EPA-460/3-73-004
LOW EMISSION
COMBUSTOR/VAPOR GENERATOR
FOR AUTOMOBILE
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-460/3-73-004
LOW EMISSION
COMBUSTOR/VAPOR GENERATOR
FOR AUTOMOBILE
RANKINE CYCLE ENGINES
Prepared by
T. E. Duffy, J. R. Shekleton,
R. B. Addoms, and W. A. Compton
Solar Division International Harvester Company
2200 Pacific Highway
San Diego, California 92138
Contract No. 68-04-0036
EPA Project Officers:
P.P. Hutchins, S. Luchter, 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
October 1973
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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 Techni-
cal Information Center, Environmental Protection Agency, Research
Triangle Park, North Carolina 27711 or may be obtained, for a
nominal cost, from the National Technical Information Service,
5285 Port Royal Road, Springfield, Virginia 22151.
This report was furnished to the U.S. Environmental Protection Agency
by Solar Division International Harvester Company in fulfillment of
Contract No. 68-04-0036 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 correspon-
dingly. Mention of trade names or commercial products does not
constitute endorsement or recommendation for use.
Publication No. EPA-460/3-73-004
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FOREWORD
This is a final technical report on the work performed under contract
68-04-0036. The opinions, findings, and conclusions expressed are those
of the author and not necessarily those of the Environmental Protection
Agency.
Contract 68-04-0036 with Solar Division of International Harvester
Company, San Diego, California, was sponsored by the Environmental
Protection Agency, Division of Advanced Automotive Power Systems Dev-
elopment, Ann Arbor, Michigan. It was administered initially under the
direction of Mr. F. P. Hutchins, and subsequently by Mr. S. Luchter and
Mr. E. Beyma, EPA Project Officers.
The program was conducted at Solar Division of International
Harvester Research Laboratories with Mr. W. A. Compton, Assistant
Director of Research and Technical Director and Mr. T. E. Duffy, Research
Staff Engineer, the Project Manager. Others contributing to the program
are: Mr. J. R. Shekleton, Dr. R. B. Addoms and Mr. J. C. Napier.
Geoscience Ltd. provided analysis used in Appendix VII and VIII.
Solar's internal report number is RDR-1708-6.
111
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CONTENTS
Section
1 SUMMARY .
2 INTRODUCTION 7
3 SYSTEM REQUIREMENTS 9
3.1 Test Bed Combustor/Steam Generator 9
3.1.1 Combustor Requirements 9
3.1.2 Parallel Flow-Bare Tube Steam
Generator Goals 10
. i 3.1.3 Controls 11
3.2 High Efficiency Finned Tube Steam Generator 11
3.2.1 Combustor 11
3.2.2 Steam Generator 11
3.2.3 Controls 11
3.3 Units Designed for EPA System Contractors 12
3.3.1 Fluorinol-85 System 12
3.3.2 AEF-78 Vapor Generator ' 13
3.3.3 SES Steam Generator 14
4 SYSTEM CONFIGURATION ANALYSIS 15
4. 1 General Constraints and Selection of a Rotating
Cup System 15
4.2. Configuration Analysis 18
5 COMBUSTOR 21
5.1 Overview of Combustor Section 21
5.2 Combustor Rig Tests 22
5.2. 1 Description of Combustor Rig 22
5.2.2 Control of NOX in Combustor 26
5.2.3 Combustor Rig Test Results 27
5.3 Integrated Combustor Performance 34
5.3.1 Design 34
5.3.2 Air Supply System 39
5.3.3 Flame Performance 51
5.3.4 Integrated System Emissions Evaluation 57
5.3.5 Temperature Pattern 77
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CONTENTS (Contd)
Section Page
6 PARALLEL FLOW STEAM GENERATOR (TEST BED
UNIT) 81
6. 1 Steam Generator Core Matrix 81
6.2 Construction 82
6.3 Performance Tests 87
6.3.1 Mechanical Integrity 90
6.3.2 Flow Stability 91
7 HIGH EFFICIENCY SINGLE FLOW PATH STEAM
GENERATOR 99
7.1 Construction 102
7.2 Steady State Performance . 105
8 CONTROL SYSTEM 109
8.1 System Description 110
8.2 System Components 112
8.2.1 System Flow Arrangement 112
8.2.2 Air Valve and Triple Valve
Mechanization 112
8. 2.3 Air Valve 115
8.2.4 Fuel Valve 115
8.2. 5 Water Metering System 120
8.3 Electronic Control 124
8.4 Test Results With Parallel Flow Steam Generator 125
8.4. 1 Explanation of Test Results 125
8.4.2 Cycling and Step Transient Performance 136
8. 5 High Efficiency Monotube Steam Generator
Performance 141
9 ORGANIC SYSTEMS 153
9.1 Fluorinol-85 System Design 153
9.1.1 Design Constraints 154
9.1.2 Core 154
9.1.3 Manifolds 155
9.1.4 Combustor and Air Valve Design 156
9.2 AEF-78 System Design 162
9.3 Internally Finned Tubing 165
VI
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CONTENTS (Contd)
Section Page
10 SYSTEM NOISE 169
APPENDIX I 181
APPENDIX II 199
APPENDIX III 203
APPENDIX IV 235
APPENDIX V 241
APPENDIX VI 245
APPENDIX VII 249
APPENDIX VIII 269
NOMENCLATURE 337
REFERENCES 343
VII
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ILLUSTRATIONS
Figure Page
1 Final Combustor/Steam Generator Configuration 2
2 Emission Measurements Versus Air Valve Position
(Power Level) 4
3A 20 Percent to 40 Percent Low Frequency Cycling of
Steam Flow 4
3B Step Changes in Steam Flow From High Power 5
3C Maximum Amplitude and Frequency Cycling 51'
4 Preliminary Combustor Test Rig 23
5 Schematic of Gas Temperature Sensor and Mechanical
Support 24
6 Combustor Side of Vapor Generator Simulator Showing
Location of Thermocouples 25
7 Emission Probe ' 25
8 Test Combustor Installed in Rig 26
9 Emission of Nitric Oxide as a Function of Air-Fuel
Ratio and Time • 28
10 Rig Test Combustor 29
11 Upstream Side of Combustor 29
12 Combustor Ready for Installation on Rig 30
13 NO£ Emissions Versus Fuel Flow 32
14 CO Emissions Versus Fuel Flow 33
IX
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ILLUSTRATIONS (Contd)
Figure Page
15 Hydrocarbon Emissions Versus Fuel Flow 33
16 Combustor Exit Plane Radial Temperature Distribution
at Maximum Flow 34
17 Initial Integrated System Test Configuration With Conical
Cup . 35
18 Combustor/Air Supply System (Drive Motor Side) 35
19 Top View of Combustor Air Inlet to Fan 36
20 Fan Mounted to Control Valve Backup Plate 37
21 Rotary Shear Plate 37
22 Combustor/Air Supply System (Vapor Generator Side) 38
23 Conical Fuel Cup 38
24 Air Supply Fan 40
25 Calculated Characteristics of Fan 41
26 Total Pressure Rise Calculated Compared to Test Points 42
27 Actual Static Pressure Rise Compared to Calculated
Total Pressure 43
28 Primary and Secondary Air Metering Port Configuration
Shown in 50 Percent Power Position 44
29A Air Valve in 2 Percent Position 45
29B Air Valve in 50 Percent Position 45
29C Air Valve in 100 Percent Position 46
30 Air Valve Backup Plate 46
31A Fan-Air Valve Flow Test Rig 47
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ILLUSTRATIONS (Contd)
Figure Page
3IB Fan Test Rig Diffuser Air Valve Section 47
32 Cd Versus Fan Flow for Several Metering Plate
Orifice Areas 49
33 Fan Drive Motor 50
34 Fan Drive Motor Characteristics 51
35 Combustor/Air Supply System Test Arrangement 52
36 Flame at 1.5 PPH 53
37 Flame at 60 PPH Fuel Flow 53
38 Flame at 80 PPH Fuel Flow 54
39 Flame at 100 PPH Fuel Flow 54
40 Flame at 130 PPH Fuel Flow 55
41 Emission Probe Location and Configuration 58
42 Cylindrical Cup Configuration for Emission Tests 60
43 Cylindrical Cup With Double Heat Shield 61
44 Auxiliary Air Swirler Ports 62
45 Tangential Slots at Base of Auxiliary Air Swirler 62
46 Recirculation Fan on Conical Cup 63
47 Conical Cup Emissions With Recirculation Fan (CO2
Adjusted to Maintain NOX at 1.38 GM/KGM) 63
48 Emissions With JP-4 and "A" = 0. 125 and 2 Heat Shields 65
49 Emissions With JP-4 and "A" = 0. 125 With Auxiliary
Air Swirler 66
50 NO2 Emissions With A = 0. 192, 2 Heat Shields and
Auxiliary Air Swirler 67
xi
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ILLUSTRATIONS (Contd)
Figure Page
51 CO Emissions With A = 0. 192, 2 Heat Shields and
Auxiliary Air Swirler 67
52 HC Emissions With A - 0. 192, 2 Heat Shields and
Auxiliary Air Swirler 68
53 Combustor Post Tests Recorded in Figures 50, 51 and 52 68
54 Emissions With Gasoline, A = 0. 192 and 2 Heat Shields 69
55 Emissions With Gasoline, A = 0.03, Auxiliary Air
Swirler and Recirculation Fan 70
56 Auxiliary Air Swirler Configuration Variations 73
57 Combustor After 70 Hours of Operation on EPA Reference
Gasoline 75
58 Auxiliary Air Swirler After 70 Hours of Operation With
EPA Reference Gasoline 75
59 Combustor and Steam Generator Outlet Gas Temperature 78
60 Test Bed Vapor Generator - Water Working Fluid 82
61 Vapor Generator Coil Connectors 83
62 Vaporization Coils (6 Per Row) 83
63 SpecialBox Connection After Burst Pressure Test 84
64 Vapor Generator Flow and Instrumentation 85
65 Assembly of Two Preheater Coils With Special Connector
Welded at Inside of Coils 86
66 Superheater Outlet Row Showing the Six Thermocouples 87
67 Test Cell Flow Schematic 88
68 Test Cell Steam Generator Control Panel 89
Xll
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ILLUSTRATIONS (Contd)
Figure Page
69 Steam Generator After 500 Hours of Emission and
Control Systems Tests 91
70 Vaporizer Spacers After 500 Hours of Operation 92
71 Superheater Outlet Tube Wall Temperature (Six Flow
Paths) 94
72 Test Bed Vaporizer Steady State Performance 97
73 Test Bed Vaporizer. Tube Wall Temperature at
Superheater Outlet 98
74 Preheater Coils With Copper Fins 103
75 One of the Two Dryer Coils 103
76 Assembled Steam Generator 104
77 Steam Generator U Connections 104
78 High Efficiency Finned Tube Steam Generator Steady
State Efficiency Measurements 106
79 Top Coil (Vaporizer) After 70 Hours of Operation 107
80 Bottom Coil (First Row Preheater) After 70 Hours
of Operation 107
81 Combustor/Vapor Generator Control System 111
82 Controls System Flow Schematic 113
83 Triple Valve Actuator Showing Air Valve Actuation Tab 114
84 Triple Valve Actuator With Fuel and Water Valves
Installed 114
85 Exploded View of Fuel Valve 116
86 Fuel Valve Components 116
Xlll
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ILLUSTRATIONS (Contd)
Figure Page
87 Metering Slot 1 17
88 Assembled Fuel Valve 117
89 Reynolds Number Vs. Percent Fuel Flow for JP-4
and Gasoline 118
90 Fuel Valve Calibration 119
91 Fuel Valve Calibration 1 to 10 ppm Range 120
92 Water Metering Valve Orifice 121
93 Component Parts of the Water Metering Valve 121
94 Assembled Water Metering Valve 122
95 Water Metering Vavle Calibration at 50 psi Differential 123
96 Differential Pressure Control Valve Components 123
97 Differential Pressure Control Valve Assembled With
Electric Actuator 124
98 Electronic Components Required to Perform Control
Functions 125
99 Steam Generator Control Transients, Step Changes
Prior to Installation with 44 Ib/in. Spring in AP Valve 127
100 Low Frequency 10 to 30 Percent Ramp Cycles After
Installation of High Gain AP Valve 129
101 Full Power Steady State Performance With Step
Transients in Steam Flow 131
102 High Amplitude and Maximum Frequency Steam Flow
Cycling System Response 133
103 Transient Emissions During Peak Steam Generator
Cycling 140
xiv
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ILLUSTRATIONS (Contd)
Figure Page
104 Open Loop Characteristics to Step Inputs in Water Rate 143
105 Open Loop Sinusoidal Frequency Response 145
106 Closed Loop Transient Response With and Without
Derivative Compensation 147
107 High Frequency Cycling 149
108 Response to Step Changes in Steam Flow 151
109 Fluorinol-85 Combustor/Vapor Generator (Front View) 157
110 Fluorinol-85 Combustor/Vapor Generator (Top View) 157
111 Fluorinol-85 Combustor/Vapor Generator (Side View) 158
112 Fluorinol-85 System Air Metering Valve Shown in 56.7
Percent Position 161
113 AEF-78 Combustor/Vapor Generator (Front View) 163
114 AEF-78 Combustor/Vapor Generator (Top View) 163
115 AEF-78 Combustor/Vapor Generator (Side View) 164
116 Tube Cross Section . 165
117 Photomicrographs of Tube-Fin Wall 166-7
118 Noise Emission Microphone Location in Test Cell 171
119 Sound Level, dB(A), Versus 1 /10 Octave Frequency,
Water Pump Only - 600 psi Outlet Pressure - 88 dB(A)
Overall 172
120 Sound Level, dB(A), Versus 1/10 Octave Frequency,
Fan Only, 5700 rpm - 30 Percent, 92 dB(A) Overall 173
121 Sound Level, dB(A), Versus 1/10 Octave Band Frequency.
Full System, 20 Percent Fan, 380 Ib/hr Steam, 96 dB(A)
Overall (Background Included) 174
xv
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ILLUSTRATIONS (Contd)
Figure Page
122 Sound Level, dB(A), Versus 1/10 Octave Frequency.
Full System, 20 Percent Fan, Boiler Flooded, 94 dB(A)
Overall (Background Included) 175
123 Sound Level, dB(A), Versus 1/10 Octave Frequency.
Full System, 50 Percent Air Valve, 780 Ib/hr Steam,
101 dB(A) Overall (Background Included) 176
124 Sound Level, dB(A) Versus 1/10 Octave Frequency.
Full System, 50 Percent Fan, Boiler Flooded, 96 dB(A)
Overall (Background Included) 177
125 Sound Level, dB(A), Versus 1/10 Octave Frequency.
Full System, 77 Percent Fan, 1200 Ib/hr Steam, 108
dB(A) Overall (Background Included) 178
xvi
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TABLES
Table Page
I Measured Levels at Simulated Driving Cycle
Steady State Points 3
II Emission Level Goals 10
III Fan Test Results With No Bypass - Fan Speed
5700 RPM 48
IV Initial Combustor Plus Fan Air Supply Test Results 56
V Fan Power With High Ratio Pulley 56
VI Residue Removed From First Row of Steam Generator 92
VII Fluid Conditions 100
VIII Summary of Performance Parameters 101
IX Overall "A" Scale Weighed Sound Level for Steam
Generator System and Components 170
xvn
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1
SUMMARY
This is the final report on a program to demonstrate a low emission
vapor generator for automotive Rankine cycle power plants. Prog-F-a-m_g,o_aJL§_
were to design and test a low emission system that required low parasitic
power, had compact packaging, high steam generator efficiency and suit-
able controls for fuel, air and water for the regulation of steam pressure
and temperature. A steam generator with an output of 1200 pounds per hour
at 1000°F and 1000 psia was demonstrated by tests to have weighed emissions
below the 1976 emission standards over a simulated driving cycle. Figure 1
is a cross section of the unit tested.
Parasitic power limitations were the major factor in determining
the overall configuration. Maximum available electrical power requirements
established a maximum ideal input to the system of 1. 2 HP. To flow the
necessary air through the system, a large frontal area steam generator was
integrated with a low pressure drop combustor (5 inches of water). Low
emissions required the maximum possible mixing velocities at the low flows.
To achieve this'a unique variable flow comtnistor design-was used in which an
air valve splits t'fie flow between primary and secondary air injection as a
function of power level required. In this manner the pressure drop, and
thus velocity was maintained relatively high at low firing rates.
Major design features include: rotating cup fuel atomization and
injection, fully modulated controls for air, fuel and water, symmetrical air.
distribution with no exhaust recirculation, flat spiral steam generator tube
matrix with extensive use of finned tubing and a straight-through combustion
gas flow path. A high degree of component integration was used to allow
. co-axial mounting of the fan and fuel cup with a single belt drive from a DC
side mounted electric motor. Performance tests on this unit showed that
"it met major performance goals of the program.
Table I summarizes the results of emission tests. All tests were
performed with the configuration shown in Figure 1, while operating the
steam generator at 1000°F and 1000 psia. Emissions were measured by an
averaging probe located two inches downstream of the vapor generator in the
exhaust duct. The characteristics of the emissions as a function of flow are
shown in Figure 2.
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, - , BLADES\ , , 1 J
1 BELT
FAN
24
V"~
I ( !
ooooo--
TRIPL
VALVE
ACnjA
TrTT
AJIA
E~
OL
TOR
;AMS
WATER
AND
FUEL
VALVES
WATER
TO
STEAM
GEN
1
, 1
j
EXHAUST
FIGURE 1. FINAL COMBUSTOR/STEAM GENERATOR CONFIGURATION
Two different types of steam generator tube matrices were fabricated
and tested. One unit was a test bed for initial combustor development and
was used to investigate the feasibility of parallel flow passages for compact
forced circulation steam generators. Parallel flow development tests were
to provide background for future organic system designs normally requiring
multiple flow passages. Parallel flow instability was observed and analyzed
in the first unit but could be controlled by correct scheduling of the startup
procedure. At near design conditions, serious instabilities disappeared. A
steady state flow imbalance between passages caused outlet temperature
differences between each of the six parallel flow paths, but did not present
a serious operational problem. A second unit made extensive use of extended
surfaces and had a measured efficiency of 86 percent LHV at full steam flow
and 93 percent LHV at 25 percent steam flow.
Both the parallel flow unit and the single flow path high efficiency
steam generator were integrated and tested with fully modulated electronic
control system tailored for each unit.
Steady state control system performance was within ±50 psi and ±50°F
of the setpoints. Transient response to the steam flow demand changes
resulting from rapid opening and closing of a throttle valve are shown in
Figure 3. Pressure and temperature outlet errors were maintained at values
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TABLE I
MEASURED LEVELS AT SIMULATED DRIVING'CYCLE
STEADY STATE POINTS
Fuel Flow
9pph
12
15
25
Time
(%)
3
76
20
1
N02
(gm/kgm)
1.31
1.32
1.35
1.25
HC
1.0
0.8
0.9
0.5
CO
9.8
9.5
9.0
6.2
TIME INTEGRATED LEVELS FOR
DRIVING CYCLE (10 MPG)
N02
HC
CO
Actual
0. 38 gm/mile
0. 24 gm/mile
2.75 gm/miie
*ER - Emissions Ratio =
Limit
0.40 gm/mile
0.41 gm/mile
3. 4 gm/mile
Actual
Limit
ER*
0.95
0.59
0.81
of less than ±50 psia and ±50°F during "normal" driving steam flow demands
shown in Figure 3A. Under severe steam flow demands (Figs. 3B and 3C)
steam conditions were maintained within a ±150 psia and ±100°F control
band. An open loop schedule of fuel-air-and-water valve position as a
function of steam flow was the primary control mode. A closed loop electronic
control of firing rate provided a trim regulation of pressure. Temperature
trim was also done electronically with a thermocouple in the superheater
outlet providing a signal to regulate feedwater rate.
Sound level measurements were made on the system and showed levels
that can be made acceptable in a closed engine compartment given noise
suppression treatment.
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1.4
.8
.4
0
E
01
3.0
2.0
1.0
0
0
LIMIT"
.4
.3
.2
.1
0
O)
LIMIT —
.8
20 40 60
AIR VALVE POSIT ION %
80
100
.4
0
FIGURE 2. EMISSION MEASUREMENTS VERSUS AIR VALVE
POSITION (POWER LEVEL)
_
1 1 1 1
1
— 20 SECONDS
1 1 1 I 1 I 1 I I 1
1 I
i i i i
i i I I
1250
1000
750 OUTLET PRESSURE
500 PSIA 250 PSI/CM
250
0
100
80
60 . STEAM THROTTLE
40
20
o/
1°
6g5
POSITION 20% CM
800 SUPER HEATER OUTLET
585 TEMPERATURE °F TYPE K
oF 5 MV/CM
FUEL-AIR-WATER
% METERING VALVES
POSITION 20%/CM
FIGURE 3A. 20 PERCENT TO 40 PERCENT LOW FREQUENCY
CYCLING OF STEAM FLOW
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1 1 1 1
V
1
^
^20SECONDS I
1 1 1 1 1 1 1 1 1 1 1 1
1250
1000
750 OUTLET PRESSURE-
500 PSIA 250 PSI/CM
250
0
100
80
60 . STEAM THROTTLE
40
20
o/
/»
905
- 695
POSITION 20% CM
800 SUPER HEATER OUTLET
585 TEMPERATURE °F TYPE K
op 5 MV/CM
100
8° FUEL-AIR-WATER
6° % METERING VALVES
40 POSITION 20%/CM
0
FIGURE 3B. STEP CHANGES IN STEAM FLOW FROM HIGH POWER
f\
(_
1 1 1 1
1
^ A A A A A f\ f\ /— ^__ n -f
^yvWWVV/v \j\J\s — J_
— 20SECONDS I
II
i i i i i i i i i i i i
1250
1000
750 OUTLET PRESSURE
500 PSIA 250 PSI/CM
250
0
100
80
60 „, STEAM THROTTLE
40 POSITION 20% CM
20
0
1115
905
'1010
695 800 SUPER HEATER OUTLET
585 TEMPERATURE °F TYPE K
°F 5 MV/CM
100
80
60
40
20
0
FUEL-AIR-WATER
% METERING VALVES
POSITION 20%/CM
FIGURE 3C. MAXIMUM AMPLITUDE AND FREQUENCY CYCLING
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Two organic fluid vapor generators were designed as backup for EPA
system contractors. AEF-78 and Fluorinal-85 organic fluid vapor generators
were designed to interface with the same type of combustor tested with the
steam generators described above. Both used a rectangular straight tube
array with machined manifolds for return bends. An all brazed construction
was used with extruded internal fins incorporated to increase the heat trans-
fer on the organic fluid side.
The results of the program proved that a low emission compact vapor
generator could be developed to meet the emission and control requirements
for automotive vehicles. The emission levels obtained, however, were
borderline. Further development work in this concept will be required
before sufficiently large emission margins were available for the normal
tolerances required for vehicle applications. Among the paths available
for lowering emissions are added pressure drop and addition of exhaust
gas recirculation. Since these changes would increase the parasitic power
demands, future work should be based upon a vapor generator optimized
for a particular engine system.
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2
INTRODUCTION
Rankine cycle engines are the most highly developed external combus-
tion system with much of the world's power being generated by such power
plants. However, little development progress had been made to reduce
emissions and improve the system and components for application at power
level suitable for automobiles. Emissions, efficiency, packaging, controls,
and cost problems required solutions that could only be fully evaluated by
combined component technology advancements and automobile system demon-
strations. A compact low emission vapor generator with controls compatible
with low emissions and automotive duty cycles is one of the major components
requiring advancements. Solar, under contract to EPA, has demonstrated
two different steam generators designed to meet the requirements of an
automotive system. A low emission combustor based upon the results of an
earlier Solar combustor demonstration program contracted by EPA (Ref. 1)
has been the basis of the design described in this report. Integration of these
combustor control concepts with specially designed compact vapor generators
forms the basic problem attacked by this project.
The purpose of this work was to demonstrate by means of test results
that a practical fully integrated steam generator had emissions below the 1976
Federal Standards. Major goals of the program are:
• High turndown range
• Compact packaging
• Low emissions during transients
• Stable vapor generator operation
• Accurate steady state control of outlet pressure and superheater
temperature
• Sufficient control capability to prevent excessively high or low
pressures and temperatures at the superheater outlet or burn-
out failures during the large amplitude and frequent steam flow
demand changes required by normal automotive stop and go
driving
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• Low parasitic power
• Low noise
Three different fluids were required to be evaluated through the design
phase of the program. Water, Fluorinol-85 and AEF-78 vapor generators were
designed specifically to interface with Steam Engines Systems, Thermo Electron
and Aerojets' vehicle packages. Each of these companies is on contract to the
EPA to develop a complete automotive Rankine Cycle system. Solar's steam
generator program was to provide backup capability for this critical component.
At the end of the design phase it became apparent that the greatest gain to the
technology would be for Solar to concentrate its development tests on water
systems alone. This decision in no way was based on technical tradeoffs
between organic based fluids and water for the automotive engine, but simply
on development cost analysis. However, one of the two steam (water) gener-
ators tested was designed to specifically obtain test data on a potentially
significant problem of organic vapor generators' parallel flow stability.
8
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3
SYSTEM REQUIREMENTS
Two steam generators have been designed and tested in this program.
The first unit was a test bed steam generator to evaluate its effects upon
emissions, parallel flow, and the control system. Efficiency was not stressed
in this unit in order to expedite construction (no extended surfaces). A second
unit designed and tested placed emphasis on efficiency and elimination of
parallel flow instabilities. It incorporated extensive finned tubing and single flow
path in the critical vaporization and superheater tube sections. Three other
units were designed for the automotive systems being developed by EPA.
Each of these systems had a different fluid and required an envelope designed
to interface with each of the three system contractors.
3. 1 TEST BED COMBUSTOR/STEAM GENERATOR
This unit incorporated six parallel flow paths in the steam generator
and bare tubes. Efficiency requirements were made sufficiently low to allow a
bare tube unit that could be fabricated entirely at Solar with a minimum of lead
time.
3.1.1 Combustor Requirements
• Heat release 2. 5 x l.O6 BTU/hr based on LHV
• Turndown ratio 40:1
• Fuel: kerosene or JP-5, EPA reference unleaded gasoline was
also used when it became available
• Size: 24 inches square by 20 inches high not including exhaust
ducting
• Weight: 175 pounds
• Startup time to full vapor generator in less than 15 seconds from
a cold start.
• Maximum electrical input to be less than 2. 5 HP electrical
equivalent
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Emissions to be below the 1976 Federal Standards. Test
results are to be compared to the limits assuming a 10 MPG
fuel consumption as shown in Table II.
TABLE II
EMISSION LEVEL GOALS
Constituent
Hydrocarbon (HC)
Carbon Monoxide (CO)
Oxides of Nitrogen NO + NO2
reported as (NO2)
Grams/Mile
0.41
3.4
0.4
Grams /Kilo gram*
. 1.42
11.8
1.38
"'Calculated assuming 10 MPG
3.1.2 Parallel Flow-Bare Tube Steam Generator Goals
• Fluid: water
• Pressure: 1,000 psia
• Flow: 1525 Ibs/hour of steam
• Temperature: 1000eF
• Efficiency: greater than 80% based on LHV, 75% based on HHV
(Note actual efficiency was 78% LHV at full flow)
• Core diameter: 21.5 inches
• Core length: less than 8 inches
• Maximum air side pressure drop: 3 inches of water
• Maximum water side pressure drop: 250 psi
• Construction: bare tubes to allow rapid construction
10
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• Flow arrangement: to be similar to an organic vapor generator
with respect to air side and working fluid side temperature and
flow distributions.
3.1.3 Controls
• Air-Fuel Control. An automatic control of air and fuel flow is
required to maintain the equivalence ratio at optimum levels
across the entire 40 to 1 turndown range.
• Steam Generator. An automatic control of feed water, fuel and
air flow is required to maintain the superheater outlet at 1000°F±
100°F and 1000 psia ± 100 psi during steady state and transients
expected in an automotive operation. Although exact flow rate
changes are not precisely known, high rates can be anticipated.
Flow rate is basically determined by engine speed, (a relatively
slow factor) and cutoff valve actuation (a fast parametei- requi re-
ing only a few milliseconds for full travel). Actual test results
were obtained with transients rates of as high as 900 pounds per
hour steam flow change per second. These transients were
made in steps as large as 80 percent of full flow.
3.2 HIGH EFFICIENCY FINNED TUBE STEAM GENERATOR
3.2.1 Combustor
Identical to the test bed combustor, see paragraph 3.1.1.
3.2.2 Steam Generator
• Efficiency: 85% based on LHV, 79.6% based on HHV (Note
actual full flow was 86% based on LHV).
• Construction: finned tubes
• Flow arrangement: no parallel flow passages in superheater or
vaporizer sections. This prevents "chugging" flow instabilities
and simplifies the control since only a single outlet tube can be
monitored for overtemperature.
• All other constraints the same as the parallel flow unit, see
paragraph 3.1.2.
3.2.3 Controls
Same as test bed unit, see paragraph 3.1.3.
11
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3.3 UNITS DESIGNED FOR EPA SYSTEM CONTRACTORS
3.3.1 Fluorinol-85 System
Fluorinol-85 Side
Flow
Pressure Drop
Outlet Pressure
Inlet Temperature
Outlet Temperature
Heat Transfer to Fluorinal-85
Gas Side
Air-Fuel Ratio
Flow
Pressure Drop
Inlet Temperature
Efficiency
Outlet Temperature
10,000 Ibm/hr
130 psi maximum
700 psia
287° F (at max. power)
550°F
2.25 x 106 BTU/hr (reference)
25:1 (JP-5 fuel)
3740 Ibm/hr
3.0 inches water column
25008F (mean)±250°F
81% based on HHV (reference)
86. 5% based on LHV (reference)
427°F (reference)
NOTES:
The maximum tube wall temperature during steady state operation
shall be less than 575°F. The maximum temperature was based upon an
assumption that the gas side inlet temperature is 2750°F and a velocity
deviation at the inlet plane of the vapor generator is 10 percent above the design
mean design value.
Envelope requirements: Outside envelope dimensions were within
the following configuration. A 20 inch diameter combustor section shall be
oriented in the vertical firing down position with a six inch air inlet at the top.
The 20 inch diameter shall extend vertically downward 10 inches to a rectan-
gular vapor generator section. The vapor generator is designed to fit within
dimensions of 1 5 inches by 25 inches by a core height of 10 inches (including
transition). It shall be symmetrically installed below the vapor generator to
give an overall height of 20 inches. Envelope parameters were held
constant in the design phases.
Efficiency is defined by the relationship
Q = WfT|H
12
-------
where Q = heat transfer rate to working fluid
T) = efficiency
Wr = fuel rate
H = heating value of fuel
The fuel is assumed to be JP-5 with HHV = 19,800 BTU/lb, LHV = 18, 550
BTU/lb.
3.3.2 AEF-78 Vapor Generator
AEF-78 Side
Flow
Pressure Drop
Outlet Pressure
Inlet Temperature
Outlet Temperature
Heat Transfer Rate to AEF-78
Gas Side
Air-Fuel Ratio
Flow
Pressure Drop
Inlet Temperature
Efficiency
Outlet Temperature
19,300 Ibm/hr
50 psi goal, 100 psi maximum
1000 psia
396°F
650°F
2.02 x 106 BTU/hr (reference)
25:1 (JP-5 fuel)
3400 Ibm/hr
3.0 inches of water column
2500«F (mean) ± 250°F
80% based on HHV (reference)
85. 5% based on LHV (reference)
455°F (reference)
NOTES:
The maximum tube -wall temperature during steady state operation
shall be less than 720°F. This shall be a local temperature and shall not
represent more than a small percentage of the total fluid temperature. Maxi-
mum temperature shall be based upon the assumption that the gas side inlet
temperature is 2750°F, and the gas velocity at the inlet plane is 10 percent of
the design mean value.
Envelope requirements. Outside envelope requirements shall be
within the following configuration. A vertically downward firing combustor
shall be mounted on top of the vapor generator.. The assembled unit shall
have an outside diameter of less than 26 inches and an overall height to the
exit plane of the vapor generator of 17 inches.
13
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3.3.3 Steam Generator
Preliminary specification for a steam generator for vehicle installa
tion are listed below.
Water Side
Flow
Pressure Drop
Outlet Pressure
Inlet Temperature
Outlet Temperature
1200 Ibm/hr
150 psi goal, 250 psi maximum
1000 psia
220°F
1000°F
Heat Transfer Rate to Water 1. 58 x 106 BTU/hr (reference)
Gas Side
Air-Fuel Ratio
Flow
Pressure Drop
Inlet Temperature
Efficiency
Outlet Temperature
25:1
2670 Ibm/hr
3 inches of -water column
2500° F mean ± 250° F
85% based on LHV
79.6% based on HHV
462 F (reference)
The preliminary estimate of the maximum envelope allowed for this system
is a horizontally firing (towards back of vehicle) combustor with a maximum
diameter of 19.5 inches. Total length of the combustor/vapor generator
shall be less than 16 inches.
14
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4
SYSTEM CONFIGURATION ANALYSIS
4. 1 GENERAL CONSTRAINTS AND SELECTION OF A ROTATING CUP
SYSTEM
Achieving emission levels below the 1976 Federal emission levels is
a challenging objective for a combustor rig operating independently. Con-
straints imposed by addition of practical automotive operations significantly
complicate the task. Automotive requirements that have a major influence
on the design include:
• Compact packaging
• High turndown ratios
• Low parasitic power
• Virtually continuously varing power level demands
• Average steam rate of approximately 10 percent of maximum in
the Federal driving cycle
• Cold startup
• Rapid warm up
* Minimum superheater temperature overshoot (to prevent
lubricant and seal degradation)
• Narrow band superheater outlet temperature control to main-
tain temperature high as is safe for maximum cycle efficiency
• Minimum overshoot in pressure control to prevent damage or loss
of fluid through relief valve
• Minimum water hold up to minimize energy release hazards
15
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A typical example of the design constraints imposed by the high
turndown ratio can be seen in a brief review of the air-fuel ratio requirements.
To maintain the air-fuel ratio within ±10 percent of the optimum at the critical
power output of 10 to 15 percent, the valves cannot leak more than ±1 percent
of their maximum rated flow values if all other errors are ignored. As a
consequence new and unconventional approaches have been incorporated to
allow accurate metering of the three input fluids, air, fuel and water across
wide turndown ranges with compact components.
Two basic approaches are available to obtain low emission. One
attractive method is to premix the air and fuel in a vapor form prior to com-
bustion. If the fuel and air are homogeneously mixed, this method can ensure
low emissions by maintaining the overall air-fuel ratio sufficiently lean (and
thus low flame temperatures) to prevent the formation of significant NOX.
Emission levels an order of magnitude below the Federal level are theoretically
possible with this approach. Several major problems indicate that this may
not be an acceptable approach for an automotive system. Initial startup emissions
of HC with cold walls are difficult to keep within limits and may represent
the same order of magnitude startup problem as with the present automotive
spark ignition engine. To ensure rapid vaporization and uniform mixing of
air and fuel in a premixed system, large mixing volumes, high air velocities,
small fuel droplets and large heat input values are required. This equates
into the need for a high power consumption air atomization system, high
power consumption in the combustor fan, recirculation of high temperature
exhaust or combustor gases (with associated pumping losses) and the need for
long mixing volumes. Since large volumes of premixed gases are necessary
for correct mixing, a flashback explosion hazard is present. The required
high degree of turndown necessary (40 to 1) for low fuel consumption in the
driving cycle makes design of flame arresters difficult and unreliable at
very low velocities. Transient performance, as with startup emissions, is
also questionable since the heat of vaporization added to the mixture must be
matched to the instantaneous changes of fuel flow. If this dynamic balance
is not maintained relatively closely, high inlet gas temperatures may occur or
"wet wall" operation will occur. Both of these conditions are undesirable since
they affect transient air-fuel ratios and can cause higher emissions.
Direct liquid fuel injection into the combustor to circumvent the above
problems minimizes parasitic power and bulk. Analysis and development
tests have demonstrated that fuel injection, droplet vaporization, air-fuel
mixing, rich reaction and final lean burnout can achieve low emissions in an
overall lean direct liquid fuel injection system.
16
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All types of injection and atomization systems were considered. These are
listed with their limitations.
• Pressure Atomizer - This system cannot form a fuel spray at
the low fuel flows required (down to approximately 3 pounds per
hour), it needs small orifices subject to contamination, and the
high fuel pressures mean an expensive fuel pump.
• Dual Orifice Pressure Atomizer - As above but with more
small orifices to reduce reliability and added filtration costs.
• Air Assist Pressure Atomizer - Requires auxiliary high pressure
air compressor that can consume more parasitic power than the
combustor fan. This is a key element since fuel economy is
directly and strongly dependent upon the combustion system.
In the Federal Driving Cycle the average roadload power is
approximately 10 percent. Thus, addition of a typical 1 HP air
assist compressor can cause fuel economy drops of approxi-
mately 10 percent. Air assist is also limited in turndown range
thus requiring control of the air assist compressor flow as well
as the main air flow into the combustor. This adds additional
complexity and cost over and above the relatively expensive
compressor and its drive system. It also has the limitation of a
variable drop size dependent on fuel flow and a relatively narrow
spray angle.
• Air Assist (Ultrasonic) - Same as above but with less proven
reliability. The spider wire that holds the resonant chamber
can and has, at Solar, been observed to collect spray and form
large drops.
• Electrostatic and Vibrating - Auxiliary power source required,
have not demonstrated capability of operating over the wide range
of fuel flows. Performance not proven.
Solar's initial program for a low emission combustor (completed
June 1971, Ref. 1}, forms the technical basis for selecting the design config-
uration. A 2 x 10° BTU/hr combustor plus integrated air and fuel control
system was demonstrated to have low emissions across a wide range of fuel
flows. Design features determined to be of significance for low emissions
were used in synthesizing the design being reported on:
• Rotating cup atomization due to its capability for wide turndown,
low power consumption, insensitivity to contamination, wide
spray angle uniform droplet size, low fuel pressure and good
ignition characteristics.
17
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• Fully modulated air and fuel control systems mechanically
synchronized to maintain desired air-fuel ratios across the
entire 40 to 1 turndown range. On-off controls result in ignition
and shutdown hydrocarbon emissions during the continuously
cycling operations for automotive usage.
• Emphasis on symmetric air flow from inlet through fan combus-
tor and vapor generator to ensure the best possible uniformity
of air-fuel mixtures.
• Control of air flow rate into combustor by means of a valve
rather than speed changes of the fan-motor drive system since
inertia lag could cause significant delays or mismatch between
air and fuel flow.
The rotating cup atomization and fuel injection system selected has
demonstrated that it has the following features:
• Requires essentially zero parasitic power to operate if carefully
packaged into system (only several watts are necessary to
accelerate and overcome fuel friction with this system). If it is
coaxially driven by the combustion air fan, the additional load is
difficult to measure.
• Has proven turndown capabilities of greater than 100 to 1
(Ref. 1). Spray quality and angle control do not change over
this large turndown range.
• No small orifices are required thus eliminating sensitivity to
contamination and improving reliability.
• No pressure drop required other than to overcome steam
generator gas side pressure drop. Thus low cost standard
automotive fuel pumps can be used.
• Ideal spray angle for ingition. Excellent ignition characteristics
have been obtained since the fuel spray pattern can be accurately
repeated and the spark plug located at the point the fuel impinges
on the wall.
4.2 CONFIGURATION ANALYSIS
Integrating the above features with the specified goals (Section 3)
determined the overall system configuration (Fig. 1). Emission goals dictated
that the maximum possible pressure drop be utilized for mixing velocities in
the combustor. Emission levels are also dependent upon combustor exit
18
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temperatures. NO formation rates are negligible below 3200°F. From
previous tests (Ref. 1) an exit temperature of 2500°F ± 250°F provided an
adequately conservative exit temperature to meet emission goals. With a
design exit condition of 2500°F and a specified heat release of 2. 5 x 10°
BTU/hr an equivalence ratio of 0.62 is obtained. An air flow of 3400 pounds
per hour can be calculated using these factors. During the preliminary design
it was assumed that an overall electric motor-fan efficiency of 50 percent
could be obtained. Therefore, ideal air horsepower available is only 1.25 HP.
At 3400 pph this limits the fan pressure rise to less than 10. 5 inches of water.
Preliminary analysis and sizing of the air valve determined that 2.0 inches of
water pressure drop was needed for air metering (see Sec. 5). Thus a total
of 8. 5 inches pressure drop was the maximum available for the combustor,
steam generator and exhaust ducting. An allowance of 0.75 inch for exhaust
leaves 7.75 inches of water.
Emission goals require that the largest portion of the available 7.75
inches of water pressure drop be allocated to the combustor. Since the air
pressure drop across the steam generator is exponentially dependent upon
its frontal area, it is essential to maintain the maximum frontal area. Prev-
ious experience has indicated that a straight through flow arrangement with a
minimum of flow transition between the combustor and steam generator would
result in the most uniform distribution of air in the combustion chamber. As
a consequence, the full diameter available was used to design a tube matrix
that required essentially no turns or flow transition between the combustor
and steam generator. Using the maximum diameter and a compact bare tube
matrix (for the test bed system), the pressure drop is 2.75 inches of water.
The remaining 5.0 inches of water pressure was used to establish the com-
bustor pressure drop.
From the above steps the steam generator and combustor diameter
were established. Efficiency and pressure drop criteria determined the tube
matrix depth (6.20 inches). An overall height limit of 20 inches is the only
major design constraint remaining to be satisfied. A combustor emissions
goal is also a major criteria at this point. To prevent combustion products
from quenching before reaction is completed, it is good design practice to
keep the combustor length as long as possible (particularly with low pressure
drop systems). Design studies emphasizing low parasitic power, compact
and symmetrical flow distribution finalized the configuration around an inte-
grated radial flow fan coaxially mounted with the combustor and steam gener-
ator (see Fig. 1 and Section 5 for detail description). This configuration
resulted in the minimum of parasitic power consumption since the same motor
that drives the fan is used to drive the fuel atomization and injection cup.
Fan bearing and rotating cup bearing friction losses are minimized as is
mounting complexity. No external ducting is required since the fan diffuser
discharges directly into a large diameter air metering valve. A large dia-
meter valve is essential to provide low flow losses and the high degree of
19
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symmetry essential to uniform mixing at low pressure drops. The air valve
uses 20 ports circumferentially spaced around the combustor outer diameter
to direct air into an air swirler section. All flow is completely symmetric
in the air and fuel injection systems. Non-symmetrical turns, flow areas,
pressure gradients, or velocity gradients either upstream or downstream
of the combustion chamber require especially careful attention to prevent
maldistributions that would normally increase emissions. As a consequence
of this design investigation, the 20 inches of total height was divided as
follows:
• Steam generator 6.20 inches
• Fan, diffuser and air valve 1.60
• Fan inlet and drive pulley 2.00
• Primary air swirler and insulation 2.20
• Combustion chamber 8.00
20
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COMBUSTOR
5. 1 OVERVIEW OF COMBUSTOR SECTION
System restraints required that an integrated approach be taken in
the design of the following combustor subsystems:
• Air supply distribution and control
• Fuel atomization, injection and distribution
• Combustion chamber configuration
It was necessary to treat combustion as a systems integration problem in
order to meet the following major constraints:
• Ideal power limit of 0.83 HP (7 inch pressure drop at
3400 pounds per hour air flow)
• Envelope limits of 24 by 24 inches square (including electrical
drive motor) and 20 inches high
• Aerodynamic and geometric physical interface with a large
frontal area steam generator
As discussed in the previous section, these constrains when coupled with
the approach of using a rotating cup atomization system predetermine the
following combustor characteristics:
• Pressure drop: 7 inches of water total (5 inches in combustor
and 2 inches across air valve)
• Diameter: 20 inches
• Length: 8 inches
• Air Flow: 3400 pounds per hour
21
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• Fuel flow: 135 pounds per hour with JP-4, JP-5 or unleaded
gasoline
These parameters were derived at the preliminary design phase of
the program and remained consistent through combustor rig tests and
evaluation of the integrated system with both steam generators. Since the
combustor and steam generator were designed to directly interface with
essentially no transition, the geometrical configuration of the combustor was
essentially frozen from the beginning of the program. Pressure drop at
maximum flow was similarly frozen. Emissions development thus concentrated
on improving performance by relatively minor changes in mixing, flow splits,
air-fuel ratios and other adjustments not requiring basic configuration
modifications. This section describes the work performed to obtain the final
emission test results summarized in Section 1. Initial combustor rig tests
are discussed first. From these tests the total air-flow and flow split
between the primary and secondary air injection ports was obtained. Next
follows an analysis of the air supply fan, air control, and distribution system
designed to be compatible with the combustor test rig results. A discussion
of the combustor development tests for a fully integrated air supply fuel and
steam generator system concludes this section on combustion.
5.2 COMBUSTOR RIG TESTS
5.2. 1 Description of Combustor Rig
In order to verify all basic design assumptions, a series of full
scale combustor tests were performed prior to design finalization. Figure 4
schematically depicts the combustor test rig setup. High pressure air enters
the air flow measurement sections from a high pressure compressor. Air
flow into the combustor is measured by means of one of three parallel orifice
runs. Because of the wide turndown requirements, 40 to 1, three different
orifices of 0.75, 1.5 and 3 inches diameter are used. Each orifice is sized
for optimum Reynolds number at air flows corresponding to the respective
fuel flows midway in the ranges of 109 to 23, 23 to 5.0 and 5.0 to 1 pounds
per hour. The orifices are sharp edged, with corner pressure taps, and
have upstream flow straighteners to minimize aerodynamic influence of the
metering section from the upstream shutoff valves. These valves are
manually operated so that air can be diverted to the metering orifice appro-
priate to the air flow required.
Fuel flow is measured by three variable area float type meters
having ranges of 1 to 10, 10 to 110 and 50 to 170 pounds per hour. Each flow
meter has been calibrated prior to use in these tests by means of a weight
flow versus time measurements at three or more points across the flow
range.
22
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_ FROM
TANK
TO NDIR AND
CHEM LUM
COOL LINES —•
MASS AVERAGE
EMISSION PROBE
SIMULATED VAPOR
GENERATOR, AIR-
COOLED AND
INSTRUMENTED
•^—i EXHAUST
VAPOR GENERATOR
BACK PRESS
ORIFICE PLATE
48-INCH LONG
EXHAUST DUCT
AIR METERING
ORIFICE PLATE
AIR INLET
3.0 ORIFICE
(SHARP EDGED
AIRFLOW
MEASURING
ORIFICES)
FLOW STRAIGHTENER
COOLING AIR
SUPPLY
(WALL TEMP
CONTROL)
SIMULATED VAPOR
GENERATOR SECTION"
COMBUSTOR
"SECTION
AIR MEASUREMENT
"SECTION
FIGURE 4. PRELIMINARY COMBUSTOR TEST RIG
Metered air is delivered into the combustor from a plenum chamber
that has provisions for installation of air metering orifices to simulate a
part of the aerodynamic interaction of the air metering valve. Air is di-
rected from this valve plate into the combustor outer case for distribution
into the primary air (swirl plate) and secondary air holes located around
the outside of the cylindrical portion of the combustor liner. Fuel is
delivered by means of a dynamic seal into the drilled shaft of the cup motor.
After passing through the rotating shaft to the cup which both atomizes and
distributes the fuel in a flat spray into the 20 inch diameter combustor.
Initial testing was performed with a perforated Inconel radiation shield across
the exhaust plane of the combustor. The majority of tests performed including
the emission results described in this report, were run without the radiation
shield.
Immediately downstream of the radiation shield an asperated thermo-
couple probe has been installed to monitor temperature. The tip consists of a
four inch long triple shielded platinum rhodium (type S) thermocouple. It is
mounted on a long cooled stainless steel tube that allows the probe to be
traversed across the entire diameter of the tail pipe (see Fig. 5). One-half
23
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STAINLESS STEEL
SUPPORT
GAS FLOW TO
ASPIRATOR
AIR COOLING
INLET
•TEMPERATURE
PROBE
JUNCTION
RADIATION SHIELD
'2ND RADIATION SHIELD
'3RD RADIATION SHIELD
FIGURE 5.
SCHEMATIC OF GAS TEMPERATURE SENSOR AND
MECHANICAL SUPPORT
an inch downstream of the thermocouple probe is the simulated air cooled
vapor generator. Welded to the surface of the coils are twelve type K
thermocouples (see Fig. 6 for thermocouple location).
An emission probe is installed at plane 4 inches downstream of the
radiation shield. Figure 4 shows its installation in a twenty inch diameter
duct. A total of 36 holes are used to provide equal area sampling across the
20 inch diameter tail pipe. A large diameter tube collects the gas sample
for delivery to the gas analysis equipment sampling lines. Temperature of
the sampling probe is controlled by cooling water flowing through passages
around the probe. Three thermocouples are located along the length of the
probe to ensure that the sampled gas is quenched to a sufficiently low tem-
perature to prevent further reactions (Fig. 7). It is approximately 42 inches
from the outlet of the tail pipe to prevent recirculation of fresh air into the
probe.
A back pressure orifice plate is installed immediately behind the
emission probe. It is a flat plate orifice with 60 holes located at radial
positions to simulate a uniform flow restriction equal to the remaining flow
restriction of the actual vapor generator. A 48 inch long tail pipe was used
to minimize buoyancy circulation and swirl recirculation of outside air.
24
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FIGURE 6. COMBUSTOR SIDE OF VAPOR GENERATOR SIMULATOR SHOW-
ING LOCATION OF THERMOCOUPLES
DUCT,
• 1
COOLING
EXHAUST
36- 0.03 HOLES
EQUAL AREA SPACED
n
CENTER LINE
OF COMBUSTOR
r<
0.03 IN. HOLE
n
THERMOCOUPLE
TC JACKS
TYPE K
11=-.=.==,
"COOLING PASSAGE
FIGURE 7. EMISSION PROBE
f AIR PURGE
1 CONNECTION
TO FID
SEAL TUBE HEATED LINE
25
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FIGURE 8. TEST COMBUSTOR INSTALLED IN RIG
Figure 8 shows the overall test rig in operation. The large diameter
flange on the right is at the combustor exit plane and is used to seal the tail
pipe to the combustor to prevent possibilities of air leakage into the tail pipe.
Any leakage at this point could cause errors in emissions as measured by
probe in tail pipe. The emission probe location is immediately downstream
of the simulator (Fig. 4). All emissions were measured as described in
Appendix I.
5.2.2 Control of NOX in Combustor
Since liquid fuel (unleaded gasoline) is directly injected into the com-
bustor the combustion process is complex. A considerable effort has been de-
voted to sophisticated computer modelling of the reaction kinetics of rotating
cup combustors. Reference 1 summarizes this analysis. What analysis and
hundreds of hours of low emission combustor development testing proves is
NOX emissions are the critical species and must be minimized by preventing
reaction from occurring at high temperatures. This rate of formation of
NO is:
67000
dt
"(N2)(02)1/2P1/2
1/2
where NO, N2 and O2 are concentrations of nitrogen, oxygen, and nitric oxide
in mole fractions, t is time in seconds, T is temperature °K, and P is
26
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pressure in atmospheres. Although this equation only holds for a homogeneous
mixture, it helps explain the importance of rapid vaporization and mixing.
Droplet burning and combustion during the transition from rich to lean burning
both occur with local zones at stoichiometric temperatures. Only a small
percentage of the fuel reacting at these temperatures will cause unacceptable
high NOX emissions.
Since the local reaction zone air-fuel ratio determines the tempera-
ture, the formation of NO as a function of air-fuel ratio and mixing versus
reaction time rates are as shown in Figure 9. Because NO must be maintained
well below 50 ppm, it is seen that the mixing time must be kept very low and
the air-fuel ratio must be generally leaner than 20 to 1 to maintain reaction
zone temperatures below 3200°F and above 2500°F to keep excessive NO from
forming, but still sufficiently high to complete the oxidation of HC and CO in a
reasonable volume. The important design rules and the methods of implem-
enting used throughout the combustor development are:
• Provide rapid and uniform vaporization prior to reaction (no
droplet burning). The cup causes a uniform droplet spray to be
directed in a 180 degree spray pattern that has a large geometric
surface area at the injection point. Use of swirl flow cause
recirculation of high temperature combustion products to assist
in vaporization. Cup rotation is opposite the tangential swirl
component from the air injection port causing high relative
velocities and consequent rapid vaporization.
• Rapid mixing of vapor with swirl air to obtain (consistent with
parasitic fan power requirements) on overall lean mixture. It
is a design goal to minimize reaction in this zone since the
mixture must change from rich to lean in this zone. Reaction
at near stoichiometric conditions would produce large amounts
of NO thus mixing rates must be made faster than reaction
rates to keep flame temperatures below 3200°F. By main-
taining high air velocities and a high swirl cone angle around
the cup's fuel spray pattern, both of the above conditions are
attempted to be optimized.
• Elimination of small pockets of slowly mixing zones or high
temperature reaction. High air velocities and design with com-
plete symmetry are incorporated to eliminate this cause of NO
formation.
5.2.3 Combustor Rig Test Results
Figure 10 shows the arrangement of the combustor used for rig
test. Photographs of both sides of the rig test unit are shown in Figures 11
27
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10,000
i.ooo U
I
o
z
100 h
- 67000 I- ,1/2
(O2)
("IN;. (02)'/2 P'/Z 1
L T'« J
NO, N2 and O2 = CONCENTRATIONS
GRAM MOLS/CM3
I = TIME, SECONDS
T = TEMPERATURE. °K
P = PRESSURE
10 12 14 16 18 20 22 24 26 28
AIR/FUEL RATIO
FIGURE 9. EMISSION OF NITRIC OXIDE AS A FUNCTION OF AIR-FUEL,
RATIO AND TIME
28
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SECONDARY AIR (92) .359 IN. DIA HOLE
SIMULATOR TUBES
I'll .
TO EMISSION CART
FIGURE 10. RIG TEST COMBUSTOR
FIGURE 11. UPSTREAM SIDE OF COMBUSTOR
-------
FIGURE 12. COMBUSTOR READY FOR INSTALLATION ON RIG
and 12. Rig air is supplied to an annular opening around the inlet to a swirler
section. Air is accelerated inward across the swirler obtaining a rotary
component as it is turned by the swirl blades. Twelve curved sheetmetal
blades were used in the tests reported on in this section. Figure 11 shows
the location and general shape of these swirl blades as traced by their res-
pective weld oxidation lines on the combustor dome. A swirler exit nozzle
is found by the annular opening between the outside diameter of the cup and a
circular hole cut into the combustor dome. In the rig tests a 3 inch diameter
cup driven by variable speed electric motor was used to atomize and dis-
tribute the fuel. A hollow armature shaft was used to supply fuel to the cup.
Both JP-4 and JP-5 fuel was used with similar results. Gasoline was not,
at this stage of the program, specified as the development fuel. A 20 inch
diameter flame tube with a total axial length of 8 inches before the first
row was used in all tests. No film cooling is provided on either the vapor
dome or flame tube to avoid possible emissions complications. All cooling
is by convection from the air passages on the outside or by means of radiation.
Inconel 600 was used for both the dome, flame tube and flexible flame tube
seal at the exit end of the flame tube. Inconel held up relatively well even
at gas temperatures well above the design point of 2500°F. Oxidation was
no problem but some serious distortion of the dome and cracking of the flex-
ible flame tube to outer housing seal was observed after less than 50 hours
of operation. An earlier prototype combustor fabricated from 321 stainless
steel showed rapid deterioration due to oxidation and distortion.
JO
-------
Two basic configurations were tested on this rig combustor. Initially
all of the air was supplied over the cup in the annular port around the cup.
This would permit the fastest and most uniform vaporization, mixing, and
reaction. However, to achieve relatively good emissions, the
annular air injection port had to be reduced in area to minimize internal
recirculation into the swirler. Thus the pressure drop at high flows was
much greater than the maximum limit of 5 inches. In order to allow more air
to be injected into the combustion zone a "secondary air" admission system
was devised. A series of ports were machined into the flame tube to increase
the total flow area of the combustor. Tests were performed with both con-
figurations. Initial tests were performed without the secondary flow ports in
the flame tube. All air entered around the cup through the primary or "high
pressure air injection arrangement". After completion of these tests the
secondary ports were machined into the combustor, thus allowing injection of
air through both the primary and secondary ports simultaneously. This split
flow configuration or "low pressure loss air injection arrangement" yielded
data for the high fuel flow end of the power range.
Figure 13 shows the overlapping NOX emission results with the two
configurations analyzed by rig tests. With air admission through both the
primary and secondary ports, the emission characteristics are seen to
steadily rise as the fuel flow is decreased from 135 to 60 pounds per hour
(pph). At a point near maximum allowable pressure drop of 5 inches of water,
the NOX level was well below the goal of 1.38 gm/Kgm. As the pressure drop
and consequently the mixing velocities dropped with lower fuel flows, the
emissions increased. At approximately 80 pph fuel flow and an air side
pressure loss of 1 .4 inches of water, the NOX level exceeded the program's
goals. During these tests the air cooled vapor simulator could not be kept
supplied with enough air (because of rig air limitations) to keep its tempera-
ture sufficiently low to prevent it being damaged. As a consequence, it
required removal at all test points above 60 pounds per hour of fuel flow.
Since the emission averaging probe was fixed downstream of the simulator
coils, their removal at high fuel flows effectively increased the actual com-
bustion length by approximately 4 inches. All test points below 60 pph were
with the two coils that comprised the simulator cooled to below 1000°F to
provide both quenching and radiation cooling of the reaction products 8 inches
from the dome.
With only the primary air being admitted into the combustor over the cup,
the "S" shaped "high pressure loss" emission characteristic was obtained.
At a pressure drop of 4.25 inches of water good NOX emission levels could be
obtained at a fuel flow of 70 pph. (It should be noted that a pressure drop of
approximately 16 inches of water would be necessary to allow enough air into
the combustor at the maximum rated fuel flow of 135 pph with this high pressure
loss configuration.) As in the case of the low pressure loss configuration,
the emissions increased as fuel flow and corresponding air pressure loss
31
-------
Lt.
E =
01
-X
01
CM
2.0
1.8
1.6
1.4
1.2
1.0
0.8
0.6
0.4
0.2
1.38 N02 LIMIT
HIGH PRESSURE LOSS
AIR INJECTION ARRANGEMENT-
LOW PRESSURE LOSS
AIR INJECTION ARRANGEMENT'
0 20 40 60 80
FUEL FLOW (pph)
100
120 140
FIGURE 13. NO2 EMISSIONS VERSUS FUEL FLOW
was decreased. A reversal of this trend occurred when the vapor generator
simulator was reinstalled into the rig at flows below 60 pph. A decrease in
NOX emissions resulted until again the characteristic increase in emissions
as air velocities fell to extremely low levels. NOX emissions exceeded the
limit at 10 pph fuel flow and an air side pressure drop of approximately 0. 1
inch of water .
The drop in NOX emissions when the vapor generator was installed was
thought to be caused by radiation cooling by the cold walls of the combustor.
Another factor that could have been important is the effects upon swirl air
flow patterns and recirculation zones within the reaction zones.
CO and HC emissions were low throughout all these tests (Figs. 14 and
15). HC emissions were generally below the background levels and thus
showed little or no effects of quenching problems on the vapor generator
simulator tubes. At full fuel flow a temperature traverse (Fig. 16) showed
good uniformity with deviations well within the ±250°F goal. The average
temperature is approximately 200°F below the theoretical exit temperature.
This is again within the expected limits since the combustor has a highly
radiant flame that loses energy by direct radiation.
Results of these tests established the approach to be taken in the fully
integrated system using fan air, control valves and the steam generator.
Since integration with these components was certain to have significant
effects upon emissions, further development on the rig could not be justified
because of program scope limitations and lack of adequate flexibility to simulate
all of the important fan, valve and steam generator effects upon performance.
-------
25
LJ
cn
E
cn
O
O
^U
18
16
14
1?
10
8
6
4
2
n
-
-
11.8 CO LIMIT
-
-
-
-
-
• ^**~~ i — ._«- i • ii r1 , , , , , *
20
40 60 80 100
FUEL FLOW(pph)
120 140
FIGURE 14. CO EMISSIONS VERSUS FUEL FLOW
LL
E
cn
E
cn
O
T
£ .U
1.8
1.6
.4
1.2
1.0
0.8
0.6
0.4
0.2
_
1.42 HC LIMIT
-
-
-
-
-
>>^^
ll II l V l «l • 1 1 1 1 T-
20 40 60 80 100 120 140
FIGURE 15. HYDROCARBON EMISSIONS VERSUS FUEL FLOW
33
-------
O.D. OF
FLAME TUBE-
10
CENTERLINE OF
COMBUSTOR
CO
3
Q
<
o;
C£.
o
h-
00
^
CQ
2
o
o
135 LB/Hrt 8.9% C02
AVERAGE
TEMPERATURE 2329°F
DISTRIBUTION +96°F
-124°F
i i
2100 2300 2500
VAPORIZER INLET TEMPERATURE (°F)
FIGURE 16.
COMBUSTOR EXIT PLANE RADIAL TEMPERATURE
DISTRIBUTION AT MAXIMUM FLOW
As a consequence of these rig tests a variable geometry combustor was
required. From design studies a mechanically simple system, the split
flow combustor, was designed and put into final integrated systems tests.
5.3 INTEGRATED COMBUSTOR PERFORMANCE
5.3.1 Design
Figure 17 is a cross section of the integrated combustor, air supply
and steam generator. All test data reported from this section was with the
fully integrated system. Air was supplied by the systems integral fan and the
steam generator was operating at approximately 1000°F and 1000 psia. As
discussed in the previous sections, parasitic power limitations required the
use of a type of variable geometry combustor defined here as a "split flow"
arrangement. Practical mechanization of this requirement was achieved by
utilization of a simple shear valve immediately upstream of the combustor"s
injection passages. A detailed description of the system is given below.
Air is drawn into the unit by a fan mounted on the top of the unit on the
centerline axis of the combustor tube and steam generator coils. Coaxial
with the fan is the rotating cup for atomization and distribution of the fuel.
Both the fan and cup are belt driven by a DC electric motor mounted to the
side of the combustor within the 24 by 24 inch square envelope (see insert at
bottom left of Fig. 17 and Fig. 18 and 19). Between the blade tips of the fan
34
-------
1 PRIMARY AIR
r ^^ 1 1 \ ,--,
[BELT DRIVE \ \
FAN MOTOR
^24 IN.—
Z / H
^ ro
CNJ V (xj
^-~J
' \
,r7r^
T
1
v.
L^j
L--
J
^>
«— i
T"
X
IR IN
I
i
/
-ET
hFUEL TO CUP
— i
[raj "^ FUEL VALVE |
1 J )
\(%~Ll-
^ ^ CUP
ooooo
ooo
^
• —
;
c*
O O O O 0 r
VAPORIZER 2 ROWS
/
OOOOOOOOQ)/X
oooooo
oooooo
ooooo
ooo
oo
f~\
»
^-SUPE
-^
X
:
i OCOOO i
oo
R HEATER
3 ROWb
^PREHEATE
5 ROWS
R
\ '
—
,,
|
t
EXHAUST
FIGURE 17. INITIAL INTEGRATED SYSTEM TEST CONFIGURATION
WITH CONICAL CUP
FIGURE 18. COMBUSTOR/AIR SUPPLY SYSTEM (DRIVE MOTOR SIDE)
35
-------
FIGURE 19. TOP VIEW OF COMBUSTOR AIR INLET TO FAN
and the air control ports a radial flow diffuser section converts velocity head
to 10 inches of water (Fig. 20). Complete symmetry is maintained by spacing
the ports equally around the circumference of the main valve plate. A rotary
shear plate (Fig. 21) meters air to the primary at a ratio and schedule that is
a function of fuel flow. After being metered by the valve, air flows radially
inward through the swirler and then radially outward around the rotating cup
with swirl flow. At above 50 percent power the secondary passages are also
opened allowing air to flow into the combustion zone through 90 holes located
4 inches from the dome in the 20 inch diameter flame tube. Figure 22 shows
the relative location of the 8.8 inch diameter primary port around a conical
cup and the 90 secondary injection ports. A radiation shield is installed
(Fig. 22) directly below and attached to the ID and OD of the dome plate. It
is made from Hastelloy X with eight radial expansion slots to prevent thermal
expansion forces being transmitted into the dome and the swirler blade
assembly.
A four inch diameter conical cup (Fig. 23) was used in the initial stages
of the integrated systems testing. It was a compromise size matched to the
rotational speed of the fan. In combustor rig tests the best results were
obtained at 10,000 rpm with a 3.0 inch diameter cup. However, satisfactory
results were obtained at as low as speed as 4500 rpm. An existing fan was
found to have ideal characteristics (except it operated at 5800 rpm) for both
-------
FIGURE 20. FAN MOUNTED TO CONTROL VALVE BACKUP PLATE
FIGURE 21. ROTARY SHEAR PLATE
7
-------
FIGURE 22. COMBUSTOR/AIR SUPPLY SYSTEM (VAPOR GENERATOR SIDE)
I I !
6 INCHES
I ' I ' 1 ' \.
CHES \
FIGURE 23. CONICAL FUEL CUP
-------
packaging and aerodynamic integration into the system. It was decided to
increase the cup diameter by one inch to more closely match the tangential
velocity and droplet size used in the test rig phase of the program. Two
possibly important side effects of this change were not considered as being
important. One was the heat input surface area of the cup increased by over
70 percent as did the volume of reaction products in the zone immediately
below the cup. As development tests progressed several modifications were
made to improve emissions possibly associated with this change in diameter.
5.3.2 Air Supply System
At maximum rated conditions the air supply and control system must
deliver 3400 pounds of air per hour to the combustion system. Flow control
into the combustor must be symmetrical to prevent hot spots and higher
emissions. A fully modulated control of air mass flow synchronized with the
fuel is required across the entire 40 to 1 heat release range. In addition to
these features, the air supply system must be integrated into the overall
design to minimize the completed package's overall length and volume.
Length can be made a minimum by using a radial fan coaxial with the com-
bustor axis. Figure 17 shows a cross-section of a radial fan with the
desired characteristics integrated into the unit. Figure 24 is a photograph
of the unit after modification required for installation into the combustor.
It is a standard automotive fan designed to produce minimum noise. The
initial unit used was brazed to allow transmission of drive load through its
shroud. Subsequent units did not appear to require brazing. It has 28 fully
shrouded blades with the following parameters:
Tip diameter = 9. 1 inches
Tip width = 1.15 inches
Inlet blade angle, /?j = 57 degrees (to a radius)
Outlet, /?2 = 25 degrees (backsweep)
Number of blades, Z = 28 blades, 0.05 inch thick
A detailed analysis of its performance characteristics at several speeds is
shown in Figure 25. The correct combustor flow and pressure ratio (3400
pounds per hour at 10 inches of water) can be obtained by operating the fan at
5740 rpm. At this speed and flow the efficiency map calculated for this unit
indicates a total efficiency of 66 percent. Figures 26 and 27 are plots of the
pressure and efficiency characteristics at 5700 rpm. Experimental pressure
versus flow test points are also included in this curve.
39
-------
I ' I ' I I I I \
(=; INCHES \
FIGURE 24. AIR SUPPLY FAN
A mechanically simple shear valve has been designed to provide the
split flow characteristics required for combustion. Figure 28 schematically
shows the principle of the valve. A series of twenty contour ports (Fig. 29)
have been cut into a moveable plate that rotates about the axis of the com-
bustor case. A corresponding series of matching ports have been machined
into the backup plate that forms one wall of the diffuser and the support for
the fan bearing mount (Fig. 30). By rotating the metering plate orifice a
total of 34 degrees, the air flow is regulated from 100 percent to less than
2 percent. The arrangement and shape of the ports has been designed to
provide separate control of primary and secondary air into the combustor
case (secondary air) and swirl vane compartments (primary air). A seal
ring is used to isolate and prevent leakage between the primary and secondary
air passages. At the 50 percent flow position, the secondary air is completely
shut-off and all air is directed into the primary zone through the swirler.
Basically for this geometry of combustor, it was determined that a flow split
of 0. 55 pounds per second into the primary zone and 0.42 pounds per second
of air into the secondary zone gave optimum emission performance at full
power conditions (135 pounds per hour fuel flow). A maximum pressure drop
of 5 inches of water was used for full flow conditions. At lower power levels
tests demonstrated that mixing rates could not be maintained due to low
pressure drops if fixed flow areas were maintained across the combustor.
One approach that was shown satisfactory by rig tests was to inject all the air
40
-------
EFFICIENCY
20
11,470 RPM
456 fps
tip dia. 9.1 in.
tip width 1.15 in.
^56.6° (to rad.)
9 25° backsweep
28 blades-.050 in. thk
15
8,600 RPM
342 fps
Desired Operating
point for
Combustor
psia
FIGURE 25. CALCULATED CHARACTERISTICS OF FAN
41
-------
3472
AIR FLOW IN POUNDS PER HOUR
FIGURE 26.
TOTAL PRESSURE RISE CALCULATED COMPARED
TO TEST POINTS
through the swirler at low flows. This approach could also provide low
emissions at the high flow rates but an unacceptably high pressure drop would
be necessary (16 inches of water) if 100 percent of the air was injected through
the swirler.
Ten primary and secondary ports are used to obtain good symmetry
and uniform air distribution within the combustor. Design of the valve is
dependent upon matching it with the fan, diffuser, combustor and vapor
generator characteristics. Rig tests were performed to obtain the basic
data necessary for design. Figure 31 was the flow rig designed to simulate
the actual metering valve flow interaction with fan and diffuser. Tests
results are compiled in Table III. Figure 27 gives static fan pressure rise
versus flow. Figure 32 shows the variation of the coefficient of discharge, Cd,
for the air metering orifices versus flow and orifice open area Am (in.^).
The term, Cd, is used in the following equation for calculating volume flow
across the orifice.
W
m
= (7.617) (60) CdA
m
- PV)
(i)
42
-------
MAXIMUM REQUIRED
FLOW
3485 4646
AIR FLOW IN POUNDS PER HOUR
FIGURE 27. ACTUAL STATIC PRESSURE RISE COMPARED TO CALCULATED
TOTAL PRESSURE
where W is flow across metering orifice (Ib/hr)
•j
y is density of air in lb/ft^
Ppe - Py is pressure difference across metering plate (inches of
water column)
Cd is found through the above equation (1) and the relation:
W
m
where Wy is flow calculated from the downstream plenum duct pressure Py
exit area (Ay) data. The coefficient of discharge for the plenum duct orifices
is assumed to be 0.65 for all cases. Figures 27 and 32 contain the important
interface data necessary to design the metering valve. From Table III and
Figure 27 it is seen that no significant stall characteristics exist with the
fan- diff user arrangement.
Provisions had been made to fabricate a bypass valve to prevent fan
stall instabilities from causing maldistribution of air into the combustor.
However, since the fan exhibits a near uniform static pressure characteristic
43
-------
BVCKUP PLATE
OPENING
SECONDARY
PORTS
PRIMARY
/ PORTS
CENTER OF
COMBUSTOR
FIGURE 28. PRIMARY AND SECONDARY AIR METERING PORT
CONFIGURATION SHOWN IN 50 PERCENT POWER
POSITION
44
-------
FIGURE 29A. AIR VALVE IN 2 PERCENT POSITION
FIGURE 29B. AIR VALVE IN 50 PERCENT POSITION
45
-------
FIGURE 29C. AIR VALVE IN 100 PERCENT POSITION
FIGURE 30. AIR VALVE BACKUP PLATE
-------
BYPASS
PORTS
FAN
FIGURE 31 A. FAN-AIR VALVE FLOW TEST RIG
FIGURE 3 IB. FAN TEST RIG DIFFUSER AIR VALVE SECTION
-------
TABLE HI
FAN TEST RESULTS WITH NO BYPASS
FAN SPEED 5700 RPM
Fan Exit
Total
Pressure
PT
(in.H20)
13.5
10.3
11.1
8.7
11.0
12.8
11.5
10.9
10.2
11.6
11.2
13.9
12.2
10.0
10.0
10.4
13.3
12.8
12.2
11.8
13.1
12.8
12.6
12.4
12.4
12.4
Dynamic
Pressure
PD
(in. HO)
Zi
4.9
3.3
3.9
3.8
2.4
4.2
4.0
4.1
2.1
4.4
2.2
5.0
3.6
2.5
2.5
3.6
4.4
3.9
3.2
2.7 .
4.2
4.1
4.2
4.2
4.2
4.1
Static
Pressure
Upstream
Metering
Orifice
PFS
(in.H20)
10.9
9.75
9.15
7.35
11.25
11.0
10.2
9.4
10.35
11.05
10.2
10.95
10.5
10.05
9.9
9.9
10.95
10.65
10 . 45
10.35
10.85
10.55
10.45
10.35
10.35
10.4
Metering
Orifice
Area
AM
(in.2)
40
40
40
40
20
20
20
20
20
20
20
10
10
10
10
10
2
2
2
2
1
1
1
1
1
1
Static
Pressure
at Duct
O.D.
Pv
(in.H20)
10.45
8.7
6.85
3.05
7.5
9.5
4.45
1.55
7.4
9.6
8.05
8.2
10.4
4.75
2.1
0.7
1.7
6.05
8.4
10.1
6.5
4.15
0.95
0.3
0.1
0.0
Duct
Exit
Open
Area
Av
(in.2)
4.71
9.42
18.8
37.7
9.43
4.72
18.9
37.7
9.43
4.72
7.07
4.72
0
9.43
18.9
37.7
4.71
1.57
0.785
0
0.785
1.57
4.71
9.42
18.8
37.7
Flow
Leaving
Duct
wv
(Ib/hr)
1230
2244
3975
5319
2100
1180
3240
3820
2085
1190
1630
1100
0
1670
2330
2560
496
312
184
0
162
258
371
417
480
.___
Coefficient
of Discharge,
Cd, for
Metering Orifice
.34
.44
.53
.52
.433
.385
.5397
.545
.485
.395
.444
.530
.580
.667
.674
.656
.585
.517
.625
.821
.969
1.05
1.21
48
-------
FIGURE 32.
1000 1500 2000 2500 3000 3500 4000 4500 5000
VOLUME FLOW LEAVING DUCT, Wy (LB/HR)
Cd VERSUS FAN FLOW FOR SEVERAL METERING
PLATE ORIFICE AREAS
from 100 percent down to complete shutoff of the air valve, it was possible
to design a simple single shear plate valve without the complexity of a bypass
valve. Prior to final sizing of the valve ports, it was necessary to establish
the actual coefficient of discharge and its variation with flow and port opening
(Fig. 32). Since the velocity of approach is relatively high from the diffuser
to the ports and at 90 degrees to their flow orientation, a nonuniform Cd is
expected. Figure 32 indicates the variation between a 40 to 1 area ratio of
port sizes is acceptable and is within ±10 percent of a Cd of 0. 5 at the area
size approximately proportional to the required flow shown by dotted lines.
Port sizing is based upon a Cd of 0. 5 and application of the flow
equation (1). To permit a reasonably small radial valve dimension a full
flow pressure drop of 2 inches of water has been used. With this pressure
drop the full flow areas for primary air (0. 55 pounds per sec) is: 22.3 in^
and secondary air (0.45 pounds per sec) is 16.9 in^ for a total flow area of
39.2 in^. Ten ports each are used for the primary and secondary flow con-
trol areas. Ports are equally spaced around the valve to provide symmetrical
air flow. In the case of the primary air the ports are located to individually
feed air into 10 separate swirl chambers to further ensure uniform air
distribution (see Fig. 28). A linear response to valve position is provided
by contouring the ports to account for the variation of pressure drop across
the valve at reduced flow conditions. A constant fan discharge static
pressure of 10 inches of water is maintained across the entire flow range by
49
-------
use of a nearly constant speed 28 volt compound wound DC motor. At full
power the pressure drop is 10 inches of water, 2 inches across the air metering
valve, 5 inches across combustor and 3 inches across the vapor generator. At
the 3 percent power level there is nearly zero pressure drop across combustor
and vapor generator with the full ten inches pressure drop taken across the
metering valve. Linear response is obtained by calculating the open area for
each valve position using the power level dependent pressure drop across the
valve. Resulting contours are thus high at the 100 percent flow side and low
at the reduced power positions (Fig. 29).
A DC electric motor drives the fan through a belt. A search for a
compact motor with good efficiency resulted in selection of a Model 1000
Hoover Electric unit. Availability, small size (Fig. 33) and light weight
(13 pounds) were the primary factors used to finalize its selection. Efficiency
at maximum rated conditions is only slightly better than 60 percent (see Fig. 34).
Since the ideal air horsepower is 1.2 HP and the fan is 68 percent efficient,
the shaft power into the fan is 1.8 HP. From Figure 34 the electrical power
input for 1.8 HP is 2120 watts or 2.8 HP (electrical equivalent). This is
slightly above the goal of 2. 5 HP but could be improved to less than 2. 5 HP
if a motor with an efficiency of greater than 72 percent were installed.
Although aircraft motors can be made in this range, no automotive motor of
this efficiency has been located and may require special design.
FIGURE 33. FAN DRIVE MOTOR
50
-------
10000
8000
3
2
hQ OVEFJLI.'-I£TJ? 1C CPUPANJf
MODEL 1000^PERFORMANCE CURVE
28 VOLTS D.C.
60
o
u.
6^00 u 60
in O
z ^
o 4000 £
P Q.
3 1
2000 g
' £C
o:
3
O
— , . __
/
^-
\/'
R.P.M.^
:FF:C;E>:CY
x
X
x
^-"
Out
~X
^*"
,-- — ^^
3Ut POW
„.,-•-"
^
.T^TM.
3r Requi
E
red at F
^
lectrica
an Effic
-^"
Input P
iency =
zs
^
^•^
/
ower =
=
_ —
O.fifi
1.8 fTp*
i^f
7 —
/
2120 W
2.8 HP
02 0.4 06 08 1.0 1.2
OUTPUT HORSEPOWER
I 6
! 8
FIGURE 34. FAN DRIVE MOTOR CHARACTERISTICS
5.3.3 Flame Performance
An initial step in the evaluation of the integrated system was a visual
examination of the flame. Symmetry and flame length can be quickly evalua-
ted by this procedure. These tests were performed in a horizontal position
with all system components installed except for the steam generator.
Figure 35 is a schematic of the test arrangement and instrumentation
used to make a preliminary evaluation of the closely coupled air supply and
combustor. At a given air valve setting fuel flow was varied to obtain
optimum flame performance as determined by visual inspection. Static
pressure rise across the fan and pressure drops across the primary and
secondary ports of the combustor were measured by three manometer read-
ings each (located approximately 120 degrees around the circumference of
the unit). Current and voltage to the drive motor were recorded as was the
fan speed. Figures 36 through 40 are records of flame performance during
the integrated systems tests. In general the results were acceptable from
visual symmetry and flame length considerations. Potential problems such
as diffuser instabilities and flow separation due to interaction between fan,
diffuser, air valve and swirler did not appear to occur, or were not significant
enough to cause apparent asymmetry or downstream hot or cold patterns in
the flame. Good correlation between design point and actual pressure drops
across metering valve and combustor were also recorded. One major
51
-------
ROTATIONAL
AIR VALVE
POSITION 0 to 100%
SECONDARY
AIR
PRIMARY
AIR
FAN BLADES
PULLEY
DRIVE
FUEL IN
Wf
2 to 135 PPH
48-INCHES LONG
EXHAUST DUCT,
FIGURE 35. COMBUSTOR/AIR SUPPLY SYSTEM TEST
ARRANGEMENT
52
-------
FIGURE 36. FLAME AT 1. 5 PPH
FIGURE 37. FLAME AT 60 PPH FUEL FLOW
;
-------
FIGURE 38. FLAME AT 80 PPH FUEL FLOW
FIGURE 39. FLAME AT 100 PPH FUEL FLOW
-------
4 Inch Cup
FIGURE 40. FLAME AT 130 PPH FUEL FLOW
problem discovered was with the matching of fan and motor characteristics.
Electrical input power (see Table IV) ranges from 1100 watts (1. 5 HP) at 3
percent flow to 2360 watts (3.2 HP) at 96 percent flow. At 22 percent power
(near the average duty cycle requirements) the electrical power required was
1.7 HP. All of the input power levels were well above calculated values.
Analysis of the problem indicates that a mismatch between the fan load speed
characteristic and the motor speed-voltage-curvent relationships was the
cause. In order to operate the fan at 5700 rpm voltages below design point
were necessary. As a consequence, higher than design currents were recorded,
increasing I^R losses and reducing efficiency. A new pully drive ratio was
installed to obtain a better speed match and has reduced the input power
requirements by allowing the motor to operate at higher speeds and voltages
(28 volts).
Table V lists results of several of the test cell operations with com-
plete system including the vapor generator. Current values range from 55 to
73 amps with the new pulley arrangement. Previous results in the same flow
range varied from 70 to 100 amps or a reduction of approximately 30 percent.
-------
TABLE IV
INITIAL COMBUSTOR PLUS FAN AIR SUPPLY TEST RESULTS
Power
Level
%
3
7.5
22
48
55
59
74
85
96
Fuel
Flow
(Ib/hr)
4
10
30.3
65
75
80
100
115
130
Air Valve
Position
% Open
5
10
20
40
60
60
80
90
97
Fan Discharge
Static Pressure
(in. of water)
10.3
11.2
10.5
9.5
8.4
8.1
7.6
7.1
7.0
Combustor Pressure
Drop (in. of water)
Primary Secondary
Plp P2s
<0.1 <0.1
<0.1 <0.1
<0.2 <0.1
1.2 <0.1
5
4.5 0.4
4.5 2.4
4.9 3.9
4.9 4.6
Fan
Speed
(rpm)
5900
6150
5900
5750
5500
—
5400
— .
—
Motor Input
Power
(Volts) (Amps)
21 53
22.8 60
23 55
23 70
23 80
22.5 83
22.5 94
22.5 100
22.5 105
TABLE V
FAN POWER WITH HIGH RATIO PULLEY
Fan
Water
Air
Air Valve Position
Fuel Flow
Fan Speed
Volts
Amps
Flow
Press In
Press Out
Temperature Out
Fan Out
P rimary
Secondary
V.G.
%
PPH
RPM
Volts
Amps
PPH
PSIG
PSIG
OF
In. ofW.C.
In. of W.C.
In. of W.C.
In. of W.C.
Run No.
#1
46
56.5
5700
24
55
1600
370
260
400
8.7
1.0
0.25
0.4
#2
69
83
5700
25
65
1600
580
460
454
8.6
5.5
1.0
0.8
#3
82
94
5680
25
71
1600
660
530
468
8.5
5.5
2.3
1.0
#4
90
101.5
5680
25
73
1600
690
590
473
8.5
5.5
3.0
1.15
56
-------
5.3.4 Integrated System Emissions Evaluation
Initial combustor test and development was performed with all the
systems and components assembled as shown in Figure 17. Steam conditions
at the outlet of the unit were kept at design conditions to ensure quenching
radiation effects and aerodynamic recirculation of the steam generator were
correct for emissions analysis.
The emission probe was located two inches downstream of the vapor
generator in an exhaust plenum (Fig. 41). It has 36 sampling ports located
across the major diameter of the vapor generator. A Beckman Model 315A
infrared analyzer (using a 41 inch long 150 ppm full range nitric oxide NDIR
analyzer) was used for NO, CO and CO2 measurements. Beckman's Model
402 FID hydrocarbon analyzer was used for HC emissions. Thermo Electron
Corporation's Chemiluminescent Analyzer Model 10A equipped with a high
temperature-thermo-reactor -was used to measure NO and NO + NO2« Values
of N©2 reported were recorded by the chemilumine scent analyzer. Beckman's
NDIR nitric oxide analyzer was operated in parallel with the Chemiluminescent
unit as a double check by an independent measurement method. Good correla-
tion between the two methods of measurement were observed throughout the
tests. In general, the NDIR would be two or three parts per million higher
than the Chemiluminescent analyzer. This bias is caused by the much greater
sensitivity of the NDIR to water vapor in the exhaust. By replacing the
"Drierite" filter immediately upstream of the NDIR analyzer every few minutes
the high bias of the NDIR could be eliminated. Difficulties due to water conden-
sation with the Chemiluminescent analyzer have been resolved by use of a water
preheater and an electrically heated sampling line. Water condensation on the
last rows of the vapor generator and in the sampling and analyzing portions of
the system were an emissions measurement problem during the first weeks
of testing. Addition of a water preheater to the feed water circuit increased
the inlet temperature to over 185°F. This temperature simulates the condenser
outlet temperature in an automobile. At 185"F the tube wall temperature was
well above the dew point of water vapor in the exhaust and no further problems
were observed with either NOx of HC measurements.
Little or no NO£ formation has been observed in the exhaust from the
vapor generator. Two or three parts per million of NO2 have been observed
in some of the initial tests but normal test results show only a scatter that is
probably inherent in the instrument's basic accuracy. (See Appendix I for a
more detailed discussion of the emission measurement instrumentation.)
The initial rig tests used a three inch diameter cup normally operating
at 10, 000 rpm. Design of the integrated combustor air valve, fan and vapor
generator required the fan and cup shaft to operate at a speed between 5, 000
and 6,000 rpm to match fan characteristics. From variable speed tests and
analysis of the three inch diameter cup, it was determined that good atomization
57
-------
DUCT
LOOOOOOOOOOOOOOOO
o
COOLING
EXHAUST
36- 0.03 HOLES
EQUAL AREA SPACED
TC TC | TC
CENTER LINE
OF COMBUSTOR
GAS FLOW
\ \
0.03 IN. HOLEX /THERMOCOUPLE
SEAL TUBE
FLEXIBLE
HOSE
TO NDIR
HEATED LINE
4 AIR PURGE
1 CONNECTION
TO FID
HEATED LINE
COOLING PASSAGE
FIGURE 41. EMISSION PROBE LOCATION AND CONFIGURATION
could be obtained at lower speeds but a larger diameter cup would be re-
quired. A four inch diameter conical cup was incorporated and tested in
the complete system. Emissions of HC and CO were well within limits,
however, NOX emissions were slightly above limits. At fuel flows of less
than approximately 20 pounds per hour, it was not possible to lower emissions
of NOX below the 1976 limits. Analysis of these results indicated that the
probably cause of the problem was a lack of adequate mixing rates because
of the low pressure drop across the combustor's air swirler at low air
flows. In particular it was theorized from flame observations, that a
relatively slow burning mass of fuel vapor was reacting in a "dead" region
immediately downstream of the cup. It was believed that fuel vapor was
trapped in this low velocity zone and consequently reacting stoichiometrically
thus driving the NOX above limits as the fuel vapor reacted from rich through
to overall lean. Correction of this problem has been based upon the basic
requirements to minimize NOX. Essential for low emissions in a combustor
in which fuel is introduced as a liquid, is the need to complete combustion
with an overall lean air-fuel ratio. Production of NOX is exponentially
dependent upon local temperatures, and availability of oxygen. Since fuel is
injected as a liquid, some zones of stoichiometric (maximum temperature)
will occur prior to transition to an overall lean condition unless mixing rates
58
-------
are made sufficiently fast. A key feature of the combustor is to obtain a
large area (wide spray angle) interface between liquid fuel and air at a zone
that has very high shear velocities. Under these conditions, vaporization and
mixing can occur extremely rapidly. Since chemical reaction rates are
relatively slow compared to the high mixing rates possible, when a high
pressure drop is used, NOX production is minimized by reducing stoichio-
metric conditions. In the actual combustor, the mixing rates and thus the
formation of NOX is, to a large extent, dependent upon the pressure drop
available. Thus a fundamental tradeoff between parasitic power limitations
and acceptable NOX levels has been made. Ideal power required for this com-
bustor has been established at 1.2 HP (approximately 2.5 HP electrical input).
In addition to the basic pressure drop-mixing rate tradeoff, it is essential
that little or no droplet burning take place and that no pocket of vaporized
fuel is allowed to mix slowly and react stoichiometrically.
Improvements of emissions at the low fuel flows is thus dependent
upon achieving three goals.
•Rapid and uniform vaporization proper to reaction (little or no
droplet burning)
•Elimination of pockets of stoichiometric fuel and slow mixing
•Increasing mixing and vaporization rates
All of these goals can be completely dependent upon mixing rates and recir-
culation or they can be independent as in the case of atomization and droplet
control. Three modifications have been incorporated to assist to reduce
NOX formation.
•Double heat shield (see Figs. 42 and 43)
Two heat shield discs with a dead air space between have been
incorporated to prevent excessive vaporization from cup surfaces
and consequent carbon buildup. This may also assist in the mix-
ing since excessive fuel vapors generated in the cup are not
accelerated tangentially at the lip of the cup as are the liquid
droplets. This gaseous fuel may mix slowly down stream of the
cup in a dead zone or be entrapped in local recirculation zones
inside the swirler.
• Auxiliary Air Swirler
Approximately 3 percent of rated air flow has been routed through
an auxiliary air swirler (Fig. 42). High velocity mixing rates
can be provided by this method since the full fan discharge air
59
-------
AIR AT FAN
DISCHARGE PRESS
AUXILIARY AI
SWIRLER
FUEL —
REFERENCE LINE IN
LINE WITH COMBUSTOR
DOME
AUXILIARY AIR
SWIRLER EXIT
DOUBLE HEAT SHIELD
CONFIGURATION
DEAD AIR SPACE
_ £ COMBUSTOR
RECIRCULATION FAN
CONFIGURATION
117 PPH AUXILIARY AIR
SWIRLER FLOW AT 120 DEC.
CONE ANGLE WHEN B = ZERO
CUP DIAMETER = 4 INCHES
NORMAL SPEED RANGE = 5000
TO 6000 RPM
FIGURE 42.
CYLINDRICAL, CUP CONFIGURATION FOR
EMISSION TESTS
pressure (10 inches of water) is maintained across this auxiliary
air swirler. In addition, the swirler provides a wide spray
angle and recirculation in the critical fuel to air interface region
adjacent to the lip of the cup. Measured air flow at 10 inches of
water pressure drop is 117 pounds per hour at a 120 degree cone
angle when "B" dimension equals zero. Figure 1 shows schema-
tically the flow route from fan discharge to auxiliary air swirler
passage. Air short circuits the air valve through 20 ports (see
60
-------
FIGURE 46. RECIRCULATION FAN ON CONICAL CUP
•
CO LIMIT 11.8 urn Kgm
E
tc -
NOTE: NOX FOR ALL POINTS A I I IS
LIMIT OF 1. 3H urn Kirm
FUEL: JP--4
11C LIMIT 1. 1^ gin Ka
10 :>0
\V, 1.11 MR FUEL
'
FIGURE
47. CONICAL CUP EMISSIONS WITH RECIRCULATION FAN
(CO2 ADJUSTED TO MAINTAIN NOX AT 1.38 GM/KGM)
63
-------
maintained at its limit of 1.38 gm/kgm by adjustments to the air-fuel ratio.
CO and HC emissions were good down to fuel flows as low as 2.8 pounds per
hour. This is the lowest fueling rate at which all emissions have been brought
within the limits. An emissions peak at 30 pounds per hour has not been fully
explained but may be an atomization problem associated with fuel impingement
on the cup or its shroud. Another possible problem is the counter rotation of
the fan swirl from the main air swirier which will produce lower mixing
velocities at a specific air flow.
To obtain a flexible cup atomizer and incorporate heat shields, auxiliary
air swirler and a recirculation fan the cup was redesigned to a cylindrical
configuration shown in Figure 42. A more stable bearing stack was also
incorporated to eliminate a wobble that was noted with the conical cup. A
cylindrical cup configuration was chosen for better control of vaporization and
to allow wakes from the center support spider be uniformly distributed prior
to atomization. A number of different fuels have been tested with this final
configuration. Auxiliary swirler dimensions A and B have been found to be
critical variables. With "A" small it has been observed that the fuel can be
drawn back into the auxiliary swirler causing impingement on the outer lip of
the auxiliary air swirler. Since the fuel accumulates into large drops and then
reatomizes from the swirler with relatively large droplets, the emissions
increase.
Test results are graphically shown in Figures 48 through 55. Initial
tests and developments, including rig tests were mainly performed with
JP-5 and JP-4. JP-4 was the initial fuel used for combustor development
on the integrated combustor steam generator tests. Figures 45 through 49
show emission test results with JP-4. Since the NOX emissions are close to
the limit, most of the test points recorded were obtained by adjustments of
the fuel-air ratio to obtain best NOX or CO emission levels. In general, this
was accomplished by moving the air valve to the desired flow position. The
mechanically linked fuel valve would also move to the same percent flow
position giving a nominal air-fuel ratio of 25 to 1 or approximately a CO2
reading of 8.3 percent. By adjusting the fuel pressure across the fuel meter-
ing valve, the air-fuel ratio could be changed to a richer or leaner point.
A direct tradeoff between NOX and CO emissions can be made by adjustments
of the fuel-air ratio at a particular air valve setting. As the air-fuel ratio
decreased (CO2 increased) the level of CO emissions can be drastically
reduced. This is due to an increase in temperature in the CO burnout zones
resulting in more complete reaction. A corresponding but less sensitive
increase in NOX results from a decrease in air-fuel ratios. In these tests
the data points were obtained by adjusting the fuel-air ratio while monitoring
and trading-off NOX against CO levels. In general, two readings were obtained
at each power level. One data point recorded was the NOjj level with the CO
level allowed to approach its maximum specified limit. A second point was
64
-------
LIMIT
10
o
o
.9
.8
? .7
IS .6
§
.4
2.2
2.0
1.8
-------
.10
__ 16
1
a
o 8
o
4
3.0
# 2.0
a
o
1.0
.2
^ 1.8
a
fi. 1.4
-------
9 19
FUEL FLOW - POUNDS PER HOUR
41 46.9 62.6 75 80
3.0
2.0
Ol
CM
O
1.0
Configuration:
T
T
Fuel JP-4
Auxiliary Air Swirler = 3%
2 Heat Shields
A =0.125
B = Zero Numbers Indicate CO9 in
Carbon: Clean v
.40 gm/mile N02 LIMIT'
20 40 60
AIR VALVE POSITION"/*
80
124
1.38 N02 LIMIT gm/kgm
.1
0
100
FIGURE 50. NO2 EMISSIONS WITH A = 0. 192, 2 HEAT SHIELDS AND
AUXILIARY AIR SWIRLER
o
o
01
14.0
12.0
10.0
8.0
6.0
4.0
2.0
0
FUEL FLOW - POUNDS PER HOUR
9 19 41 46.6 62.6 75 8O
124
l l
11.8 CO LIMIT gm/kgm
3.4
680» gm/mile
-Configuration:
Auxiliary Air Swirler
2 Heat Shields
Numbers Indicate
• 7.70
in
O
o
Ol
0 10 20 30 40 50 60 70 80 90 100
AIR VALVE POSITION %
FIGURE 51. CO EMISSIONS WITH A = 0. 192, 2 HEAT SHIELDS AND
AUXILIARY AIR SWIRLER
67
-------
1.4
1.2
1.0
CJ
I -8
£
*>S
* .6
.4
.2
FUEL FLOW - POUNDS PER HOUR
9 19 41 ^46.6^62^ 75 80
1.42 HC LIMIT cjm/kgm
Configuration:
Fuel JP-4
Auxiliary Air Swirler =
2 Heat Shields
A = 0. 125
B = Zero
Carbon: Clean
.41 gm/mile
Numbers Indicate CO2 in
x
.40
.30
.20
.10
o
-
E
•
"0 10 20 30 40 50 60 70 80 90 100
AIR VALVE POSITION %
FIGURE 52. HC EMISSIONS WITH A = 0. 192, 2 HEAT SHIELDS AND
AUXILIARY AIR SWIRLER
FIGURE 53. COMBUSTOR POST TESTS RECORDED IN FIGURES 50,
51 AND 52
68
-------
20
16
12
O *
LIMIT
O
a
2.8
2.4
2.0
3
IN
1.2
NOT RECORDED
LIMIT
1O 20 30 40
FUEL FLOW POUNDS PER HOUR
CONFIGURATION
Fuel - Gasoline
Auxiliary Air Swirler = 3%
2 Heat Shields
A = 0.192
B = Zero
Carbon: Inside of Cup and
indication of combustion on
outer lip of cup
FIGURE 54. EMISSIONS WITH GASOLINE, A = 0. 192 AND 2 HEAT SHIEX.DS
69
-------
16
14
12
1O
8
6
4
2
0
0.8
I0'6
O
o
O
s
0.4
0.2
LIMIT = 11. 8
HC LIMIT = 1.42
CONFIGURATION
Fuel: Gasoline
Auxiliary Air Swirler = 3%
Recirculation Fan
A = 0.03
B = 0.5
Carbon: Clean
1.5
1.0
g
0.5
LIMIT = 1.38
5 10 20 30 40 50 60 7O 80
FUEL FLOW POUNDS PER HOUR
FIGURE 55. EMISSIONS WITH GASOLINE, A « 0.03, AUXILIARY AIR
SWIRLER AND RECIRCULATION FAN
70
-------
the NOX level with low CO reading. By analyzing these limits it is possible
to obtain approximate values of the tolerance band within which the controls
are required to hold the air-fuel ratio to obtain low emissions across the
full range.
The first tests using the cylindrical cup and dual heat shield were
disappointing. Figures 48 and 49 show the emission levels for fuel flow
ranges between 16.8 and 70 pounds of fuel per hour. The first data point on
Figure 48 was at 16.8 pounds per hour with the air-fuel ratio adjusted to move
the CO level to near its maximum allowable limit of 11. 8 gm/Kgm. The CO£
reading at this point was 4.8 percent. At this condition the emissions of
NOX were well above the limit (1.65 compared to a limit of 1.38 gm/Kgm).
The next points recorded were obtained by decreasing the air-fuel ratio to a
CO2 reading of 5.45 resulting in an order of magnitude drop in CO levels
while relatively small increases in NOX were recorded. The general appearance
of widely scattered data points is a result of the continual adjustments made
to CO2 levels to obtain best emission levels or control band information. No
auxiliary air swirler flow was used for the test points of Figure 48. Addition
of 3 percent auxiliary air swirl is shown in Figure 49. Many of the NOX
emission data points were well below the limit indicating a significant gain by
addition of the auxiliary air swirler. However, it was not possible to reduce
NOX emissions below the limits with this configuration at fuel flows below 25
pounds per hour. Visual examination of the combustor after the test indicated
carbon on the auxiliary air swirler. This accumulation was indicative of fuel
impingement caused by excessive recirculation over the cup lip due to the
auxiliary air swirler. By moving the cup away from the immediate discharge
of the swirler, this problem appeared to be resolved as shown by the emission
results plotted in Figures 50 through 52. The "A" dimension (Fig. 42) was
increased from 0. 125 to 0. 192 inch to permit emission levels to be held
within limits down to 9 pounds per hour of fuel flow. Although this change
reduced the beneficial effects of the air swirler, it prevented impingement.
By adjustment of the air-fuel ratio (as shown by the %CO£ readings adjacent
to the data points) emissions of NOX and CO were traded off to obtain the
"best" results shown in the curves on Figures 50, 51 and 52. In order to
maintain NOX below the limits the air-fuel ratio was increased (low CO2
values on curves) at the low fueling rates. A rapid increase in the main
tradeoff parameter of CO is illustrated by the rapid rise in CO at low fueling
rates as the overall air-fuel ratio is made leaner. In the mid-fuel flow ranges,
ample margins exist to allow NOX and CO to be well below the limits. At
fuel flows above 80 pounds per hour it was no longer possible to adjust the
air-fuel ratios and keep both CO and NOX below their limits. At 100 percent
air valve setting the unit was operated at its design air-fuel ratio of 25 to 1 to
do full flow vapor generator calibrations of efficiency and stability. At these
conditions, the CO was less than 1/2 of its limit but NOX was approximately
twice. In general it can be assumed that the emissions at flows above
approximately 25 pounds per hour have little effect upon the steady state
71
-------
simulation of the Federal Driving Cycle. If an assumption of 10 mpg fuel
economy is used, the emissions at a fueling rate of 12 to 15 pounds per hour
dominate the results. As a consequence the greatest effort has been placed
upon lowering the emissions at these low fuel flow rates. Inspection after
emission tests shown in Figures 50, 51 and 52 showed the cup and swirler to
be free of carbon. Figure 53 is a post test inspection after approximately 15
hours of operation with this configuration that results in the lowest emissions.
Upon completion of tests with JP-4, the best configuration (Figs. 50
through 52) was tested with commercial automotive unleaded gasoline. Results
are shown in Figure 54. Emission levels were worse. This was unexpected
since the greater volatility of gasoline should have been an asset in vaporization
and rapid mixing. Inspection after these tests revealed carbon on the swirler
and burn marks on the outer diameter of the cup as far back as the distance it
extended from the swirler. To correct this tendency for gasoline vapors to be
sucked back toward the swirler, the inside ring of the swirler was cut back
0. 5 inch (dimension "B" on Fig. 42) and the cup was extended only 0.03 inch
from the swirler ("A" dimension). In addition, a recirculation fan was in-
stalled in place of the dual heat shield to assist in controlling the flow of
vaporized gasoline and recirculated products of combustion. Results of these
tests showed a considerable improvement (Fig. 55). Emissions were well
within limits at low fuel rates of 6. 5 pounds per hour. However, at the critical
fuel flows of 12 to 15 pounds per hour, the emissions of NOX could not be brought
within limits even though CO was allowed to increase towards its limits. At
these fuel flows the NOX was approximately 10 percent above the limits with
CO approximately 20 percent below its limit. It appears that the recirculation
fan is, as in the tests with JP-4 and the conical cup, producing good emissions
performance at the very low fuel flows as anticipated. However, a hump in
the emissions at fuel flows from 10 to 25 pounds per hour exists that is not
aleviated by the use of the auxiliary air swirler. Cold flow tests with water
through the cup instead of fuel has shown an interaction between auxiliary air
swirler's jet and the fan discharge jet that causes liquid to impinge on the
swirler. Modifications incorporated to eliminate this interaction by changing
the swirl angle of the auxiliary air swirler are discussed below.
Efforts to improve the systems emission in the low fuel flow range
(5 to 15 pph) were attempted. A series of cold flow tests with water used in
place of fuel resulted in the selection of an optimum auxiliary air swirler and
recirculation fan configurations. Visual observation of the water spray pattern
was made by removing the combustor section from the vapor generator coils.
An optimum configuration was defined as one in which no water impingement
was observed on any combustor surface except the side walls. Spray angles
were observed visually with the widest angle being considered the best. Two
configurations had excellent spray patterns. Figures 42 and 56 show the
geometry arrangements between the cup, the auxiliary air swirler and the
recirculation fan. With the "C" dimension at 0.06 and "A" at 0.075 inch a
72
-------
-B-
o.:
AUXILIARY AIR
CUP EDGE
FIGURE 56. AUXILIARY AIR SWIRLER CONFIGURATION
VARIATIONS
180 degree spray pattern of air and water could be obtained with only a slight
impingement on the combustor dome. This condition was with auxiliary air
flow but no fan. When the fan was added the "A" dimension had to be
increased to 0. 137 inch to prevent impingement on the swirler. With that
configuration no impingement on the dome was observed. An ideal spray and
air flow pattern results from high down to flows of 2 or 3 pounds per hour.
As the "A" dimension is decreased, water impingement on the outer lip of
the auxiliary air swirler was observed. As "C" dimension is increased, the
swirl angle is increased (since the axial velocity component goes down) and
a greater tendency to send water back into the auxiliary swirler was observed
Emission tests with both secondary air swirl, recirculation fan and
combination of both were performed, results were not as good as the emis-
sion levels reported above. With "A" at 0. 075, "B" at 0. 5 and "C" at 0. 06
inch without a recirculation fan but with a double heat shield, emissions with-
in limits were obtained at fuel flows above 16. 5 pounds per hour (tests run
up to 43 pph). At lower flows NO was above limits. Inspection of the unit
after runs indicated considerable carbon on the cup and on the outer lip of
the secondary swirler. It was thought that the carbon formed as a result of
high velocity recirculation of partially burnt products across the edge of the
cup. A baffle 0. 1 inch greater than the OD of the cup lip was installed to
eliminate this edge effect (see dimension "D" in Figure 56). Results of emis-
sion tests with the baffle installed were good with respect to carbon formation.
A small amount of carbon formed within the cup but all exterior surfaces re-
mained clean. However, emissions were not as good as without the baffle.
Tests with the "best" swirler configuration and with the recirculation
fan were also disappointing. A large amount of carbon formed and emissions
were all above limits. Some of these poor results were believed to be
73
-------
associated with the discontinuity caused by the auxiliary swirler on the main
air swirlers flow path. The auxiliary air swirler was removed and tests
were performed with and without the recirculation fan. In both cases the
emissions of NO were generally borderline or well above the limits.
None of these tests showed an improvement of the configurations
reported in Figures 50 through 52. In all the tests without the auxiliary air
swirler, there was a carbon buildup on the inside and outside of the cup.
This indicates the main swirler is not flowing full and that a central recircul-
ation across the edge of the cup back into the swirler exit draws fuel vapor
back into this zone. Carbon deposits on the outside of the cup indicate that
fuel vapor burns in this area. Heat from the recirculation flow and the com-
bustion reaction on the outside of the cup compounds the problem by generat-
ing more fuel vapors prior to leaving the cup as a spray. Oxide layer buildup
has indicated cup temperatures in excess of 600°F.
Another series of tests with the main swirl vanes 80 percent blocked
allowed measurements to be performed with high pressure drop and low fuel
flows. As in the above series of tests, this resulted in excessive recirculation
and burning around the fuel flows of 10 to 15 pounds per hour. However, the
safety margin was only abour 10 percent. It was decided to rebuild the com-
bustor back to the configuration that resulted in the best emissions with
gasoline (Fig. 55). Although these results were not entirely satisfactory a
combination of the auxiliary air swirler and recirculation fan resulted in
below requirement from 6. 5 pounds per hour to 10 pounds per hour emis-
sions. From 10 to 20 pounds per hour the emissions of NO were approxi-
mately 10 percent above limits (Fig. 55). This configuration has been
retested with the addition of the double heat shield using the EPA reference
gasoline. Test results show a similar pattern with emissions of NOX 4 or 5
ppm within limits at 8. 5 pounds per hour and reaching the limit at 14 pounds
per hour. From 14 to 30 pounds per hour the emissions of NOX were
approximately 3 or 4 ppm above the limit (23 ppm limit).
Inspection after approximately 70 hours of operation with the EPA
reference gasoline showed an extremely clean combustor and steam generator.
Figure 57 shows the overall view of the combustor after 70 hours. All
surfaces were clean with the exception of a small accumulation of carbon
near the exit of the main swirler and on the outer tube of the auxiliary swirler
(Fig. 58). Local recirculation patterns into the inner core of the main swirler
appear to have been the cause of these carbon formations. Time limitations
have not permitted the necessary redesign to eliminate these slow reacting
pockets. By eliminating these slow mixing, rich pockets (indicative of
near stoichiometric temperatures), a considerable improvement in the
emission characteristics will be achieved.
74
-------
FIGURE 57. COMBUSTOR AFTER 70 HOURS OF OPERATION ON
EPA REFERENCE GASOLINE
FIGURE 58. AUXILIARY AIR SWIRLER AFTER 70 HOURS OF OPERATION
WITH EPA REFERENCE GASOLINE
75
-------
One approach to significantly reduce emissions below the limits is to
use exhaust gas recirculation or water injection. Injection of water can
also give a qualitative analysis of the mechanism of NOX formation. Water
injection was used with the best configuration (Fig. 50, 51, 52) to assist in
analysis.
The rotating cup was well suited to modification for water injection.
A water line was tied into the fuel line just ahead of the pickup on the cup
shaft, as shown in the schematic below. The water was run through a flow-
rator, then through a needle valve, which controlled the water flow rate, just
ahead of the tee.
Control Valve
Flowrator
Cup Pickup
Fuel
City Water
It was found that this simple injection device was capable of producing
drastic reductions in NO emissions, especially at low fuel flows, where
combustor pressure drop is low. At fuel flows on the order of 10 Ibm/hr,
the injection of one pound of water per pound of JP-4 fuel was sufficient to
reduce NOX levels down to less than 1/2 the level required by 1976 standards.
Water injection allowed the average flame temperature at low loads to be
greatly increased over its permissible value for acceptable NOX emissions
without water injection. CO levels were down to about ten percent of their
values without water injections, as would be expected from the higher
average flame temperatures.
The mechanism by which water injection reduces NOX emissions is
fairly well known. Primarily, the water acts to reduce the local flame
temperature by absorbing heat as it vaporizes and as its temperature in-
creases. At high flame temperatures it also absorbs heat as it dissociates.
The dissociation reaction
H2°
O
OH + OH
also tends to reduce NO formation by competing with nitrogen for oxygen
radicals, thus interferring with the Zeldovich mechanism through which NO
is formed, that is,
76
-------
O + N —*- NO + N
N + O2 —*- NO -I- O
From the foregoing, it is apparent that water injection cannot inter-
fere with the formation of NOX unless the water is present in the gas phase
in the regions where NOX is being formed.
Water is considerably less volatile than the JP-4 fuel which was used
for these tests, so it is reasonable to assume that in regions where the water
is primarily in the gas phase, so is the fuel. If diffusion burning from
excessively large droplets was a significant factor in NOX formation in the
present burner, as was thought to be one of the possible problems, water
injection would be unlikely to significantly reduce NOX emissions. In zones
where significant fuel is present in the liquid phase, there would be unlikely
to be much water present in the gas phase.
The water tests therefore seems to confirm that NOX formation in
the test combustor is due to poor mixing in the gas phase, allowing near
stoichiometric pockets of fuel to persist for significant amounts of NO to
form. This problem is probably compounded by the formation of recir-
culation zones in the main swirler close to the cup producing NO through
much the same mechanism as near stoichiometric pockets in the flow out-
side the main swirler.
5.3.5 Temperature Pattern
Evaluation of the combustor outlet temperature pattern was made
during performance analysis of the steam generator. A triple shielded and
asperated thermocouple (Fig. 5) was used to measure the temperature. It
was positioned at one inch increments across the full 20 inch diameter of the
flame tube. Axial position was 8 inches downstream of the rotating cup
(immediately upstream of the first row of the steam generator). Fuel flow
and steam conditions (1000°F and 1000 psi) were held constant throughout
the temperature traverse. The top curve of Figure 59 records the tempera-
ture pattern at 100 percent steam flow (1200 pounds per hour), the maximum
rated condition for the high efficiency monotube unit (see Section 7). EPA
reference gasoline was the fuel, with a measured fuel rate of 104 pounds
per hour. An average temperature of 2350°F was recorded with the
maximum peak temperature of 2625°F measured 4 inches from the centerline
of the combustor. A minimum temperature of 2135°F was recorded near the
wall of the flame tube. A total temperature spread of +275 and -215°F
around the mean results at full flow. This is slightly above the initial goal
of ±250°F but an acceptable deviation when considering the steam generator
heat transfer assumptions. The theoretical flame temperature at these conditions
(neglecting radiation effects) is approximately 2500°F. Radiation heat transfer
77
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DOME
2800
2600
2400
2200
2000
1800
1600
1400
1200
1000
800
600
400
200
\
CUP
-FLAME TUBE
\
MAIN AIR SWIRLER
AUXILLARY AIR
SWIRLER
SIN. (REF)
COMBUSTOR OUTLET
TEMPERATURES
104 PPH
STEAM GENERATOR OUTLET
TEMPERATURES
i i i i i I I I i
-SECONDARY AIR
2350 °F AVERAGE
TEMPERATURE
12 11 10 9
5432101234567
INCHES (FROM CENTERLINE)
10 11 12
FIGURE 59. COMBUSTOR AND STEAM GENERATOR OUTLET
GAS TEMPERATURES
from the highly luminescent flame into the relatively cool (540°F) first row
of tubes was expected to be high. Thus the gas temperature drop of 150°F
below the theoretical level is indicative of the expected high flame emissivity.
A typical temperature distribution at lower flows with leaner air-fuel
ratios is shown by the second curve. It was recorded at the same station
(8 inches from the cup) at the identical positions across the combustor's
diameter. Fuel flow was 31 pounds per hour. The temperature profile at
this flow is considerably different than at the higher flows. A major factor
in explaining this difference is the entirely different air flow arrangement
between the two flows. At 104 pounds per hour fuel flow, approximately 50
percent of the air is being admitted into the combustor through a ring of 90
ports located 4 inches downstream of the fuel injection cup station. The
lower temperatures near the flame tube wall at high flows is probably due to
inadequate penetration and mixing from these jets. Temperatures recorded
at the fuel flow of 31 pounds per hour were with no secondary flow. All air
78
-------
is admitted through the main air swirler. As a result, the characteristic
distribution of temperatures is considerably different. The low temperature
recorded at about 0. 1 inch from the flame tube wall may be a result of a
small amount of secondary air flow due to air valve leakage.
Although the temperature profiles are acceptable with regards to
steam generator performance, they do not look ideal for low emissions.
Of particular concern is the lack of symmetry. As was discussed earlier in
this section, emissions at 100 percent flow is well above the goal. The
asymmetrical shape of the temperature distribution indicates that some
degree of non-uniform fuel-air ratio distribution is taking place, and thereby
contributing to formation of NOX.
The four temperature curves at the bottom of Figure 59 were taken
at a station one inch downstream of the steam generator. Two thermocouples
mounted at 90 degrees to each other recorded temperatures from approxi-
mately 3 inches (a six inch diameter insulated manifold is in center of
steam generator to the wall of the exhaust plenum). Good distribution was
observed except for one temperature spike in the x plane at 104 pph. This
may be indicative of a local gas blowby causing a reduction in the efficiency
of the steam generator (see Section 7).
79
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6
PARALLEL FLOW STEAM GENERATOR
(TEST BED UNIT)
At the preliminary stages of this program it was decided to incorporate a
compact steam generator into the combustor development as soon as possible
to evaluate its effects upon emissions. A bare tube design was selected since
high efficiency for this "test bed" steam generator was not critical and a bare
tube unit could be fabricated quicker. In addition to acting as a test bed for
emissions tests, the unit was designed to provide information on parallel
flow stability and controls systems interface requirements. Parallel flow
was an important factor because the organic systems being considered require
approximately an order of magnitude greater flow than the steam system. At
this high flow rate, a compact vapor generator would normally require a
large number of parallel flow passages to have reasonable pumping losses.
6. 1 STEAM GENERATOR CORE MATRIX
Geoscience Ltd. , on a subcontract to Solar made the heat transfer
analysis of the parallel flow steam generator (see Appendix VII for details of
the analysis).
Output conditions of the steam were selected to correspond to the
temperature and pressure (1000°F and 1000 psia) being used in Steam
Engine Systems Rankine cycle engine being developed for EPA.
One of the important test goals of the program was to obtain low
emissions on a full scale vapor generator/combustor combination. Emission
effects of cold walls, flame quenching, geometry, temperature and velocity
distribution with a compact vapor generator interfaced with combustor were
of prime importance. Figure 60 is a diagram of the selected configuration.
Design details are as follows:
•Steam flow: 1525 pounds per hour at 1000°F and 1000 psi
• Ten flat spiral coils, 5 preheater, 2 vaporizer and 3 superheater
coils
81
-------
VAPORIZER
SUPERHEATER
PREHEATER
Tq=2500°F
Wg = 1.04 Ib
/sec
55
a 1 0 X 10^
Btu/hr
(2 coils,
6 parallel
passages)
6°F 1100 psig
It
M«^H
335°F
r
STEAM
1000 ps
y
q=0.48 X 106
BtU/hr
(3 coils,
6 parallel
passages)
1
OUT 1000°F
ig
^ 550°F
RESTRICTION ' 105° psig
194°
q=0.52 X 106
Btu/hr
(5 coils,
single passage)
694°F EXHAUST
230°F, 1135 psig
1.53 X 103 Ib/hr
WATER INLET
FIGURE 60. TEST BED VAPOR GENERATOR - WATER
WORKING FLUID
• Tube size throughout unit: 0.5 inch OD, 0.035 inch wall, 321
SS all welded construction
• Heat transfer area (gas side) 62 ft
• Tube nest size: 21.5 inches OD, 6.25 inches high
• Gas side pressure drop: 2.8 inches of water
•Water side pressure drop: 135 psi
• Tube weight: 106 pounds
• Water holdup: 30 pounds
• Triangular axial and transverse tube pitch with the pitch equal
to 1.25 tube diameter
Each of the five preheater rows consists of five, flat spiral coils. Connections
at each end of this coil are made by specially designed fittings (bottom of
Fig. 61). Both vaporizer and the three superheaters rows are each con-
structed from six flat spiral coils (Fig. 62) connected in parallel by the
fittings shown in Figure 63.
6.2 CONSTRUCTION
A test bed unit was constructed to minimize lead times. All tubing
is 321 stainless steel (0. 5 OD, 0.035 inch thick wall) with welded construction
throughout. A total of 10 rows of flat spiral coils each 21. 5 inches outside
diameter make up the tube matrix.
82
-------
I
FIGURE 61. VAPOR GENERATOR COIL CONNECTORS
FIGURE 62. VAPORIZATION COILS (6 PER ROW)
8
-------
FIGURE 63. SPECIAL BOX CONNECTION AFTER BURST PRESSURE TEST
Figure 64 schematically identifies the flow arrangement and row numbering
system. Inlet water enters the first row (row number 1) on the exhaust gas
side of the flow system at the center of the coil. It then flows radially out-
ward along the single tube spiral coil that forms row 1 for a total length of
approximately 50 feet. Tubes within each row and the next row are staggered
with 1.25 times outside tube diameter axial and radial pitch. Two of the pre-
heater rows are staggered with respect to each other in Figure 65. Special
welded box fittings (Figs. 61 and 63) connect each row to the next row on both
the outside and inside of each coil to form a compact preheater section con-
sisting of rows 1 through 5 with a total stack height of 3. 125 inches and a
total monotube length of approximately 250 feet. A monotube flow arrange-
ment was acceptable because water velocities can be maintained low in the
preheater and thus a reasonable pressure drop (35 psi) was the maximum
required. A counter flow arrangement has been utilized in the first five
rows of tubing.
A more complex arrangement was necessary in the remaining five
rows. To ensure against potential burnout and hotspots in the superheater,
the counter flow arrangement was modified to place the two vaporizer rows
10 and 9 immediately adjacent to the combustor. Additionally the need for
large surface area with compact tubing arrangement dictated that small dia-
meter tubes be maintained in the vaporizer and superheater sections. As a
result, water side velocities were high and a parallel flow arrangement be-
came necessary (a total pressure drop of 100 psi was used for rows 6 through
10). Each row 6 through 10 consists of six parallel flow passages connected
in a manner to form a single flat spiral. Construction of double rows 6-7 and
9-10 are identical. Each row consists of six flat spiral coils connected at
the center to the next row by means of a special welded box fitting (Fig. 62
views the downstream side of rows 9 and 10). A manifold in the exhaust duct
distributes water from the last row of the preheater to six feed lines connec-
ted to the six coils forming row 10. Water flows radially inward and connects
to the inside of row 9 giving a parallel flow arrangement in the two vaporizer
coils. From the outlets of row 9 six 0.25 inch diameter lines connect to six
adjustable restrictor valves used to balance flow in each of the flow paths
(flow paths are not interconnected except for final outlet of superheater).
84
-------
-1ST
THERMOCOUPLE
-3RD
THERMOCOUPLE
FROM MANIFOLD
(12) THERMOCOUPLES SPACED EVERY 8.00 INCHES ON
ALTERNATE FLOW PATH TUBES TYPICAL ROW 6 AND 9
(6) THERMOCOUPLES ON EACH OF THE SIX
OUTLET TUBES TWO INCHES FROM MANIFOLD
Tw SUPERHEATER
OUTLET
(SIX TUBES)
V(
INLET WALL
TEMPERATURE
TO FLOW PATH - B -
00
Ul
nNr
PRESSURE OUT TH,
PRESSURE TRANSDUCER-^L^l
PRESSURE
TRANSDUCER —
(6) PRESSURE
TRANSDUCER
0-1500 psi
TG OUT 9
*10
\
#8
(1) THERMOCOUPLE
IMMERSED
(12) WALL THERMOCOUPLES
EQUALLY SPACED ON ALTERNATE FLOW
PATHS PLUS (10) THERMOCOUPLES
EQUALLY SPACED ON FLOW PATH- B-
FROM MANIFOLD
TO VALVES
iL&r^ *7 f\
AjSr— *6
/a — «
' ( #4
( *3
( #2
s, ( *1
(2) THERMOCOUPLES
IMMERSED
(COMPRESSION FITTING)
-^
\
\
\
\
\ TG OUT 6
(6) EXHAUST AIR THER
EVERY 60 DEGREES TC
IMMERSION PROBE
(COMPRESSION FITTING
TWO PLACES)
"FROM VALVES
"(12) WALL THERMOCOUPLES
EQUALLY SPACED ON ALTERNATE
FLOW PATHS
H.. OUTLET ROW 3
OUT ROW 1
-(21 THERMOCOUPLES IMMERSED
INLET COMPRESSION FITTING)
1/8 SS TUBES (INLET
PRESSURE TO EACH
SUPERHEATER COIL)
PRESSURE
TRANSDUCER
(2) THERMOCOUPLES
IMMERSED
(COMPRESSION FITTING
FIGURE 64. VAPOR GENERATOR FLOW AND INSTRUMENTATION
-------
XX.*——Nv
W
O
FIGURE 65.
ASSEMBLY OF TWO PREHEATER COILS WITH SPECIAL
CONNECTOR WELDED AT INSIDE OF COILS
From the restrictor valves a 0.25 inch tube connects the flow into the first
row of the superheater (row 6). Flow in this row is again inward and back-
outward through row 7 where special welded connectors direct the flow into
row 8 where a central collection manifold finally connects each of the six
flow paths to a single 0. 75 inch diameter outlet tube (Figs. 66 and 61).
A proof pressure of 3000 psi was applied for 20 minutes as a leak
check prior to installation in the test cell. Weight of the assembled vapor
generator is 104 pounds including instrumentation and restrictor valves.
Instrumentation was installed to monitor steam generator performance;
the most important and useful measurements are listed below:
• Input and Output Conditions
Two immersed thermocouples, closed Inconel sheath, 0.06 inch,
type K thermocouples were located at the inlet and outlet of the
vapor generator. Two were used for redundency for these
critical parameters. A pressure tap is also located at the inlet
and outlet.
•Outlet Temperature in Each of the Six Parallel Superheater Flow
Paths
Six surface thermocouples are installed two inches from the
outlet of each tube in the superheater row 8 (Fig. 66). These
.so
-------
FIGURE 66.
SUPERHEATER OUTLET ROW SHOWING THE SIX THERMO-
COUPLES
were used to monitor outlet temperature in each flow path
for stability.
• Superheater Thermocouples
Four immersion type thermocouples are located in the super-
heater. Two are at the 33 percent and the other two at the 66
percent point along the flow path.
• Water Flow Rate
A high response turbine flow meter with direct digital readout
in pph installed in the inlet line of the vapor generator was used
to measure water flow rates.
6.3 PERFORMANCE TESTS
Steam generator performance, emissions and control system tests
were all performed with the same basic test loop. Operation of the test
loop is shown by the flow schematic (Fig. 67). Untreated city water in
the test cell supply line is passed through four deionization bottles.
87
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CONTROL
CONSOLE
AMP H
IGN I -K
EXHAUST
PLENUM
IPREHEATER
_, [THROTTLE!
EXHAUST
SYSTEM
R \/ I—I . (
ly^X! I |~DE-IONIZATIol\r
T ! _SYSTEM_
FIGURE 67. TEST CELL FLOW SCHEMATIC
After deionization the water goes to a preheater that discharges it at a
regulated temperature between 180 to 190°F. An insulated suction line is
connected to the inlet of a Union DX-10 triplex pump. A variable diameter
pulley belt drive from an electric motor allows remote control of the pumps
speed. Prior to the installation of the automatic electronic control system
pumps, speed was used to adjust the water flow control steam outlet tempera-
ture. Accumulators on the inlet and outlet reduce pressure and flow pulsations
from the piston pump to acceptable levels for accurate control. A relief
valve protects the pressure side of the pump circuit from overpressure. Some
of the outlet flow from the pump is bypassed by a valve (B.P. valve of Fig. 67)
to the inlet to the pump. A bypass was necessary at this point to allow a
match between the pump speed characteristics and the speed range of the
variable speed drive system. Water inlet flow is measured by turbine flow
meters (FM) immediately ahead of the inlet to the steam generator. Two
turbine meters are necessary to cover the flow range of interest. Both are
checked periodically by means of a weight of flow versus time to ensure their
accuracy. Under steady state conditions the inlet water is identical to the
steam rate. Difficulty with exact measurements of steam flow directly have
resulted in the use of water inlet flow measurement for the steam rate. The
only difficulty with this approach has been the extremely long times at low
flows required to obtain true steady state values. At some control conditions,
a true steady state condition did not ever appear to be established. Generally,
when the rate of change of temperature for a specific flow was only 1 or 2°F
per minute the condition was considered steady state.
88
-------
The blocks labeled COMB/V. G. schematically represents the test
unit. Inlet water temperature are measured in the 0. 5 inch diameter feed
line downstream of the reverse flow check valve. Exhaust gases, after
passing through the unit, discharge into an exhaust plenum and then into a
15 inch duct for venting into the test cell's exhaust stack. Steam rate is
controlled by the setting of a throttle valve. Between the steam generator and
throttle valve a safety relief valve (R.V.) and thermocouples (TC) are installed.
From the throttle valve the superheated steam is vented into the test cells
exhaust system.
All functions of the system were regulated from the test cells control
panel (Fig. 68). Fuel flow, water flow, and air flow can each be independ-
ently regulated remotely from the control panel. Steam flows were established
at the control panel by manually adjusting the setting of the throttle valve.
Ignitor and fuel solenoid controls are also manually controlled from the face
of the panel. After installation of the steam generators control system, the
position of the air valve could be controlled automatically or manually from
the control room. Water rate was also capable of being manually or automat-
ically regulated by either automatic or manual remote positioning of the input
to the differential pressure regulator (see Section 8). Fan speed was adjust-
able by changing the voltage to the fan drive motor . Important instrumentation
FIGURE 68. TEST CELL STEAM GENERATOR CONTROL PANEL
89
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and control functions in the control room included:
• Speed of water feed pump
• Fuel pressure and flow
• Superheater outlet temperature
• Six parallel flow paths outlet temperatures
•Inlet and outlet pressure of steam generator
• Water flow
• Fan speed
• Fan discharge pressure
•Primary and secondary air pressures
• Combustor pressure
• Temperature of inlet air to fan
• Combustor outlet temperature
• Air valve position
• Voltage and current to fan motor
• Pressure regulator position
6.3.1 Mechanical Integrity
The six parallel flow passage steam generator has been operated more
than 500 hours in emissions, controls and performance development tests.
In all these tests the unit was operated at near rated temperature and pressure.
Two leaks developed after approximately 70 hours of preliminary tests. Both
leaks were in the vaporizer section at weld joints made to the tubes accidently
by a weld arc-over from a tube spacer. Figure 65 shows the six tube spacers
installed in each row. Each end of the tube spacer is welded together to hold
the spacer closely about the tubes. In both leakage failures a small crack
formed about the weld nugget between the tube spacer and the tube wall. It
had been anticipated that the thermal strain would cause this type of problem
and the design required that the tubes be free to move slightly within each
spacer. Errors in welding proved the importance of not restricting the tubes
90
-------
with a limited number of highly stressed weld joints. After repair, no fur-
ther leaks developed during more than 400 hours of testing. The only physical
damage noted during the final inspection of the unit was distortion and failure
of several of the spacers and flow blockage plate at the center (Fig. 69 and 70).
These failures were due to insufficient depth to the first spacer row and use
of a weld joint between the center plate and the six spacers. A slip joint and
an increase in depth of 0.25 inch on the second high efficiency unit has correc-
ted this problem. The deposits shown on the center sections of the coils were
rust color and apparently were residual elements in the fuel and air. Table
VI shows the results of a spectrographic analysis of the deposits on the first
row of tubes. Carbon deposits are seen on the outside of the coils and were
thought to be the result of ignition failures.
6.3.2 Flow Stability
It has been demonstrated that it is possible to operate a six parallel
flow steam generator at water flows up to rated conditions without serious
water flow maldistributions in the superheater tubes, and without the con-
sequent overheating of the tubes. Successful operation has been achieved
without any flow restrictors in the water flow passages. Balanced flow
seems to depend on maintaining near rated outlet pressure at all flows, and
FIGURE 69. STEAM GENERATOR AFTER 500 HOURS OF EMISSION
AND CONTROL SYSTEM TESTS
9 1
-------
FIGURE 70. VAPORIZER SPACERS AFTER 500 HOURS OF OPERATION
TABLE VI
RESIDUE REMOVED FROM FIRST ROW OF STEAM GENERATOR
(Chemical Analysis by X-ray Fluorescent Spectrography)
Elements Detected (Approximate %)
Al
Ca
Cd
Cr
Ca
Fe
K
>0.20
>1.00
>0.08
>0.25
>0.25
>0.75
>0.03
Ni >0.15
P >0.50
Pb >1.00
S >6.00
Si >11.00
Zn >0.30
92
-------
on adjusting water flow so that the preheater outlet liquid is within a few
degrees of the saturation temperature. It appears to be impossible to achieve
stable operation of the unit unless these two conditions are simultaneously
met.
The effect of preheater outlet temperature is illustrated quite clearly
in the traces of Figure 71. At the left hand side of the trace, the preheater
outlet temperature is only 410°F, about 75°F below the saturation temperature
at the outlet pressure of 610 psi. The six traces shown represent steam
temperature at superheater outlet for each of the six parallel flow superheater
tubes. It can be seen that two of the tubes have a much lower temperature
than the other four, and that the temperature in these two tubes undergoes
a sustained oscillation in temperature. It is believed that there is considerable
liquid phase flow in the tubes shown by these temperature oscillations, even
near the superheater outlet. This is also shown by the mixed mean super-
heater outlet temperature, which was quite low, in spite of the high wall
temperature, which fluctuated between 480 and 620°F. It -was found that with
oscillations such as those shown on the left hand side of the trace, it was
impossible to achieve anything like the rated superheater outlet temperature
of 1000°F without producing excessive wall temperatures in the superheater
tubes.
At the right hand side of the trace, one sees the effect of an increase
in the preheater outlet temperature. Preheater temperature is now 455°F,
only about 45°F below saturation temperature. As preheater outlet tempera-
ture is increased, the wall temperature oscillations disappear, and the wall
temperatures of all six superheater tubes become nearly equal. Superheater
temperature was steady, without the 40°F oscillation observed at the left
hand side of the trace.
An explanation of the importance of the preheater outlet temperature
and other parallel flow stability criteria has been discussed in the literature
A brief review helps explain the performance of the unit. There are several
types of two phase flow instability which must be considered and, in general
avoided, if the performance of the steam generator is to be considered
satisfactory. They are to be avoided for two reasons. First, they can cause
disastrous overheating of the tube walls, even though steady state calculations
indicate that wall temperatures should be satisfactory. Second, they can
produce oscillations in steam generator outlet temperature and pressure even
though there has been no change in steam demand, air flow, or fuel flow.
This tends to complicate the controls problem. Flow instabilities can be
classified according to whether they are periodic or aperiodic.
A periodic instability results when the flow conditions are such that
an increase in flow produces a reduction in pressure drop. For this type of
instability, a high loss factor for the flow tends to be stabilizing, as does a
93
-------
P in =640 psi
P oul=610 psi
Tmout410°F
47 PPHTUEL
580to620°F
39% AIR FLOW
TSHout=902°F
P in = 670 psi
TF;2275 F
TSH out=908°F
39% AIR FLOW
I I I I I I I I I I I I I I I I I I I I I I
J—I i I—1_J—I—L_l—I—1—I—I—I—I—I—I—I—I—I—I—I—I—I—I—I—I—1—I—I—I—I—I—L,
440 PPH WATER
40% AIR FLOW
Pout =640 psi
Tp'H="430lo440 F
P out= 630 psi
P in= 660 psi
I I I I I I I I I I I I I I I I I I I I I I I I I I I I I I I I I I I I I ' I I
400 PPH WATER
47 PPH FUEL.
TPHout= 455° F
h
HI-
Time ~*| r~ 20 Seconds
FIGURE 71. SUPERHEATER OUTLET TUBE WALL TEMPERATURE (SIX FLOW PATHS)
-------
small difference in specific volume between the liquid and the vapor phases.
Inlet subcooling has a definite destabilizing effect. A good discussion of these
tendencies is given in chapter 7 of Tong (Ref. 2). It should be mentioned
that any inflection point in the curve of the pressure drop versus flow is likely
to cause difficulties in parallel flow units. If the stability is marginal, then
large changes in flow rate can result from small differences between the
tubes in heating rate or loss factor. This can cause serious flow maldistribu-
tions with its accompanying overheating of some of the tubes, as they become
starved for water.
Periodic flow instabilities are usually the result of coupling between
thermal and hydrodynamic forces. They tend to be more troublesome in
parallel flow units than they are in monotube units. In a parallel flow unit,
oscillations can occur in one tube without significantly affecting the others.
When this is the case, other parts of the loop, such as accumulators and
throttle, have less chance to damp the oscillations. Periodic instabilities
can occur in systems which are stable against aperiodic instabilities. They
can have highly undesirable interactions with the boiler automatic controls,
if this possibility is not carefully eliminated in the design. In general, those
factors which increase the aperiodic stability also increase the periodic
stability. Some other considerations are: (1) a high heat input aggravates
instability; and (2) an orifice at the inlet strongly increases the stability,
while increased resistance at the outlet is strongly destabilizing. Periodic
flow instabilities can sometimes be caused by a shifting back and forth
between different two-phase flow regimes, for example, slug flow and
annular flow.
While the analysis of flow instabilities is quite complicated and, at
best, only approximate, it is nevertheless fairly clear what the designer must
do to eliminate the danger of two phase flow instabilities. The approach to
the solution of this problem in the high efficiency unit discussed in the next
section is outlined below:
1. The preheater is designed so as to give a very slight amount of
vaporization before the flow enters the vaporizer. The inlet
subcooling for the vaporizer is therefore zero. Under such
circumstances, it is theoretically impossible for aperiodic .
instabilities to exist, because the inlet subcooling is negative.
Collier and Pulling (Ref. 3) s-ay that their experiments indicate
that with two phase flow at the inlet, periodic instabilities are -
unlikely to occur.
2. There is a restriction at the inlet to the vaporizer section. This
has a further stabilizing effect.
95
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3. The number of parallel passages in the vaporizer and superheater
sections will be kept as low as possible, so that other parts of the
loop will have the maximum possible damping effect.
4. The flow will be mixed in manifolds at both the inlet and exit of
the vaporizer tubes. This will minimize the effect that flow
instabilities in any one section have on the other sections. It
also ensures a very stable flow in the superheater section, where
tube wall temperatures are expected to be highest.
5. Mass flow rates are chosen so that the transition from one type of
two phase flow to another is not likely to take place over a large
length of tube all at the same time.
6. The final design will be analyzed by the method presented by
Quandt (Ref. 4) in order to determine the degree of periodic
flow stability, and the frequencies involved. Such information
will be useful in the design of the controls.
Experience with the parallel flow steam generator tends to confirm the
results which would be expected from a study of the literature. Inlet subcool-
ing has been seen to have a definite destabilizing effect. The flow is less
stable at high heat inputs than it is at low heat inputs. Increasing the outlet
pressure, hence reducing the volume change due to vaporization, has tended
to stabilize the flow. Since test results confirm results predicted in the
literature, this literature is believed to be a reliable basis for design.
Steady state performance data have been obtained at fuel valve settings
ranging from 10 to 98 percent. The results are shown in Figure 72. At full
load, the performance is very close to theoretical predictions based on
2500°F flame temperature and zero heat loss to the surroundings. The rise
in efficiency with decreasing load is much less than theory predicts, however.
There is considerable scatter in the data at low load.
It is believed that the low efficiency at low load is primarily due to
the fact that it is necessary to lean the combustor at low fuel flows to reduce
NOX causing a drop in flame temperature as the load is decreased. While at
full load the flame temperature based on CO2 measurements is in the range
2500-2700°F, at 20 percent air valve position it drops down to 1800-2000°F.
The lower temperature would cause a noticeable loss in boiler efficiency.
The scatter in the low load data points is thought to be primarily due
to the fact that backlash in the fuel-air valve linkage used with the six parallel
flow passage unit occasionally cause rather large differences in flame tem-
perature corresponding to the same setting. A contributing factor to the
scatter may be the difficulty of obtaining steady state operation at low loads.
96
-------
95 -
90
85
o
Q
u
U
z
u
b.
b.
30
75
70
65
LEGEND
O DATA POINTS
BEST CURVE FIT TO DATA
THEORETICAL PREDICTION
BASED ON 2500* F FLAME
TEMPERATURE + ZERO
HEAT LOSS TO SURROUNDINGS
8
10
20
30 40 50 60 70
FUEL VALVE POSITION - % OF FULL LOAD
80
90
100
FIGURE 72. TEST BED VAPORIZER STEADY STATE PERFORMANCE
A change in superheater temperature of only 2°F per minute can apparently
mean that the unit is still well away from steady state at low loads. A
continual high frequency fluxuation in the water inlet flow recorded by the
turbine flowmeter used to measure water rate is also thought to be a possible
factor contributing to the scatter of the data (see Section 8).
The unit seems to have only limited hydrodynamic stability at high
loads. Even at superheater temperatures in excess of 1000°F, a definite
sustained oscillation in superheater temperature can frequently be observed.
at loads in excess of 80 percent of the design load. The poor stability also
seems to be reflected in worsening flow distribution in the superheater tubes
as the load is increased. This is shown in the charts of Figure 73. At the
high load point, the difference between hottest and coldest tube is slightly
97
-------
1300
b-
°i 1200
w
OS
EH
K
TUBE WALL TEMPE
M M
O H*
O O O
v\ ° ° <=>
Tube
No.
DATA POINT 23%
I
1
TF •» 1880° F
TpH = 512° F
TSH = 1004°F
I
I
^
$$
^
^
$$
$$
^
._.
l%^m%^^
1
23456
DATA POINT 81%
Tp • 2675° F
TPH " 475°F
TSH - 988" F
!
NX
ss
^
C?^
1
^
s§
^
^
^
^
^
XX
^
^
XV
LEGEND
TF - AVERAGE FLAME
TEMPERATURE BASED
ON CO2 MEASUREMENT
TpH = PREHEATER OUT
TEMPERATURE
Tcu - MIXED MEAN SUPER-
SH
HEATER OUT TEMPERATURE
' (INDICATED BY DOTTED
LINE ON BAR CHART)
1 23456
FIGURE 73. TEST BED VAPORIZER. TUBE WALL TEMPERATURE
AT SUPERHEATER OUTLET
over twice what it is at the low load point. The difference between hottest
tube wall temperature and mixed mean superheater temperature is also
slightly over twice as great at high loads. Although these stability
characteristics were present, hundreds of hours of emission tests and
closed loop control system tests do indicate that a practical parallel flow
system can be achieved. The main problem is controls. Steady state and
most transient controls do not present much of a problem. It appears from
the results that only an initial startup stability is of major concern. Although
not worked out in demonstration tests, it does appear as though a controlled
schedule in the rise of pressure and temperature would allow satisfactory
automatic startup of a parallel flow unit. Additional control measurements
would probably be required to give adequate reliability. By measuring each
of the parallel flow outlet temperatures and restarting if chugging occurred
during transient operation, a high degree of reliability against burnout
failures would be ensured. Although it turned out to be unnecessary with the
automatic control system, the six parallel flow passages temperatures were
continuously monitored to prevent damage to the unit during all test cell
operations.
98
-------
HIGH EFFICIENCY SINGLE FLOW PATH
STEAM GENERATOR
A second steam generator incorporating fins for high efficiency and
light weight was designed and tested. It also incorporated essentially a
monotube flow arrangement (only two parallel flow passages in the dryer)
for improved stability. To achieve high efficiency, 'low weight, low water
holdup and low gas side pressure drop, small diameter finned tubing was
used in the design. The dominant thermal resistance is on the gas side.
Fins allow this resistance to be reduced by adding heat transfer area which
is not highly stressed, as are the tube walls. Since the fins are unstressed,
they can be made of very light gage metal thus saving •weight. In the pre-
heater, the fins account for only 61.9 percent of the total weight, while
providing 85.9 percent of the total heat transfer area. Because of the small
hydraulic diameters attainable with fins, heat transfer coefficients are much
higher, providing a further saving in weight. Heat transfer coefficients in
the finned tube parts of the exchanger are about half again as high as the
bare tube parts of the exchanger.
Gas side pressure drop is low because fins present a surface with a
high ratio of drag to form factor. Friction drag allows a much higher heat
transfer per unit pressure drop than does form drag, since the viscous
dissipation in the turbulent wake does not contribute to the heat transfer
process.
Finned tubes help reduce water hold up, because they allow a large
gas side heat transfer surface for a given tube length with only a small
volume of tube.
The effect of changing tube diameter on water hold up and matrix
weight is derived in Appendix II. Both hold up and matrix weight increase
slightly more rapidly than in direct proportion to the tube diameter. For
finned tubes, fin weight increases in proportion to the square of tube diameter.
The basic constraints used for the unit were to have an efficiency of
at least 85 percent LHV at full load of 1200 pounds per hour of steam and
that it fit in the same combustor used for the first steam generator (21. 5
inch steam generator OD).
99
-------
Table VII shows the flow arrangement and fluid conditions of the steam
generator. In general, the overall flow arrangement that proved out well on
the test bed unit was incorporated in this improved steam generator. Place-
ment of the vaporizer tubes upstream of the dryer and superheater gives
protection against burnout failures and high tube wall temperatures that
would result in a pure counterflow unit. Table VIII is a summary of the units
performance and construction features.
TABLE VII
FLUID CONDITIONS
0
Vaporizer
Gas In
2670 Ibm/hr
Superheater
©
Dryer
©
75*
0
Preheater
0
0
Water In
1200 Ibm/hr
Station
1
2
3
4
5
6
7
8
9
10
11
Temperature
(°F)
2500
2017
1985
1714
1090
306
160
561 (3. 9% quality)
559 (57. 3% quality)
696
999
Pressure
Atmospheric
Atmospheric
1194 psia
1148 psia
1126 psia
1064 psia
1000 psia
A comprehensive summary of the monotube steam generator design
including sizing, heat balances, part load-performance, dryer heat transfer,
burnout, air side pressure drop, and parallel flow-passage stability is con-
tained in Appendix III.
100
-------
TABLE VIII
SUMMARY OF PERFORMANCE PARAMETERS
Maximum Steam Rate 1200 Ibm/hr
Maximum Heat Transfer Rate 1.651 x 106 BTU/hr
Boiler Efficiency
/ API 56 Gasoline, S.G. = 0.755 \
\HHV = 20, 160 BTU/lbm; LHV = 18, 840 BTU/lbm'
100% load 84.4% (HHV); 90. 1% (LHV)
5% load 90.0% (HHV); 96.3% (LHV)
Gas Exit Temperature
100% load 306 F
5% load 172 F
Gas Inlet Temperature 2500 F
Gas Side Pressure Drop 1.318 in. H2O at 100% load
Feedwater Inlet Pressure 1194 psia
Steam Outlet Pressure 1000 psia
Maximum Fin Temperature 999 F
Maximum Tube Wall
Temperature 1031 F
Total Matrix Metal Weight 66.0 Ib
Matrix OD 21. 5 in.
Total Matrix Depth 6. 65 in.
Water Hold up 6.76 Ibm
Construction
Preheater: Three rows of single flow passage flat spiral
coils in cross counter flow, 0.375 inch outside dia-
meter, 0.028 inch thick 321 stainless steel tubes with
15 copper fins per inch. Fin outside diameter is 0.625
inch with a thickness of 0.012 inch. Transverse pitch
0.717 inch, axial pitch 0.784 inch.
Dryer: One row of two parallel flow passage flat spiral coils
0.375 inch outside diameter, 0.028 inch thick Hastelloy
X tubes with 15, 304 stainless steel fins per inch. Fin
outside diameter 0.625 inch with a thickness of 0.012
inch. Transverse pitch 0.717, axial pitch 0.784 inch.
Superheater and Vaporizer: Both are two rows of bare 0. 627
inch, 0.042 inch wall thickness 3Z1 stainless steel tubes
in single flow passage arrangement. The transverse and
axial pitch is 0.9375 inch. Superheater is in cross counter
flow and the vaporizer is in cross parallel flow.
101
-------
7.1 CONSTRUCTION
Flow arrangements, matrix spacing and materials are defined in
Tables VII and VIII. All welded construction is used with the exception of
the fins. Fins were brazed to tubes after coiling. Figure 74 shows a copper
finned preheater coil prior to brazing, Three coils make up the preheater.
Flow is cross counter flow with each of the 0.375 tubes being joined in series
by special welded connectors. The only parallel passage row in the unit is
the dryer located immediately above the preheater. Two coils with twice
the transverse pitch (Fig. 75) of the preheater are intermeshed to form a
single dryer row with the same transverse pitch as the preheater but with
two parallel flow passages. Tube material in the dryer was Hastelloy X.
Under normal conditions stainless steel would be adequate but the unit was
designed to be failure proof with regards to parallel flow instabilities.
Stress temperature characteristics are sufficiently great to allow successful
operation with zero flow through one of the two dryer coils. Analysis of
stability characteristics indicated a high degree of stability in this area.
Tests confirmed the analysis but did show a 60°F difference in the outlet
temperature between the separate flow paths.
Outlet water from the preheater bypasses the dryer and superheater
and enters the first row of the vaporizer adjacent to the combustor. A
vertical tube in the center core (Fig. 76) brings the preheater outlet water
up to a special welded connector on the inside end of the top coil. At the
outside of the spiral coil a special welded U-connection (Fig. 77) into the
second row of the vaporizer section. At the inside core a jumper line brings
the outlet from the vaporizer section down past the two superheater rows to
the dryer. Flow is divided at this point into the two dryer coils where it
flows from the center out to two jumper tubes. These tubes bring the steam
down around the outside (see tube at far right of Fig. 76) of the preheater
section and across the bottom of the unit and back up to the first row super-
heater through the center core. In the center core the flow is combined in
a manifold and flows through the superheater coils in a single flow path.
Steam from the first superheater row is coupled by a "U" connection (Fig. 77)
to the final superheater row for discharge through an outlet tube that directs
superheated steam down the center core and out across the bottom of the
unit.
Immersion thermocouples were installed at the inlet and outlet of
each section of the unit. Figure 77 shows the installation of a thermocouple
that records temperature on the outlet of the first row vaporizer and one
that measures temperature at the outlet of the first superheater row. Two
wall temperature probes were installed within 3 to 4 inches of the outlet of
the superheater's last row. Both of these units agreed well with the
immersed superheater thermocouple, but always indicated slightly higher
as was expected.
102
-------
FIGURE 74. PREHEATER COILS WITH COPPER FINS
FIGURE 75. ONE OF THE TWO DRYER COILS
103
-------
FIGURE 76. ASSEMBLED STEAM GENERATOR
FIGURE 77. STEAM GENERATOR U CONNECTIONS
104
-------
7.2 STEADY STATE PERFORMANCE
The high efficiency finned tube steam generator has been calibrated
for steady state efficiency at fuel flows from 15 to 105 pounds per hour fuel
flow, corresponding to steam rates from 240 to 1200 pounds per hour at
1000 psia and 1000°F. The unit was stable with little or no difficulty in
holding a particular steady state point. As a consequence, little scatter in
the data was observed.
The design goals for this unit are listed below.
• Maximum Steam Rate: 1200 pph @ 1000°F and 1000 psia
• Efficiency (LHV) at full load 85%
• Water Side Pressure drop at full load - 250 psi maximum
The steady state efficiency is shown on the curve of Figure 78. As
can be seen, the design goal is slightly exceeded at full load, and well
exceeded at fuel flows representative of those encountered during the Federal
Driving Cycle.
At the maximum steam flow measured, 1245 pph, the water side
pressure drop was 320 psi. This corresponds to a water side pressure drop
of 297 psi at a flow of 1200 pph, which is slightly higher than the design goal
of 250 psi. It is thought that this increase is due to the fact that the mani-
folding system in the present unit is slightly more restrictive of the flow than
was originally estimated.
Although the unit meets the design goals regarding efficiency, it does
not measure up to the efficiency predicted for it by analysis, and gas side
pressure drop was also significantly lower than predicted. These three
factors; low measured efficiency, high exhaust temperature, and low gas side
pressure drop combine to indicate that the poorer than expected performance
is due to significant blowby being short circuited around the tube coils, so
that not all the combustion gases pass through the coils. It seems quite
possible that if the blowby can be stopped, full load efficiency can be brought
close to the 90 percent level.
As was predicted by analysis in Appendix III, flow in the parallel
passages of the dryer was quite stable. No tendency toward chugging was
observed. Flow through the boiler was much more nearly constant at
steady state than it was for the original six parallel path unit. The maximum
temperature spread between the dryer outlets was 57°F, as compared with
300°F for the six parallel path unit, as reported in Section 6.
105
-------
1.02
1.00
0.98
0.96
0.94
u 0.92
UJ
y 0.90
L_
m 0.88
>
5 0.86
0.84
0.82
0.80
EPA REFERENCE FUEL- LHV 18,500 BTU/LB (UNLEADED GASOLINE)
® DATA POINTS
D DESIGN GOAL
A ANALYTICALLY ESTIMATED POINTS
®
I I I
1 I
DESIGN GOAL
I I I I I I I i I I
0 5 10 15 20 25 30 35 40 45 50 55 60 65 70 75 80 85 90 95 100
FUEL FLOW »~PPH
FIGURE 78. HIGH EFFICIENCY FINNED TUBE STEAM GENERATOR
STEADY STATE EFFICIENCY MEASUREMENTS
Inspection of the steam generator after approximately 70 hours of
operation revealed no damage or accumulation of deposits. EPA reference
unleaded gasoline was used as the fuel. Both steady state efficiency measure-
ments and severe transient cycling tests were performed during the operating
period. No leakage or distortion occurred during operation of this unit,
although temperatures and pressures in excess of 1200°F and 1200 psi
occurred during some transients. Figure 79 illustrates the condition of the
top row just below the combustor. Vaporization takes place in this and the
next row at an average temperature of 550°F. As a consequence, the tubes
still retain a bright appearance with little traces of oxidation even though gas
temperatures in excess of 2600°F (see Fig. 79) entered the tube matrix at
this row. Two rows below the last superheater coil had a relatively dark
but thin oxide coating typical of operation of 321 stainless steel at about
1100°F wall temperature. A 1/8 inch thick Hastelloy X center core baffle
plate indicated a relatively heavy oxide layer but little distortion. No direct
connection between this plate and the three tube spacers was made in this
installation. Additionally, the depth of the tube spacers was greater than
the unit described in Section 6. As a result of these two improvements no
breakage or distortion occurred in the first row tube spacers on this unit.
Figure 80 illustrates the condition of the last row preheater from the
exhaust duct. Both the stainless steel tubes and copper fins were in a brigh*:
metal condition with little indication of either oxidization or deposits.
106
-------
FIGURE 79. TOP COIL (VAPORIZER) AFTER 70 HOURS OF OPERATION
FIGURE 80. BOTTOM COIL (FIRST ROW PREHEATER) AFTER
70 HOURS OF OPERATION
107
-------
8
CONTROL SYSTEM
A unique automatic control system was evaluated as an integral part
of the combustor steam generator test program. Requirements for an auto-
motive steam generator control can be summarized as:
• Maintain a fixed schedule of air-fuel ratios across the full
operating range of the combustor to ensure low emissions,
correct outlet temperature and prevent flame out.
• Maintain outlet steam pressure at constant value of 1000 psia
(tolerance goal of ±100 psi)
• Maintain steam outlet steam temperature at a constant value of
1000°F (tolerance goal of ±100°F).
An automatic control was synthesized to maintain these functions under the
virtually continuously varying steam demands of an automotive city driving
cycle. Inherent in the requirements are large turndown ratios with rapid and
frequent power level demands.
System synthesis was based upon first providing a positive synchro-
nization between the fuel and air flow control subsystems. Secondly, it was
determined that the transient requirements for steam flow and the slow
thermal response of the unit required an open loop schedule of fuel and water
to be instantaneously responsive to steam flow. By providing this anticipatory
feature in the control loop, a relatively limited authority was needed in the
closed loop temperature and pressure trim control circuits. Response
characteristics observed during manual operation of the test bed steam
generator indicated that the best control response occurred when firing
rate (fuel-air flow) was used as the trim control input to correct pressure
errors. Water flow rate changes were used to correct temperature errors.
In both cases outlet steam pressure and temperature were the only parameters
in addition to steam rate required as inputs to the control system. Fuel flow,
water flow and air flow were the only outputs of the control system. Com-
ponents to allow open loop synchronized scheduling of air, fuel and water
were specially designed to be integrated into the system. Electronic sensing,
positioning and compensation was used in the integrated system since its
flexibility lends itself to a highly developmental controls program.
109
-------
8.1 SYSTEM DESCRIPTION
Operation of the basic open loop schedule control (anticipatory type
steam generator control) is simple and straight-forward. Output steam flow
rate is controlled by a throttle valve that has a pressure ratio above critical
flow at all times. It is a simple variable orifice contoured to provide a
linear change in flow area with input position. Thus, if the pressure in the
steam generator is relatively constant, the steam flow is directly proportional
to the "manual throttle position". An electrical position transducer measures
this position and thereby provides an electrical signal directly proportional
to flow. In an actual automobile installation this signal could be synthesized
from the engine speed and cutoff valve position or a direct measurement of
flow. A closed loop position control subsystem (Fig. 81) automatically
moves the triple valve actuator to a position that will give the desired steady
state temperature and pressure while balancing the heat input to the steam
rate. The gain of the control system is approximately 600,000 BTU/sec.
(It takes approximately two seconds to move from 0 to full flow.)
Closed loop trim control on output steam pressure is provided by an
electronic system. Outlet pressure is sensed by a strain gage transducer.
Comparison between actual pressure and the desired setpoint of 1000 psi is
made in an electronic controller having a simple proportional output. A
wide adjustment of gain is provided in this controller to assist in establishing
the best gains to be used with the system. From the controller the signal is
operated on by a pressure trim gain compensator that scales the control
signal proportionally to steam flow (position of the steam throttle valve).
For example, if operating at 5 percent steam flow the pressure circuit would
have an actual authority of ±2 percent. At 50 percent steam flow the corres-
ponding authority of the pressure trim circuit would be ±20 percent. The
compensated pressure trim signal is then summed together with the main
open loop position control signal (proportional to steam flow) to trim the
triple valve actuator position from its basic position assumed as a function
of steam flow.
Temperature trim is provided by a completely independent electronic
system. Outlet steam temperature is sensed by a thermocouple and a pro-
portional plus integral and derivative controller which trims the temperature
by adjustments to the feedwater rate. For a given set of conditions, reducing
the feedwater rate decreases the water level. When the water level (dryout
zone) drops the most significant effect is to increase the length of the super-
heater, thereby increasing outlet steam temperature. Since the feedwater
metering valve is mechanically linked to the air and fuel valves, it is
necessary that a controller independent of the triple valve actuator be incor-
porated. The method selected was to independently regulate the pressure
drop across the water metering valve.
110
-------
POSITION
FEED BACK
TEMPERATURE
TRIM TO AP REG
VALVE
ACTUATOR
AMP
R 1
1
1
r APREG — -,
/" X ' 1 Ww
/ WATER \1±. MAE.TERING -•••• • ',
\ PUMP J VALVE _ O, ,
\__^ 1 MOTOR !
ACTUATOR T
MECHANICAL - •
LINK BETWEEN
THREE VALVES
1
1 — v
COMBUSTION
FAN
AIR -|0j
VALVE ! 1
•»-
1
/^~^\ I_LJ r"*"
{FUEL)-r FUEL :«i j
^^ I «
L- APREG —
F
F
C
STEAM I
TEMP ;
OPEN I
POSITI
/CA CONTR
AMP (Vl^
vv
POSITION ^
FEED BACK
I 10V
T ^-\ m
i yVAAAA-4-V >— TH
POSITION po
•i? TRANS
T S
.OOP
ON
OL
NUAL
ROTTLE
SITION
TEAM
VAPOR WlL'" ^ THROTTLE EXHAUST
V«rUK — •-+ — . ^ . — ^ — . ^^ ... . .._ . _ . — . -»-
GENERATOR STEAM J p VALVE
OUI 1
pn<;
COMBUSTOR j^
LV^i
I 1 PRESSURE * /
EXHAUST TRIM GAIN /
tXHAUbT COMPENSATION-
'ROPORTIONAL STEAM
RESSURE — JSIILv
ONTROLLER
PRESSURE
TRIM SIGNAL
ITION
NS
NA^ '
FIGURE 81. COMBUSTOR/VAPOR GENERATOR CONTROL SYSTEM
111
-------
8.2 SYSTEM COMPONENTS
8.2.1 System Flow Arrangement
Figure 82 schematically describes the operation of all major fluid
flow components. The basic control approach is to mechanically schedule
fuel, air and water as a function of steam demand. A triple valve actuator
operates each of the metering systems from a signal from the electronic
controller. Linkage between the large diameter rotary air shear valve (see
Section 5) drives through cams the fuel and water metering valves. Both
valves are spool type with slotted orifices. The water metering valve is
pressure balanced to minimize actuation loads. A constant pressure drop of
16 psi is used to control fuel. Thus flow is a function of the valve's position
only. The mechanical linkage and cam drives allow a fixed schedule of fuel
flow to air flow for emissions requirements. Water flow orifice area is
controlled by the triple valve actuator but the pressure drop is regulated by
an electronic temperature trim actuator. A bypass differential pressure
regulator controls the pressure drop across the water metering orifice.
Pump flow must be at least 20 percent greater than the actual steam demand
to allow proper functioning. As pressure is increased on the upstream side
the piston of the valve moves to the right against the balance spring force.
The bypass orifice opens up until the pressure drop across the piston and
metering valve is equal to the spring force balance. By adjusting the spring
force with an electric position actuator the pressure drop and thus the flow
can be adjusted. An electronic temperature trim circuit adjusts the flow to
maintain a constant outlet temperature.
8.2.2 Air Valve and Triple Valve Mechanization
A complete description of the rotary shear valve is given in Section
5.3.2. Rotary positioning of the valve is accomplished by the attached
actuation tab shown on the upper right hand corner of Figure 29. This tab
projects through a seal in the combustor case wall. The tip of the air valves
actuation tab is visible at the top of Figure 83. Rotation of the air valve is
accomplished by a linear screw jack actuator driven by an electric motor.
A bolt drives the rotary shear valve through a bearing mounted in a slotted
cutout in the tab. Two cams on the bottom of the linear actuator carriage
operate the fuel (inboard) and the water metering valves. In Figure 83 the
fuel metering valve is installed. Valve position is controlled by the con-
tour of the cam which drives the spool down reducing the fuel metering area
as the air valve reduces the flow area into the combustor. Since constant
voltage is maintained to the fan motor, its speed and pressure are relatively
constant. Thus air flow is always in correct synchronization -with fuel flow;
the ratio depending upon the shape of the cam. Figure 84 shows the triple
valve actuator with both the fuel and water metering valves in operating
positions of 100 percent power output. Microswitches installed on the actuator
112
-------
rs
OPEN LOOPJCHEDULE AND
PRESSURE TRIM CONTROL
AIR VALVE
AIR METERING PORTS
SUPERHEATER
PRESSURE
SUPERHEATER .
TEMPERATURE
STEAM FLOW '
VARIABLE
DISCHARGE
PUMP
(FUNCTION
OF STEAM
FLOW)
FUEL PUMP
DISCHARGE
FUEL
REGULATOR
TRIPLE VALVE
ACTUATOR
WATER METERING VALVE
FUEL METERING VALVE
K WATER TO
% STEAM
v GENERATOR
FLOW METER
SENSING PORT
AP REGULATOR
FUEL TO
ROTATING
CUP
HYD OR ELECT
TRIM TEMPERATURE
ACTUATOR
o
PUMP
SUCTION
TEMPERATURE TRIM CONTROL
FIGURE 82. CONTROL SYSTEM FLOW SCHEMATIC
-------
Air Valve Actuation Tab
FIGURE 83. TRIPLE VALVE ACTUATOR SHOWING AIR VALVE ACTUATION
TAB
FIGURE 84
TRIPLE VALVE ACTUATOR WITH
FUEL AND WATER VALVES
INSTALLED
114
-------
carrier electrically limit travel at full and low flows. A rotary position
transducer attached to the screw drive shaft provides a position feedback
signal.
8.2.3 Air Valve (see Section 5)
8.2.4 Fuel Valve
A technical challenge exists in providing a linear fuel valve having a
40 to 1 turndown range. A goal of maintaining at least ±10 percent accuracy
in the fuel-air ratio schedule means that the valve must be repeatible within
better than 10%/40 or ±0.25 percent of the full range at the final turndown
position. To make the valve insensitive to contamination and allow use of low
pressure automotive pumps, a low pressure drop (15 psi) was used in its
design. Figures 85, 86, 87 and 88 illustrate the design of the valve. A
rectangular slit forms the fuel metering orifice. Flow area is controlled
by the position of a cam driven plunger.
For accuracy at these low flows (down to 3. 5 pounds per hour) it is
important to size the valve metering orifice to make it insensitive to viscosity
changes caused by temperature and blending. For normal practice a Reynolds
number greater than 4000 will ensure a relatively constant coefficient of
discharge.
Reynolds number is:
Wf 4A
R :
e n
•where
W£ is mass flow (Ibm/sec)
A is flow cross section area (ft )
P is perimeter of flow slot (ft)
M is kinematic viscosity of fuel
(M = 1.70 x 10"5 Ib sec/ft2 for JP-4 and M = 1.02 x 10~5 Ib
sec/ft2- for gasoline at 70°F)
4A/P is the hydraulic diameter of flow orifice
One hundred percent fueling rate is
137 Ib/hr = 0.0381 Ib/sec
115
-------
FIGURE 85. EXPLODED VIEW OF FUEL VALVE
im
I I I I I
6 INCHES
FIGURE 86. FUEL VALVE COMPONENTS
116
-------
FIGURE 87. METERING SLOT
E.
I I I I 1
6 INCHES
<
FIGURE 88. ASSEMBLED FUEL VALVE
117
-------
Assume JP-4 as a fuel and slot dimensions of 0.008 in. x 0.40 in. at 100
percent
4 x
Reynolds =.
0.038
32.17
= 4087
12
Figure 89 shows Reynolds number for JP-4 fuel at the low range lies
in the transition zone (R = 2000 through 4000) from laminar to turbulent
flow. The valve was sized this way such that slot dimensions did not become
too small and impair accurate function of valve. Low flow valve operation
will be experimentally investigated to determine if this has an unacceptable
effect on fuel flow. For gasoline (the normal fuel to be used) the Reynolds
number is well above the transition zone across the entire operating range.
Flow test calibrations confirmed the analysis of the valves character-
istics. Lapsed time to accumulate an accurately measured weight of fuel
was used to provide an absolute calibration of the valve. Figures 90 and 91
plot the data points against axial position of the valve's spool. Linearity
with ±5 percent is exhibited across the 40 to 1 range. Relatively linear
characteristics are maintained down to 0.9 pph. Valve gain was 368 pph
per inch of stroke.
8000
20
40 60 80
PERCENT OF FULL FLOW
100
FIGURE 89.
REYNOLDS NUMBER VS. PERCENT FUEL FLOW FOR JP-4
AND GASOLINE; 40 to 1 Turndown Ratios; Slot Dimensions
0. 40 Inch by 0. 008 Inch
118
-------
FUEL MIL-F-7024A
S.G. - .764
VISCOSITY » 1.17 CENTISTOKES
AP= 16.4 PSI
FLOW PATH: INSIDE TO OUTSIDE
GAIN • 368 PPH PER INCH
.15 .2
VALVE POSITION INCHES
FIGURE 90. FUEL VALVE CALIBRATION
119
-------
10 u
368 PPH PER INCH
0.
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—40 TO 1 RANGE
.01
.015 .02
VALVE POSITION INCHES
.025
.035
FIGURE 91. FUEL VALVE CALIBRATION 1 TO 10 PPM RANGE
8.2.5 Water Metering System
Water flow to the steam generator is controlled by a valve similar to
the fuel valve in concept. A slotted orifice area (Fig. 92) is controlled by
the position of a spool land actuated by a cam. Because the operating
pressure is over 1250 psia at maximum steam flow a pressure balanced
spool (Fig. 93) was incorporated. Water from the feed pump enters the
right hand side port of Figure 94 and flows around a reduced spool diameter
flow passage. It then passes through the variable orifice to the steam
generator. Flow rates are a function of the position of the spool and the
pressure drop across the valve. Calibrations of the valve were made at a
120
-------
FIGURE 92. WATER METERING VALVE ORIFICE
FIGURE 93. COMPONENT PARTS OF THE WATER METERING VALVE
121
-------
FIGURE 94. ASSEMBLED WATER METERING VALVE
constant 50 psi differential pressure with water. For each position of the
spool calibrated, an absolute flow rate was established by measuring the
weight of water flowing through the valve during a known period of time.
Figure 95 is a plot of the flow rates measured versus valve spool position.
From 1507 to 34 pounds per hour the standard RMS deviation from a straight
"best fit" line was 18 pounds per hour.
As was discussed earlier in this section, closed loop temperature
trim is providid by controlling the pressure differential across the water
metering valve. As a consequence the temperature of the outlet steam can
be regulated independent of the firing rate. A specially designed differential
pressure control valve was designed to perform this function. Water from
the pump enters a chamber on the high pressure side of the valve's piston
from the pump discharge. Pressure from the downstream side of the metering
valve is connected by a small sensing line to the opposite side of the piston.
Downstream pressure plus the spring force balance the piston causing it to
bypass excessive pump flow into the pump's suction line through a needle
orifice valve. A needle orifice and a large diameter piston were used in
this prototype system to ensure stability of the valve in the pulsating flow of
the test cells large triplex pump. Much smaller sizes appear quite practical
even with a triplex pump. An electric actuator positions the spring carrier
(thus controlling differential pressure) in response to closed loop position
control signals from a temperature trim circuit (Fig. 97). In a system that
employs a variable displacement pump, the differential pressure could be
controlled by regulating the displacement of the pump. It should be noted
that the differential pressure regulator has no direct effect upon the steam
pressure. Differential pressure across the metering valve is controlled
from 15 psi to 70 psi. Valve design is established to make the differential
pressure controller insensitive to input pressure requirements of the steam
generator. However, its operation becomes non-linear at pressures below
300 psia steam generator pressure. The inlet pressure to the steam
generator is only a function of steam demand, firing rate and flow pressure
122
-------
1600
& 1400
ID
O
oc 1200
UJ
Q.
^ 1000
2 800
600
K 400
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RESULTS ACROSS 40 TO 1 RANGE
.»•
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* BEST STATISTICAL FIT POINTS
i i i i i I i
0 .100 .200 .300 .400 .500 .600 .700 .800
VALVE INPUT POSITION (INCHES)
FIGURE 95. WATER METERING VALVE CALIBRATION AT 50 PSI
DIFFERENTIAL
FIGURE 96. DIFFERENTIAL PRESSURE CONTROL VALVE
COMPONENTS
123
-------
FIGURE 97. DIFFERENTIAL PRESSURE CONTROL VALVE ASSEMBLED
WITH ELECTRIC ACTUATOR
drop across the tube matrix. Valve performance in the system was good
with no apparent instabilities.
8.3 ELECTRONIC CONTROLS
The functional arrangement of the electronic system is shown in
Figure 81. For all tests performed, a breadboard system was used. It
consisted of two commercially available electronic controllers integrated
into a breadboard system to perform the necessary functions discussed
earlier in this section. A Barber-Colman Model 260T pressure controller
was used in the pressure controller block. A Barber-Colman Model 523C
Digital Temperature Setpoint controller was used in the block labeled as
temperature controller. Both of these units had adjustable proportional
bands and set points. Each also had an integral compensation in the forward
control loop. Summing amplifiers, actuator position, null setting position
feedback and signal conditioning circuits were breadboarded around these two
components to form the complete control system. Most of the features
required in the commercial controllers are not necessary in an integrated
system. However, their wide flexibility with respect to gain ranges and
control modes makes their use more effective in optimizing control functions
and gains. After gains and control compensations were established an
integrated electronic package was fabricated to perform all of the functions
of the electronic portions shown in Figure 81. Only four circuit boards were
necessary to obtain the necessary electronic functions. Figure 98 shows the
arrangement of these components. Program time allowed some limited
bench testing of these circuits but problems with temperature drift, com-
parator circuit operations, and thermocouple reference junction compensation
were not worked out of these circuits. As a consequence, no complete systems
tests were performed using the actual circuits shown in Figure 98. However,
they do show the relatively few number of components necessary to perform
the electronic functions necessary, and they do perform the same functions
of the system breadboard successfully used to control the steam generator.
124
-------
FIGURE 98.
ELECTRONIC COMPONENTS REQUIRED TO PERFORM
CONTROL FUNCTIONS
8.4 TEST RESULTS WITH PARALLEL FLOW STEAM GENERATOR
8.4. 1 Explanation of Test Results
Initial control system tests were performed with the combustor
discussed in Section 5 and the steam generator described in Section 6. Test
cell components were essentially the same as described in Section 6.3.
Important transient control system data was recorded on a high response
six channel Sanborn chart data system. Figures 99, 100, 101 and 102 are
reproductions of the Sanborn strip chart traces. At the top outlet steam
pressure is recorded from a strain gage transucer connected several inches
from the outlet tube of the steam generator. Each major ordinate division
is 0. 5 cm, scales are all based on Sanborn calibrations. Pressure is scaled
at 250 psia per cm. Next trace is the steam throttle position. Steam flow
is controlled by the position of the throttle valve. A contoured needle
plunger allows the throttle valve to control steam flow directly proportional
to valve position. Throttle valve position (steam flow rate) is the control
system "disturbance" for transient tests and also establishes steady state
flows. Total system volume between the outlet of the steam generatorand
the throttle valve is 40 cubic inches. Two linear position transducers are
mounted on the throttle valve drive mechanism. Output voltage from one
transducer supplies the main input to the electronic control system. This
-------
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1250
1000
"750 OUTLET PRESSURE
500 PSIA 250 pSI/cM
250
STEAM THROTTLE
POSITION 20% CM
TEMPERATURE ERROR
SIGNAL 1V/CM
FEED WATER FLOW
200 PPH/CM
1010
800
585
SUPER HEATER OUTLET
TEMPERATURE °F TYPE K
5 MV/CM
FUEL-AIR-WATER
METERING VALVES
POSITION 20%/CM
T
\
TIME IN SECONDS PER MARK
FIGURE 99. STEAM GENERATOR CONTROL TRANSIENTS, STEP CHANGES
PRIOR TO INSTALLATION WITH 44 LB/IN SPRING IN AP VALVE
127
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250
STEAM THROTTLE
POSITION 20% CM
TEMPERATURE ERROR
SIGNAL 1V/CM
FEED WATER FLOW
200 PPH/CM
1010
800 SUPER HEATER OUTLET
585 TEMPERATURE °F TYPE K
5MV/CM
FUEL-AIR-WATER
METERING VALVES
POSITION 20%/CM
\
TIME IN SECONDS PER MARK
FIGURE 100. LOW FREQUENCY 10 TO 30 PERCENT RAMP CYCLES AFTER
INSTALLATION OF HIGH GAIN AP VALVE
129
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TIME IN SECONDS PER MARK
FIGURE 101.
FULL POWER STEADY STATE PERFORMANCE WITH STEP
TRANSIENTS IN STEAM FLOW
131
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500 PSIA 250 PSI/CM
250
STEAM THROTTLE
POSITION 20% CM
TEMPERATURE ERROR
SIGNAL 1V/CM
FEED WATER FLOW
200 PPH/CM
1010
800 SUPER HEATER OUTLET
585 TEMPERATURE °F TYPE K
°F 5 MV/CM
FUEL-AIR-WATER
METERING VALVES
POSITION 20%/CM
TIME IN SECONDS PER MARK'
FIGURE 102. HIGH AMPLITUDE AND MAXIMUM FREQUENCY STEAM
FLOW CYCLING SYSTEM RESPONSE
133
-------
signal represents "steam flow input signal" to the basic open loop position
control circuit of the triple actuator servo. A feedback position transducer
on the triple valve actuator provides a nulling signal to the loop. Its position
is recorded on the bottom trace with the ordinate labeled fuel-air-water
metering valve position 20 percent per cm. As it can be seen from these
three traces, pressure control is obtained by operation of the triple valve
actuator in response to an open loop signal from the steam flow error signal
(second trace labled steam throttle position) and a closed loop pressure error
signal providing a proportional trim adjustment of the triple valve actuators
position.
Analysis of the ramp steam flow change in the middle of Figure 99
will illustrate the basic pressure loop control sequence. Pressure prior to
the steam flow change was constant at an average value shown on the data
trace of 890 psia. For the controls development tests the normal setpoint
used for the pressure system was 900 psia. A value lower than the design
value of 1000 psia was selected to prevent possible test disruptions from
overpressure safety circuit shutdowns or relief valve blowdowns. Tempera-
ture setpoint for controls experimentation was also reduced slightly below
the design value to 900°F to prevent possible damage in these highly develop-
mental phases of the program. Neither of these changes is believed to have
had significant effects on the conclusions drawn from the dynamic or steady
state controls tests. From the steady state operation at 890 psia a ramp
disturbance in steam flow was introduced by opening the steam throttle. At
a rate of approximately 5 percent per second the steam flow was dropped
from 50 percent to 32 percent in approximately four seconds. It should be
noted that the time scale is the bottom line on the trace. Each space between
marks on this line is one second of time. Portions of this line that show up
as a heavy solid line indicate the chart speed was reduced to more than one
second per mm (each small division is one mm). From the chart it is seen
that as soon as the steam valve is moved (thus changing steam flow) the fuel-
air-water (triple valve actuator) follows with the same slope, 5 percent valve
movement per second. A delay or dead time of 1. 5 seconds in the steam
generator's pressure response can be seen in the top trace. At 1. 5 seconds
from the initiation of steam flow reduction a rise in steam pressure starts
its maximum overshoot in 6 seconds. At 1. 5 seconds the closed loop pressure
error signal adds to the open loop ramp function command and increases the
slope of the triple valve actuator to 8. 5 degrees per second. A maximum
overshoot to 950 psia produces a sufficient error signal to cause the triple
valve actuator to move to 10 percent position. Gains in the pressure loop were
adjusted to have a 50 psia error signal result in a 50 percent trim correction
signal applied to the triple actuator. As pressure error is corrected, the
fuel-air-water metering valve proportionally follows and settles itself in
approximately 16 seconds to a zero error position of 20 percent with the
steam pressure stable at 890 psia.
135
-------
Three of the record traces in Fig. 99 were not discussed in the above
analysis of the response characteristics of the pressure loop. These three
charts are basically associated with the steam temperature control loop.
Superheater outlet temperature measured immediately at the outlet of the steam
generator is recorded on the second from bottom trace. A type K thermocouple
with an Inconel sheath 0.06 inch in diameter inserted in the steam flow is used
in this measurement. It also provides a feedback signal to the temperature
trim circuit. A non-linear temperature scale (for type K thermocouple) is
superimposed on the mV data recorded from the temperature sensor.
Temperature error (the difference between the setpoint and feedback signal)
is displayed in the third from top trace. It is scaled as 1 V per cm and was
used to investigate control system characteristics. Voltage recorded on this
trace essentially duplicates the superheater outlet temperature variations
but multiplied by a gain factor. Feedwater flow is the last trace to be dis-
cussed and is located fourth from the top. It records the linearized output
from a turbine water flow meter located in the steam generator inlet water
line. It is a high response measurement system and the oscillations shown
in the trace appear to be actual water flow. Weight calibrations of the tur-
bine flow meter with all conditions similar except no steam being generated
in the unit do not show these oscillations. Also during startups and shut-
downs, the high amplitude oscillations start and disappear with the beginning
and termination of boiling in the steam generator. In addition, when the
high efficiency steam generator (described in Section 7) was installed, the
amplitude of these oscillations were observed to be lower by a factor
between 2 to 5. A 50 cubic inch displacement accumulator is installed
down stream of the pump to dampen pressure oscillations. Normal pump
speed is approximately 120 rpm and being a triplex pump, its output pressure
and flow oscillation is normally 6 pulsations per second. Figure 99 shows a
flow oscillation an order of magnitude slower (approximately 0. 5 per second).
The accumulator did an effective job of damping out the normal pump oscilla-
tions. Below the prechange pressure of 400 psia, high amplitude pump
oscillations were observed. When the accumulator was unseated at 400 psia
oscillations disappeared. Water flow can be seen (Fig. 99) to follow the
triple valve actuator position. Temperature trim effects are difficult to
detect in Figure 99 since temperature variations are relatively small.
Overall performance of the closed loop control system as shown in Figure 99
is a control band of ±50 psi about a pressure setpoint and a band of ±40°F
about a setpoint. These are well within the goals of the program and are
considered good since the transients are representative of normal driving
requirements.
8.4.2 Cycling and Step Transient Performance
Figure 100 records the closed loop performance with cycling and
steady state dwells. Steam flow rate is the only external disturbance to the
136
-------
system. Output steam flow is manually varied by opening and closing the
steam throttle valve. Steam rate was 8 and 40 percent six times within
70 seconds in the center trace that produced the most severe excusions in
pressure and temperature. It should be noted the time base for Figure 100
is 1 mm per second. Maximum pressure overshoot in these traces was
120 psi and 75°F above the setpoint. It should be noted that a mechanical
stop was reached on the main metering valve (shown by the saturation lines
at -2 percent valve position) during some of the cycles. At this condition
the pressure loop is saturated, thus causing an increase in pressure.
Because of introduction of auxiliary air swirler with fixed flow and the desire
to maintain a fixed air-fuel ratio, fuel flow was 8 pounds per hour under this
condition. As a consequence, the steam rate was actually less than the
controlled corresponding mechanically limited firing rate. This was a
significant contributor to the pressure overshoot at the low steam flows.
Steady state control from 10 to 100 percent is recorded in Figure 101.
The time base on the left side of the trace was 0. 1 mm per second for steady
state performance and 1.0 mm per second during cycling transient analysis.
Steam flow was stepped up from 18 to 100 percent in 5 to 10 percent incre-
ments. Steam flow was left constant for approximately 100 seconds to allow
the temperature and pressure controls settle. Pressure remained within
a ±25 psi band around the setpoint during this calibration. Temperature
control exhibited a positive offset in the steady state control of 100°F
between the setpoint at 10 percent and the controlled temperature at 100 per-
cent. If the set point were defined at the 50 percent flow condition, a steady
state temperature control accuracy of ±50°F could be assigned to the system.
This offset change as a function of load was typical of all testing performed.
It was apparently caused by the fact that the water metering cam had signifi-
cant steady state errors with respect to the fuel metering cam profile. A
high proportional loop gain would eliminate this mechanical error, however,
a relatively low gain was necessary to prevent oscillations during transients.
Addition of integral mode control also gave significant improvements (some
results showed control within ±10°F), however, it was also destablizing and
was only used at very low gains.
Transient response tests were performed by stepping the steam rates
from high flows. The feedwater flow meter was disconnected from the
Sanborn during these tests since a manual valve was required to connect two
different turbine flow meters into the system to measure both high and low
flow ranges. Flow was only passed through the large turbine meter to pre-
vent damage to the low flow unit. As it is seen from the trace, the noise
output from the large meter is of high amplitude when the flow goes below
300 pph. Step changes show the typical overshoot in pressure expected in
a test loop with low volume (40 cubic inches) between the steam generator
and throttle valve. Upon closing the throttle steam flow drops almost
instantaneously. However, generation of steam is still at original levels
137
-------
until the fuel flow is dropped and energy distributions in the metal tube
matrix can be rebalanced. As a result steam pressure rapidly rises causing
an error signal to trim the triple valve actuator to further reduce fuel flow.
A pressure overshoot of 150 psi above the 900 psia setpoint occurs when
rapidly closing the steam throttle from 85 to 20 percent. The bottom trace
shows the overshoot to saturation that occurs due to the high pressure signal
driving the valve into its mechanical stop. Exactly the opposite sequence of
events occur when the throttle valve is rapidly opened. From Figure 101 the
magnitude of the pressure drop is seen to be -150 psi before the control
system restores the output pressure to its new steady state valve. From
the traces it is apparent that the pressure loop is highly damped with no
tendency to oscillate.
An explanation of the major cause of the pressure overshoot and under-
shoot can be seen in the transient cycles at the right of Figure 101. The
throttle valve was opened and closed as fast as the screw jack actuation mech-
anism would allow. Throttle valve rates as high as 50 percent per second
were recorded. The first five cycles the triple valve actuator speed was 15
percent per second (or only about 30 percent as fast as the throttle valve).
A small increase in actuator velocity gain was made in the last few cycles
by changing the drive motor output speed from the control panel. A small
but significant reduction in the magnitude of the pressure overshoot resulted
by this gain change. It is theorized that the magnitude of the pressure over-
shoot is associated with the lag between the rate at which steam flow is changed
and the corresponding response of the triple valve actuator. As the actuator
velocity is increased, a better instantaneous match between output steam flow
and input firing rate is achieved. Variation in temperature and pressure are
normal and cannot be fully eliminated with a single control mode. Water
level for steady state operation moves towards the inlet of the unit as steam
rate is decreased (see Appendix VII). Thus a rapid closing of the throttle
valve from high steam flow conditions to low flow results in the steam gener-
ator operating at a low power condition with a water level higher than normal
steady state levels. As a consequence, outlet steam temperature will be low-
er than normal until the controls correct the situation. The opposite sequence
occurs in going from low to high power rapidly. Water levels with an increas-
ing power step transient are now lower than steady state thereby giving effec-
tively greater length to the superheater. As a consequence, outlet steam
temperature will increase until corrections to the closed loop temperature
control system raise the water level back to the normal high power level. It
should be noted that these corrections require many seconds since the changes
in water level are dependent on feedwater flow rates. Another factor to be
noted is that with a fully modulated control the response time at low powers
is much lower since feedwater trim control rates are made proportionally
lower to allow stable operation. As can be seen from the temperature traces,
these are inherent characteristics of the system with a single control mode
138
-------
(f eedwater rate only). These can be partially eliminated by use of a separate
trim control with an auxiliary water injection for desuperheating. However,
the control is sufficiently accurate to not require the added complexity and
efficiency loss associated with desuperheating.
Maximum possible steam flow cycling rates were imposed on the
system as a final proofing for the controls. Operation of the steam throttle
was limited to the cycling rates shown in Figure 102 by its manual screw
jack actuation fixture. However, these cycles are considered more severe
than could be imposed by an automotive service. Dwells at high and low
flows after a series of cycles were maintained until the controls stabilized
pressure and temperature. From the performance trace it is seen that peak
to peak pressure excusions were at levels of as high as ±200 psi with normal
peaks about ±150 psi. Temperature variations were less than ±100"F about
a mean temperature line although they reached a peak of 150°F above the
nominal setpoint of 900°F.
Evaluation of the transient response characteristics of the fuel and
air control systems performance was made during the high velocity cycling.
Each of the emissions was plotted by a strip chart recorder during the eight
high velocity steam flow transients in the center of Figure 102. Figure 103
shows the results with zero time starting on the right hand side. The
relatively constant emission line at the right of each figure is a record of
the emissions during the steady state operation of the unit with the steam
throttle at 21 percent (see Fig. 102). Eight cycles between 21 and 96 per-
cent steam throttle position follow. As was discussed in Section 5, a cam
drive between the air valve and fuel valve schedules the air-fuel ratio
leaner at low powers. This is a programmed variation and is best analyzed
by reference to the CO^ trace. As it is seen, the syncronization between
the air and fuel control system is relatively good. The rounding off of the
minimum and maximum peaks is indicative that the CO2 cell cannot respond
rapidly enough to give an exact value of the instantaneous CC>2, but it does
fall within the bounds normally expected for the steady state values of CO?
at the extremes of the cycling. It should be noted that the variation in CC>2
(air-fuel ratio) peak values follows very closely the triple valve actuator
position on the bottom trace of Figure 102. The first peak (left one on
Fig. 102 and right one on Fig. 103) are both the highest with the fourth one
being the next highest. This type of correlation indicates good syncroniza-
tion between air and fuel flow controls. Emissions are placed on Figure 103
in order of response of (the particular analytical system. NOX measure-
ments by chemiluminescent techniques (Appendix I) were the slowest and
are shown at the top of the figure. As discussed in Section 5, decreasing
the air-fuel ratio (increasing CO?) causes an expected increase in NO with
a decrease in CO and HC. Corresponding data points are shown on the
curves with emission limits shown in the table. As it can be seen, only NO
is a problem during these severe transients. However, its characteristics
139
-------
70 PPM
75 PPM
9.0%
0-1000PPM
—I [~9
SEC.
co2 8.0%
J L
3.5%
150 PPM
0-15%
J L
9 SEC.
J U
20.0 PPM
0-500 PPM RANGE I I
9 SEC.^ I—
Power
Level
(%)
90
20
C02
(%)
7.7
6.15
CO (ppm)
Actual
290
350
Limit
462
377
NO (ppm)
Actual
60
35
Limit
31
26
HC (ppm)
Actual
30
60
Limit
120
84
FIGURE 103. TRANSIENT EMISSIONS DURING PEAK STEAM GENERATOR
CYCLING (See Fig. 102)
140
-------
are very similar to the steady state situation in which the emissions are
borderline at low flows and exceed limits at high flows. From this it is con-
cluded that the air-fuel ratio control is basically satisfactory and can main-
tain reasonably accurate control of air-fuel ratios during severe and con-
tinuous transient operation.
8. 5 HIGH EFFICIENCY MONOTUBE STEAM GENERATOR PERFORMANCE
Initial tests with the monotube steam generator indicated slower than
expected response characteristics. Open loop tests were performed to
measure some of the characteristics. A steady state condition was obtained
on the steam generator with both temperature and pressure loops opened.
While at a steady state temperature and pressure the water flow was
manually increased by 50 percent for approximately 10 seconds. Through
these step tests the triple valve and the throttle valve were maintained at
fixed positions of 20 percent. Figure 104 shows the results. Temperature
had a 30 to 40 second dead time before effects of the step change •were
observed. Pressure responds in approximately 10 seconds. A sinusoidal
open loop response test was also performed (Fig. 105). Both pressure and
temperature loops were opened. A sine wave voltage signal was imposed on
the amplifier controlling the position of the differential pressure regulator.
Feedwater input to the steam generator follows the sinusoidal in phase.
Output response of the steam generator is shown by the sinusoidal
disturbance to the outlet temperature. Input frequency was set at 0.01 cycle
per second to cause a sustained response from the steam generator. Note
that a high frequency input had no effect upon temperature. The amplitude
of the temperature response to the flow variation input was ±15°F. A phase
lag of 135 degrees in the temperature response was measured. These tests
and closed loop transient response indicated the need for a lead compensation
in the temperature control loop. A derivative amplifier operating the pro-
portional error signal added anticipation into the circuit to improve
performance and stability. Figure 106 shows the improvements made by
addition of rate compensation in the fore ward temperature control loop. In
the center of the trace a step change in throttle position from 15 to 52 per-
cent shows a typical temperature transient characteristic with a relatively
low gain derivative circuit. During the next step up in steam flow the
derivative circuit was removed leaving only a proportional amplifier in the
system. As it is seen, the addition of a relatively low gain derivative com-
pensator significantly reduced the magnitude of the overshoot and reduced
the settling time to approximately one half of the value in the uncompensated
characteristic. Further development optimization of this circuit was not
possible due to time schedules, but significant improvements seem likely.
141
-------
Figure 107 shows the response of the control system on the monotu.be
unit to high velocity cycling of the steam throttle. As in the discussion of
the parallel flow unit, the speed of response of the triple valve actuator is
not sufficient to prevent a large mismatch between steam flow rates and
firing rates. This factor combined with the low discharge volume between
the steam generator and throttle valve produce the pressure oscillations in
the output steam. Higher response and a larger volume (test volume was
40 cubic inches in tubing to throttle valve) will reduce the magnitude of the
pressure excusions. Temperature responded in a similar manner to the
parallel flow with the inherent water level mismatch between high and low
flow contributing to temperature fluxuations. Response to step changes in
throttle valve position are shown by Figure 108. Good pressure loop
characteristics are shown in these results but a tendency to oscillate in the
temperature loop. It should be noted that some experiments with integral
compensation were being performed in this test sequence causing greater
tendencies to oscillate.
142
-------
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143
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147
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1250
1000
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500 PSIA 250 PSI/CM
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POSITION 20% CM
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TIME IN SECONDS PER MARK
- 695 800 SUPER HEATER OUTLET
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149
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OUTLET
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250 PSI/CM
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151
-------
9
ORGANIC SYSTEMS
To provide a possible backup system for the EPA contractors, two
organic systems were designed. A Fluorinol-85 system was designed to fit
within the packaging constraints of TECO's power plant. AEF-78 was used
in the second unit integrated into Aerojet's vehicle constraints. Vapor
generator core analysis, combustor design modifications, and detail designs
of the steam generators were completed in this portion of the program.
Geoscience Limited performed the core analysis for the organic systems
(see Appendix VIII). No parts of the system were assembled, only critical
long lead time tubing for the core matrix was procured.
9. 1 FLUORINOL-85 SYSTEM DESIGN
The general flow arrangement for the vapor generator is shown in
the schematic below:
Superheated Vapor
Gas Out
-^
]
i
Preheater
r-*--i
i
r-—J
4
Superheater
1
^Saturated V
''4 -^
ape
Vaporize:
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Gas In
-^
Liquid In
Saturated Liquid
As shown, the working fluid flow arrangement is counterflow in the pre-
heater and the superheater, but is parallel flow in the vaporizer. The
superheater is located between the vaporizer and the preheater. This
arrangement requires a larger heat transfer area to achieve a given
efficiency than a pure counterflow arrangement, but it makes the problem
of preventing overheating of the superheater tube walls much more tractable.
153
-------
O 1
/ • * I
i r»«
Design Constraints
The design constraints for this unit were as follows:
F-85 Side
Flow
Pressure Drop
Outlet pressure
Inlet temperature
Outlet temperature
Heat transfer rate to organic
Gas Side
Air-fuel ratio
Flow
Pressure drop
Outlet pressure
Inlet temperature
Efficiency
Outlet temperature
10,000 Ibm/hr
130 psia maximum
700 psia
287°F (at max. power)
550°F
2.25 x 106 BTU/hr (reference)
25:1 (JP-5 fuel)
3740 Ibm/hr
3.0" H2O
atmospheric
2500°F (mean) ±250°F
81% based on HHV
86.5% based on LHV
427°F (reference)
(reference)
Maximum tube wall temperature was restricted to 575°F. Maximum
temperature for all other parts of the core was to be consistent with structural
requirements for the materials used. Maximum temperatures were to be
based on the assumptions that the gas side inlet temperature is 2750°F, and
the velocity at the gas side inlet plane of the vapor generator is 10 percent
above the design value.
The vaporizer unit maximum outer dimensions were to be 15 inches
width by 25 inches length by 8 inches core depth.
9. 1.2 Core
The preheater consists of 10 rows of 45 tubes each. The tubes are
1/4 inch OD by 0.020 wall thickness, and are made of low carbon steel. The
tubes centers have their longitudinal and transverse spacing equal to 1.25
tube diameters, or 5/16 inch. The tube rows are staggered, as shown in
the sketch below. 5/32"
O © JO
©00
-H 5/16" h
i
5/16"
T
154
-------
All tubes, in preheater, vaporizer, and superheater, are 22 inches long,
leaving approximately 1-1/2 inches on each end for headers and manifolds.
The total outside area of the preheater tubes is 53.9
The vaporizer consists of 2 rows of 21 tubes each. The tubes are
1/2 inch OD by 0.030 wall thickness, with 16 internal longitudinal fins 0.030
inch thick by 0.056 inch high. The tube rows are staggered, as in the pre-
heater. The longitudinal and transverse spacings of the tube centers are the
same, at 0.675 inch. The tubes are made of low carbon steel. Low carbon
steel was chosen as the tube material for three reasons: it improves the
heat transfer characteristics of the internal fins, it is possible to
fabricate in the internally finned configuration required, and it is amply
strong at the low tube wall temperatures required. The internal fins were
necessary in order to keep the tube wall temperatures below the specified
575°F maximum. As an additional measure to prevent tube wall overheating,
the tubes are protected by a 0.010 inch radial thickness air gap, formed by
placing a 0. 530 inch OD by 0.005 inch wall thickness AISI type 310 stainless
steel tube concentric with the internally finned tube.
The superheater uses three rows of 21 tubes each. The tube spacing,
tube material, and internal fin arrangement are the same as for the vaporizer.
The superheater tube walls are protected from overheating by a 0.020 inch
air gap, formed by placing a 0. 550 inch OD by 0.005 inch wall thickness AISI
type 321 stainless steel tube concentric with the internally finned tube. The
gas side heat transfer area in the vaporizer and superheater, based on the
outside area of the 1/2 inch OD tubes, is 25.4 ft .
The working fluid side pressure drop is estimated as 89 psi. This
includes an allowance for pressure drop in the manifolding. The gas side
pressure drop is estimated to be 4.3 inches of water. It appears that the
working fluid side pressure drop will be well within 130 psi maximum
allowable, while the gas side pressure drop will be in excess of the 3.0
inches of water initially specified.
9.1.3 Manifolds
The manifolds are subject to several conflicting requirements. They
must fit within the required envelope, they must be able to withstand the
temperatures and pressures to which they will be subjected, and they must
direct the flow in the manner required by the core design.
The flow arrangement must satisfy the conflicting requirements that
the working fluid velocities be high enough to prevent overheating of the tube
walls, but low enough to prevent excessive pressure drop on the working
fluid side.
155
-------
For the first eight rows of its passage through the preheater, the
working fluid flows in parallel through the 45 tubes of each row. At the end
of each of these rows it passes through a simple return bend manifold, which
directs it through the next succeeding row.
For the last two rows of the preheater, local gas temperatures are
high enough so that liquid flow velocities must be increased in order to pre-
vent overheating of the tube walls. The flow is redirected so that it passes
through only fifteen tubes in parallel, and each tube row consists of three
liquid flow passes.
To accomplish the change from 45 to 15 flow tubes in parallel, a
simple reducing flow manifold is used. Groups of three adjacent tubes are
manifolded across each other and then up to single tube in the next row.
Within the row simple return bends then connect each subsequent tube to
provide the 15 parallel flow tubes in rows 8, 9, and 10. From 10 the fifteen
parallel passages must be connected to the seven parallel flow passages in
the vaporization section. The flow from the 15 tubes on the preheater outlet
is connected by 6 passages to a common header that feeds the seven vaporiza-
tion tube inlets. Experience from the test bed system assisted in the
decision to place the seven flow restrictors at this point in the manifolding.
Liquid phase flow is ensured at this position and thus controllable pressure
drops are more likely than the inlet to the superheater (as in the test bed
system).
There is no collector manifold at the vaporizer outlet. Instead, the
flow is carried in seven parallel passages to seven tubes of the first row of
the superheater. (The "first" row is here defined as the first row through
which the fluid passes, which here is the row with the coolest local gas
temperatures.) The superheater flow is then in seven parallel passages with
three passes in each tube row, just as in the vaporizer. As in the vaporizer,
this flow is maintained by a complex system of simple return bends.
A collector manifold is provided at the superheater outlet, from which
the working fluid is carried to the expander.
9. 1.4 Combustor and Air Valve Design
In order to package within the engine compartment of the TECO
power plant, it was necessary to redesign the combustor and air valve to fit
within a 20 inch diameter envelope. A rectangular 15 by 25 inch vapor gener-
ator was also the largest frontal area unit that could be coupled to the 20
inch diameter combustor. Figure 109, 110 and 111 are envelope layouts of
the front, side and top view compared to the TECO vapor generator out-
line drawing. Basic to the approach taken is the goal of obtaining the largest
possible diameter for the combustor combined with the largest frontal area
156
-------
r
L FAN^^
1 MOTOR
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COMBUSTOR- >
VAPOR
GENERATOR
AIR INLET JjpjU FAN
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// GENERATOR \
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AIR-FUEL
~ V VALVE
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FIGURE 109. FLUORINOL-85 COMBUSTOR/VAPOR GENERATOR
(FRONT VIEW)
r
TECO COMBUSTOR-VAPOR
GENERATOR OUTLINE
25.0
J
FIGURE 110. FLUORINOL-85 COMBUSTOR/VAPOR GENERATOR
(TOP VIEW)
157
-------
FAN MOTOR
"TECO"COMBUSTOR-VAPOR
GENERATOR OUTLINE
FIGURE 111. FLUORINOL-85 COMBUSTOR/VAPOR GENERATOR
(SIDE VIEW)
vapor generator. Symmetrical aerodynamic flow between the combustor and
vapor generator with a minimum area transition are design features incor-
porated to minimize emissions and provide uniform stable air flow. The
basic configuration and design approach for the air supply, air metering,
atomization and combustion subsystems is the same as tested on the "test
bed system" (See Section 5). Air enters at the fan inlet coaxially mounted
to the combustor and fuel atomization rotating cup. A symmetrical air valve
having 20 ports (10 each for primary and secondary air) controls the air flow
circumferentially and axially into the combustor. Immediately before enter-
ing the primary zone, combustion air is caused to swirl by ten swirl vanes.
A 28 volt DC motor is mounted at any circumferential position around the
combustor to assist in packaging or installation. A second electric motor
drives the air valve positioning gear to control air and fuel flow into the
combustor. In a similar fashion to the fan drive motor the rotary fuel air-
valve actuator can be placed in any circumferential position around the com-
bustor to assist with envelope or reduce condenser fan air blockage.
As shown both the fan motor and fuel-air valve actuator would be at the rear
of the package thereby reducing their effects on condenser fan air blockage.
158
-------
Air Metering Valve Design
Because of space limitations with a 20 inch diameter combustor, it
was not possible to design a linear air valve. Although several advantages
are available with a linear design, one of the most important is a relatively
low sensitivity to leakage at the 40 to 1 turndown flow point. The close-off
height of the ports in this valve have been made as small as possible to help
reduce sensitivity at the low flows. Thus a given angular displacement will
produce a relatively small flow change. Since equal flow is essential through
each port, this feature is of key importance in the valve's design.
The design parameters are:
• Fan diffuser section remains uninterrupted up to 13.5 inch
diameter
• 11.3 inches water pressure is available at fan discharge
• Combined combustor/vapor generator pressure drop (excluding
metering orifice) is equal to 9.3 inches of water
• The range 0-56.7% full flow is comprised of primary flow only
and primary flow remains constant at 56.7% of full flow at
56.7-100% power setting
• Coefficient of discharge is assumed 0. 5 for the entire range
• Full air flow is to be 3740 Ib/hr
• At 100% flow, 43% of the air flows through the secondary ports
From these constraints the following was decided upon:
• Primary port height should remain constant for 0-56.7 full flow
with the exception of the low end of this range which was modified
for more sensitivity
• Secondary port height was made constant in the 56.7-100% full
flow range
• Primary port should gain increments of area in the 56.7-100%
range to compensate for combustor pressure loss
• Primary port will not close at leading side in range of 56.7-
100% flow
159
-------
• Ten sets of metering ports will be used
• Primary 0-56. 7% shall include a travel of 7 degrees and
secondary (56.7-100%) valve rotation of 10 degrees
Sample Calculation
The port shape is obtained by an incremental flow area analysis that
proportions secondary and primary flow with corresponding pressure drops
across air valve and combustor /vapor generator pressure drops.
• Area sizing of 65.6 percent power setting
• Pressure across metering orifice is
2
AP = 11.3-9.3 li^6- = 7.3 inches water
100
• Secondary flow is to be:
162° lb/hr = (0'20) 162° lb/hr
= 5.4 Ib/min
The following flow relationship is used in area calculation
Wa = 7.617 KAm v/r(P1 - PZ) (66)
where
Am = in (area of orifice)
y = lb/ft3 (density at standard conditions)
Wa = Ibs/min
AP = inch of H2O
K = coefficient of discharge
_ Wa _ _ 5.4 _
Ams = 7.617 K/X(P - P) ~ 7.617(0.5) 9-076x7.3
= 1. 94 sq. inch
Flow area for each of the secondary ports is 0. 194 in2 in the 65.6 percent
valve position. At 65.6 percent port height, hg, of secondary is 1.125 inch
and thus to opening width is 0.172 inch. The corresponding height of the
primary port is therefore equal to hs(Rs/Rp)(Amp/Ams) = 1. 125 x (9.313/
8.00)(0.072/0. 19) where Am is incremental area to be added to existing
primary area to maintain 35.4 Ib/min flow. This is determined through
equation (66) .
160
-------
R and R are effective radii of secondary and primary port areas
respectively.
The air metering port configuration determined through this method
is shown in Figure 112. The valve is positioned at 56.7 percent full flow.
Exhaust Gas Duct
Since the Solar unit fires down, while the TECO unit fires upward, it
was necessary to do some preliminary design work to ensure that Solar's
exhaust ducting would be compatible with envelope constraints imposed by
the engine compartment and by other components in the system. One of the
most severe of these restraints is imposed by the number two cross member,
just aft of the vapor generator, and the system components just above it. It
appears that the most convenient and acceptable location for the exhaust duct
to pass aft is through the shear web of this cross member.
A preliminary calculation was made to determine if it would be
possible to turn the exhaust flow aft in the space available between the
vaporizer core outlet and the number two cross member. In addition to
space constraints, the exhaust duct must satisfy the requirements that it
produces low pressure losses and causes no serious gas flow maldistributions
in the vaporizer core.
BACKUP PLATE
OPENING
SECONDARY
PORTS
PRIMARY
PORTS
FIGURE 112.
CENTER OF
COMBUSTOR
FLUORINOL-85 SYSTEM AIR METERING VALVE SHOWN
IN 56. 7 PERCENT POSITION
161
-------
It was found that the maximum dynamic head to which the flow could be
accelerated without a danger of flow maldistribution in the core was 0.43 in.
H^O. This corresponds to a duct area of 45. 1 in , or 22 in x 2.05 in inside
dimensions. It appears that a duct of these dimensions could easily pass
through the shear web of the number two cross member. The minimum bend
dimensions necessary for such a duct to achieve low losses are shown below:
15 in.
Core outlet face—*
The duct is a constant, 22 in wide. It is seen that, because of the accelerating
flow, very little turning space is required in order to achieve low losses.
9.2 AEF-78 SYSTEM DESIGN
Aerojet's engine compartment envelope constraints allow the use of a
combustor configuration of 26 inches outside diameter. Operation of the test
bed system's combustor at fuel flows up to 140 pph indicated the unit was
adequate for the 135 pph fuel flow required by Aerojet's AEF-78 combustor
and air supply control system. Figures 113, 114 and 115 show the integration
of the 23 inch combustion system with the rectangular vapor generator. A
simple aerodynamic flow system has been emphasized with a minimum of flow
area changes and no bends or turns in the combustor or vapor generator. As
in all of the designs, aerodynamic symmetry is emphasized to the greatest
extent possible, while using the largest possible face area for the vapor
generator.
Solar completed a detailed design of the heat exchanger incorporating
a core matrix established through a subcontract to Geoscience Ltd (see
Appendix VIII). The heat exchanger tube arrangement is a simple cross coun-
ter flow with all plain tubes on the gas side (no insulation or extended surfaces
are used). From inlet to outlet the matrix consists of seven rows of plain
0.25 inch diameter, 0.02 inch wall, carbon steel tubes. These tubes are
arranged in a staggered array with an axial and transverse pitch of 1.25 times
the diameter of the tube. Sixty tubes are in each row for a total of 420,
quarter inch tubes. All tubes in each row are manifolded together by simple
return bends to provide sixty parallel flow passages. The remainder of the
matrix consists of four rows of internally finned 0. 5 diameter tubes, 30 to a
row. Each row is manifolded in a manner to provide three sets of ten
162
-------
AIR INLET
FAN
1"DIA INLET
\-
"AEROJET" COMBUSTOR-VAPOR
GENERATOR OUTLINE
FIGURE 113. AEF-78 COMBUSTOR/VAPOR GENERATOR (FRONT VIEW)
"AEROJET"
COMBUSTOR-VAPOR
GENERATOR OUTLINE
26.4"-
FIGURE 114. AEF-78 COMBUSTOR/VAPOR GENERATOR (TOP VIEW)
163
-------
IIP L
23"DIA—
"AEROJET"
COMBUSTOR-VAPOR
GENERATOR OUTLINE
FIGURE 115. AEF-78 COMB US TOR/VAPOR GENERATOR (SIDE VIEW)
parallel flow paths per row. Fins on the inside of the tubes are straight and
parallel to the axis. Sixteen fins are equally spaced inside the tube. Fin
height is 0.056 inch by 0.03 inch thick. These fins and the higher velocity
fluid flow ensure that the tube wall temperature at the outlet row (adjacent
to the combustor) will not exceed 720°F. Other characteristics of the unit
are:
AEF-78 Side
Flow
Outlet pressure
Pressure drop
Inlet temperature
Outlet temperature
Heat transferred to AEF-78
Efficiency (based on LHV)
Gas Side
Outlet temperature
Flow
Pressure drop
Inlet temperature
19,300 Ib/hr
1000 psi
20 psi
396°F
650°F
2.02 x 106 BTU/hr
82.2%
550°F
3500 Ib/hr
2. 9 inches of water
2500°F
164
-------
9.3 INTERNALLY FINNED TUBING
The French Tube Division of Noranda Metal Industries has produced
samples of internally finned tubing which appear to be satisfactory. These
tubes are necessary in the design of the Fluorinol-85 and AEF-78 unit.
These tubes required a special fabrication development effort since they are
unique in design.
Tubing samples were examined under magnification. Some small
cracks were found, but they did not appear severe enough to affect either
the rupture strength or the low cycle fatigue life of the tubing. Photographs
of some of the worst cracks are shown under 50X and 500X magnification
in Figures 116 and 117.
Three samples of the internally finned tubing were heated to 900°F
for 10 minutes and then quenched in water repeatedly for a total of 18 cycles,
then pressurized at room temperature until they burst. The bursting
pressures were 8, 000 psi for two of the samples, and 7, 500 psi for the third.
Based on nominal wall thickness and typical room temperature properties
for low carbon steel, the samples should have burst at 7, 020 psi. All
samples showed considerable gross yielding before they burst.
FIGURE 116. TUBE CROSS SECTION
165
-------
A.
Magnification: SOX
Etchant: Nital
'+
-.
B.
Magnification: 500X
Etchant: Nital
FIGURE 117. PHOTOMICROGRAPHS OF TUBE-FIN WALL
(page 1 of 2)
166
-------
'
c.
Magnification: 500X
Etchant: Nital
D.
Magnification: SOX
Etchant: Nital
FIGURE 117. PHOTOMICROGRAPHS OF TUBE-FIN WALL
(page 2 of 2)
167
-------
1O
SYSTEM NOISE
Table IX lists overall dB(A)* noise levels of the Rankine system
with various components operational and different power settings. The data
was taken in an enclosed test cell with equipment positioned as shown in
Figure 118.
An estimation of semi-free field levels can be made by assuming the
room to be constructed with 100 percent absorbing material on all •walls and
ceiling (but not floor) and use of the following equation:
a2
Noise Level Reduction (db) = 101og._ (Ref. 16)
1U a -I
where a, is the actual composite room absorption coefficient given by
N
5- _ (% area x absorption coefficient)
_ K. — 1 K
1 Total Room Surface Area
where there are N types of surfaces in the actual test room and a^ is the
composite room absorption coefficient assuming 100 percent absorbing
material on all surfaces except the floor. This is determined by a similar
expression as used for ai .
From dimensions and surface materials given in Figure 118, a, is
found to be 0. 0362 and a?, 0. 812. Then the correction factor becomes:
10 Iog10 22.43 = 13.5 dB
The resultant dB(A) values are also given in Table IX.
* USA Standard for Sound-Level Meters, SI. 4, 1961,
169
-------
TABLE IX
OVERALL "A" SCALE WEIGHED SOUND LEVEL FOR
STEAM GENERATOR SYSTEM AND COMPONENTS
Operational
Components
Water Pump
(Outlet pressure
600 psi)
Fan
Water Pump, Fan
/*"* 1~ 4- ~1
Steam Generator
Percent
Air
Valve
Setting
-
30
20
20
50
50
77
Fan
Speed
(rpm)
0
5700
5780
5780
5500
5500
5300
Steam
Flow
(Ib/hr)
-
-
380
steam
gener-
ator
flooded
780
steam
gener-
ator
flooded
1200
Steam
Pressure
(psig)
-
-
900
-
900
400
900
Actual
Noise
Level
[dB(A)]
88
92
96
94
101
96
108
Approximate
"Semi-Free
Field" Cor-
rected Noise
Level
[dB(A)]
75
79
83
81
88
83
95
In order to extrapolate these sound measurements to that expected at
50 feet, six decibels should be substracted for each doubling of distance from
actual microphone position. However, this rule applies only if the position
was in the "far field" from the sound source. The semi-reverberent surround-
ings used in this test do not allow for determination of a far-field reference
required in this sort of correction. All that can be said is that the 50-foot
level could be less than indicated in the "semi-free field" column of Table IX.
Installation in an engine compartment with normal noise suppression and
treatment would produce further significant reductions of noise.
A sound level versus frequence trace (1/10 octave bandwidth) was
made for each of the conditions listed in Table DC. They are shown in Fig-
ures 119 through 125. All of these figures use the dB(A) frequency weight-
ing.
170
-------
H—6 FT. 10 IN.—I
26 FT. 9 IN.
4 FT.
FAN,
COMBUSTOR,
BOILER
PLASTER CEILING X^.
CONCRETE FLOOR Xa.0175
PLAN VIEW
—12 FT.—
-1 FT.
PAINTED
-BLOCK
BRICK
ys.0173
CEILING AREA
-WITH
ACOUSTIC-CELOTEX
1 IN. THICK
7=0.57
7 FT.
'4 MICROPHONE
END VIEW
FIGURE 118. NOISE EMISSION MICROPHONE LOCATION IN TEST CELL
171
-------
100
25 30 40 50 60 7080 100 150 200 250300 400 500600 800 1000 1500 2000250030004000 6000 10,000
FREQUENCY CPS
FIGURE 119. SOUND LEVEL, dB(A), VERSUS 1/10 OCTAVE FREQUENCY, WATER PUMP
ONLY - 600 PSI OUTLET PRESSURE - 88 dB(A) OVERALL
-------
•100
-J
OJ
25 30 40 50 60 70 80 100
150 200 250300 400 500600 800 1000 1500 2000250030004000 6000 10,000
FREQUENCY CPS
FIGURE 120. SOUND LEVEL, dB(A), VERSUS 1/10 OCTAVE FREQUENCY FAN ONLY -
5700 RPM - 30 PERCENT, 92 dB(A) OVERALL
-------
100
30
25 30 40 50 607080 100 150 200 250300 400 500600 8001000 1500 2000250030004000 6000 10,000
FREQUENCY CPS
FIGURE 121. SOUND LEVEL, dB(A), VERSUS 1/10 OCTAVE BAND FREQUENCY. FULL
SYSTEM, 20 PERCENT FAN, 380 LB/HR STEAM, 96 dB(A) OVERALL
(BACKGROUND INCLUDED)
-------
Water pump only
25 30 40 50 607080 100 150 200 250300 400 500600 8001000 1500 2000250030004000 6000 10,000
FREQUENCY CPS
FIGURE 122. SOUND LEVEL, dB(A), VERSUS 1/10 OCTAVE FREQUENCY. FULL
SYSTEM, 20 PERCENT FAN, BOILER FLOODED, 94 dB(A) OVERALL
(BACKGROUND INCLUDED)
-------
100
25 30 .40 50 60 70 80 100
150 200 250300 400 500600 8001000 1500 2000250030004000 6000 10,000
FREQUENCY CPS
FIGURE 123. SOUND LEVEL, dB(A), VERSUS 1/10 OCTAVE FREQUENCY. FULL
SYSTEM, 50 PERCENT AIR VALVE, 780 LB/HR STEAM, 101 dB(A)
OVERALL (BACKGROUND INCLUDED)
-------
100
CD
Water pump only
25 30 40 50 60 7080 100
150 200 250300 400 500600 8001000 1500 2000250030004000 6000 10,000
FREQUENCY CPS
FIGURE 124. SOUND LEVEL, dB(A), VERSUS 1/10 OCTAVE FREQUENCY. FULL SYSTEM
50 PERCENT FAN, BOILER FLOODED, 96 dB(A) OVERALL (BACKGROUND
INCLUDED)
-------
100
00
Water pump only
30
25 30 40 50 60 70 80 100
150 200 250300 400 500600 800 1000
FREQUENCY CPS
1500 2000250030004000 6000 10,000
FIGURE 125. SOUND LEVEL, dB(A), VERSUS 1/10 OCTAVE FREQUENCY. FULL SYSTEM,
77 PERCENT FAN, 1200 LB/HR STEAM, 108 dB(A) OVERALL (BACKGROUND
INCLUDED)
-------
The fan and its drive motor appear to be a dominant noise source
only at the 20 percent air valve setting and no steam (frequency = 2200 cps).
The water pump is sufficiently low in noise level that it contributes no more
than 1. 1 decibel to any of the full system measurements. The maximum
levels are due to changing from flooded to steam flow operation. This noise
is very broad-band with no specific peaks and occurs between 2,000 and
10,000 cps. Noise contributions during steam flow operation included throttle
valve sonic flow and high velocity venting into the test cell light gage metal
exhaust ducting. Elimination of this noise source would lower maximum
system power noise by approximately 16 db. By flooding the steam generator
at a fixed firing rate the noise contribution of the sonic flow throttle valve
could be eliminated. In an automotive system, a throttle valve would not
normally be used, and thus this contribution would be reduced. For normal
driving power the noise from the unit would be estimated to be 81 dB(A)
based on a free field correction. Installation in a properly treated enclosed
engine compartment would significantly reduce this level in a vehicle.
179
-------
APPENDIX I
181
-------
APPENDIX I
EMISSIONS SAMPLING EQUIPMENT AND DATA REDUCTION
The emissions analysis is based principally on four pieces of equipment;
namely, the Beckman Nondispersive Infrared Analyzer Model 315A for the analysis
of CO2, CO and NO; the Beckman Flame lonization Detector Model 402 for the analysis
of unburnt hydrocarbons, the Thermo Electron Company Chemiluminescent Detector
Model 10A for NOX and a Von Brand Smokemeter in conjunction with a Photovolt
Model 610 reflectometer for participates (smoke).
A schematic of the equipment is shown in Figure 197.
1. The Nondispersive Infrared (NDIR) Analyzer
The Beckman NDIR Analyzer Model 315A shown in Figure 198 consists of a
group of three separate instruments for the analysis of CO, CO£ and NO in parallel
through which the sample flows concurrently.
Each of the instruments consists of a reference cell, which is mechanically
chopped to a frequency of 10 Hz and directed into the cells. At the bottom of the cells
is a detecting chamber subdivided into two separate sections by a moveable metal
diaphragm. Both sides of the detecting chamber are fitted with the gaseous constituent
to be measured.
Asa result of the absorption of the infrared radiation in the sample cell,
more radiation reaches the reference side of the detecting chamber than on the sample
side hence the gas in the reference chamber expands and distends the diaphragm.
Positioned next to the diaphragm is a stationary metal button which, together with the
diaphragm, constitutes a variable capacitor.
When the chopper blocks the radiant beam, the gas cools and the diaphragm
returns to the neutral position. Thus the detector emits a variable capacitive signal,
which is superimposed on a high frequency carrier wave. This latter wave is then
modulated, rectified, and filtered to remove the radio frequency. After the signal is
amplified, it is again modulated, rectified and filtered. A DC signal results, which
is directly proportional to the concentration of the gas to be analyzed in the sample.
In this fashion, each of the three components, namely CO2» CO and NO is analyzed
183
-------
.-HYDROCARRON ANALYZER.
COMBUSTOR
EXHAUST
AIR IN
CHEMILUMINESCENT ANALYZER-
MULTI-POINT
SAMPLING
PROBE
NO A
PUMP
~*
DRYER
— »
NO CELL
CO
C02 CELL
CO Al
CO CELL
AIR-
FIGURE 197. SCHEMATIC OF SOLAR RESEARCH EMISSION ANALYSIS
EQUIPMENT
in their respective instruments. The capability of the equipment includes the measure-
ment of NO from a few parts per million up to about 1 percent; CO and CO2 are
measured up into the percent range.
2. The Beckman Model 402 Flame lonization Detector (FID)
In this equipment, shown in Figure 199, the sensor is a burner where a
regulated flow of sample gas passes through a flame sustained by a metered flow of
hydrogen and air. Within the flame, the hydrocarbon components of the sample gas
undergo a complex ionization that produces electrons and positive ions. Polarized
electrodes collect these ions, causing a current to flow through an electronic measur-
ing circuit. This ionization current is proportional to the rate at which carbon atoms
enter the burner, and is therefore a measure of the concentration of hydrocarbons in
the sample stream. The range of this equipment is linear from 5 ppm up to 5 percent.
It operates at an oven temperature of approximately 375° F.
184
-------
I
Nitric Oxide
11 Inch Coll
Flow Meters
Carbon Dioxide
).25 Inch Cell
Carbon Monoxide
2.5 Inch C
10.0 Inch Cell
FIGURE 198. BECKMAN NONDISPERSIVE INFRARED (NDIR) ANALYZER FOR
MONITORING CO, CO2, AND NO
185
-------
05
Electronic I'nits
FIGURE 199. BECK1VLAN MODEL 402 HYDROCARBON ANALYZER
-------
WINDOW
SAMPLE
PHOTOMULTIPLIER
TUBE
- DRY AIR IN
FIGURE 200. SCHEMATIC OF NO CHE MIL UMINE SCENT DETECTOR
3. The Thermo Electron Model 10A Chemiluminescent NOx Detector
The Chemiluminescent Detector which is shown schematically in Figure 200
operates on the basis of the reaction: NO + 03 — * NO£ + NO2* + O% whereby about
10 percent of the NO2 formed is at a higher than base energy level. As this excited
NO2 decays in the ground state energy on the form of photons is emitted.
The sample gas and an excess of ozone are reacted in a low pressure
reaction chamber. A photomultiplier tube measures the resultant emissions of
Chemiluminescent radiation through a window and provides an electrical signal which
is proportional to the-.NO concentration on the sample gas.
The instrument (see Fig. 201) is equipped with a converter which when operated
at a temperature of about 1200° F reduces any NO£ present on the sample gas to NO
hence NO£ can be monitored using a different method.
4. The Von Brand Smokemeter
Exhaust smoke is sampled with a Von Brand Smokemeter (Fig. 202). This
meters a quantity of the exhaust, set at 0. 108 cubic feet per square inch of filter
187
-------
FIGURE 201. THERMO ELECTRON CORP. CHEMILUMINESCENT ANALYZER
MODEL 10A
188
-------
•<-
FIGURE 202. VAN BRAND SMOKEMETER
-------
paper, through a Whatman #4 filter paper tape which moves at a constant speed of
4 inches per minute under the sample head. A small vacuum pump draws the sample
through the sample head and filter paper. The instrument enables traces to be made
during steady state and transient running conditions with the particulate carbon being
retained on the filter paper". The resultant trace is analyzed with a Photovolt 610
Reflectometer. With this instrument the filter paper smoke stain is assigned a
reflectance number by adopting a standard of 100 for a clean paper trace and zero for
black. The filter tape is positioned under a. search head which measures the amount
of light reflected from the strained tape when subjected to a standard amount of
incident light. The Von Brand Smoke number is defined as 100-% (relative reflectance
in %).
The sampling probe is mounted at any convenient station and is uncooled. The
probe design is either single or multipoint for averaging. The sample line is generally
kept as short as possible with the minimum of bends or kinks.
5. Calibration
The NDIR, FID and CL instruments depend upon calibration gases for their
operation. Calibration gases can be divided into two categories:
• Primary standards which have an accuracy of between 0.02 and 1
percent for the component being monitored present in the gas. (The
smaller the quantity of the component in the gas, the larger the
error becomes.)
• Certified standards in which the error in the component monitored is
between 2 and 5 percent. (This is again a function of the quantity of
concentration of the monitored component in the gas.)
In practice, certified standard gases are used as span gases and zero gases for the
calibration of the emission monitoring equipment noted above.
The certified standard gases were carefully analyzed against the ten
standardization procedures of the Environmental Protection Agency at the Mobile
Source Pollution Control Lab., Willow Run Airport, Ypsilanti, Michigan. Zero
gas is defined as a gas which does not contain the component being monitored and
therefore sets the zero point on the analytical equipment. The span gas is defined as
a certified standard analyzed against the primary standard which contains a certain
concentration level of the component being monitored, to which concentration value
the analyzing equipment is set. In the case of the flame ionization detector and the
190
-------
chemiluminescent detector, the relationship between concentration and readouts is
linear. When calibrating the NDIR several span gas ranges are necessary since the
relationship of signal to concentration is not quite linear.
6. Sample Treatment
From the sampling probe the NDIR gas sample passes through a particulate
filter and a water vapor drop which dries the sample by freezing out the water vapor.
Additional dehydration filters use chemical adsorbants to remove all traces of water
on the sample flowing to the NO cell which is extremely sensitive to water vapor inter-
ference. In the drying process the temperature of the sample drops to below 100°F.
The samples to the FID and CLD are routed to the instruments through
heated lines maintained at temperatures of about 250 and 350°F respectively after
passing through particulate filters.
7. Data Reduction
The data reduction procedure is based on IBM 360-50 Program EP-415; a
sample data point is given as Figure 203. The various steps are described below.
Raw Data
The raw data are given in parts per million (ppm) on a dry basis for CO and
NO by volume; ppm on a wet basis for hydrocarbons by volume (measured as carbon
atoms), and percent on a dry basis for CO2 by volume. If the NO/NOX is measured
on the chemiluminescent instrument, the reading is ppm wet.
Correction for Interference
The NO and CO observed readings are corrected for interference effects.
This does not apply if the NO/NOX is measured on the chemiluminescent instrument.
• The NO is corrected for the presence of CO, CO£ and HC in the
sample.
• The CO is converted for the presence of CO£ in the sample.
• No corrections are required to the CO£ and HC observed values.
Conversion to "Dry" Volumetric Concentrations
This is required for the HC and NO/NOX if measured on the chemiluminescent
instrument. The correction factor is a function of the air-fuel ratio (as measured or
calculated from a carbon balance) and ultimate analyses of the fuel (5C,%H£).
191
-------
awn-i. HL'1.1! F
TATA PP ff>!T
FUEL * CAPRON
I HV <*TU/LR
HUMIDITY GN/L1*
THF APOVF VAIUF
V'ATFR VAPOUR
FQUIV. PATIO
HUMIDITY FED
HUMIDITY CAL
.D.C. *3r'."> RRAYTPN 6-^-72
CftC r (\ir, I «gp SPFFD *
84.70000 r.O? 7
15.30000 HORSFPPWFP
13700.0 WA LP/SFC.
75.0000 WF I.P/HR
JP4 JIC3
1.0000
7.9^000
399.00000
0.4310?
22.00000
PF WA WAS CALniLATFD RASED ON AN ASSUMED
- 0 . 9f. 9 1 o
0.71314
1 .00000
O. QOQO]
NO
PPM qgsfRvpn 773.OOOOOO
PPM CORP. INT 773.000000
PPM WFT 773.00000O
PPM DRY ?«l. 705811
PPM DRYf, STDICH 1280.363781
PPM DRYt STHICCHUM 1730. 863281
nOM WFTEHUM 773.000000
273.000000
273.000000
773.000000
281.70581 1
1280.86328 1
1280.863781
273.000000
AS AS
( N(12 ) ( N02 J
GM/KG FUFL 31.003621 31.008671
GM/HP 310.03618? 310.08618?
GM/HR PHP 0.344973 0.344923
LB/HR 0.633612 0*633612
L3/HD PHP 0.000760 0.000760
MICPOGM/CUP. MFTER 5^1569.9375 531569.9375
F COMR en
c fpM R H/C
F COMB COEH/C
99. 0771 73
99.960360
99.04651 3
CO? RATIO MFAS/CALC WITH CH£H/C
0. 999003
loor C02 PATIO.
CO
16.^00000 •
14.500801
14.05266?
14.500801
68.034958
68.034958
14.05266?
CD
0.971808
9.718081
0.010810
0.021424
0.000024
16659.3633
H/C ASCI
9.000000
9.000000
9.000000
9.287009
47.226257
42.226257
9.000000
( H/C I
AS
CHI. 85
0.308270
3.082703
0.003429
0.006796
O.OOO008
5284.5703
FIGURE 203. DATA REDUCTION
Conversion to "Equivalent Stoichiometric" Volumetric Concentrations
This step involves a multiplication by the ratio
Actual A/F ratio
Stoichiometric A/F ratio
which is the inverse of the equivalence ratio. The numerator is either obtained from
direct measurements of the air and fuel flows or a carbon balance. The demoninator
is calculated from the ultimate analysis of the fuel.
192
-------
Correction of NO/NOy for Ambient Humidity Effects
The formation rate of NO is sensitive to the humidity of the test air supply.
The NO results are therefore normalized to a standard humidity condition of 75
grains/pound dry air in line with the correction formula laid down in the Federal
Register of July 2, 1971, Vol. 35, No. 128, Part II, EPA "Exhaust Emission Standards
and Test Procedures".
Mass Concentrations
From the wet basis, volumetric concentrations (corrected for humidity in
the case of NO/NOX), and the mass concentrations are calculated using the densities
of the various species as laid down in the Federal Register of July 2, 1971, Vol. 35,
No. 128, Part II, EPA "Exhaust Emission Standards and Test Procedures". The
NOX is expressed as NO2 and the HC is assumed to be CH^ 85.
Combustion Efficiency and Carbon Balance
Combustion efficiencies are extracted from the amounts of CO and HC. (The
HC in this case is assumed to be completely unburnt fuel.) A carbon balance is made
on the air and fuel flows into the combustor with the CO, CO2 and HC flows out of the
combustor to check on the measurement accuracies. This can only be carried out of
the air, and fuel flows are measured directly; otherwise the carbon balance must be
assumed to obtain a test air-fuel ratio.
193
-------
SOLAR
RESEARCH
MEMORANDUM
R73J-3441
January 11, 1973
To:
cc:
T. E. Duffy. Research Staff Engineer
W. A. Compton J. Urick, Eng.
J. V. Long File (PER)
D. J. White Ref. File
J. Stice
From: P. B. Roberts
Senior Analytical Engineer
Subject: ANALYSIS OF EPA GAS SAMPLES. S.O. 6-3827-7,
EWO 6033425
I. BACKGROUND
As part of a continuing effort to ensure the use of uniform gas analysis
instrumentation, calibration and sampling methods among its contractors working
on low emission programs, the Environmental Protection Agency delivered to
Solar last November a number of cylinders containing gaseous constituents of
unreported concentration for our analysis.
The objective of the exercise was for Solar Research to label the
concentrations of the various cylinders and to then return the cylinders to the
EPA for presumably subsequent analysis by other contractors. A listing of the
cylinder constituents and their approximate ranges of concentration was supplied
by the EPA.
As a secondary effort during this investigation, the cylinders were
analyzed by Engineering on their own separate instrumentation for comparison
purposes (using their own charge number).
The work was carried out during the month of December 1972.
II. INSTRUMENTATION
The Research instrumentation comprises:
• Beckman 315-A NDIR for CO2 and CO
. Beckman 402 FID for UHC
SOLAR
DIVISION OF INTERNATIONAL HARVESTER COMPANY
195
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R73J-3441
Page 2
• Thermo Electron 10A CLD for NO/NOX with molybdenum coil converter
for NO 2 reduction
The Engineering instrumentation is essentially identical to the above with
the exception that the NDIR instrument is the Model 315-B which incorporates solid
state electronics and has different CO and CO ranges.
The N©2 converter is of the standard stainless steel coil type.
IH. PROCEDURE
The procedure adopted was for Research to initially calibrate the instrumen-
tation with its available Gold Standard gases. These have been labeled by the EPA and
are normally retained for use in labeling secondary standards only. The NDIR curves
were checked at four points on each range.
The EPA samples were then analyzed by introducing them into the instrumen-
tation through the calibration gas input points rather than through the various heated
sample lines. As it was know that the diluent of each sample was either dry air or
nitrogen, either method would give identical results with clean sample lines.
Engineering's results were obtained after only preliminary calibration with
their normal (secondary standard) gases and a 'one point span check on the NDIR
ranges. It should be understood that Engineering utilizes the Research Golden
Standard gases for initial labeling of calibration gases (secondary standards) upon
delivery.
IV. RESULTS
The results are shown in tabular form in Table I. Column #1 lists the
approximate ranges as reported by the EPA, #2 shows the Research results and
#3 gives the Engineering results.
Total time charged for Research to complete the analysis, including reporting
is 35 hours.
In comparison, Engineering with its reduced calibration effort completed
the analysis within 8 hours.
The expected level of accuracy for the Research results is high except for
the three low concentration CO samples. Due to an absence of suitable gold standards
for this instrument range only a single point calibration could be made with a secondary
standard. Any error however is unlikely to be more than ±10 percent.
196
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R73J-3441
Page 3
TABLE I
ANALYSIS RESULTS
#1
EPA Reported
#2
#3
Cylinder
No.
B-1094
B739
37562
37853
37730
37758
37769
37793
37699
37749
B-880
37768
37581
37723
37847
37764
Gas
NO
NO
C3H8
C3H8
CO
CO
CO
CO
CO
CO
co2
co2
C02
C02
—
—
Diluent
N2
N2
Air
Air
N2
N2
N2
N2
N2
N2
N2
N2
N2
N2
N2 zero
N2 zero
Concentration
Range
0-100 ppm
0-50 ppm
0-50 ppm
0-200 ppm
0-100 ppm
0-100 ppm
0-100 ppm
0-500 ppm
0-500 ppm
0-2000 ppm
0-5%
0-5%
0-15%
0-15%
gasO. 1 ppm HC
<0. 1 ppm HC
Research Results
96. 5 (3)
47.1
33.3 (1)
150.7 (1)
21.0
46.5
78.0
271.0
447.0
-(4)
2.07
4.28
8.82
13.60
0.0
0.0
Engineering
Results
94.0
47.0
30.7 (1)
127.5 (1)
21.0
49.0
82.0
279.0
463.0
>1000.0 (2)
2.0
4.2
8.6
13.4
0.1
0.0
(1) measured as methane -
(2) Engineering instrumentation has maximum CO scale 0 -1000 ppm
(3) 2.0 ppm NO2 analyzed in addition
(4) Not analyzed. Available scales are 0-1000 ppm and 0-2.5%
V. RECOMMENDATIONS
•Obtain sufficient gold standard gases to adequately calibrate every range
of each instrument
•Attempt to fall into line more with previous informal EPA recommendations
as to procedures, materials, calibration gases, etc. Ref. : S. I.C. "Trip
Report, EPA Ann Arbor, Michigan", P. B. Roberts to J. Watkins,
31 Jan. 1972
•Adopt the procedure of regular cross checks between the Engineering
and Research instrumentation.
PBRrst
197
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APPENDIX II
199
-------
EFFECT OF SCALING ON MATRIX METAL WEIGHT AND WATER HOLD UP
Kays and London (Ref. 5), on page 10 of their book, show that for a heat
exchanger with the gas side resistance dominating,
C \
*-" (i>
6 mm
If NOg. and C /Cmi_ remain constant,
o o
stg
g
Ordinarily, St can be described by an equation of the form
St = KRe~a (3)
/W\"a -a
DH 4
NTU oc DH (5)
The exponent a is normally in the range 0.4 < a < 0.7.
Now if tube stress is to remain constant, then tube wall thickness is
proportional to tube diameter, which means that for geometrically similar bare tube
arrangements, both matrix metal weight and hold up volume are proportional to core
volume. The core volume is given by
„ -1
201
-------
From equation (5), if NTU is maintained constant;
55 * »
Combining equations (6) and (7), one sees that if NTU and W are maintained
constant,
W V1 + a 1 + a
1 (8)
So for geometrically similar bare tube matrices, both core weight and hold up volume
tend to be proportional to tube diameter to the (1 + a) power.
For a finned tube, the fin thickness must increase more rapidly than the tube
diameter, if fin effectiveness is to remain constant. As is shown on page 9 of Kays
and London, if fin effectiveness is to remain constant, then, for geometrically
similar arrangements,
2
d « DH h (9)
/ W \ / W Y ~ a -a
but h«st_oc_ DH. (10)
-
so 6 « DH — (11)
Since, for a geometrically similar matrix, spacing between the fins is
proportional to DH, core weight for the fins is scaled by
M « (12)
2
so M « DH (13)
From the above, it is seen that it is advantageous from the standpoint of core
weight and hold up to keep the tube diameters as small as possible.
202
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APPENDIX III
203
-------
SIZING ANALYSIS
The method used to size the various parts of the unit was essentially
the same in all cases. First, inlet and outlet gas and water conditions were
estimated, using the design requirements and an overall heat balance. This
fixed the required effectiveness, and the thermal capacitance rate ratio for
the two streams. Next, the amount of heat transfer surface needed to
produce the required effectiveness was estimated using the NTU-effectiveness
approach described in Chapter 2 of Kays and London (Ref. 5). This approach
uses the functional relationship
€ = f(NTU, C . /C , flow arrangement) (1)
1111,11
where NTU = (2)
C = WCp (3)
ATmin
(4)
T, . - T .
hi ci
s
Various gas side surfaces were evaluated on the basis of ease of manufactur-
ing, water hold up volume, gas side pressure drop, and the number of tube
rows needed to produce the required NTU. It was known that in all areas
the dominant film resistance would be on the gas side, so the actual config-
uration of the water side was relatively unimportant from a heat transfer
viewpoint.
With the gas side surface chosen, the number of parallel passages
on the water side could be determined on the basis of water side pressure
drop considerations. With the water side flow arrangement known, a check
was made to ensure that the water side thermal resistance did not reduce the
NTU below the required limits.
Preheater
The fluid stream temperature estimates used to size the preheater
are tabulated below.
Gas Side Water Side
Inlet 1090°F 160°F
Outlet 306°F 56l°F (saturated)
205
-------
These conditions required
€ = 0.851 C • /C = 0.537
mm max
Since most of the actual water hold up in the vaporizer is in the pre-
heater, it is important to minimize the tube volume in this section. For this
reason, a finned tube.configuration was chosen for the pr cheater. The pre-
heater is in a cold part of the exchanger, so copper fins were chosen because
of their high fin effectiveness and the ease with which they are manufactured.
It was decided to use a monotube arrangement if pressure drops permitted,
since this would tend toward simplicity in the manifolding arrangements and
easy, low-cost assembly of the unit.
The final configuration was three rows of 3/8 OD 321 stainless steel
tubing, with 1.8 inch high by 0.012 inch thick copper fins, with 15 fins per
inch. Pertinent data for this surface are shown in the table below.
S = 0.717 inch D = 0.784 inch
a = 0.297 a = 151.1 ft2/ft3
DH = 0.00785ft fin area/total area = 0.859
This surface is quite similar to the surface of Figure 96 in Kays and
London (Ref. 5).
Gas side heat transfer was evaluated using the data provided in
Figure 96 of Kays and London. Pertinent results were:
Re = 404
h = 32.4 BTU/ft2°F
tube length/row = 39.0ft
area/row = 22.3ft2
fin effectiveness = 0.98
gas side AU/row = 710 BTU/hr°F
Water side friction factor and pressure drop were evaluated using the
standard Darcy-Weisbach friction factor. It was found that a monotube
arrangement would allow the pressure drop to be held down to a reasonable
level. The water side pressure drop results were:
Re = 1.329 x 105
f = 0.01804
Ap - 45-3 Psi
Heat transfer coefficient on the water side was estimated using the
correlation on page 399 of Bird, Stewart, and Lightfoot (Ref. 6).
206
-------
(St) (Pr)
2/3 _
0.026
175
(5)
The following results were obtained for the water side heat transfer.
h = 5070 BTU/hr ft2°F
water side AU/row = 17,120 BTU/hr°F
area/ row = 3.38
With the above results for the gas and water side, the results for
overall heat transfer were as shown below.
AU/ row
Cmin
NTU/row
682 BTU/hr°F
705 BTU/hr°F
0.967
The effectiveness for each row of tube was estimated using equation
15 on page 12 of Kays and London for the crossflow arrangement with the
maximum capacitance side mixed and the minimum side (in this case, the
gas side), unmixed.
r =
-NTU
(6)
max
'mm
- e
/C
'min max
(7)
The number of rows necessary to achieve the required effectiveness
was determined from equation 17b on page 13 of Kays and London, which
applies to multipass cross counterflow exchangers with each fluid mixed
after each pass.
€ =
1 - £ C . /C
p mm max
1 - e
n
- 1
(8)
I-e C . /C
p mm max
1-6
n
mm
max
207
-------
The results obtained for the preheater were:
€p = 0.527
f3 = 0.843
62 = 0.737
C4 = 0.903
Although three rows do not quite give the effectiveness of 0.851 which was
desired, it was decided that three rows would provide adequate preheating.
This was indeed the case, as is shown by Table VI. Water leaves the pre-
heater with only about 5° F of subcooling. Experience with the present
test bed unit and the available literature on the subject (discussed in Section
6), indicate that such a small degree of subcooling is unlikely to produce
any hydrodynamic stability problems in the vaporizer, especially in the
monotube configuration finally chosen for the vaporizer.
Vaporizer
The estimates of inlet fluid conditions shown below were used to
size the vaporizer.
Gas Side Water Side
2500°F Saturated at
545°F, 1000 psia
It was desired that the vaporizer have an ample safety margin for burnout,
and that it produce an outlet steam quality of at least 50 percent. This
would allow dryout to take place in a low temperature part of the boiler, at
a high quality, so that the reduction in water side heat transfer associated
with dryout would not be severe. A wide safety margin for burnout at full
load should help keep the safety margin adequate at part load, even when
operating at high flame temperatures. As in the preheater, it is desirable
that the tube volume be kept down to a minimum, and that the manifolding
arrangement be as simple as possible.
It was felt that the best arrangement for satisfying these requirements
was two rows of 5/8 OD 321 stainless steel tubing, arranged similar to the
surface of Figure 48 of Kays and London. While this is certainly not the most
compact possible bare tube arrangement, it was felt that it was the most
compact which would allow reasonable ease of manufacturing. It was decided
to use a monotube arrangement in the vaporizer for improved stability, if
pressure drop and dryout considerations would permit it. Pertinent data for
the gas side heat transfer surface are shown below.
208
-------
S = 0.9375 inch
a = 0.333
DR = 0.0496 feet
D = 0.9375 inch
a = 26.9 ft2/ft3
Gas side heat transfer was evaluated by the same method as used in
the preheater design. The results were
Re = 1343
h = 24.2 BTU/hr ft2°F
tube length/row = 29.8ft
area/row = 4. 88 ft2
gas side AU/row = 118. 1 BTU/hr°F
It was assumed that the water side thermal resistance would be negligible,
as long as dryout did not occur.
With the above information for heat transfer, the effectiveness of
the vaporizer was determined from
6-1- e-NTU (9)
This is the relation given on page 12 of Kays and London for heat exchangers
with C,,. 1C =0. The results were
mm' max
cmin = 823 BTU/hr°F
€ = 0.249
NTU = 0.287
exit quality = 51.3%
With these results, the probability of dryout in the vaporizer was estimated
using the criteria described in Appendix IV. The Baker plot indicates that
the flow is indeed in the annular flow regime. The Westinghouse Atomic
Power Division correlation, presented as equation (4) in Appendix IV,
x , =(
cr" do)
- _L* 0.548
209
-------
was applied with the input data and results shown below. In equation (10),
must be input in inches, and G in Ibm/hr ft2'.
De = 0. 561 inch
G = 0.750 x 106 Ibm/hr ft2
xcr.t = 76.1%
L = 59.6 ft (for two monotube rows in series)
L/D = 1273
(VV/VL) = 20.6
It is seen that there is amply safety margin for dryout, so a monotube
arrangement in the vaporizer is certainly satisfactory from that standpoint.
Using Tippets' correlation, described in equations 5 through 9 of
Appendix V,
x = 0.513
qcrit = 2'746 x 1()5 BTU/hr ft2
a = 0.808
q = 4.23 x 104 BTU/hr ft2
The margin of safety indicated by Tippets' correlation is not so large as it
seems. An increase in flux increases the exit quality, which in turn,
sharply reduces the critical heat flux predicted by Tippets' correlation.
Using the values of quality and void fraction shown above, the vapor
velocity can be calculated from
Gxvv
Vv =—T
with vv = 0.445 ft3/lbm
Vv = 59.1 ft/sec
This is about the maximum velocity which Bennet and his co-workers (Ref. 7)
found could be tolerated without danger that the liquid film would be blown
off the wall.
210
-------
After considering the results of all of the above correlations, it was
decided that the vaporizer would almost certainly have an ample margin of
safety for dryout at full load, and that a monotube arrangement would not
produce high enough vapor velocities to cause premature dryout.
Pressure drop in the vaporizer was estimated by the method presented
by Martinelli and Nelson (Ref. 8). Their correlation is
P
, (P, x ) (12)
and r = r(p, XG) (14)
The functions f1 and r are tabulated in Reference 8. Applying the method of
Reference 8 to the vaporizer tube rows, the following results were obtained:
G = 0.750 x 106 Ibm/hr ft2
=. 0.01734
p = 1000 psia
f = 9.13
ReLQ = 1.452 x 10. 5
Ap = 1.118psi/row
LO
xe = 0.513
r = 0.155
Ap = 21.9 psi
For automotive boilers, it appears that vaporizer velocities low enough to
prevent premature dryout will almost always ensure that the pressure drop
in the vaporizer will be quite acceptable low.
211
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Dryer
It is in the dryer that the working fluid is changed from saturated
vapor to superheated steam. Since this is the case, the liquid film on the
wall must at some point become so thin that it breaks up, with a consequent
reduction in the water side heat transfer coefficient. The dryer tube is
placed in a relatively cool part of the exchanger, so that a decrease in water
side heat transfer coefficient will not produce excessive increases in tube
wall temperatures. The tube mass flow rates are such that wall dryout will
occur at relatively high quality. When this is the case, fairly high heat
transfer coefficients are maintained on the water side even after dryout,
thus keeping tube wall temperatures within reasonable limits.
In order to keep the dryer tube pressure drops within reasonable
limits, it was necessary to use either two parallel passages of 3/8 OD tube,
or a 5/8 OD monotube. In order to achieve reasonable compactness, a finned
tube arrangement was desired. Because of the high fin tip temperatures
expected in the dryer, it was felt that the fins should be stainless steel rather
than copper. In order to achieve reasonably high fin effectiveness with
stainless steel fins, short fins were necessary. In order to get high fin area/
total area with short fins, tube diameters must be small. For this reason,
the parallel flow 3/8 OD finned tube arrangement was chosen.
The dryer therefore consists of one row of 3/8 OD finned tube, arranged
in two parallel flow passages. The tube spacing and fin arrangement are
exactly the same as for the preheater, but the fins are 304 stainless steel
rather than copper.
For one row of 3/8 OD tube with stainless steel fins, the gas side
heat transfer results were:
Re = 296
h = 42.0 BTU/hr ft2°F
tube length/row = 19.5ft
area/row = 22.3 ft2
fin effectiveness = 0.76
gas side AU/row = 743 BTU/hr°F
As a preliminary estimate, it was assumed that the gas side heat transfer
would be the dominating resistance. An initial guess of both gas and water
side entrance and exit temperatures was made, as shown below.
212
-------
Gas Side Water Side
Inlet 1768CF 545 F, 1000 psia, 51.3% quality
Exit 1061°F 707 F, 1000 psia
With these assumed conditions,
C . = 760 BTU/hr°F
mm
Cmi_/C = 0.229
11:1111 max
NTU/row = 0.966
Putting these values into equation (7),
6 = 0.576
When this value for the effectiveness of the dryer was factored into a
more precise heat balance for the overall exchanger, the values shown in
Table VII resulted.
Pressure drop in the dryer was estimated by the method of Reference 8
for the two phase flow part of the dryer (also see Appendix VI).
The functions f^ and r of Reference 8 allow the calculation of two
phase pressure drop when the heat input to the water is uniform with length
and the water is saturated at inlet. It follows, then, that if the heat input is
uniform but the water has some quality at inlet, equation (12) and (13) can be
generalized to
G* ' ' - *>e
Friction pressure drop in the superheated regime was estimated using the
standard Darcy-Weisbach friction factor approach. Acceleration pressure
drop in the superheated regime was estimated from the assumed saturation
and exit conditions. The results were:
Two Phase Flow
G = 1.004 x 106 Ibm/hr ft2 ReLO = l-145xl05
fT _. = 0.01848 ApT ^ = 1.648psi
J-iVj T .O
p = 1000 psia xj = 0.513
xe = 1.00 f = 9.13
213
-------
f = 14.9 r- = 0.155
C I
r = 0.424 Ap = 14.03 psi
6
Superheated
R = 4.41 x 105 f = 0.01544
Apf = 20.3 psi Ap = 2.8 psi
I cLC C
Ap = 23. 1 psi
Stability of the parallel flow paths was investigated using Quandt's
(Ref. 4) approach. The approach to designing for stability in the dryer
has been extremely conservative. Quality is quite high at entrance to the
dryer at 57.3 percent. As mentioned in Section 6, this should have a
strongly stabilizing effect on the flow. There are only two parallel passages
in the dryer. The dryer tube material has been selected as Hastelloy X,
as alloy with excellent strength at temperatures much higher than those
anticipated for the dryer tube walls. To prevent possible development test
burnout failures the dryer material and wall thickness selected allows
operation at the maximum flow distortion. Even with no flow through one
of the passages the tube wall would not fail at maximum system pressure
and the wall in equilibrium temperature with the combustion gases. This
precaution was necessary as demonstrated by the test results. No serious
maldistributions of flow occurred in the dryer during steady state or transient
tests. Detailed analysis of the stability of the dryer also confirmed the high
degree of stability (see page 226).
Superheater
The fluid temperature estimates used to size the superheater are
shown below.
Gas Side Water Side
Inlet 2030°F 707 F, 1000 psia
Outlet 1768°F 1000 F, 1000 psia
214
-------
These conditions required
£ = 0.221
C . /C = 0.894
mm max
Unlike the rest of the unit, the maximum capacitance side in the superheater
is the gas side, not the water side.
In keeping with a conservative approach to the problem of
parallel flow stability, and in order to simplify the manifolding as much as
possible, a monotube arrangement was chosen for the superheater. The
surface used was the same as for the vaporizer, a 5/8 OD 321 stainless steel
monotube with the tube lateral and depth spacing equal to 1 . 5 times the tube
OD. The pertinent data for this surface are presented in the vaporizer section.
Gas side heat transfer results for this surface were
Re = 1448
h = 23.2 BTU/hr ft2°F
tube length/ row = 29.8 ft/ row
area/row = 4.88ft^/row
gas side AU/row = 113. 1 BTU/hrJF
Water side friction factor and pressure drop were evaluated using the
Darcy-Wiesbach friction factor. Acceleration pressure drop was estimated
from the assumed inlet and exit conditions. The results were:
Re = 5.58 x 105
Apf = 62. 2 psi
Ap = 64. 2 psi
f = 0.01433
Heat transfer coefficient on the water side was estimated from a
modified form of the MacAdams equation
0.023
st =
The results were
h = 727 BTU/hr ft2°F
area/row = 4.22 ft /row
water side AU/row = 3070 BTU/hr°F
215
-------
ii,ven in the superheater the gas side thermal resistance greatly dom-
inates that on the water side. This means that the superheater tube tempera-
tures will be quite close to the steam outlet temperature. A pure counterflow
unit would probably have been feasible, although slightly less conservative,
for this unit.
With the above results for the gas and water sides, the results for
overall heat transfer were
AU/row '= 109.1 BTU/hr°F
C . =718 BTU/hr°F
mm
NTU/row = 0.1519
The effectiveness for each row of tube was estimated using the same
method as described for the preheater . The minimum capacitance side is
now the water side, (the mixed side), so that the pertinent equations become
-NTU(C . /C )
r = 1 - e mm maX (17)
-77C . 1C )
6p = 1 -e (18)
The number of rows necessary to achieve the required effectiveness was then
determined from equation (8) of this report.
The results obtained for the superheater were
€ = 0.1324 £_ = 0.235
P 2
*3 = 0.318
It appeared that two tube rows would be just barely adequate for the
superheater, since the required effectiveness was 0.221.
HEAT BALANCE
As has been pointed out, sizing of the various parts of the unit required
that assumptions be made about entering and leaving gas and water side con-
ditions. In order to check these assumptions, a heat balance was made for
the whole exchanger. By assuming constant heat capacities and heat transfer
coefficients, it was possible to write this heat balance in the form of 12 linear
equations in 12 unknowns. These equations and their derivations are presented
below. The numerical subscripts refer to the numbered stations shown in
Table VII.
216
-------
Pr eheater
From the definition of effectiveness, and the calculated value of the
preheater effectiveness,
. 0.843! ,1,)
A second preheater equation if formed from a heat balance on the pre-
heater. The gas side specific heat is assumed to be 0.264 BTU/lbm. The
water inlet enthalpy at ^60°F is 127.9 BTU/lbm. The ratio of gas side to
water side flow rate is 2.225. The energy balance equation for the preheater
is then
hg = 127.9 + (T5 - T6) x 0.264 x 2.225 (20)
A third preheater equation is written to determine the state of the
water leaving the preheater. At the calculated preheater exit pressure of
1 124 psia,
h = 560. 9 BTU/lbm Ah, = 625. 9 BTU/lbm
sat ig
(hg - 560.9)
X8 V 625.9
Vaporizer
The water in the vaporizer is assumed to be in equilibrium with its
vapor at the vaporizer pressure. The average temperature of the water in the
vaporizer is therefore 558°F. The effectiveness equation for the vaporizer
then becomes
2500 - T
- = 0.249 (22)
2500 - 558
The average gas side specific heat for the vaporizer was taken as
0.308 BTU/lbm°F. The heat balance equation for the vaporizer is then
hQ = h0 + (2500 - T~) x 0.308 x 2.225 (23)
/ o <-
The outlet pressure from the vaporizer was calculated as 1103 psia.
At this pressure,
h . = 557. 8 BTU/lbm Ah, = 629. 8 BTU/lbm
sat tg
217
-------
The state equation is then
(h - 557.8)
x = —f (24)
9 629.8 v '
Dryer
The water entering the dryer is assumed to be in equilibrium with its
vapor at the dryer entrance pressure of 1103 psia. At this pressure,
T t = 557°F
sat .
The effectiveness equation is then
T - T
= 0.576 (25)
T4 - 557
The mean gas side specific heat for the dryer was taken as 0.289.
The heat balance equation for the dryer is then
h!0 = ho + (T4 - T) x 0.280 x 2.225 <26)
In calculating the outlet temperature of the dryer, the dryer pressure
was taken as 1000 psia, and an average specific heat of 0.851 BTU/lbm°F
was assumed for the superheated vapor. At 1000 psia,
h = 1191.8 BTU/lbm T . = 545°F
v sat
The state equation for the dryer is then
(h - 1191.8)
T1Q = 545 + ^n (27)
Superheater
The effectiveness equation for the superheater is
T - T
^-^ = 0.235 (28)
With Cmin/C =0.894, and the water side as the minimum capacit-
ance rate side, the heat balance equation for the superheater is
218
-------
T - T
3 4 = 0.894 (29)
Tll ~ T10
An additional heat balance equation was written in order to allow for
some heat loss from the unit to ambient. This heat loss was idealized as a
drop in gas temperature taking place between the vaporizer and the super-
heater. The temperature drop used was consistant with heat loss observations
on the test bed unit now in operation. The temperature drop used was 32°F,
so the heat loss equation is
T2 - T = 32
The solution to these equations is shown in Table VII. The values
shown in Table VII agree reasonably close with those assumed in the design.
An approach similar to that presented above is used to investigate
off-design performance of the unit in the next section.
PART LOAD PERFORMANCE
Part load performance of the single flow path steam generator has
been estimated at 5 percent of the rated air flow and 2500°F flame tempera-
ture. The results of the part load performance estimation are shown in
Table X.
It is evident that as load is reduced, a larger proportion of the total
heat transfer takes place in the vaporizer tubes. This causes the quality at
exit from the vaporizer to increase as load is reduced. The ratio of water
flow to gas flow remains nearly constant as load is reduced. This is because
the effectiveness of the unit is quite high, even at full load. Reduction in load
therefore produces only a very slight increase in effectiveness, hence only a
very slight increase in the ratio of water flow to gas flow.
In general, the method used to estimate part load performance was
to assume that the gas side was the dominating heat transfer resistance. If
this is the case, and if the gas side is the minimum capitance rate side, as
is usually the case for the unit being considered here, then NTU is proport-
ional to the Stanton number, which in turn can usually be considered to be
inversely proportional to Reynolds number to some power, depending on the
flow regime and heat transfer matrix geometry. In equation form,
NTU oc St « Re~x oc W~x (30)
219
-------
TABLE X
CONDITIONS AT 138 Ibm/hr GAS FLOW AND 66.4 Ibm/hr WATER FLOW
D
k
©
©t
Vaporizer
©
\
i
m
l) • - '
Qi)
Superheater
©1
(5
Dryer
TvJv '
^ — —
*
(5
*D
Preheater
©
'Cf) 1
I©
Gas in 138 Ibm/hr
Station
1
2
3
4
5
6
7
8
9
10
11
Temperature (°F)
2500
1281
1249
990
693
172
160
418
545 (90. 6% qua
700
962
Pressure
Atmospheric
\
r
Atmospheric
1000 psia
lity)
1
1000 psia
For the finned tubes of the preheater, the exponent x is 0.624. The
full load NTU per row is 0.967, so at five percent of the rated gas flow,
NTU/row = 20°'624 x 0.967 = 6.27
For the three cross counterflow passes of the preheater, with an NTU of
6.27 per row, the overall preheater effectiveness is 0.978. The full load
preheater effectiveness was 0.843. As was expected, the very large increase
in NTU at part load has produced only a slight increase in effectiveness.
For the bare tubes of the vaporizer, the exponent x is 0.409. The
full load NTU is 0.287, so at five percent gas flow,
NTU = 20°'409 x 0.287 = 0.977
220
-------
For a heat exchanger with the minimum to maximum capacitance rate ratio
equal to zero, this NTU produces an effectiveness of 0.624. This effective-
ness compares to an effectiveness of 0.249 at full load. It is this increase
in effectiveness with reducing load which causes a larger proportion of the
total heat transfer to occur in the vaporizer at part load.
In the dryer, it is not valid to assume that the gas side resistance
will still dominate the heat transfer process even at low load. The large
proportion of fin area/total area on the gas side, the rather low fin effective-
ness at full load, the very large ratio of gas side area to water side area,
and the fact that for the finned tubes of the dryer, the gas side exponent x
is much larger than it is on the water side, all combine to make the heat
transfer resistance on the water side increase much more rapidly than it
does on the gas side as load is reduced. At five percent of rated gas flow,
the gas side and water side resistances are of the same order of magnitude.
Heat transfer coefficient on the gas side can be estimated by an
equation very similar in form to equation (30)
h oc W(1 ~ x) (31)
where the exponent x has the same value as it does in equation (30). Equation
(31) implies that thermal transport properties undergo negligible changes as
the flow rate is changed. This is an adequate assumption for initial perfor-
mance estimates i
At full load, the gas side heat transfer coefficient was 42.0 BTU/hr
ft2°F. At five percent of the rated gas flow, with x equal to 0.624 as in the
preheater,
h = (0.05)0'376 x 42.0 = 13.62 BTU/hr ft2
With this h, the fin effectiveness increases to 0.90, which is considerably
above its full load value of 0. 76. The hA/row at five percent load then
becomes 278 BTU/hr°F. The full load gas side hA/row was 743 BTU/hr°F.
The gas side hA decreases much more slowly than the flow rate.
It was expected that at part load, dryout on the water side would occur
quite close to the entrance of the dryer tube row. Part load heat transfer
estimates in the dryer were therefore based on the assumption that the entire
length of the dryer was filled With superheated steam. At full load, the water
side heat transfer coefficient in the superheated regime was 1080. 5 BTU/hr
ft °F. For turbulent flow inside tubes, x is 0.2, so at five percent flow,
h = (0.05)0'8 x 1080.5 = 98.4 BTU/hr ft2°F
221
-------
With this h, the water side hA/row becomes 332 BTU/hr0!', which is only
slightly larger than the gas side hA/row. The overall hA/row for the dryer
then becomes 151.3 BTU/hr°F, giving an NTU for the dryer of 3.93. Since
at part load the dryer operates almost entirely in the superheated regime,
8 much higher than it is at full load. After a few iterations, the
final value of C -n/C used was 0.844, which gave an effectiveness of
.. iii.Ji.il TTicLjt
0.667.
NTU in the superheater was estimated by the same method used for
the preheater and vaporizer, except that allowance must be made for the
fact that, in the superheater, while the dominant thermal resistance is on
the gas side, the minimum capacitance side is the water side. The exponent
x for the superheater tubes is 0.409, as it is for the vaporizer. The full load
NTU/row for the superheater was 0. 1519. The gas flow is at five percent of
rated flow, but the water flow is 5. 53 percent of the rated flow. The NTU/
row therefore becomes
NTU/row = 200l4°9 x 0. 1519 x (0.05/0.0553) = 0.468
With this NTU/row, the effectiveness for the superheater becomes 0.477.
With effectiveness for each of the components in the exchanger esti-
mated, linearized effectiveness, state, and energy balance equations were
written for each of the components, just as was done for the full load per-
formance estimation, with the results shown in Table X.
DRYER HEAT TRANSFER
The location of the dryout point in the dryer at full load was
estimated, and water side heat transfer coefficients were estimated
for the dryer both at the dryout point and in the superheated region.
The method presented by Tippets (Ref. 10) was used to estimate the
location of the dryout point. The results were:
Zcrit = 6'44 ft/(L = 19'5 ft " total)
x = 0.745
5 2
crit - x 10 BTU/hr ft
where Zcr^ is the flow path distance into the tube at •which dryout occurs, x
is the quality at the dryout point, and qcrit ^s ^ne critical heat flux.
The heat transfer coefficient at the dryout point was estimated from
the correlation of Bishop, Sandberg, and Tong (Ref. 9). Their correlation is
222
-------
0.0193GC f(p/pj-(p ).'
h = - Pf G b - Ll _ (32)
_ 0.2 . 0.068
(Re) (V/V)
They recommended using a homogeneous flow model for the estimation of the
ratio (P/PV.)- For homogeneous flow,
P x((v /v )-!)+!
-G- = . G L (33)
In equation (32) the transport properties, such as viscosity, thermal con-
ductivity, specific heat, and Prandtl number, are evaluated for the gas
phase. Mass flux is evaluated for the entire flow, both liquid and gas. The
input quantities and results for equations (32) and (33) were:
C = 1. 162 BTU/lbm°F k = 0. 031 BTU/hr ft'F
P
\i = 0.0481 Ibm/hr ft Re = 5.76xl05
Pr = 1.803 PG/Pb = °-757
V-,/VT =20.6 h = 1224 BTU/hr ft2'F
Lr -Li
It appears that dryout occurs at a sufficiently high quality and at sufficiently
high mass fluxes to prevent any drastic overheating of the dryer tube wall.
The heat transfer coefficient in the superheated regime of the dryer
was estimated using a modified form of the MacAdams equation.
0.023 GC
h = —o~i—*~27a (34)
(Re)f°-2(Pr)f2/3
The input quantities and results for equation (34) were
C = 0.851 BTU/lbm°F P = 1.486
p r
H = 0.0521 Ibm/hr ft h = 1080 BTU/hr ft2°F
k = 0.0298 BTU/hr ft°F
223
-------
It is interesting to see that the average heat transfer coefficient in the super-
heated regime is slightly lower than that at the dryout point. This is because
of the high quality at dryout, and also because the non-ideality of the vapor
phase at saturation at the dryer pressure produces quite high specific heat
and Prandtl number at the dryout point.
With the heat transfer coefficients on the water side known, the
overall dryer heat transfer was re-established. The results were:
Gas Side AU/row = 743 BTU/hr°F
Overall AU/row = 659 BTU/hr°F
NTU/row = 0.857
C • /C = 0.222
mm' max
e =0.540
The € computed with the water side thermal resistance factored in was only
very slightly less than the 0.576 estimated without it.
BURNOUT
The high quality calculated at exit from the vaporizer at five percent
gas flow indicated that as load is reduced, a point must be reached where
dryout occurs in the vaporizer tubes rather than the dryer. In such a case,
tube wall temperatures in the vaporizer are subject to two conflicting ten-
dencies. The reduction in water side heat transfer coefficient as dryout occurs
tends to increase the tube wall temperature. On the other hand, especially
in the second row of the vaporizer, the gas temperature seen by the tubes
becomes lower as load is reduced, tending to reduce the tube wall tempera-
ture. If dryout can be delayed to a sufficiently high quality, so that post-
dryout heat transfer coefficients are high, and if dryout in the vaporizer can
be delayed until load is so low that gas temperatures seen by the vaporizer
tubes are low, then dryout in the vaporizer will not produce excessive tube
wall temperatures.
The possibility of excessive vaporizer tube wall temperatures at low
load was checked by investigating the probability and consequences of dryout
in the vaporizer tubes at five percent of rated gas flow. It was found that at
this load, the water flow rate was so low that it was outside the range of all
of the correlations used for investigating burnout at full load. The Westing-
house APD correlation for example, predicted a dryout quality in excess of
100 percent. The Baker plot shows that the water flow rate is so low that
even at the vaporizer exit, the flow is probably in the stratified flow or
224
-------
possibly the wave flow regime, rather than the annular flow regime. In any
case, it appears that at this low flow rate, dryout is likely to be delayed to a
very high quality.
In the full load dryout investigations, it was found that when dryout
occurs at high quality, post dryout heat transfer coefficients could be fairly
well estimated by simply using the modified MacAdams pipe flow equation
used to estimate heat transfer coefficient in the superheated regime. This
approach was used for the vaporizer tubes, using thermal transport proper-
ties for saturated steam at 1000 psia. Gas side heat transfer coefficient was
estimated using the same approach as was used for the dryer tubes, with
exponent x equal to 0.409 for the bare tubes of the vaporizer. The results
for heat transfer coefficient were 90.4 BTU/hr ft^"F on the water side, and
4. 12 BTU/hr ft2°F on the gas side.
At five percent of rated gas flow, the gas temperature at exit from the
first row of vaporizer tubes is already down to 1744°F. An effective gas
source temperature for any point along a tube row can be determined from
the following equation, which is derived by considering an elemental slice
of tube as a miniature heat exchanger with the gas side capacitance rate
negligibly small compared to the water side capacitance rate.
-NTU C . /C \
T • = T + i-§i !1L_\ '_ (35)
source w • NTU
From equation (35) the effective source temperature for the second row of
vaporizer tubes is only 1494°F.
Since gas side and water side heat transfer areas are nearly equal,
the tube wall temperature in the vaporizer was estimated from
TW = hwTw + hgTsource <36>
It was found that even if dryout did occur in the second row of the vaporizer,
the maximum tube wall temperature would be only 586°F.
From the above investigation the following conclusions were drawn.
1. At some part load condition, probably less than five percent load,
the dryout zone will move into the vaporizer tubes.
2. When this occurs, water side heat transfer coefficients in the
dried out zone will be high enough, and effective source tempera-
tures will be low enough, so that excessive tube wall temperatures
will not result.
225
-------
AIR SIDE PRESSURE DROP
Air side pressure drop at full load has been estimated, using a slightly
mofified form of Kays and London (Ref. 5) equation (24b) presented on page
21 of their book.
G2v
i
2
A . 4fL
v
m
DH v7
i
(37)
gc
The results of the pressure drop analysis are summarized below:
Component AP (in. H^O)
Vaporizer 0.151
Superheater 0.138
Dryer 0.387
Preheater 0.642
Total 1.318
The pressure drop is quite low, considering the high effectiveness of the unit.
PARALLEL FLOW PASSAGE STABILITY
Theory
A computer program for preducting the hydrodynamic stability of
flow in parallel passages was written. The program has been used to
investigate the hydrodynamic stability of the two parallel flow tubes in the
dryer section of the steam generator. It was found that these passages
should be quite stable, with a damping coefficient in excess of one.
The program uses the method of Quandt (Ref. 4) which is essentially
a perturbation analysis used to find the transfer function for the perturbation
in the water rate as a response to an input perturbation in the heating rate.
Quandt's result for the transfer function is:
226
-------
(4W./W.)
-H(Ds + Es +C)
s [£(A+D)- HD] + s[Q'(A + D) + 0(B + E) - HE]
+ [a(B + E) - HC]
(38)
where
W = water rate, or its La Place transform
g = average surface heat flux for the entire channel, or its
La Place transform
H = steady state change in coolant enthalpy as it flows through
the channel
A, B, C, D
E, a, ft
parameters which depend on the unperturbed conditions in
the channel. One of the main functions of the program is
to evaluate these parameters
i = at the inlet to the heated passage
Each parallel flow path is idealized as shown below
(Constant
Pressure)
Heated Length
INLET
EXIT
(Constant
Pressure)
The heated passage is connected at both ends to manifolds which are main-
tained at constant pressure. There are unheated inlet and exit lengths
connecting the heated passages with the manifolds. Quandt indicates that
fluid friction, entrance losses to the heated channel, and mass inertia of
the fluid are important terms in the entrance length pressure drop. However,
exit losses from the heated channel are important in the exit length pressure
drop. The exit pressure drop is a function of exit losses only because in
many cases the flow either discharges directly into a plenum or into a large
227
-------
diameter pipe which later joins the effective exit plenum. Exit friction,
elevation, and time acceleration losses are not included since these are
usually very small and the time lags associated with large steam volumes
are usually much longer than the transients which are of interest in the
stability analysis.
Although the inlet enthalpy is considered to be constant, the flow
is assumed to be perturbed in such a way that the resistance time of the
flow in the channel is very small when compared to the time scale of the
transients considered. If this is the case, then
AW = AWt + (AWe - AW^-^ (39)
Ah = Ah --' (40)
(41)
where z = distance from the inlet of the heated channel
L = length of heated channel
z1 = a weighted distance through the heated channel
e = at the exit of the heated passage
i = at the inlet of the heated passage
With the assumptions described above, it is possible to make a linearized
perturbation analysis, with the end result as shown in equation (38);
The computer program input consists of sufficient information to
completely define the flow passage geometry, the state of the fluid within
the passage, and the friction loss pressure gradient within the passage.
This information consists of specific volume as a function of enthalpy, in
the form of a table consisting of two corresponding vectors [(vl» ni)>
(v£, 1*2), (v.,, h^). . . . (vm, hm)] , enthalpy as a function of distance through
the heated channel in the form of two corresponding vectors, loss factors
at inlet and exit, channel areas, friction loss factors for the heated channel,
areas for the inlet, main, and exit channels, lengths for the inlet and the
heated channels, and the steady state flow rate. The program incorporates
an interpolation subroutine which allows functional values to be determined
given a value of the independent variable and its functional dependence
expressed in the form of two corresponding vectors. For instance, for a
value of v between v^ and v, above, the subroutine will interpolate the value
of h which is between t and h.
228
-------
The program assumes that in the liquid and vapor phases, friction
pressure drop can be described by the usual Darcy-Weisbach relation, while
in the two phase regime it can be described by the method described in
Appendix VI. The function f, of Appendix VI is input as a function of quality
in the form of two corresponding vectors.
In the liquid and vapor phases, the interpolation of specific volume
as a function of enthalpy is quite straightforward. In the two phase regime,
Quandt recommends the following "fog flow" model for calculating specific
volume and the partial derivative of density with respect to enthalpy.
where
where
v = VL f XVLG
v = specific volume (=l/p )
x = quality
(42)
L
LG
liquid phase
change associated with change from liquid to vapor phase
R/ 1 -R \
x- I 1 - •«.„)
P =
VG
(43)
P = average static density
R,-, = void fraction
O
G = gas phase
_ _ -2
(44)
The void fraction, R , is determined from the method suggested by Yamazaki
and Shiba (Ref. 11) where
(45)
(46)
•where
X = liquid-vapor volume flow rate ratio
229
-------
The program itself is primarily a straightforward evaluation of the
parameters A, B, C, D, E, Of, and ft , as shown in Equation (38) and in
Quandt. These parameters in turn depend on the channel geometry and loss
factors and on six integrals of fluid properties and pressure gradients, as
shown below:
D =..L/2gcAf
(47)
A = D + L./gc A
fi
(48)
B,!^(^U-X-1+K.
M s Ke
fi'
+ M * !
(49)
(50)
C = -
(W/A) M
f; M /dv''
.I + I
2 3
(51)
Wv M
n
c f
(52)
W
(53)
-I.
(54)
L f
L _2_( L
aw \ az >
-- dz
(55)
if T- dz
(56)
= f J^ fl
•^ ^>i T.
dz
(57)
230
-------
I B/-L _JL U-ilJi dz (58)
4 y aw \ az / L
o
T dZ (59)
o
/•L _ah z dz
./ az L
•L
6
o
where
g = acceleration due to gravity, 32.2 ft/ sec
g = gravitational mass-force-time conversion constant, 32.2 ft
Ibm/sec2 Ib
Ar = flow area of heated channel
L. = inlet channel length
A£- = flow area of inlet channel
A£ = flow area of exit channel
F = Darcy-Weisbach friction factor
K^ = entrance loss factor for heated channel
Ke = exit loss factor for heated channel
5P£/5Z = pressure gradient due to friction losses
= density
All of the integrands in I, through L. can be determined as a function of z,
the position in the channel, so the numerical evaluation of the integrals is
quite straightforward. The program uses Simpson's rule.
With the integrals I through I, and the parameters A, B, C, D, E,
Q, and /S determined, it is possible, through examination of equation (38)
to determine any of the properties of interest which are usually considered
in the study of a second order system, such as frequency response, steady
231
-------
state gain, step function response, natural frequency, and system damping..
A study of these factors and how they change with changes in the system
geometrical and thermodynamic inputs is often much more informative than
any number of particular solutions could be. The program determines steady
state gain, undamped natural frequency, and the system damping coefficient.
From SL knowledge of La Place transforms, the transfer function can
be written in the generalized form as:
( A Wi/Wi ) K
<§ ) S2 + 2 /7 w S + a> 2
Rearranging Equation (38) into this form gives:
+ D)-HD (61)
a(A +D) +^(B + E) - HE
(62)
2y[a(B + E) - HC] [£(A + D) - HD]
G ~ a(B + E) - CH (63)
where
w = undamped natural frequency
77 = damping factor
G = steady state gain
I
Chugging
The phenomenon called "chugging" is associated with a diverging
oscillatory response, that is, with the damping coefficient in the range
-1 < *7 < 0 (64)
The conditions for which this can occur can be determined from a study of the
denominator of equation (38). Beta is always positive for a heated channel,
and A and D must always be positive. The first term in the denominator of
equation (38) must therefore be positive. The "chugging" stability boundary
is therefore defined by the conditions causing the bracket multiplying s in the
denominator of equation (38) to become negative.
The second bracket of Equation (39) may be rearranged to the form
a-(A + D) +/9B + E(£- H) (65)
232
-------
Q is always positive, so the first term is always positive. An examination of
equations (49) and (55) shows that high inlet losses tend to make ft more
positive, thus stabilizing the flow. Friction pressure losses also stabilize
the flow. The fact that inlet losses tend to stabilize the flow suggests why
orifices at the inlet to parallel flow passages are frequently mentioned as a
means of eliminating chugging. An examination of equations (52) and (58)
show that E must always be positive, and that high exit losses and high friction
pressure losses tend to drive E more positive. This is not necessarily a
stabilizing factor, however, since E multiplies the term (ft- H).
It seems that in practice, the only -way for a chugging response to
develop is for /Jto become small, so that the stabilizing effect of the ftB term
is minimized while the E()9- H) term may become negative. Equation (54)
shows that this can happen if the integral Io, which is always negative, has a
large absolute value. The absolute value of 1^ can become very large if there
is a substantial region of flow in which the quality is low, as is shown by
equations (57) and (44). This is especially true if the pressure is low, so that
VLG *s ^ar§e» as *s shown by equation (44). A large degree of subcooling
tends to accentuate all the above effects, because it increases the value of
the weighting multiplier z'/L in equation (58) for regions of low quality. All
of the above trends have been observed in operation of the present test bed
boiler, as has been discussed in Section 6 of this report.
As seen from the above discussion, the best way to avoid chugging is
to ensure that there is some quality at entrance to the parallel flow passages,
If this is not practical, the flow can always be stabilized if the inlet loss
factor is made sufficiently high.
The dryer tubes were used as a test case for the stability program.
The approach taken was to design for very high quality at entrance to the
parallel flow passages of the dryer, as shown in Table XI. The stability
results from the computer run are shown in Table XI.
As shown, the value of rj is greater than one, that is, the system is
critically overdamped. For such a system, dynamic response to any
oscillatory input over the entire frequency range is quite poor. Calculations
indicated that the dryer tubes in the improved test bed unit would not exhibit
any tendency toward chugging.
The reason for the high damping coefficient in the dryer tubes is
evident from a study of Table XI. The high value of the quality at inlet
to the dryer has kept the absolute value of the integral I_ quite small, thus
allowing ft to become large enough so that the damping coefficient becomes
strongly positive.
233
-------
TABLE XI
COMPUTER STABILITY RESULTS - DRYER TUBES
L = -3.902 x 104 Ib secAbm ft2
I2 = -11. 579 Ib lbm/ft2 BTU
I3 = -0.12847 Ibm2/ft2 BTU
I4 = -3.917 x 104 Ib secAbm ft2
I5 = 22.17 lbm/ft2
Ig = 195.63 BTUAbm
A = D = 507. 2 Ib sec2/lbm ft2
B = 3.134 x 104 Ib secAbm ft2
C » 1.8854 x 105 IbAbm ft2
E = 5. 662 x 104 Ib secAbm ft2
a = 2173 BTUAbm sec
ft = 368.2 BTUAbm
-------
APPENDIX IV
235
-------
BURNOUT
The vaporizer is ordinarily required to produce dry working fluid at consider-
able superheat. This means that at some point in the unit the quality will become so
high that there will be a transition from nucleate to film boiling, with its character-
istic drastic reduction in fluid side heat transfer coefficient. Flame temperature in
the combustor is ordinarily above the melting point of most high temperature alloys.
It is obvious that the location of the point of departure from nucleate boiling must be
accurately predicted and carefully controlled in order to prevent disastrous failure
of the tubes.
Design approach used here in the high efficiency steam generator is
to place the vaporizer closest to the combustor, using it to substantially
cool the combustion gases and thus protect the superheater tubes. Burnout
in the vaporizer section is prevented by ensuring that the vapor leaving the
vaporizer is quite moist over the entire 40 to 1 turndown ratio. The transi-
tion from nucleate to film boiling takes place in the superheater coils, in
a region where gas temperatures are quite low.
The departure from nucleate boiling has been estimated by a combination of
four different methods.
The first of these was a simple estimation of the changes in types of two
phase flow in the vaporizer. According to Tong (Ref. 2), page 147, the departure
from nucleate boiling is associated with a 'drying out" of the liquid film on the wall.
A transition from annular to dispersed flow as the quality increases is therefore likely
to cause burnout. Flow rate per unit area should be chosen so that the flow is annular
throughout the vaporizer, if at all possible.
The type of two phase flow existing was determined by the method of Baker
(Ref. 12). A good description of the various types of flow discussed by Baker is given
on page 170 of Reference 13.
The second method was to simply estimate the vapor velocity. Bennett, et al.
(Ref. 7) found that steam velocities in excess of 50 to 60 ft/sec tended to create suffic-
ient splash and wave action to literally blow the liquid film off the wall. Mean vapor
velocity is estimated from:
237
-------
LrXV
Vv =
Where G is the total mass flow rate per unit area, x is the quality, Vv is the specific
volume of the vapor phase, and a is the void fraction. The void fraction was estimated
from equation (4) of Yamazaki and Shiba (Ref.ll),
1
"= 2
(2)
where Rv is the volumetric flow ratio, given by
'' '
The third method was the Westinghouse Atomic Power Division Correlation
for critical enthalpy rise presented by Tong, et al. (Ref.14). If the water is assumed
to enter the vaporizer as saturated liquid, this correlation gives the critical quality as:
(4)
n -17De -1.5G/106
x . = (0.825 + 2.3e )e
crit
n ,, -0.0048L/De 1.12 n r/io
-0.41e - 7 + 0.548
(v /v )
v L
Where De is the tube diameter in inches L is the tube heated length, and G is in
(\
units of Ibm/hr ft .
The fourth method was correlation presented by Tippets (Ref.10). This
method has a solid theoretical basis and is probably the most reliable of the four.
When Tippets' equations (35a) and (35b) are combined with information about the two
phase friction factor presented on page 81 of Tong (Ref. 2) and in Yamazaki and
Shiba (Ref.ll), there results
\7/4
238
-------
0.667
-------
APPENDIX V
241
-------
A MODIFIED VERSION OF THE TIPPETS BURNOUT CORRELATION
The original correlation presented by Tippets is
qc = ^j— (1-1)
a PL (1+ PL/PV)
with ^ - ^—r- (1-2)
PL
\ i-V PV
and « - //n v : (1-3)
The two phase flow friction multiplier, 0-ppF' and the two
factor, fw, are related to the two phase pressure drop by
'dp\ "TPF "F
dL/
(1-4)
TPF * "L "
Tippets determined the constants in his correlation by using the method of
Martinelli and Nelson (Ref. 8 ). Their equation is
1.75 ^2
It Vl t/1 ft C\
\ "^7 ^^ \ 7
/TPF \^/LO L
where 0y is defined by
n \ / Hn \
(1-6)
243
-------
Yamazaki and Shiba (Ref.ll) present a very useful empirical relation for
7/8
, = (1 - a) (1-7)
/dp\ G LO
But» br~ = T-T- ~7T (i-s)
Combining equations 1-4 through 1-8 gives
, , 'LO /i-x7/4
*TPF *F = ~ \T
When this is substituted into (1-2) there results
?/4
= Bi +
0.667 cr (1 + (v /v ))
- - — (1-11)
when a is in Ib/ft, G2/gcPL is in psi, and D is in inches. Equations (1-10) and (1-11)
are equations (6) and (7) of the main body of this report.
Substituting equation (1-9) into equation (1-3) yields
ft Y
1+[1 + C X
B
2' 1 -Q/
1/2
Equations (1-12) and (1-13) are equations (8) and (9) of the main body of this report.
244
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APPENDIX VI
245
-------
GENERALIZATION OF THE MARTINELLI-NELSON METHOD FOR
ESTIMATING TWO PHASE PRESSURE DROPS
Martinelli and Nelson' (Ref. 8) method for two phase pressure drop depends
on their observation that the friction pressure drop can be described by the following
equation.
(If) -(£) •'<".»
\ * /TPF \° /LO
Integrating equation (1) from inlet to exit
i, P)dZ (2)
z. z.
If the absolute pressure is nearly constant, and the heating is uniform,
x = A +K(Ze-Z.) • (3)
where x - x. .
K = -~ 4
e i
then J (Z - Z.)
dx s i
dz = ¥ = fc IT^o
6 1
Substituting equation (5) into equation (2)
V
AP
LO e i Jx P = const e i
Y
A
J(x, P) dx
P = const
247
(6)
-------
If Z. = 0 = x (7)
11
then equation (6) reduces to
'AP,
LO / . "e I P = const
j<
„ p-
T'P'R
f = - = — J(x' P)dx
The function f-, is presented in tabular form by Martinelli and Nelson (Ref. 8).
Rearranging equation (8)
r
J(x, P) dx = x f (9)
0
Now
y T y
7f(x)dx = / f(x)dx - I :
^a J 0 J 0
f (x) dx (10)
Substituting equations (9) and (10) into equation (6) gives
AP
TPF
APLO Xe-Xi
Equation (11) was used to predict two phase pressure drop in the dryer.
248
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APPENDIX VII
TEST BED STEAM GENERATOR CORE ANALYSIS
Appendix VII was prepared for Solar by Geoscience
Ltd. to provide a core analysis of the parallel flow
steam generator (test bed unit) described in Section 6,
249
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GLR-95
VAPOR GENERATOR TECHNOLOGY SUPPORT
FOR SOLAR DIVISION OF INTERNATIONAL HARVESTER
(PO 3739-32134-FO3)
C. M. Sabin
H. F. Poppendiek
G. Mouritzen
R. K. Fergin
GEOSCIENCE LTD
410 S. Cedros Avenue
Solana Beach, California 92075
251
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INTRODUCTION
All of the vapor generators presently under development or being
proposed for use in mobile Rankine cycle engines are of the forced-convection
once-through type. Such vaporizers receive subcooled or saturated liquid
at the inlet of a continuous tube, and discharge dry, superheated vapor from
the outlet. In steady state operation, the flow past each point in the vaporizer
is described by a fixed vapor quality. Typical flow regimes in such a vaporizer
can be (in order, from inlet to outlet); (1) all liquid; (2) a foamy, quasi-
homogeneous mixture of liquid and vapor; (3) separated annular flow with
liquid on the wall and vapor in the core; (4) wetted liquid rivulets or droplets
on the wall, wet or dry vapor in the core; (5) film boiling droplets; (6) fog
flow; (7) vapor superheating.
This sequence of processes does not appear in every vaporizer. In
some cases, flow regimes listed do not occur, or are replaced by others. A
number of transition processes can occur which fall in between the flow regimes
listed above.
The preheater-vaporizer-superheaters to be designed for the low emission
burner are intended to accommodate three distinctly different fluids; pure
water; Fluorinol-85, an organic-water mixture; and PID Fluorocarbon*, a pure
organic. The first two of these liquids are to be vaporized and superheated,
while the third will be exited from the heat exchanger in a super-critical con-
dition, so that no change of phase takes place.
The two fluids containing organics are, of course, restricted in their
temperature, so that careful control of the interior -wall temperature and heat
flux distribution throughout the heat exchanger must be maintained to prevent
fluid damage.
The heat exchanger units are subject to a number of design constraints
other than those imposed by fluid properties. There are restrictions on total
heat exchanger volume, and on the shape of this volume, on pressure drop,
and on weight.
In order to originate the most optimum vapor generator designs, and in
order to verify predicted performance in the laboratory, certain support
studies must be performed. Those being investigated are briefly summarized
below:
1. Test and prototype instrumentation for vaporizer temperature and
pressure control.
2. Flow changes during power level transients.
*PID was the original working fluid in the Aerojet system. This was replaced
by AEF-78 at a later date.
-------
3. Radiation transfer between combustor, vaporizer and gases.
4. The effects of nonuniform gas temperature distribution on the
vaporizer performance.
5. Performance of the vaporizer at reduced power levels.
6. Influence of duct geometry and possible inserts on working fluid heat
transfer.
7. Prevention of vaporizer tube overheating by temperature monitoring.
SUMMARY
Geoscience "s support role in the subject program consists of two study
areas, (1) vapor generator design, and (2) heat transfer and fluid flow support
for the vapor generator-combustor system. The vapor generator design work
consists of parameter investigations of the relationships between the fluid
properties and the requirements imposed on the vaporizer by other components
of the power system. Three different working fluid types and a wide power
level range are specified. The support studies are principally concerned with
the coupling between the combustor and vapor generator; questions involving
instrumentation, power transients, radiation transfer, nonuniform gas tem-
perature distribution, vaporizer performance variation with power level, tube
geometry effects, and working fluid overheating are being considered.
During the last quarter, Geoscience's work related to five different tasks:
(1) the design of the test bed vapor generator to be used with water as a working
fluid; (2) the prediction of part load performance of the test bed water vapor
generator, (3) the parametric design study of a Fluorinol-85 vapor generator;
(4) the analysis of peripheral heat flux in vaporizer tubes; and (5) radiant heat
exchange analyses. The results of these studies are presented in this report.
DESIGN STUDIES
Test Bed Water Vapor Generator Design
A test bed water vapor generator has been designed for use with the
Solar combustor. The conditions and constraints to which this design was
subjected in order to meet the objectives of Solar's combustor development
program are presented in Tables I and II.
Because of the lack of space in which to install flow transition sections,
the vapor generator cross-section should match the size and shape of the
253
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TABLE I
OPERATING CONDITIONS
Combustion Side:
Air flow
Air-fuel ratio
Combustor outlet temperature
Heat release In combustor
Water Side:
Inlet temperature
Outlet temperature
Outlet pressure
1.0 Ib/sec
26 to 1
2500 "F + 250 "F
2.5 x 10 Btu/hr
250 °F
1000 °F
1000 psia
TABLE II
WATER TEST BED VAPOR GENERATOR DESIGN CONSTRAINTS
Physical Maximums:
Core diameter
Core length
Tubing weight
Tube wall tempcraUu e
Water side pressure drop
Air side pressure drop
Other Constraints:
Plain tubing without extended fins
Essential!;/ :'••.?. K;;.71. air sH-= pressure drop and flow distribution as an
optimum system
Thermal efficiency 80 percent or greater baaed on fuel HHV
21.5 inches
Unspecified, preferably less than 8 inches
110 Ibs
1200°F
250 psig
3 inches of H O
254
-------
combustor outlet as closely as possible and, since the combustor outlet is
circular, this cross section was chosen for the heat transfer matrix. The
vapor generator diameter was enlarged from the 18-inch diameter of the
combustor to the maximum allowable 21.5 inches, however, so that a short
adaptor will be required. The gas side pressure drop is a strong function of
frontal area and the pressure drop requirement could not be met with an 18-
inch diameter matrix. A matrix made up of a stack of flat-wound spirals
was chosen, with the combustion products flow parallel to the spiral axis.
The water-steam side heat transfer conductances in a unit of this type
are relatively high, and the controlling heat transfer resistance is on the gas
side.
Plain, unfinned tubing was a necessary choice for this heat exchanger
in order to avoid the delays associated with special procurements and, since
the pressure drop associated with a given heat transfer rate is usually higher
for bare tube banks than for-extended surfaces, a careful consideration of the
heat exchange matrix geometry must be made.
There are a number of criteria by which a candidate heat exchanger sur-
face configuration can be judged for a particular service. For the present
purpose, where volume and pressure drop constraints are both important, the
configuration must, when compared to others, have a high heat transfer con-
ductance, a high heat transfer area per unit of volume, and relatively low fluid
friction. A figure of merit for comparison of heat transfer configurations is the
ratio of friction factor to Stanton number. This ratio, which related the
momentum transfer to the heat transfer for a given heat transfer rate. The
lower is this ratio, the more effective is the configuration in utilizing momentum
losses to enhance heat exchange.
Data for several staggered tube arrays in cross flow are presented in
Table III , based on data taken from Reference 17. It may be seen that
configurations 2 and 3 are the most compact, ana that number 2 has the highest
conductance and lowest friction factor. The friction factor-Coiburn modulus
ratio is by far the lowest of the six configurations listed. Configuration 2
appears to be the most attractive for the present purpose.
The tube diameter has a strong effect upon the heat transfer area per
unit of volume, and in the heat transfer conductance. For the configuration 2
of Table III , which has tubing on a 1.25 diameter spacing in both the trans-
verse and longitudinal direction, calculations have been made for the change
in these parameters as functions of tube size. These data are presented in
Table IV. Since a circular cross Section arrangement adapts most easily
to the combustor, a staggered tube matrix made up of flat spiral coils is
the geometry of choice. Therefore, the length of tubing in a flat coil of each
.tubing diameter is also shown in Table IV. It is clear that from considerations
255
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TABLE III
SOME CHARACTERISTICS OF PLAIN TUBE ARRAYS
OF STAGGERED TUBES
(Data Taken From Selected Figures in Reference 17)
Configuration
Number
1
2
3
4
5
6
Spacing
Longitudinal Transverse
Inches Inches
.468 .563
.468 .468
. 375 . 563
.563 .563
.375 .750
.282 .938
Beat Exchange Area Per
Cubic Foot of Volume
ft2
53.6
64.4
67.1
44.8
50.3
53.6
Heat Transfer Conductance
Btu/R2-F
34.6
46.1
36.2
32.4
35.3
38.2
Friction
Factor
.080
.051
.082
.079
.140
.124
(h/Gcp)pr2/3
4.5
3.5
4.3
4.7
6.1
5.7
TABLE IV
EFFECT OF TUBE DIAMETER UPON MATRIX PARAMETERS FOR A
STAGGERED TUBE ARRAY WITH A 1.25 DIAMETER GRID
Tube Diameter
Inches
1/4
3/8
1/2
5/8
3/4
Mini'num Kiov Area
:Y;.::t;:t Arer.
•:>.?.
0.2
0.2
0.2
0.2
Hea', 1 Vansfei Area
Pfu U'-it Volume
f -'/f*3
96.6
64.4
48.3
38.6
32.2
H«:it Transfer
Cou'luctanci:
Btu/ft^ hr °F
5f.
4ii
41
38
36
Tubiug Spacing
Inches
.312
.468
.625
.781
.937
Turns in Spii-al
Coil
21.5 in. O.D.
29.6
19.8
14.8 .
11.9
9.9
Length r,f Spir .
21.5 in. O. E
Indies
1)62
775
580
4G4
387
;>f overall heat exchanger volume, the smallest possible tubes are the best
choice, since the product of heat transfer area and conductance is maximized
by the small diameters. However, the internal pressure drop also must be
considered, and to a first approximation it changes inversely as the fifth
power of the tube diameter.
The choice of one-half inch diameter tubing allows the heat exchanger
to fit into the required volume, and the water-side pressure drop can be brought
256
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within the required limit with a reasonable number of parallel passages in
the vapor phase and two-phase regions of the vapor generator. A choice of
smaller tubing would require such a large number of parallel passages, that
the flat spiral coil arrangement would have to be abandoned in favor of the
more easily manifolded rectangular cross section arrangement, with tube
sheets. A choice of larger tubing would decrease the number of flow passages
somewhat, but could not reduce the system to a single flow channel within the
desired total volume, because the amount of heat transfer area per unit
volume decreases rapidly with larger diameter tubing.
The simplest arrangement of flat coils from the point of view of flow
passage interconnections is the counterflow arrangement, in which the com-
bustion products enter the matrix from the same end the superheated steam
exits. This is, however, an intolerable geometry, since superheater wall
temperatures would exceed 1200°F by a significant margin even in steady
state.
The internal flow passage must, therefore, be modified to reduce tube
wall temperatures. A more sstisfactory geometry has the vaporizer first,
the superheater next, and the preheater last, when listed in the direction of
combustion products flow. With this arrangement, the first rows of the
matrix, which are immersed in gases near 2500°F, then contain boiling
water at a relatively low temperature (~550°F) and very high internal heat
transfer conductances, so that the wall temperatures are far below 1200°F,
and the last tubing in the superheater, which contains 1000°F vapor with
relatively low heat transfer conductances (compared to boiling) is immersed
in much lower temperature combustion gases. This water flow path, which
places the superheater section between the vapor section and preheater
section, is the arrangement of choice.
Forced convection boiling systems, which receive saturated liquid at one
end of a passage and discharge dry or superheated vapor at the other all have
some location along the passage at which the wall is no longer covered by a
liquid film. At this location the conductances on the vaporizing surface change
from those characteristic of boiling to those characteristic of gaseous heat
transfer, a decrease which can be several orders of magnitude. For fixed
conditions the location of this point can be established with tolerable precision.
However, in a boiler which will be subjected to sudden and rather large changes
in operating power level, the location of the end of the liquid film can be expected
to move significant distances upstream or downstream, and the abrupt change in
wall temperature associated with the conductance change will also move. In
order to avoid potential problems, the expected location at which the liquid
film ends has been placed in a location of relatively low heat flux. This is
insured by two means. First, the vaporizer coils have been arranged so that
the flow from the preheater enters the coil adjacent to the combustor, and the
vaporizer is operated in cross-parallel flow. Second, the vapor exits the
257
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vaporizer and enters the superheater with a significant amount of moisture,
so that vaporization takes place in the superheater under design full load
conditions.
Based on the considerations given and detailed calculations of the
requirements for each of the three sections (preheater, vaporizer, super-
heater), the water test bed vapor generator thermal and hydrodynamic
design was established.
The physical description is given in Table V. A schematic diagram
of the flow paths is shown in Figure 1.
TABLE V
PHYSICAL CHARACTERISTICS OF THE WATER TEST BED
VAPOR GENERATOR
21.5 Inches
6.25 Inches
62ft2
Dimensions Overall:
Matrix diameter
Matrix thickness
Heat transfer area (outside)
Geometry:
Flat spiral coils arranged with axes parallel to combustion products flow
direction
10 flat coils; 2 in vaporizer, 3 in superheater, 5 In preheater
Preheater and superheater, cross-counterflow; vaporizer, cross-parallel flow
Tubing arranged on a 1.25 tube diameter staggered grid when viewed on a
radial cut through core
Tubing 1/2-inch outside diameter by 0.035-Inch wall thickness
6 parallel passages in vaporizer and superheater; one passage In preheater
Pressure Drops
Combustion products 2.6 inches of water
Water 175 psi
Water Design Flow Rate:
Weight of Tubing:
Weight of Water Hold Up:
Design Power Output:
1525 Ibs/hr
106 Ibs
30 Ibs.
2 x 106 Btu/hr
258
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VAPORIZER
SUPERHEATER
PREHEATER
C PARALLEL
PASSAGES
« PARALLEL
PASSACES
SINCLE PASSA6E
COMBUSTION
PRODUCTS
INLET
MOIST
VAPOR
TCcTs OF COILS
SUPERHEATED VAPOR
SUBCOOLED VAPOR
SATURATED LIQUID
(MANIFOLD WITH ADJUSTABLE
PRESSURE DROPPING ORIFICES )
FIGURE 1.
SCHEMATIC DIAGRAM OF WATER FLOW PATH
THROUGH VAPOR GENERATOR
This design could be improved significantly (although not necessarily
optimized), by the use of extended surfaces in the preheater. The total
length of tubing in the exchanger could be reduced by at least one third, so
that water hold-up volume, water-side pressure drop, and possible gas-side
pressure drop could be reduced. There would probably not be a great
reduction in metal weight, however, since high temperature resistant
materials would have to be used for the fin material.
Part Load Performance of the Test Bed Water Vapor Generator
The part load performance of the test bed water vapor generator has
been computed for the range of operating levels between full load and idle
(2.5 percent of full load). The calculations were performed with the ideal-
izations that neither the water inlet and outlet conditions, nor the combustion
products inlet temperature change over this range.
As the power level is decreased, a larger portion of the total heat is
transferred in the first two coils (the "vaporizer section") of the matrix. As
a result, there are significant shifts in the location of the vaporization
process over the power range. However, no operational difficulty should be
encountered because of these shifts.
The heat exchanger effectiveness of the water vaporizer as a function of
combustor load is shown in Figure 2 . As expected, the effectiveness
increases rapidly as the power level decreases. The thermal efficiency,
based on the lower heating value of the fuel and a 908F air stream into the
combustor is also shown. The efficiency is necessarily lower than the
effectiveness because the water inlet temperature is above the assumed
259
-------
IOO
90
SO
70
rrEJtliw
ENCU*
I MEAT
IHftf TIUMFEimeD W 1
AM INPMMTK CXCHANMRj
2O 4O CO
FCKEMT OF COMftUfTO* Of tWN
«O
100
FIGURE 2
TEST BED WATER VAPORIZER PERFORMANCE
AT PART LOAD
combustor inlet air temperature. The higher the water inlet temperature,
the lower the efficiency will be. At 250°F water inlet, and 90°F air inlet to
the combustor, the thermal efficiency of an infinitely large vaporizer would
be 94 percent. The two curves in Figure 2 approach 100 percent and 94
percent at zero power level. They are, however, not simple curves in the
region below 2. 5 percent of full load.
There is a strong influence between the water-stream flow path through
the heat exchanger and the performance of the superheater at part load. At the
design load, approximately 26 percent of the total heat flow is in the super-
heater, while at idle load (2.5 percent of full power), the superheater trans-
fers only 3 percent of the total heat. At that low power level, the vapor
discharged from the vaporizer is already at 930°F and, therefore, rises only
70eF in its passage through the superheater. If the same vapor generator
were arranged in cross-counterflow from preheater through superheater (an
intolerable geometry at full load with unprotected tubing), the part load
performance would be significantly better.
The temperatures of the two streams at various points in the vapor
generator are shown in Figure 3. It may be seen that the steam coming out
of the vaporizer (and into the superheater) is saturated from the full load
power level down to about 45 percent of full power. From this level on
down, the vapor superheats in the vaporizer section. At the design full load,
260
-------
FIGURE 3.
2400
3000
3
5 12 OO
a
• oo
40 O
COM* PROD INLtT TO VAP
-------
Ul
bi
8
U) tt
t IU
25
t- IM
0.
3
I I OO
IOOO
900
1
1
20 4O 6O «O
PERCENT OF COMBUSTOM OCSIGM FULL LOAD
IOO
FIGURE 4.
TEST BED WATER VAPORIZER TUBE WALL
TEMPERATURE
Base Design No. 1 was optimized to.obtain the smallest size vapor
generator which could be designed to meet the pressure drops and temperature
limits specified in the work statement. Base Design No. 1 is not a conserva-
tive design and was not recommended for construction. Rather this design
serves as a size reference for the smallest obtainable vapor generator,
from which the envelope may be expected to expant to meet practical con-
siderations for a manufactured vapor generator, or to meet new limitations
for Fluorinol-85.
In the Fluorinol-85 vapor generator extended surfaces may be employed.
However, the stringent limitation upon maximum fluid temperature, com-
'bined with the relatively poor heat transfer properties of the Fluorinol (com-
pared to water) limit the heat fluxes so that only relatively shallow fins may
be utilized, and these only in portions of the heat exchanger. The temperature
limit on the internal wall was established at 600°F for Base Design No. 1.
Because of the relative magnitudes of the gas side and organic side heat
transfer conductances, some sort of heat flux limiting for the inner wall
is required. Two means to accomplish this end are to use internal fins in the
vaporizer and superheater, which would increase the inside heat transfer
area, or to use external insulation. The internal fins used alone would
increase the internal pressure drop significantly and would be more costly to
manufacture in the spiral .coil tube design. To compensate for the increased
pressure drop, it would be necessary to use about 12 parallel passages in
262
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TABLE VI
CONDITIONS FOR FLUORINOL-85 VAPOR GENERATOR
Specified Flow Conditions
Combustion side:
Fluorinol-85 side:
Design Constraints
Physical maximums:
Same as for Test Bed Water Vapor Generator
(see Table I).
Inlet temperature
Outlet temperature
Outlet pressure
Core diameter
Core length
Fluorinol-85 pressure
drop
250 °F
550 "F
700 psia
21.5 inches
8.0 Inches
100 psi (50 psl
preferred)
Air side pressure drop 3 inches of H O
Tubing weight 110 Ibs
Tubo wull tompuruturo 600°F
Other Constraints:
Full rated output from a cold start in 15 seconds
Maximum fluid temperature shall be kept sufficiently low during transients
to prevent decomposition
each coil of the vaporizer and superheater. The external insulation requires
more coils in order to obtain the same total heat transfer, but is a mechanically-
simpler alternative. External insulation must be of fairly low thermal
conductivity, two to three orders in magnitude lower than stainless steel.
This is clear, since the insulation is added to the outside of the tubes and
must be reasonably thin, otherwise the effective diameter of the tubing
increases so much as to cause a significant decrease in the inside wall heat
transfer area per unit volume.
Design Considerations. Near the outset of the work on Fluorinol-85
Base Design No. 1, a configuration similar to that of the water vaporizer
was chosen. It was also decided to use the same combustion side heat trans-
fer configuration wherever external finning was not required.
263
-------
For a given heat exchanger configuration, pair of fluids and inlet and
outlet flow conditions, the amount of heat to be transferred in the exchanger
is fixed, and depends only on the product of heat transfer area and overall
conductance. The overall conductance, which normally may be computed in
a straightforward manner from the properties of the two fluid streams and
the wall, is used to establish the required heat transfer area. However, it
may be shown that the overall conductances computed in the normal manner
for the Fluorinol-85 vapor generator lead to tube wall temperatures which
greatly exceed the temperature limit for the fluid throughout a significant
portion of the heat exchanger. There is, thus, a heat flux limit imposed by
the fluid temperature limit. For example, consider the superheater outlet
condition. At this location the bulk vapor temperature is 550°F, and the
maximum acceptable wall temperature is 600°F, so that the maximum
acceptable temperature difference, from wall to bulk, is only 50°F. This
temperature difference, combined with the vapor side heat transfer con-
ductance, yields a maximum acceptable heat flux which is two to three times
smaller than that which is computed from the overall heat transfer conductance
for a matrix containing plain metal tubing. To avoid fluid overheating, the
local overall heat transfer conductance must be decreased to a value which
yields a heat flux less than the maximum acceptable value with the existing
local temperature difference, i.e. , the heat flux must be tailored locally in
the heat exchanger.
There are basically three approaches which may be utilized to accom-
plish this decrease in overall conductance, which are: decrease the outside
(combustion products) conductance, decrease the wall conductivity, or
increase effective organic side conductance. The decrease in external con-
ductance is probably the simplest, but requires changes in heat transfer
configuration which lead to very bulky heat exchangers, and this choice does
not appear to be a useful one for the present application. An effective increase
in the internal conductance may be accomplished by an unacceptably large
increase in pressure drop, or by increasing the internal wall heat transfer
area by use of extended surfaces. Extended surfaces on the interior lead to
the most compact heat exchangers, but have a significant effect on pressure
drop, are very difficult to obtain in materials usable for the present purpose,
and may be impossible to form into spirals as required. The third alternative
is to decrease the effective thermal conductivity of the tube wall, by the
addition of an insulating layer.
In the case of Base Design No. 1, the insulating layer approach to heat
flux tailoring was chosen. Although half-inch outside diameter tubing with an
0.020-inch wall was used to perform the calculations involving the organic
side, the outside diameter used for the combustion products side calculations
was taken to be 5/8-inch. Therefore, a one-sixteenth inch insulating layer
on the outside of the tubes was visualized.
264
-------
With these choices, an insulation conductivity which yielded the
maximum acceptable heat flux (or 600°F internal wall temperature) was
specified throughout the vaporizer and superheater sections. It was found
that insulation was not required on the vaporizer first row (next to the com-
bustor), nor on the superheater first row (farthest from the combustor).
The preheater does not require insulation, but may instead be finned
externally to increase the local heat flux (based on the bare tube surface
area). However, because of the relatively poor thermal properties of
Fluorinol-85 compared to water, only shallow fins may be used. These
increase the external surface area by somewhat over a factor of three.
This increase yields preheater wall temperatures which are well below the
600 F maximum, so that somewhat larger fins could possibly be used.
However, the design is constrained to integral numbers of spiral coils, and
a full coil could not be eliminated by increasing the fin area to bring the wall
temperature to 600°F.
As stated previously, no insulation was specified for the first vaporizer
coil for Base Design No. 1. There are several reasons why a bare coil would
probably actually be unacceptable for this application. For one example, at
part load, the location at which vapor superheating begins is expected to move
from the superheater into the vaporizer. The organic vapor could overheat
under this condition.
The Fluorinol-85 is to be vaporized at a pressure quite near the critical
point, so that the latent heat of vaporization is a relatively small portion of the
total heat added. The density change upon vaporization is also relatively
small. It is expected that the forced convection vaporization processes in
this heat exchanger will be modified enough from those usually encountered
to warrant a rather conservative vaporizing section design.
Base Design No. 1. Some of the physical characteristics of Fluorinol-
85 Base Design No. 1 are given in Table VII . A flow diagram is shown in
Figure 5. Under design full load condition the vapor exits the vaporizer
coil with considerable moisture, and vaporization is completed in the first
superheater coil. This feature is similar to that in the water vapor generator.
The preheater is required to have four parallel passages in order to
control pressure drop, while the vaporizer and superheater have six.
The size and weight of this vapor generator could be further reduced by
extensive use of internally finned tubes but this change would increase the
internal pressure drop.
265
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TABLE VII
FLUORINOL-85 BASE DESIGN NO. 1
Vaporizer
Superheater
Preheater
Pressure drops
Axial length
Computed thermal efficiency
(based on LHV)
1 coil 1/2-inch O.D. x 0.020 wall tubing
6 parallel passages.
4 coils in cross-counterflow, 6 parallel passages. First
three coils effectively 5/8-inch O. D., built up of 1/2-inch
O. D. x 0.020-inch wall tubing. Effective thermal con-
ductivity of insulation in each coil is tailored to bring inside
wall temperature to 600°F. The fourth coil is bare 1/2-inch
tubing.
Two coils in cross-counterflow, 4 parallel passages. The
tubes are 1/2-inch O. D. x 0.020 wall with shallow fins on
the outside to give an area ratio (based on the bare tube area)
of 3.45. Slightly larger fins could be used without exceeding
the wall temperature limit. However, a complete coil could
not be eliminated, and partial coils are unacceptable.
Fluorinol-85: 36 psi
Combustion products:
5.36 inches
80 percent
3.1 inches of HO
2
O.SI9 • I
FLUORINOL
44O°F
— •»,
»*»TU
>^— -•«*-—
MM
~l
***;**+•
5,o-r
roots*
J
/
f
1
:
/
4
.
OO4
!
0 57
/
/
/
t
1
t '
f
5 26
FIGURE 5.
FLOW DIAGRAM OF FLUORINOL-85 BASE DESIGN NO. 1
266
-------
A significant decrease in structural complexity might be realized by
changing from the spiral coil configuration to a conventional rectangular
cross section matrix utilizing lengths of straight tubing. Although this alter-
native would require a transition section between the combustor and vapor
generator, the tube interconnections, manifolding for parallel passages, and
the construction of the matrix would be greatly simplified. This change might
also make the use of internally finned tubing acceptable, since the tubing would
not have to be coiled.
267
-------
APPENDIX VIII
Appendix VIII was prepared for Solar by Geoscience
Ltd to provide a core analysis for the organic vapor
generators described in Section 9.
269
-------
GLR-95
VAPOR GENERATOR TECHNOLOGY SUPPORT
FOR SOLAR DIVISION OF INTERNATIONAL HARVESTER
(PO 3739-32134-FO3)
C. M. Sab in
H. F. Poppendiek
G. Mouritzen
R. K. Fergin
GEOSCIENCE LTD
410 S. Cedros Avenue
Solana Beach, California 92075
271
-------
SUPPORTING STUDIES
Methods of Controlling Wall Temperatures in Vapor Generators
Because organic fluids in both liquid and vapor phases can decompose
(thereby adding thermal resistances to vapor generator tubes), the wall
temperatures must be controlled. For fixed hot gas and working fluid
temperatures, the inner tube wall temperature (maximum working fluid
temperature) can be reduced by a number of ways; several being considered
are:
1. Increase the working fluid convective conductance
2. Decrease the gas convective conductance
3. Add a thermal resistance on the outer tube surface
4. Combinations of items 1, 2 and 3
An increase in the working fluid conductances can be best achieved by
using internal extended surfaces, new boundary layer development, turbulence
promotion and rectangular ducts. This method leads to increases in heat
flux and thus decreases in heat exchange volume.
A decrease in gas conductances can be obtained by such limited pro-
cedures as operating under in-line rather than staggered flow conditions and
by increasing the heat exchanger matrix compactness with some gas flow by-
pass. This method, as well as the subsequent one, yields decreases in heat
flux and corresponding increases in heat exchange volumes.
The approach of utilizing added thermal resistance at the outer tube
surface for the purpose of reducing inside tube wall and •working fluid
temperatures has been reviewed.
In specifying insulating layers to be added to the outside of vapor
generator tube walls, it is important to consider the relative arrangement of
the components involved. Consider a two-component system (one component
being a good thermal insulator). Three different models have been considered.
One model is based on the postulate that the constituents are positioned in
laminae parallel to the heat flow; in the second one the laminae are positioned
normal to the heat flow; the third one is based on the postulate that small
particles are uniformly distributed in a second component and that the
volume of all the particles is small compared to the total (the Eucken model).
The results for the three models are:
272
-------
Parallel Model: k = k (1 - v) +
Series Model:
Eucken Model:
1 _
TT
k
1 - v^
ki
2£
i
_ i
+ "2
k2
- 1) + v
2
-------
1.000
0.100
0010-
u
j,
0.001 •-"
0
Series Model
Parallel Model
Eucken Model
0.4 0.6
VOLUME FRACTION,
O.I
1.0
FIGURE 1. COMPARISON OF TRANSFER CONDUCTION MODELS FOR A
TWO COMPONENT INSULATOR
Nonuniform Heat Transfer Considerations in a Vapor Generator Matrix
1. Radial and Tangential Gas Temperature Variations
In order to be able to design and fabricate vapor generators that are
as compact as possible, uniform combustion gas temperature distributions
at the vapor generator inlet must be realized. Otherwise, the design must
be based on minimum gas temperatures yielding larger heat exchanger
274
-------
matrices. Nonuniform gas temperature fields can result fron nonuniform
air and gas velocity fields or nonuniform fuel droplet addition to the air flow.
A number of idealized convection models have been outlined that can be used
to bound such temperature differences in the combustion gases at the entrance
of the vapor generator. Variations in the velocity fields and the volumetric
heat sources are being considered; thermal and fluid flow boundary layer
development and jet mixing processes are involved.
2. Peripheral Tube Wall Variations
For the water test bed vapor generator, nonuniform heat rates do not
pose a problem as the tube wall temperaturs are well below the material
temperature limits and water does not present a decomposition problem.
However, for the vaporizer design for Fluorinol-85, nonuniform heat rates
require careful consideration as wall temperatures must be controlled to a
maximum of about 6008F to prevent decomposition of the Fluorinol-85. The
analyses include the review of nonuniform radiation and convective conductances
which control peripheral tube wall temperature variations.
Values for the peripheral variation of the convective conductance are
available in the literature only for in-line tube banks, single tubes or staggered
tube banks with large spacing. However, by considering the probable effects
of the closer spacing, based on photographic observations, a close approxi-
mation can be made by using stagnation values at the leading edge, based on
the maximum velocity through the minimum flow area and then using a flat
plate boundary layer convective conductance variation to slightly less than
90 degrees (except for the front row, which can be approximated by using the
stagnation value based on the on-flowing stream velocity and then using
boundary layer conductance values based upon changing freestream velocity).
Convective conductances for the rear half of the tube are approximated by
extrapolating from 90 degrees by use of profile geometry for cylinders at the
same Reynolds number (based on the maximum velocity and tube diameter).
In view of the above discussion and the geometry of the tube bank, the
maximum heat flux is believed to occur at the leading edge of the second row
of the tube bank where both the convective conductance and radiation total
shape factor are at their maximum values (although the leading edge of the
front row in the tube bank has a higher total shape factor for radiation, when
considered with the convective conductance, the flux is not as great as in the
second row). Figure 2 shows a typical convective conductance distribution
for which wall temperature profiles are to be calculated.
275
-------
! 6
1.4
I 2
0.8
0.6
0.4 .
02
I
I
45 90 135
i, ANGLE TROM LEADING EDGE
ISO
FIGURE 2. TYPICAL TUBE OUTER CONVECTIVE CONDUCTANCE
PERIPHERAL, VARIATION - TRANSITION REYNOLDS NUMBER
RANGE - STAGGERED TUBE BANK - INTERIOR TUBE
The combined effects of nonuniform convective conductance and
radiation will be used to provide design criteria for tube insulation or internal
finning in those areas where tube wall temperatures could exceed the allowable
decomposition limit for Fluorinol-85. Radiation considerations are further
discussed in the following section.
276
-------
Radiant Heat Transfer Considerations in a Combustor-Vaporizer System
A grid plate located between the combustor and the vapor generator
entrance has been used at times by Solar in some of its experimental com-
bustor work. One reason for using the grid is that it can act as a fluid flow
mixer. With this component, it may be possible to obtain more uniform
temperature and velocity fields at the vapor generator entrance. Another
reason for considering a grid is that changes in radiant heat transfer within
the combustor can be made; such changes in net radiation fluxes can be sig-
nificant. Therefore, a study of the radiation and convection heat transfer in
an idealized perforated radiation shield system was performed.
Consider first only the radiant energy transfer in the system shown
in Figure 3 (by two infinite parallel planes separated by a perforated radia-
tion shield). The two parallel planes (which represent the combustor wall
and the vaporizer surface at its entrance) are at two different uniform and
constant temperatures. The radiation shield is infinite in extent, thin, a
gray body, a diffuse reflector, a diffuse emitter, at some uniform and con-
stant temperature, and uniformly perforated. The two parallel planes are
gray bodies. Gaseous absorption in this radiation system is small and
neglected. The temperatures of the two infinite planes, all gray body
absorptivities in the system, and the amount of shield perforation are known;
the net radiant heat transferred and the shield temperature are unknown and
to be determined.
The net radiant heat transferred at the surface of either of the two
parallel planes in the radiation system under consideration is equal to the
difference between the emission power of that surface and the absorbed amount
of all radiation falling upon it. This incident radiation (irradiation) consists
of direct and reflected radiation from the various radiating surfaces of the
system. For instance, the net radiation exchange at surface 1 is expressed
as follows:
/ T. \ . _ f» r\i I f~• | /-« | /•"« | /^ \ / i \
\S/ ~ 11 ' 1- 12 1 3L 1 °" ' *
nl
where,
E = emissive power of surface 1
o; , = gray body absorptivity or emissivitv of plane 1
GI = irradiation on surface 1 as a result of all reflected
radiation originating at surface 1
G2 , = irradiation on surface 1 as a result of all direct and
reflected radiation originating at surface 2
277
-------
PERFORATED
RADIATION
SHIELD
FIGURE 3. PERFORATED RADIATION SHIELD SYSTEM
, T ,
3 J_i - 1
= irradiation on surface 1 as a result of all direct and
•*
reflected radiation originating at the left face of the
shield
i
= irradiation on surface 1 as a result of all reflected
radiation originating at the right face of the shield
The irradiation terms indicated in Equation (1) have been evaluated
considering the multiple inter-reflection processes involved (details not
shown here). These terms are functions of wall and shield temperatures,
absorptivities and reflectivities and the shield fraction. Upon making
these substitutions, Equation (1) becomes:
(1 - PjCj
4 4
- a
- P? +aP7)
- P3- P.- P3-aP2P3)
-------
where
(1 -
-a
a = shield fraction equal to the nonperforated shield area per
unit area (e. g. , a = 1 for a shield with no perforations,
and a = 0 for complete perforation (no shield))
P - 1 -. a
Similarily, the net radiation flux at plane 2 can be expressed as,
q
(A)
n
a a
- P2C3)(1 -
a(TJ
- p
- T
1 2
a(P -
- T )
3 2
(3)
where
(1 - a) P.
For the specific case where no convection exists (a vacuum system),
(q/A)n = (q/A) ; the simultaneous solution of Equation (2) and (3) yields the
1 n2 .
net radiation flux and shield temperature.
For the more general case where hot gases flow through the radiation
shield (thereby transferring convective heat to it), the sum of the net radiant
heat flow at planes 1 and 2 must be equal to the convective heat flow from the
gases to the shield, namely,
(4)
The net radiation flows q and q are defined by Equations (2) and (3) and
the convective heat flow term q , is given by the defining expression,
279
-------
. 'conv ' h As
where
h = convective conductance of shield
Ag = total shield surface area
(T - T ) = gas-shield temperature difference
8 *
Upon substituting Equations (2), (3), and (5) into Equation (4), one
equation in one unknown, T?, results; its solution yields the shield tempera-
ture, with which the net radiation and convection heat flow terms are
determined.
Calculations were made for the following representative conditions:
Combustor wall temperatures 1200°F
Vaporizer tube wall temperature 600°F
Gas temperature flowing past shield 2500°F
Solid fraction of radiation shield 0. 5
Convective conductance for radiation 53 Btu/hr ft2 °F
shield (flow through grid)
Emissivities of all radiating surfaces 0.8
For the conditions given, the following results were obtained:
t u. ,„ = 2350°F
shield
(q/A)n = 10, 600 Btu/hr ft2
(q/A)n = 22, 200 Btu/hr ft2
2
This example was representative of some recent experiments per-
formed at Solar. On the basis of temperature measurements made in the
experimental system, the predicted shield temperature of 2350°F appears to
be reasonable.
280
-------
The idealized heat transfer model described has been further general-
ized to include the effect of the radiant energy emitted by the flame within
the combustor; although the flame emissivity is low, the inclusion of this
term will give a more exact representation of convection and radiation in the
combustor-vaporizer system.
281
-------
APPENDIX VIII (Contd)
GLR-98
VAPOR GENERATOR TECHNOLOGY SUPPORT
FOR SOLAR,
DIVISION OF INTERNATIONAL HARVESTER
(PO 3739-32134-FO3)
C. M. Sabin
H. F. Poppendiek
G. Mouritzen
R. K. Fergin
GEOSCIENCE LTD
410 S. Cedros Avenue
Solana Beach, California 92075
283
-------
I. SUMMARY
Geoscience's support role in the subject program consists of two
study areas, (1) vapor generator design, and (2) heat transfer and fluid flow
support for the vapor generator-combustor system. The vapor generator
design work involves an investigation of the relationships between the fluid
properties and the requirements imposed on the vaporizer by other components
of the power system.
\
During the last quarter, Geoscience performed the following tasks:
(1) Fluorinol-85 vapor generator design studies; (2) transient analyses of the
water vapor generator; (3) radiant heat transfer analysis for the vapor gen-
erator-combustor system; (4) studies involving nonuniform heat transfer in
vapor generator matrices; (5) investigation of methods for controlling excessive
wall temperatures in vapor generators; and (6) test instrumentation reviews.
II. INTRODUCTION
The vapor generators presently under consideration for use in mobile
Rankine-cycle engines are of the forced-convection, once-through type. In
such generators, a working fluid in the subcooled or saturated state flows
into the generator and is discharged in a dry, superheated vapor state at the
outlet. In steady state operation, the flow past each point in the vaporizer is
described by a fixed vapor quality. Typical flow regimes in such a vaporizer
can be (in order, from inlet to outlet); (1) all liquid; (2) a foamy, quasi-
homogeneous mixture of liquid and vapor; (3) separated annular flow with
liquid on the wall and vapor in the core; (4) wetted liquid rivulets or droplets
on the wall, wet or dry vapor in the core; (5) film boiling droplets; (6) fog flow;
(7) vapor superheating. This sequence of processes does not appear in every
vaporizer. In some cases, flow regimes listed do not occur, or are replaced
by others. A number of transition processes can occur which fall in between
the flow regimes listed above.
The preheater-vaporizer-superheater systems being designed for low
emission combustors are to service three different Rankine cycle systems.
One utilizes water, a second Fluorinol-85, and the third uses AEF-78, a
pure organic. The first two liquids are to be vaporized and superheated, -while
the third will be exited from the heat exchanger in a super-critical conditions,
so that no change of phase takes place.
The two organic fluids are, of course, restricted in their temperature,
so that careful control of the interior wall temperature and heat flux distribution
throughout the heat exchanger must be maintained to prevent fluid damage.
The heat exchanger units are subject to a number of design constraints other
than those imposed by fluid properties. There are restrictions on total heat
exchanger volume, and on the shape of this volume, on pressure drop, and on
weight.
284
-------
A number of heat transfer and fluid flow studies are being performed
for the purpose of supporting the design studies noted above. Examples of
the main tasks being undertaken are given below:
1. Test and prototype instrumentation for vaporizer temperature and
pressure control.
2. Flow changes during power level transients.
3. Radiation transfer between combustor, vaporizer and gases.
4. The effects of nonuniform gas temperature distribution on the
vaporizer performance.
5. Performance of the vaporizer at reduced power levels.
6. Influence of duct geometry and possible inserts on working fluid
heat transfer.
7. Prevention of vaporizer tube overheating by temperature monitoring.
III. DESIGN STUDIES
A. Fluorinol-85 Vapor Generator Design Considerations
1. Effects of Lowering the Maximum Fluid Temperature Limit on
Flourinol Base Design No. 1
The size and characteristics of Fluorinol Base Design No. 1 were
established with 600°F as the maximum temperature limit for the working
fluid. However, new information on fluid decomposition indicates that this
temperature limit of 600°F should be considered as a peak to be reached
only during transients, and that the steady state limit should be reduced to
575°F. The desired superheated vapor outlet temperature is 550°F, so that
this reduction in maximum fluid temperature (or maximum wall temperature)
decreases the available temperature difference, from wall to bulk, in the
superheater by a significant amount. The final portions of the superheater,
in which the bulk fluid temperature approaches 550°F, must have nearly
double the inside surface area.
The overall conductances in Base Design No. 1 have been tailored to
bring the inside wall temperature near 600°F throughout the heat exchanger,
in order to produce the most compact design possible without internal finning.
Some discussion of the changes required to accommodate this new criterion
is required.
285
-------
The design changes which can be made to control the wall temperatures
are: (1) addition of internal fins; (2) vary the external fin sizes on the pre-
heater; (3) vary the number of passages in each coil; and (4) vary the insulation
on the superheater coils.
a. Preheater
If the internal wall temperature limit is to be lowered to 575*F,
it might be necessary to add internal fins to the preheater coils.
The number of passages would then have to be increased to com-
pensate for the additional pressure drop as a result of the internal
finning. If eight longitudinal fins (height 0.067 inches, thickness
0.020 inches) were equally spaced around the periphery of the half
inch tubing, the number of passages required would increase from
four to six. The internal finning would make it possible to main-
tain the two coil preheater arrangement with external finning, as
in Base Design No. 1, without exceeding an internal wall tempera-
ture of 575 F.
If more than eight internal fins were required, the fluid side
pressure drop would exceed the specified limit or the number of
passages required would become impractical.
With the addition of internal fins, it would be possible to increase
the external finning. However, it does not seem possible to reduce
the preheater section from two coils to one coil by this method.
From a manufacturing point of view, the combination of internal
and external finning would make it more desirable to design a
vapor generator with a square cross section.
b. Vaporizer
If the in-side wall temperature limit is lowered from 600°F to
575°F, either internal finning or outside insulation would be required
on the vaporizer coil, which was previously bare tubing. If outside
insulation is chosen, the number of vaporizer coils would have to
increase from one to two. Internal finning would require an in-
crease in the number of passages from six to ten in order to
compensate for the increased pressure drop.
c. Superheater
Internal fins could be added to the superheater coilst for the purpose
of lowering the wall temperature. It would then have been possible
to reduce the number of superheater coils from four to two for the
obsolete 600°F maximum temperature. The overall internal pressure
drop could also have been maintained below 50 psig. No external
286
-------
insulation would be necessary with internal finning if the wall tem-
perature limit is 600°F. External insulation would be necessary
in addition to the internal finning if the wall temperature limit is
575°F, unless the total pressure drop could be increased to, say,
100 psig. In that case, sixteen diagonal fins across the entire
tube cross section would lower the temperature to nearly 575°F.
The pressure drop in two superheater coils would be 68 psi using
six passages in each coil.
Base Design No. 1 has not been analyzed for part load operations
or overload operation. The latter would result in excessive wall
temperatures. It is possible that certain part load operations
could also result in excessive temperatures because the liquid-
vapor distribution would be different from the design point opera-
tion. However, since other changes in the operating characteristics
are sure to take place before the Flourinol design is frozen, it is
felt that these changes should be established before further detailed
analysis is performed.
2. Special Fluorinol-85 Vapor Generator Problems
The Fluorinol vapor generator operates at a high pressure so that
vaporization occurs near the critical point. The latent heat of vaporization is,
therefore, a relatively small portion of the total heat added, as seen in
Figure 4. The vaporizer is, therefore, the least bulky portion of the vapor
generator, in contrast to a water system. The largest heat addition occurs
in the preheater where the temperature difference is smallest. Furthermore,
the thermal conductivity of Fluorinol liquid is relatively low so that external
as well as internal finning is desirable in order to obtain a compact design.
Because the operating temperature of the superheater (550°F) is very
close to the allowable limit for the Fluorinol (575°F to 600°F), it is the most
difficult and most critical component to design.
Because of the low temperature limit, it is also necessary to control
the liquid-vapor cycle so that no superheating occurs in the vaporizer at part
load operating conditions. In Base Design No. 1, only the superheater would
provide the temperature protection required. In a practical design such
protection will have to be provided in the vaporizer as well.
A few observations may be made concerning the vaporization process
in Fluorinol. As the critical point is approached, the change in; fluid volume
with a change of phase approaches zero, and for Fluorinol vaporizing at 700
psia, the volume change is approximately five times. The familiar energetic
bubble action in nucleat boiling, which is frequently visualized, is associated
with density changes of several orders of magnitude, and one may expect this
process to be suppressed with small density change. For vaporization under
287
-------
100
200
ENTHALPY, BTU/LB
300
FIGURE 4. ENTHALPY, PRESSURE AND TEMPERATURE CHART FOR
FLUORINOL
forced convection conditions in tubes, the flow through a large portion of the
vapor quality range is annular, with the vapor flowing down the core. The
thickness of the liquid layer on the wall is inversely related to the ratio of
vapor velocity to liquid velocity, a ratio which is, in turn, related to the
density ratio. In the case of vaporization near the critical point, thick liquid
layers with small mixing due to bubble generation may very well occur, and
under such conditions heat transfer conductances may not be particularly
high. In the complete absence of bubble generation at the wall, one model of
the vaporization heat transfer process visualizes the heat as being convected
across a superheated liquid layer to vaporize liquid at the interface with the
vapor. Such a process may very well be the dominant heat transfer mechan-
ism in the vaporizer.
It is well established that vapor bubble generation next to liquid-solid
interfaces in boiling occurs at particular nucleation sites, which are surface
cavities and crevices containing inert gas or vapor. In the absence of suitable
cavities or in the absence of inert gases to render the cavities active, many
liquids can superheat without phase change far above their equilibrium satur-
ation temperature. Metallic tubing is usually adequately provided with pot-
ential nucleation cavities, because the manufacturing processes enhance
their formation. However, in sealed, carefully cleaned and evacuated systems,
these potential sites may not be active.
288
-------
The preceding considerations indicate that for a conservative design
for organic vaporization, it would be appropriate to use a relatively low heat
transfer conductance, based on an annular flow model without bubble genera-
tion.
3. Alternative Geometries
External insulation and external and internal fins will be required for
the Fluorinol vapor generator. A conventional rectangular array substituted
for the spiral coils, would simplify the installation of insulation and eliminate
the possibility of collapse of internal longitudinal fins during coiling. In
addition, a rectangular array is much easier to manifold, so that the use of
many parallel channels to control pressure drop would not present serious
difficulties. A short flow transition section would be required between the
combustor and vapor generator to adapt the change in cross section, but if
the actual flow area change -were not large, this adaptor could probably be
simple in form.
B. Transient Behavior of the Water Vapor Generator
There are two distinct transient operations of major importance in the
operations of the water vapor generator. These are; the startup process, and
varying load during operation. In the former case, the heat exchanger is
initially full of water, and the liquid must be cleared from the vaporizer and
superheater coils during the startup. In the latter case, the heat exchanger
responds to changes in combustor output, steam demand at the outlet, and
water flow at the inlet. Since these two transient operations are very difficult
in nature, they can be discussed independently.
1. Vapor Generator Startup
The idealized startup process proceeds in the following manner. With
the heat exchanger completely full of liquid at ambient air temperature, the
combustor is impulsively brought to full rated output and maintained at that
level until the superheater outlet conditions are at the design point, 1000 psia
and 1000°F. During this process, the boiler tube may be dry, filled with
stagnant or flowing fluid.
a. Dry Tube
In the case of the dry tube, the temperature history of the first
coil wall is similar to that of the voltage on the capacitor in a
series resistance-capacitor circuit impulsively subjected to a
steady voltage at time equal to zero.
289
-------
The governing equation is
fm = 1 - e^ (1)
where
T - T.
T = —^
t =
R C
g
values of the constants appropriate to the vapor generator are as
follows
Tm = metal wall temperature
T^ = initial temperature (a constant), 70°F
T = gas temperature (a constant after t = 0), 2500°F
&
t = time
C__ = metal wall heat capacity Me = 1.2 Btu/°F for the first coil
m r J pm
M = metal mass, 11 Ibs
c = metal specific heat, 0.11 Btu
pm r
R = 1/h A, the heat transfer resistance on the gas side
hg = gas side heat transfer conductance, 43 Btu/ft hr°F
A = gas side heat transfer area, 6.2 ft
The time constant RgC has the value 16.1 seconds. In the
absence of cooling, the metal temperature will go to the gas
temperature. However, at some time after combustor startup
the water flow would presumably begin. The time taken for the
wall of the first coil to reach the saturation temperature of water
at 1000 psi, 545°F, is, from Equation (1).
t = 3. 52 seconds
290
-------
The test bed water vapor generator contains six coils between the
preheater inlet and the end of the first vaporizer coil, representing
a length of 280 feet of tubing. The design liquid velocity is approx-
imately eight feet per second, so that the transit time from the
inlet of the preheater to the outlet of the first vaporizer coil is
about 35 seconds. It seems clear that a startup process with the
vapor generator dry would have to have significant operational
advantages to justify the necessary careful timing.
b. Tube Filled With Stationary Fluid
In the case of startup of the boiler containing stationary liquid, the
temperature history of the first coil wall is similar to that of the
voltage on the first capacitor in a series arrangement of resistor-
capacitor and resistor-capacitor. The temperature history of the
liquid mixed mean is similar to that of the voltage on the second
capacitor. The governing equations are
1+r2
T = 1 +
m r r.
1 +
and
=
(1 - K) -
rl - r2
(2)
+ Ke
"*
1 - K
(3)
where
T -
ig
K =
It . •»-
14-:;—
- 1 *
1 - 4K
= liquid mixed mean temperature
= 1/htfAfl heat transfer resistance on liquid side
= liquid heat capacity for entire first coil, 3 Btu/°F
= 1/hgA, the heat transfer resistance on the gas side
291
-------
Ag = heat transfer area on liquid side of first coil, 5.4 ft"
hg = liquid side heat transfer conductance
Evaluation of these solutions depends upon the establishment of a
conductance for the liquid side. For a stationary liquid, the
classical solution for the flow of heat in a cylinder subjected to a
uniform external heat flux may be used. The appropriate solution
may be found on page 203 of Reference 18, and elsewhere. The
conductance extracted from this solution is a function of time, but
may be approximated by a constant for the early part of the heat-
ing process. For the thermal properties of water and the dimen-
sions of the water vapor generator, the water side conductance
may be computed to be approximately
U0 = 128 Btu/ft2 hr F
Based on the indicated constants, Equation (2) for the vaporizer
tube wall becomes
f = 1 - 0.765 e~°'3lT - 0.235 e'3'24* (4)
The composite time constants for the two terms are 52.0 seconds
and 4. 98 seconds, respectively. The vaporizer tube wall reaches
the temperature of 545°F when Tm = 0. 196, t - 0. 263, and t = 4. 23
seconds.
Equation (3) for the mixed mean liquid temperature becomes, with
introduction of the indicated constants,
T^ = 1.0 +40 e'1-025* - 41 e-f (5)
The time constants for the two terms are 15.7 seconds and 16. 1
seconds, respectively. At t = 4.23 seconds, when the tube wall
has come to the boiling point, the liquid mixed mean temperature
is 187°F.
The idealizations upon which Equation (4) and (5) are based become
invalid after subcooled boiling begins at the wall. At this time,
the effective inside conductances and the mixing of the liquid in-
crease to fairly high values. However, the possibility of film
boiling, with a consequent decrease in the inside wall conductance
to a very low value is a distinct possibility if boiling continues
without forced convection. It, therefore, appears that the water
flow will have to be established soon after the wall reaches the
water boiling point in order to insure that film boiling does not
292
-------
occur. With this constraint, it appears unlikely that there will be
any net vapor generation possible before water flow must be started.
c. Tube Filled With Moving Fluid
These cases are more complex than the preceding two. The
analyses presented below strictly apply only to such short lengths
of tubing that the average temperature for heat transfer from the
gas may be taken as the arithmetic mean temperature between
inlet-and outlet liquid flows. They are further restricted to water
flows large enough that the liquid side heat transfer resistance is
negligible compared to that on the gas side.
Other restrictions consistent with the above qualifications are:
1. Water and metal temperatures are identical
2. Conductances are invariant
3. Gas temperature is invariant
4. Water flow in invariant
(1) Constant Inlet Water Temperature
The governing differential equation for the water outlet tempera-
ture is
A
T + ^± = 0 (6)
dt
R
T
T - WO
~rl~
The reference temperature is
T —
TR
2 hgA
Ci
(
T_l_ T
g wi
|2wc
P
C.
Ci -
T
R
The reference time is
Ct
TR ~ 2wc + h A,
P g I
293
-------
The other new symbols are
A. = gas side heat transfer area for chosen tube length
C» = heat capacity of both water and metal for chosen tube length
JL
w = water flow rate
c = water heat capacity
Twi = water inlet temperature
TWQ = water outlet temperature
The solution to Equation (6) for the appropriate boundary condition
$ = Tj, A = 0, is
A
JL -t
For conditions appropriate to the entire water vaporizer first
coil and full design water flow, the time constant is
TR = 4.59 seconds
The reference temperature is
TR = 458°F
The resulting temperature history of the water outlet is, for
an initial temperature of 70°F,
458°F - T A
-° = a'* (8)
388°F
After one time constant, 4.59 seconds, the temperature has
increased to 3l6°F.
It may be observed that in this case the time to boiling tempera-
ture cannot be computed, since the water will not reach boiling
with a 70°F inlet to the first coil. The limit temperature is
458°F.
The restriction of this case to fixed inlet temperature invalidates
the solution for times great enough so that the preheater outlet
temperature has begun to rise. For cases with significant
294
-------
variation in first vaporizer coil inlet temperature, the following
analysis is presented.
(2) For the case/of a variable water temperature of the form
Twi = X(l - e~^ (9)
similar to Equation (8), the differential equation is of the form
wo ~
(10)
dt
where the appropriate temperatures are normalized according
to the relationship
T - T
initial
T =
T - T
boil initial
•2wc 2h A
P g
Equation (10) has the solution
wo = (kl + k2> - e~ C +
where C is an arbitrary constant dependent upon the initial
conditions .
This equation has not been evaluated in detail. The essential
information with regard to the time constant has already been
obtained. The time constant for this case is identical to that for
the constant water temperature inlet case.
Although the time constants are the same, the actual water flow
temperature from the preheater is not described well by the
relationship used. The preheater gas side temperature varies
during startup. At the initial instant, while the heat exchanger
coil is at a uniform temperature, the gas temperature profile is
295
-------
easily computed, and its variation is shown in Figure 5. Also
shown in the figure is the gas temperature distribution for the
steady state full load condition.
d. Startup of the Test Bed Water Vaporizer Generator as a Whole
The transient behavior of the entire heat exchanger during startup
can be computed numerically subject to some simplifications. The
early part of this process has been worked out and the calculations
indicate that the time constants obtained from Equations (8) and (11)
are representative of the behavior of the entire heat exchanger core.
Therefore, one may expect the heat exchanger to be near the design
full load output within two to three time cons tants, or ten to fifteen
seconds. However, the design full load conditions have been com-
puted with a liquid flow into the preheater of 250°F, and since after
fifteen seconds the preheater inlet flow will probably be still near
the initial temperature of the condenser, hot well and pump
assemblies, the steam output will probably be somewhat below the
full design superheat.
The principal difficulty with this numerical calculation (and the
one which makes transient analysis by means of tables available
=! 3
. STEADY STATE
I
23456789 10
COIL NUMBER (COUNTING FROM COMBUST OR END)
II (OUTLET)
FIGURE 5.
GAS TEMPERATURE DISTRIBUTION AT FACE OF EACH COIL
AT STARTUP AND STEADY STATE FULL LOAD
296
-------
in the heat exchanger literature impossible) is caused by the com-
plex water flow path. The flow path is also responsible for the
serious problem associated with actual startup of the vapor
generator. The onset of boiling will almost certainly occur near
the outlet end in the first coil of the vaporizer section (next to the
combustor). In order to accommodate the vapor expansion, the
liquid will have to be expelled from the superheater coils. This
process could be quite violent under some circumstances.
The analyses presented for temperature history of the first coil
indicate a temperature rise in the absence of boiling, which is in
the order of 25 to 50°F per second. At the temperature level of
545°F (the design boiling point), this temperature rise rate corres-
ponds to a saturation pressure change of in the range of 200 to 500
psi per second. If the pressure in the vapor generator is not to
exceed 1200 to 1500 psi, then the superheater must be cleared in
a few seconds. Since each of the six passages is some 32 feet
long and the design water velocity is of the order of eight feet/
second, this clearing process requires some significant changes
in flow velocity.
If adequate flow accelerations cannot be attained within the pressure
limits of the vapor generator, the combustor heat release would have
to be limited to something less than full power on startup. This
change would lead to longer startup times.
2. Transient Performance Characteristics
Perturbations around the full load design point have several aspects.
At a change in load, the water flow must change as well as the combustor
firing rate in order to bring the system into equilibrium at the new demand
level.
The transient times for the two streams through the heat exchanger
core are very different. At full load, the gas stream transit time is 14 milli-
seconds. The water stream total transit time is 36.4 seconds, of which 35.0
seconds is required for the water to pass through the preheater, 0.8 seconds
for the vaporizer, and 0.6 through the superheater. The long time required
for a particle of water to flow through the vapor generator is indicative of the
time required for a thermal change in the inlet condition to make itself felt
at the steam outlet. However, flow rate accommodation times are much
shorter. Since the liquid in the preheater is substantially incompressible, a
flow rate change at the inlet to the preheater is instantaneously transmitted to
the location of the onset of boiling.
297
-------
The principal time lags associated with a change in steam demand are
probably caused by the inertia of the water in the preheater, which affects the
rate of change of flow; the flow response time of the steam vapor space con-
sidered as a long, narrow storage reservoir; and the thermal time constant
of the wall.
A simple analysis of the vaporizer wall in the first coil can be used to
illustrate the magnitude of time lags associated with energy storage in the
vaporizer metal. Consider a tube initially in steady state at the design full
load conditions, which at time t = 0 is subjected to an impulsive change in gas
temperature while all other conditions are held constant. This perturbation
is approximately that which occurs when the combustor output is suddendly
changed to a new level. The governing equation for the tube wall temperature
is:
— = f , - T (12)
dT f
where
t = t/T
/-L + X T + A- T
0 - .
s ^ el
Tf = - -- — = normalized final tube wall temperature
T = T/Tgd
T = tube wall temperature
T , = inlet gas temperature
T ^ = design inlet gas temperature
T = water saturation temperature
t = time
T = time constant =
Crn = tube wall heat capacity (for entire first coil)
w_ = gas flow rate
O
298
-------
c = gas heat capacity
re~t
U = gas side conductance based on gas side heat transfer area
o
U, = boiling side conductance
A = gas side heat transfer area for first coil
This equation is subject to the following idealizations and approxi-
mations:
1. Conductances are invariant with time
2. Gas side temperature is always greater than tube wall temperature
3. The mean gas side temperature for heat transfer may be approxi-
mated by the average of the incoming and outgoing gas tempera-
tures for the single coil
4. The saturation temperature is constant and the fluid is always
boiling
The approximate boundary condition for Equation (12) is
I = 0, T = Td
The solution is simply
T = T T
_f = e'* . (13)
Tf - Td
The time constant, T, is of particular interest. The following constants are
appropriate to the water vapor generator
wgc = 1.2 x 103 Btu/hr°F
Ug = 43 Btu/ft2 hr°F
Cm = 1.2Btu/°F
A = 6.2ft2
299
-------
Ub = 3000 Btu/£t2hr
A! =4.5
\2 = 0.0645
These yield the time constant value
T = 0. 23 seconds
It appears that the wall temperature will have accommodated itself to a new
level within one second of onset of a temperature change.
IV. SUPPORT STUDIES
A. Radiant Heat Transfer Considerations for the Combustor-Vapor Generator
Design
In the previous quarterly report, an equation set was defined which
accounts for thermal radiant heat transfer processes between the combustor
and entrance of the vapor generator. The system consisted of two planes
separated by a perforated radiation shield. Plane one of this idealized system
represented the combustor walls, plane two represented the entrance of the
vapor generator, and the perforated radiation shield represented a flow dis-
tribution grid for the system. The analysis accounted for radiant exchange
from the various surfaces identified, including multiple interreflection as
well as convective heat addition to the flow distribution grid from the hot
gases. This system has subsequently been extended to include radiant heat
transfer from the luminous flame. The luminous radiation was postulated
to be absorbed completely at the first impingement on a surface (reasonable
approximation). This extension consisted of adding a constant thermal input
term to the shield as a result of the absorption of the luminous heat transfer
from the flame. The general equation set that defines the thermal radiation,
convection and luminous radiation shield absorption for this system follows:
Iconv + lium = qn + qn
J. Ct
\ = *!<*!' T2' a» a'S' T3>
qn = f2(Tl' T2' a» a'S' T3)
£*
Icon = hAs
300
-------
where
q , convective heat transfer to shield
conv
q, , luminous heat transfer from flame
q , net thermal radiation exchange at plane 1
q , net thermal radiation exchange at plane 2
T , known temperature of plane 1
T? , known temperature of plane 2
T- , unknown temperature of the shield
T , known gas temperature
5
T, , known flame temperature
a , solid fraction of shield
a , gray body absorptivity or emissivity of radiation surface
h , convective conductance of shield
A , total shield area
s
f.r , flame emissivity
In order to calculate the total net radiant fluxes at planes 1 and 2, one
would add the net thermal flux and the absorbed luminous flux. The net thermal
fluxes would be obtained from equation set (14) and the luminous flux from
classical procedures; at plane 2, however, only that fraction of the luminous
radiation that passes through the perforated radiation shield would be involved.
In order to illustrate the effects of varying the porosity and emissivity
of the radiation shield on the net radiation flux at plane 2 (at the first row of
tubes in the vapor generator) and on shield temperature, parameter evalua-
tions of the equation set were performed neglecting the effect of flame
luminosity. Some of the results are shown in Figure 6 and 7. Note that the
net radiant flux at plane 2 (vapor generator entrance) significantly increases
as the solid fraction of the shield increases. Also note that it is possible to
reduce the net radiant flux at the vapor generator entrance by decreasing the
emissivity of the shield.
301
-------
30OO —
2500 -
20OO
CONSTANTS:
o,= 02 = a 3 = 0.8
h =53 BTU/HRFT2°F
SO ,OOO
20,000 —
10,000
0 0.5
SOLID FRACTION OF SHIELD, 1
FIGURE 6. (q/A) AND T VERSUS a
2
B. Nonuniform Heat Transfer in Vapor Generator Matrices
1. Effects of Combustion Gas Velocity and Temperature Variations
From Their Mean Values on Tube Wall Temperatures
When dealing with vapor generator systems that utilize organic
working fluids such as Fluorinol-85, it is important to be able to define
quantitatively the effect of combustion gas velocity and temperature variations
(from their mean values) on tube wall hot spots. An elementary analysis has
been performed which establishes the criteria for defining the hot spots.
Consider the thermal circuit shown in Figure 8. In this system it is
postulated that the convective conductance of the working fluid, hc, the tube
wall thickness, 6 , the tube wall thermal conductivity, kw, and the working
fluid bulk temperature,
ture,
T , are all constants. The combustion gas tempera-
Tn, and velocity, u (and, therefore, the convective conductance, h ),
&
are variable and control the wall temperature, tw«
A heat flow equation can be written for the thermal circuit in terms
of the unknown wall temperature, t ,
namely,
302.
-------
3000 -
250O
2000
1500
a = 0. 5
Tg= 25OO°F
h=53BTU/HRFT2 "f
I
0.5
EMISSIVITY OF SHIELD, a 3
30,000
- 20,000
- 10,000 i
FIGURE 7. (q/A)n AND ?3 VERSUS a
(U)
hc
tw
•Tc
w
-COMBUSTION
GASES
.TUBE WALL
-WORKING
FLUfD
1 w
FIGURE 8. THERMAL CIRCUIT DESCRIBING HEAT FLOW FROM COMBUS-
TION GASES TO THE WORKING FLUID
303
-------
A heat flow equation can be written for the thermal circuit in terms of
the unknown wall temperature, tw, namely,
T - T t _ T
8 C --=-^- (»)
J
h
—
ha * k
g w
Upon regrouping Equation (14) in a dimensionless form, one obtains
t -
w
T
go
T
C
T
C
h
(**
h
g
go
h
c
h 6
g w
k
w
+ .
+ i
h /
c
(r -
' K
i T
^ go
T
C
- T
C
(16)
where
h , the uniform or design value of h
go 8
, the uniform or design value of
O O
o
The convective conductance on the combustion gas side is a function of the
local gas velocity (dependent upon the Nusselt-Reynolds modulus function for
flow over tubes). For the systems under consideration, it can be shown that
the conductance varies as the 0.6 power of the velocity, namely;
h / 0) . 6
o
Therefore, the dimensionless wall temperature can be expressed as
t - T C T - T
_w _ c _ i _ g c
T -r = ~I - ~ r -T
g C
o
(t)
304
-------
where hgQ
°1 = h~
c
V- \
2
w c
Equation (18) has been used to predict mean wall temperature varia-
tions in tubes of a typical Fluorinol-85 superheater. The results are shown
in Figures 9 and 10 in terms of gas velocity ratio, U./UQ, and temperature
ratio, T - TC/ T - T , typical of this system.
e BO
Some experimental information available from Solar's combustor gas
temperature traverses indicate that the temperature ratio might vary from a
low value of 1.05 to a high value approaching 1.3. Clearly by good combustor
desigri it is not expected that the high value will be typical, but that the low
value can be achieved. No experimental information is available on the velocity
ratio, u/uQ.
It was thought appropriate to determine from Figure 10 how large a
wall temperature increase would result for the hypothetical condition that
both the combustion gas velocity and temperature ratios are equal to 1.2.
For this case, the wall temperature increase above the design (uniform)
value was t - t = 38°F. The results indicate that such temperature
vv ^^O
asymmetries at the hot spots could cause decomposition and deposition
problems; thus, lower gas velocity and temperature ratios should be set.
It is believed that it would be fruitful to analyze combustion processes
and secondary air addition processes for idealized flow geometries repres-
entative of the Solar combustor for the purpose of estimating the limiting
gas velocity and temperature ratios that may occur and what can be done
about reducing them. Some possible mathematical model studies that would
give such information have been outlined for future evaluation.
2. Effects of Peripheral Variations in Heat Transfer Around Vaporizer
Tubes
The matter of peripheral variations in heat transfer in vapor generator
tubes is also of interest in connection with excessive wall temperatures, if
organic working fluids are used.
The peripheral variation in heat flux is composed of convection and
radiation terms. The latter mode is only present to any practical extent in
the first or second row at the vapor generator entrance, of course.
305
-------
O.15
0.12
0. II
O.IO
0.09
WHERE:
hc= 3S3BTU/HR FT2 °F
««.-_ 0.0002 HRFT'-F
*w BTU
T^ = 440 *F
T,= 1540'F
0.8 1.0 1.2 1.4 1.6 1.8 2jO 2.2
FIGURE 9. DIMENSIONLESS TUBE WALL TEMPERATURE VERSUS
GAS VELOCITY AND TEMPERATURE RATIOS
Maximum peripheral wall temperature variations (without considering
peripheral tube wall conduction) at the vaporizer inlet of the Fluorinol-85
vapor generator were calculated. At the stagnation point of the first tube, the
ratio of the local convective heat flux of the tube is greater than 1. 5 (depending
on how large the radiation flux is as shown in Figures 6 and 7, for example).
This means that the inner tube wall temperature at the stagnation point can be
as much as 49°F greater than a mean inner tube wall temperature. Similar
calculations have been made for the superheater. In this case, the maximum
inner wall temperature excess at the stagnation point is a little higher than the
value for the vaporizer tubes, namely, 54°F. Although the radiation flux does
not exist and the gas temperature is only 1540°F (rather than 2500°F) for the
superheater, the inner wall-working fluid temperature difference is higher
in this component because of the relatively low working fluid convective con-
ductance that exists.
306
-------
-10 -
- 20 -
FIGURE 10.
TUBE WALL, TEMPERATURE VERSUS GAS VELOCITY
AND TEMPERATURE RATIOS
Geoscience has previously studied the effects of peripheral heat con-
duction in tube walls for asymmetrical heat transfer situations. An analytical
solution has been developed that can be used to predict the peripheral wall
temperature distribution when there is a step function variation in the heat
transfer conductance on one side of the wall and a uniform value on the other.
The solution also accounts for wall thickness and its thermal conductivity.
It is possible to use this work to predict how much the wall temperature hot
spots described in the first part of this section will be reduced as a result
of tube wall conduction. In addition to this method of analysis, it is also
possible to perform two-dimensional flux plots with a thermal analog appara-
tus, if necessary.
C. Methods for Controlling Excessive Wall Temperatures in Vapor Generators
There are a number of methods by which excessive wall temperatures
in vapor generators can be reduced, two of which have been studied by
Geoscience to date. They involve rather direct and relatively uncomplicated
307
-------
approaches to the problem. One method involves adding thin thermal
resistances to the outside of vapor generator tubes and the second one con-
sists of internal tube finning.
1. External Thermal Resistance
The addition of a thermal resistance to the outer tube surface has the
additional advantage that it does not increase the internal fluid pressure drop.
Further, if the resistance layer is thin, it does not significantly change gas
flow patterns or increase the heat exchanger size.
The overall heat transfer conductance, U, for the insulated tube
system is,
1
(19)
where the subscripts o, r, w, and i, refer to outside, added thermal resis-
tance, wall and inside, respectively. The other symbols have been described
previously. It is clear that the overall heat transfer coefficient can be reduced
by increasing the thickness of the added resistance in addition to decreasing
its thermal conductivity. Figure 11 illustrates this feature for typical
Fluorinol-85 superheater conditions.
CONSTANTS : h0 = 40 BTU/HR FT2 °F
0.0
0.01 0.02
0.03 0.04 0.05
6 , INCHES
0.06
0.07 0.08
FIGURE 11. THE EFFECTS OF kr AND <$r ON THE CONDUCTANCE U
308
-------
It is seen that a gas layer would provide a high thermal resistance to
the wall. In this case, the tube could be wrapped with a thin foil which is
spaced by dimples pressed into, say, one percent of the foil area. By varying
the depth of the dimples, various thermal resistances can be obtained to
satisfy the requirements in each row of tubes. For example, the material
gaps required for each row of the superheater in the system illustrated in
Figure 12 would be:
Overall Material Material
Heat Transfer . Having a Having a
Row Conductance, kr = 0.035 (gas) kr = 0.10
No. Btu/hr ft2°F 6, Inches 6, Inches
2
3
4
5
U
U
U
U
2
3
4
5
= 12
= 17
= 25
= 36
6
6
6
6
2
3
4
5
= 0.
= 0.
= 0.
= 0.
031
017
006
000
6
6
6
6
2
3
4
5
= 0.
= 0.
= 0.
= 0.
075
040
015
000
The above calculations illustrate that a gas gap would increase the
overall tube diameter less than twelve percent for a half inch tube. In com-
parison, an insulator with k = 0. 10 Btu/hr ft°F would increase the overall
tube diameter up to 30 percent.
ROW NO. I
GAS
2500°F
3740LB/MR
Cfc-r O.S29 » 1
FLUORINOL
440 °F
36 BTU
2092 °F
MR
550°F
1 I 7OOPSI4
/
' ' /
f
1 1
'/
,
^
t
CL ;
rji. /
'-^?
/
K
' N
0.576
/
t
t
1
1
f
' f
/
/
f
i
rf
(
f
•06BTU
t
t
/
f
/
_ /
/
f t
'""*/
/
A
/HR
f
t
f
i
f
' f
f
^^
/
i
t
~~
X.
1543°F I
+**^* »_
-
-
•
-
-
E
;
440°F ~
—Xi
1= 0
: i
—
^*-
| 690°F
- '~«-^-^
-
-
888 > I06BTU/HH
_ _
\_ ~
: ;
r :
_
-
:FLUORINOL
-
; 300° F
- 736PSIA
9450LB/HR
FIGURE 12. FLOW DIAGRAM OF FLUORINOL,-85 BASE DESIGN NO. 1
309
-------
In order to demonstrate the insulating effectiveness of various types of
tube insulators, a one-half inch diameter steel pipe, covered with test insul-
ation samples, was heated electrically to a glowing red color. The test
samples consisted of:
(1) A dimpled stainless steel foil,
(2) A single stainless steel foil wrapped directly on the pipe,
(3) A stainless steel foil separated from the tube by a fine gauge
screen,
(4) A stainless steel foil separated from the tube by a coarse gauge
screen,
(5) A ceramic coating.
Except for the ceramic coating, each test sample had a small hole
drilled through the foil, through which the pipe wall color-temperature could
be estimated. It is noted that the reverse direction of heat flow in these tests
in comparison to that to a vapor generator tube is unimportant because one is
only interested in the thermal insulating effect. The dimpled foil showed
significant insulating characteristics because of the air gap between it and the
pipe wall. The thermal insulating effect of the foil wrapped directly on the
wall was not very good (as expected because of the thin air gap). Both foil
samples separated by the screens demonstrated good insulating characteristics,
the one with the coarse screen being somewhat better. The ceramic coating
was not as good an insulator as the other test samples because the coating
was in direct contact with the wall. These tests verified that gas gaps are
very effective as insulators as would be expected from the previous calculations.
2. Internal Finning
Another method of reducing the inner tube wall temperatures in the
Fluorinol vapor generator is to increase the internal tube wall area by adding
fins. However, increased surface area also means additional pressure
losses, particularly in viscous fluids. In order to avoid profile pressure
losses, the fins should be longitudinal and with a cross sectional geometry
which reduces the equivalent tube diameter the least possible. Ring or spiral
fin arrangements are, therefore, undesirable. In order to minimize pressure
losses, sudden flow changes should be avoided at the tube inlets and outlets.
In designing internal finning, the pressure losses and the heat transfer
effectiveness must be optimized for each component in the vapor generator so
that allowable pressure losses are distributed most advantageously for obtaining
desired heat fluxes and temperatures in each tube section. The number of flow
310
-------
passages in each section must be large enough, within practical limits, to
effectively reduce the pressure drop to a minimum in each section.
In some cases, external finning or external resistance is required
together with internal finning in order to control optimum heat fluxes. To do
this, it is necessary to optimize the external and internal geometries to
obtain the best design.
Design trade-offs using internal finning on the Fluorinol vapor generator
for the purpose of lowering wall temperatures are given in the design section
of this report.
D. Instrumentation Support
Geoscience has also reviewed the water test bed vapor generator
that has been fabricated by Solar and made recommendations concerning
primarily thermocouple instrumentation. It is important that the actual
steady state and transient performances of the system are determined and
compared to design values.
311
-------
GLR-102
VAPOR GENERATOR TECHNOLOGY SUPPORT
FOR SOLAR
DIVISION OF INTERNATIONAL HARVESTER
(P03739-32134-F03)
C. M. Sabin
H. F. Poppendiek
G. Mouritzen
R. K. Fergin
GEOSCIENCE LTD
410 South Cedros Avenue
Sol ana Beach, California 92075
313
-------
I. SUMMARY
Geoscience's support role in the subject program consists of two
study areas (1) vapor generator design, and (2) heat transfer and fluid flow
support for the vapor generator-combustor system. The vapor generator
design work involves an investigation of the relationships between the fluid
properties and the requirements imposed on the vaporizer by other compon-
ents of the power system.
During the last quarter, Geoscience performed the following tasks:
(1) final Fluorinol-85 vapor generator design, (2) final AEF-78 vapor genera-
tor design, (3) water test bed vapor generator performance experiments,
and (4) several support studies. Also there are summarized project meetings
and test cell experiments.
II. PROJECT MEETINGS AND TEST CELL EXPERIMENTS
A. Ann Arbor Rankine Cycle Contractor's Meeting (January 20-21, 1972).
Geoscience participated in the Solar presentation by reporting on its
vapor generator design and support studies for Solar.
B. Progress Report to S. Luchter, P. Hutchins, H. Naser and A. Kreeger
(January 27, 1972)
Geoscience reported progress to date and reviewed questions relative
to AEF-78 physical properties and vapor generator design efficiencies.
C. Technical Discussions With Aerojet Staff Members (February 7, 1972)
Solar and Geoscience staff visited Aerojet to review physical proper-
ties and thermal stability questions in addition to considering vapor generator
efficiency limitations.
D. Progress Report to P. Hutchins (February 23, 1972).
The water test bed vapor generator was operated for Mr. Hutchins at
Solar. Geoscience reviewed and interpreted the steady state and transient
performance characteristics of the generator. Additional steady state water
test bed vapor generator experiments were scheduled to be performed within
the next several weeks.
E. Progress Report to G. Thur and W. Mirsky (February 24, 1972).
The water test bed vapor generator was operated for Messrs Thur and
Mirsky. Geoscience reviewed and interpreted the steady state and transient
314
-------
performance characteristics of the generator. Additional steady state water
test bed vapor generator experiments were scheduled to be performed within
the next several weeks.
F. Test Cell Experiments
In this quarterly period, Geoscience staff assisted Solar in the opera-
tion of the water test bed vapor generator. During the many tests performed,
Geoscience proposed modifications to the test cell equipment to simplify
system operation; these modifications were subsequently made. Steady state
and startup performance data were obtained during these tests.
III. VAPOR GENERATOR DESIGN STUDIES
Two vapor generator designs have been completed during this quarter.
These are for the fluids Fluorinol-85 and AEF-78. Descriptions of these two
vapor generatores are given below. The test bed water vapor generator
operation has shown that it performs as expected. Some details on the opera-
tion are presented in the following paragraphs, also.
A. Fluorinol-85 Vapor Generator Design
A demonstration prototype vapor generator was designed for a
Fluorinol-85 Rankine cycle engine system. The resulting design conditions are
Fluorinol-85 Side:
Flow 10,000 Ibm/hr
Pressure drop 89 psi
Outlet pressure 700 psia
Inlet temperature 287°F (at max. power)
Outlet temperature 550°F
Heat transfer rate 2.25 x 106 Btu/hr
Efficiency 81% based on HHV of JP-5 (19800 Btu/lb)
Max tube wall temperature 506°F
Gas Side:
Air-fuel ratio 25 to 1 (JP-5 fuel)
Flow (gas) 3740 Ibm/hr
Pressure drop 4.3 inches H^O
Outlet pressure atmospheric
Inlet temperature 2500°F (mean) ± 250°F
In the direction of the gas flow, the vapor generator consists of a
vaporizer, a superheater and preheater. The prcheater and superheater are
in counterflow to the gas and the vaporizer is parallel.
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The preheater consists of ten rows of 45 tubes each (i/4-inch (JlJ
tubing with 0.016 inch walls). Longitudinal and transverse spacing is 5/16-
inch in a triangular array. Tube heat transfer length is 22 inches with an
outside area of 53. 9 ft .
Material, stainless steel or carbon steel. First eight rows of preheater
measured from cold end, are 45 tubes in parallel. Remaining two rows are
15 tubes in parallel.
Flow exiting preheater passes into first vaporizer row adjacent to com-
bustor, then into second vaporizer row, in a cross parallel flow arrangement.
From second vaporizer row, saturated vapor flows into 1/2-inch tube row
farthest from combustor, passes through two remaining 1/2-inch tube rows in
a cross-counterflow arrangement.
All 1/2-inch tubes are arranged seven tubes in parallel, so that each
parallel path makes three passes in each row.
The five vaporizer and superheater rows consist of 1/2-inch OD tubing,
0.030-inch wall, 16 internal longitudinal fins 0.030-inch thick by 0.056-inch
high. Material is carbon steel, longitudinal and transverse spacing is 0.675-
inch. Active heat transfer length is 22 inches with a total outside heat transfer
area of 25.4 ft2.
In order to control the amount of heat transferred through the vaporizer,
the tubes are insulated with a foil tube forming a 0.010-inch thick insulating
layer of gas around the vaporizer tubes. The combined conduction and radiation
effects result in all saturated vapor at the vaporizer outlet.
All three rows of the superheater are, likewise, insulated with a 0.020-
inch thick gas layer to obtain superheated vapor of 550°F at the outlet. An
experiment was performed at Geoscience to support the analysis of the foil tube.
The maximum tube wall temperature was calculated to occur at the
superheater outlet. This temperature is 566°F as compared to the design limit
of 575°F. It is not feasible to lower this temperature for safety reasons since
the outlet temperature of the Fluorinol is 550°F. Tube wall temperatures at
the vaporizer inlet and outlet and at the superheater inlet are about 470°F.
I
Carbon steel tubing was used in the design in order to obtain minimum
peripheral tube wall temperature variations. Also, internal finning was used
in designing the vaporizer and superheater in order to obtain minimum tube
wall temperature.
316
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B. AEF-78 Vapor Generator Design
An AEF-78 vapor generator designed for 85.5 percent thermal
efficiency (based on the LHV) has been completed. This unit is of rectangular
cross section, utilizing straight tubes in tube sheets. The heat exchanger is
made up of four rows of 15 one-half inch diameter internally finned tubes at
the combustor end, followed by ten rows of 60 one-fourth inch diameter plain
tubes at the cooler end. The total gas side heat exchange area is approximately
106 square feet. Although the most critical location in the matrix with respect
to fluid overheating is at the outlet, the combination of the high temperature
tolerance of AEF-78 (the maximum acceptable tube wall temperature is 810°F)
and the internally finned tubing allow this heat exchanger to be arranged com-
pletely in a cross counterflow arrangement, without the complex paths required
in the test bed water vapor generator and the Fluorinol-85 vapor generator.
This AEF-78 vapor generator is expected to have a maximum tube wall tem-
perature near 720°F.
Physical characteristics of this design are listed in the table below.
With the specified flow conditions imposed by the other engine compon-
ents on this design, a thermal efficiency of 85.5 percent based on the lower
heating value requires a heat exchanger effectiveness of 0.97. In this range
of effectiveness, an increase of only one percent in effectiveness requires a
heat transfer area increase of over ten percent. In a joint meeting with
Aerojet personnel, it was decided that the decrease in size, complexity, and
weight of the vapor generator justified a decrease in the specified thermal
efficiency. It was, therefore, decided to decrease the efficiency to 82.2 per-
cent, and thereby eliminate three rows of quarter inch tubing. With this
change, the total number of quarter inch tubes decreases from 600 to 420,
arranged in seven rows of 60 tubes each.
A change such as this does not affect the vapor flow rate or heat
transfer rate. Instead, the combustion products flow rate increases (by some
three percent) and the exhaust temperature rises about 100°F. The fuel-air
ratio is, of course, held constant, so the combustion products inlet to the
vapor generator remains at 2500°F.
C. The Water Test-Bed Vapor Generator Performance
During the past quarterly period, the water test bed vapor generator
has been operated with the Solar combustor for the purpose of determining
experimental steady state and transient performance characteristics.
A number of experiments have been performed with the vapor generator
covering the following ranges of parameters:
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TABLE I
Characteristics of AEF-78 Vapor Generator for 85. 5 Percent Thermal
Efficiency (Based on the Lower Heating Value)
AEF-78 Side:
Flow 19,3001b/hr
Pressure drop 20 psi
Outlet pressure 1,000 psi
Inlet temperature 396°F
Outlet temperature 650°F
Heat transfer rate to organic 2.02 x 10^ Btu/hr
Gas Side:
Outlet temperature 455°F
Flow 3,400 Ib/hr
Pressure drop 2.9 inches of water
Inlet temperature 2500°F mean
Tube Arrangement:
Four rows of 30 tubes each, internally finned 1/2 inch OD carbon
steel tubing at combustor end of heat exchanger. Internal fins
identical to those for Fluorinol-85 vapor generator. Exterior of
tubes bare.
Ten rows of 60 tubes each, bare 1/4 inch OD tubing at combustion
products outlet end.
AEF-78 flow path through exchanger is cross counterflow. Cold
fluid enters at combustion products outlet end and passes through
1/4 inch tubing with 60 tubes in each row in parallel. Return bends
are manifolded. The half-inch tubing is to be ten parallel paths,
so each group of three tubes in each row is in series. Superheated
vapor outlet collector manifold is at row next to combustor outlet
face.
Physical Characteristics of Tube Bank:
Tubes 23 inches long (heat transfer length)
Face area 23 x 19.1 inches
Total thickness 5.6 inches
Tube bank weight (without
tube sheets or headers) 92 Ibs
Liquid hold up weight 38 Ibs.
318
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(1) Water flow rate range 400-1700 Ibs/hr
(2) Percent air 22-51 percent
(3) Fuel rate 40-75 Ib/hr
(4) Inlet water pressure 150-1100 psia
(5) Outlet water pressure 15-1100 psia
(6) Exit water temperature up to 1100°F
(7) Degree of superheat 9-580°F
When operating the system so that the temperature out of the preheater
was close to the saturation temperature, the flows and the temperature pro-
files in the six parallel passages (through the vaporizer and superheater) were
satisfactorily balanced.
IV. SUPPORT STUDIES
The vapor generator support activities carried on by Geoscience during
this quarter included analysis of two phase flow in parallel channels, some
assistance to Solar on fabrication sources, and proposals for the fabrication
or mechanical simplification of the vapor generators. Typical examples of
this work are given in the following paragraphs.
A. Two Phase Flow in Parallel Channels
Analyses have been made for the differences in mass flow rate through
parallel tube vapor generators for the cases of liquid flow in some channels
and vapor in others (hypothetical -worst case). In particular, one is interested
in the difference in maximum wall temperature at the vapor generator exit for
the case where half the tubes are filled with vapor and another half liquid. The
results of the analyses show that a number of the important system parameters
control this process. An important variable is the fluid density. One can show
that a water vapor-filled tube can have approximately twice the temperature
rise that a liquid-filled tube has for a typical set of design conditions. In an
actual situation, however, this temperature difference would be less because
the tubes would contain two phases rather than either vapor or liquid alone. It
is pointed out that for the Fluorinol-85 case, this hypothetical result would be
much smaller. These calculated performance features were noted during the
water test bed vapor generator tests. For example, flow asymmetries could
be generated by reducing the vapor generator back pressure (to create a large
liquid-vapor density ratio). It was also possible to show that flow asymmetries
could be created by increasing the water flow rate to such a value that the exit
temperature of the preheater was well below the saturation temperature.
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When the water test bed vapor generator is operated with very large
water flow and low combustor power, the liquid from the preheater passes
into the vaporizer rows very much subcooled. Under this condition, the
location at which vapor generation begins is not well established, and small
variations in flow from one channel to the next can cause chugging and erratic
behavior. If in the startup process the combustor power level and water flow
rate are reasonably well matched, startup occurs without difficulty, since the
first vaporization occurs at nearly the same location in all tubes.
B. Proposals For the Simplification of the Fluorinol-85 Vapor Generator
Return Bends
The tube connection arrangement for the Fluorinol-85 vapor generator
which was proposed in the sketches submitted with the design utilized mani-
folded return bends through the preheater. In the preheater rows 9 and 10,
where the flow makes three passes instead of one, the manifolded return
bends led to a header design which is bulky and expensive to produce. It is,
therefore, suggested that an individual return bend arrangement be made for
the preheater rows 8, 9, and 10. The three-to-one flow area reduction
required in going from row 8 to 9 would then be accomplished locally, with 15
identical reducing return passages, so that these returns need not have a
passage depth from the tube sheet greater than that of the manifolded return
bends on rows 1 through 7. This arrangement also allows the use of 15 (or
less) small tubes as connectors between the preheater outlet and the vaporizer
inlet manifold, so that less space is required for this connection.
The construction of the tube bend assembly can be accomplished in
several ways. It is possible to machine the assembly from a. solid block. It
is also possible, however, to braze the assembly from a number of small
subassemblies in several stages. This latter approach may decrease the
machining costs significantly, so that a net savings is realized in spite of the
additional steps in assembly. The individual return bends in preheater rows
8, 9, and 10 and in the vaporizer and superheater sections could, for example,
be built up from short pieces of drawn tubing reformed in cross section to
provide the required passages. The row-to-row pressure drop is small, so
that pressure across these dividers need not be a serious consideration in
choice of their wall thickness. They would, however, be loaded in tension,
since they would provide the link between the cover plate and the tube sheet.
If the manifolds were built up of tubing, there would be voids in the assembly
which were not part of the flow passages. However, these would be sealed
and need not present a problem.
It is also possible to form the return bend arrangement for the super-
heater (or vaporizer) from reformed tubing sections. These would have to be
in the order of 1/2 to 5/8 of an inch thick to provide adequate flow area. If
the preheater return assembly were made up of the same thickness, a continuous
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cover plate could be used for the entire assembly. This cover plate would be
perforated only by inlet, outlet, and section interconnection lines. It appears
that a rather large decrease in machine work would be accomplished by this
means.
The manifolding could also be made up with return bends, separators
machined from a single plate, but using a cover plate rather than machining
the cover integral. In this manner, a large portion of the metal removal
could be accomplished by through drilling, rather than by use of an end mill.
The cover plate would then be brazed on in a separate step.
V. DESIGN PROCEDURES FOR COMBUSTION HEATED FORCED CONVEC-
TION VAPOR GENERATORS
A. Principal Design Steps
The vapor generator has performance criteria specified by require-
ments of the system. This information usually includes the following items:
Inlet thermodynamic state of both streams
Mass flows of both streams
Outlet vapor thermodynamic state
Minimum acceptable thermal efficiency of burner-vaporizer
combination at full load
Maximum acceptable pressure drops in both streams
Matrix volume limits
If these specifications are made arbitrarily, it may be impossible to
satisfy them all. More discussion of this aspect appears in Section B.
The first five steps in the design procedure are straighforward
thermodynamic calculations. However, it is frequently necessary to use
some arrangement for superheater and vaporizer other than counterflow to
limit tube wall temperatures so these steps may be done several times.
Step 1 By energy balance and with properties tables for the fluids,
establish for the preheater, vaporizer, and superheater sections:
Enthalpy changes in each stream
Heat flow
State points between sections
321
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If pressure drops are expected to alter fluid state points
significantly, assume some pressure drop distribution for the
above calculations.
The next four steps may be done by several equivalent procedures.
The notation used here is from Kays and London, Compact Heat Exchangers,
2nd Edition.
Step 2 For both streams compute the capacity rates for each of the
three sections. The capacity rate for a vaporizing fluid is
infinite. Compute the capacity rate ratios for each section.
Step 3 Compute heat transfer effectiveness of each section.
Step 4 Compute the NTU requirement for each section, utilizing a
reasonable choice for the heat exchanger arrangement, such
as multipass counter cross flow.
Step 5 Using the appropriate value of Cmjn, compute the product AU
from the NTU values for each section.
The establishment of the overall conductance requires computation of
hot and cold side heat transfer conductances. For normal combinations of
fluids, the controlling heat transfer resistance will be on the hot, combustion
products side. It is at this point in the procedure that the major choices are
available to the designer. There are several tradeoffs which must be con-
sidered.
a. For most heat exchange surfaces, pressure drop goes up more
rapidly with increasing velocity than does heat transfer conduct-
ance. If pressure drop is an important criterion, as it is with
most automotive power plants, then one wants to use the maxi-
mum available frontal area, and in addition consider configura-
tions which have relatively large free flow area.
b. Neither the conductance nor the area per unit of volume of a sur-
face configuration is to be maximized but, instead, the product
ah must be maximized for a compact matrix.
c. Extended surfaces on the hot side increase heat fluxes on the
vaporizing fluid side. If the vaporizing fluid is an organic and
subject to thermal decomposition or if the possibility of film
boiling exists, extended surfaces cannot be used without con-
sideration of the cold side conditions.
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An example of such a condition occurs in the Fluorinol-85 vapor
generator near the preheater outlet. The overall U at this
location (which depends only weakly on the internal conductance),
together with the local AT, specifies a heat flux which could lead
to unacceptably high wall temperatures. The internal velocity
was tripled at this location in order to raise the internal conduc-
tance and bring the wall temperature down. If an extended surface
configuration had been used which held the overall ah constant but
decreased the internal wall area, the internal velocity would have
to be increased still more to control wall temperature. Pressure
drop goes up rapidly with velocity increase so such a change would
probably make the internal pressure drop too high. An extended
surface configuration which had a lesser ah product would, of
course, lead to a preheater with larger volume.
Step 6 Choose a hot side heat exchanger surface appropriate to the hot
side surface and determine appropriate conductance relationships.
Proper heat transfer relationships for the cold side flows may be
chosen with the aid of the following guidelines.
a. For the all-liquid (preheat) and all vapor (superheat) locations
use standard single phase relationships. Beware of using
relationships for fully established flow for computations in-
volving short runs of tubing.
b. Subcooled boiling may occur near the preheater outlet.
c. The point of the so-called "boiling burnout" or "transition
boiling" which corresponds to the end of the wetted wall in the
vaporization region, marks the change from boiling or liquid-
like conductances to vapor-like conductances. This point is
usually found at vapor qualities significantly less than unity.
The location depends upon the fluid properties, the mass flux,
heat flux, orientation in the gravitational field, etc. A wealth
of literature is available which describes this phenomenon
with varying degrees of accuracy.
d. The actual boiling conductance depends upon the fluid the local
thermodynamic conditions, and the phase distributions.
e. The maximum heat flux which can be applied during boiling
without film boiling is limited. If the computed wall tempera-
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tare under wetted wall boiling is too high, the liquid will not
remain in contact with the wall, since the liquid in contact
with the wall must clearly be superheated.
Step 8 Compute cold side conductances as functions of pressure drop,
number of parallel passages, or tube diameter (in the case of a
monotube circuit). Choose particular combinations for each
section or subsection which meet the design requirements.
Step 9 Compute overall conductances using proper corrections for fin
effectiveness, wall resistance, etc. Beware of fin effectivenesses
which are too low. Fin effectiveness relationships are based upon
the idealization that the gas side conductances are uniform. Since
they are in actuality not uniform, the designer depends upon con-
duction in the fin to transfer heat from one area of the fin to
another. Low fin effectiveness implies that the heat flow into the
tube wall from the fin will not be uniform.
Step 10 Compute the required heat transfer area and the heat exchanger
volume from the required AU product determined in Step 5. If
the volume is too great, choose a more compact surface (larger
ah product).
Step 11 Compute the gas side pressure drop.
It may be observed that the frontal area, heat transfer surface type
and array, and flow paths have been determined, but that the construction of
the matrix is not defined. All of these calculations are equally valid for a
round cross section or a tube-in-sheet geometry. If a tube sheet geometry is
chosen, it must have a number of tubes in each row which is evenly divisible
by the number of parallel flow paths. A nested spiral arrangement may have
any number. The tube sheet arrangement has a high cold side pressure loss
in tube bends, and this pressure loss does not contribute to the heat transfer
performance.
Step 12 Choose configuration, manifolding connections, tube wall thick-
nesses and other mechanical details. Since an approximate tube
wall thickness was necessarily made before internal conductances
and pressure drops could be calculated, a large change in tube
wall thickness could require recalculation of some of the above
steps.
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B. Vapor Generator Example Design for AEF-78
Initial Data:
AEF-78
Flow 19,300 Ib/hr
^p 50 psi goal, 100 psi max.
Outlet pressure 1000 psia
Inlet temperature 396°F
Heat transfer rate to organic 2.02 x 10° Btu/hr
Gas Side
Outlet temperature
Air-fuel ratio
Flow
Outlet pressure
Inlet temperature
General Constraints
Efficiency based on HHV
Efficiency based on LHV
Maximum tube wall
temperature
Maximum face size
Maximum core thickness
Step 1
Step 2
455°F
25:1 (JP-5)
3400 Ib/hr
3.0 inches H2O
atmospheric
2500 F (mean) ±250°F
80 percent
85. 5 percent
810°F
26 in. diameter circle, or 26 in.
by 21 in. rectangle
6 inches
In this case there is no change of phase, so energy balance
specified all information for Step 2.
Overall gas capacity rate:
mm
2.02 x 10 Btu/hr
2045°F
= 988 Btu/hr°F
max
2.02 x 10 Btu/hr
254°F
= 7950 Btu/hr°F
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Step 3 For the entire matrix:
_ 2000 F - 455°F _
6 ~ 2500 F - 396°F ' °'9?
Step 4 A reasonable first choice of flow arrangement for this unit is
multipass counter cross flow. For the probable number of passes,
this arrangement is indistinguishable from counterflow. From
the appropriate equation in Kays and London,
A TT
NTU = U = 3.9
min
Step 5
AU = 3.9 (988 Btu/hr°F) = 3870 Btu/hr°F
Step 6 Choose 1/2 inch OD plain tubes in a staggered array, 5/4 dia-
meter on centers. It has been shown that this configuration yields
compact vapor generators. For more details on the basis for this
choice among other staggered bare tube arrays, see Geoscience
Ltd quarterly report to Solar for June to August 1971, GLR-95.
The heat transfer and flow friction characteristics of this geo-
metry are to be found in Figure 10-6, page 185 of Kays and
London's Compact Heat Exchangers, 2nd Edition.
Choose a rectangular matrix of 10.09 inches by 23 inches.
Af = 3.05 ft
This area corresponds to 30 tubes in each row of one-half inch
tubes, each 23 inches long.
For average gas properties,
h = 31.3 Btu/ft2 hr°F
O
This value ranges from a maximum (at the combustion products
inlet) of 38.9 Btu/ft2 hr°F to a minimum (at the combustion
products outlet of 24.8 Btu/ft2°F.
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For one quarter inch tubing in a geometrically similar array,
the corresponding values are
h =42.9 Btu/ft2 hr°F
mean
h. , . =51.5 Btu/ft hrcF
inlet
h .. . = 32.2 Btu/ft2 hr°F
outlet
Step 7 The internal conductances may be computed based on standard
correlations. No details can be shown because of the proprietary
nature of the fluid AEF-78. The fluid throughout this vapor
generator may be treated as single phase. There are, however,
some temperature levels at which the transport properties vary
rapidly with small temperature changes.
Step 8 For ten parallel paths, the internal conductance at 650°F is
hr = 895 Btu/ft2 hr°F
This value is based on the Dittus-Boelter correlation for heat
transfer to a single phase fluid in a tube. The half inch tubing
is internally finned and, including the fin effectiveness, approxi-
mately doubles the heat transfer over that of a bare tube.
Step 9 If the gas side conductance is computed for the inlet state of the
combustion products (the same location where the vapor exits),
the gas side conductance may be determined to be
h = 38.9 Btu/ft2 hr°F
O
The overall conductance, including the wall infludence is,
U = 38.0 Btu/ft2 hr°F
The overall AT at this location is 1850°F, so that the local heat
flux is (based on the outside wall area)
.) = 70,300 Btu/ft2hr
outlet
The inside wall to vapor temperature difference is 43°F. This is,
of course, a mean value and does not take into consideration the
peripheral variations in heat flux due to radiation from the com-
bustor and circumferential distribution of gas side heat flux. This
mean value does not include possible variations in local flux from
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combustor hot spots or local high velocities. All of these varia-
tions have been discussed in previous quarterly reports.
The apparent mean wall temperature at the superheated vapor exit,
based on gas convection alone, is 693°F. This value, which is very
conservative compared to the 810°F maximum allowable, therefore
allows the design to be tolerant of local high fluxes and variations
in fluid properties from those so far measured for this fluid.
A typical relationship between number of flow passages, heat
transfer, and pressure drop is shown in Table II (for a one-half
inch OD by 0.020 wall thickness tube which is internally smooth,
and unf inned).
TABLE II
Number of
parallel
passages
1
2
5
10
15
20
25
h
Btu/hr ft2 F
5640
3240
1560
895
648
513
434
P dyn
psi
157
39
6.3
1.6
0.7
0.4
0.25
1/Amax
Based on Twall = 810°F
0.9 x 10*?
0. 52 x 10^
0.16 x 106
0. 14 x 10
0. 10 x 10^
0.082 x 10
0.070 x 10
Step 10 The quantity, a , for this staggered tube configuration and one-
half inch tubes is 48.3 ft2/ft , so that
UQ = 1500 Btu/ft3 hr F
The total volume available for this vapor generator is 1. 53 ft ,
so that based on the above figure, the maximum AU product with
this configuration (neglecting wall and liquid heat transfer resis-
tances) is 2290 Btu/hr°F, less than the 3870 Btu/hr°F required to
meet the specified performance. Clearly the entire matrix cannot
be made up of one half inch plain tubing.
One quarter inch tubing in an array similar to that of the one half
inch tubing has an a of 96.6 ft2/ft3, so that
Ua = 3660 Btu/ft3 hr°F
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based on properties at temperatures typical of the cooler end of
the exchanger. The matrix may be made up of a combination of
one-half inch and quarter-inch tubing. Such a combination is
four rows of one-half inch tubing, 30 tubes per row, and ten rows
of one-fourth inch tubing, 60 tubes per row.
Step 11 The gas side pressure drop for this configuration may be computed
to be 2.9 inches of water. This is within the design maximum limit.
Step 12 It can be shown for this geometry that the inside wall temperature
in the last row of the one-fourth inch tubing is not critical with all
60 tubes in parallel, and since this flow path configuration yields
a simple manifolded return bend arrangement for the entire small-
tube section, it is chosen for the entire quarter inch tube section.
At the connection between the one-fourth and one-half inch sections,
the flow goes into a 10-tubes-in-parallel arrangement, so that six
one-fourth inch tubes feed each of 10 one-half inch tubes. Each
of the 10 streams in the one-half inch section therefore makes
three passes across the exchanger in each row, for a total of 12
passes between the last one-quarter inch tube row and the super-
heater vapor outlet.
Commentary on the Design:
The one-fourth inch section of this exchanger contains 60 tubes.
Because the effectiveness is very near one, a small decrease in performance
yields a large decrease in tube number, and in weight. After consultation
with Aerojet, Solar, and EPA personnel, the effectiveness and efficiency were
decreased, so that the one-fourth inch tube section was reduced to seven rows.
The consequences of this change are discussed in Section C of this report.
Portions of this exchanger could very well be designed with finned
tubing as well as plain tubing. Some precautions associated with such a
change are noted in the description of the Design Steps. It should be clear
that there are many configurations which will satisfy the particular design
requirements, and that choices must be made frequently on the basis of cost,
materials and component availability, fabrication techniques, and other con-
siderations not related to the heat transfer calculations.
C. Effects of Thermal Efficiency Specification
The usual definition of thermal efficiency of a combustor-vapor
generator assembly is the ratio of the heat transferred to the working fluid
divided by the heat of combustion of the fuel added. This definition may be •
formed either with the lower or higher heating value. Therefore,
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C
LHV ~ W (LHV)
(20)
where
7? is the thermal efficiency
C is the combustion products capacity rate
Tjn is the combustion products temperature into the vapor
generator
T . is the combustion products temperature out of the vapor
generator
Wp is the flow
LHV is the lower heating value of the fuel
For an efficient combustor with small heat loss, Equation (2) may be
written
C (T. -
*I*V = C(T. -T h) (21)
v in comb'
where
T i is the air (and fuel) temperature at the combustor inlet
The heat exchanger effectiveness is defined as the ratio of the heat trans-
ferred to that which would be transferred in an infinite counterflow heat
exchanger, i.e. , for the condition of zero temperature difference between
the inlet cold side temperature and the outlet hot side temperature. Therefore,
C(T. - T )
C = " _°"t (22)
where
C
T • is the inlet cold side temperature
The quantity (. cannot exceed unity.
These two expressions, (21) and (22), can be combined to yield a
relationship between €. and 77.
(T. - T . )
<23>
.
in comb
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The efficiency of a boiler-vapor generator combination is limited by
the difference between the ambient air and the working fluid inlet, even with
very large heat exchangers, where € approaches unity.
A graph of Equation (23) utilizing a 2500°F peak combustion products
temperature, a 70°F ambient and using inlet working fluid temperature as a
parameter, is shown in Figure 13.
The relationships between effectiveness and heat exchanger size are
all exponential in form, so that for effectiveness near unity, a small increase
in effectiveness requires a large increase in exchanger size. Figure 14 is a
graph of heat exchanger size as a function of inlet liquid temperature and
T) for a counterflow arrangement.
-L/rl V
A detailed analysis of these considerations has been worked out for the
AEF-78 vapor generator, which is a cross counterflow tube bank made up
partly of one-half inch tubing (in the hot section) and partly of one-fourth
inch tubing (in the cool section). Figure 15 shows the percentage change in
total gas side heat transfer area as a function of thermal efficiency for the
design operating conditions. The design operating liquid inlet temperature
is 396°F. It may be seen that the slope of this curve is very steep. At the
1.0
09
08
0.7
06
O.5
04 -
0.3 -
02 -
ILHV = '
IOBTC,N=TCO«>
200
300
400
5OO 60O
INLET LIQUID TEMPERATURE, 'f
FIGURE 13. RELATIONSHIP BETWEEN THERMAL, EFFICIENCY, EFFECT-
IVENESS, AND INLET LIQUID TEMPERATURE
331
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200 300 400
INLET LIQUID TEMPERATURE ,*F
FIGURE 14.
VAPOR GENERATOR SIZE AS EFFECTED BY EFFICIENCY
AND LIQUID INLET TEMPERATURE 70°F AMBIENT AIR
85 percent efficiency level, a one percent change in thermal efficiency causes
a ten percent change in required heat transfer area.
For this particular vapor generator, the design thermal efficiency was
very difficult to meet within the required volume, and the one-fourth inch tube
section consisted of a rather formidable number of tubes. Discussions
between Geoscience, Solar, Aerojet and EPA led to a reduction in the thermal
efficiency requirement to simplify the matrix. A graph of the relationship
between heat transfer area, thermal efficiency, and number of one-fourth
inch tubes is shown in Figure 16.
D. Factors Affecting Gas Side Pressure Drop
Two equations are given by Kays and London for heat exchanger
pressure drop. For simplicity, the equation appropriate to tube banks will
be used here. The following arguments hold for both.
The pressure drop relationship is
(24)
332
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too
96
96
94
92
90
88
66
84
62
80
78
76
f 74
72
70
68
66
64
I I I I I I I I I I
68 70 72 74 76 78 8O 82 84 86 86 90
n THERMAL EFFICIENCY, PERCENT
FIGURE 15. EFFECT OF VAPOR GENERATOR THERMAL, EFFICIENCY
UPON REQUIRED HEAT TRANSFER AREA OF AEF-78 MATRIX
m
>c
where
Ap is pressure drop
G is mass velocity, w/Ac (It is based on minimum free flow area)
P-. is inlet gas density
P^ is outlet gas density
is mean gas density
is the proportionality constant in Newton's second law
is the ratio between free flow area and frontal area for the
O matrix
f is friction factor
A is heat exchange area
A is minimum free flow area
w is mass flow rate
For simplicity, density variations are neglected in this discussion
(first term in the square brackets). In a real combustion heated vapor genera-
tor, flow accelerations and density changes may not be neglected.
333
-------
120
no
100
95
90
85
80
75
70
65
60
55
50
45
40
35
ROWS «GO TUBES
76 77 78 79 80 81 82 83 84 85 86 87
IJ THERMAL EFFICIENCY, PERCENT
FIGURE 16. AEF-78 VAPOR GENERATOR EFFECT OF THERMAL
EFFICIENCY UPON HEAT TRANSFER AREA REQUIRED
Since the design mass flow rate and not mass velocity is specified,
make this substitution. Then Equation (24) becomes
AP =
w
fA
2g~
(25)
where
A, is the heat exchanger frontal area
It may be clearly seen that both frontal area and contraction coefficient a have
a very strong effect upon the pressure drop.
For typical heat exchange configurations, the conductance variation
with Reynolds modulus may be written in the form
= KRe
'm
(26)
334
-------
where
h is heat transfer conductance
K is a constant dependent upon the configuration, but also
combining some gas properties dependent upon the heat
exchanger operating conditions
m is approximately constant, and varies between 0.38 and
0.55 for must surfaces
Re the Reynolds modulus, is 4 r.
Equation (26) in its expanded form, is
h =
(-m)
- m)
rhY 1 / 1 \(1
T) J(£)
Equation (27) indicates the manner in which the gas side conductance depends
upon the gas flow rate, exchanger frontal area, and the matrix contraction
coefficient.
From this relationship, the principal reason that the effectiveness
(and efficiency) of a vapor generator rises at part load can be seen. The
effectiveness is a function of the quantity NTU, which contains conductance
in the numerator and flow rate in the denominator. The overall conductance
is controlled by the gas side conductance in the usual vapor generator. If
the flow rate is decreased, the gas side conductance decreases to a lesser
extent, so the conductance-flow rate ratio (and the value of NTU) increases.
Consequently, the heat exchange effectiveness rises.
The ratio of pressure drop to heat transfer conductance may be con-
structed from Equations (25) and (27)
AP J1 * "> tA
K
Pressure drop rises with a decrease in free flow area much more rapidly
than does the heat transfer conductance.
335
-------
NOMENCLATURE
q heat flax
Cp specific heat at constant pressure
Pr Prandtl number
\l viscosity
k thermal conductivity
€ effectiveness
AU overall thermal conductance
Z distance along the tube
P pressure
J a function of quality and pressure
A, K constants
NTU number of transfer units
St Stanton number
Re Reynolds number
W mass flow rate
x an exponent which describes, in part, the dependence of
Stanton number on Reynolds number, also quality
h heat transfer coefficient, also enthalpy
C capacitance rate
T temperature
337
-------
4P pressure drop
G total mass flow rate per unit free flow area, also steady state
gain
v specific volume
O ratio free flow/frontal area, or surface tension in Ib/ft
f fanning friction factor, or Darcy-Weisbach fraction factor
Ot ratio gas side heat transfer area/volume, also void fraction
S transverse spacing, center to center
D spacing between tube rows, center to center
£p effectiveness for one tube row
Vv mean vapor velocity
Rv volumetric flow ratio
p density
C"C', m empirically determined constants
L matrix depth, also tube length
DH hydraulic diameter
No overall surface effectiveness
gc gravitational mass-force-time-length conversion factor (4. 17
x 108 ft/lbm/lb hr2), or 32.2 ft Ibm/lb sec2
a exponent describing dependence of Stanton number on Reynolds
number
AF frontal area
V core volume
M core weight, also a loss parameter defined by equation (20)
6 fin thickness
338
-------
4J a small perturbation in the quantity J
© average surface heat flux for one channel of a parallel flow
system, or its La Place transform
H steady state change in coolant enthalpy as it flows through
the channel
A, B, C.D, parameters which depend on the unperturbed conditions in a
E,0t, B parallel flow channel
z1 a weighted distance through the heated channel, defined by
equation (11)
fj • a function used to evaluate friction pressure drop in two phase
flow, defined in equation (8) of Appendix VI..
P average static density
RQ void fraction
A liquid-vapor volume flow rate ratio
g acceleration due to gravity
Af flow area of heated channel
Li inlet channel length
Afi flow area of inlet channel
Afe flow area of exit channel
F Darcy-Weisbach friction factor
KJ entrance loss factor for heated channel
Ke exit loss factor for heated channel
3P£/3z pressure gradient due to friction losses
B, B£ defined in text
Hf latent heat of vaporization
339
-------
tube radius
TBF
I, through
16
(JO
r, f,
h
Subscripts
sat
crit
f
G
b
min
max
TPF
LO
source
w
two phase flow friction multiplier
integrals used in evaluating hydrodynamic stability, defined
in equations (25 through 30)
undamped natural frequency
damping factor
functions tabulated in Reference (8)
enthalpy
for saturated liquid
critical (for dryout)
evaluated at the mean between bulk and wall temperature
gas phase
average bulk property
minimum capitance side
maximum capitance side
two phase flow
computed assuming the entire flow is in the liquid phase
inlet, or at inlet to the row
exit
liquid phase
source
water
340
-------
g
w
m
h
c
gas side
tube wall
mean value
hot side
cold side
v vapor
LPF computed assuming that only the liquid phase is flowing,
that is, total flow rate is G(l-x)
341
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343
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