus (position shown in right-hand
sketch of Figure 5.5). Further slight upward motion causes the
pressure under the servo spools to decrease to a value which will
achieve force balance, at which point all motion will cease and the
servo spools will be in their "closed" position. If in this condition
there is leakage into the bleed line, the servo spools will move slowly
upward until they contact a mechanical stop (not shown). During
operation the spools will never contact the stop, due to insufficient
time for this leakage to occur. When the servo valve is in this
closed position, the area above the drive plunger will be switched
from high pressure to "tank" pressure, and the area below the valve
actuating piston will be switched from "tank" pressure to high pres-
sure. The resultant pressure drop across these members will now
drive and hold the engine valve closed.
5-11
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THURIDiO BIBCTROM
CORPORATION
Control of the valve actuating mechanism described herein is
accomplished by a speed-sensitive phototransistor timer which ener-
gizes the capacitor discharge circuit that feeds the solenoid at the
proper time in the expander cycle. A circuit is also required to
turn this signal off at the proper time to control the length of time
that the valve is open. The timer is equipped with a speed-sensitive
automatic advance mechanism which maintains the valve opening at
the optimum point relative to crank angle as a function of speed. Also
built into the valve duration control is an electronic speed sensor
which is used to limit the maximum intake ratio as a function of speed,
so that the expander cannot overdraw the capacity of the boiler. The
control schematic is shown in Figure 5. 6 along with the other com-
ponents required for operating the valves. Note that the hydraulic
reservoir for the valving is separate from the lube oil sump in the
crankcase. The lube oil pump provides only enough oil for make-up
of leakage out of the valve system. Leakage occurs only down the
valve stem and around the vane pump and is quite small. A separate
reservoir is used to insure that the oil used for the valving system is
free of Fluorinol-85, which may be present in the crankcase, par-
ticularly during start-up.
The vane pump is of the variable displacement type with a speed
sensing system to control its output pressure as a function of speed,
as shown in Figure 5. 6. Use of a variable displacement pump minimizes
the power required to operate the intake valves. Fast valve response
is required only at higher expander speeds and the servo pressure can
therefore be allowed to drop at lower expander speeds from 1500 to
800 psi without sacrificing performance but saving almost half the power.
5-12
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DRAIN
r-
1
r~ "
VAL.VC. C,VtNI
DURATION
CONTROL
1 1 '
! ! c\^
.....
1
1
SEPARATOR
a
RECEIVER
SOLENOID
VALVE
VALVE
ACTUATO
VALVE
ACTUATOR
VALVE
ACTUATOR
VALVE
ACTUATOR
ACCUMULATOR
SOLENOIOC7
VALVE t
PRESSURE
CONTROL
(VARIABLE
DISPLACEMENT)
EXPANDE
LUBE
PUMP
EXPANDER
CRANKCASE
ACCELERATOR
PEDAL
VSTART
~~" SWITCH
U1
ELECTRICAL
HYDRAULIC
Figure 5. 6 Schematic of Control and Hydraulic System for American Bosch Intake Valves.
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T H • R MO ELECTRON
CORPORATION
An accumulator is required to provide the high instantaneous flow
rates during the stroking of the inlet valve.
The system is equipped with a solenoid valve which holds pressure
in the accumulator during shutdown to provide fast system cold startup.
During startup, this pressure closes any expander intake valve that
might have opened due to leakage during the period the expander is shut
down. The closed valves permit faster buildup of boiler pressure and
temperature during startup cranking, since no vapor is passed through
the expander until the boiler pressure reaches a predetermined level.
Operation of a 5 hp system at Thermo Electron has demonstrated that
the procedure results in a significant reduction in the cold startup time.
During the initial evaluation of the hydraulically-actuated valve, the
American Bosch Company performed a design analysis of the actuator
illustrated in Figure 5. 5. The object of this study was to size the
various actuator and servo spools, determine the servo pressure
required and calculate the response time of the valve. Table 5. 1
summarizes the results of this study. During the second phase of
the subcontract with American Bosch, a complete actuator was built
and its response was measured in bench-testing. A photograph of the
actuator tested, is presented in Figure 5. 7. The valve head was not
built with a seat and balancing piston, but was fitted with a weight
to simulate the mass of the real valve (0. 3 Ibs). The actuator was
tested in air; no attempt was made to simulate the pressure loads or
thermal situations which might occur in a real expander. The primary
object of this work was to see if the actuator could be made to operate
the valve with the proper time response.
The opening and closing time response of the experimental unit was
measured as a function of servo pressure with the results presented
in Table 5. 3 and Figure 5. 8.
5-14
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THERMO ELECTRON
CORPORATION
TABLE 5.2
DESIGN CHARACTERISTICS OF
AMERICAN BOSCH HYDRAULIC ACTUATOR VALVE
Supply Pressure
Supply Flow
Valve Drive Piston Diameter
Valve Return Piston Diameter
Valve Stroke
Valve Weight
Stroke Time
Required Force
Flow Rate During Stroke
Servo Valve Stroke
Upper Servo (Small) Diameter
Lower Servo (Large) Diameter
Servo Stroke Time
Force to Drive Servo
Solenoid Valve Diameter
Lower Servo Feed Diameter
Total Valve Event
- Up to 2000 psi
- 1. 75 gpm/valve assembly at
2000 rpm
- 0. 650 inch
- 0.695 inch
-0.3 inch
- 0.3 Ib
- 0.001 second
- 466 Ib
- 5. 78 x 10"2 ft3/sec (26 gpm)
- 0. 200 inch
- 0. 250 inch
- 0. 281 inch
- 0.001 second
- 30 Ibs
- 0.071 inch
- 0. 086 inch minimum
- 0.003 second
5-15
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01
IS)
-J
Figure 5. 7 American Bosch Hydraulic Valve Actuator
Used in Bench Testing.
-------
Ul
I
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«
10
UJ
O
UJ
Q.
O
vO
Ul
I
I
I
500 1000 1500
SERVO PRESSURE (PSI)
2000
Figure.5. 8 Inlet Valve Opening Time versus Servo Pressure
for Servo-Actuated Inlet Valve.
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THBRMO ELECTRON
CORPORATION
TABLE 5.3
MEASURED OPENING AND CLOSING TIMES
FOR AMERICAN BOSCH HYDRAULIC VALVE ACTUATOR
Servo Supply Pressure
(Psi)
600
800
1000
1200
1500
2000
Time to Open
(Millisecs)
4.0
3.0
2.3
2.2
2.0
1.7
Time to Close
(Millisecs)
4.0
4. 0
3.5
3.5
3.0
2.5
These data illustrate that the valve response time is levelling off
with increasing supply pressure above about 1500 psi; the increased
pressure is being used primarily to accelerate the fluid in and out of
the mechanism rather than to move the valve.
Oscilloscope traces of the valve response were made with the
different events occurring described in Figure 5.9. The effect of
supply pressure on the valve motion is illustrated in Figures 5. 10
and 5. 11. These traces were used to determine the valve opening
and closing times.
The part stroke performance of the valve was also measured
with the results presented in Table 5.4. The corresponding valve
motion traces are shown in Figure 5.12. The cycle time is 1. 5
milliseconds greater than the sum of the opening and closing times
because of a dwell at the top of each event, as illustrated in Figure
5.12. This dwell results from the deceleration-acceleration times
5-18
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1-2230
r
End of valve opening at snubbing point.
. 050 x . 687 dia. chamber on lower piston
softened snubbing action and reduced
maximum valve stem load from 6000 Ibs
to approximately 1200 Ibs.
Valve continues to sink slowly
into snubbing area.
Start of valve closing.
ms
Full valve
stroke.
Current to
solenoid.
Snubbing action as
valve approaches
closed position.
Start of
event signal.
"0" valve
stroke.
Engine valve motion
measured with a
variable inductance
pickup.
Valve closed.
•End of event signal.
-Start of valve
opening.
Figure 5. 9 Valve Opening and Closing Motion for
Servo-Actuated Inlet Valve.
5-19
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1-2657
Solenoid current
Valve travel
U.—5 ms ,
4.1 j- 1
it > i
».l. J.f.
I 1 1 1
^-
-H-H-
g-.l~".r- r-
Full stroke
Supply Pressure
600 psi
— "0". Stroke
Supply Pressure
800 psi
-H-H-
2 ms
I.I-:
rrrr,
• C -
I ^jETL-h'-LVW'^.
\
•1-1
L
\ \
_:j j ____ | ^. ...... _
Supply Pressure
1000 psi
Figure ?. 10- Effect of Supply Pressure on Valve Motion.
5-20
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1-2658
•H-
H-H-
2ms
•H-H-
Tr-eg^-ye^
Supply Pressure
1200 psi
2 ma
Supply Pressure
1500 psi
-H-H-
•H-H-
/H
-v^
2 ms
•t,^^J
•K-
4-44*
R
X
XL
Supply Pressure
2000 psi
Figure 5. 11 Effect of Supply Pressure on Valve Motion.
5-21
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Full stroke
U1
ro
"0" stroke —
Full stroke •
"0" stroke-
2 ms (Typical)
Full stroke
Min. time for full
stroke cycle.
Fig.. 8
• 2 ms
++H
Fig. 9
"0" stroke
-
_J.
x.
^F- •
f
/
» I'J
^ :
v
fJJJ.WC
Full stroke
"0" stroke-
Full stroke .
110" stroke
Fig. 10
*| I* 2 ms
_J.
•s
^
Fig. 11
2 ma
**
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THBRMO BLBCTRON
CORPORATION
for the valve. The valve motion can be modulated down to zero lift if
desired. Some cycle-to-cycle variation occurs at less than full lift
and is shown in the last trace of Figure 5. 12. This variation is not
sufficient to affect the expander performance. Valve motion measure-
ments were made with and without the bypass piston in the servo feed
line and with orifices from . 027 inch to . 040 inch diameter. The best
combination was a . 040 inch orifice with the bypass piston included.
Figures 5. 13 and 5. 14 illustrate the results with the . 027 inch and
the . 040 inch bleed orifice, respectively. Note that the delay to
initiate closing after the end of the electrical signal is approximately
6 milliseconds with the . 027 inch bleed, and is reduced to approximately
4. 5 milliseconds with the . 040 inch bleed. The opening delay was in-
creased approximately 0. 5 millisecond by this change.
TABLE 5. 4
PART-STROKE PERFORMANCE
OF SERVO-ACTIVATED VALVE
Supply Pressure = 1500 psi
Fraction of Full Stroke
Full
.9
.6
.35
.08
Cycle Time (milliseconds)
6: 5
5. 5
4. 5
3.5
2. 0
The valve displacement versus time curves for the various con-
ditions shown here were curve fitted and the resulting expressions
were used in the expander cycle simulation program. The valve flow
area-crank angle curve is shown in Figure 5. 2 for the design point.
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THERMO ELECTRON
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This experimentally determined curve was used to arrive at the design
point condition previously discussed. A peak supply pressure of 1500
psi was found to be adequate for reasonable performance. The power
to operate the four (4) valves at the design point is estimated at 3. 6
hydraulic horsepower.
The integration of the hydraulic actuator with the expander cylinder
head is discussed in Section 5. 1. 1. 3. The timing device and vane pump
and controls have been designed and their integration with the expander
is also discussed in Section 5.1.1. 3.
b. Mechanically Operated Two Valves-in-Series: This concept
in its simplest form with poppet valves is shown in Figure 5. 1. The
valve event is determined by the amount of overlap between cams
No. 1 and No. 2. Cam No. 1 determines the beginning of the event
and has fixed timing with respect to the expander crankshaft; cam
No. 2 determines the end of the valve event and must have variable
timing with respect to the crankshaft in order to effect variable cut-off.
Since valve No. 2 determines cut-off, all of the volume from this
valve to the cylinder is clearance volume. The roles of the two valves
cannot be interchanged, because if valve No. 1 •were the "cut-off valve,"
it would open somewhere on the exhaust stroke of the expander, re-
leasing the volume of high pressure vapor between the two valves
directly to exhaust. This would constitute an unacceptable efficiency
penalty.
The unavoidable clearance volume resulting from use of poppet
valves as illustrated in Figure 5. 1 results in the undesirable charac-
teristics. If adequate flow areas are used for a reasonable pressure
drop, the clearance volume is sufficiently large to seriously degrade
5-24
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1-2660
—
rf( 1
BHE
f
2 ms.
1 —
KLJ.J
vW
S5*8E*
AALt.
'
f
1 L^
a^=
>Ai/ 1-^.VUiJ.
•Vr {r^-VvT"r
~J
J --ij-TGTy..
1
"\
^
K^S
i
1..
.jjjj
TTTl
,.
\i
0. 027 bleed
Figure 5. 13 Valve Motion with 0. 027 inch Bleed
Orifice and Bypass Piston.
5-25
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1-2661
/N.
•' I "c^-
Zjfl9SI!»!*-Lt]
2 ras .
:.C-Uk
/V^-v^-,,. !.._,._ >\
J1J
H-H-
JL
0. 040 Bleed
Figure 5. 14 Valv« Motion with 0. 0^0 inch Bleed
Orifice and Bypass Piston.
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THKRMO BLBCTRON
CORPORATION
the expander efficiency. In addition, the minimum expander ratio
is the volume between the two valves. For adequate flow area, this
minimum intake ratio is large relative to that required for idle and
light load conditions. A throttle valve would thus be required between
the boiler and expander to permit operation of the expander under
these light-load conditions.
In addition to these problems, the poppet valves are pressure
unbalanced. The pressure forces and resultant cam loading is ex-
cessive and renders the use of poppet valves as illustrated in Figure
5. 1 unfeasible.
An extensive evaluation of various approaches to the two-valve-in-
series concept was carried out in both the initial and current programs
with EPA in order to eliminate or alleviate these problems. This
evaluation resulted in the use of concentric and annular valves as
illustrated in Figures 5. 15 and 5. 16. The inner valve is operated
with fixed timing (cam No. 1) and the outer valve with variable timing
(cam No. 2), so that the function of the valves is identical to that
illustrated in Figure 5. 1. Use of the concentric valves eliminates the
volume trapped between the two valves so that the intake ratio can be
reduced to zero. The clearance volume is also reduced to a level
small enough so that no significant effect on expander efficiency occurs.
Pressure imbalance forces are reduced by use of ports in the transition
section from the valve stem to the valve sleeve thereby equalizing the
pressure across the valve. Pressure forces are thus reduced to those
due to the valve stem and due to any pressure differences on the valve
sleeve. A detailed analysis of the mechanical valve design presented
in Figure 5. 15 is given in Appendix I; this analysis demonstrates that
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THERMO ELECTRON
CORPORATION
all stress levels, including cam stresses, are acceptable. The maxi-
mum contact stress between the cam and follower is 198, 400 psi
necessitating a roller follower. Seal rings are used to reduce leakage
of vapor by the valve sleeves.
The differential gear arrangement for changing phase between
the cam shafts for the variable intake function is illustrated in
Figure 5. 17.
(i) Valve Opening Force. Due to the finite size of the valve seat,
a pressure force unbalance occurs before the valve lifts, and a rela-
tively high force is required to initiate valve movement. Figure 5. 18
illustrates the unbalanced pressure force situation when the valves
are both seated and unseated. (See Appendix I for pressure force evalu-
ation.) To reduce this force and the stresses and vibration it produces,
two steps were taken in the design of the valve gear. First, to reduce
the magnitude of the force required to initiate valve movement, the
thickness of the valve seat is made less than the . 0930 inch thickness
of the valves as illustrated in Figure 5. 19. With the valve ends shaped
as detailed, the unbalanced pressure force will be greatly reduced. As
a second measure, the cam curve is designed so that the ramp will
take up the lash in the system, and then initiate the movement of the
valve off its seat while there is a low constant velocity in the valve
train. This will reduce any jerk amplification of the valve lifting force
and initiate the motion of the valve smoothly into the main cam event.
(ii) Cam Curve and Return Springs. Since the event is controlled
by the opening of one valve and closing of the other, the sharper the
opening and closing curves, the better the overall flow coefficients will
be. The maximum positive acceleration is effectively limited by the
5-28
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1-2718
NUTS FOR CLEARANCE ADJUSTMENT
Figure 5. 15 Mechanically-Driven Variable Cut-Off Intake Valve.
5-29
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1-2735
VALVE STEM
KEEPER
'C-G'
Figure 5. 16 Mechanically-Driven Variable Cut-off Intake Valve.
5-30
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1-2720
r
DRIVEN GEAR
FROM CRANK
BEARINGS
CUT-OFF
CONTROL ARM
Figure 5. 17 Differentia.! Gear Arrangement for Mechanically-Driven
Variable Cut-off Intake Valve.
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1-2664
OUTER VALVE
348 Ibs.
BDC
118.5 Ibs.
349 Ibs
I
79.4 Ibs.
TDC
INNER VALVE
349 Ibs
BDC
Figure 5. 18 Pressure Forces on Sleeve Valves with
0. 0930 inch Valve Seat Thickness.
5-3Z
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Lol
VALVE SEATING
SURFACE
.040
Figure 5. 19 Valve Seat Profile to Reduce Opening
Pressure Force on Valve.
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THERMO ELECTRON
CORPORATION
requirement to keep the cam convex to minimize grinding cost. The
minimum acceleration that can be obtained is limited by the magnitude
of the spring force required to keep the cam and follower in contact.
The resulting cam curve given in Figure 5.20 is a compromise between
these factors. The cam velocity and cam acceleration in/degree
2
and in/degree , respectively, are also presented in Figure 5. ZO.
The cam characteristics are given in tabular form in Appendix I.
Concentric coil springs are used in the design to obtain the required
spring force. Consideration was also given to the use of torsion bar
return springs mounted in the rocker arm pivots, but they are more
difficult to package, although a higher spring force could be obtained
by their use.
(iii) Valve Stem Leakage. An analysis was carried out to deter-
mine the magnitude of the leakage of working fluid past the valve stems.
Laminar flow assumptions were used, and the effect of grooves in the
valve stem was examined. Two predictions were made. The first
assumed that the stem was concentric in the bore and the second
assumed that the stem lay on one side of the bore. The second
geometry gave a higher leakage rate.
The analysis showed that if the diametrical clearance is kept
below ~ . 0003 inches, the leakage rates are acceptable (~ 15 Ib/hr
for four valves). It also showed that with clearances of this magnitude,
the effects of grooves in the valve stem were negligible or even in-
creased the leakage rate. They would also act as stress risers. It
was therefore concluded that grooves should not be cut into the valve
stems.
5-34
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Ul
I
OJ
Ul
I0*3in/o
AX|0+4in/02
t
-1.0
-2.0
X - DISPLACEMENT (INCHES)
V - VELOCITY (INCHES/DEGREE)
A - ACCELERATION (INCHES/DEGREE2)
I
I
I
I
I
I
I
I
I
I
I
ro
o
o^
ts)
I
0 5 10 15 20 25 30 35 4O 45 50 55 60 65 70 75 80 85 90
CAM ANGLE
0 5 10 15 20 25 30 35 40 45 50 55 60 65 70 75
CRANK ANGLE
Figure 5.20 Cam Profile, Velocity, and Acceleration for Mechanicallv-Driven Intake Valve.
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THERMO ELECTRON
CORPORATION
(iv) Assembly. Several schemes for construction and assembly
of the valves were considered. These were: (a) thread or bolt the
bell-shaped cylindrical valves onto the valve stems; (b) weld the bell
shaped cylindrical valves to the stems after assembly; or (c) make
the valves and stems in one piece and use the split collet arrangement
shown in Figure 5. 15 to connect the valve stems to the rocker arms.
Scheme (c) was chosen for several reasons. Welding the stems and
valve together would mean that the assembly could not be dismantled.
Also, serious difficulties might be encountered with distortion of the
valves during heating and alignment of the two valves. The fatigue
durability of any threaded joint is questionable, especially since it
is located at a point of stress concentration between the stem and the
valve. Threading the two components together also results in a large
increase in the height and weight of the assembly.
Making the valve and stems in one piece eliminates alignment
problems. The taper onto which the lower disc is placed, and the
upper disc forced onto the taper of the split collets by the Belville
washers, add very little either to the weight or height of the assembly.
The valve gear is also easily dismantled.
b. Choice of Valving Approach for Preprototype System Testing;
The American Bosch system has been selected over the mechani-
cally driven valves as the prime intake valve approach for use in the
preprototype system testing. Primary emphasis in this selection was
based on the following:
(1) The American Bosch system should provide a higher expander
efficiency, particularly at part-load, medium-speed expander
conditions.
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THKRMO ELECTRON
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(2) The American Bosch system is more fully-developed and
should require less development time and effort than the
mechanically-driven valving approach. Use of the
American Bosch system should thus provide satisfactory
expander operation at the earliest possible date in
accordance with the EPA development schedule.
(3) The American Bosch system results in a smaller expander
size facilitating packaging of the complete engine in the
1972 Ford Galaxie.
No experimental testing has been carried out on the mechanical valving
system. This approach will undoubtedly have vibration problems due
to the rapid change in the pressure force on the valves (see Figure 5. 18).
There are also numerous leakage paths (see Figure 5. 15) as well as
assembly and machining problems with the mechanical approach.
During the next phase of the program, the Ford Motor Company will
be working with Thermo Electron Corporation in resolution of these
problems and testing of the mechanical valving system. Effort on
the BICERI hydraulic valving approach (funded solely by Thermo
Electron Corporation) will also be continued. Both of these alternate
approaches are simpler than the American Bosch system and will
have a lower manufacturing cost, but are in an earlier stage of
development.
5.. 1. .1. 2 Exhaust Valving
The exhaust valve used in the expander is similar in operation
to that used in the 5-1/2 hp TECO expanders. The method of operation
of the automatic exhaust valve is illustrated in Figure 5.21. The poppet
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THERMO ELECTRON
CORPORATION
type exhaust valve is connected to the cylinder by three ports; the main
exhaust port near bottom dead center which just starts to be uncovered
by the piston at 130 degrees ATDC and which is completely uncovered
at BDC; an auxiliary exhaust port which is completely covered by the
piston at 34 degrees BTDC; and a smaller port near TDC through which
cylinder pressure is applied to the top of the valve to insure closure
of the valve before the auxiliary exhaust port is uncovered by the piston
during the power stroke. During the power stroke (sketch 1 of Figure
5. 21) the cylinder pressure is applied to the top of the valve and keeps
the exhaust valve closed until the main exhaust port is uncovered and
the cylinder pressure vented to the pressure in the expander exhaust
line. When the main exhaust port is uncovered, the pressure force
across the valve is equalized and the spring force opens the exhaust
valve so that the auxiliary port is also vented to the expander exhaust
line (sketch 2 of Figure 5.21). During the return stroke (sketch 3 of
Figure 5.21), vapor in the cylinder is pushed through the auxiliary
exhaust port until this port is closed by the piston. When this port
is closed, recompression of the remaining vapor in the cylinder
occurs; this pressure is applied to the top of the exhaust valve through
the small port near TDC, and the pressure imbalance initiates closing
of the exhaust valve as illustrated in sketch 4 of Figure 5.21. This
type of exhaust valving gives an ideal P-V diagram as illustrated in
Figure 5.21 and maximizes the power per unit displacement of the
expander.
In scaling this valve to the auto motive-size expander, much larger
flow areas are required. As a result, the valve travel and the valve
diameter are considerably larger than in the 5-1/2 hp expander. This
5-38
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POWER STROKE
EXHAUST VALVE
OPENED
RETURN STROKE
EXHAUST VALVE
CLOSED
EXHAUST VALVE SEQUENCE
I
3
DISPLACEMENT
INDICATOR DIAGRAM
ro
00
I
O
Figure 5. 21 Schematic of Exhaust Valve Function.
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THBRMO ELECTRON
CORPORATION
greater travel results in a requirement for higher average velocity
and the larger valve diameter results in higher pressure forces on
the exhaust valve than for the 5-1/2 hp expander. As a result of these
differences, bench-testing of the full-size exhaust valve was carried
out as part of this phase of the EPA program to establish the operating
characteristics of the exhaust valve to be used in the preprototype
expander. Since it is difficult to exactly simulate in the bench test
unit operating conditions of the expander, additional development of
the exhaust valve may be required in testing of the preprototype ex-
pander to attain optimum expander performance.
The test fixture used for testing of the exhaust valve is illus-
trated in Figure 5. 22. The test fixture consists of a rotary valve
and drive arrangement to provide timed pressure pulses for operating
the exhaust valve. Compressed air is fed into one end of the rotary
valve which is rotated by a variable speed drive, and a square-wave
pressure pulse is then generated and applied to the top side of the
exhaust valve, causing it to alternately close and open. As in the
expander, the opening force is supplied by a spring. Originally, the
exhaust valve was fitted with a velocity transducer, as shown in
Figure 5. 22, but it proved impossible to keep the magnetic core
attached to the valve, due to vibrations caused by the exhaust valve
hitting its seat. This method was abandoned in favor of using a
stroboscope to plot valve displacement versus time.
The valve development during the testing involved primarily
(1) reduction of the seating velocity to an acceptable level while still
maintaining the required closing time for the valve, and (2) selection
of the appropriate spring to provide the proper valve opening time.
5-40
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Ul
I
NOTEV.
I A.LIOM VALVE (iT
%E«.T ( ITEM n)
SCRE.>WI (ITEM
1C*") IN
L
Figure 5. 22 Exhaust Valve Test Apparatus.
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THERMO BLBCTROM
CORPORATION
The required opening and closing times are established by operation
of the expander at its maximum speed of 1800 rpm. For closing of
the exhaust valve, closing action starts at 34" BTDC and the valve
must be closed at 34° ATDC when the auxiliary exhaust port is again
uncovered. Leakage by the exhaust valve guide and piston at the top
of the valve is small enough to have negligible effect on expander
performance.
For the specified travel of 0. 3 in. , the minimum average valve
velocity during closing is:
x (1800)(360) -- xr-x-= 3. 97 ft/sec.
,
68 degrees mm. 12 in. 60 sec
The maximum closing time is 6. 30 milliseconds. During opening of
the valve, opening action starts at 130° ATDC and the valve must be
completely open at 130" BTDC (the main exhaust port is open during
this period). The minimum average velocity during opening is thus
2. 70 ft/sec and the maximum opening time is 9.25 milliseconds.
The primary experimental effort was concentrated on develop-
ment of a damping method to reduce the seating velocity to a level
of 10 - 12 ft/sec. Also, since the pressure differential varies across
the valve during the recompression, and for different expander oper-.'
ating conditions, it was desirable that the method of damping provide
a relatively small variation in the valve velocity for different pressure
differentials. A large number of damping methods to accbmplish these
requirements were evaluated or tested during the program. These
can be divided into two categories:
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a. Pressure Restriction Dampers. This type of damper is
illustrated in Figure 5. 23. A check valve with a small opening is
placed between the valve and expander cylinder. This valve allows
unrestricted flow during exhaust valve opening and limits the rate of
pressure buildup during valve closing. A number of variations on
this principle were evaluated, but all fell short of the mark for one or
both of the following reasons. Either the time to close was excessive
for a reasonable seating velocity, occupying more than the allotted
68 crank degrees at 1800 rpm (6. 30 milliseconds) and/or a particular
design operated properly at a particular operating conditions, but did
not operate properly at another operating condition where the shape
of the P-V diagram (i. e. , the forcing function on the valve) was differ-
ent. With variable intake valving, the P-V diagram varies greatly,
depending on the intake ratio and expander rpm..
b. Hydraulic Dampers. A hydraulic type of damper was devel-
oped which operated satisfactorily and this design, illustrated in
Figure 5.24, was selected for incorporation in the expander design;
the key dimensions of the exhaust valve are summarized in Table 5. 5.
During exhaust valve opening, the check valve opens permitting a
large flow area for the lubricant so that the opening characteristics
of the exhaust valve are not affected by the hydraulic valve. During
closing of the exhaust valve, the check valve closes so that the lubri-
cant must flow through the orifice down the center of the check valve.
Since the flow rate through the orifice (which is directly proportional
to the valve velocity) varies as the square of the pressure differential,
the valve velocity is relatively insensitive to the pressure applied to
the top of the valve, and thus to variations in the P-V diagram of the
expander.
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TABLE 5. 5
EXHAUST VALVE DIMENSIONS
Table Seat Diameter
Valve Piston Diameter
Valve Travel
1. 5 inches
0. 75 inches
0. 30 inch
Damping Plunger Diameter 0.40 inch
Valve Return Spring 71 Ibf/in
0. 3 inch recompression with
valve in open position.
Damper Plunger Return
Spring
Orifice Size
20 Ibf/in
0.4 inch precompression
1/32 inch diameter
5-44
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Ol
CONDENSER
CHECK
VALVE
ORIFICE
RESTRICTION
I
IS)
O
Figure 5.23 Pressure Restriction Damper for Exhaust Valve.
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1-2667
EXHAUST
VALVE
VALVE RETURN
SPRING
DAMPING
PLUNGER
DAMPING PLUNGER
RETURN SPRING
CHECK VALVE
a ORIFICE
LUBE OIL FROM
PUMP AT 50
psig
Figure 5. 24 Exhaust Valve Design with Hydraulic Damper.
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THERMO ELECTR O N
CORPOPHtlON
The top of the exhaust valve to which vapor pressure i's applied
is made smaller than the valve seat diameter to limit the pressure
force and the closing velocity of the valve. A seal ring is incorporated
in the top valve guide to reduce leakage by the valve guide, since the
exhaust valve will be open during the initial part of the power stroke
during some operating conditions. This seal ring thus eliminates
significant leakage of vapor from the cylinder through the exhaust
valve under these conditions.
Typical valve motion during closing is illustrated in Figure
5. 25 for a pressure pulse of 60 psig applied to the valve piston.
Curves are presented for no damping and with damping with 1/32 inch
and 1/64 in. diameter orifices. At the maximum travel of 0.3 inch
for the exhaust valve, the valve velocity and total travel time are
summarized in Table 5. 6. An orifice size of 1/32 inch diameter was
selected for the initial design; the optimum orifice size will be ex-
perimentally determined during testing of the preprototype expander.
Measured valve motion during opening is illustrated in Figure
5.26. The opening time is about 5 milliseconds for 0. 3 inch lift
which meets the requirement for an opening time of less than 9. 25
milliseconds at 1800 rpm expander speed. The opening time is
determined primarily by the spring characteristics used in the valve.
5.1.1.3 Expander Layout
The detailed layout of the expander with American Bosch intake
valving and hydraulically damped exhaust valve is shown in Figures
5.27 through 5. 33. The main cylinder block and cylinder head are
of cast iron, the crankshaft is of cast steel, and the pistons and
connecting rods are of cast aluminum. The primary forces are
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TABLE 5.6
MEASURED EXHAUST VALVE CLOSING VELOCITY
AND TOTAL TRAVEL TIME
FOR 0. 3 INCH LIFT AND 60 PSIG PRESSURE PULSE
Case
Seating Velocity
(ft/Bee)
Total Travel Time
(milliseconds)
No Damping
With Damping
1/3Z inch Orifice
1/64 inch orifice
15. 7
12
10
3.25
3.7
3.9
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1-2668
1/32 BLEEDHOLE
DAMPER
1/64 BLEEDHOLE
DAMPER
01 2 3 4 5
TIME FROM START OF OPENING, MILLISECONDS
Figure 5.25 Exhaust Valve Closing Travel vs. Time.
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1-2669
246
TIME(MILLISEC)
8
Figure 5. 26 Exhaust Valve Opening Travel vs. Time.
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o
Figure 5. 27 Expander Layout with American Bosch Hydraulic Valving
Cross Section Through Front Cylinders.
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01
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ro
Figure 5. 28 Expander Layout with American Bosch Hydraulic Valving -
Cross Section Through Rear of Expander Showing Feedpump
and Oil Pump Drive.
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DO
ts)
Figure 5. 29 Expander Layout with American Bosch Hydraulic Valving -
Side Cross Section Through Cylinder Bank and Crankcase.
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1-2673
Figure 5. 30 Expander Layout with American Bosch Hydraulic Valving -
Horizontal Cross Section through Crankcase Along Crankshaft.
5-54
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1-2674
Figure 5. 31 Expander Layout - Cross Section at Rear of Expander
Showing Variable Displacement Vane Pump for
American Bosch Hydraulic Valving System.
5-55
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1-2675
Figure 5. 32 Expander Layout - Cross Section at Rear of Expander
Showing Accessory Drive Bell Housing and Transmission
Interface.
5-56
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1-2676
Figure 5. 33 Expander Layout - Rear View Showing
Accessory Drive.
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THBMMO ELECTRON
- naturally balanced^ while the primary moments are balanced by the
counterweights (note that the crankpins are hollow to reduce rotating
unbalance). The secondary forces are unbalanced in the horizontal
plane and have a peak value of 381 pounds. There is no secondary
pitching moment.
Needle or roller type bearings are utilized throughout to mini-
mize difficulty with bearings during initial development of the pre-
prototype expander. The bearings are sized for a life of approxi-
mately 1500 hours based on expander speed and load equivalent to a
vehicle speed of 60 mph.
In Figures 5. 27 and 5. 31, the location and method of driving
the feedpump and lube oil pump are illustrated. The variable dis-
placement vane pump for operating the inlet valve, shown in Figure
5. 31,' is driven off the same shaft as the feedpump. Figure 5. 28
shows the oil separator and reservoir for the valving oil supply.
The accessory drive bell housing is shown in Figures 5. 32
and 5. 33. The electric starter is shown in Section B-B and the
accessory drive shaft in Section C-C. The timer which controls the
inlet valve opening is attached to the accessory drive shaft housing.
Since the accessory drive is driven at 2. 7:1 speed ratio, the timer
speed is reduced back to expander speed.
In Figure 5. 27, the method of pressure balancing the poppet
intake valves is illustrated. A balancing piston is incorporated in
the valve design as illustrated to balance pressure forces, thereby
greatly reducing, the force required to operate the valve. A port
through the center of the valve provides cylinder pressure on top
of the balancing piston equalizing forces due to the cylinder pressure.
5-58
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THERMO ELECTRON
The pressure of the incoming vapor is applied to the bottom of the
balancing piston and top of the valve equalizing these forces. Piston
type rings are used to reduce leakage around the balancing piston.
The balancing hole is 0. 25 inch in diameter which is large enough
so that flow rates through the valve stem do not cause excessive
pressure differentials.
For ease in manufacturing and assembly, the connecting rod
is split on a bias so that it can be installed and removed through
the bore.
5.1.2 Feedpump Design and Test Results
Many of the features incorporated in the feedpump are dictated
by the overall system requirements. The pump is directly driven
from the expander, since it represents a significant power require-
ment. A positive displacement, piston type of feedpump is the
most suitable type to provide high overall efficiency at high pressure
for a wide range of speeds, with a variable flow rate requirement at
any speed. A multiple piston pump is used to minimize pressure
pulsations without the use of accumulators.
Two feedpump designs, a 5-cylinder axial and a 7-cylinder
radial, were designed, fabricated, and tested in the period covered
*
by this report. Testing of the axial 5-cylinder pump design, as
described in Appendix II, has been terminated in favor of the 7-cylinder
radial feedpump. The design features and test results of the radial
feedpump, which was selected for the Rankine-cycle power system,
are described in the following sections.
*
The axial 5-cylinder pump was developed as part of the EPA
program. The radial 7-cylinder pump development was financed
by Thermo Electron Corporation funding.
5-59
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THERMO ELECTRON
CORPORATION
5.1.2.1 Performance Requirements
The performance requirements for the feedpump are summarized
in Table 5.7.
TABLE 5.7
PERFORMANCE REQUIREMENTS FOR FEEDPUMP
Outlet Pressure
Flow Rate
Inlet Pressure-Operation
Overall Efficiency
Non-operating Temperature
Range
Operating Temperature
Range
Speed Range
850 psia
0 to 17 gpm, modulated to any con-
dition under the peak flow curve
(Figure 5. 40)
4-90 psia
75% at full flow
60% at 30% full flow
-40°F to 150°F ambient
Fluid inlet temperature from
100 - 250 "F; Start-up temperature
from -20°F.
300 to 1800 rpm. The pump is
directly driven from the expander
and uses variable displacement for
flow control.
5.1.2.2 Pump Design
The pump is a reciprocating piston type with variable displace-
ment. Views of the pump are shown in Figures 5. 34 and 5. 35. The
pump has seven cylinders located radially about the axis of the rotating
shaft. Each cylinder houses seal ring pistons which can be varied
from zero to full stroke. The cylinders receive fluid from a common
inlet plenum and discharge into a common outlet plenum. Both the
inlet and outlet valves are simple spring-loaded washer types. The
5-60
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1-2811
Figure 5. 34 Reciprocating Piston Pump with Variable Displacement
Cross Section.
5-61
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1-2678
SECT\OK' C-C
Figure 5. 35 Reciprocating Piston Pticip with Variable
Displacement - Section A-A.
5-62
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THERMO ELECTRON
CORPORATION
piston shoes ride on a septagonal ring with a circular ring used to
return the pistons. These rings can be seen in Figure 5.35.
Variable displacement of the pump is achieved by axially
moving the angled portion of the shaft through the center eccentric
ring. Figure 5. 34 shows the pump in its maximum displacement
position with the shaft fully engaged. As the shaft is pulled out
axially, the stroke of the piston decreases until full extension is
achieved at which point the stroke and, therefore, displacement
are zero.
The feedpump has an aluminun housing with steel pistons and
drive mechanisms. Other important characteristics are sum-
marized in Table 5. 8.
5.1.2.3 Feedpump Test Facility
The feedpump test stand is comprised of two systems, the
mechanical drive system and the fluid system. The feedpump
itself is the connecting link between these systems.
Since the pump is to be tested over a speed range of 300 - 1800
rpm, a variable speed drive (Reeves Model 400) is used to drive the
feedpump. The drive unit is supplied with a tachometer for speed
indication and a strobotac is used to check the tachometer. A
rotating through-shaft torque sensor is used to measure driving
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THERMO ELECTRO
CORPORATION
TABLE 5. 8
FEEDPUMP CHARACTERISTICS
Bore - 1. 50 inches
Stroke - 0. 466 inches
Maximum Displacement - 5. 76 inches
Speed Range - 300 - 1800 rpm
Design Point Discharge
Pressure - 850 psia
Maximum Pumping Rate - 17 gpm
Design Point Pumping
Rate - 15.4 gpm
5.64
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THBRMO ELBCTROM
CORPORATION
torque. To protect the torque sensor from overload, a torque
limiting clutch is used to couple the drive unit to the torque sensor.
The drive unit system is shown in Figure 5. 36.
The pump test loop is shown schematically in Figure 5. 37.
Some of the instrumentation readouts associated with the sensors
in the test loop are shown on the extreme right of Figure 5. 36.
The test loop can be isolated from the pump by closing the ball
valve in the discharge line and the shutoff valve in the intake
line. The reservoir is heated by strip heaters so that the pump
suction pressure can be set by control of the fluid temperature
in the reservoir.
A critical aspect of operation of the test loop is to ensure
that all air has been removed from the loop, particularly on the
intake side of the system. Vents are provided at all possible
traps to ensure that any air may be removed from the system.
A turbine flowmeter with an associated digital read-out is
used to measure the flow rate output of the pump. The rating of
the turbine flowmeter is 1. 8 to 18.4 gpm; suitable calibration
curves have been provided by the manufacturer, Fischer and
Porter Instrument Company.
Variable displacement of the pump was achieved by utilizing an
external double-acting piston actuator with a controlling servo valve
to maintain the set position of the displacement shaft and, therefore,
5-65
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-n
i
Figure 5. 36 Feedpump Test Facility.
-------
(Jl
I
COOUlklGa
WATE.S.
Figure 5. 37 Schematic of Feedpump. Test Loop
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THERMO ELECTRO
CORPORATION
the displacement of the pump. A micrometer dial is used to measure
the axial setting on the displacement mechanism, which allows for
precise indication of actual pistion stroke.
In summary, the test loop provides the following capabilities:
• Variable inlet and discharge pressure.
• Variable inlet temperature.
• Variable pump speed.
• Variable pump displacement.
• Measurement of inlet and discharge static and dynamic
pressures.
• Measurement of inlet and discharge temperature.
• Measurement of pump delivery rate.
• Measurement of actual pump displacement.
• Measurement of pump speed and torque to drive the
pump.
5.1.214 Feedpump Testing
A test version of the pump was fabricated for development testing.
It is essentially identical to the system pump shown in Figures 5. 34
with a main bearing added to the drive shaft side. In the system where
the feedpump is integrated with the expander, this bearing is part
of the expander accessory drive housing. The pressure servo control
was not incorporated into the test pump.
Initial tests on the radial feedpump were performed at conserva-
tive operating conditions with discharge pressures set from 200 psi to
400 psi and speeds not exceeding 700 rpm. The initial testing was per-
formed to accumulate "run-in" time on the pump, to assess the mech-
anical and basic functional operation of the feedpump, and to accurately
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THBRIMO ELECTRON
CORPORATION
correlate the axial movement of the displacement mechanism with the
actual stroke of the pistons. The displacement measurement was es-
tablished to achieve accurate pump performance in subsequent testing,
since volumetric efficiency is a direct function of actual pump displace-
ment. After these initial tests, the pump was disassembled and in-
spected for excessive wear patterns and other mechanical disorders.
The feedpump was found to be in excellent condition and was subsequently
reassembled and installed back in the loop for performance testing.
In Figure 5. 38, the volumetric and overall pump efficiency
(hydraulic power/shaft power) is presented as a function of displace-
ment for discharge pressures from 415 psia to 715 psia and pump
speeds of 500 rpm and 600 rpm. Suction pressure for these tests
varied from 9 to 18 psia. From Figure 5. 38, it is evident that the
pump volumetric and overall efficiencies are insensitive to pump
discharge pressure. Also the efficiencies maintain a high level down
to low displacements (or pumping rates). This characteristic is im-
portant in maintaining a high system efficiency at part-load operating
conditions. The peak overall efficiency of the pump is 78% at these
speeds.
The effect of pump speed on efficiency is indicated in Figure 5. 39
at a discharge pressure of 515 psia and for various displacements. Both
the volumetric efficiency and the overall efficiency decrease as the pump
speed is increased up to the maximum speed of 1800 rpm. At a speed
of 420 rpm, the volumetric efficiency is very close to 100% and the
overall efficiency is 83% . At the maximum speed of 1800 rpm, the
volumetric efficiency has decreased to 72% and the overall efficiency
to 63% . At all pump speeds, the efficiencies are relatively insensitive
to the pump displacement.
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THERMO ELECTRON
CORPORATION
In Figure 5.40, the maximum pumping rate required for the
system is presented as a function of feedpump (or expander speed).
This curve is based on the system performance calculations at the
peak power output. The fraction of total pump displacement and
required horsepower to achieve these maximum pumping rates are
also presented at several pump speeds in Figure 5.40, based on
the experimental performance at a discharge pressure of 515 psia.
In Figure 5.41, a plot is presented of the horsepower, pump displace-
ment, volumetric efficiency, and overall efficiency as a function of
pump speed for the maximum pumping rate, as presented in Figure
5.40, and for a discharge pressure of 515 psia. The peak power
requirement of 6. 55 hp occurs at the maximum speed of 1800 rpm.
The displacement required to achieve the maximum required pumping
rate at 1800 rpm is 48%of full stroke. Since the pump efficiency is
relatively insensitive to discharge pressure, the pumping power
required is closely proportional to the pressure differential across
the pump at a given rpm and displacement. At the maximum discharge
pressure of 830 psia at 1800 rpm, these results indicate a required
pump shaft power of:
830-49psia
6. 55 hp — =* 10. 6 hp.
r 515-35 psia
Testing of the pump will be continued to establish performance up to
the design discharge pressure of 850 psia and to establish acceptable
life for the pump. The tests to date have demonstrated the feasibility
of the concept.
5-70
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1-2680
OPERATING CONDITIONS-
P|N = 9 to 18 PSIA
POUT* 415 PSIA"
- R
515 PSIA
o
X
Q-POUT= 715 PSIA_
A-PQUT» 515 PSIA N«600RPM
100
90
? 80
T 70
^ 60
UJ
S£ 50
u_
u.
uj 40
30
20
10
i i i i i i i i
1.0
2.0
3.0
4.0
DISPLACEMENT ( IN0)
5.0 5.57
Figure 5. 38 Efficiency vs. Displacement for 7-Cylinder Feedpump.
5-71
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cv
IOO
90
801—
70
p-
>: 50
5
30
20
10
TESTS 23-29
OPERATING CONDITIONS
P,N = 35PSIA
POUT«5.5PSIA
O- 100% DISPL. D-46 % DISPL.
i
ro
00
^-84 %
X-68 %
•-57 %
O-38%
® -33 %
200
400
600
800 1000 1200
SPEED N (RPM)
1400
1600
I80O
Figure 5. 39 Efficiency vs. Speed for 7-Cylinder Feedpump.
-------
1-2682
6.10 HPp—^6-20 HP, 73% DISPL
^87%\,6.20 HP,62% DISPL
niqpi ^v.
uioru ^-3^6.25 HP, 54% DISPL
^.55 HP, 48%
DISPL
OPERATING CONDITIONS
PIN = 35 PSIA
POUT=5I5PSIA
I
I
I
I
400
800 1200
SPEED-N (RPM)
1600
2000
Figure 5. 40 Flow Rate vs. RPM for 7-Cylinder Feedpump.
5-73
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10
9
8
7
Ul
o:
3
2
no
100
90
LJ
S
LJ
O
80
_i70
a.
060
<0
p-50
40
o
UJ
u
IZ30I—
U.
LJ
20
10
''VOL
Pm=35 PSIA
PoujS5l5PSIA
* * * DISPL
00
200
400
600 800 1000
SPEED-N ( RPM)
1200
1400
1600
1800
Figure 5. 41 Efficiency, Displacement, and Shaft Power
Input vs. Speed for 7-Cylinder Feedpump.
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THBRMO ELECTRON
CORPOPATIOH
The test pump delivers the required range of flow rates at high
pressure over the operating range of shaft speeds. The efficiency
of the pump remains high over a large variation of displacement.
Pressure pulsations are very small and acceptable over the entire
range of operating speed and displacement tested.
5.1.2.5 Feedpump Control
The function of the feedpump displacement control is to maintain
a constant boiler discharge pressure over the dynamic system opera-
ting range from idle to full power by varying the organic flow rate to
the boiler in response to the boiler outlet pressure. The feedpump
control is illustrated in Figure 5.42 and the position of the control on
t
the .feedpump is illustrated in Figure 5.34. .The control em-
ploys a spool control valve operated by a spring-biased diaphragm to
which the boiler outlet pressure is directly applied. This spool valve
controls application of the feedpump discharge fluid to a power piston,
which is connected to the stroke control shaft of the pump and is used
to vary the displacement of the feedpump. The power piston is re-
quired because of the large force (~6001b) required to vary the pump
displacement.
Steady-state positioning of the shaft is achieved when a force
balance is reached between the pump shaft force and the power piston
force. Force from the pump pistons acts on the inclined section of
the control shaft to yield an axial force component that is applied to
the shaft and power piston assembly. The motion resulting from this
force tends to reduce the pump displacement, unless an opposing
pressure force is applied on the power piston side of the control
shaft. Controlled movement of the shaft and power piston assembly
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TMBRIMO ELECTRON
CORPORATION
is obtained by varying the working fluid pressure applied to the power
piston from the spool control valve (3). In Figure 5.42, the two-land
spool valve is shown in the null position, which indicates that the
required boiler design pressure has been reached. An increase in
boiler pressure will move the diaphragm (4), load spring (5), isolating
bellows (6), and spool valve (3) assembly downward to a new equilibrium
position. This motion ports the power piston cylinder to the low pres-
sure return port (7) of the spool valve, thus decreasing the power piston
pressure and, consequently, the pump displacement. Shaft motion
ceases when the spool valve is again nulled at the boiler discharge
pressure design point.
A decrease in boiler discharge pressure will cause the power
piston pressure to rise as the spool valve assembly moves upward
to meter flow from the supply port (8) to the power piston chamber.
The resulting control shaft motion will increase the pump displacement
until a new equilibrium position is reached at the boiler discharge
pressure design point.
5.1.3 Transmission
The transmission requirements are:
• Completely automatic operation.
• Permit expander to idle at zero vehicle speed.
• A minimum of two-speed ratios for improved
performance.
• High efficiency under both part load and WOT conditions.
• Small enough to install in existing transmission tunnel
with expander next to fire wall.
5-76
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SUPPLY PRESSURE
PORT (8)
LOW PRESSURE
RETURN PORT (7)
DIAPHRAGM
(4)
Ul
BOILER DISCHARGE PRESSURE
LOAD SPRING (5)
BELLOWS (6)
SPOOL VALVE (3)
\\XXXX\\ XXX
\\x\\xxx\
CSJ
STROKE CONTROL
SHAFT (I)
Figure 5. 42 Feedpump Control.
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THERMO ELECTRON
CORPORATION
• Must be available with minimum modifications required
for integration with Rankine-cycle powerplant.
Several transmission choices are available. As part of the previous
work, the Dana Corporation has developed, under subcontract, a layout
design of a transmission specifically for the Rankine-cycle powerplant.
The transmission design is described in Appendix VII. This trans-
mission is a two-speed unit with a hydraulically-operated, slipping
wet clutch used to permit both expander idle at zero vehicle speed
and low speed operation of the vehicle. Above a speed of 8 mph,
the clutch locks up; the transmission then operates as a direct-coupled
unit to provide high efficiency and to use the desirable torque charac-
teristics of the Rankine-cycle expander with variable cut-off intake
valving. A control system was also designed for the low-speed slipping
mode of operation, as well as to provide the desired shifting map
between the two speed ratios. Since this is a special transmission and
would require considerable development it is not presently being con-
sidered for the prototype cars. However, the performance and fuel
consumption characteristics presented in Chapter 4 were calculated
using the characteristics of this transmission.
Another alternative is use of a three-speed manual transmission.
This approach has not been seriously considered due to the strong
preference of the American consumer for an automatic transmission.
The requirement for minimum development cost thus restricted
the choice to off-the-shelf or conventional transmissions. Two choices
exist:
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THERMO ELECTRON
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• Conventional three-speed automatic transmission with
torque converter coupling.
• Conventional three-speed automatic transmission with fluid
coupling.
To use the torque converter coupling with the relatively low-speed
Rankine-cycle expander, a speed-up gear is required between the
expander and torque converter to bring the speed of the input shaft
to the torque converter to a level corresponding approximately to the
speed of the I/C engine. The Ford Motor Company has made per-
formance calculations for several different combinations to provide
a basis for preliminary selection of the transmission and the results
are summarized in Table 5.9. T^hese calculations were based on the
characteristics of Ford's C-4 automatic transmission. System No. 4
from Table 5.9 was selected as providing performance equivalent to
that obtained with the 351 CID I/C engine without exceeding the maxi-
mum torque rating of the C-4 transmission. The performance of the
fluid coupling was definitely poorer, at least for the conditions run.
In Figure 5. 43, the WOT drive shaft torque output of the RCS with
System No. 4 is compared with that of the I/C engine. The RCS
system produces higher torque output below about 30 mph; above
30 mph, the torque is slightly less than that obtained with the 351
CID I/C engine. The characteristics of the selected transmission are
summarized in Table 5. 10; a layout drawing of the overdrive-gear
torque converter portion of the transmission is given in Figure 5. 44.
5.1.4 Rotary Shaft Seal
The system has been designed so that only one dynamic shaft
seal, that on the 3" diameter expander shaft at the rear of the expander,
5-79
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TABLE 5.9
i
oo
o
WOT ACCELERATION PERFORMANCE OF 1972
FORD GALAXIE - EFFECT OF TRANSMISSION
CALCULATIONS BY FOMOCO
Performance Weight = 4486 Ibs
Transmission Rates = 2.40 Low, 1.47 Intermediate, 1.00 High
Rear Axle Ratio
System
No.
1
2
3
4
5
Engine
351 1C
RCS
RCS
RCS
RCS
Axle
Ratio
2.75
2.75
2.75
2. 75
1.50
Speedup
Gear
Ratio
1.00
1.75
2.00
2.25
1.00
Coupling
12" D Torque
Converter
12" D Torque
Converter
12" D Torque
Converter
12" D Torque
Converter
1 1 " D Fluid
Coupling
0-60
MPH Time,
Sec.
14.6
13. 8
14. 1
14.4
16.2
0-10 sec
Distance
ft.
406
450
443
435
389
25-70
MPH Time,
Sec.
15.5
16. 1
16.0
16.3
18.5
50-80
MPH Time,
Sec.
16.0
17.0
17.2
17.2
20. 1
I
0
0
H
a
o
z
-------
1-2702
1000
900
800
700
UJ 600
<— PCS, 2.75 AXL
J* 2.25"O.D., 12"!
\ Tnpmic rriKiwi
RCS, 2.75 AXLE
DIA
TORQUE CONVERTER
3 SPEED AUTOMATIC
S
l-
u.
CO
uj
E
o
500
400
351 CID I/C ENGINE
2.75 AXLE, 12" DIA
TORQUE CONVERTER
3 SPEED AUTOMATIC
300
200
100
I
10 20 30 40 50 60 70
VEHICLE SPEED, MPH
80
90
100
Figure 5. 43 Comparison of Wide Open Throttle Torque with Torque
Converter Coupling and Three-Speed Automatic Trans-
mission, TECO RCS and 351 CID Engine.
5-81
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THKRMO BLBCTRO
TABLE 5.10
CHARACTERISTICS OF SELECTED TRANSMISSION
Overdrive Gear Between Expander and Torque Converter
Type Planetary
Gear Ratio 1:2. 25
Ford C-4 Transmission
12 inch diameter Torque Converter
Gear Ratios 2.40 Low
1.47 Intermediate
1. 00 High
Driveline
Standard Ford Motor Company
Axle Ratio 2. 75:1
5-82
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Figure 5.44 Layout of Torque Converter and Overdrive Gear for Transmission
(12" diameter torque converter).
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THERMO ELECTRON
is required. During system shutdown, the crankcase pressure is
subatmospheric; there is therefore a tendency for air to leak into the
system. During operation, the crankcase pressure will vary from
subatmospheric to approximately 100 psia, depending on operating
conditions and the ambient temperature level. To positively prohibit
either air leakage into the system or working fluid leakage out of the
system, a double seal is used with a prepressurized buffer fluid
between the two seals, as illustrated in Figure 5.45. The pressure
of the buffer fluid between the two seals is maintained above the crank-
case pressure at all times so that any leakage through the inboard seal
is the buffer fluid leaking into the crankcase.
The buffer fluid is the lubricant used in the system. The buffer
pressure is always maintained above atmospheric pressure by the
springs in the buffer fluid reservoir so that any leakage through
the outboard seal is leakage of lubricant to the atmosphere. Leakage
through the outboard seal is collected in the power take-off housing
on the rear of the expander and can be drained at intervals if required.
To reduce the pressure differential across the inboard seal in order
to minimize buffer fluid leakage into the system, the buffer fluid
pressure is controlled by the crankcase pressure, as Illustrated in
Figure 5.45. Provision is also made for charging make-up oil to
the reservoir when required.
The actual seal construction is illustrated in Figure 5.46. Two
face seals are used as illustrated, with the hardened steel mating
ring rotating with the shaft and two stationary, spring-loaded carbon
rings. This seal is manufactured by the Chicago Rawhide Company
and is a modification of a standard seal. Testing on this seal, as
5-84
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U1
I .
00
Ul
BUFFER
ZONE
CRANKCASE
CHARGING
VALVE
SHAFT SEAL
ASSEMBLY
ATMOSPHERE
XX X XX NT
X X XX XX
TO
CRANKCASE
ATMOSPHERE
BUFFER
FLUID
RESERVOIR
-j
o
Figure 5. 45 Double Seal and Buffer Fluid Concept.
-------
Ul
30
STATIC
SEAL
("0-RING)
ATMOSPHERE
BUFFER
FLUID -
SEAL
HOUSING
ASSEMBLY
^ MATING
RING
-STATIC
SEAL
("0-RING)
WASHER
SPRING
SEAL
CARTRIDGE
Figure 5.46 Layout Drawing of Chicago Rawhide Dotble-Shaft Seal.
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THBRIMO ELECTRON
CORPORATION
well as on an alternative type of seal manufactured by the Crane
Company, was carried out at Thermo Electron Corporation under
simulated operating conditions; the detailed test program and data
are presented in Appendix III. The Chicago Rawhide seal was
selected for the preprototype design because of its short length and
acceptable leakage rates. In Table 5. 11, measured leakage rates
are presented over a 3187-hour test with this seal. The leakage rate
when the system is not operating is much lower (by a factor of approxi-
mately 10) than when the seal is operating. The measured leakage
rates are considerably lower than those that can be tolerated in the
system.
The buffer fluid reservoir design is presented in Figure 5. 47.
Bellofram seals are used as illustrated to provide a leaktight reser-
voir and to insure that the buffer lubricant is maintained free of
dissolved air. A level indicator is used to indicate the buffer fluid
level when in the static condition so that make-up buffer lubricant
can be charged to the reservoir when required.
5-87
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TABLE 5, 11
ROTARY SHAFT SEAL TEST
SEAL
TYPE
CHICAGO
RAWHIDE
TYPE OF
OPERATION
CONTINUOUS
ELAPSED
TIME
(HOURS)
3187
TOTAL
0.438
LEAKAGE RATE
PINTS/1000 HOURS)
CRANK CASE
0. 183
OUTBOARD
0. 255
a
2
0
H
a
o
z
oo
oo
LEAKAGE RATE GOALS:
TOTAL (BUFFER) LEAKAGE RATE: 2/2 PINT/1000 HOURS
OUTBOARD LEAKAGE RATE: 1/2 PINT/1000 HOURS
-------
ATMOSPHERE
y x>i|
1 j
ft s
] 1
i 1
/w. . .
FROM CRANKCASE
STATIC CONDITION
FROM CRANKCASE
DYNAMIC CONDITION
Figure 5.47 Buffer Fluid Reservoir.
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THERMO ELECTRON
5. 2 COMBUSTION SYSTEM - BOILER SUBASSEMBLY
In Figures 5.48 and 5.49, the complete assembly cross sections
of the combustion system-boiler subassembly are presented. Two
partial sectional views are presented in Figure 5. 50. These drawings
include the boiler, burners, combustion air blower and motor drive,
atomizing air compressor, and fuel/air controls. Two burners firing
upward are used with control components and the combustion air blower
packaged between the two burners. Two burners operating in parallel
are used in order to obtain a low burner pressure drop within the
packaging constraints for the system and to facilitate obtaining uniform
combustion gas flow through the rectangular tube bundle. The boiler
tube bundle is positioned at the top of the subassembly because of
packaging constraints. Combustion gases flow upward through the
boiler tube bundle, leaving at a temperature of 600 °F or less, depending
on the system power level. The exhaust gases are collected in the outlet
chamber and ducted to the bottom of the engine compartment and to the
rear of the vehicle if required. Provision is also made (not shown in
these figures) for directing part of the exhaust gases to the combustion
blower inlet for control of NO emissions by exhaust gas recirculation.
x
The combustion chamber walls are air-cooled by the incoming com-
bustion air, as illustrated.
A prime consideration in design of the burner-boiler has been
that of minimizing the total power required to operate the unit within
the packaging constraints, not only to limit the parasitic load on the
system, but also to make practical the use of completely electric
drive for the combustion blower and atomizing air compressor. For
rapid cold startup, it is necessary to operate the burner at the
5-90
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COMBUSTION AIR
CONTROL VANE
ATOMIZING
AIR COMPRESSOR
BOILER TUBES
COMBUSTION
CHAMBER
COMBUSTION
BLOWER
AIR ATOMIZING
NOZZLE
FUEL
SOLENOID VALVE
i
ro
-J
o
CD
c
T
i.
Figure 5.48 Front View of Combustion System-Boiler Subassembly
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1-2709
- -
• • •
AIR/FUEL RATH
CONTROLLER
WORKING
FLUID
EXIT
DC MOTOR
Figure 5.49 Side View of Combustion System-Boiler Subassembly.
5-92
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1-2710
Atomizing
Air
Compressor
».uw C-C
Figure 5. 51 Views of Layout Drawing of Figure 5.48
Illustrating Position of Fuel/Air Control.
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THERMO ELECTRON
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maximum burning rate on startup. Since the only power available
on startup is that stored in the battery, it is necessary that both
the atomizing air compressor and combustion blower be driven by
electric motors. Another consideration is design of the combustion
blower so that the shaft power required decreases as the burner is
operated at low firing rates corresponding to part-load operation.
5.2.1 Boiler Design
The boiler design point requirements and characteristics are
summarized in Table 5. 12. In Figure 5. 51, a cross section of the
boiler tube bundle across the narrow dimension is presented; Figure
5. 52 is Section A-A of Figure 5. 51. In Figure 5. 53, cross sectional
views of the tube bundle across the wide dimension are presented.
The organic flow path through the boiler tube bundle is presented
in Figure 5. 54. The liquid organic flow enters the preheat stage at
the middle of tube row No. 2. The flow is split into two parallel
passes through the preheat stage in order to use a smaller tube
diameter, thereby reducing the preheat stage size while still main-
taining an acceptable organic pressure drop. Flow through the preheat
stage is single phase only, so that flow instability with parallel flow
is not a problem. The combustion gases exit from the preheat stage,
which is the lowest temperature portion of the boiler, so that the
boiler efficiency can be maximized without use of air preheat and with
a reasonable overall size. The organic flow enters tube row No. 2
rather than No. 1 in order to have the lowest organic temperature
where the combustion gases enter the preheat stage, and thus minimize
the possibility of overheating the organic.
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THBRMO ELECTRON
TABLE 5.12
BOILER DESIGN POINT CHARACTERISTICS
Combustion Rate
Heat Transfer Rate to FL-85
Fuel
Fuel Flow Rate
Primary Air Flow Rate
Recirculation Gas Flow Rate
AirrFuel Ratio
Combustion Gas Temperature
at Inlet
Combustion Gas Temperature
at Outlet
Efficiency at 100% Load
Efficiency at 10% Load
Combustion Gas Pressure Drop
FL-85 Flow Rate
FL-85 Temperature at Inlet
FL-85 Pressure at Inlet
FL-85 Temperature at Outlet
FL-85 Pressure at Outlet
Maximum Tube Wall Temperature
on FL-85 Side
Maximum Fin Tip Temperature
Maximum Water Jacket Pressure
at 100% Load
Maximum Water Jacket Pressure
at 10% Load
Core Dimensions
Overall Dimensions
Core Weight
Weight of Water
Total Weight (Burner, Boiler,
Insulation, Headers, Water,
Expansion Tanks, etc. )
2. 78 x 10 Btu/hr
2.25 x 108 Btu/hr
JP-4 (HHV = 20096 Btu/lb)
138. 1 Ibs/hr
2468 Ibs/hr
521 Ibs/hr
17.85:1
2975°F
600°F
0.81
0.88
3.40" W. C.
9860 Ib/hr
287°F
826.5 psia
558°F
700 psia
569°F
877°F
1486 psia
1130 psia
34 x 18 x 6. 3 inches
40 x 20 x 8. 3 inches
226 Ibs
8. 15 Ibs
330 Ibs
5-95
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1-2713
(•TD GE0 GUD GTD GL£> GID
GHD GHD
GUDx
Figure 5. 5 1 Cross Section of Boiler Tube Bundle Across Narrow Dimension.
5-96
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1-2714
5E-OTION A "A.
Figure 5. 52 Section A-A of Figure 5. 51 Illustrating Header Construction.
5-97
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1-2715
SCOION B-B
E
•t
•3* .Of
Figure 5. 53 Cross Section of Boiler Tube Bundle
Across Wide Dimension.
5-98
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EXHAUST GAS FLOW
T
Ul
i
vO
PRFHFAT
STAGE
LIQUID ^
INLET *
Ol IDCDLJf AT
oUrtKntAI
STAGE
BOILING >
s _ . , _ ,
A.
^ *
l > *
A o o o 9 q> q> o o o o o o " c> o"o
*- r> r> o r> r> r> o r>— r^ r
TUBE
FLOW
NO.
3C\ C\ 1
^Vx
^ ^^ lg^ C*
•S VAPOR
0^0 <() OUTLET 3
0 0 4
A
> o ^ .s
-J
U)
ro
STAGE
t
COMBUSTION GAS FLOW
FROM BURNER
Figure 5. 54 Flow Path Schematic of Working Fluid Through Boiler.
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THERMO ELECTRON
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The flow leaves the preheat stage close to the boiling tempera-
ture and goes to the boiling stage or tube row No. 5. The boiling
stage is monotube, with straight-through flow to eliminate any possi-
bility of flow instability. This stage is located where the combustion
gases enter, since the organic heat transfer coefficient is highest
in the boiling stage and it is thus located where the combustion gas
temperature is highest.
The working fluid leaves the boiling stage as high-quality vapor
and goes to the first row of the superheat stage (tube row No. 4).
Parallel flow is used in the superheat stage, as illustrated in Figure
5. 55, since flow is single phase only and no problems with flow
instability will be encountered. As in the preheat stage, parallel
flow permits use of a small diameter tube with an acceptable pressure
drop and minimizes the boiler size. The flow passes through three
tubes in parallel except at the exit of each tube row (Nos. 3 and 4),
where parallel flow through two tubes occurs. The superheated vapor
leaves the boiler from tube row No. 3 and goes directly to the expander.
The stagewise characteristics of the boiler at the design point
conditions are presented in Table 5. 13. The boiling and superheat
stages of the boiler are bare tube bundles with a water jacket buffer
to positively prevent either gross or local overheating of the organic
working fluid. Dual tube boiler construction is used for these stages,
as illustrated in Figure 5. 55, with organic flowing through the inner
tube. The annular space between the tube bundles is sealed and filled
with water, with an external thermal expansion tank to permit thermal
expansion of the water. Heat transfer from the hot combustion gases,
flowing around the outer tube, to the inner tube carrying the organic
5-100
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I-27Z1
HOT
COMBUSTION
GASES
BOILING AT
THIS SURFACE
ORGANIC
FLUID
WATER
CONDENSATION
AT THIS SURFACE
Figure 5. 55 Illustration of Water Jacket Operation.
5-101
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TABLE 5. 13
BOILER DESIGN POINT CHARACTERISTICS
Stage
Boiling
Superheat-I
Superheat-II
Preheat
Total
Tube
Row
No.
5
4
3
1 and 2
Heat Transfer
Rate
Btu/hr
625,000
444, 800
224,400
958,000
2.25 x 106
Combustion Gas
Temperature
Inlet
°F
2975
2371
1939
1709
Outlet
°F
2371
1939
1709
607
FL-85
Temp.
In
•F
437
445
502
287
Out
°F
445
502
550
437
Pressure Drop
Gas Side
in w. c.
2.38
0.74
0. 11
0. 17
3.40
Organic
Side
psi
40
30
32.5
24
126.5
Water
Jacket
Pressure
1470
1405
1431
-
Mass of
Core
Without
Water
Ibs.
66
40
40
80
226
Mass of
Water
Ibs
2. 7
2. 7
2. 7
-
8. 1
TUBE SPECIFICATIONS
Boiling Stage
Inner Tube - 1.00" O. D., 0.083" wall
Outer Tube - 1. 313 " O. D. , 0. 093 " wall
Superheater I and II Stages
Inner Tube - 5/8" O. D. , 0.049" wall
Outer Tube - 7/8" O. D. , 0.058" wall
Preheat Stage
5/8" tube expanded (0.577" O. D. , 0. 035" wall)
18 fins/inlet (rippled)
Tube and Header Material = AISI 4130 Steel
ISI
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THBRMO ELECTRON
CORPORATION
occurs by boiling of the water on the inner surface of the outer tube
and condensation of vapor on the outer surface of the inner tube. The
annular gap width is approximately 60 mils. The organic tube wall
temperature can thus not exceed the saturation steam temperature
corresponding to the pressure in the water jacket; the water jacket
pressure provides a convenient and sensitive means of controlling
the maximum temperature to which the organic is exposed. The water
jacket pressure is also used in the startup sequencing to indicate when
the boiler has been heated sufficiently to start cranking the expander-
feedpump assembly. Use of the water jacket permits safe system
startup, even if the boiler is initially dry (empty of working fluid).
Freezing of the water buffer is prevented through use of an inorganic
salt as an antifreeze agent.
The superheat and boiling stages are brazed construction, with
the tubes furnace-brazed into machined steel headers, using a nicro-
coat braze alloy. Each row of these stages is constructed separately
to facilitate leak-checking and to reduce throwaway cost if any
irreparable leak is encountered. The headers are designed, however,
so that any leaks can be repaired. The brazed tube rows are finally
connected by welding connector tubes between the rows. The steps
for fabrication are summarized in Table 5. 14. For external corrosion
protection in the test boilers, it is planned to coat the entire assembly
externally with a brazing alloy.
The preheat stage is a conventional finned tube heat exchanger.
A water jacket is not used in this stage, since the combustion gas
temperature is relatively low and the tubes are filled with liquid
Fluorinol-85; both of these features minimize the possibility of
5-103
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TABLE 5.14
STEPS FOR BOILER FABRICATION
1. Apply the braze compound and assemble the tubes and headers for each row.
2. Furnace - Braze each row subassembly.
3. Leak test. (Mass Spec. ) through water jacket charging port. (In the event of a leak,
isolate the leak location, rebraze with a lower melting nicrobraze and leak test again. )
4. Weld the end caps.
5. Leak test the inner tube circuit for leaks in the welds.
6. Normalize and stress relieve.
7. Weld row connections.
Materials are:
- Tube and header material: - A1S1 4130 Steel (0.9% Cr, 0.27% Mo)
- Tubes coated with nicrocoat-6 for corrosion protection and thermal
fatigue characteristics.
- Nicrobraze-200 as brazing alloy (in the event of a leak use nicrobraze-130).
n
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o
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THBRMO ELECTRO
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overheating in this stage. The fins and tubes are made of AISI 4130
steel, with brazing used to insure good thermal contact between the
fins and tubes as well as to provide external corrosion protection.
5.2.2 Burner and Fuel/Air Supply Designs
The burner design is presented in the layout drawings (Figure
5.48); the burner design point characteristics are summarized in
Table 5. 15. The burner design is based on the burners tested at
Thermo Electron Corporation for emission characteristics, as out-
lined in Appendix VI. The tested burners were designed for a 100 shp
system, so that scale-up to the size for 131. 1 hp was required. Some
modification to the combustion chamber was required for packaging
and for insuring uniform flow into the boiler tube bundle. Steady-
state testing has recently been initiated at Thermo Electron Corpora-
tion on the full-size burner for 131.1 hp system.
With reference to Figure 5.48, the cylindrical combustion
chamber is air-cooled, with the combustion air entering at the top
of the burner and flowing down the space between the combustion
chamber and outer wall of the burner. The combustion chamber
wall operates at approximately 1500°F. At the bottom of the burner,
the air flow direction reverses and flows through swirl vanes into
the combustion chamber. The fuel nozzle is an air-atomizing
Sonicore nozzle, of the same type as used in the burner on which
the emission measurements were made. The section of the com-
bustion chamber close to the nozzle is lined with a ceramic insert
for improved combustion and to prevent local overheating of the
combustion chamber wall in this region at low firing rates. The
combustion chamber wall is flared outward, as illustrated in
5-105
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THERMO ELECTRON
TABLE 5. 15
BURNER DESIGN POINT CHARACTERISTICS
Combustion Rate (HHV)
Number of Burners
Fuel
Fuel Flow Rate
Excess Air
Air Flow Rate
Recirculation Gas at 600°F
Recirculation Gas Flow Rate
Total Combustion Gas Flow Rate
Combustion Gas Adiabatic
Temperature
Combustion Gas Pressure Drop
Ideal Combustion Gas Fan
Power
Ideal Atomizing Air Pumping
Power
Ideal Fuel Pumping Power
Combustion Space Rate
Overall Dimensions
2.78 x 10 Btu/hr
2
JP-4 (HHV = 20098 Btu/hr)
138. 1 Ibs/hr
20%
2468 Ibs/hr
20%
521 Ibs/hr
3127. 1 Ibs/hr
2975°F
4.5" W. C.
1. 14 hp
0.26 hp
0.005 hp
6 ^
1.5 x 10 Btu/hr ft
34 x 17 x 19. 5 inches
Salient Features
1. Sonicore-air atomizing nozzle for fuel spray.
2. Swirl blades for enhanced air-fuel mixing.
3. Recirculation of exhaust gases to control oxides of nitrogen.
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THERMO ELECTRON
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Figure 5.48, above the main combustion zone to diffuse the combustion
gases and to provide a uniform gas velocity over the entire area of the
rectangular boiler tube bundle.
The burner can also be constructed with ceramic lining the entire
combustion chamber wall and without air cooling other than natural
convection. Air-cooling was utilized, however, to minimize the weight
of the burners. The combustion chamber wall is constructed of Type
316 stainless steel; the outer can operates at low temperature and
is constructed of carbon steel.
The experimental testing indicates satisfactory operation of this
type of burner over a turndown ratio of 20:1 (70,000 Btu/hr to 1.4 x
10 Btu/hr). This turndown ratio should be adequate for the system.
As indicated earlier, packaging constraints and the requirement for
minimum burner pressure drop required the use of two burners as
illustrated. Two major problem areas must be anticipated in operation
of two burners in parallel:
• Unbalance of fuel and/or air flow between the two burners,
resulting in non-uniform combustion gas flow through the
boiler and off-optimum fuel/air ratio, causing excessive
emissions.
• Occurrence of flow instability between the two burners,
with pulsing of the burning rate in each burner.
To insure proper balance of fuel/air flow between the two
burners, the air and fuel flow paths to the burners have been designed
to be symmetrical so that the flow paths from the common fuel control
and common air control are identical. In construction of the two burners,
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THERMO ELECTRON
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the characteristics of the two fuel nozzles and the two combustion
chambers will be matched to insure identical fuel/air flow to the
two burners. This procedure will provide adequate balance between
the two burners. If necessary, trim controls can also be incorporated
in the fuel and air supplies to each burner.
To evaluate flow instability, the Northern Research Company
of Cambridge, Massachusetts, provided a stability analysis of the
system. This study involved a computer stability study using an
existing program. The complete system was divided into zones for
the purpose of the calculation. A slight air pressure perturbation
was then introduced in one burner, and the calculation performed
to indicate if the pressure disturbance was damped or continued to
oscillate and grow in magnitude. It was assumed that the fuel flow
to each burner was not influenced by the pressure oscillations, based
on the fuel control design. The conclusions reached were:
• The hot gas side of the burner configuration is stable as
designed. While the calculation indicated the configuration
is stable, the partition between the two burners should be
carried as close to the boiler tube bundle as possible to
reduce interplay between the two burners on the hot side
and reduce the potential for difficulties.
• The main source of instability results from interplay
between the two burners on the cold air side if a common
air blower and air supply is used for both burners. The
calculations indicated this procedure would result in
unstable operation. Use of a separate air supply to each
burner eliminates this source of instability.
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THBRMO BLBCTRON
• A blower whose head-flow characteristic has a negative
slope over the burner operating range should be used.
These conclusions were used in the final design iteration, as presented
in Figures 5.48 and 5.49 and in selection and design of the combustion
air blower.
The combustion air blower requirements and characteristics
are summarized in Table 5. 16, and the design is illustrated in the
layouts of Figures 5.48 and 5.49. The type of blower selected is the
transverse blower, which provides uniform air flow across the im-
peller length. A splitter is constructed as part of the housing to
separate and isolate the air flow to each of the burners while still
retaining use of a single blower wheel and modulating control. Two
separate ducts carry the air to the top of each burner. The blower is
motor-driven at constant speed, with air flow controlled by means of
a pivoted control vane in the exhaust duct. Pivoting of this vane moves
the vortex position in the impeller, thereby modulating the air flow.
This type of blower and control results in reduced shaft power at
low flow rates with constant blower rpm, thereby reducing the blower
parasitic load at low firing rates corresponding to part-load system
operation.
The characteristics of the combustion blower motor are sum-
marized in Table 5. 17. These characteristics are based on a com-
mercially-available dc motor. The fuel pump and atomizing air com-
pressor are also driven by this motor. The fuel pump, whose charac-
teristics are summarized in Table 5. 18, is the gerotor type with
standard rotor elements and a special housing. This pump provides
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CORPORATION
TABLE 5. 16
COMBUSTION AIR BLOWER REQUIREMENTS AND CHARACTERISTICS
Design Point Performance Requirements
Mass Flow Rate of Air at 60 °F 2468 Ibs/hr
Mass Flow Rate of Exhaust Gas at 600°F 521 Ibs/hr
Mean Mix Temperature 165°F
Volumetric Flow Rate at 165°F 770 CFM
Pressure Head 9 inches w. c.
Ideal Fan Power 1. 14 hp
Blower Efficiency 50%
Requirements for Stability and Part-Load Operation
Negative slope on head flow characteristics required independent air
supply to each burner.
Constant speed operation with control for modulating air flow/power
characteristic such that reducing air flow reduces shaft power input.
Design and Construction Specifications
Type Transverse Flow
Impeller Length 7 inches
Impeller Diameter 4 inches
Overall Housing Dimensions 7x7x6 inches
Impeller Construction Die Cast Aluminum
Integral flow splitter used to isolate air flow to each burner.
Pivot control vane in exhaust duct to modulate air flow by modifying
blower characteristic.
Zero flow shaft power = 20% of design point power.
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THERMO ELECTRON
TABLE 5.17
DC MOTOR FOR COMBUSTION SYSTEM
PERFORMANCE REQUIREMENTS
Rpm 7000
Shaft Horsepower 2.68hp
Voltage 12 Vdc
Efficiency 70%
SPECIAL REQUIREMENTS AND QUALITY ASSURANCE
Open-air frame motor.
Provide insulation to work in ambient temperature range of
-20°F to 200°F
Continuous duty, shunt wound.
DESIGN AND CONSTRUCTION SPECIFICATIONS
Diameter 5-9/16 inches
Length 8-1/8 inches
Weight 181bs
5-111
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THBRMO BLBCTRON
CORPODATION
TABLE 5.18
FUEL PUMP DESIGN POINT CHARACTERISTICS
PERFORMANCE REQUIREMENTS
Fuel
Flow Rate
Delivery Pressure
Ideal Pumping Power
Pump Efficiency
Delivery Pressure Control
MATERIAL AND PROCESSES
Rotor Elements
Housing
MAXIMUM OVERALL DIMENSIONS
JP-4
138 Ib/hr
25 psig
0.005 hp
70%
Pressure Bypass Valve with
Accumulator
Steel
Cast Iron
1. 5 inches diameter x 0. 5 inch thick
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THBRIMO ELECTRON
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constant pressure (25 psig) fuel to the burner fuel metering valve as
well as to the combustion air servo-controller; fuel is used as the
actuating medium for these hydraulic-mechanical controllers.
The atomizing air compressor characteristics are summarized
in Table 5. 19. The compressor is identical in construction to
commercially-available vane compressors; however, a special size
is required for this application.
5. 2. 3 Fuel/Air Controller Design
The location of the fuel/air controls are given in Figures 5. 48,
5. 49 and 5. 50, and the functional schematic is presented in Figure
4. 2 of Chapter 4. In this section, the detailed layouts of the control
components are presented. These component designs are similar to
those used in the transient burner emission measurements described
in Appendix VI, with modifications made for the different blower
characteristics and for packaging in the space between the two burners.
The combustion air servo-control is illustrated in Figure 5. 56.
This unit controls the fuel pressure applied to the diaphragm actuator
illustrated in Figure 5.60. Motion of this actuator positions the blower
vane controlling air flow. The fuel pressure to the diaphragm is con-
trolled by the spool valve with constant 25 psig fuel pressure applied
at inlet port A. The motion of the spool valve is controlled by a set
of diaphragms actuated by pressure signals from the boiler inlet
flow rate and boiler outlet temperature and by a cam-operated roller
actuated by a cam on the shaft connected to the blower vane. The
cam-operated roller provides the feedback in the displacement-con-
trolled servo loop. The orifice AP generated by the organic flow rate
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CORPORATION
TABLE 5..19
ATOMIZING AIR COMPRESSOR CHARACTERISTICS
PERFORMANCE REQUIREMENTS
Mass Flow Rate of Air at 60°F 25.3 Ibs/hr
Pressure Head 15 psig
Ideal Pumping Power 0. 26 hp
Efficiency 65%
SPECIAL REQUIREMENTS
Continuous constant speed operation
Oil-free flow output
DESIGN AND CONSTRUCTION SPECIFICATIONS
Type Rotary Vane
External Length 5 inches
External Diameter 3 inches
Speed 3450 rpm
Vane Material Graphite
Drive Shaft and Pulley
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to the boiler is applied at ports B and C. A buffer fluid is used to
transmit the pressure from metal diaphragms located at the orifice
to the controller so that high pressure working fluid is not applied
directly to the controller. Bellows seals are used in the high pressure
part of the controller. The pressure signal proportional to organic
outlet temperature is applied through a port (not shown) to the shaft
side of the diaphragm on the far right of the layout of Figure 5. 56.
The fuel control valve is illustrated in Figure 5. 57. Constant
pressure (25 psig) fuel is supplied at port A, with port B being the
discharge flow to the fuel solenoid valve. The control valve stem is
operated by a cam-operated roller, controlled by a cam connected
to the blower van controlling the combustion air flow rate. This
procedure insures that the fuel flow rate directly follows the com-
bustion air flow rate, providing the desired fuel/air ratio over all
transient conditions. If desired, the fuel/air ratio can be changed
as a function of burning rate by proper shaping of the cam. This type
of control maintained tight tolerance of the fuel/air ratio over all
transients encountered in the transient burning tests at Thermo
Electron Corporation. The full-swing time response of the controller
with a step change in the orifice AP is approximately 200 millisec.
The boiler outlet organic temperature sensor is illustrated in
Figure 5. 58. The prime sensor consists of a stainless steel tube
immersed in the organic flow with a central, low thermal expansion
inner rod passing through the middle of the tube and fastened at the
bottom of the tube. Differential thermal expansion provides movement
of the central rod in response to organic temperature changes. Move-
ment of the central rod operates a small valve to control the signal
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THERMO ELECTRON
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pressure applied to the servo control described earlier. The atomizing
air compressor is used to supply constant pressure air for operation
of this unit. The constant pressure air enters the supply port (4) and
flows through an orifice so that the pressure in the chamber above the
valve, or the signal pressure, depends on the air flow rate. The air
flow rate is controlled by the valve, with the discharge flow vented to
the engine compartment. As the organic temperature increases, the
outlet stainless steel tube expands, moving the central rod downward
and closing the control valve. This change results in a reduction in
the air flow rate through the valve and increases the signal pressure
to the servo control unit, thereby reducing the burning rate. A de-
crease in the organic temperature results in the inverse of this
action. The control valve unit is thermally isolated from the
thermal expansion unit carrying the organic flow to minimize drift
of the controller.
5-116
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01
I
ro
Overall Dimensions:
6-3/4 x 2-1/8" diam. (max)
Materials:
Body - Aluminum Alloy
Metering Elements -
Stainless Steel
Figure 5. 56 Layout Drawing of Combustion Air Servo-Control.
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01
I
CD
Overall Dimensions:
3-1/2 x 2-1/16" Diam. (max)
Materials:
Body- Aluminum Alloy
Metering Elements -
Stainless Steel
Figure 5. 57 Layout Drawing of Fuel Control Valve.
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1-2725
Materials:
Body - Stainless Steel
Core Rod - Invar
Figure 5.58 Boiler Outlet Organic Temperature Sensor.
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5.3 REGENERATOR
The regenerator layout design Is illustrated in Figures 5. 59
and 5. 60, and the regenerator design point requirements and charac-
teristics are summarized in Table 5.20. The liquid and vapor flow
paths are illustrated in Figure 5.61. The exchanger can be classed
as a multipass, cross-counterflow exchanger with both vapor and
liquid mixing between stages, but unmixed within stages. Four stages
are used, with the liquid flowing in six parallel circuits in each stage.
The core is brazed aluminum construction with flat tubes
carrying the liquid and with fins on both the liquid and vapor sides
of the core. The fins on the liquid side serve the dual function of
increasing the heat transfer area and acting as stays to permit the
flat tube walls to withstand the design pressure of 850 psig on the
liquid side with reasonable tube wall thickness. The design is
adaptable to one-step brazing for inexpensive high volume production.
The fins are based on standard fins produced by the Garrett-AiResearch
Corporation, with the basic dimensions summarized in Table 5.20.
Experimental measurements of the heat transfer coefficient and
friction factor as a function of Reynolds' number are available for
the specific fins selected.
The regenerator is also used as an oil separator to collect and
return to the crankcase a major portion of the oil droplets in the
exhaust vapor from the expander. With reference to Figure 5. 60,
a mesh screen (item 10) is placed across the vapor inlet to remove
an appreciable fraction of the oil droplets before the vapor enters
the first stage of the exchanger core. This minimizes fouling of the
5-120
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-- -
1
-
1
1
.
" i! " ' t~ ' 5 " ;; " ~"'i" '~^~ "~"
" (l 1 ' " a
-_t .- _ .-.a. _JL- - ^ - ..It .8 _
F.-. ,.,„.„,,.-_-_-.,_-. -.^-^_-_-_^ ,.,_„__„ ^m .._.-., ,-^, ,„,_.. — „„-„,_
— -"^_ 1 ^_j. —
^^ ^—
I--''
-\,
s
I
i
),
\
*•!•»
X
—
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-------
1-2727
Figure 5. 60 Transverse Cross Section of Regenerator.
5-122
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THBRMO ELBCTRON
TABLE 5. 20
REGENERATOR DESIGN POINT REQUIREMENTS
Rate of Heat Transfer 414, 000 Btu/hr
Effectiveness at 100% Load 81.5%
Vapor Temperatures
Inlet 375°F
Outlet 239°F
Vapor Pressure
Inlet 42. 5 psia
Pressure Drop 1.328 psia
Liquid Temperatures
Inlet 208 °F
Outlet 287°F
Liquid Pressure
Inlet 831 psia
Pressure Drop 4.38 psi
Fluorinol-85 Flow Rate 9924 Ibs/hr
Number of Stages 4
Number of Parallel Liquid Circuits 6
Overall Core Dimensions 24 x 10. 25 x 2. 706 inches
Overall Core Weight I6.38lbs
Liquid Side
Plate Finned Tube-Inside Dimensions 2" x 24" x 0. 025" wall x 0. 160"
Inside Gap
Fins - Plain Plate Fins
20 FPI.
Thickness 0. 009"
Height 0. 160"
Vapor Side
Strip Fin Surface - Strip Fin
25 FPI
Thickness 0.004"
Fin Offset Length (Flow Direction) 1/9"
Height 0.200 inches
Two .screens of 5 mil wire 50% open area for oil separation.
All fabrication out of aluminum alloy 6061.
5-123
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1-2728
VAPOR OUT
c
LIQUID IN
VAPOR
IN
0
i.
LIQUID
OUT
Figure 5. 61 Liquid and Vapor Flow Paths
in Regenerator.
5-124
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THBRMO ELECTRON
CORPORATION
vapor fins of the first two stages by an oil film. The interception
efficiency of the screen is given in Figure 5. 62 as a function of oil
droplet diameter. The vapor fins of the first two stages are used
as an additional separator to remove a fraction of the oil droplets
not collected by the screen separator. The vapor plenums between
the first two stages are connected by a drain line; oil is collected
from both plenums and the oil drains by gravity through a common
line to the crankcase. A fraction (~5%) of the vapor is bypassed
around the first two stages in the regenerator, through the oil drain
line connecting the two plenums, due to the vapor pressure drop
across the core of the first two stages. The vapor velocity is low
enough so that the oil will drain by gravity from the second plenum
against the vapor upflow in this line. Any oil droplets passing the
first two stages will'pass through the remainder of the regenerator
and on through the condenser, pumps, and boiler back to the expander.
The flow path has been designed so that no oil traps exist in these
components.
The regenerator is directly mounted to the exhaust part of the
expander by means of a flanged connector, with a metal "O" ring
seal on the vapor inlet duct to the regenerator.
5-125
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THERMO ELECTRON
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5.4 CONDENSER SUBASSEMBLY
The design of the condenser-condenser fan-condenser fan drive
and controls subassembly has a very important influence on the system
performance and fuel economy. The condenser fan power represents
the largest parasitic power loss in the system. Furthermore, the
expander power output and system efficiency are dependent on the
condenser pressure; each psia decrease in the condenser pressure
represents about 1. 5 hp increase in the expander power output. The
overall goal, approach followed, and the design point selection for
this subassembly, are summarized in Table 5. 21. The condenser
and fans are designed to meet the design point conditions representing
the peak power condition at high vehicle speed. The condenser fan
drive and control is then designed to optimize the fan speed for opti-
mum horsepower and efficiency under part-load and low vehicle speed
operating conditions. An additional consideration has been use of
an inducer to maintain the condenser free of condensed liquid,
so that the entire condenser core is effective for condensation. The
pump subassembly described in Section 5.6 requires no liquid sub-
cooling for proper operation.
5.4.1 Condenser Design
The condenser design parameters are summarized in Table
5.22 and the condenser design is presented in Figure 5.63. An im-
portant consideration has been maximizing the condenser frontal
area within the packaging constraints of the 1972 Ford Galaxie engine
compartment without any major modifications to the vehicle frame
and engine compartment. This large frontal area is particularly
important since, for a given core design and condensing heat load
5-126
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100
90
80
o
? 70
UJ
o
u. 60
UJ
50
gj 40
UJ
I—
20
10
0
25
I
to
SCREEN- 2 USED
-5 MIL WIRE,
50% OPEN
FLOW AREA
50 75
MICRON SIZE OF OIL DROPLET
Figure 5. 62 Interception Efficiency by Regenerator Screen
Separator for Oil Droplets.
IOO
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HERMO ELECTRON
CORPORATION
TABLE 5.21
DESIGN CRITERIA FOR CONDENSING SUBASSEMBLY
OVERALL GOAL
• Reduce Parasitic Fan Power and Maximize System
Performance within Packaging Restraints.
APPROACH
• Maximize Condenser Frontal Area.
• Utilize Condenser Fan Speed Control to Optimize System
Part-Load Performance.
• Minimize Condensing Side Heat Transfer Resistance Relative
to Air Side Heat Transfer Resistance.
• Use air side fin with optimum characteristics for minimum
fan power.
DESIGN POINT FOR SIZING
• Peak System Power Output.
• Ram Air Equivalent to Vehicle Speed of 90 mph.
• Ambient Air Temperature = 85 °F.
• Condenser Average Pressure = 40. 5 psia.
• Equivalent Condensing Temperature = 217°F.
5-128
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THERMO ELECTRON
TABLE 5.22
CONDENSER DESIGN PARAMETERS
HEAT TRANSFER RATE
CONFIGURATION (to maximize A within
packaging constraints)
CORE DIMENSIONS - Central
Side
1.88 x 10 Btu/hr
34.0"x21.0"x4.3"
18, 0"xl3. 0!'x4.3»
Frontal Area - Central 4, 96 ft2
l,625.ft2
8.21 ft2
,757
Side
Total
O
AIR SIDE
- Fin - Garrett
Heat Transfer Area
Air Flow
ORGANIC SIDE
Ideal Air Power
Fin - Garrett
Heat Transfer Area
Organic Flow
P entering
OVERALL HEAT TRANSFER PERFORMANCE
RESISTANCES, % of TOTAL - Air 73%
Organic 26.
__ Wall
UA
Effectiveness
Entering Air
Temperature
Exit Air
Temperature
22R-.326-PERF 9-13)
-.004(AL)
1450 ft2
75, 300 lb/hr
17,300 CFM at 85°F
20, 540 CFM at 188°F
29.8 Btu/in-ft2-°F
3.. 47 in W. C.
9. 38 hp at 85 °F
11,17 hp at 189°F
20 f/in, 0. 004(AL), 1 Strip
270 ft2
9860 lb/hr
5. 5 ft^/sec entering
690 Btu/in-ft2-°F
5 psi
40 psia
0. 2%
20.3 Btu/hr-ft - *F
0,80
85 °F
188. 5°F
5-129
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—I
in
i
i—•
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O
.*;•
OJ
Ul
Figure 5. 63 Condenser Design Layout.
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THERMO ELECTRON
CORPORATION
and temperature, the fan power required is approximately inversely
proportional to the square of the frontal area. A T-shaped configura-
tion has thus been used for the condenser design because of vehicle
packaging constraints. With reference to Figure 4. 8 of Chapter 4,
the top part of the condenser extends across the width of the engine
compartment above the frame; the lower part of the condenser sits
between the frame members and is brought as low as possible within
the limit set by the vehicle driveline illustrated in Figure 4.4 of
Chapter 4.
The condenser design is based on the air-side fin and the
condenser construction recommended by Garrett-AiResearch;
Garrett-AiResearch has concluded an analytical and experimental
study to optimize the air-side fins for Rankine-cycle condensers and
will provide the condenser, shroud, and condenser fans for the pre-
prototype system testing at Thermo Electron. They will also be
performing condensing tests with Fluorinol-85. The condenser is
brazed aluminum construction and is designed for one-step furnace
brazing. The design is thus suitable for high volume production.
The vapor enters an integral top header over the entire top
length of the condenser. The header distributes the vapor flow to
vertical flat tubes operating in parallel with downflow condensation.
The flat tubes have internal fins which serve two purposes: to
increase the heat transfer area on the condensing side thereby
minimizing the condensing side heat transfer resistance, and to
act as stays on the flat tubes to minimize the tube wall thickness
for the condenser design pressure of 150 psia. The air side fins
are slotted fins and have been optimized by Garrett-AiResearch to
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THERMO ELECTRON
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minimize the condenser thickness and condenser fan power. Both
the internal and air-side fins are adaptable for high volume mass
production.
The condensate is collected in bottom headers on the condenser;
the condensate headers are directly connected to the inducer.
Because of the T construction of the condenser, the tubes have differ-
ent lengths in the central portion and outer portions of the condenser.
The tube internal gap is different in these two portions of the con-
denser, as indicated in Figure 5.63, for proper balancing of the
condensing rate with the same pressure drop across both portions.
The condensing side pressure drop at the design point condition is
5 psi.
The ideal fan horsepower with the fans on the rear of the con-
denser is 11.17 hp. The system is designed for maximum utilization
of ram air so that at the design point with a vehicle speed of 90 mph,
7. 0 shp is the required fan power input. A model for the grill and
engine compartment losses, along with the fan characteristics, has
been incorporated in the system performance prediction program
and has been used in calculating the condensing system performance
at the design point as well as at all other operating conditions.
5.4.2 Condenser Fan Design
The condenser fan requirements and characteristics of the
selected fans are summarized in Table 5.23 and the condenser fan
arrangement and drive is illustrated in Figure 5.64. The fan design
was based on recommendations by a consultant, Flowtron, Inc. of
Newton, Massachusetts.
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•RMO ELECTRON
TABLE 5.23
CONDENSER FAN REQUIREMENTS AND CHARACTERISTICS
REQUIREMENTS
• Air Flow Analysis Indicated 1.5" w. c. Required from Fan
at 90 mph Design Point - 2. 0" w. c. Supplied by Ram Air -
Velocity Head Available at 90 mph = 3. 73" w. c.
• Air Flow Rate (Hot Side) = 20, 540 CFM
• Size and Placement Consistent with Good Air Flow Through
Condenser
• Efficiency Maximized Within Packaging Constraints
• Low Noise Level
FAN CHARACTERISTICS
• Tube Axial, Air Foil Blades, Cast Aluminum
• Efficiency at Design Flow, 55%
• Characteristics at Design Point
Number
Required
1
2
Outer
Diameter
inches
24
16
RPM
2700
4050
CFM
10700
4920
each-
Hub
Diameter
Inches
5.5
8.0
Thickness
inches
2.4
2.4
Number
of
Blades
10
16
5-133
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Figure 5. 64 Condenser Fan Mounting on Condenser.
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THERMO ELECTRON
CORPORATION
At the design point, the condenser pressure loss on the air side
is 3. 5 inches W. C. The air flow analysis indicated that 2. 0 inches
W. C. was supplied by ram air at the discharge air flow rate of
20,540 CFM. At 90 mph, the velocity head of the air at 85°F
ambient temperature is 3. 73 inches W. C. , with grill and engine
compartment losses of 1. 73 inches W. C.
The axial length available for the condenser fans is limited by
the available space between the condenser rear face and the burner-
boiler subassembly, and by the requirement for sufficient air-flow
area for exhaust of the condenser cooling air flow. The selected
fans were of the tube-axial with air-foil blades for high efficiency.
Three fans are used to provide uniform air flow over the entire
condenser area. As indicated in Figure 5.64 and Table 5.23, a
24 inch diameter fan is used in the central portion of the condenser,
with two 16-inch diameter fans used on the outer portions of the
condenser. The hub thickness of the fans is 2.4 inches. A shroud
is used on the back of the condenser and the fans are mounted to the
shroud. The largest fans possible were used in the design to provide
good coverage of the condenser as well as to minimize the fan speed
and fan noise.
5.4.3 Condenser Fan Drive and Speed Control
A condenser fan speed control is required for optimum system
performance. At low vehicle speeds, ram air is not effective and
the condenser cooling air must be supplied completely by the con-
denser fans. For wide-open-throttle operation at low vehicle speeds,
a large ratio of fan speed to expander speed is required. If this speed
ratio is maintained up to high expander speeds, the power consumption
5-135
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THERMO ELECTRON
by the fans is excessive; it is therefore desirable to reduce the fan/
expander speed ratio at high expander speeds to optimize the system
performance.
For part-load operation, the condensing load is greatly reduced.
Maintaining the condenser fan speed at that required for the wide-open-
throttle condition results in a larger than optimum air flow rate; the
condenser fan power is again excessive and degrades the system
efficiency and fuel economy. It is thus desirable to have an additional
control responsive in some way to the condensing rate under part-load
conditions. An additional factor is the influence of ambient tempera-
ture. At low ambient temperatures, a lower fan speed is required at
a given operating condition than at high ambient temperatures for
optimum system efficiency.
A preliminary study of the optimum fan speed for various
operating conditions was carried out and the results are indicated
in Figure 5. 65. For the wide-open-throttle condition (maximum intake
ratio) the optimum fan speed is approximately constant above an
expander speed of 700 rpm, and decreases below 700 rpm down to the
idle speed of 300 rpm. For part-load conditions, the optimum fan
speed decreases as the power level drops at any expander speed. The
lower limit for the speed control is set at the optimum corresponding
to an intake ratio of approximately 0. 025. The lower limit is not
crucial since the shaft power to the fan is much less than the maximum
fan power at the low fan speeds.
Many alternative approaches were evaluated for the condenser
fan speed control. In Figure 5. 66, an illustration is given of the
selected control which approximates the desired condenser fan speed
5-136
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3200
2400
Q
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800
i I r
i i i
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MAXIMUM INTAKE RATIO
I I
tV)
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0 200 400 600 800 1000 1200 1400 1600 1800
ENGINE SPEED (RPM)
Figure 5. 65 Optimum Fan Operating Range.
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TEMPERATURE
SENSOR
16 INCH C
FAN £.
(BELT 1.5=1)
Ul
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I L
MORSE ADJUSTABLE
SHEAVE DRIVE ~
EATON TEMPATROL VISCOUS CLUTCH
24 INCH FAN (MOUNTED)
16 INCH FAN (BELT 1.5:1)
CENTRIFUGAL SHEAVE
LJ
INPUT
TORQUE
SHEAVE
ro
^j
v>)
00
Figure 5. 66 Condenser Fan Variable Speed Drive.
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THBRIMO ELECTRON
CORPORATION
variation. Figure 5. 67 shows the actual range of operating conditions
with the control. The control is made up of off-the-shelf components.
To provide the wide-open-throttle speed ratio control, a standard
Morse variable-speed belt drive is used with a centrifugally-controlled
sheave provided to vary the speed ratio. With an Input speed propor-
tional to the expander rpm, this control provides a constant fan shaft/
expander speed ratio of 4. 91:1 up to 550 rpm expander speed. Above
550 rpm expander speed (or 1490 rpm accessory drive shaft speed),
the centrifugal sheave maintains a constant fan shaft speed of 2700
rpm irrespective of expander speed. The maximum fan shaft speed
is thus controlled near the optimum except at expander speeds between
300 and 500 rpm, where the fan shaft speed is somewhat below the
optimum.
To provide control for part-load conditions as well as ambient
temperature variation, a standard Eaton Tempatrol viscous clutch
is used; these units are currently used for controlling the fan speed
in some I/C engine-powered automobiles. This clutch modulates the
fan speed by sensing the air temperature leaving the condenser and,
if the air temperature is low, allows the clutch to slip, thereby
reducing the fan speed. The amount of slip depends on the air tempera-
ture from the condenser. The upper and lower curves of Figure 5. 67
illustrate the operating range of the Eaton clutch, which is satisfactory
for this application.
The viscous clutch is constructed integral with the central
24-inch fan. The 24-inch fan blades are mounted directly on the
clutch rim, and the two 16-inch fans are belt-driven from the clutch
rim at a 1.5:1 speed ratio. The belt drive configuration is illustrated
in Figure 5. 64.
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THERMO ELECTRON
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5.5 STARTUP SEQUENCING, SAFETY CONTROLS, AND
ACCELERATOR PEDAL LINKAGE
The system is designed for completely automatic startup and
operation. The required driver functions are identical to those of
current I/C engines with automatic transmission, that is, ignition
switch, gear shift lever, accelerator pedal, and brake pedal.
The system startup and safety control logic diagram is illustrated
in Figure 5. 68 with the nomenclature defined in Table 5. 24. The
operation of the logic is outlined below:
Startup
When the key is turned on initially, the following events take
place:
1. Gate (1) Energizes the atomizing air compressor, fuel pump,
and combustion air blower.
2. Gate (2) Energizes the accumulator solenoid valve in the
intake valve hydraulic circuit. This applies hydraulic
pressure to the intake valves and keeps the expander intake
valves closed during startup, permitting a faster buildup of
boiler pressure and faster startup.
3. Gate (3) is energized if Gate (4) is not on.
4. Gate (5) is energized when the combustion air and atomizing
air pressures are correct.
5. Gate (6) is energized if the fuel pressure is correct and the
remaining inputs are true.
5-140
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Q 2400
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1600
800
E o
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8%
OPTIMIZED
i i
MORSE ADJUSTABLE SHEAVE DRIVE
OPERATING RANGE
EATON TEMPATROL VISCOUS CLUTCH
200 400
1400
600 800 1000 1200
( ENGINE RPM
810 2700
INPUT RPM VARIABLE SPEED DRIVE
1600 1800
4800
I
PJ
-J
k>>
00
Figure 5.67 Fan Drive Characteristics with Fan Speed Control.
-------
COMBAB KEON
KEON
RESTART -N 2
"OR" GATE
V—"AND" GATE
»"NOT" GATE
Figure 5. 68 System Startup and Safety Control Logic Diagram.
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THERMO ELECT ROM
CORPORATION
TABLE 5.24
NOMENCLATURE USED IN SYSTEM LOGIC DIAGRAM
KEON - KEY ON
COMBAB - COMBUSTION AIR BLOWER ON
CONHIP - CONDENSER HIGH PRESSURE
BOHIP - BOILER HIGH PRESSURE
BUFHIP - BUFFER FLUID HIGH PRESSURE
RUN - STARTING RPM
BUSTAP - BUFFER FLUID STARTING PRESSURE
RESET - RESET
ORHIT - ORGANIC FLUID HIGH TEMPERATURE
COGHIT - COMBUSTION GAS HIGH TEMPERATURE
FUP - FUEL PRESSURE
FLASEN - FLAME SENSOR
RESTART - RESTART
IGNR - IGNITER
FUSOL - FUEL SOLENOID
ACSIG - ACCELERATOR SIGNAL
AC SOL - ACCUMULATOR SOLENOID
IVCON - INTAKE VALVE CONTROLLER (ENABLE)
IVAP - INTAKE VALVE ACTIVATING PRESSURE
STAMO - STARTER MOTOR
BLOP - BLOWER PRESSURE
ATAP - ATOMIZING AIR PRESSURE
5-143
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T H B R IM O ELECTRON
CORPOR4TION
6. Gate (7) is energized and the igniter turned on if Gate (8) is
de-energized.
7. Gate (9) energizes the fuel solenoid.
8. Gate (11) is energized by the flame sensor. This holds the
fuel solenoid on through Gate (9) when the igniter is turned
off if ignition has been obtained. A manual restart is provided
through NOT Gate N2.
9. Time Delay TDl energizes Gate (8), which locks up through
Gate (10), since the other inputs to Gate (10) are normally
true. If a signal from the flame sensor is not received during
the time delay period, TDJL^ opens and the fuel solenoid is closed.
Restart can be attempted using the manual restart button.
10. Gate (8) energizes the NOT Gate Nl, which de-energizes the
igniter through Gate (7).
11. Gate (12) energizes the starter motor until the idle rpm or
Run Speed is reached. This input is taken from the intake
valving system controller. % The starter motor is energized
when the buffer fluid pressure, BUSTAP, reaches a preset
level.
12. Gate(13) energizes the intake valve controller, which enables
the intake valve hydraulic circuit.
13. Gate (14) is energized when the idle rpm is reached. This
enables the accelerator signal to the intake valve controller.
5-144
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THERMO ELECTRON
'-^» IM ^H_^^
T:
1. Gate (15) de-energizes the fuel solenoid if the maximum
pressures are exceeded in the boiler discharge, buffer fluid,
or condenser. Lock-up, requiring a manual reset is provided
with Gates (17). (4), and NOT Gate N3.
2. Gate (16) de-energizes the fuel solenoid when the buffer fluid
pressure drops below the starting pressure during a running
operation. The same reset circuit is used, as above.
3. Gate (6) de-energizes the fuel solenoid when the fuel pressure
is low, or when the temperatures of the combustion gas or
organic fluid are high. Restart is automatic.
4. Gates (17), (18). (19), N4, and N5_lock out the starter motor
circuit after it has been de-energized. The key must be turned
off to reset the circuit.
In Figure 5.69, the circuit schematic which provides the startup
sequencing and safety shutoff of the system is presented. The control
box with this circuit is illustrated in Figure 5. 70; the entire control
box occupies a space 3-5/8" x 2-1/4" x 3".
An electrical interface through a LVDT is used between the
accelerator pedal and the intake valving control system. The schematic
of the valving control system is illustrated in Figure 5. 71 and the
accelerator linkage with the LVDT is illustrated in Figure 5. 72. The
intake valving control system includes an automatic advance, intake
ratio upper limit as a function of expander rpm to prevent exceeding
the boiler capacity and a hydraulic pressure control to reduce the
hydraulic pressure at low engine speeds when extremely rapid intake
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THERMO ELECTRON
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valve events are not required, thereby minimizing the hydraulic
pump power for the intake valves. Other than the accelerator
linkage, these controls will be supplied by American Bosch with
the intake valving assembly for the four-cylinder expander.
5-146
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IJ1
I
Figure 5.69 Circuit Schematic for Startup Sequencing and Safety Controls.
-------
1-2741
Figure 5. 70 Startup Sequencing and Safety Control Box.
5-148
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DURATION
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ACCELERATOh
PEDAL
>s START
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ACCUMULATOR
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(VARIABLE
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EXPAMDEF
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37
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Figure 5. 71 American Bosch Valving Schematic.
-------
SECT B-B
Figure 5.72 Accelerator Linkage to LVDT for
Intake Valve Timing Control.
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THBRIMO ELECTRON
CORPORATION
5. 6 BOOST PUMP-INDUCER-RESERVOIR SUBASSEMBLY
The components of this subassembly provide the following
functions:
• Produce the net positive suction head (NPSH) required by
the feedpump for normal operation.
• Produce the NPSH required by the feedpump at the start
condition, when condenser pressure is low and all working
fluid is at the ambient temperature.
• Eliminate required condenser subcooling, thereby making
more efficient use of the available frontal area for condensing.
• Provide reservoir capacity for inventory transfer during
start condition and transient operation.
• Prevent separation of lubricant from working fluid in
condenser and reservoir.
In Figure 5. 73, a flow schematic is presented illustrating the position
of these components between the condenser and the feedpump. The
inducer is located at the bottom of the engine compartment and serves
as the condensate sump pump to maintain the condenser drained of
liquid. Part of the flow from the boost pump is used for operation of
the inducer. The inducer discharges into the system reservoir located
at the top of the engine compartment. This reservoir insures sufficient
suction head for proper operation of the centrifugal boost pump during
startup and transient operation when working fluid inventory changes
in the system may occur. The top of the reservoir is vented to
the condensate header at the condenser, so that the reservoir and
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THERMO ELECTRON
CORPORATION
condensate header pressures are equal. The boost pump provides
sufficient pressure to the main feedpump suction line to insure proper
inlet valve operation and to insure no cavitation.
This pumping assembly is similar to that used on the 5-1/2 hp
systems under test at Thermo Electron, except no inducer is used.
In these systems, a horizontal condenser is located on top of the
powerplant; the condensate drains directly into the reservoir located
under the condenser. In the automotive system, the condensate header
is the lowest part in the system, and the inducer was added for
draining the condenser and pumping the condensate into the reservoir.
5.6.1 Boost Pump Design
The boost pump requirements are summarized in Table 5.25.
The boost pump is driven by the output from the Morse variable speed
drive used for the condenser fan speed control (see Section 5. 4). The
requirements are thus given for two conditions, one representing
operation at the expander idle speed (boost pump speed = 1470 rpm)
and the other at expander speeds above 550 rpm where the boost pump
speed is constant at 2700 rpm. The bottom of the reservoir is 10 inches
above the boost pump suction and insures 10 inches head to the boost
pump when the reservoir has a low liquid level due to transfer of the
working fluid inventory to other locations in the system.
The boost pump layout drawing is illustrated in Figure 5. 74. The
pump is a centrifugal type with an impeller diameter of 3. 5 inches.
A screw inducer is used on the pump to meet the NPSH requirement
of 10 inches at 2700 rpm. To eliminate the requirement for a dynamic
shaft seal, a permanent magnet drive is used. The magnet drive is
commercially available and is the type used for the centrifugal boost
5-152
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VENT
(JI
CONDENSER
llNDUCER
I BOOST
RESERVOIR
FEED PUMP
I
N
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Figure 5.73 Boost Pump-Inducer-Reservoir Subassembly.
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1-2750
TABLE 5. 25
BOOST PUMP PERFORMANCE REQUIREMENTS
• BOOST PUMP AT 1470 RPM (EXPANDER 300 RPM)
Maximum Flow Rate to Feedpump . 7. 5 gpm
Maximum Flow Rate to Inducer 8. 5 gpm
Maximum Total Boost Pump Flow Rate 16 gpm
Head Rise 7. 88 feet
Minimum Available Net Positive Suction
Head 5 inches
Shaft Power . 0. 125 hp
• BOOST PUMP AT 2700 RPM (EXPANDER 550- 1800 RPM)
Maximum Flow Rate to Feedpump 17 gpm
Maximum Flow Rate to Inducer 12. 4 gpm
Maximum Total Boost Pump Flow Rate 29. 4 gpm
Head Rise 26. 6 feet
Minimum Available Net Positive Suction
Head 10 inches
Shaft Power 0. 5 hp
5-154
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1-2745
DIA
Figure 5. 74 Cross Section of Centrifugal Boost Pump.
5-155
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THBRMO ELECTRON
CORPORATION
pump on the 5. 5 hp TECO system. In Figure 5. 75 the impeller vane
and volute construction are illustrated. The impeller uses backward
curved vanes as illustrated.
The shaft power required to drive the boost pump is 0. 5 hp at
2700 rpm and 0. 125 hp at 1470 rpm. The parasitic load from the
boost pump is thus very low.
The pump is constructed primarily of aluminum.
5.6.2 Induce r
The inducer performance requirements are summarized in
Table 5. 26 and the inducer design is presented in Figure 5. 76. The
required head output is 2. 0 feet to pump the liquid from the bottom
of the condenser to the reservoir located at the top of the engine
compartment. The pump is designed to operate with about 2 inches
NPSH. As evident from Figure 5. 76, the inducer construction is
very simple. Brazed aluminum construction is used and the design
is suitable for high volume production. The inducer is brazed to the
condenser and reservoir lines.
5. 6. 3 Reservoir
The reservoir (or receiver) is illustrated in the system packaging
drawing of Figure 4. 7 of Chapter 4. It is an aluminum tank with a
capacity of 1. 16 gallons and dimensions 5" diameter by 10" length
plus headers. The capacity for the design was based on the total
condenser internal volume, so that the system can be started with
the condenser completely filled with liquid. The final reservoir size
required for the system will be determined experimentally in operation
of the complete system to insure adequate capacity for all startup and
transient conditions encountered by the system.
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Figure 5. 75 Boost Pump Vane and Volute Construction.
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TABLE 5-26
INDUCER PERFORMANCE REQUIREMENTS
• BOOST PUMP AT 1470 RPM (EXPANDER 300 RPM)
Primary (Nozzle) Flow Rate 8. 5 gpm
Primary Head Available 7. 88 feet
Secondary (Condensate) Flow Rate 7. 5 gpm
Inducer Head Rise 2. 0 feet
• BOOST PUMP AT 2700 RPM (EXPANDER 550 - 1800 RPM)
Primary (Nozzle) Flow Rate 12.4 gpm
Primary Head Required 16. 8 feet
Secondary (Condensate) Flow Rate 17 gpm
Inducer Head Rise 2. 0 feet
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Figure 5. 75 Boost Pump Vane and Volute Construction.
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TABLE 5-26
INDUCER PERFORMANCE REQUIREMENTS
• BOOST PUMP AT 1470 RPM (EXPANDER 300 RPM)
Primary (Nozzle) Flow Rate 8. 5 gpm
Primary Head Available 7. 88 feet
Secondary (Condensate) Flow Rate 7. 5 gpm
Inducer Head Rise 2. 0 feet
• BOOST PUMP AT 2700 RPM (EXPANDER 550 - 1800 RPM)
Primary (Nozzle) Flow Rate 12.4 gpm
Primary Head Required 16. 8 feet
Secondary (Condensate) Flow Rate 17 gpm
Inducer Head Rise 2. 0 feet
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r— i.so—i
ui
i
ui
\O
.035 TYP
'..K
SECTION
Figure 5.76 Inducer Design.
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THERMO ELECTRON
CORPORATION
5. 7 ACCESSORY AND AUXILIARY COMPONENTS
The automotive accessories for passenger comfort and con-
venience selected for the system and incorporated in the packaging
described in Chapter 4 are:
• Power Steering - Identical to that on 1972 production Ford
Galaxie.
• Power Brakes - Identical to Ford preproduction hydraulic
power brake unit.
• Air conditioning Compressor - Identical to that on 1972
production Mach IV Lincoln (swash plate type).
• Heater and Air Conditioning Package - Identical to that on
1972 production Ford Galaxie. To provide hot water to
the heating coil, a small heat exchanger will be located in part
of the exhaust gas stream from the boiler with circulatory
water for transfer of heat.
It was possible to retain the production heating-air conditioning package,
even though this package extends into the engine compartment.
The battery-alternator supplies power to both the system and
the normal automotive functions requiring electrical power, such
as headlamps and the heating - air conditioning blower. A detailed
analysis was carried out to insure selection of an adequate alternator
and battery capacity for the system. The primary electrical power
demands for the Rankine-cycle system are for operation of the com-
bustion system and for startup. The electrical system is 12 V dc as
in current automotive practice.
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THERMO ELECTRON
CORPORATION
The starting power requirements are summarized in Table 5. 27.
For the first 25 seconds, the combustion system only is operating
with a total electrical power requirement of 2. 75 hp. At the end of
25 seconds, the boiler is heated and the starter motor is cranking the
expander and valve drive pump, feedpump, boost pump, condenser
fans, and automotive accessories. If sufficient boiler pressure exists,
as it would if the boiler contained working fluid, the expander will
immediately take over. If the boiler is dry, working fluid will be
pumped into the boiler, generating the required pressure to operate
the system. To handle the latter situation, a maximum of 10 seconds
operation of the starter motor is required. The electrical power
input for the starter motor is 2. 0 hp.
In Table 5.28, the battery supply requirements for the starting
sequence are summarized for a nominal 12 V dc system based on
these power requirements. The required amps include allowance for
voltage drop with the current draw.
The selected battery to meet these starting requirements is a
standard AABM size 24C battery with a capacity of 84 amp-hrs and
a high rate discharge at 0°C of 500 amps. To insure that this size
was adequate for at least two sequential startup attempts, a simulated
startup test was made with a 96 amp-hr battery. The results are out-
lined in Table 5.29. The battery supplied basically the same voltage
over both starts, with the terminal voltage dropping to 10. 5 volts.
in the pre-start operation and to 7.8 volts in the cranking operation.
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TABLE 5. 27
STARTING POWER REQUIREMENTS
PRE-START (25 sec) TOTAL POWER: 2. 74 HP
• COMBUSTION AIR BLOWER (FULL POWER): 2. 28 HP
• ATOMIZING AIR COMPRESSOR: 0.41 HP
• IGNITER, FUEL PUMP AND CONTROLS: 0.05 HP
STARTER MOTOR (10 SEC) TOTAL POWER: 2. 00 HP
(CRANKING SPEED: 300 RPM)
• EXPANDER: 0. 38 HP
• FEEDPUMP: 0. 55 HP
• VALVE DRIVE PUMP: 0. 70 HP
• BOOST PUMP: 0. 17 HP
• ACCESSORY DRIVE (CONDENSER FANS, ALTERNATOR,
POWER STEERING PUMP): 0.20 HP
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TABLE 5. 28
BATTERY SUPPLY REQUIREMENTS
(FOR NOMINAL 12 V SYSTEM)
PRE-START (25 SEC)
• CURRENT DRAW: 240 AMPS
STARTER MOTOR (10 SEC)
• CURRENT DRAW: 180 AMPS
COMBINATION OF PRE-START AND STARTER MOTOR (10 SEC)
• CURRENT DRAW: 420 AMPS
BATTERY SELECTION
• AABMSIZE: 24 C
• CAPACITY: 84 AMP - HR
• HIGH RATE DISCHARGE AT 0°F: 500 AMPS
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TABLE 5.29
SIMULATED START-UP TEST
FOR BATTERY
FIRST START - TERMINAL VOLTAGE: 12.5 VOLTS
• PRE-START (25 SEC) - V = 10. 5 V: I = 225 AMPS
• CRANKING (10 SEC) - V= 7.8V: I = 450 AMPS
INTERVAL TIME BETWEEN FIRST AND SECOND START: 20 SEC
SECOND START - TERMINAL VOLTAGE: 12. 5 VOLTS
• PRE-START (25 SEC) - V = 10. 3 V: I = 220 AMPS
• CRANKING (20 SEC) - V = 7.8V: I = 450 AMPS
NOTES:
• BATTERY CAPACITY: 96 AMP-HRS
• ALTERNATOR DISCONNECTED DURING TEST SEQUENCE
• TEST PERFORMED ON 1971 FORD LTD
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CORPORATION
The alternator must be sized to handle the maximum sustained
power demand to be encountered in vehicle operation as well as to
handle the worst transitory power requirements resulting from system
operation on long grades. For these transitory conditions, it is
assumed that power will be drawn from both the battery and the
alternator; the alternator must have sufficient capacity to prevent
total discharge of the battery. In Table 5. 30, the alternator specifica-
tions and the alternator requirements for sustained driving conditions
are presented. The alternator selected is a standard, heavy-duty
14 V dc alternator with 130 amp output at 5000 rpm. This size was
based on a continuous current draw of 85 amps for continuous system
operation at 70 mph on a 0% grade, plus 35 amps required for normal
operation of the vehicle (headlamps, etc.).
The worst transitory condition is operation on long grades re-
quiring high system power output and a resultant high electrical load
for operation of the combustion system. For these conditions, power
would be taken from both the battery and the alternator. For operating
conditions requiring 80% of full system power, the total current re-
quirement is; 236 amps; 130 amps are supplied by the alternator and
106 amps are supplied from the battery. Two operating conditions
were evaluated to determine the allowable time and total distance
traveled with the selected battery-alternator combination, with the
results indicated in Table 5.31. At 15 mph on a 30% grade, 80% of
full system power is required. The vehicle could travel 3. 5 miles
continuous at these conditions and the selected alternator-battery is
more than adequate. For 70 mph vehicle speed on a 5% grade, 77%
of full power is required. The vehicle could travel 17. 5 miles at these
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TABLE 5.30
SUSTAINED POWER REQUIREMENTS
ALTERNATOR SPECIFICATIONS
• STANDARD HEAVY DUTY 14 VOLT ALTERNATOR
• 130 AMP OUTPUT AT 5000 RPM
• 90 AMP OUTPUT AT 2000 RPM (IDLE)
FEDERAL EMISSION DRIVING CYCLE RATIOED FOR 200 MILES
• AVERAGE BURNING RATE: 12% FULL POWER
• CURRENT DRAW: 56 AMPS
CRUISE AT 70 MPH ON 0% GRADE FOR 200 MILES
• BURNING RATE: 30% FULL POWER
• CURRENT DRAW: 85 AMPS
NORMAL ELECTRICAL LOADS
• ACCESSORIES (AIR CONDITIONING, HEADLIGHTS, ETC. )
• CURRENT DRAW: 35 AMPS
MAXIMUM SUSTAINED POWER REQUIREMENTS
• CRUISE AT 70 MPH PLUS NORMAL LOADS
• CURRENT DRAW: 120 AMPS
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TABLE 5.31
TRANSITORY POWER REQUIREMENTS
BATTERY YIELD TAKEN AT 0°F
15 MPH VEHICLE SPEED ON .30% GRADE (80% FULL POWER)
• ALLOWABLE TIME AT THIS CONDITION: 13. 8 MINUTES
• MILEAGE TRAVELED 3. 5 MILES
70 MPH VEHICLE SPEED ON 5% GRADE (77% FULL POWER)
• ALLOWABLE TIME AT THIS CONDITION: 15 MINUTES
• MILEAGE TRAVELED: 17. 5 MILES
CONDITION OF 80% FULL POWER BURNING RATE
• TOTAL CURRENT REQUIREMENT: 236 AMPS
• ALTERNATOR SUPPLY: 130 AMPS
• BATTERY CURRENT DRAW: 106 AMPS
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conditions before difficulties were encountered with the battery draw.
In Table 5.32, the worst grades in the U. S. A. are summarized, with
the most difficult being 11.2 miles with an average grade of 5. 1% . In
Table 5. 33, the recorded grades on a cross country round trip between
Chicago, Illinois and Portland, Oregon are presented. For the 3058. 5
mile trip, 24. 0 miles total had a grade of 5% , 10. 5 miles total had a
grade of 6% and 3. 0 miles had a grade of > 7% .
The selected alternator-battery should thus be adequate for all
sustained and transitory driving conditions to be encountered by the
vehicle.
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TABLE 5.32
HIGHWAY GRADES IN SOUTHWEST UNITED STATES
LOCATION
SUPERIOR, ARIZONA
DAVIS DAM, ARIZONA
GRAPEVINE, CALIFORNIA
BAKER, CALIFORNIA
JACOB LAKE, ARIZONA
FARNEL, ARIZONA
GRADE LENGTH
(MILES)
4.0
11.2
13. 8
17.0
13.3
2. 1
MINGUS MOUNTAIN, ARIZONA 2. 6
AVERAGE
GRADE
(%)
5.2
5. 1
3.4
3.4
3.5
6.0
6.0
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TABLE 5.33
HIGHWAY GRADES FOR INTERSTATE ROUTES
PORTLAND TO CHICAGO CROSS-COUNTRY RUN IN 1966
GRADES RECORDED FOR 3058. 5 MILE ROUND TRIP
RECORDED GRADES
(+ or - 1/2 % )
0%
1%
2%
3%
4%
5%
6%
7+%
MILES
1,399.0
1,090.0
371.0
118.0
43.0
24.0
10. 5
3.0
PERCENT
45.7
35.6
12.1
3.9
1.4
0. 8
0.3
0.2
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