EPA-460/3-73-007
CONDENSER
AND FAN DEVELOPMENT
FOR RANKINE
CYCLE ENGINES
U.S. ENVIRONMENTAL PROTECTION AGENCY
Office of Air and Water Programs
Office of Mobile Source Air Pollution Control
Alternative Automotive Power Systems Division
Ann Arbor, Michigan 48105
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EPA-460/3-73-007
CONDENSER
AND FAN DEVELOPMENT
FOR RANKINE
CYCLE ENGINES
Prepared by
J. Killackey, R. Morgan, C. Morse,
B. Foster, and C. Lee
AiResearch Manufacturing Company
2525 West 190th Street
Torrance, California 90509
Contract No. 68-01-0407
EPA Project Officers:
W. Zeber, P. Sutton, and E. Beyma
Prepared for
U.S. ENVIRONMENTAL PROTECTION AGENCY
Office of Air and Water Programs
Office of Mobile Source Air Pollution Control
Alternative Automotive Power Systems Division
Ann Arbor, Michigan 48105
November 1973
-------
This report is issued by the Office of Mobile Source Air Pollution Control,
Office of Air and Water Programs, Environmental Protection Agency, to report
technical data of interest to a limited number of readers. Copies of this
report are available free of charge to Federal employees, current contractors
and grantees, and non-profit organizations - as supplies permit - from the
Air Pollution Technical Information Center, Environmental Protection Agency,
Research Trianale Park, North Carolina 27711, or may be obtained, for a
nominal cost, from the National Technical Information Service, 5285 Port
Royal Road, Springfield, Virginia 22151.
This report was furnished to the U.S. Environmental Protection Agency by
AiResearch Manufacturing Company, Torrance, California, in fulfillment of
Contract No. 68-01-0407 and has been reviewed and approved for publication
by the Environmental Protection Agency. Approval does not signify that the
contents necessarily reflect the views and policies of the agency. The
material presented in this report may be based on an extrapolation of the
"State-of-the art." Each assumption must be carefully analyzed by the
reader to assure that it is acceptable for his purpose. Results and
conclusions should be viewed correspondingly. Mention of trade names or
commercial products does not constitute endorsement or recommendation for
use.
Publication No. EPA-460/3-73-007
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PREFACE
This report (AiResearch Report 73-903*0 describes development conducted
by the AiResearch Manufacturing Company of California, a division of The
Garrett Corporation, for the Environmental Protection Agency, Advanced
Automotive Power Systems Development Division, under Contract No. 68-01-0407-
The work was performed under the direction of Mr. K.O. Parker of AiResearch
and Messrs. W. Zeber, P. Sutton, and E. Beyma of EPA.
i i i
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CONTENTS
Sect ion Page
1 INTRODUCTION AND ABSTRACT 1-1
Introduction 1-1
Abstract 1-1
2 CONCLUSIONS AND RECOMMENDATIONS 2-1
Conclusions 2-1
Recommendations 2-2
3 PERFORATED FIN DEVELOPMENT 3-1
Surfaces 3-1
Test Cores 3-6
Test Procedure 3-6
Test Results 3-9
4 CONDENSING HEAT TRANSFER TESTS 4-1
Test Objectives 4-1
Test Core Descriptions 4-1
Open-System Tests 4-4
Closed-System Tests 4-11
5 CONDENSER DESIGN, FABRICATION, AND TEST 5-1
Heat Transfer Design 5-1
Structural Considerations 5-16
Detail Des ign 5-30
Fabrication 5-34
Heat Transfer Performance Test 5-66
6 FAN DESIGN, FABRICATION, AND TESTING 6-1
Design 6-1
Fabrication 6-10
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CONTENTS (Continued)
Sect ion Page
7 CONDENSER AND FAN AIRFLOW TEST 7'1
8 INSTALLATION AIRFLOW TEST 8-1
9 REFERENCES 9-1
Appendix SPECIFICATION A-1
VI
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ILLUSTRATIONS
Figure No.
1-1 Aerojet Condenser and Fan Assembly (Front View) 1-2
1-2 Aerojet Condenser and Fan Assembly (Back View) 1-2
1-3 Aerojet Condenser Assembly (Top View) 1-2
1-4 Thermo Electron Condenser and Fan Assembly (Front View) 1-3
1-5 Thermo Electron Condenser and Fan Assembly (Top View) 1-3
1-6 Thermo Electron Condenser and Fan Assembly (Back View) 1-4
1-7 Steam Engine Systems Condenser and Fan Assembly 1-5
(Front View)
1-8 Steam Engine Systems Condenser and Fan Assembly 1-5
(Back View)
3-1 Perforated Fin Materials for Cores -21 and -23 3-2
3-2 Effect of Air-Side Fin Geometry on Condenser Length 3,-k
and Requi red Fan Ai r Horsepower
3-3 Effect of Fin Height on Fan Air Horsepower at a Fixed 3~5
Condenser Size
3-4 Typical Test Core for f and j Factors (22 Fins per 3"7
Inch, Perforated, 0.004-in. Aluminum Material)
3-5 Perforated Fin Test Setup Schematic 3-8
3-6 Test Setup Showing Water Flowrate Measuring System J>-\Q
and Thermocouple Measuring Apparatus
3-7 Test Setup Showing Test Core Mounted in Ducting 3"H
3-8 Turbulator Rods Design 3-12
3-9 Performance Parameters for the -21 Perforated Fin 3-13
3-10 Performance Parameters for the -23 Perforated Fin 3-14
3-11 Condenser Solutions for Tested Fins 3~15
4-1 Test Core Assembly Condensing Heat Transfer Test 4-2
v i i
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ILLUSTRATIONS (Continued)
Fi gure No. Page
4-2 Steam-Side Fins in Variable Geometry Test Core 4-3
4-3 Condensing Pressure Loss vs Flow Rate in Each 4-5
Condenser Segment of Variable Geometry Core
4-4 Distribution of Vapor Flow Rate in Variable 4-6
Geometry Core
4-5 Open-System Test Schematic 4-7
4-6 Open-System Performance of Condenser Test Modules 4-8
4-8 Closed-System Test Schematic 4-10
4-9 Test Setup Used for Closed System Condensing Tests 4-13
4-10 Closed-System Performance of Condenser Test Modules 4-15
as a Function of Effectiveness
4-11 Closed-System Performance of Condenser Test Modules 4-16
as a Function of Subcooling
5-1 Model Used in Oil Film Analysis 5-4
5-2 Estimated Condensing Heat Transfer Coefficient and 5-7
Oil Film Thickness for Fluorinol-85 Condenser with/
wi thout Oil Fi1m
5-3 Estimated Condensing Heat Transfer Coefficient and 5-8
Oil Film Thickness for Steam Condenser with/without
Oi1 Film
5-4 TECO Condenser Performance Calculations, 5~9
Desuperheat and Condensing Sections
5~5 TECO Condenser Performance Calculations, Condensing 5~10
and Subcooling Sections
5-6 Variation of Heat Rejected with Airflow Length for 5-11
Condensing Section of TECO Condenser
5-7 Aerojet Condenser (190390) Structural Schematic 5-20
5-8 TECO Condenser (190370 Structural Schematic 5-22
VI I I
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ILLUSTRATIONS (Continued)
Figure No. Page
5-9 SES Steam Engine Condenser (190640) Structural 5-23
Schemati c
5-10 Aerojet Condenser with Truss 5-26
5-11 Aerojet Support Frame for Fans 5-2?
5-12 Air-Side Perforated Fin (-13) 5-48
5-13 Vapor-Side Offset Fin (TECO and SES) 5-48
5-14 Vapor-Side Perforated Plain Fin (Aerojet) 5-48
5-15 Condenser Core Module 5-49
5-16 Condenser Core Module Detail Component Arrangement 5-50
5-17 Failed Module 5-52
5-18 Aerojet Condenser, Partially Assembled 5-53
5-19 Aerojet Condenser Assembly--Front View 5-55
5-20 Aerojet Condenser, Back View 5~56
5-21 Aerojet Condenser--Top View 5-57
5-22 Aerojet Condenser—Fan Support Frame 5~58
5-23 Thermo Electron Condenser—Front View 5-59
5-24 Thermo Electron Condenser—Back View 5~60
5-25 Thermo Electron Condenser—Three-Piece Fan 5-61
Support Frame
5-26 Core Module Structural Test 5-62
5-27 Steam Engine Systems Condenser—Front View 5-63
5-28 Steam Engine Systems Condenser—Back View 5-64
5-29 Steam Engine Systems Condenser—Three-Piece Fan 5-65
Support Frame
5-30 Heat Transfer Performance Test Module 5-67
5-31 Heat Transfer Performance Test Module 5-68
IX
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ILLUSTRATIONS (Continued)
Figure No. Page
6-1 Fan Velocity Triangles 6-4
6-2 Effective Stress (KSl) on Plate at 4000 RPM 6-6
6-3 Effective Stress (KSl) on Disk at 4000 RPM 6-7
6-4 Blade Vibration Interference Diagram 6-8
6-5 Aerojet Fan Assembly 6-17
6-6 Aerojet Fan Assembly 6-18
6-7 Thermo Electron Fan Assembly 6-19
6-8 Steam Engine System Fan Assembly 6-20
6-9 Hydraulic Motor Calibration Test Setup 6-21
6-10 Motor Calibration--HPI PN M20-90155-01, SN 20047 6-22
6-11 Motor Calibration—Vickers Motor, Model No. 3911-30 6-23
6-12 Fan Aerodynamic Performance Test Setup 6-25
6-13 Aerojet Fan Calibration Test Setup 6-26
6-14 Performance of the Aerojet Hydraulic-Motor-Driven 6-28
Fan, AiResearch PN 605972-1-1
6-15 Performance of the TECO Shaft-Driven Fan, 6-29
AiResearch PN 605977-1-1
6-16 Performance of the SES Belt-Driven Fan, 6-30
AiResearch PN 605982-1-1
6-17 Fan Noise Test Setup 6-31
6-18 Sound Pressure Levels for Aerojet Fan Assembly 6-33
P/N 605972-1
6-19 Reference Background Sound Pressure Level for 6-41
Aerojet Fan Assembly P/N 605972-1
6-20 Hydraulic Motor Sound Pressure Level Test Setup 6-43
6-21 Sound Pressure Level for HP I Hydraulic Motor 6-45
M20-90155-01, S/N G20047
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ILLUSTRATIONS (Continued)
Figure No. Pacle
6-22 TECO Fan Assembly Installed in Noise Test Setup 6-51
6-23 Sound Pressure Level for Thermo Electron Fan 6-52
Assembly P/N 605977-1
y-1 Condenser and Fan Airflow Test Setup 7~2
7-2 TECO Assembly Installed in Test Rig 7~3
7-3 SES Assembly Installed in Test Rig 1~^
J-k Duct Configuration at Condenser Inlet 7~5
7-5 Duct Configuration at Condenser Inlet 7~6
7-6 Hydraulic Motor Fan Drive 7~7
7-7 Hydraulic Motor Fan Drive 7~8
7-8 Pressure Tap Locations 7~9
7-9 SES Condenser and Fan Airflow Test Results 7-10
7-10 TECO Condenser and Fan Airflow Test Results 7~11
8-1 Thermo Electron Engine Mockup 8-3
8-2 Thermo Electron Engine Mockup B-k
8-3 Thermo Electron Engine Mockup 8-5
8-^t Adaptor Plate with Simulated Bumper 8-6
8-5 Air Flow Ducting 8-7
8-6 Thermo Electron Installation Air Flow Test Setup 8-8
8-7 Thermo Electron Installation Air Flow Test Results 8-9
8-8 Aerojet Installation Airflow Test Setup 8-11
8-9 Aerojet Installation Air Flow Test Setup 8-12
8-10 Aerojet Installation Airflow Test Results 8-1^
XI
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TABLES
Number Page
3-1 Fin Perforation Geometries 3-1
3-2 Design Conditions for Fin Optimization Study 3~3
5-1 Condenser Design Requirements 5-2
5-2 TECO Condenser Design Summary 5-12
5-3 Aerojet Condenser Design Summary 5-14
5-4 SES Condenser Design Summary 5-15
5-5 Summary of Pressure Stresses in the 190390 Condenser 5-17
for Aerojet
5-6 Summary of Pressure Stresses in the TECO Condenser 5-18
5-7 Summary of Pressure Stresses in the SES Condenser 5-19
5-8 Condenser Weights 5-24
5-9 Summary of Structural Dynamic Stress in the 190390 5-28
Aerojet Condenser
6-1 Fan Design Requirements 6-2
6-2 Fan Blade Parameters 6-5
6-3 Fan Des ign Summary 6-9
6-4 Noise Test Results 6-44
8-1 Aerojet Installation Design Point Comparison 8-16
x i i
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NOMENCLATURE
A Heat transfer area, sq ft
A Cross-sectional flow area, sq ft
\j
D Tube or duct diameter, ft
f Fanning friction factor, dimensionless
2
F Wall shear stress multiplied by g , Ib /ft hr
o ' 3c m
F Tensile stress at material yield point, psi
g Conversion factor, k.\15 x 10 Ib ft/lb, hr2
c m t
G Mass velocity = W/A , Ib /hr ft2
' cm
h Heat transfer coefficient, Btu/hr ft F
k Thermal conductivity, Btu/ft hr F
L Total condensing length between qualities x = 1 and x = 0, ft
M Torque, i n.-Ib
N Rotational speed, rpm
N Design Point operating speed, rpm
P , Ambient pressure, in. H00
amb K ' 2
Pr Prandtl number, dimensionless
P Static pressure, in. H_0
PT Total pressure, in. H?0
P Dynamic pressure, in. H..O
Q Heat transfer rate, But/hr; Volumetric flow, cfm
Rey Reynolds number based on total flow rate and local vapor density
dimensi on 1 ess
Circumference, ft
XI I
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NOMENCLATURE (Continued)
T Absolute temperature, R
u Fluid velocity, ft/hr
U Overall coefficient of heat transfer, Btu/hr ft °F
DA Thermal conductance, product of U x A, Btu/hr F
W Fluid flow rate, Ib /hr
m
x Fluid quality, dimension less
y Total liquid layer plus oil film thickness, ft
z Location along condensing length, ft
01 Void fraction, dimensionless
5 Liquid layer thickness, ft
APs Static pressure rise, in. rLO
APT Total pressure rise, in. HO
T|r Fan total efficiency, percent
(j, Vi scosi ty, Ib /ft hr
p Fluid density, Ib /cu ft
o~ Ratio of air density to standard air density
6 Angle that vapor flow vector makes with the horizontal, deg
XI V
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Subscr ipts
a Acceleration field
av Average
d Discharge
eq Equivalent
f Fr iction
i Inlet
SL L iqu id
m Momentum
0 Evaluated at the wall
T Total
v Vapor
z Evaluated at location z along the condensing length
xv
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SECTION 1
INTRODUCTION AND ABSTRACT
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SECTION 1
INTRODUCTION AND ABSTRACT
INTRODUCTION
Condenser size and cooling air fan power consumption are two significant
factors that limit the application of-Rankine cycle power systems to auto-
mobiles. The Compact Condenser program recently completed by AiResearch
(Reference 1) resulted in a substantial reduction in condenser size compared
to that obtained using present day radiator technology. 'Reduction in fan
power consumption is a definite possibility because a substantial increase in
fan efficiency can be realized by using a sophisticated aerodynamic design.
The primary goal of this program is, therefore, to develop a high
performance condenser and fan system for each of three Rankine cycle auto-
motive power plants now under development by the Environmental Protection
Agency. A condenser and fan package is to be delivered to each of three system
contractors--Aerojet Liquid Rocket Company, Thermo Electron Corporation, and
Steam Engine Systems--at the completion of the program.
This development program is supported by other tasks including: (1) the
further optimization of the air-side perforated fin, (2) a measurement of the
condensing heat transfer performance, and (3) a determination of the air-side
pressure losses in a typical automobile installation.
ABSTRACT
The three condenser and fan assemblies were fabricated, leak checked,
performance tested, and delivered to the three system contractors. Final
specifications for each system are presented in Appendix A. The completed
Aerojet assembly is shown on Figure 1-1, a front view, and Figure 1-2, a
back view. A T-shaped design is used for the Aerojet unit to provide clear-
ance between the condenser and the automobile frame rails. The assembly is
designed to fit within the engine compartment of a 1972 Chevrolet Impala. A
top view of the condenser assembly show ing the flared -vapor inlet duct is
shown on Figure 1-3.
Photographs of the completed Thermo Electron condenser and fan assembly
are shown on Figure 1-4, a front view, Figure 1-5, a top view, and Figure 1-6,
a view of the back face. The assembly is designed to fit into the engine com-
partment of a 1972 Ford Galaxie.
The completed condenser and fan assembly built for Steam Engine Systems
is shown on Figures 1-7 and 1-8. The assembly is designed to fit within a
Plymouth "C" body engine' compartment.
The optimization of the perforated air-side fin which was developed in
the Compact Condenser program (Reference 1) was continued. Two additional sur-
face geometries were tested. Additional optimization studies were conducted
to establish the optimum fin height and fins per inch for a given perforation
geometry.
1-1
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72207-2
Figure 1-1. Aerojet Condenser and Fan Assembly (Front View)
' , 72Z07-6
Figure 1-2. Aerojet Condenser and Fan Assembly (Back View)
72207-3
Figure 1-3- Aerojet Condenser Assembly (Top View)
1-2
-------
72375-'*
Figure 1-4. Thermo Electron Condenser and Fan Assembly
(Front View)
72375-1
Figure 1-5- Thermo Electron Condenser and Fan Assembly
(Top View)
1-3
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IS*
-.', ':.-:Sl^
S^teiftiS£?S^f:;^^.-C^
'•MS^-diR^yi• •- • Vlrf;^i -4^"-^ ^ •u K;; '-• '-?.s ^:^^'t'' v ,, „-¥ *-^ H ,; "'^^) \ ,
^ -": .. ": "'%- "iil*S-^^p^l^^^^*?^?^fV:^ : "'
• • • '- •' ' vl - •- ^- v:.'v: • :v *•'. -
-,•'.-• , •^^:"^^NsVM^.-^t-'--^-;.-v.-t;n - •> 1"
.', :->":-.'"Sg;g^iailii;??g.t.-i''^:j:,;fe;j''' >^'-' - •'•;,?•.'•-• , -,t m—ii.,_-^»^-=
Figure ^-6. Thermo Electron Condenser and Fan Assembly
(Back View)
1-if
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72*423-2
Figure ^-7• Steam Engine Systems Condenser and Fan Assembly
(Front View)
72^23-1
F-17^31
Figure 1-8. Steam Engine Systems Condenser and Fan Assembly
(Back View)
1-5
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Condensing heat transfer tests were performed using steam and Fluor-
inol-85, the TECO organic working fluid, condensing in three different vapor
passage fin configurations. These tests were run primarily to establish the
effect of a nonuniform condensation rate on the performance of a cross flow
condenser.
Three different condensers were designed to meet the specific performance
requirements of the three Rankine cycle engines. Each unit was configured to
fit within the confines of the specified engine compartment. Detailed structual
analyses were conducted to ensure that the units would withstand the expected
pressure and temperature operating conditions. Analyses were also conducted
to establish the required structure and mounting arrangement that would be
necessary in a readable design. An identical optimized perforated fin was
used on the air-side of each condenser. The condenser core assembly was fabri-
cated using a new fluxless brazing process to obtain a high level of cleanliness
and structural reliability. All assemblies were treated with an epoxy ester
coating to insure meeting a rigid vapor-side helium mass spectrometer leakage
requi rement.
All fans were designed to use a common high performance impeller design.
The tip diameter and rotational speed were varied to meet the demands of three
different installations. Aerodynamic performance and noise tests were per-
formed on the completed fan assembl ies.
Air flow tests were conducted on the Thermo Electron and Steam Engine
Systems assemblies. The condensers were assembled with the fans and system
airflow capability as a function of fan rotational speed was determined.
Qualitative checks were made with regard to the system noise and vibration
levels.
Cooling air flow through a Rankine cycle engine is about 3 to 5 times
greater than that of a conventional 1C automobile engine radiator. Thus, it
was expected that the air-side pressure losses could become critical in a
Rankine cycle engine installation. Since no test data was available, mockups
of two of the Rankine cycle engines were installed in actual engine compart-
ments and the pressure drop across the bumper and grille and through the engine
compartment was experimentally determined.
1-6
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SECTION 2
CONCLUSIONS AND RECOMMENDATIONS
-------
SECTION 2
CONCLUSIONS AND RECOMMENDATIONS
CONCLUSIONS
Condenser and fan performance should meet all specification performance
goals. The assemblies have been configured to fit within their respective
engine compartments. The assemblies are not considered to be a fully readable
design, however, because additional structure and/or supports are required to
withstand the expected vibration and shock environment. The installation could
be modified at a later date to provide the required support.
Air flow tests performed on the condenser and fan assemblies revealed that
the airflow through the Steam Engine Systems unit was slightly higher than pre-
dicted whereas the airflow through the Thermo Electron unit was nine percent
lower than predicted over the fan speed range. The low performance in the
TECO unit is attributed to the close spacing between the condenser and fans
which was dictated by the engine compartment space limitations. The TECO design
point airflow can be achieved, however, by merely increasing that fan speed
and accepting a small increase in fan power consumption. Alternatively, the
spacing between the condenser and fans could be increased.
The air-side perforated fin geometry appears to be close to optimum for
this application. The selected fin height (0.326 in.) and fins per inch (22)
result in a minimum in fan power consumption. The -13 rectangular slot per-
foration geometry as described in Section 3 was found to provide superior
performance and should be used in the final condenser designs.
Vapor-side heat transfer performance should be as predicted based on the
condensing heat transfer test results. The best vapor-side fin is the offset
configuration because it permits redistribution of vapor flow as required to
match the variable condensation rate in a crossflow condenser. A variable fin
geometry designed to match the condensation rate could be used to obtain max-
imum performance but the small performance gain does not warrant the increased
fabrication complexity. If vapor-side pressure drop is limited, then a perfo-
rated plain rectangular fin should be considered.
Fluxless brazing has been established as a practical fabrication method
for the condenser plate-fin core assembly. A hydrostatic pressure test per-
formed on a condenser core module reached a burst pressure of about 5 times the
maximum operating pressure. The unit structural design is believed to be
adequate. Braze and weld assembly of helium leaktight vapor passages will
always present problems, especially with regard to detection and repair of
individual leaks. It is believed that an epoxy ester coating as used on the
present assemblies can be effectively used as a production technique to seal
minute leaks and to provide a protective coating on the air-side fins.
2-1
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The cooling air fans have met all performance requirements including air
efficiency goal of greater than 70 percent. The measured noise levels were
exceptionally low. It is expected that, if the fans are the major noise
source on the engine, the EPA vehicle noise specifications can be met. The fan
design is feasible for mass production and its cost can be minimized by injec-
tion molding of the fan impeller and by using a glass-filled nylon material.
Airflow pressure losses measured in the engine installation mockup were
less than expected and, in fact, came reasonably close to meeting the specifi-
cation allowance. These low losses were achieved by removing the grille from
the 1971 Ford Galaxie and other vehicle modifications. The engine mockups
were not complete and the addition of all equipment including wiring, instru-
mentation, and ancillary equipment would, no doubt, increase the pressure loss.
However, by providing additional air flow exit area, e.g., louvers in the fender
wells, the increase in flow restriction could be counteracted.
The Ford Galaxie had a high pressure drop and it must be modified to
provide additional airflow area. It was found that small increases in airflow
area such as obtained by removing 2 of the k vehicle headlights produce a
significant reduction in pressure drop. The effect of undercar velocity as
used to simulate vehicle motion on the compartment pressure drop was found to
be mi nor--compartment pressure drop increased by only 15 percent when the
undercar flow velocity increased from 0 to 60 mph. Qualitative observations
indicated that the existing bumper design causes a significant flow maldistri-
bution at the condenser inlet face and this effect will degrade the condenser
performance.
RECOMMENDATIONS
Before road testing, the vehicle frame should be modified to provide a
center support for the condenser and fan assembly. As an alternate and pos-
sibly better solution, the fans could be mounted to the vehicle frame with a
flexible duct between the condenser and fans.
A specific test should be run to evaluate the full-size condenser heat
transfer and pressure drop performance. A 1A section of the condenser could
be used for this purpose. This data would be of value in determining the
condenser behavior after it is installed in the vehicle.
Flow blockage at the fan discharge should be minimized. Engine equipment
should be relocated to avoid blockage insofar as possible because the velocity
maldistribution caused by blockage will significantly affect fan performance.
Because of the limited space within the engine compartment, an axial flow
fan will always be subjected to flow blockage. It does appear that a radial
flow or centrifugal machine would fit the installation better. Less axial
length is required and a diffuser section can be incorporated to recover a
portion of the swirl energy. The fan can be designed independent of the rest
of the engine because the outlet diffuser directs the flow down and away from
the engine. The installed efficiency would be higher than that of an axial
machine because there would be no flow blockage effects. A centrifugal machine
2-2
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turns slower than any other type of fan, and hence, it is more quiet. It is
recommended that the system contractors review the results of their initial
tests and determine the effectiveness of the installed axial flow fans. If
performance is lacking then a radial flow machine should be considered.
The existing bumpers and grille should be modified to reduce flow blockage
at the condenser inlet.
2-3
-------
SECTION 3
PERFORATED FIN DEVELOPMENT
-------
SECTION 3
PERFORATED FIN DEVELOPMENT
SURFACES
During the previous program (Reference 1) perforated fins with 0.015-,
0.030-, and 0.060-in. wide slots and open area ratios of 12.5 and 25.0 percent
were tested, and the best overall performance was obtained for the 0.030-in.
slots at 25.0 percent open area. This particular core was designated as the
-13 configuration; details of the perforation geometry are shown in Table 3-1.
On the basis of this work, it was decided to test the same 0.030-in. slot
width at open area ratios of 16.7 and 33-3 percent during the current program.
The 22 fins/in., 0.326-in. fin height and Q.OOk-ln. fin thickness of the pre-
vious cores were maintained.
TABLE 3-1
FIN PERFORATION GEOMETRIES
Core
-13
-21
-23
Slot
Width,
i n.
0.030
0.030
0.030
Slot
Length,
i n.
0.250
0.250
0.250
Slot
Spaci ng,
i n.
0.120
0.180
0.090
Theoret ical
Open Area,
percent
25.0
16.7
33.3
Actual
Open Area,
percent
22.0
16.0
32.0
The theoretical percent open area values are for the perforations in an infi-
nite heat exchanger where edge effects are negligible. The actual percent
open area values are for the perforations in the test cores where the perfora-
tions had to be in a total airflow length of 3-38 in. and the air fin actually
consisted of two separate pieces, each 1.69-in. long. Figure 3-1 shows the
perforation layout for the -21 and -23 fins.
It was originally planned to select a third perforation geometry on the
basis of a study of heat transfer from perforated plates at the University of
Michigan (Contract No. 68-04-0019). Based on a meeting held in December of
1971 with Dr, Wen-Jei Yang, who conducted the study, it was concluded that
one of the two geometries already selected represented very nearly the optimum
geometry (fin spacing and percent open area) indicated by the results of the
University of Michigan study.
Therefore, it was decided to conduct a heat exchanger optimization study
to obtain the fin geometry (fin height and fins/in.) yielding minimum fan air
horsepower for a typical condenser problem statement and set of envelope
limitations. The purpose of this study was to determine whether it would be
profitable to test one of the previously tested perforation geometries in a
configuration involving either a new fin height or a different number of
fins/in. The design conditions used for this analysis were essentially those
of the TECO problem statement as listed in Table 3-2. The -13 fin perforation
geometry was used, since this was found to be optimum in the Reference 1 study.
3-1
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(18 SLOTS)
(18 SLOTS)
0.090
|« 0.
"T
0.250
TANGENT LINE
-0.030
053
(a) -23 FIN SHEET
(9 SLOTS)
.69-
u
0.180
A
-TANGENT LINE
0.250
-0.030
053
«
AIR
AIR
FLOW
(b) -21 FIN SHEET
NOTE: ALL DIMENSIONS ARE IN INCHES
Figure 3-1- Perforated Fin Materials For Cores-21 and-23
S-6809I
3-2
-------
TABLE 3-2
DESIGN CONDITIONS FOR FIN OPTIMIZATION STUDY
Working fluid
Total heat rejection
Condensing temperature
Air inlet temperature
Working fluid flow rate
Condenser core frontal area
Fluorinol 85
1.88 x 106 Btu/hr
213°F
85°F
9860 lb/hr
8.46 sq ft
The fin configuration in the condensing side passage was a 0.050 in. high
rectangular offset surface with twenty 0.004-in. thick aluminum fins/in. An
average condensing heat transfer coefficient of 640 Btu/hr sq ft °F was used
and the analysis was based upon a "no DA margin" design. The air side effec-
air side fin height and number of fins/in, were varied. The
from the -13 perforated fin were used for all the air
At each air side effectiveness, air side fin height, and
the condenser flow length and air AP were calculated. The
tiveness and the
experimental f and j
side fins analyzed.
number of fins/in.,
pressure drop associated with getting the air out of the engine compartment,
AP.
installat ion'
was calculated at each value of air side effectiveness from
AP.
installation
= 2.80 x
i n.
where W is the airflow rate through the condenser in Ib/min. The TECO design
point is at a vehicle speed of 90 mph which gives a ram pressure rise (assuming
100 percent recovery) of 3-73 in. H«0. The required fan pressure rise was
calculated as the condenser air side AP plus AP. . ,, .. minus the ram AP.
r installation
The fan air horsepower could then be calculated from the fan pressure rise and
the air volumetric flow rate. The results of the analysis were plotted as fan
air horsepower vs condenser flow length. A typical plot is shown in Figure 3-2
for a fin height of 0.326 in. It is seen that with an increasing air-side
effectiveness (decreasing airflow) the power requirement decreases at the
expense of increased condenser airflow length. At a particular condenser
length, the power requirement decreases with increasing number of fins/in, over
the range analyzed. Similar data were generated for fin heights of 0.250 in.
and 0.400 in. and cross-plotted to obtain the required fan air horsepower as a
function of fin height, as shown in Figure 3-3- The comparison in Figure 3-3
is made at a flow length of 4.25 in. as this was the design objective for the
TECO condenser. It is seen that for 22 or 26 fins/in., there is no horsepower
3-3
-------
26 FINS PER INCH !!
FIN HEIGHT = 0.326
4
6.
3.8
4.0
4.2 4.4 4.6
AIR FLOW LENGTH, INCHES
4.8
.0 5.2
S-68453
Figure 3-2. Effect of Air Side Fin Geometry on Condenser Length and
Required Fan Air Horsepower
-------
o
Q_
LU
CO
01
O
01
16
12
10
!26 FINS PER INCH
; ]
I
art
Iji; AIR FLOW LENGTH = 4.25 INCHES
}
it
lliil
iir
-1
0.20 0.24
0.28 0.32 0.36
FIN HEIGHT, INCH
0.40
0.44
S-68452
Figure 3-3- Effect of Fin Height on Fan Air Horsepower at a
Fixed Condenser Size
3-5
-------
incentive to go to fin heights other than the baseline 0-326 in. At 18 fins/in.
slightly less power is required at a fin height of 0.250 In. than at 0.326 in.
As noted above, the required power decreases with increasing number of fins/in.,
but because of the potential problems of contamination and plugging of the con-
denser fins, it was decided that the 22 fins/in, baseline should not be ex-
ceeded. On the basis of the 22 fins/in, limit and the horsepower-fin height
relationship of Figure 3~3, it was concluded that the baseline fin (22R-.326-
PERF(-13)--004 Al) was still optimum. Based on these results, it was decided
not to test a third fin.
TEST CORES
As a result of the considerations discussed above, two cores were tested.
The cores were tested without and with turbulence upstream of the core.
A typical perforated fin aluminum test core is shown in Figure 3-^- The
flow configuration is single-pass crossflow. The core is constructed from
ten air-side sandwiches using the selected perforated rectangular fin surface
and eleven water-side sandwiches using a 20 fin/in, offset rectangular fin
surface.
As in the case of the eleven previous cores, the test data were reduced
on the basis of a solid fin both in heat transfer area and in fin efficiency.
In addition, the small variations between cores were neglected and the same
dimensions were used for all the cores. As a result, the curves can be com-
pared directly. That is, at the same Reynolds number, the airflow in each
core would be the same and the heat transfer and friction factors can be used
to compare the heat transfer conductances and the pressure drops, respectively.
It would be impossible to make such direct comparisons if the effect of the
perforations on the hydraulic diameter and the heat transfer area was included
in the curves. Since the principal purpose of this study is to obtain the
best air side surface, this procedure is indicated.
TEST PROCEDURE
Figure 3-5 is a schematic of the test setup. The air source was a blower
instead of the pressurized air supply used in the previous program.
The air weight flow rate was determined upstream of the core by a square-
edged orifice plate in a ^-in. diameter pipe flow measuring section; both met
the requirements of the 1959 ASME Power Test Code for this type of flowmeter.
Four orifice diameters were used: 3-0, 2.5, 2.0 and 1.5 in. The inlet air
temperature was measured by two thermocouples, and the outlet air temperature
was measured by four thermocouples after the air had been mixed in an insulated
mixing device.
On the water side, the water was pumped from an open sump through a heat
exchanger where it was heated by steam. The water then passed through a fil-
ter and a mixing device to the unit. After flowing through the core, the water
passed through another mixing device and through a weighing system back to
the sump. The weight flow was determined by the weight-time method where the
flow time for a specific weight of water was determined automatically to
0.01 sec.
3-6
-------
** 69494-2
Figure 3-k. Typical Test Core for f and j Factors
(22 Fins per Inch, Perforated, O.OOA-in,
Aluminum Material)
3-7
-------
oo
PUMP
TEMP
CONTROL
STEAM
OUT
SUMP
STEAM
IN
HX
MIXING "T"
FILTER
INLET THERMOCOUPLE
\
BOX
TO
AMBIENT J, \
OUTLET THERMOCOUPLES
SCALE
MIXING "T1
t
UNIT
VALVE
OUTLET THERMOCOUPLE
INLET
T.C.
T
BLOWER
AIRFLOW
A IN.
MEASURING
SECTION
S-69006
Figure 3-5. Typical Heat Exchanger Test Setup Schematic
-------
Thermocouples were of special grade copper-constantan, which has an allow-
able tolerance of ±0.75°F. Temperatures were read in millivolts by a potentiom-
eter. Water temperature drop was measured by a differential thermocouple in
most runs.
Air inlet static pressure and static pressure drops were measured by water
manometers; inclined manometers were used to read all but the highest pressure
drops to an accuracy of 0.01 in. of water. Wall pressure taps were located
3.0 in. upstream and 5-0 in. downstream of the faces of the test cores. Two
taps located on opposite sides of the duct were used at each location and were
connected to a common manometer to average the pressures. Water-side pressure
drops were measured in inches of water by differential pressure gauges and were
used in the data reduction to calculate the frictional heating of the water.
Accuracy of the data was indicated by the heat balances achieved. For all
airflows except one, at least one run was obtained with a heat balance of not
over 2.7 percent. Average heat balance for all runs was approximately 1.6
percent.
The test setup did not have the traversing pressure probes upstream of the
core or the thermocouple grid downstream of the core as in the previous tests.
In addition, the orifice pipe measuring section was changed. Photographs of
the previous test setup are presented in Figures 3-6 and 3~7-
In the tests to determine the effect of turbulence, turbulator rods were
installed upstream of the test section to determine the effect of turbulence
on fin performance. Three 1/2-in.-diameter rods were placed k in. upstream of
the face of the core as shown in Figure 3-8. The size and spacing of the
turbulator rods were selected on the basis that the wakes of the air stream
would be intersecting upstream of the test unit, resulting in a minimum varia-
tion in air velocity at the face of the core.
TEST RESULTS
Figures 3-9 and 3-10 present the heat transfer factors (j) and friction
factors (f) for the two perforated surfaces. As previously, the surface
identification, 22R--326-PERF( )-.004(Al), is based on the AiResearch system
for plate-fin surfaces where the 22 refers to the fins/in., the R is for
rectangular fin, the .326 is the plate spacing in inches, PERF refers to a
perforated fin, .004 is the fin thickness, and Al is the fin material. In
addition, for the purposes of this work, the perforated core identification
is placed in the parentheses after PERF. Fluid properties including the den-
sity were evaluated at the bulk average temperature. The heat transfer area,
hydraulic radius, and fin efficiency are based on those of a solid fin of the
same geometry. The hydraulic diameter (D.) is 0.073^5 in., the area density
(3) is 589 sq ft/cu ft and the fin area-to-total area ratio (A /A ) is 0.886.
With upstream turbulence, the two cores had lower friction factors and
equal or higher heat transfer factors than the original or no-upstream-
turbulence runs, except for the Reynolds number of 1200 in the -21 core where
the test point was essentially the same for the two cases. The -21 core curves
in the 800 to 1200 Reynolds number range are not well defined in this transition
range due to the lack of sufficient data points.
3-9
-------
Figure 3-6. Test Setup Showing Water Flowrate Measuring
System and Thermocouple Measuring Apparatus
3-10
-------
Figure 3-7- Test Setup Showing Test Core Mounted in Ducting
3-11
-------
3.30 IN.
AIR
FLOW
1.98
IN.
1/2 IN.
DIA
3.96 IN.
TEST
UNIT
r"
L— DUCT WALL
•4 IN.
S-64774
Figure 3-8. Turbulator Rods Design
3-12
-------
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o
o
<
o
i—i
CC
UJ
u_
(/)
10
08
-06
,05
.04
.03
.02
.010
.008
.006
.005
.004
.003
,002
O TURBULENCE TEST POINTS
I. SURFACE 22R- .326 - PERF (-21)- .004 (A.? )
2. FLUID PROPERTIES EVALUATED AT BULK AVG TEMP
3. AREAS AND FIN EFFICIENCY BASED ON FIN AS A SOLID FIN
,001
200
300 400 500 600 800 1000
2000 3000 4000 5000
8000
REYNOLDS NUMBER
Figure 3-9- Performance Parameters for the -21 Perforated Fin
3-13
-------
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" 3 AREAS
HWiW^
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PROPERTIES EVALUATED
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200 300 400 500 600 800 000
ORIGINAL TEST PO
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FIN
2000 3000 4000 5000 8C
REYNOLDS NUMBER
Figure 3-10. Performance Parameters for the -23 Perforated Fin
-------
3 FROM PREVIOUS STUDIES
0.98
.02
1.04
1.06
1.08
I . 10
AIR SIDE TOTAL FRONTAL AREA
AIR SIDE TOTAL FRONTAL AREA GOAL
S-68991-A
Figure 3-11. Condenser Solutions for Tested Fins
-------
The f and j data were used to size condensers at the previous parametric
study (Reference I) design point heat rejection of 1.5 x IO6 Btu/hr. The non-
dimensional ized core sizes are shown in Figure 3~11- The cores were non-
dimensional ized by dividing the air total frontal area and the air side volume
by the area and volume design goals, respectively. The variation of volume
with frontal area for each fin geometry is obtained by varying airflow and
air pressure drop in such a way as to maintain a constant air power require-
ment of 9^0 hp.
The selection of the best fin involves a tradeoff between air frontal
area and air volume because a larger volume means a longer airflow length,
which is not desirable due to the restricted space available for the fan and
condenser. As shown in Figure 3-11, there are many combinations of air fron-
tal area and volume which will yield condensers for the given heat load and
air horsepower requirement. For designs that lie near the minimum frontal
area solution, it is possible to achieve significant decreases in volume with
small increases in air frontal area by operating at somewhat higher airflows.
Introduction of upstream turbulence improved the fin performance such
that slightly smaller condensers are obtained. However, there is some
uncertainty in the turbulent data results in that (1) the reason for the
decreased f and higher j with turbulence is not fully understood, (2) it is
possible that the lower f was partly due to the pressure tap being located
in the wake of the middle turbulence generating rod so that a lower than actual
upstream pressure was measured, and (3) the changes in f and j in the range of
interest for this program are estimated to be in the order of the experimental
uncertainty of the data. Therefore, for conservatism, the no-turbulence data
were used in the design of the condensers for this program.
On the basis of the no-turbulence data, the -13 core is still the best
core for this application. The -21 core does not merit any consideration.
The -23 core has a smaller minimum frontal area than the -13 core but the
difference is less than 0.5 percent. For frontal areas above the minimum,
the -13 core has a significantly smaller volume (and thus shorter airflow
length) than the -23 core. The smaller volume characteristic of the -13 core
is definitely preferred for this application due to the critical envelope
restriction on overall fan/condenser airflow length.
3-16
-------
SECTION 4
CONDENSING HEAT TRANSFER TESTS
-------
SECTION 4
CONDENSING HEAT TRANSFER TESTS
TEST OBJECTIVES
As the result of a condensing heat transfer analysis, it was revealed
that a large degree of uncertainty existed in the prediction of overall con-
denser performance due to a variation in condensation rate with axial position
within the heat exchanger. This variation occurs because the temperature
differential between condensing fluid and air decreases by a factor of five
(corresponding to an air-side effectiveness of 0.80) between the air inlet
face and the air outlet face. As a result, there is a tendency for the outlet
vapor flow to be subcooled in the passages nearer the air inlet face and par-
tially uncondensed in the downstream passages. This situation, which leads
to a lower overall heat transfer rate for a given vapor inlet temperature, may
be partially offset by vapor flow redistribution, vapor crossflow (if the vapor
fins are offset or perforated), axial conduction of heat in the airflow direc-
tion, and condensation of vapor on liquid in the outlet manifold. Because of
the complexity involved in trying to account for each of these effects analyt-
ically, the condensing heat transfer tests were oriented toward establishing
the magnitude of the loss in overall condenser performance as a result of
variable condensation. In addition, tests were performed to determine the
improvement in condensing performance obtained by using a variable flow resis-
tance on the vapor side.
TEST CORE DESCRIPTIONS
Three test cores were tested. Each core incorporated three vapor-side
passages alternating with four air-side passages. The interior air-side
passages were of double-sandwich configuration, whereas the passages forming
the core sides were single-sandwich and thus contained half the heat transfer
area of the interior passages. This arrangement provides essentially uniform
cooling of each of the vapor passages. The air-side fin sandwiches were 1 6R-
0.153-0.143(0)-0.004(A1) (16 rectangular fins/in., 0.153 in. high, 0.143 in.
offset length, 0.004 in. fin thickness, aluminum).
Each of the three cores incorporated a different vapor-side fin
configuration. A plain rectangular fin (20R-0.050-Plain-0.040) was used in
the first core, an offset rectangular fin (20R-0.050-0.100(0)-0.004) was used
in the second core, and the third core incorporated a variable-fin geometry
designed to the SES problem statement. Each core was 18.0 in. long in the
vapor flow direction and 4.25 in. long in the airflow direction. A picture
of one of the completed test core assemblies is shown on Figure 4-1. Large,
cylindrical manifolds are incorporated at each end of the assembly to provide
uniform air flow distribution across the face of the test core.
The design of the vapor passages in the variable fin geometry core is
shown in Figure 4-2. In this design, which is based on the SES problem state-
ment, the steam-side fin geometry is varied in such a manner as to obtain a
steam flow distribution that matches the steam-to-air temperature differential.
4-1
-------
Figure 4-1. Test Core Assembly Condensing Heat Transfer Test
k-2
-------
AIR
C3
LU
CO
SATURATED STEAM
I I I
CJ
LU
CO
LU
CO
CO
LU
CO
FLOW PASSAGE
DIVIDER STRIP
r
HEAT EXCHANGER
SIDE STRIP
AIR
I I I I
SATURATED WATER
Segment
1
2
3
4
Fins/In.
10
12
16
28
Fin Offset
Length ( in. )
None
0.500
0.125
0. 100
Fin Height
(in.)
0.050
0.050
0.050
0.050
Fin Thickness
(in.)
0.004
0.004
0.004
0.004
Figure k-2. Steam-Side Fins in Variable Geometry Test Core
4-3
-------
The heat exchanger is divided into four parallel passages of equal width, and
a different fin geometry is selected for each passage. The variation in flow
resistance obtained by varying fins per inch and fin offset length is suffi-
cient to obtain a steam flow rate ratio of approximately three to one from
the first to the last passage. Figure 4-3 shows the variation in pressure
loss with flow rate for each of the parallel passages. The design point flow
in each passage is obtained from the requirements that (1) the pressure drops
in all four passages must be equal, and (2) the total flow rate must equal
the design flow of 21.33 lb/min. These two requirements define the dashed
line labelled "Equi1 ibrium AP" in Figure 4-3, and the intersections between
dashed and solid lines yield the flow rates in each passage. Figure 4-4
shows both the desired steam flow distribution, which is proportional to the
steam-to-air temperature differential, and the calculated actual distribution
obtained with the variable fin geometry. The actual is very close to the
desired distribution and should yield a nearly uniform fluid state at the
steam-side outlet.
Two series of tests were conducted, the first in an open-system test
setup utilizing steam as the vapor-side fluid, and the second in a closed-
system test setup utilizing Fluorinol 85 as the condensing fluid. The open-
and closed-system tests are discussed below.
OPEN-SYSTEM TESTS
Test Setup and Procedure
The open-system test setup is shown schematically in Figure 4-5 and a
picture of the completed setup in Figure 4-6. Steam was drawn from the plant
steam system and regulated to an inlet pressure of approximately 15 psig.
Because the plant steam is wet, a superheater was installed upstream of the
core. The superheater used hot air as the energy source and was adjusted to
provide 10 to 25 F of superheat. A calibrated sight glass was installed in
the steam line downstream of the test core, followed by a valve for back-
pressuring the system. During operation, this valve was adjusted to maintain
the vapor-liquid interface at an approximately constant level in the sight
glass. Downstream of the valve,the water was cooled and then weighed in a
weigh tank. The flow rate, as measured by the weigh tank, was corrected by
the measured change in the sight-glass liquid level between the start and
finish of each test run.
The cooling fluid was air at ambient temperature. Instrumentation was
installed to obtain test core inlet and outlet pressures and temperatures on
both the steam and air sides. Airflow rate was measured using a standard
orifice section.
Testing was conducted on each core over a range of airflows (approximately
15 to 80 lb/min) corresponding to an air-side effectiveness range of about
0.70 to 0.90. The data were reduced to obtain test UA at each set of flow
conditions, based on measured air-side inlet and outlet temperatures, the
measured inlet temperature and average pressure on the vapor side, and the
assumption of saturated liquid conditions at the vapor-side outlet. The
4-4
-------
4 6 8 10
VAPOR FLOW RATE, LB/MIN
12
14
16
S-69044
Figure k-3- Condensing Pressure Loss vs Flow Rate in Each
Condenser Segment of Variable Geometry Core
-------
1 ^
10
1 — 1
i 8
i — i
CD
0 6
1 — i
3
i — i
•z.
UJ
Q_
UJ
t—
3 2
0
\
\
\
X
X.
T«
*
'
V.
APTI
DESI
•-^c
AL FLOW DIS-
RED FLOW DIE
— ^
"RIBUTION
>TRIBUTION
"^^-— ^
I 2 3
DISTANCE FROM AIR INLET FACE, IN-
Figure k-k. Distribution of Vapor Flow Rate in Variable Geometry Core
S-69043
-------
HOT AIR HOT AIR
OUT IN
PLANT
STEAM
» fih
* ^y
PRESSURE
REGULATOR
t
1
SUPERHEATER
—©
0©
1 1
COOLING /
AIR IN "~^ /
•^
1
1
©-
=x ®—
1 j
TEST
CORE
~^L^
*— —
^=
FLOW ORIFICE =
CONTROL SECTION
VALVE
. M^
I Tnnrr* i FUFI _^
/^
A —
i— J. X,1-^ if l~ L— u ^ L_ fc»— — i
L
-i P i
\. /
SIGHT
^©
1
COOLING
~~ ^AIR OUT
GLASS '
"©
^?W
BACKPRESSURE'*-*'
COOL
rn
tK \
VALVE 1 A 1
COLD COLD \ /
WATER WATER \/FTr/
°"T i- ?™
S-69313
Figure k-$. Open-System Test Schematic
-------
71781-2
Figure 4-6. Test Setup Used for Open-System Condensing Test
4-8
-------
latter assumption, by ignoring the possibility that portions of the condenser
passages may be filled with liquid and therefore act as subcoolers, results
in the calculation of an equivalent (rather than actual) UA. The equivalent
DA defines the cooling capacity of the heat exchanger based on the assumption
that the total core is used for desuperheat and condensing.
Results and Discussion
Results of the tests, plotted as (Test UA)-*-(Predicted UA) versus air-side
effectiveness, are shown in Figure k-7. Of a total of 29 test data points
taken on the three cores, five were rejected due to poor heat balances (greater
than five percent difference between calculated heat rates in the hot and cold
streams). Of the remaining 2k test points, four exhibited a high degree of
subcooling (greater than 50°F) in the outlet liquid and are shown circled with
subcool ing as noted on Figure 4-7- The remaining 20 points were used in draw-
ing the curves of condenser performance for each of the three units.
These curves show that the variable-fin core has the best performance,
with a test UA ten percent greater than predicted at the design air-side
effectiveness of 0.80. The offset-fin core is shown to be almost as good,
with a test UA eight percent greater than predicted at design airflow; the
plain-fin core exhibits the poorest performance, with a test UA three percent
above predicted at design. All three cores show a rapid decrease in test UA
as a percent of predicted UA at air-side effectiveness above 0.8. This is
believed to be due to the increasing variation in condensation rate across
the vapor flow passage as air-side AT increases. In addition, the variable-
fin core shows a slight tendency toward reduced performance at lower air-side
effectiveness, as can be expected because this core achieves optimum vapor
flow distribution at only the design effectiveness of 0.80.
A possible explanation of the poor performance of the high subcooling
points of Figure k-1 is that there was a buildup of noncondensible gases
above the liquid-vapor interface in these runs. The presence of non-
condensibles in the core would reduce performance, and noncondensibles in the
outlet manifold and sight glass could increase the measured subcooling by pre-
venting uncondensed vapor from exiting the core and condensing on the liquid.
The buildup of noncondensibles was minimized by flushing the system with steam
prior to each test run data point.
No definite explanation can be given as to why the measured performance
was generally higher than predicted. The high conductance ratio (vapor side-r
air side) of these cores, which ranged from about 15-to-1 to 40-to-1 over the
test flow ranges, precludes the possibility that this could be due to better-
than-predicted condensing coefficients. The measured air-side pressure drop,
however, was 30 to 40 percent above predicted, which could be consistent with
an eight to ten percent increase in heat transfer performance if this were due
to irregularities in the air-side fins. To expedite fabrication of the test
cores, an available offset fin was used on the air side. A perforated air-
side fin, as is used on the preprotytype condensers, is still expected to meet
predicted performance based on previous tests.
4-9
-------
0.9C
0.70
O PLAIN FIN
• OFFSET FIN
D VARIABLE FIN
OHIGH SUBCOOLING
0.75 0.80 0.85
AIR-SIDE EFFECTIVENESS
0.90
0.95
S-69305
Figure 4—7- Open-System Performance of Condenser Test Modules
4-10
-------
Based on the open-system test results, the following conclusions were
drawn:
(a) The offset vapor fin performs as well or nearly as well as the
variable fin in the flow range of interest and should therefore
be used in the SES and TECO condenser designs.
(b) The plain vapor fin exhibits a performance penalty at design
effectiveness that is probably due to variable condensation and
resultant liquid fill-up of portions of individual passages. A
penalty of approximately seven percent in heat rejection capability
(at design flow) would be expected if this fin were used in the
Aerojet condenser. The use of an offset or perforated fin in the
Aerojet design would be a way of reducing this penalty.
CLOSED-SYSTEM TESTS
Test Setup and Procedure
The test setup for the closed-system condenser module tests with
Fluorinol 85 as the vapor-side fluid is shown in the schematic diagram in
Figure 4-8. The completed setup is depicted in Figure 4-9. The cold
Fluorinol 85 fluid was pumped through a calibrated turbine flowmeter to a
boiler that evaporated the liquid. The vapor was superheated approximately
50° to 60°F by a superheater before the dry vapor entered the test core.
Both the boiler and the superheater were heated by hot air. A sight glass
was in the Fluorinol 85 line downstream of the test core, and one was used
for the boiler. During operation, a va
-------
HOT AIR HOT AIR
OUT IN
t 4
HOT AIR
IN
HOT AIR
OUT ^_
SUPERHEATER
BOILER
r
SIGHT GLASS
©-
COOLING
AIR IN "
7
I
FLOW
CONTROL
VALVE
ORIFICE
SECTION
i
LIQUID LEVEL
h-Q
-©
TEST
CORE
RD
©
SIGHT
GLASS
FLOW
METER
BACKPRESSURE
VALVE
COOLING
'AIR OUT
COOLER
VALVE
* t
COLD COLD
WATER WATER
OUT IN
PUMP
S-69517
Figure if-8. Closed-System Test Schematic
4-12
-------
Figure k-S. Test Setup Used for Closed System Condensing Tests
4-13
-------
Results and Discussion
Figure 4-10 shows the test results as (Test UA)-r(Predicted UA) plotted
against air-side effectiveness. A total of 16 points were taken on the three
cores. One was rejected because of a very poor heat balance due to an extremely
low Fluorinol 85 flow. All of the remaining 15 points were plotted; although
two had heat balances between 5 and 6 percent, they plotted satisfactorily
compared to the rest of the data. The curves are not similar to the curves
for the open-system tests. These curves are concave instead of convex, are
lower at low effectiveness rather than high effectiveness, are higher at high
effectiveness, and the plain vapor fin results are farther above the predicted
values than the other two fins. At the design air-side effectiveness of 0.8,
the offset and variable fin results are essentially as predicted, but the plain
fin results are nearly 12 percent higher than the prediction.
The amount of subcooling of the condensate was higher than in the
open-system tests. To check this effect, the data points of Figure 4-10 were
plotted in Figure 4-11 as a function of the subcool ing. The offset fin curve
of Figure 4-11 is smoother than that of Figure 4-10, although exactly the
same values are plotted. Since the other two fins do not cover as wide a
range of subcool ing and are not as smooth, curves are not drawn through the
data for these two. Figure 4-11 shows that the spread between these points
is due to the difference in air-side effectiveness, which corresponds to the
difference in airflow.
It is not clear why the test performance was so much higher than
predicted at the high air-side effectiveness. The high air-side effectiveness
points correspond to low airflows and low Fluorinol 85 flows. At airflows
corresponding to the design point effectiveness of 0.8, the heat transfer con-
ductance ratio (vapor-side divided by air-side) is on the order of 3-to-1 for
the plain fin and 5-to-1 for the other two. In the case of the plain fin at
high effectiveness, the Fluorinol 85 flow is in the range where the Nusselt
equation for laminar condensation, as modified by McAdams, gives higher heat
transfer conductances than the Soliman correlation (Reference 2). The calculated
condensing heat transfer coefficient (based on Soliman) therefore could be
conservative in this range of air-side effectiveness for the plain fin.
Factors influencing the data points are airflow, Fluorinol 85 flow, inlet
temperature, condensing temperature, amount of subcool ing, and heat balance.
The curves shown in Figure 4-11 are at attempt to reduce the effect of the
first four parameters. This plot involves an estimate of the desuperheat por-
tion that varied from 8 to 13 percent of the core in the closed-system tests.
Control of the subcooling was attempted by regulating the liquid level in the
sight glass at the core outlet. While most of the test points had heat
balances under 5 percent, heat balance is still a significant factor at the
high air-side effectiveness because a 5 percent difference in the design point
effectiveness of 0.80 would be 0.04, which would significantly affect the
calculated test UA.
4-14
-------
<
•s
o
<
O
O
PLAIN FIN
OFFSET FIN •
VARIABLE FIN D
0.5
0.6 0.7 0.8
AIR-SIDE EFFECTIVENESS
S-69519
Figure 4-10. Closed-System Performance of Condenser Test Modules
as a Function of Effectiveness
-------
LU
Cf.
Q_
<
Z5
•V
O
<
CC.
O
^
O
O
O
I .2
0.9
0.8
0.7
0.6
PLAIN FIN O
OFFSET FIN •
VARIABLE FIN D
CD
0 20 30
40
50
60
70
80
90
SUBCOOLING, °F
100 110
S-69518
Figure 4-11. Closed-System Performance of Condenser Test Modules
as a Function of Subcooling
4-16
-------
Although the plain fin overall thermal conductance ratio, UA/UAn
is
PRED
higher than that for the other two fin configurations, the overall heat trans-
fer performance of all the fins is about the same. Consider the following
test data for an airflow that yields approximately the same air-side effective-
ness as that at the design point:
Vapor Fin
Plain
'Offset
Variable
Airflow,
Ib/min
27.1
27.0
26.8
F-85 Flow,
Ib/min
3.06
3.28
3.40
Average
Heat Transfer,*
Btu/min
667
733
716
Initial
Temperature
Di f fere nee, **
°F
173.5
189
186
Average
Heat Transfer
•
Initial
Temperature
Di fference,
(Btu/min)/°F
3.84
3.88
3.85
-•'-Total of desuperheat ing, condensing, subcooling
-,v*p-85 inlet temperature - air inlet temperature
As shown in the last column, the heat transfer rate per unit temperature
difference at the heat exchanger inlet is about the same for all fins. That
is, overall thermal conductance, DA, of the heat exchanger is not affected by
the vapor fin configuration. This result is somewhat surprising since it was
believed that the offset fin would yield the best performance based on the
known performance of plain and offset fins in air.
While the shape of the curves in Figure 4-10 cannot be completely
explained on the basis of present theory, two factors are obviously important.
In the closed-loop tests, as opposed to the open-system steam tests, the ratio
of vapor-side to air-side conductance is such that an error in prediction of
condensing film coefficient could lead to a significant error in predicted
overall UA. Since the Sol iman condensing heat transfer correlation is believed
to be increasingly conservative at low flow (dropping below the laminar film
prediction at a vapor flow of about 2 Ib/min in the offset fin core), the UA
prediction at high effectiveness is probably significantly in error from this
effect. Secondly, the relatively high subcooling experienced at high vapor
flow, while not explicable by present theory, is sufficient to significantly
reduce performance in the low effectiveness region. An unknown in these tests
is the amount of noncondensible gas in the system, which could affect both
condenser performance and amount of subcooling. Since the tests were generally
run in order of increasing airflow, a consistent buildup of noncondensibles
would have the greatest effect on the high-flow, low-effectiveness points.
Since both of the above sources of error (i.e., the effect of inaccurate con-
densing heat transfer prediction and the possible presence of noncondensibles)
were minimized during the open-system steam tests, it is believed that the
4-17
-------
open-system test results represent the more accurate prediction of the effect
of variable condensation rate on condenser performance. The closed-loop tests
serve to verify that predicted performance is obtained at the design effective-
ness of 0.80 but are less useful in identifying the effect of a single factor
such as variable condensation.
f-18
-------
SECTION 5
CONDENSER DESIGN, FABRICATION, AND TEST
-------
SECTION 5
CONDENSER DESIGN, FABRICATION, AND TEST
HEAT TRANSFER DESIGN
Condenser design requirements as established by the system contractors
are summarized in Table 5-1- Based on the requirements, a specific condenser
was designed for each contractor, accounting for the differences in working
fluids, heat rejection rates, operating temperatures and pressures, and
installation requirements.
A computer program for the calculation of condensing heat transfer
coefficients and two-phase pressure drop was written. This program was then
combined with an existing AiResearch performance prediction program to provide
a tool for analysis and design of condensing heat exchangers.
All three condensers utilize the same perforated fin on the air side, its
selection being based on the test results presented in Section 3 in conjunc-
tion with the previous testing reported in Reference 1. The air-side fin has
a fin pitch of 22 fins per in., a fin height of 0.326 in., a fin thickness of
0.004 in., and the -13 perforation geometry. The -13 perforations are rec-
tangular slots with a slot width of 0.030 in. and a slot spacing such that an
open area of approximately 25 percent is obtained. Reference 1 should be con-
sulted for a complete description of the geometry and heat transfer performance
of this surface.
Selection of the condensing side fin geometry was based on obtaining
maximum heat transfer performance within the allowable pressure drop specified
for each condenser. In addition, the condensing test described in Section k
indicated the desirability of using a surface that allows some transverse
flow (crossflow) of the vapor within the heat exchanger core. Vapor crossflow
improves overall heat exchanger performance by increasing the rate at which
vapor is supplied to the areas of high condensation rate (i.e., passages near
the air inlet face). For the TECO and SES condensers, an offset fin was found
to be optimum for this application. For the Aerojet condenser, where pressure
drop is limiting, a perforated fin is used in the condensing section and an
offset is used in the subcooler.
The following paragraphs present a discussion of the condenser computer
program that was written during this task, the heat transfer analysis, and the
design of each of the three condensers.
Condensing Heat Transfer and Pressure Loss
This program for the calculation of condensing coefficients and two-phase
flow pressure drop was written using the correlation of Soliman et al (Refer-
ence 2). Correlations presented in the referenced paper are used in the
program to calculate local and average condensing coefficients with modifica-
tions for plate fin heat exchanger applications. A brief description of the
program is given in the following paragraphs.
5-1
-------
TABLE 5-1
CONDENSER DESIGN REQUIREMENTS
System Contractor
Vapor S ide
Working fluid
Total heat rejection, Btu per hr
Flow, Ib/hr
Inlet temperature, F
Inlet pressure, psia
Condensing temperature (avg) , F
Condensing pressure (avg), psia
Liquid outlet temperature, F
Subcool ing , F
Core pressure drop, ps i
Over-all pressure drop, ps i
Maximum operating conditions
Maximum allowable leakage, std cc per
sec of hel ium
Air S ide
Flow, Ib/hr
Inlet temperature, F
Inlet pressure, psia
Outlet temperature, F
Temperature effectiveness
Overall core pressure drop, in. HLO
Aerojet
AEF-78
1.50 x 106
20,000
241
32.7
235
31.3
192.5
38.5
2.8
3.1
50 ps ig
at 300 F
1 x 10~6
52,000
85
14.7
205
0.80 Condenser
0.62 Subcooler
4.0
Thermo
Electron
Fluorinol 85
1.88 x 106
9,860
238
40.0
212
36.4
193
17.0
2.6
5.0
100 ps ig
at 300°F
1 x 10~6
75,300
85
14.7
189
0.80
4.1
Steam Eng i ne
Systems
Water
1.21 x 106
1,285
258
34.0
256
33.3
256
0.0
0.6
1.0
35 psiq
at 280°F
No visible
leakage
38,200
85
14.7
217
0.763
2.1
5-2
-------
The heat transfer and pressure drop equations are as follows;
(I)
Fn = F, + F + F
0 r m a
(2)
,0.0526
,0.261
f. N0.47 1.33
(I - x) x
vO.105
,0.522
(I - x)
0.94 0.86 / Mv
f = friction factor evaluated at Re,.
v T
-0.2.
[For round tube, f - 0.045 (Re ) ]
,2 v / v >. v2/3
1/3
(2x - I - I.25x) —
o.s(x - i) hr
(3)
,5/3
= g sin 9 (
\ u
- «)
or =
/« \2/3
a
^P
dz
gcD o
(6)
(7)
For the derivation of the above equations reference may be made to the
original paper.
5-3
-------
The average condensing heat transfer coefficient is calculated by
integrating the local coefficient between the inlet and exit qualities.
av
- x,
x
/
h dz
z
(8)
I
This program can be used as an individual program and has also been in-
corporated as a subrouti'ne into the AiResearch plate-fin heat exchanqer per-
formance program. The resultant program, called HIOOO, was used for final
sizing of the three condensers.
Effect of Oil Fi1m
In the Thermo Electron and Steam Engine Systems designs, a small amount
of oil is expected to blow by the pistons in the expander and become entrained
in the condenser inlet vapor. It is expected that this oil will form a film
on the condenser surfaces thereby introducing an additional thermal resistance
on the condensing side. An analysis was performed to estimate the oil film
thicknesses and the magnitude of the additional resistance.
With an oil film present in the condensate, the wall shear stress F (x)
given in Equation 2 is assumed at the interface between the oil film and
the condensate. Figure 5-1 shows the model used in the analysis.
CONDENSING
SURFACE
*• Y
CONDENSATE
LAYER
Figure 5-1. Model Used in Oil Film Analysis
-------
A force balance on the oil film gives
- y)
or
The oil film flow rate at any location z is
W . . ,. . = p .. S I u dy
o i 1 fi1m oil * '
or
oil film oil
F 6 p .. g sin
o oil
(9)
The local oil film flow rate is a function of the total oil flow (W .,)
01 I
entering the condenser which must be given in the problem statement. There
are two assumptions made in relating the total oil flow to the local oil film
flow rate:
(a) Constant oil film flow rate (W .. ,:-. = W ..):
^ v 011 fi1m oil
This model assumes
that all the oil mist is separated from the vapor at the inlet of
the condensing channel. This will give a maximum calculated oil
film thi ckness.
(b) Constant oil deposite rate [W .. ,.. = (I - z)w . ]: This model
assumes that the oil mist is completely mixed with the condensing
vapor at the entrance to the condenser. The deposit rate of the oil
mist is assumed to be equivalent to the vapor condensing rate.
Calculation of the oil film thickness was incorporated into the condensing
computer program. The program calculates oil film thicknesses and equivalent
condensing heat transfer coefficients which account for the oil film based upon
both of the above assumptions.
5-5
-------
The average equivalent heat transfer coefficient is
f2
I h dz (10)
av,eq x - x J z,eq
X|
, I I 01 I | , N
where T = T~ + 1 \ ' ' >
The program was used to calculate the average oil film thickness and the
average equivalent heat transfer coefficient as a function of the percentage
of oil in the inlet flow for the Thermo Electron and Steam Engine Systems
preliminary condenser designs. The results are shown in Figures 5-2 and 5~3-
The oil film thickness is shown for both of the assumptions regarding local
oil film flow rate. The average equivalent heat transfer coefficients are
shown with and without oil films. The actual percentage of oil in the inlet
vapor has not been established but it is thought to be less than 0.1 percent.
The analysis predicts reduction in condensing coefficients of 10 to 20 percent
at this oil concentration.
TECO Condenser
The final problem statement for the TECO condenser is shown in Table 5-1.
The Fluorinol-85 working fluid enters the condenser as superheated vapor at
238 F and 40 psia and leaves as subcooled liquid at 193 F. The total heat
load of 1.88 x 10 Btu/hr is divided into approximately 0.065 x 10 Btu/hr
desuperheat, 1.70 x 10 Btu/hr condensing, and 0.118 x 10 Btu/hr subcooling
(referred to the initial condensing pressure of 40 psia). In addition to the
thermodynamic design requirements of Table 5-1, packaging requirements impose
dimensional limits of 22.5 in. on the height, 58 in. on the width, and 4.5 in.
on the airflow length of the condenser.
The fractions of the condenser required for desuperheat, condensing, and
subcooling were determined from separate performance calculations of each
section. The performance calculations were made for a constant airflow length
of 4.25 in. and assumed maximum total core frontal dimensions of 58 in. by
20.8 in. (allowing 1.7 in. for the Fluorinol-85 manifolds). Results of the
calculations were plotted as (UA-ca leu lated) -r (UA-requ i red) vs the condensing
section fraction, as shown in Figures 5-4 and 5-5- The intersection points
in these figures provide the estimated ratios of condensing section size to
desuperheat section size and condensing section size to subcooling section
size. Combining the results from the two figures yields a desuperheat frac-
tion of 0.042, a condensing fraction of 0.879, and a subcooling fraction of
0.079.
Having established that 87-9 percent of the vapor-side flow length is
available for condensing, the condenser performance was calculated as a func-
tion of condenser airflow length. The results are plotted in Figure 5-6.
5-6
-------
ASSUMED VALUES
TOTAL INLET FLOW 9860
IxlO
-T-rrrm I 0, 000
IxlO
UJ
z
*:
IxlO
IxlO
23.76 (IN. )
• L = 20.25 (IN.)
• D = 0.046 (IN.)
h
• OIL: k = 0.088 (BTU/HR-FT-°F)
=5.19 (LB/FT-HR)
p = 0.0298 (LB/IN.3)
AVERAGE H (WITHOUT OIL FILM)
TTt
• f = 0.0296
• CONDENSING TEMP 217.2°F
(P = 40 PSIA)
S d L
\ .11 AVERAGE EQUIVALENT H
(WITH CONST. OIL DEPOSIT RATE) £3
AVERAGE EQUIVALENT H
(WITH CONST OIL FILM FLOW)
OIL FILM THICKNESS
(WITH CONST OIL FILM FLOW)
OIL FILM THICKNESS
(WITH CONST. OIL DEPOSIT RATE)
0.01
O.I 1.0
PERCENT OF OIL IN TOTAL INLET FLOW
0.0
S-68448
Figure 5-2. Estimated Condensing Heat Transfer Coefficient and Oil Film
Thickness for Fluorinol-85 Condenser with/without Oil Film
5-7
-------
ASSUMED VALUES
LB
• TOTAL INLET FLOW 1280 ~
IxlO
m10,000
IxlO
1000
CO
LLJ
IxlO
IxlO
= 15.84 IN.
L = 19.0 (IN.)
o
• D, = 0.046 (IN.)
n
i • DILI k = o.os (BTU/HR-FT-°F)
10.0 (LB/FT-HR)
p = 0.0318 (LB/IN.3
AVERAGE H (WITHOUT OIL FILM)
• f = 0.043
j AVERAGE EQUIVALENT H
F (WITH CONST OIL DEPOSIT RATE)
:= • CONDENSING TEMP 258UF
(P = 34.25 PSIA)
AVERAGE EQUIVALENT H
(WITH CONST OIL FILM FLOW)
OIL FILM THICKNESS
(WITH CONST OIL FILM FLOW)
OIL FILM THICKNESS
(WITH CONST OIL DEPOSIT RATE)
O.I
PERCENT OF
1.0
OIL IN TOTAL INLET FLOW
10.0
S-68449
Figure 5-3. Estimated Condensing Heat Transfer Coefficient and Oil Film
Thickness for Steam Consenser with/without Oil Film
5-8
-------
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CONDENSING SECTION 0_, 10 BTU/HR
70
-n
i—
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HEAT EXCHANGER AIR AP, IN.
-------
With a required condensing heat load of 1.70 x 10 Btu/hr, the required
airflow length is 4.66 in. and the air pressure drop is 4.23 in. HLO. The
required airflow length of 4.66 in. exceeds the specified envelope and it was
therefore decided to reduce this dimension to the specified maximum length of
4.50 in. A performance analysis of the 4.50-in. condenser indicates that it
meets performance on a theoretical basis but does not contain the amount of UA
margin usually required to compensate for such factors as flow maldistribution
and manufacturing tolerances. As a result, it is considered marginal as to
whether this condenser will reject the full required heat load of 1.88 x 10
Btu/hr under specified design conditions. The indicated heat rejection rate
is instead 1.68 x 10 Btu/hr for the condensing section alone and 1.85 x 10
Btu/hr total.
The final design of the TECO condenser is summarized in Table 5~2. The
Fluorinol-85 core pressure drop in this design is 2.6 psi and the estimated
manifold loss (vapor plus liquid) is 2.3 psi, yielding a total Fluorinol-85
pressure drop of 4.9 psi. The air pressure drop, from Figure 5~6, is 4.09 in.
H_0. The values of air pressure drop given here and for the other condensers
include a one-half velocity head loss at the inlet face and an exit loss based on
on an exit duct area equal to the frontal area of the core.
TABLE 5-2
TECO CONDENSER DESIGN SUMMARY
Airflow length
Vapor-side flow length (core only)
Width
Air-side fins (nominal)
Height
Thickness
F i ns/i n.
Vapor-side fins
Height
Th ickness
F i ns/i n.
4.50 in.
20.8 in.
57.8 in.
Perforated (-13)
0.326 in-
0.004 in-
22
Offset rectangular
0.050 in-
0.004 in.
20
5-12
-------
Aerojet Condenser
The Aerojet condenser consists of two separate units, arranged to form a
T-shaped frontal area, with the upper unit sized for the condensing load only,
with no subcooling, and the lower unit sized to have the same airflow length
as the condenser and utilize the maximum available air-side frontal area. The
problem statement for the condenser is given in Table 5-1. The AEF 78 enters
as saturated vapor at 241°F and 32.7 psia. The condensing heat load is 1.25
x 10 Btu/hr, and a minimum subcooling of 38 F is required at the outlet from
the subcooling unit.
Detail design of the Aerojet condensers resulted in the adoption of a
modular approach to accommodate manufacturing requirements. The condenser was
brazed as four separate modules and the modules welded together to form a
single unit. Similarly, the subcooler consists of three modules. Each module
in the condenser consists of alternating air and vapor sandwiches. The pas-
sage heights are 0.326 in. and 0.050 in. for air and vapor, respectively, with
the exception that the passages adjacent to the module side plates are air pas-
sages with a height of 0.163 in. (i.e. , 22R-. 163-Perf (-13)-. 004). For calcu-
lation purposes, the equivalent sandwich structure was assumed to be 116 vapor
sandwiches and 116 air sandwiches of nominal height. Each module in the sub-
cooler consists of alternating air and liquid sandwiches, with air sandwiches
of half nominal height (i.e., 22R-.163-Perf (-13)-.004) again adjacent to the
side plates. The equivalent sandwich structure for calculating subcooler per-
formance was taken as 87 vapor sandwiches and 87 air sandwiches of nominal
height.
Based on the results of the Task 2 tests, it was decided to use a
perforated fin for the vapor-side heat transfer surface in the condenser. This
fin provides open area for transverse vapor flow redistribution with less of a
pressure drop penalty than would be incurred with the use of an offset fin.
The perforation geometry used is the -15 geometry (Reference 1), which has a
slot width of 0.030 in. and an open area of approximately 12.5 percent. In
the subcooler, where pressure drop is less critical, an offset fin is used.
The airflow length for the final design, as required to reject the
condensing load in the condenser, is 4.65 in. With this airflow length, the
liquid outlet temperature from the subcooler is 192.5°?- Vapor-side pressure
drops are 2.64 psi in the condenser and 0.06 psi in the subcooler. Additional
manifold losses are estimated to be approximately 0.3 psi, resulting in a
total heat exchanger pressure loss of 3-0 psi and an outlet pressure of 29.7
psia. The saturation temperature corresponding to this outlet pressure is
231 F, giving a subcooling AT of 38.5°F. Air-side pressure loss is 3-95 in.
HLO. The final design of the Aerojet condenser is summarized in Table 5-3.
^ f
Total heat rejection at design conditions is 1.50 x 10 Btu/hr, including 0.25
x 10 Btu/hr in the subcooler.
5-13
-------
TABLE 5-3
AEROJET CONDENSER DESIGN SUMMARY
Condensing Unit
A i rf1ow 1ength
Vapor flow length (core only)
Width
Air-side fins (nominal)
Height
Thickness
F i ns/in.
Vapor-side fins
Height
Th ickness
Fins/in.
Subcooli ng Un it
Airflow length
Vapor flow length (core only)
Width
Air-side fins (nominal)
Height
Th ickness
F i ns/i n.
Vapor-side fins
Height
Th ickness
F i ns/in.
4.65 in.
14.65 in.
48.24 in.
Perforated (-13)
0.326 in.
0.004 in.
22
Perforated (-15)
0.050 in.
0.004 in.
20
4.65 in-
4.90 in.
36.22 in.
Perforated (-13)
0.326 in.
0.004 in.
22
Offset rectangular
0.050 in.
0.004 in.
20
-------
SEj Condenser
The SES design requirements are shown in Table 5-1- The steam enters the
condenser as saturated vapor and leaves as saturated liquid, with no subcooling
called for by the problem statement. Subcool ing is not desirable in this sys-
tem due to the possibility of freezing that would occur with the existence of
a liquid level in the condenser core. The absence of any subcooling poses the
requirement for a uniform fluid state (saturated liquid) at the fluid outlet,
independent of transverse position. The vapor-side fin geometry shown in
Figure 4-1 was proposed as one method of achieving this result. The variable
flow resistance of the Figure 4-1 passage is such that the transverse vapor
flow distribution matches to the vapor-to-air AT variation, resulting in a
vapor flow in each parallel flow path proportional to the local heat rejection
rate in that path. By matching local vapor flow rate to local heat rejection
rate, uniform fluid conditions are obtained at the fluid outlet. It was
inferred from the condensing heat transfer test results, however, that with
the use of a single offset fin on the vapor side there is a sufficient amount
of transverse redistribution of vapor flow within the core to accomplish
essentially the same result. For this reason, an offset fin geometry was
selected in preference to the more complex variable geometry of Figure 4-1.
The overall heat exchanger design is summarized in Table 5-4. Steam-side
pressure drop for this design is 0.6 psi for the core and approximately 1.0 psi
total. Air-side pressure loss is 2.09 in. " "
H20.
TABLE 5-4
SES CONDENSER DESIGN SUMMARY
Airflow length
Vapor-side flow length (core only)
Width
Air-side fins (nominal)
Height
Th ickness
Fi ns/in.
Vapor-side fins
Height
Thickness
F ins/in.
4.0 in.
19.0 in.
48.0 in.
Perforated (-13)
0.326 in.
0.004 in.
22
Offset rectangular
0.050 in.
0.004 in.
20
5-15
-------
STRUCTURAL CONSIDERATIONS
A structural analysis was performed on the three system condensers. This
structural analysis consisted of the following elements:
(a) Pressure containment
(b) Thermal stresses
(c) Supporting structure
The information on each of the condensers is presented in the following para-
graphs :
Pressure Containment
Pressure containment strength has been based on the maximum operating
temperature and pressure in the condenser. The strength criteria for design
based on short-time material properties were that stresses at proof pressure
not exceed the material yield strength and that the stresses at burst pres-
sure not exceed the material ultimate strength. For creep limited design,
the long-term load conditions are compared with the life of the material based
on a Larson-Miller plot of the material stress rupture at the operating tem-
perature. The design of the condensers was based on the following conditions
for the three condensers:
(a) Aerojet condenser (I 90390)
Maximum operating pressure = 50 psig at 300 F
Proof pressure = 100 psig at 70°F
(b) TECO condenser (I90370)
Maximum operating pressure * 85 psig at 300 F
Proof pressure * 183 psig at 70 F
(c) SES condenser (190640)
Maximum operating pressure = 35 psig at 280 F
Proof pressure = 86 psig at 70 F
The calculated pressure stresses for the three condensers are summarized in
Tables 5-5 through 5-7.
1. Aerojet Condenser (I 90590)
A structural schematic of this condenser is shown in Figure 5~7- For
the normal operating conditions of 18 psig at 24! °F for a life of 3500 hr,
the highest operating stresses are 5.4 ksi in the middle pan and 3.I ksi in
the flat inlet duct. For these low operating stresses at 24I°F, the creep
stresses will have a factor of safety of 6.5 on the stresses and a factor of
better than 100 to I on the operating life of the condenser.
5-16
-------
TABLE 5-5
SUMMARY OF PRESSURE STRESSES IN THE
190390 CONDENSER FOR AEROJET
PROOF PRESSURE STRESSES, 100 PSIG AT
70° F
Part
(See Figure 5-7)
Top pan at weld
Top tie bar at weld
Top header bar at
weld
Bottom pan at weld
Middle pan away from
weld
Flat inlet duct away
from weld
Inlet duct tie rods
at weld
High pressure fins
Tube sheets
Header bars
Aluminum Al loy
and Dimens ions
in.
6061 T4
t = 0.10
6061 T4
0.05 x 0.15
6061 T4
0.12 x 0.326
6061 T4
t = 0.37
6061 T6
t = 0.23
6061 Ik
t = 0.10
6061 T4
3003-0
t = 0.004
No. 22 Brazed Stock
t = 0.020
No. 22 Brazed Stock
0.12 x 0.12
Load
Weld
Tens ion
Tension and
Bend ing
Tens ion and
Bend ing
Tension and
Bend ing
Tension and
Bending
Tension and
Bend ing
Tens ion
Tens ion
Tens ion
Tens ion
Stress
ks i *<'-
4.0
16.7
3.6
11.6
30.0
16.9
15-9
2.3
0.3
0.4
Min.
Ftv
ksi*
18.0
18.0
18.0
18.0
33.2
18.0
18.0
6.0
14.4
18.0
Marg in
of
Safety
3-5
0.08
4.0
0-55
0.11
0.07
0.13
1.60
Adequate
Adequate
*MIL-Hdbk-5A Values of F after a soak at 300°F for 1000 hr
''"'Weld joint efficiency factor of 0.70 used to modify stress where relevant
5-17
-------
TABLE 5-6
SUMMARY OF PRESSURE STRESSES IN THE TECO CONDENSER
PROOF PRESSURE STRESSES, 183 PSIG AT 70°F
Part
(See Figure 5-8)
Top pan and weld
Top tie bar at weld
Top header bar at
weld
Bottom pan at outlet
port
Bottom pan at center
Bottom pan at under-
cut for thermal
rel ief
Flat inlet duct at
span center
Inlet duct tie rods
at weld
High pressure fins
Tube sheets
Header bars
Alumi num Al loy
and Dimens ions ,
in.
6016-T4
t = 0.125
6061-T4
0.05 x 0.15
6061 -T4
0.12 x 0.326
6061 -T4
t = 0-50
6061 -T6
t = 0.31
6061-T6
t = 0.20
6061 -T4
t =0.128
6061 -T4
d = 0.1875
3003-0
t = 0.004
No. 22 brazed stock
t = 0.020
No. 22 Brazed stock
0.12 x 0.12
Load
Weld
Tens ion
Tension and
Bend ing
Tension and
Bending
Tension and
Bend i ng
Tension and
Bend ing
Tens ion and
Bend ing
Tension and
Bend ing
Tens ion
Tens ion
Tens ion
Tens ion
Stress
ks i **
4.9
12.6
5-31
15-11
31. **
30.8
15-3
17-0
4.2
0.6
0.7
Min
Fty
ksi*
18.0
18.0
18.0
18.0
35-0
35-0
18.0
18.0
6.0
14.0
18.0
Margin
of
Safety
2.67
0.43
2.39
0.19
0.118
0.14
0.19
0.06
0.43
Adequate
Adequate
*MIL-Hdbk-5A Values of Ft after a soak at 300°F for 1000 hr
->-Weld joint efficiency factor of 0.70 used to modify stress where relevant
5-18
-------
TABLE 5-7
SUMMARY OF PRESSURE STRESSES IN THE SES CONDENSER
PROOF PRESSURE STRESSES, 86 PSIG AT 70°F
Part
(See Figure 5-9)
Top and bottom pan at
weld
Tie bar at weld
Header bar at weld
Inlet duct at span
center
Inlet duct tie rods
at weld
High pressure fins
Tube sheets
Header bars
Alumi num Al loy
and Dimens ions ,
in.
6016-T4
t =0.100
6061 -Jk
0.05 x 0.15
6061 -T4
0.12 x 0.326
6061 -T4
t =0.100
6061-T4
OD = 0.200
ID - 0.150
3003-0
t = 0.004
No. 22 Brazed Stock
t = 0.020
No. 22 Brazed Stock
0.12 x 0.12
Load
Weld
Tens ion
Tens ion and
Bending
Tension and
Bend ing
Tension and
Bend ing
Tens ion
Tens ion
Tens ion
Tens ion
Stress
ksl**
2.88
5-93
2.50
15-5
9.8
1 .98
0.29
0.3^
Min
Ftv
ks I*
18.0
18.0
18.0
18.0
18.0
6.0
14.0
18.0
Margin
of
Safety
5.26
2.04
6.20
0.16
0.84
2.03
Adequate
Adequate
*MIL-Hdbk-5A Values of F after a soak at 300°F for 1000 hr
•-'-Veld joint efficiency factor of 0.70 used to modify stress where relevant
5-19
-------
PAN TIE RODS
TOP PAN
HEADER BARS
000
FLAT INLET DUCT
TIE RODS
TIE BARS
i
ho
O
MIDDLE PAN
S-6'?588
WELD AT BOTTOM PAN OUTLET
Figure 5-7. Aerojet Condenser (190390) Structural Schematic
-------
2. TECO Condenser (I 90570)
A structural schematic of this condenser is shown in Figure 5~8. For
the normal operating conditions of 25 psig at 238°F for a life of 3500 hr,
the highest operating stresses are 2.33 ksi in the tie rod and 2.3 ksi in the
flat inlet duct. For these low operating stresses at 238°F, the creep stresses
will have factors of safety of at least 6.5 on the stresses and a factor of
better than 100 to I on the operating life of the condenser.
3. SES Condenser (190640)
A structural schematic of this condenser is shown in Figure 5~9- For
the normal operating conditions'of 20 psig at 258°F for a life of 3500 hr,
the highest operating stresses are 2.3 ksi in the tie rod and 3.6 ksi in the
flat inlet duct. For these low operating stresses at 258°F, the creep stresses
will have a factor of safety of at least 5.0 on the stresses and a factor of
better than 100 to I on the operating life of the condenser.
Thermal Stresses
Temperature differences are developed within the condensers which produce
thermal strains. The maximum normal operating transient temperature difference
between the bar and tube sheet is calculated to be 20°F for 1000 cycles and
I5°F for 2000 cycles. A design life of 3500 cycles was selected for the design
point under normal condenser operating conditions. The condensers have adequate
life capability for this requirement, however^ some further thought should be
given to the possibility of higher metal temperature differentials occurring
when a hot condenser core is subjected to a cold water splash condition.
Supporting Structure
An analysis was conducted on the support structure for the three conden-
ser and fan assemblies. The support structures were analyzed for a 10-g shock
load in three axes and a 2.0-g vibration input load in three axes with a re-
sponse of 5 times the input to give a maximum vibratory response output os
10 g. The structures should be provided with isolators to limit the vibra-
tory response to 10 g.
The strength criteria for shock loads are that the stresses do not
exceed the ultimate strength of the material. For vibratory loads, the struc-
ture should be designed so that the alternating stresses do not exceed the
allowable endurance limit fatigue properties of the material.
The component weights listed in Table 5-8 were used in this analysis.
5-21
-------
TOP PAN
vn
M
~\
HEADER BARS
FLAT INLET DUCT
RODS
S-6'?b87
OUTLET DUCT
OUTLET DUCT
Figure 5-8. TECO Condenser (190370) Structural Schematic
-------
TOP PAN
vn
i
NJ
HEADER BARS
TIE BARS
BOTTOM PAN
OUTLET
DUCT
S-69590
Figure 5-9- SES Steam Engine Condenser (190640) Structural Schematic
-------
TABLE 5-8
COMPONENT WEIGHTS
Part
Condenser
Subcooler assembly
Liquid in condenser
Liquid in subcooler
Wet weight of condenser
Two motors
Two fans
Support structure
Fan and structure
Total weight condenser and fan
TECO
190370
Weight, Ib
137
0
10
0
147
0
70
17
89
236
Aerojet
190390
Weiqht, Ib
80
27
0
18
125
70
46
24
! 140
265
SES
I 90640
,, Ib
93
0
I I
0
104
0
66
20
86
190
1. Heat Exchangers
Vibration isolators equivalent to the Lord BTR elastomeric mountings,
HT2-80 or 200P-45, should be used to limit the maximum response output to
10 g. The heat exchanger fins are very flexible and are not capable of
supporting the heat exchanger core weight in either the vertical or the longi-
tudinal direction of the airflow. For vertical loads the air fins act as
very flexible guided cantilever beams since the thickness per fin is 0.004
in. For longitudinal loads, the fluid fins act in a similar manner; the
thickness per fin is also 0.004 in. As a result, the top and bottom pans are
not capable of acting as an integrated beam due to insufficient shear carrying
ability of the core.
The top and bottom pans must act as separate beams to support the longi-
tudinal and vertical loads. For three isolator supports on each side of the
heat exchanger and no middle support, the unsupported pan beam span will be
50.0 in. for the Aerojet, 60.19 in. for the TECO, and 51.62 in. for the SES
units. The top and bottom pans are insufficient in themselves to carry the
lO-g vibratory loads in the vertical direction for all three designs.
To make the heat exchanger core more rigid and to reduce the vibration
amplitude and strengthen the structure so that its alternating stresses will
be
less
than the endurance limit of the material, a truss structure shown for
5-24
-------
the Aerojet condenser in Figure 5-10 is recommended, and a similar truss may be
used for the other condensers. This truss may be located in the discharge
shroud of the condenser or it may be located at the inlet to the condenser.
If the truss fs located in the shroud between the condenser and fan, this
shroud will have to be lengthened to prevent obstruction of airflow through the
outer sides of the shroud. For the best structural design, the truss should
be located as close as practical to the core.
An alternate and preferred structural design is to provide the condenser
core with at least one support at its middle. This support is shown with an
isolator in Figure 5~10 (adjacent to liquid outlet port). Use of the center
mount eliminates the need for a truss structure. This will divide the beam span
in half and reduce the alternating bending stresses to one quarter such that
these stresses would be less than the endurance limit of the material.
The top and bottom pans and cores of the TECO, Aerojet, and SES condensers
are sufficiently rigid and strong to take the 10-g vibratory loads in the longi-
tudinal direction so the condenser core may be supported on three vibration
isolators at each end. An additional longitudinal truss must be provided be-
tween the cores to strengthen the middle pan in the Aerojet condenser to carry
the longitudinal ±10-g vibratory loads. By using this structural configuration,
the alternating stresses in the components of the Aerojet unit will be less
than the endurance limit of the material. No additional reinforcing is re-
quired to carry longitudinal vibratory loads in the TECO and SES units.
To isolate the condenser and fan package from external vibration inputs
and to maintain compatibi1ity with the isolation mounting, the inlet ducts to
the condenser must be provided with flexible connections. Also, the outlet
duct and the motor lines must be flexible to eliminate vibration input through
these 1ines.
2. Frame Assembly
For automotive installations with a 10-g response output, the condenser
and fan frame assembly should be reinforced as follows: on the frame assembly
for the Aerojet condenser and fan, the top of the two support rings (which are
machined away to provide for the inlet ducts to the condenser) should be fab-
ricated as, sol id 0.625-in. by 1.75-in. rectangular rings, and the inlet ducts
should be indented to accommodate the rings. As shown in Figure 5-11, these
solid rectangular rings should have a tapered thickness which increases from
0.625 in. at station (T) to 1.00 in. at station (T). This tapered thickness
is necessary to support the torsional stresses that are transmitted through
the upper arches of these support rings. Also, to carry the torsional stresses
through the lower arches of the rings, rectangular box sections must be incor-
porated between stations (V) and (6j.
For the automotive installation of the TECO condenser, the fan annulus
support ring, spokes, and hub act as an integrated structure for supporting the
fan and belt loads. The channel ring section between the upper support bolts
should be a rectangular box section to prevent excessive twist of the fan
annulus which will occur when subjected to vibratory loads. Likewise, the
channel ring section between the lower support bolts should be a rectangular
box section.
5-25
-------
.STATION NO. 8
un
I
NJ
BELLOWS COUPLING BETWEEN CONDENSER AND
REGENERATOR FOR VIBRATION ISOLATION
TOP PAN
VIBRATION
ISOLATORS
3 ON EACH
SIDE
USE FLEXIBLE COUPLING
LONGITUDINAL TRUSS
BOTTOM PAN
BOTTOM SUPPORT AND ISOLATOR MAY REPLACE TRUSS
S-69585-A
Figure 5-10. Aerojet Condenser with Truss
-------
VJ1
I
to
-J
NOTES:
• STATION NUMBERS ARE CIRCLED
• DIMENSIONS IN INCHES
H 1.75 I—
STATION NUMBER (TYP)
S-69586
Figure 5-I I. Aerojet Support Frame for Fans
-------
Table 5-9 summarizes the alternating stress calculations for the various
structural components in the Aerojet unit. Vibratory stresses in the TECO and
SES units are approximately the same.
TABLE 5-9
SUMMARY OF STRUCTURAL DYNAMIC STRESS IN THE
190390 AEROJET CONDENSER
2-g input with a response of 5-to-1 to
yield a 10-g maximum response output
Station
1
2
2
3
k
4
5
6
7
8
9
10
Di rect ion
of Load
Longitudinal
Longitudinal
Long! tudinal
Longitudinal
Long i tud inal
Long i tud inal
Longi tud inal
Long i tud inal
Longi tud inal
Longi tud inal
Longi tud inal
Vert ical
Part (Refer to Figure
5-10 and 5-11)
Ring at upper center
Ring at upper center bolt
Ring at upper center bolt
Ring at lower center
Ring at lower center bolt
Ring at lower center bolt
Center of truss
Ring at lower outer edge
Lower outer beam
Top pan at center
Bottom pan at center
1/4-in. dia truss rods
Type
Stress
Bend ing
Pr inc ipa 1
Shear
Pr incipal
Tens ion
Bend ing
Pr incipal
Shear
Principal
Tens ion
Bending
Shear
Bending
Bend ing
Bending
Tens ion
Vi bratory
Stress
ksi
10.45
7-98
11.00
10.45
7-60
12.00
9-25
6.50
9-60
2.48
1.00
10.00
For 6061-T4 Al the endurance limit F = 12.00 ksi
5-28
-------
Structural Summary
1. Aerojet Condenser (190390)
The 190390 condenser and subcooler meets the strength criteria for design
based on short-time material properties, and the stresses at proof pressure
will not exceed the material yield strength. The normal operating stresses
are low and the maximum long-term operating temperature is only 2^1°^; there-
fore, it should not rupture during the 3500 hour operating life because it has
a factor of safety of 6.5 on creep stress and a factor of 100 on creep life
based on a Larson-Miller plot of the material stress rupture at the operating
temperature. The thermal stresses are low and in the elastic range of the
material,
The Aerojet condenser inlet ducts may take static loads of 500 Ib in the
longitudinal and lateral axes and 150 Ib in the vertical direction for handling
and installation; however, these ducts are not designed to take any dynamic
loads and should be provided with bellows or a flexible coupling which has a
spring rate of 1000 Ib per inch or less in the three axes with a maximum dis-
placement capability of 0.50 in. In the static systems test program of the
Aerojet condenser, the above handling loads should not be exceeded.
To limit the maximum response output to 10 g or less, vibration isolators
should be ut i1ized.
The present condenser core and pan structure of all three units needs to
be reinforced to support a vertical vibratory response of 10 g if no midspan
mount is provided. One of the following structural changes is recommended:
(a) The recommended installation is to provide all three units with a
midspan support of the condenser core with vibration isolators at
all supports, as shown by Figure 5-10.
(b) In lieu of (a) above, an alternate design would be to provide the con-
denser core with a truss structure, as shown in Figure 5-10. Additional
space may be required in the shroud between the condenser and the
grille to reduce the cooling airflow losses.
(c) The least desired alternate design would be to weld gussets and
plate beams on the top, middle, and bottom pans. This would be
expensive and heavy alternative.
an
For the automotive installation of the Aerojet condenser and fan frame
assembly., the top two support rings should be fabricated as solid l.75-in.-
wide rectangular rings having a tapered thickness that increases from 0.625
in. at station (T) to I.00 in. at station (2) of Figure 5-11. The lower ring
arches between stations (Z) and (5) of Figure 5-11 should be fabricated as a
box sect i on.
5-29
-------
2. TECO Condenser (190370)
The I90370 condenser and subcooler meets the strength criteria for design
based on short-term materia1 propert i es, and the stresses at proof pressure
will not exceed the material yield strength. The condenser was not analyzed
for a burst pressure because no definition for the burst requirement is speci-
fied. The normal operating stresses are low and the maximum long-term operat-
ing temperature \s only 238°F; therefore, it should not rupture during the design
life of the engine (3500 hr) because it has a factor of safety of 3.0 on creep
stress and a factor of 45 on creep life based on a Larson-Miller plot of the
material stress rupture at the operating temperature. The thermal stresses are
low and in the elastic range of the material.
As outlined in the Aerojet condenser discussion, the TECO condenser and
fan annulus support rings should be fabricated as rectangular box sections
on the upper and lower segment to prevent excessive twist of the fan annulus
due to the 10 g response output loads.
For the automotive installation of the TECO condenser, spring loaded idler
pulleys should be provided on the belt between the fan pulley and the drive
pulley on the automobile frame. These idler pulleys should be designed to
take up a racking motion between the fan and auto frame.
3. SES Condenser (190640)
This condenser and subcooler meets the strength criteria for design based
on short-time material properties, and the stresses at proof pressure will not
exceed the material yield strength. The normal operating stresses are low and
the maximum long-term operating temperature is only 258°F; therefore, it should
not rupture during the 3500-hr design life of the engine because it has a factor
of safety of 5»0 on creep stress and a factor better than 100 on creep 1 ife
based on a Larson-Miller plot of the material stress rupture at the operating
temperature. The thermal stresses are low and in the elastic range of the
material. The same comments made for the support frame of the TECO condenser
also apply to the SES condenser.
DETAIL DESIGN
Aerojet
Outline dimensions for the condenser and subcoder assembly are shown on
Drawing 190391- For manufacturing purposes the condenser case consists of
four separate brazed modules which are subsequently welded together. Likewise,
the subcooler sections consists of three separate modules.
The vapor inlet ducts are designed to mate directly to the engine recuper-
ator. The ducts are tapered to provide a constant flow area. Tension tie rods
extend across the inlet ducts to minimize the stresses in the flat sections of
the duct. The inlet face of the duct is subject to a final machining operation
to ensure that the face is parallel to the front face of the condenser (engine
installation requirement). The ends of the vapor Inlet manifold are tapered
to clear the vehicle hood.
5-30
-------
)39/ U I
IOO'l44 OLaLf ""j^
CONDENSER AND
UBCOOLER ASSY
/9O39/
-------
SECTION 13" I)
-------
Mounting brackets are welded directly to the sides of the condenser core
assembly. The threaded bosses which are welded to the manifolds are for instru-
mentation purposes. A single liquid outlet is utilized.
The manifolds are designed to minimize the overall condenser height and
thus, the intermediate and liquid outlet manifolds are flat rather than round
as would be desired to minimize metal thickness. The flat intermediate and
liquid outlet manifolds are machined from plate stock and are a maximum of
3/8 in. thick as required to withstand the specified operating pressure.
The condenser-to-fan shroud is designed to be welded to the back face of
the condenser.
Outline dimensions of the final condenser and fan assembly are shown on
Dwg No. 190390. A separate aluminum alloy frame assembly is used to mount the
fan assemblies to the condenser. The weight of the two fan and hydraulic
motor assemblies was estimated to be 110 Ib, and hence, a robust structure is
required to withstand the expected operating loads. A rubber pad is used at
the bottom mounting surface to absorb the differential thermal expansion be-
tween the frame and condenser. A description of the fan assemblies is pre-
sented in Section 6.
Thermo Electron
Outline dimensions of the condenser assembly is shown on Dwg No. 187532.
For manufacturing purposes the core is designed to be brazed in four separate
modules which are welded together to form the complete assembly in a subsequent
operat ion.
The vapor inlet ducts are tapered to minimize area changes and tension
ties are used to reduce the stress in the flat portion of the duct. Mounting
brackets are welded directly to the sides of the core. As specified by TECO,
two vapor Inlets and two liquid outlets are incorporated in the design. The
liquid outlet manifold was designed to be flat to reduce the overall height
of the condenser. The condenser-to-fan shroud is welded directly to the back
face of the condenser.
Outline dimension of the final condenser and fan assembly are shown on
Dwg No. 190370. For this design, a three-piece frame is used to mount the
fans to the condenser. A center piece is used to join to two fans together;
two side pieces secure the frame to the condenser. The side pieces incorporate
the vehicle mounts. A 1/4-in. thick pad of closed-cell foam silicone rubber
is placed between the fan front flange and the condenser-to-fan shroud to
prevent air in-leakage. No provisions are made for a center bottom support
although a cradle type of mount could be used to help support the assembly in
the actual automobile installation. A description of the fan assembly is pre-
sented in Sect ion 6.
5-33
-------
Steam Engine Systems
The condenser assembly is shown on Dwg No. 187562. The design is similar
to the TECO unit except that a single vapor inlet duct is used. The ends of
the inlet manifold are tapered to clear the vehicle hood. The outlet manifold
is cylindrical and incorporates a threaded bosses on each end which are designed
to mate with the SES noncondensible gas vent manifold.
Outline dimensions of the condenser and fan assembly are shown on Dwg No.
190640. The assembly is similar to the TECO unit and uses a three-piece frame
assembly to mate the fans with the condenser. Location of the fan drive pulley
and V-belts is shown for reference. A description of the fan assembly is pre-
sented in Section 6.
FABRICATION
Conventional aluminum heat exchangers are fabricated using a salt bath
brazing technique, which requires a flux during the brazing process for re-
moving the oxide film from the aluminum parts. Such fluxes are corrosive,
and it is necessary to wash the parts thoroughly to remove both residual flux
and brazing salts. Even minute quantities of entrapped salt and flux will
cause severe corrosion.
Techniques for salt bath brazing are well developed at AiResearch and
thousands of units are in the field and successfully operating. It was recog-
nized, however, that due to the very small vapor passages in the EPA condensers
(0.050 in. high), that salt entrapment would be likely. It was decided, there-
fore, to use a relatively new vacuum fluxless brazing process.
The key to the vacuum fluxless brazing process is the use of a clad braz-
ing sheet. The clad sheet consists of a proven heat exchanger core alloy, type
3003 or 6951, which is clad on both sides with an AI-Si-Mg alloy. Total clad
thickness is 5 to 15 percent of the brazing sheet thickness, depending on gage
or appli cat ion.
Silicon lowers the melting point of the clad to below that of the core
alloy, allowing parts to bond without thermal distortion under proper condi-
tions of clamp ing force , temperature, and vacuum. In the furnace environment,
the protective oxide film on the aluminum surface breaks up, exposing the
clean surface essential for brazing. Magnesium in the clad also vaporizes
under these conditions, helping to disrupt the oxide film. Oxygen in the
furnace atmosphere preferentially combines with the magnesium, preventing
reoxidation of the clean aluminum. The clad melts and wets joint surfaces
thus forming the brazed joint.
5-34
-------
-------
-------
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Page 5-45
-------
For the three EPA condensers the following specific alloys were selected
for the tube plates:
Aerojet and SES
Reynolds MD-133 Alloy, 0.020 in. thick, 15 percent cladding thickness
(0.003 in.). This alloy consists of type 3003 core with type X4003
cladding which contains 7-5 percent Si and 2.5 percent Mg.
Thermo Electron
Reynolds MD-151 alloy (present designation is XS) , 0.020 in. thick,
15 percent cladding thickness (0.003 in.). This alloy consists of
type 6951 core with type X^003 cladding which contains 7.5 percent Si
and 2.5 percent Mg.
A heat treatable alloy, MD-151 » was specified for the TECO core because of its
higher working pressure. All other alloys used in the condenser were conven-
tional heat exchanger aluminum alloys.
Core Assembly
All units were basically similar and used a conventional bar-and-plate
construction technique. To facilitate brazing, the condenser was divided into
four separate modules. After completing module brazing, the four units were
weld-assembled to form the complete condenser. The identical -13 perforated
air side fin was used on all assemblies and is shown on Figure 5-12. The off-
set vapor fin used in the TECO and SES assemblies is shown in Figure 5-13
and the perforated plain fin as used in the Aerojet assembly is shown in
Figure 5-1^, Figure 5-15 shows a completed core module assembly (TECO condenser),
and Figure 5-16 shows a closeup of the core construct ion.
The air-side and vapor-side fins were fabricated from type 3003 alloy;
the header and reinforcement bars from type 6951 core alloy; and the tube
sheets from Reynolds MD alloys.
When the first core modules were brazed, a considerable amount of distor-
tion was observed as a result of random movement of the header bars from their
normal vertical orientation. There was also extensive leakage at the tube-to-
header joints. The cores were mechanically straightened in a press and TIG
weld repairing was used to achieve a near leak-tight condition. The cores
successfully passed a proof pressure test after these operations.
After extensive weld repair, some of the core modules still exhibited a
series of minute header bar-tube sheet joint leaks. The leaks were discernible
as streams of approximately 0.030 in. dia bubbles when the cores were pressur-
ized to approximately 40 psig. At a pressure less than 40 psig, there were
no bubbles. It was obvious that the module would not pass the required helium
mass spectrometer leak test without additional repair.
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Figure 5-15- Condenser Core Module
5-
-------
LIQUID PASSAGE FINS
Figure 5-16. Condenser Core Module Detail Component Arrangement
5-50
-------
The cores were believed to be structually sound and useable except for
the minute leakage. Accordingly, the following repair procedure was used on
one of the leaking cores. After the outside surfaces of the core were treated
by a chem-film process, it was dipped in a thinned epoxy ester bath while a
vacuum was induced in the coolant passages in accordance with AiResearch Pro-
cess Specification FP-3&. After this treatment, the core was thoroughly
drained and the epoxy was then cured at 275°F; two coats of epoxy, each approxi-
mately 0.00005 in. thick, were applied in this manner. The unit was then leak-
tested, using a mass spectrometer. A hard vacuum was drawn inside the core
while a helium atmosphere was maintained outside. Under these conditions, a
- 8
leakage of 8 x 10 sec of helium per sec was measured. This leak rate was
well below the maximum allowable condenser leakage of 1 x 10 sec of helium
per sec.
The epoxy ester repair procedure was successful and was subsequently
adapted for use on all units. AiResearch has accumulated considerable experi-
ence in the field with this coating and it has been found to be extremely
durable. In addition to sealing minute leaks, the coating provides excellent
corrosion protection for the air-side fins. The coating does not add a
measurable amount of resistance to heat transfer.
To verify the structural integrity of the module core assembly braze
joints, the epoxy treated core assembly was subjected to a room temperature
hydrostatic burst test. The assembly withstood a pressure of k75 psig with-
out leakage. At this pressure, the joint between the pan and core failed.
No failure or distortion was observed in the braze joints. Figure 5-17 is a
photograph of the failed module. The failure is adjacent to dial indicator
No. 5.
Proof pressure for the Aerojet assembly is 85 psig while the maximum
proof pressure requirement is 183 psig for the TECO assembly. Because of the
margin of safety demonstrated by the specimen, no additional tests were
performed.
Subsequent development of the brazing operation eliminated the problems
associated with core distortion and gross leakage. To ensure meeting the
helium leakage requirement, it was necessary, however, to coat the external
surfaces of the final condenser assembly with epoxy ester as previously
described.
Aerojet
Final assembly of the Aerojet condenser consisted of the following tasks:
(a) Weld assembly of the four condenser core modules and three
subcooler core modules and testing of the weld joints.
(b) Welding of the inlet and outlet pan assemblies and shroud to
the core. The appearance of the unit at this stage of assembly
is shown by Figure 5~18.
5-51
-------
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Figure 5-18. Aerojet Condenser, Partially Assembled
5-53
-------
(c) Welding of the support bracket to the sides of the cores and
the flanged ducts to the inlet and outlet pans.
(d) Machining of flange faces and brackets.
(e) Leak testing with air at 50 psig and proof pressure testing at
85 psig. At this point, all external joints were bubble-free at
50 psig; however, there were minute leaks at three of the core
module tube header joints. These leaks were not evident at
approximately 10 psig.
(f) Epoxy ester coating of the exterior surfaces of the core and
helium leak testing. The measured leakage, with a vacuum inside
the vapor passages and a helium atmosphere outside was 1.2 x
10 sec of helium per sec.
(g) Weld assembly of the fan support frame.
(h) Painting and final assembly of the condenser and support frame.
Figure 5-19 is a photograph of the front view of the completed Aerojet
condenser assembly showing the core assembly. Figure 5-20 is a back view
showing the inlet and outlet ducts and the condenser-to-fan shroud. Figure
5-21 is a back view of the condenser assembly showing the inlet manifold
assembly. Figure 5-22 shows the final fan support frame of the condenser.
Thermo Electron
A similar assembly sequence was followed for the TECO assembly as for
the Aerojet unit. The completed assembly was subjected to a 185 psig proof
pressure and leak tested. Helium leakage, measured with a vacuum inside the
- 9
vapor passages and helium outside, was 4.8 x 10 sec per sec. Figure 5-23
shows the front view of the completed assembly and Figure 5-24 shows the back
view. The three-piece fan support frame, assembled with the two fans, is
shown in Figure 5-25.
Steam Engine Systems
A similar assembly sequence was followed for the SES assembly as for the
Aerojet unit. The completed assembly was subjected to an 85 psig proof pres-
sure test. Before final painting, the condenser was subjected to a helium
leak test and the final leakage value was 4.8 x 10 sec per sec. A front
view of the completed assembly is shown on Figure 5-26. Figure 5-27 is the
back view of the completed assembly. The three-piece fan support frame,
assembled with two fans is shown in Figure 5-28.
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Figure 5-19. Aerojet Condenser Assembly—Front View
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Figure 5-21. Aerojet Condenser-Top V.«
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72207-7
Figure 5-22. Aerojet Condenser—Fan Support Frame
5-58
-------
v_n
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Figure 5-23. Thermo Electron Condenser—Front View
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Figure 5-25. Thermo Electron Condenser--Three-Piece Fan Support Frame
5-61
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Figure 5-26. Core Module Structural Test
5-62
-------
Figure 5-27. Steam Engine Systems Condenser--Front View
5-63
-------
Figure 5-28. Steam Engine Systems Condenser-Back View
5-64
-------
Figure 5-29. Steam Engine Systems Condenser—Three Piece Fan Support Frame
5-65
-------
HEAT TRANSFER PERFORMANCE TEST
Condenser heat transfer performance can be established by testing a
section or module of the full-size condenser in striaght crossflow. The full-
size condensers for all three systems actually consists of four separate
modules which are welded together to form a complete assembly. Thus, if pro-
duction type detail parts, i.e., fins, header bars, and tube plates are
utilized in the module (1A section of the full condenser), then the module
test results can be considered as fully representative of the full-size con-
denser performance. Module testing is considered wise since a separate test
unit can be utilized rather than risk contaminating or damaging a deliverable
assembly.
To minimize test cost it was decided to eliminate testing with the
organic working fluids (AEF-78 and Fluorinol-85) and to utilize steam or hot
water as the working fluid. This procedure is believed to be reasonable
because most of the heat transfer resistance is on the air-side of the unit
and any change in the condensing-side performance would have only a small
effect on the overall performance. The condensing side resistance is calcu-
lated to range from k percent (steam) to 18 percent (Fluorinol-85) of the over-
all resistance. Moreover, the condensing side performance has been determined
as described in the condensing heat transfer test, Section k.
Thus, the remaining item to be verified is the air-side performance. The
condenser design is based on the results of the -13 test core as previously
described. This data was originally obtained using s small k in. by k in.
face area test core which utilized preprototype perforation tooling. It was
considered to be wise to test a full-size module which utilized all of the
product ion-type detail parts and fluxless brazing processer.
Test Unit
A SES module was selected to represent the typical condenser section. A
sketch of the module design is shown in Figure 5-29. Figure 5-30 is a photo-
graph of the completed assembly. The unit was designed to utilize high velocity
steam as the heat source. Thus, a fan-shaped steam inlet duct was specified
to provide uniform steam distribution across the inlet face of the test module.
Flanges, designed to mate with the laboratory air ducts, are incorporated on
the air side of the test module. The test module finishing operations, chemical
film, epoxy dip, and paint, are identical to those used for the full-size
condenser.
P e r fo rma n ce Te s t
This task was terminated after completing the test unit fabrication.
No performance tests were conducted.
5-66
-------
-11-7/8 IN.-
21 IN.
18-7/8 IN.
•14 IN.-
k IN.
AIR FIN
LENGTH -
3 IN,
-6 IN.
S-76503
Figure 5-30. Heat Transfer Performance Test Module
5-67
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Figure 5-31. Heat Transfer Performance Test Modul
5-68
-------
SECTION 6
FAN DESIGN, FABRICATION, AND TESTING
-------
SECTION 6
FAN DESIGN, FABRICATION, AND TESTING
DESIGN
Fan design requirements as generated by the three system contractors are
summarized on Table 6-1. Installation drawings supplied by the system con-
tractors all depicted a tube-axial fan configuration because of the severe
axial length restrictions in the engine compartment: i.e., space was alloted
for the fan impeller but no axial length was allowed for the inclusion of
stator vanes for removing the swirl velocity component from the fan discharge
air.
After a detailed review of the fan problem statements, the following fan
design philosophy was adopted:
(a) Design a high performance tube-axial fan to meet the specified
performance. A sophisticated aerodynamic design will be essential
to meet the high efficiency requirement.
(b) If possible, use a single basic aerodynamic design for all applica-
tions. Close to optimum performance would be achieved by varying
the tip diameter and rotational speed. This procedure was considered
essential to minimize program cost since a single aerodynamic design
would require only one impeller casting.
(c) Maximize fan diameter consistent with the installation dimensions
to improve condenser effectiveness and minimize the fan dump loss.
(d) Incorporate a slight radial component in the fan discharge velocity
(conical flow) to help reduce the fan sensivity to flow blockage
(e) Minimize rotational speeds to reduce swirl losses and acoustic
noi se
(f) Select a design that can be produced using high volume, low cost
processes.
(g) Provide ample stall margin to allow for a probable increase in the
installation pressure loss.
(h) Design for minimum axial length to fit within the alloted space
6-1
-------
TABLE 6-1
FAN DESIGN REQUIREMENTS
System Contractor
Total number of fans
Design point vehicle speed, mph
Ambient temperature, F
Ambient pressure, psia
Ae ro j e t
2
50
14.7
14.7
Air temperature at fan inlet, °F 205
Total airflow, Ib/hr 52,000
Volumetric airflow per fan, cfm 7,250
Inlet flow dynamic pressure, in. HO
Inlet recovery factor, per cent
Net ram pressure rise, in. H90
Predicted condenser air-side pressure drop, in. H?0
Installation loss, in. H 0^ '
Required fan total pressure rise, in. HO
Calculated air horsepower
Specified efficiency, percent
(2)
Fan rotation
Fan drive mechanism
Maximum overall length
Fan noise level, dbA maximum
1.17
78
0.91
3-95
1.67
4.71
5.4
70
ccw
Hydraul ic
Motor
6.4
(4)
Thermo
Elect ron
2
90
14.7
1^.7
189
75,300
10,300
3-79
100
3-79
4.09
2.80
3.10
5-0
70
ccw
V-belt
4.0
(4)
Steam
Engine
Systems
2
32
14.7
14.7
217
38,198
5440
0.44
100
0.44
2.09
1.22
2.87
2.5
Not
Speci f ied
ccw
V-belt
Not
spec! f i ed
(4)
(1) Installation loss includes losses across the bumper, grille, transition
section between condenser and fans, and the engine compartment.
(2) Fan rotation as viewed from drivers seat
(3) Hydraulic motors and V-belt drive mechanisms are supplied by the system
contractor.
(4) Fan noise levels have not been specified. However, the overall vehicle noise
level, of which the fans are probably a major portion, shall not exceed 77 ,
dbA at 50 ft according to the EPA vehicle spec fications.
6-2
-------
Impeller Des i gn
The basic fan impeller design was based on the TECO requirements because
it had the longest blades. The NACA 65 series blade section thickness distribu-
tion was used because experience has shown this section to be superior to the
older circular arc and parabolic arc sections for this class of tube axial fan.
A hub shape previously found effective on this class of machine was used and
the maximum permissible tip diameter of 22 in. used. The maximum number of
blades for trouble-free casting, 17, was selected. By iteration, the hub
diameter and approximate chord length were established. The flow paths at
various blade sections were obtained by electrical analog assuming a uniform
pressure at the blade trailing edge. The velocity triangles presented in
Figure 6-1 were obtained from this approach. The estimated rotational speed
was 2500 rpm.
Each individual blade section was carefully analyzed to allow for maximum
surge margin without compromising too much on negative stall. This results in
the selection of slightly negative incidence angles for sections near the tip
where the flow angles are very high. Blade parameters are summarized on Table
6-2.
The rotational speeds for the TECO and SES fans were selected to produce
the pressure rise required by the problem statements in the areas of maximum
efficiency and the impeller tip diameters were calculated to reduce the flow
area to that required for the specified flow rate. It was recognized that
minor speed changes could be made on prototype units to make final adjustments
to performance.
A stress analysis of the impeller was conducted on the 22 in. impeller
for a rotational speed of 4000 rpm. Effective stress levels in the blades and
hub are presented in Figures 6-2 and 6-3- All stresses were well within the
allowable stresses for the impeller material. A blade vibration study indi-
cated that the first bending mode occurs at approximately the twelfth harmonic
of blade frequency at operating speed. Results of that study are presented in
Figure 6-4. Additionally, it was found that at 4000 rpm the forward movement
of the blade tip of a 22 in. impeller would not exceed 0.015 in. max. In the
interests of safety, it was decided to limit fan operating speed to 3600 rpm.
Housing Design
In order to insure maximum fan efficiency, the impeller to housing clear-
ance must not be so close as to cause excessive boundary layer drag, nor so
great as to permit significant leakage. For fans of this class, that value
is approximately 0.030 in., which is the nominal value used for the three
designs. Clearances of this magnitude dictated that the impeller had to be
accurately positioned relative to the outer housing (shroud). Thus, four
struts were used to support the centerbody which carried the rotating assem-
bly. The impeller hub diameter was matched to the centerbody diameter to
minimize turbulence, thus reducing turbulence induced noise. The number of
struts does have an integer relationship to the number of impeller blades, so
noise was minimized by having only one blade wake striking a support strut
6-3
-------
ROTOR INLET:
70.0
240.0
72.7
MEAN
R = 8.277
68.1
180.6
105-0
87-3
CTs
I
ROTOR EXIT:
62.8
177-2
MEAN
R = 8.696
71.8 117-9
HUB
R = 5-50
S-76514
Figure 6-1. Fan Velocity Triangles
-------
TABLE 6-2
FAN BLADE PARAMETERS
Streaml i ne
Leading edge radius, in.
Trailing edge radius, in.
Camber angle, deg
Chord along streamline, in.
Thickness, percent
Incidence angle, deg
Stagger angle , deg
Hub
4.0
5-5
45.0
3.48
12.0
-3-8
21.0
Mean
8.3
8.7
19.4
3-77
8-5
1.1
57-3
Tip
11 .0
11.0
11.9
3.84
6.0
-0.8
68.6
Number of blades = 17
6-5
-------
S-76932
Figure 6-2. Effective Stresses (KSI) on Plate at 4000 RPM
6-6
-------
S-76428
Figure 6-3- Effective Stresses (KSI) on Disk at 4000 RPM
6-7
-------
I
00
1200
1000
800
>-
o
LU
Sf 600
CC
400
200
1ST TORSION MODE
1ST BENDING MODE (MAX)
1ST BENDING MODE (MIN)
1000
2000
ROTATION SPEED, RPM
3000
4000
S-76975
Figure 6-4. Blade Vibration Interference Diagram (22.0 In. OD)
-------
at any given instant in time. There is a direct load path from the centerbody
through the struts to the mounting holes. The support struts were aligned with
the average air angle leaving the impeller. There was no possibility of trying
to recover the rotational (swirl) component of velocity, so the struts were
aligned to present a minimum of flow resistance.
Fan designs are summarized on Table 6-3-
TABLE 6-3
FAN DESIGN SUMMARY
Volumetric airflow per fan, cfm
Total pressure rise, in. water
Inlet temperature, F
Number of blades
Tip diameter, in.
Speed , rpm
System Contractor
Aerojet
7,250
4.7
205
17
19.2
2,980
TECO
10,300
3-1
189
17
22.0
2,400
SES
5,440
2.9
217
17
18.85
2,360
Detail Design
Outline dimensions of the Aerojet fan are shown on Dwg No. 605972. In
this case, the fan impeller is directly mounted on the hydraulic motor output
shaft; the bearing assembly is contained within the motor. The hydraulic
motor is a commercial product, developed by Hydraulic Products Incorporated,
Sturtevant, Wisconsin. The motor ratings are: 11 gpm flow at 2800 rpm pro-
ducing 220 ft-lb torque at 2000 psi. All shaft seals are Viton elastomer.
The unit is a standard design except that the output shaft diameter was ground
to 0.8740-0.8742 in. diameter to provide a closer fit with the fan impeller.
The impeller is secured to the shaft using a 0.250 in. square key and a set-
screw. The hydraulic motors were purchased by Aerojet and supplied to
AiResearch for installation on the fan assembly.
Outline dimensions of the TECO and SES fans are shown on Dwg Nos. 605977
and 605982, respectively. Both fans are designed to be belt driven. Slots
were cut in the housing to permit belt clearance. Large diameter, sealed,
grease packed, deep groove ball bearings with maximum separation between the
impeller end and pulley end bearings are provided. Based on a 3000 hr life,
these bearings are designed to support maximum side loads of from 775 lb
applied at the face of the drive shaft step to 225 lb applied 4 in. back of
the fan centerbody (i.e., the end of the 605977-1-1 driveshaft). The bearings
6-9
-------
are lubricated with Unitemp 500A, a high temperature grease, and do not
require additional lubrication during the estimated 3000 hr life. The bear-
ings are mechanically preloaded and locked to preserve the preload and ensure
maximum bearing life.
FABRICATION
The prototype fan impellers were cast using aluminum alloy 356 because of
its excellent casting properties and previous AiResearch experience with the
alloy in similar applications. The impeller was cast using a plaster mold
process and heat treated to the T6 condition. Additional material was pro-
vided in two planes of the hub which could be removed during the balancing
process. The entire housing assembly was fabricated of type 6061 aluminum
alloy and heat treated to the T6 condition.
After balancing was completed, each impeller was oversped by a factor of
1.5 times the maximum rated speed of 3600 rpm, or 5^00 rpm.
The completed Aerojet fan assembly is shown in Figures 6-5 and 6-6; the
TECO fan assembly in Figure 6-7 and the SES fan assembly in Figure 6-8.
TESTING
Three separate tests were run to calibrate the fan assemblies prior to
shipment. Tests consisted of hydraulic motor performance calibrations for the
Aerojet unit and laboratory unit, aerodynamic performance tests of all three
fan assemblies, and a noise test which was performed on the Aerojet and TECO
assemblies. Details of these tests are described below.
Hydraulic Motor Calibration
The HPI hydraulic motor was removed from the Aerojet fan and was calibrated
with hydraulic fluid per MIL-H-5606A. The test setup was as shown in Figure
6-9- The inlet and outlet oil pressure, outlet oil temperature, and the motor
shaft torque were measured at 3000 rpm and the results are shown on Figure 6-10.
The motor torque was found to be a linear function of pressure drop across the
motor and no significant variation with speed was noted.
A similar test was performed to calibrate a Vickers laboratory motor which
was used to drive the TECO and SES fan assemblies. The results of this calibra-
tion are shown in Figure 6-11. The Vickers motor torque is again a linear
function of pressure drop across the motor.
These motor calibrations were used during the subsequent fan performance
tests to establish the shaft power input to the fans.
6-10
-------
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Page 6-13
-------
NAMEPLATE LOCATION
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COUNTER
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FLOW ADJUSTMENT
VALVE
Figure 6-9- Hydraulic Motor Calibration Test Setup
-------
30
25
7 20
o
X
QQ
t 15
10
-4-
NOTES: SPEED - 3,000 RPM
FLOW = 12.70 ± 1.0
OIL - MIL-H-5606
TEMP = 106 ± 5°F
GPM
10
15
20
25
30
PRESSURE DROP LB/SQ. IN. X 10
-2
S-76330
Figure 6-10. Motor Cal ibrat ion—HP I PN M20-90155-01 , SN 200^7
6-22
-------
J5U
2^
20
o
X
EQ
UJ
n
0 1 0
h- ' u
c:
Q
NOTES
RPM =
2500
FI ou = 6.^n r,PM + 0.11
OIL = MIL-H-5606
TEMP =
89°F ± 5°
/
jT
F
/
/
/
/
r
sf
/
/
>
5 10 15 20 25 30 35
S-76323
Figure 6-11. Motor Calibration--Vickers Motor, Model No. 3911-30
6-23
-------
Fan Performance
Fan aerodynamic performance was determined using a setup as shown in
Figure 6-12. Figure 6-13 is a picture of the Aerojet fan installed in the
performance test rig. The duct leading to the fan was designed to be the same
diameter as that at the fan exit. Because the fan exit duct is conical, the
laboratory inlet duct diameter was slightly larger than that at a fan inlet
and, hence, a fairing was used to provide a smooth entrance to the fan. Duct
length from the fan inlet to the flow measuring orifice section was equal to
at least 20 duct diameters to ensure uniform flow at the fan inlet. The fan
discharged directly to atmosphere as is the case in the actual installation.
A hydraulic motor was used to drive the fan. The Aerojet fan used the
HPI motor that is part of the fan assembly. The TECO and SES fans used a
laboratory supplied Vickers hydraulic motor. Both motors were calibrated as
previously described.
During the test, the fan speed and suction pressure were- varied until
the flow and pressure rise as specified by the fan design point conditions
were achieved. This speed was considered to be the actual fan design point
speed. This speed was then held constant and the flow varied from full flow
to an aerodynamic stall condition. The fan static pressure rise and hydraulic
motor flow and pressure drop over the range of air flow was then recorded.
Fan total pressure rise is defined as
where PT - P + P (6-2)
T s v v '
and
c 60 TT D
Equation (6-1) can now be written as
(6-4)
AP_ = P +P -/P +P
If the discharge and inlet fan duct diameters are equal the dynamic
ures, P and P , are identical neglecting th
Vd Vi
density across the fan. Thus, Equation 6-k becomes
pressures, P and P , are identical neglecting the small difference in air
Vd Vi
APT=P - Ps> (6-5)
d i
but, P = P , (6-6)
' s , amb
d
6-2k
-------
ho
\_n
© ©
LABORATORY
AIR SUPPLY
ORIFICE
SECTION
INLET
PAIRING
0
SPEED PICKUP
0 0
HYDRAULIC
MOTOR
0 „
HYDRAULIC
PUMP
UNIT
S-76317
Figure 6-12. Fan Aerodynamic Performance Test Setup
-------
72104-1
Figure 6-13- Aerojet Fan Calibration Test Setup
6-26
-------
and thus, Ap = p - p (6-7)
T amb s .
i
Using the setup shown on Figure 6-12 the fan total pressure rise can be
determined by measuring the inlet static pressure and the barometric pressure.
The fan efficiency is defined as
•n Total Ai r hp Out ,, r,\
''f = Shaft hp in. (6'8)
where Air hp - 1.57 QAP-,- x 10"^ (6-9)
N Mt
and Shaft hp=-' (6-10)
Substituting in Equation 6-9 the following expression for fan efficiency is
obta ined.
(.99) OAP
• (6-11)
f N Mt
Equations 6-7 and 6-11 were used to reduce the test data. The final
performance plots for the Aerojet, TECO, and SES fans are shown on Figures
6-14, 6-15, and 6-16, respectively. The following fan efficiencies were
achieved at the design point flow:
System Contractor Design Point Flow, CFM Efficiency, Percent
Aerojet 7,250 71.2
TECO 10,300 73.5
SES 5,440 70.2
All fans met the 70 percent design goal efficiency.
Noise Tests
Testing was conducted in a chamber which had been treated to reduce back-
ground noise with the fans operating at their respective design point conditions.
Each fan was set up in the chamber in an arrangement similar to its automobile
installation with the inlet restricted and the fan discharging directly to
atmosphere. Figure 6-17 is a sketch of the test setup. Seven points, 45 deg
apart, were established with reference to the fan centerline and each point is
five ft from the fan. These points were the microphone positions for measuring
the sound pressure level generated by the fan. A wind shield was placed over
the microphones to eliminate the noise of the air impinging on the microphone.
6-27
-------
NOTES:
Q;
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FAN VOLUMETRIC FLOW RATE |j- X 10~3 - CFM
o
S-76300
Figure 6-14. Performance of the Aerojet Hydraulic-Motor-Driven Fan,
AiResearch PN 605972-1-1
6-28
-------
CtL
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ED FAN TOTAL PRESSUF
CORRECTED INPUT
FAN EFFICI
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^-—
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- 1. OPERATING
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where P = I
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SPEED Z^tOO RP
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x
-------
CORRECTED FAN TOTAL PRESSURE RISE (N/N )
T
2 —
CORRECTED NPUT HORSEPOWER (N/N )3
iNCHES OF WATER
HP
ON
in
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FAN EFFICIENCY £ ("H )
-------
MICROPHONE POSITION (TYP)
0
FAN ASSEMBLY
os
-©--
FLOW
•5 FT.-
\
NOISE ABSORBING
STRUCTURE
s-76311*
Figure 6-17- Fan Noise Test Setup
-------
1. Aerojet
Figure 6-18 presents the automatic recordings of the sound pressure level
obtained for the Aerojet fan. Figure 6-19 shows the measurement of the back-
ground noise in the chamber indicated by sound pressure levels below 200 Hz.
The average overall sound pressure five ft from the fan was 88.5 db(A). This
value includes the noise generated by the HPI hydraulic motor.
The HPI M20-90155-01 hydraulic motor for the Aerojet fan was removed from
the fan and connected to a dummy load as shown in Figure 6-20. Microphones
were placed five ft from the hydraulic motor as shown in Figure 6-20. The
motor was operated at 2820 rpm and 1930 psi Ap (9.75 hp) which is the approxi-
mate fan shaft horsepower at rated load conditions. A background noise level
test was also run with the motor stopped. Test results are presented in
Figure 6-21. The background noise and the inconsistency between the 1/3
octave and full octave analyzer indicates that the noise below 100 or possibly
200 Hz, was background noise and not caused by the motor. A comparison between
the fan noise and motor noise shows that the dominant fan noise was in the
100 Hz octave band, at which the motor noise was quite low, indicating the
motor was not a significant contributor to the fan noise. In all octave bands,
the motor noise was at least five db lower than fan noise, indicating that the
motor did not contribute more than approximately 1-1/2 db to fan noise in any
octave band. The principal noise source was aerodynamic noise as generated by
the fan.
2. Thermo Electron
The installation of the TECO fan was as shown on Figure 6-17- Figure
6-22 is a picture of the installation; the microphone positions are indicated
by X's on the floor. A laboratory Vickers hydraulic motor was used to drive
the fan. The motor was connected to the fan via a drive shaft containing two
universal joints.
Test results for seven microphone positions are shown on Figure 6-23-
The automatic sound recorder was inoperative and the data were recorded by
hand. In this case the noise of the hydraulic motor, which was previously
calibrated as described for the Aerojet fan, was analytically subtracted from
the measured overall values. The average sound pressure level for the fan
alone was 88.6 db(A) at five ft away, operating at design point conditions.
3. Steam Engine Systems
The three EPA fans utilize the same basic impeller and differ only in the
fan tip diameter. Thus, the fans could be expected to exhibit similar noise
characteristics. Theoretically, the Aerojet fan should be 1.1 db noisier than
the TECO fan based on the following relationship.
(CFM)A , ^ , APA
6-32
-------
' POSITION NO.
Figure 6-18. Sound Pressure Levels for Aerojet Fan Assembly P/N 605972-1
6-33
-------
.50 |DB'
1/3 OCTAVE
FULL OCTAVE
(1B11/2III) A B C Lja
FULL OCTAVE
S-767W
Figure 6-18 (Continued)
6-35
-------
(INI/2111) A B C Ua
70 DB
900 1000
FULL OCTAVE
FULL OCTAVE
(1BH/2JI1) A B C Lin.
S-76?
Figure 6-18 (Continued)
6-37
-------
50 DB
1/3 OCTAVE
S-767M
Figure 6-18 (Continued)
6-39
-------
EPA FAN ASSEMBLY P/N 605972-1
1/3 OCTAVE
(1811/aiO A a c IA 50 DB
den/2111) A B C LJI.
FULL OCTAVE
S-767W
Figure 6-19. Reference Background Sound Pressure Level for
Aerojet Fan Assembly P/N 605972-1
6-itl
-------
SENSOR POSITION (TYP)
SEMI-ANECHOIC
CHAMBER
LOAD PUMP
FLOW ADJUSTMENT
VALVE
VARIABLE FLOW
AND PRESSURE
HYDRAULIC
SUPPLY CART
S-76291
Figure 6-20. Hydraulic Motor Sound Pressure Level Test Setup
-------
Where the subscripts A and T refer to Aerojet and TECO, respectively. Actual
measurements were
Aerojet = 88.5 db(A)
TECO = 88.6 db(A)
Adb(A) = -0. 1
This difference is well within the experimental error of ± 1 db on each
measurement and thus, it was concluded that the two fans obey the theory.
In view of the low sound pressure levels achieved and the conformance of
the fans to the fan lows it was decided not to test the SES fan, but to calcu-
late its sound pressure level based on the Aerojet test results. Based on the
above expression for change in noise level, the calculated noise of the SES
fan is 83.2 db.
4. Summary
Test results are summarized on Table 6-4. Assuming a spherical radiation
pattern, the sound pressure levels at 50 ft would be 20 db lower than at five
ft. To determine the noise level for two fans as is the case for the actual
installation, 3 db should be added to the noise values shown on Table 6-4.
Fan specifications do not include a specific noise limitation. However,
a noise standard is imposed on the complete automobile and a maximum level of
77 db(A) at 50 ft is cited. Since the fans are expected to be the major noise
source on the vehicle, it would appear that the vehicle noise standards can be
met.
TABLE 6-4
NOISE TEST RESULTS
System
Contractor
Aerojet
TECO
SES
Rotational
Speed
2980
2400
2360
Sound Pressure Level
Overall Average
at 5 ft, dbA
88. 5 *
88.6 *
83.2 **
Sound Pressure Level
Overall Average
at. 50 ft, dbA
68.5 **
68.6 **
63.2 *-;,-
•"Measured
-'.-'-Calculated
NOTES: (1) All data are for a single fan
(2) Calculated values at 50 ft assume a spherical radiation pattern
-------
20 Hz 50 100 200 500 1000 2000 5000 10000 20000 4
requency Scale by Zero Level 50- (1612/2IT2)
1/3 OCTAVE
5000 10000 20
(1612/2112)
J I I
SENSOR POSITION
Hz 50 100 200 500 1000 2000 5000 10000 ""jllUUU tmfltfD ABC
Tequency Scale by • Zero Level 60 (1612/2112) A B C Lm
OCTAVE
zl SENSOR POSITION 2
3 50
W^ Hi 50 100 200
. frequency Scale by
500 1000 2000
Zero Level 50
Hz 50 100
requency Scale by
Figure 6-21. Sound Pressure Level for HP I Hydraulic Motor M20-90155-01 , S/N £200^7
6-45
-------
SENSOR POSITION 3 '-I
Hz 50 100 200 500 1000 2000 5000 10000 20000T0BBO D A B C
Vequency Scale by Zero Level 50' (1612/2112) A B C Lin
1/3 OCTAVE
J-J-lUL-f-J
6000 10000 20000 <000t D A B C t
t6 Hz 50 100 200
requency Scale by
OCTAVE
500 1000 2000
Zero Level 50
r i—\ ' SENSOR POSITION 4 ~r^-_-
Hz 50
-requency Scale by
500 1000
Zero Level 50 '
5000 10000 201
(1612/2112)
" D A B C
ABC Lin.
20 Hz 50 .100
:requency Scale by
500 1000 2000
Zero Level 50
5000 10000 2000X) 4000B D ABC
(1612/2112) A B C Lin
Figure 6-21 (Continued)
6-4?
-------
SENSOR POSITION 5
2fT Hz 50 100
•requency Scale by
500 1000 2000
Zero Level 50
1/3 OCTAVE
Hz 50 100
-requency Scale by
Zero Level 50
Hi 50 100 200
requency Scale by
OCTAVE
500 1000 2000
Zero Level 50
-requency Scale by
0000 jfll!IA> TllBlJ D A B C Mr
(1612/2112) ABC Lin.
Figure 6-21 (Continued)
s-76737
6-49
-------
/2168-1
Figure 6-22. TECO Fan Assembly Installed in Noise Test Setup
6-51
-------
\j-\
NJ
100
90
21.99 DIA
17 BLADES
2400 RPM
STRUTS
60
80 10KC
S-7&5S8
Figure 6-23. Sound Pressure Level for Thermo Electron
Fan Assembly P/N 605977-1
-------
SECTION 7
CONDENSER AND FAN AIRFLOW TEST
-------
SECTION 7
CONDENSER AND FAN AIRFLOW TEST
Testing was conducted to determine isothermal airflow rate as a function
of fan speed for both the TECO and SES condenser and fan assemblies. The pur-
pose of these tests was to obtain an estimate of actual airflow to the con-
denser when the condenser and fan are joined as a single assembly. Thus, the
results include the effect on performance of any flow maldistribution, both
in the condenser core and at the fan inlet face, due to the close-coupling of
these two components.
Figure 7-1 is a schematic of the test setup showing the airflow system
and the location of the condenser and fans. Figures J-2 and 7-3 show the
completed setup with the TECO and SES assemblies installed. To ensure uniform
air flow velocity across the condenser inlet face, two perforated screens (50
percent open area) were installed in the rectangular air inlet duct and a
bellmouth contour was located just upstream of the condenser inlet face. The
bellmouth contour was adjusted as required to conform to the front face
dimensions of TECO and SES condensers. Views of the bellmouth contour (SES
configuration) and one of the perforated screens are shown on Figures 7-^ and
705. The two thermocouples which were used to record the inlet air tempera-
ture are also shown in Figures 1-k and 7~5-
A single hydraulic motor was used to drive both fans as shown on Figures
7-6 and 7~7- A spur gear was incorporated on each fan shaft to generate pulses
for the electronic speed pickup. During testing the speed difference between
the two fans was limited to less than 1.0 percent at the maximum power input
cond it ions.
The tests were run by varying the rotational speed of the fans and adjust-
ing the air flow at the air system inlet to obtain the desired static pressure
upstream of the condenser. In most cases, the upstream pressure was main-
tained at zero gage pressure. It was necessary to maintain at least zero
gage pressure at the condenser inlet to prevent the fan from operating in a
surge condition. With the SES unit, runs were also made at 1.0 and 2.0 in.
hLO positive pressure at the condenser inlet face.
Inlet air pressure was measured with eight static wall taps located as
shown in Figure 7-8, and airflow was measured with a standard orifice section
at the air inlet. The air was dumped directly to ambient at the fan outlet
face. During testing, the maximum deviation in any one inlet air static pres-
sure reading from the average of the eight measured values was less than 0.1
in. H20.
Figures 7-9 and 7-10 summarize the test results for the two fan/condenser
assemblies. In both cases, airflow is a linear function of fan rotational
speed. The points run with a positive inlet pressure to the SES condenser
show virtually the same results as those plotted in Figure 7-9.
7-1
-------
00
FACILITY
AIR FLOW
FLOW
CONTROL
VALVE
ORIFICE
SECTION
BELLMOUTH
ENTRY
FLOW
DISTRIBUTION
SCREENS
THERMOCOUPLE
(TYPICAL,
2 PLACES)
STATIC
PRESSURE
TAP (TYPICAL,
8 PLACES)
FAN
SPEED
PICKUP
5-761*69
Figure 7-1. Condenser and Fan Airflow Test Setup
7-2
-------
-2
Figure 1-2. TECO Assembly Installed in Test Rig
7-3
-------
72M32-2
Figure 7-3- SES Assembly Installed in Test Rig
-------
Figure 1
Duct
Configuration at Condenser Inlet
7-5
-------
72381-2
Figure 7-5. Duct Configuration at Condenser Inlet
7-6
-------
to
c
CD
C
o
rt
O
in
3
o
—I
<
-------
c
(0
o
4->
o
(0
1_
-o
I—.
I
13
CT)
CO
-------
-•4
VD
BELL
MOUTH
2
o
PRESSURE TAPS (1-8)
3
o
1
o
5
o
o
8
o
7
o
6
k.O IN.
TAP LOCATION
FROM HX FACE
VIEW TAKEN DOWNSTREAM
OF FANS. LOOKING FORWARD
Figure 7~8. Pressure Tap Locations
-------
1400
1200
1000
800
600
400
200
CONDENSER INLET PRESSURE = 0.0 PS IG
AIR TEMPERATURE = 108°F
1000
1500
2000
FAN RPM
2500
3000
Figure 7~9- SES Condenser and Fan Airflow Test Results
-------
CONDENSER INLET PRESSURE = 0.0 PSIG
AIR TEMPERATURE = 98F
200
1000
1500
2000
FAN RPM
2500
3000
S-76^82
Figure 7-10. TECO Condenser and Fan Airflow Test Results
-------
The predicted system performance, as represented by a dashed line on
Figures 7-9 and 7-10, was determined as follows: The fans were assumed to
perform in accordance with their respective calibration curves (see Figures
6-15 and 6-16). This is not entirely correct since any flow maldistribution
at the fan inlet would affect performance. The fan 1aws were then used to
determine the fan flow and pressure rise at speeds differing from that of
the calibration curve. It was assured that one velocity head was lost at
the fan exit. The pressure drop across the condenser was based on the pre-
dicted condenser air side pressure drop as presented in Section 5- Finally,
for a given fan speed, the airflow at which the fan static pressure rise was
equal to the pressure drop across the condenser was determined.
The airflow in the SES unit is seen to be slightly higher than predicted,
whereas the TECO airflow is nine percent less than predicted over the rpm
test range. The discrepancy between predicted and test performance for the
TECO unit is believed to be primarily due to flow distribution effects caused
by the transition between core outlet face and fan inlets. The power per-
formance of the TECO unit relative to the SES unit is probably due to (1) a
closer spacing between condenser and fan (1.9^ in. for TECO versus 2.51 in.
for SES) which restricts the redistribution of flow in the transition plenum,
and (2) the slight oversizing of the TECO fans relative to the condenser face
(i.e. , the fan diameter of 22 in. is about one inch greater than the core
height) which tends to restrict flow to the tops of the TECO fans where they
overhang the core.
Based on Figure 7-10, it established that for the TECO unit the required
fan horsepower is increased by 27 percent over the previous prediction at
design point airflow conditions. The estimated power requirement for the SES
fans is unchanged based on the close correlation between test flow and pre-
diction shown by Figure 7-9.
7-12
-------
SECTION 8
INSTALLATION AND AIRFLOW TEST
-------
SECTION 8
INSTALLATION AIRFLOW TEST
Cooling air flow through a Rankine cycle engine condenser is about 3 to
5 times greater than that through a conventional 1C automobile engine radiator.
It would be expected then, that the air flow pressure losses across the bumper
and grille and through the engine compartment, referred to as the installation
loss, could easily become excessive in the case of a Rankine cycle engine (RCE).
Installation pressure losses are of considerable interest because they
directly affect the fan design. The system contractors were contacted to
determine how they evaluated the engine installation air flow losses. Thermo
Electron determined the losses by assigning an assumed loss coefficient to
each element in the system. Steam Engine Systems assumed the losses were
equal to three condenser exit velocity heads. Aerojet assumed that the total
losses were equivalent to 25 percent of the ram air velocity head. The assumed
installation losses ranged from 0.3 to 2.0 in. HLO. It was apparent that
there was no reliable test data available from which to compute the installa-
tion loss.
An electric analog study was performed to obtain an estimate of the
losses in the TECO installation. This study indicated that, due to the limited
flow area in the engine compartment, a high compartment pressure loss could be
expected.
Test data for the pressure losses through the grille, air conditioning
condenser, radiator, fan shroud, and engine compartment of a 1971 Galaxie as
a function of air throughflow was obtained from Ford (Reference 6-1). The
data was obtained in a wind tunnel, and as indicated by Ford, does not correlate
accurately with driving conditions. However, it should give an idea of the
magnitude of the actual losses. Using loss coefficients calculated from the
Ford data, the pressure drops through the Galaxie grille, air conditioning
condenser, and the Galaxie engine compartment were calculated at the TECO
design air flow rate to be 3-6, 1.3, and 8.8 in. HLO respectively. The total
loss assigned to the installation by TECO was about 2.8 in. H20. Thus, it
would appear that installation of the TECO engine in an unmodified engine
compartment would result in excessive fan power consumption.
Because of the potential for a significant impact on the fan design, a
test program was undertaken to determine the actual loss. It was expected
that their data would help to guide the system contractors in modification of
their respective systems to provide greater cooling air flow area which would
in turn, result in a reasonable limit on fan power consumption.
The basic test plan was to use an actual engine compartment (obtained from
a junked automobile) and install wood and Styrofoam mockups of the RCE in the
compartment. A known amount of air would be flowed through the grille and
across the bumper into the compartment. The resultant pressure drops would be
recorded. Results of these installation air flow tests are described below.
8-1
-------
Thermo Electron/Ford Installation
The front end of a 1971 Ford Galaxie was obtained and a mockup of the
Thermo Electron Engine was installed. Photographs of the completed installa-
tion are shown in Figures 8-1, 8-2, and 8-3.
Figure 8-1 shows the condenser mockup which for purposes of this test,
was constructed of 2 in. diameter tubes which acted as flow straighteners.
The 58 in. wide rectangular condensers stretched across the entire width of
the engine compartment. Condenser flow resistance was designed to be simulated
by layers of tight mesh screen. The two annular openings just downstream of
the condenser represent the fan blade sweep area. The frame rails were modified
to go along the side of the condenser which required locating the bumper sup-
ports outboard of the condenser.
Figure 8-2 presents a clear view of the rectangular vapor generator which
is located immediately behind the fans. Figure 8-3 shows the close spacing
(about 3 in.) between the fans and the vapor generator.
An adaptor plate as shown on Figure 8-4 was made to connect the vehicle
inlet to the laboratory air duct. The adaptor also included a simulated
bumper. The real bumper was not used because the supports were relocated as
noted above. As shown on Figure 8-5, the adapter plate was fitted to cover
the entire area between the lower lip of the head and the bottom of the simu-
lated bumper between the fender projections. The vehicle springs were col-
lapsed until the vehicle was at a height equivalent to that of a normally
loaded automobile. A rectangular plywood duct was fitted to the adapter plate
to supply air flow to the condenser. The airflow was supplied by two engine-
driven blowers, and flow measurement was accomplished by calibrated bellmouth
inlets fitted to the blowers. The completed test setup is shown on Figure 8-6.
Test results obtained for the TECO installation are presented on
Figure 8-7- The overall loss at the design point conditioner is about 3.0 in.
HO. This installation loss includes losses across the bumper and through the
engine compartment, but does not include that of the grille because it was
removed for their test. The loss across the condenser mockups was assumed to
be zero (no flow resistance screens were used during this test).
Attempts were made to measure engine compartment and simulated bumper
losses separately. It was found, however, that there was no representative
location at which a static pressure could be measured between the bumper and
the compartment. Thus, an average volumetric flow must be used to estimate
the overall installation pressure drop.
According to the Thermo Electron fan problem statement summarized in
Table 6-1, 2.8 in. HO pressure loss is allowed for the installation. The
measured loss with no grille is about 3-0 in. H20. It would appear that the
specified installation loss could be met with some vehicle modifications and
a non-restrictive grille. This test was run, however, at a condition of zero
vehicle speed. It would be expected that with vehicle motion, the airflow
exit from the engine compartment would be restricted because of the airflow
underneath the vehicle. This effect would tend to increase the compartment
pressure drop.
8-2
-------
Q.
13
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c
O>
c
Ul
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i
oo
Ol
Ll_
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71^01-5
Figure 8-3. Thermo Electron Engine Mockup
8-5
-------
Figure 8-k. Adaptor Plate with Simulated Bumper
8-6
-------
ADAPTOR
PLATE
SEAL
VAPOR GENERATOR MOCKUP
AIR
FLOW
S-76238
SIMULATED
BUMPER
CONDENSER
MOCKUP
Figure 8-5. Air Flow Ducting
8-7
-------
Figure 8-6. Thermo Electron Installation Air Flow Test Setup
8-8
-------
AP
OVERALL CORRECTED STATIC PRESSURE
in. H20
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-------
Subsequent to this test, Thermo Electron completely redesigned the vapor
generator. The resulting cylindrical configuration provides additional flow
area at the fan exit. It would be expected that the engine compartment pres-
sure drop would be less with the new vapor generator configuration.
Aerojet/Chevrolet Installation
The compartment mockup was loaned to AiResearch by the Aerojet Liquid
Rocket Comapny for purposes of this test. The equipment consisted of the
engine compartment, bumper, grille, hood, and front running gear of a 1971
Chevrolet Impala fitted with Styrofoam mockups of the major components of the
Aerojet engine. Two different vapor generator configurations were supplied.
The first mockup was 26 in. in diameter and represented the current configura-
tion, while the second mockup was basically 23 in. in diameter and was con-
figured to show an advanced vapor generator design.
The mockup was modified by AiResearch to position two fan housings ahead
of the engine. This necessitated the complete removal of the front main sup-
port panel and radiator brackets. The two front shock absorbers were replaced
with tie down rods to compress the coil spring suspension and position the car
at the proper height above ground level. Plywood panels were added to ensure
that all condenser air entering the grille passed through the fan housing
mockup. A condenser mockup was not used in the installation because its pres-
sure drop was known and the lack of a condenser would not affect the determina-
tion of the installation pressure drop.
A rectangular duct the height of the bottom of the car bumper and
extending approximately 2 ft either side of the car was terminated 3 ft in
front of the car tires as shown on Figure 8-8. The purpose of this duct was
to supply air to the under-side of the car to simulate vehicle motion. The
duct was fitted with flow stra ighteners and a flow splitter to promote uniform
air velocity across the discharge of the duct.
The duct was supplied with low pressure air. The discharge air velocity
was controlled so that the discharge air velocity across the duct was main-
tained within 3 mph of the desired value.
A second rectangular duct was fitted on top of the lower duct to supply
the fan airflow. This duct was fitted to cover the entire area between the
lower lip of the hood and the bottom of the bumper between the fender projec-
tions, as shown in Figure 8-8. The airflow for this duct was supplied by two
engine-driven blowers, and flow measurement was accomplished by calibrated
bellmouth inlets fitted to the blowers.
Instrumentation consisted of (1) four total-static pressure probes
located approximately 3 ft from the vehicle inlet in the upper duct, (2) two
static pressure taps, two total-static pressure probes, and a thermocouple
located 6 in. from the vehicle inlet in the upper duct, and (3) a total-static
traverse probe located at the discharge of the lower duct. Figure 8-9 is a
photograph of the partially completed installation.
8-10
-------
FLOW STRAIGHTENER
FLOW SPLITTER
CO
i
HOOD REMOVED
FOR CLARITY
FAN HOUSING
MOCKUP-
AIRFLOW
UPPER DUCT SECTION
FOR CONDENSER AIRFLOW
LOWER DUCT SECTION
FOR UNDER-CAR VELOCITY
CAR MOCKUP
PITOT-STATIC TRAVERSE PROBE
TO MEASURE AIR VELOCITY
AT DUCT DISCHARGE
S-6<>589 -A
Figure 8-8. Aerojet Installation Airflow Test Setup
-------
71877
Figure 8-9. Aerojet Installation Air Flow Test Setup
8-12
-------
Attempts were made to measure the pressure drop across the grille and
bumper directly, but no representative location could be found for the
condenser-side pressure pickup to ensure accurate readings. Accordingly,
the simulated engine components were removed from the engine compartment, and
the engine compartment to atmosphere pressure drop and the pressure drop
from just upstream of the bumper and grille to atmosphere were measured at
,simulated condenser flows of 10,880, 14,500, and 20,600 cfm. In all cases,
the engine compartment pressure drop was less than 0.3 in. WC at maximum
flow, no matter where measured.
Therefore, to simplify testing, the pressure drop of the empty engine
compartment was assumed to be zero and the entire pressure drop from in
front of the bumper and grille to atmosphere was assigned to the bumper and
grille.
The grille and bumper pressure drop was run with no lower duct flow to
simulate vehicle motion. The simulated engine components were reinstalled in
the engine compartment and the pressure drop from just upstream of the grille
to atmosphere was again measured at simulated condenser flow rates of 10,880,
14,500, and 20,600 cfm with lower duct discharge velocities of 0. 20, 40, and
60 mph. The bumper and grille pressure drop was assumed to be independent of
lower duct velocity, so the bumper and grille pressure drop found earlier was
subtracted from the pressure drop with under-car airflow and the difference
was taken as the engine compartment pressure drop.
Since the pressure drop was quite high during the original bumper and
grille pressure drop tests, the two inboard headlights were removed, result-
ing in a substantial reduction in the bumper and grille pressure loss. The
engine compartment loss was determined for both the 23-in. and 26-in. dia
vapor generator configurations.
The test data are summarized in Figure 8-10. General conclusions based
on these test results are as follows:
(a) The grille on the test car had a high pressure drop, and relatively
minor increases in open area (removal of 2 headlights) produced a
very significant reduction in pressure drop.
(b) The change from a 23-in. dia vapor generator to a 26-in. dia vapor
generator produced significant changes in engine compartment pres-
sure drop. This suggests that the engine compartment is quite
sensitive to relatively minor changes in size of components and
quite possible to their placement as well.
(c) The effect of under-car velocity on the engine compartment pres-
sure drop was relatively minor. The compartment pressure drop
increased by only 15 percent when the under-car velocity was
increased from zero to 60 mph.
8-13
-------
Q-
-------
A comparison between the Aerojet fan specification requirements as outlined
on Table 6-1 and the requirements calculated on the basis of the present test
results is shown in Table 8-1. The comparison is made for a 23-in. dia vapor
generator engine configuration with two headlights removed to provide additional
airflow area. The test results indicate that with this configuration, about a
23 percent increase in fan input power would be required to overcome the
increase in the specified installation loss. The increase in fan input power
can be minimized, however, by further modifications to the vehicle grille and
bumper to provide additional inlet airflow area.
8-15
-------
TABLE 8-1
AEROJET INSTALLATION DESIGN POINT COMPARISON
Condenser airflow, Ib per hr
Condenser inlet temperature, F
Condenser inlet pressure, psia
Condenser inlet volume flow, cfm
Bumper and grille static pressure loss,*
in. H20
Fan outlet temperature, F
Fan outlet volume flow, cfm
Vehicle speed, mph
Engine compartment static pressure loss,
26- in. dia vapor generator, in. hLO
Engine compartment static pressure loss,
23- in. dia vapor generator, in. hLO
Fan exit velocity head, in. hLO
Installation total pressure loss,** in. hLO
Condenser pressure loss, in. hLO
Ram pressure rise, in. hLO
Fan total pressure rise, in. hLO
Spec! f i cat ion
52,000
85
14.7
11 ,900
206
14,5^0
50
-
™
-
1.67
3.91
0.91
4.67
Performance
Based on
Test Results
52,000
85
14.7
11 ,900
2.04
206
14,540
50
0.89
0.46
0.56
3.06
3.91
0.91
6.06
*2 headlights removed
**Total loss = bumper and grille plus engine compartment (23-in. dia vapor
generator) plus fan velocity head.
8-16
-------
SECTION 9
REFERENCES
-------
SECTION 9
REFERENCES
1. Wong, S. et al, Final Report, Compact Condenser for Rankine Cycle Engine,
AiResearch Report 71-7^64, August 1971.
2. Soliman, M., Schuster, J., and P. Berenson, "A General Heat Transfer
Correlation for Annular Flow Condensation", Journal of Heat Transfer,
Transactions of the ASME, May 1968.
3. Personal Communication from Bill Macauly, Ford Motor Company, Dec 1971-
9-1
-------
APPENDIX
CONDENSER AND FAN ASSEMBLY SPECIFICATIONS
-------
APPENDIX
SPECIFICATION
CONDENSER AND FAN ASSEMBLY
PART NUMBER 190390
AEROJET LIQUID ROCKET COMPANY
AUTOMOTIVE PROPULSION SYSTEM
Prepared for the
Environmental Protection Agency
Division of Advanced Automotive Power
Systems Development
Ann Arbor, Michigan 48105
Contract 68-01-0407
GENERAL
This specification defines the requirements for a condenser and fan
assembly to be utilized on a low emission, Rankine cycle automotive pro-
pulsion system. The assembly is designed to operate on a pre-prototype
engine which will be tested for 150 hours on a dynamometer test setup.
The condenser and fans have been designed to fit within the engine com-
partment, although the assembly is not considered to be a fully readable
des ign.
CONDENSER PERFORMANCE
Design point performance goals are as follows:
Vapor Side
Working fluid AEF-78
Total heat rejection, Btu/hr 1.50 x 10
Flow, Ib/hr 20,000
Inlet temperature, °F 241
Inlet pressure, psia 32.7
Condensing temperature (avg), °F 235
Condensing pressure (avg), psia 31.3
Outlet temperature, °F 192.5
Subcooling, °F 38.5
Core pressure drop, psi 2.8
Overall pressure drop, psi 3.1
A-1
-------
Air S ide
Flow, Ib/hr 52,000
Inlet temperature, °F 85
Inlet pressure, psia 14.7
Outlet temperature, °F 205
Temperature effectiveness 0.80 Condenser
0.62 Subcooler
Overall core pressure drop, in. hLO 4.0
2.I Condenser Design
The condenser shall be a plate-fin crossflow design utilizing bar and
plate, brazed and welded construction. The material of construction
shall be aluminum alloy. The unit shall be fluxless brazed to avoid
contamination in the vapor passages.
3. FAN PERFORMANCE
Design point performance goals are as follows:
Total air flow, Ib/hr 52,000
Total no. of fans 2
Inlet temperature, °F 205
Inlet pressure, psia 14.7
Airflow per fan, cfm 7250
Total pressure rise, in. H?0 4.7
Efficiency, percent 70
Shaft power input, hp 7.5
3.I Fan Design
The fan shall be a tube-axial type of the following configuration:
Tip diameter, in. 19.2
Number of blades 17
Fan rotation as viewed from driver's seat ccw
Maximum operation speed, rpm 3600
Material Aluminum alloy
A-2
-------
3.2 Fan Drive
The fan shall be designed to be directly driven by a hydraulic motor.
Interface requirements as they pertain to the hydraulic motor shall be
as specified in AiResearch Source Control Drawing 499-001. The motor
shall be supplied by others.
3.3 Noise Standards
No specific limit on the noise generated by the condenser and fan assembly
is specified. The fans shall be designed, however, to produce a minimum
amount of noi se.
4. INSTALLATION
Overall dimensions, mounting arrangements, provisions for differential
thermal expansion, inlet and outlet port configuration, and weight shall
be as shown on Outline Drawing 190390.
4.I Air Flow Characteristics
Condenser and fan design point airflow rating of 52,000 Ib/hr shall be
based on the following installation characteristics specified by the
system contractor:
Vehicle speed, mph 50
Inlet flow dynamic pressure, in. hLO 1.17
Inlet recovery factor, percent 78
Net ram pressure rise, in. h^O 0.91
Installation pressure loss#, in. H20 1.67
5. STRUCTURAL DESIGN
The condenser shall be designed to withstand a maximum operating pressure
of 50 psig at a metal temperature of 300°F. The corresponding proof
pressure shall be 85 psig at room temperature. The condenser, fans and
the support structure shall be designed to withstand the stresses associated
with a bench test of the pre-prototype engine.
6. FINISH
All external surfaces of the condenser and fan assembly, with the exception
of the air-side fins and the fan impellers, shall be painted in accordance
with AiResearch Process Specification FP-73F04 (epoxy paint, flat black color!
•"•Installation loss includes the pressure losses as a result of airflow
across the vehicle bumper and grille, and the losses associated with
the fan discharge airflow through the engine compartment.
A-3
-------
7. ACCEPTANCE TEST
7.I Proof Pressure
The condenser shall be pressurized to 85 psig at room temperature.
After pressure has been released, there shall be no evidence of permanent
deformat ion.
7. 2 Leakage
After completing the proof pressure test, the condenser assembly shall be
leak checked. Leakage shall be less than I x 10 sec/sec of helium with
the interior of the condenser evacuated and the exterior surfaces surrounded
by helium at one atmosphere pressure.
7.3 Fan Ca1i brat ion
The fan shall be tested as a component to establish its performance. The
fan shall be run at the design point speed, and the fan pressure rise
shall be determined over a range of airflows.
7.4 Condenser and Fan Performance
The condenser and fans shall be tested as an assembly to establish the
airflow characteristics of the system. The assembly shall be tested
without any inlet or outlet flow obstructions, and the airflow through
the system shall be determined over a range of fan operating speeds
including the design point speed.
A-4
-------
SPECIFICATION
CONDENSER AND FAN ASSEMBLY
PART NUMBER 190370
THERMO ELECTRON CORPORATION
AUTOMOTIVE PROPULSION SYSTEM
Prepared for the
Environmental Protection Agency
Division of Advanced Automotive Power
Systems Development
Ann Arbor, Michigan 48105
Contract 68-01-0407
I . GENERAL
This specification defines the requirements for a condenser and fan
assembly to be utilized on a low emission, Rankine cycle automotive pro-
pulsion system. The assembly is designed to operate on a pre-prototype
engine which will be tested for 150 hours on a dynamometer test set-up.
The condenser and fans have been designed to fit within the engine com-
partment, although the assembly is not considered to be a fully readable
des i gn.
2. CONDENSER PERFORMANCE
Design point performance goals are as follows:
Vapor Side
Working fluid Fluorinal-85
Total heat rejection, Btu per hr 1.88 x 10
Flow, Ib per hr 9860
Inlet temperature, °F 238
Inlet pressure, psia 40.0
Condensing temperature (avg), °F 212
Condensing pressure (avg), psia 36.4
Outlet temperature, °F 193
Subcooling, °F 17
Core pressure drop, psi 2.6
Overall pressure drop, psi 5.0
A-5
-------
Ai r Side
Flow, Ib per hr 75,300
Inlet temperature, °F 85
Inlet pressure, psia 14.7
Outlet temperature, °F 189
Temperature effectiveness 0.80
Overall core pressure drop, in. I-LO 4.1
2.I Condenser Des ign
The condenser shall be a plate-fin crossflow design utilizing bar and
plate, brazed and welded construction. The material of construction
shall be aluminum alloy. The unit shall be fluxless brazed to avoid
contamination in the vapor passages.
3. FAN PERFORMANCE
Design point performance goals are as follows:
Total airflow, Ib 75,300
Total no. of fans 2
Inlet temperature, °F 189
Inlet pressure, psia 14.7
Airflow per fan, cfm 10,300
Total pressure rise, in. H?0 3.1
Speed, rpm 2640
Efficiency, percent 70
Shaft power input, hp 7.2
3.I Fan Design
The fan shall be a tube-axial type of the following configuration:
Tip diameter, in. 22.0
Number of blades 17
Fan rotation as viewed from driver's seat ccw
Maximum operation speed, rpm 3600
Material Aluminum alloy
A-6
-------
3.2 Fan Drive
The fan shall be designed to be belt driven. Interface requirements
shall be as specified on Outline Drawing 190370. Maximum load imposed
on the fan bearing assembly by the belt drive during any operating con-
dition shall be limited to 600 Ib.
3.3 Noise Standards
No specific limit on the noise generated by the condenser and fan assembly
is specified. The fans shall be designed, however, to produce a minimum
amount of noi se.
4. INSTALLATION
Overall dimensions, mounting arrangements, provisions for differential
thermal expansion, i-nlet and outlet port configuration, and weight shall
be as shown on Outline Drawing 190370.
4.I Airflow Characteristics
Condenser and fan design point airflow rating of 75,300 Ib per hr shall
be based on the following installation characteristics as specified by
the system contractor:
Vehicle speed, mph 90
Inlet flow dynamic pressure, in. hLO 3.8
Inlet recovery factor, percent 100
Net ram pressure rise, in. hLO 3.8
Installation pressure loss*, in. H?0 2.8
5. STRUCTURAL DESIGN
The condenser shall be designed to withstand a maximum operating pressure
of 100 psig at a metal temperature of 300°F. The corresponding proof
pressure shall be 183 psig at room temperature. The support structure
shall be designed to withstand the stresses associated with a bench test
of the pre-prototype engine.
6. FINISH
All external surfaces of the condenser and fan assembly, with the exception
of the air-side fins and the fan impellers, shall be painted in accordance
with AiResearch Process Specification FP-73F04 (epoxy paint, flat black color'
^Installation loss includes the pressure losses as a result of airflow
across the vehicle bumper and grille, and the losses associated with
the fan discharge airflow through the engine compartment.
A-7
-------
7. ACCEPTANCE TEST
7.I Proof Pressure
The condenser shall be pressurized to 183 psig at room temperature.
After pressure has been released, there shall be no evidence of permanent
deformat ion.
7.2 Leakage
After completing the proof pressure test, the condenser assembly shall be
leak checked. Leakage shall be less than I x 10 sec/sec of helium with
the interior of the condenser evacuated and the exterior surfaces surrounded
by helium gas at one atmosphere pressure.
7.3 Fan Cali brat ion
The fan shall be tested as a component to establish its performance. The
fan shall be run at the design point speed, and the fan pressure rise
shall be determined over a range of airflows.
7.4 Condenser and Fan Performance
The condenser and fans shall be tested as an assembly to establish the
airflow characteristics of the system. The assembly shall be tested with-
out any inlet or outlet flow obstructions, and the airflow through the
system shall be determined over a range of fan operating speeds including
the design point speed.
A-8
-------
SPECIFICATION
CONDENSER AND FAN ASSEMBLY
PART NUMBER 190640
STEAM ENGINE SYSTEMS
AUTOMOTIVE PROPULSION SYSTEM
Prepared for the
Environmental Protection Agency
Division of Advanced Automotive Power
Systems Development
Ann Arbor, Michigan 48105
Contract 68-01-0407
GENERAL
This specification defines the requirements for a condenser and fan
assembly to be utilized on a low emission, Rankine cycle automotive pro-
pulsion system. The assembly is designed to operate on a pre-prototype
engine which will be tested for 150 hours on a dynamometer test set-up.
The condenser and fans have been designed to fit within the engine com-
partment, although the assembly is not considered to be a fully readable
des i gn.
CONDENSER PERFORMANCE
Design point performance goals are as follows:
Vapor Side
Working fluid Water
Total heat rejection, Btu per hr 1.21 x 10
Flow, Ib per hr 1285
Inlet temperature, °F 258
Inlet pressure, psia 34.0
Condensing temperature (avg), °F 256
Condensing pressure (avg), psia 33.3
Outlet temperature, °F 256
Subcooling, °F 0.0
Core pressure drop, psi 0.6
Overall pressure drop, psi 1.0
A-9
-------
Air Side
Flow, Ib per hr 38,200
Inlet temperature, °F 85
Inlet pressure, psia 14.7
Outlet temperature, °F 217
Temperature effectiveness 0.763
Overall core pressure drop, in. H?0 2.1
2.I Condenser Design
The condenser shall be a plate-fin crossflow design utilizing bar and
plate, brazed and welded construction. The material of construction
shall be aluminum alloy. The unit shall be fluxless brazed to avoid
contamination in the vapor passages.
3. FAN PERFORMANCE
Design point performance goals are as follows:
Total airflow, Ib per hr 38,200
Total no. of fans 2
Inlet temperature, °F 217
Inlet pressure, psia 14.7
Airflow per fan, cfm 5440
Total pressure rise, in. H?0 2.9
Speed, rpm 2360
Efficiency, percent 70
Shaft power input, hp 3.6
3. I Fan Design
The fan shall be a tube-axial type of the following configuration:
Ti p diameter, in. 18.85
Number of blades 17
Fan rotation as viewed from driver's seat ccw
Maximum operation speed, rpm 3600
Material Aluminum alloy
A-10
-------
3.2 Fan Drive
The fan shall be designed to be belt driven. Interface requirements shall
be as specified on Outline Drawing 190640. Maximum load imposed on the
fan bearing assembly by the belt drive during any operating condition
shal1 be 1imited to 600 Ib.
3.3 Noise Standards
No specific limit on the noise generated by the condenser and fan assembly
is specified. The fans shall be designed, however, to produce a minimum
amount of noise.
4. INSTALLATION
Overall dimensions, mounting arrangements, provisions for differential
thermal expansion, inlet and outlet port configuration, and weight shall
be as shown on Outline Drawing 190640.
4. I Airflow Characteristics
Condenser and fan design point airflow rating of 38,200 Ib per hr shall
be based on the following installation characteristics as specified by
the system contractor:
Vehicle speed, mph 32
Inlet flow dynamic pressure, in. H?0 0.44
Inlet recovery factor, percent 100
Net ram pressure rise, in. hLO 0.44
Installation pressure loss*, in. FLO 1.22
5. STRUCTURAL DESIGN
The condenser shall be designed to withstand a maximum operating pressure
of 35 psig at a metal temperature of 280°F. The corresponding proof
pressure shall be 86 psig at room temperature. The support structure
shall be designed to withstand the stresses associated with a bench test
of the preprototype engine.
6. FINISH
All external surfaces of the condenser and fan assembly, with the exception
of the air-side fins and the fan impellers, shall be painted in accordance
with AiResearch Process Specification FP-73F04 (epoxy paint, flat black color),
•"•Installation loss includes the pressure losses as a result of airflow
across the vehicle bumper and grille, and the losses associated with
the fan discharge airflow through the engine compartment.
A-11
-------
7. ACCEPTANCE TEST
7.I Proof Pressure
The condenser shall be pressurized to 86 psig at room temperature.
After pressure has been released, there shall be no evidence of permanent
deformati on.
7.2 Leakage
When pressurized to 86 psig, the proof pressure, there shall be no visible
1eakage.
7.3 Fan Cali brat ion
The fan shall be tested as a component to establish its performance.
The fan shall be run at the design point speed, and the fan pressure
rise shall be determined over a range of airflows.
7.4 Condenser and Fan Performance
The condenser and fans shall be tested as an assembly to establish the
airflow characteristics of the system. The assembly shall be tested
without any inlet or outlet flow obstructions, and the airflow through
the system shall be determined over a range of fan operating speeds
including the design point speed.
A-12
------- |