FINAL REPORT
   LOW EMISSION  BURNER
FOR  RANKINE CYCLE  ENGINES
     FOR AUTOMOBILES
    U.S. Environmental Protection Agency

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Low Emission  Burner for
 Rankine Cycle Engines
     for Automobiles
          FINAL REPORT
                By
             T. E. Duffy
            J. R. Shekleton
            R. T. LeCren
            W. A. Compton
             Prepared for

        Department of Motor Vehicles
        Research and Development
        Air Pollution Control Office
      < Environmental Protection Agency
          a
         H SOLAR
         OmSIOK OF INTERNATIONAL HARVESTER COMPANY
         2200 PACIFIC HIGHWAY -SAN DIEGO. CALIFORNIA 92112
                                    KDR 1695

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                                 FOREWORD
        This Final Report covers the work performed under EPA/APCO during the
work period, July 1, 1970 to March 31,  1971.  It is published for information only,
and does not necessarily represent the recommendations, conclusion, or approval of
the Environmental Protection Agency, National Air Pollution Control Administration.

        The contract with the Research Laboratories of Solar Division of International
Harvester Company, San Diego, California,  was initiated by the EPA/APCO under
contract number EHS 70-106, monitored by Mr. F. Peter Hutchins, Project Officer.

        The program was under the general direction of Mr. W. A.  Compton,  .
Assistant Director-Research, who served as Program Director.  Mr. R.  T.  LeCren,
Group Engineer was the principal investigator  at Solar.   Mr. Thomas E. Duffy,
Research Staff Engineer was the specialist on controls and systems.  Mr. Jack R.
Shekleton, Engineering Specialist was the  leader in combustion system development.

        This report is identified by Solar as RDR 1695.
                                      11

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                                 CONTENTS


Section                                                                Page

  1          INTRODUCTION                                             1

  2          SUMMARY                                                  3

  3          CONCLUSIONS                                               7

  4          RECOMMENDATIONS                                        9

  5          SYSTEM REQUIREMENT DISCUSSION                        11

             5.1   System Performance Goals                            11
             5.2   Design Approach                                      12

  6          CONTROL SYSTEM                                        13

             6.1   Air  and Fuel Control System Design                     13

                   6.1.1   System Analysis                               13
                   6.1.2   Air Metering Valve Analysis                    16
                   6.1.3   Fan Selection                                  20
                   6.1.4   By-Pass Valve Analysis                        20
                   6.1.5   Fuel Pressure Regulator Analysis               23
                   6.1.6   Fuel Metering Valve                            30

             6.2   Control System Component Tests                       31

                   6.2.1   Fuel Metering Valve  Calibrations                33
                   6.2.2   Air Metering Valve Development Tests           36
                   6. 2. 3   Delta-P Fuel Pressure Regulator Calibrations    51

             6.3   Transient Response Emissions Test                     53

                   6.3.1   Description of Demonstration Combustor System  53
                   6.3.2   Startup and Shutdown Transients                 58
                   6.3.3   Power Level Transient Emissions               59
                                      111

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                             CONTENTS (Cont)


Section                                                               Page

  7          COMBUSTOR DISCUSSION                                   63

             7.1  Reaction Kinetic Study By Computer Modelling            63

                  7.1.1  Summary                                      63
                  7.1.2  Emissions Analysis                             63
                  7.1.3  Initial Calculation                              64
                  7.1.4  Combustor Design by Computer Modelling         70

             7.2  Combustor and Test Rig.Design                         89
                  7.2.1  Combustor Design                              89
                  7.2.2  Combustor Test Rig  Design                      93
                  7.2.3  Instrumentation                                96

             7.3  Combustor Development                              100
                  7.3.1  Summary                                    100
                  7. 3. 2  Preliminary Combustor Rig Tests              100
                  7.3.3  Combustor Pressure Loss Reduction            101
                  7.3.4  Simulated Vaporizer  Tests                     105
                  7.3.5  Combustor Tests With Fan                     107
                  7.3.6  Final Combustor Tests                        116
                  7.3.7  Combustor Noise                             121
                  7.3.8  Ignition Tests                                125
                  7.3.9  Aldehydes and Smoke

  8          OPTIMUM DESIGN APPROACH FOR RANKINE CYCLE
             COMBUSTION SYSTEM                                    127

             APPENDIX A - Emission Monitoring Equipment and
                           Procedures                                131

             APPENDIX B - Test Fuel Specifications                     141

             APPENDIX C - Fan Noise Reduction Methods                147

             APPENDIX D - Emission Data Reduction                     159
                                     IV

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                             ILLUSTRATIONS

Figure                                                                 Page

  1         Control System Schematic                                     14

  2         Air Metering Valve                                           17

  3         Air Metering Valve Port Shown in the 60 Percent Power Position   17

  4         Fan Characteristics (All Curves at 27V Input Except as Noted) -
            Joy Model Number P/N 702-93                                 21

  5         Combustor and By-Pass Valve Pressure Drop Versus Power
            Demand                                             •         22

  6         Dynamic Head in Annulus                                      22

  7         Fan and Combustor Flow Versus Power Lever Position            24

  8         By-Pass Flow as a Function of Power Required                   25

  9         By-Pass Valve Area Versus Power Required                     26

 10         Fuel Metering Valve Concept for 100:1 Turndown                 31

 11         Reynolds Number as a Function of Temperature and Orifice
            Size at 1 Ib/hr for Kerosene Through a Square Orifice             32

 12         Discharge Coefficient as a Function of Orifice Size  for Kerosene   32

 13         Fuel Metering Valve Assembly                                 34

 14         Fuel Metering Valve Stroke Measurement Arrangement            36

 15         Fuel Metering Valve Final Weight Flow Calibration               37

 16         Fuel Metering Valve Performance From 0.5 to 10 Pounds
            Per Hour (Final Weight Flow Calibration)                        38

 17         Air Valve and Fan Flow  Test Mock-Up Schematic                 39

 18         Air Valve Flow Test Mock-Up                                  40

 19         Air Valve Flow Test Schematic                                 40
 20         Air Flow Test Bench                                          41

 21         Flow Control Performance of Air Metering Valve                 44

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                           ILLUSTRATIONS (Cont)
Figure                                                                  Page
  22        Air Valve AP Test Results With Mock-Up Air Valve                45
  23        Demonstration System Air Valve                                 47
  24        Sound Pressure Level Versus Frequency for Blower Without
            Inlet Flow Straightener                                         48
  25        Noise Measurement Room Geometry and Materials                49
  26        Fuel Pressure Regulator Installed on P^ - Pg Actuator             51
  27        Fuel Pressure Regulator \Vith Molded 0.012  Inch Thick PVC
            Diaphragm                                                    52
  28        Fuel Pressure Regulator Performance With 0.008 Inch Flat
            Rubber Diaphragm                                              54
  29        Demonstration System                                          56
  30        Demonstration System With Long  Mixing Duct                     57
  31        Integrated System Demonstration  Test Stand Arrangement          57
  32        Startup and Shutdown Emissions                                  59
  33        Power Level Transient Emissions (Combustor Configuration "D"    62
  34        Flowchart, Generalized Kinetics Program                        65
  35        Fuel Droplet Lifetime at Maximum Heat Release Rate (2 x 106
            BTU/hr)                                                       66
  36        Fuel Droplet Lifetime, 50 Micron Drop Size,  at 109, 45.5,
            and 1. 09  IbAr of Fuel                                          66
  37        Equilibrium Flame  Temperature as a Function of Air-Fuel
            Ratio                                                         67
  38        Equilibrium Concentrations by Volume of Carbon  Monoxide
            and Nitric Oxide as a  Function of Air-Fuel Ratio                   68
  39        Primary  Equilibrium  Composition                               69
  40        Reaction Set                                                   71
  41        Design A  - Configuration No.  1                                   75
  42        Design A  - Configuration No.  2                                   77
  43        Design A  - Configuration No.  3                                   79
                                     VI

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                           ILLUSTRATIONS (Cont)

Figure                                                                  Page

  44        Design A - Configuration No. 4                                 81

  45        Design B - Configuration No. 1                                 83

  46        Design B - Configuration No. 2                                 85

  47        Emissions at One Percent Heat Release With and Without a
            Fice Percent Heat Loss For Design A,  Configuration No. 1        87

  48        Emissions at One Percent Heat Release With and Without a
            Five Percent Heat Loss For Design A,  Configuration No. 2        88

  49        Side View of Rotating Cup Combustor Assembly                   90

  50        Front View of Rotating Cup Combustor and Case                  91

  51        Front View of Rotating Cup Combustor and Case With Rotating
            Cup Removed                                                 92

  52        Rear View of Rotating Cup Combustor and Case                   92

  53        Rotating Cup and Motor Assembly                               93

  54        Schematic of Combustor Test Rig                               94

  55        Rear View of Fan and Control Valve Assembly Showing Anti-
            Swirl Plates Installed                                          95

  56        Combustor Rig Showing Arrangement of Air Metering Orifices     95

  57        End View - High Temperature Probe With Triple Radiation
            Shield and High  Velocity Aspiration System                       98

  58        Installation of a High Temperature Thermocouple                 98

  59        Circumferential and Radial Positions of Thermocouple  at the
            Exit of the Combustor                                         99

  60        Schematic of Emission Pickup Probe                           100

  61        Air-Fuel Ratio For Minimum Emissions as a  Function  of
            Combustor Air Flow                                          101

  62        Emissions of Carbon Monoxide  as a Function of Combustor Air
            Flow, at Optimum Air Fuel for Minimum Emissions and Also
            With a ±10% Deviation of Air Fuel From Optimum               102
                                     vn

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                            ILLUSTRATIONS (Cont)

Figure                                                                   Page

  63        Emissions of NO  as a Function of Combustor Air Flow at
            Optimum Air-Fuel For Minimum Emissions and Also With
            a ±10% Deviation of Air-Fuel From Optimum                     103
  64        Effect of Fuel Flow on Emissions at Differing Air-Fuel Ratios
            Using Rig Air Supplies                                         104

  65        Combustor Rig With Vaporizer Installed                          105

  66        Effect of Varying Fuel Flow on Emissions When Air-Fuel is
            Maintained at a Constant Value (26/1) and Using a Rig Air
            Supply and Boiler                                              106
  67        Air Maldistributions Due to Unstable Diffusion                    108

  68        Average Radial Profile of Temperature Out of Combustor Using
            Rig Air  Supply                                                 110
  69        Average Radial Profile of Temperature Out of Combustor Using
            Fan Air Supply                                                 110

  70        Repeatability of Radial Profile of Temperature Out of Combustor
            Using Both Fan and Rig Air Supplies                             111
  71        Radial Profile of Te mperature Out of Combustor at Several
            Different Circumferential Locations and Using the Rig Air
            Supply                                                        112
  72        Radial Profile of Temperature Out of the Combustor at Several
            Different Circumferential Locations and Using the Fan Air
            Supply                                                        113

  73        Circumferential Variation of Combustor Outlet Temperature at
            Different Radii and Using Rig Air Supplies                        114

  74        Circumferential Variation of Combustor Outlet Temperature at
            Different Radii Using Fan Air Supply                             115
  75        Radial Profile at Various Fuel Flows, Using Rig Air  Supply        116
  76        Radial Profile of Temperature (Final Demonstration Compared
            to Preliminary)                                                117
  77        Final Air-Fuel Control System.  Discharge Temperature
            Variation With Fuel Flow                                       118
                                     viii

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                           ILLUSTRATIONS (Cont)

Figure                                                                   Page

  78        Emission of NO9 as a Function of Fuel Flow                      118
                          LJ
  79        Emissions of CO as a Function of Fuel Flow                      119

  80        Emissions of HC as a Function of Fuel Flow (Test A)              119

  81        Emissions at Various Fuel Flows, With Different Fuels and
            at Two Air-Fuel Ratios, With and Without Heat Losses            120

  82        Before and  After Modification Burner Noise at Location "A"        122

  83        Identification of Combustor Noise at Location "B"                 123

  84        Diagram of Combustor Acoustic Test Location                    124

  85        Two Fan Sketches - Present and Optimum                        130

  A-l       Beckman Model 315A Infrared Analyzer                          134

  A-2       Beckman Model 402 Hydrocarbon Analyzer                       135

  A-3       NO  Calibration Results                                         139

  A-4       CO  Calibration Results                                         139

  A-5       CO2 Calibration Results                                        140

  B-l       Variation of ASTM Distillation Temperatures for the Test
            Fuels (Average Values)                                         144

  C-l       Effect of Pressure Ratio on Rotor Alone Blade Passing
            Frequency Noise                                               153

  C-2       Effect of Number of Blades on Rotor Alone Blade Passing
            Frequency Noise                                               153

  C-3       Effect of Pressure Ratio on Interaction Noise Generated at the
            Blade Passing Frequency                                       154

  C-4       Effect of Number of Blades on Interaction Noise Generated at the
            Blade Passing Frequency                                       154

  C-5       Effect of Vane/Blade Ratio on Interaction Noise Generated at the
            Blade Passing Frequency                                       156

  C-6       Effect of Spacing on Interaction Noise Generated at the Blade
            Passing Frequency                                             15G
                                     IX

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                                   TABLES
Table                                                                   Page

 I           Fuel Metering Valve Weight Flow Calibrations                  35

 II           Air Control Valve Test Results With 2.25 Inch Wide By-
             Pass Valve (0.5-Inch Band on Orifice Side of By-Pass)          43

 III          Pressure Regulator Calibration With 0.008 Inch Flat
             Rubber Diaphragm                                           55
 IV          Power Level  Transient Emissions (Combustor Configuration
             "D")                                                        60

 V           Significant Noise Frequency Peak Levels                      124
 A-I         NDIR  Compared to Saltzman                                 136
 A-II        Reproducibility Test                                        137

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                                      1
                                INTRODUCTION
        Much of the air pollution in the United States is a result of undesirable
emissions from automotive Otto cycle engines.  Each day thousands of tons of toxic
gaseous and particulate pollutants are dumped into the limited volume of the atmos-
phere over the United States.  Because of the highly complex and transient nature of
combustion combined with the cold walls of an automotive spark ignition engine,
direct control of emissions in the combustion process is  a difficult and possibly
insoluble problem.  Control by  various post combustion schemes appears to offer con-
siderably more potential.  Several methods have been demonstrated and show promise.
However,  potential maintenance,  service life, and high cost problems make it nec-
essary to also consider alternatives to the internal combustion engine if these methods
fail to achieve the goals.   Systems utilizing continuous flow external combustors have
been demonstrated to have low emission levels under steady-state operation. A
Rankine Cycle engine has  one of the best potentials of any of the continuous combustion
engines for automotive application.  The EPA is presently supporting basic research
studies to optimize the low pollution featui es  of burner designs for Rankine engines.

        The basic problem  resolved by this program was the demonstration that a
Rankine cycle combustor system designed to be integrated into a "family car" could
meet the 1980 Advanced Automotive Power  Systems (AAPS) goals.  A major portion
of the effort for such a system has been devoted to the development of a full  scale
(2,000, 000 BTU/hr) prototype combustor system and controls with the desired low
emission characteristics.  Automotive requirements for  duty cycle, compactness,
cost and efficiency are the constraints that  challenge the  present state-of-the-art for
combustors.  Rapid starts,  high response,  frequent shutdown, with large and frequent
transients of power level demand, are required automotive performance factors that
cause existing combustor  designs to fall short of emission goals.

        This program addressed itself to these problems by applying a new and novel
fuel atomization and precise air-fuel ratio control concept to the automotive Rankine
combustion system problem. Rotating cup  atomization with a full range air-fuel ratio
control is ideally suited to the wide fuel flow variations necessary for automotive duty
cycles requiring low emissions because of all the various methods  of fuel injection,
it has the widest possible  range of operation with a single fuel atomizer.  Experience
has indicated a major portion of the emissions result from load  transients in conven-

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Lional combustors.  The problem of making an on-off system emission-free is tech-
nically difficult, if not impossible.  A combined on-off and modulated system share
some of the problems of both and has greater complexity. Thus, this program
developed a fully modulated system incorporating a rotating  cup  fuel atomization
system and all necessary controls to supply and regulate both fuel and air at high
response rates while maintaining the optimum air-fuel ratio for  lowest emissions.

        The purpose of this program was to apply modern analytical and experimental
tools to the design,  fabrication,  development and demonstration  of a low emission
combustor system for an automotive Rankine cycle engine.  The  result of this study
has been two-fold.  First, the capability of meeting the emission goals has been
demonstrated.  Second, areas, which require further development in order to make
the system usable in an automobile, have been identified.

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                                     2
                                  SUMMARY
        The primary objective of this program was to apply modern analytical and
experimental techniques to the design and demonstration of a low emission combustor
for an automotive Rankine engine.  Results would show if the 1980  AAPS emission goals
could be achieved and determine technology requiring advancement to equip a "family
car" with a low emission combustor.  Excellent progress had previously been made
with  combustors to obtain low unburned hydrocarbons (HC)  and carbon monoxide (CO)
at rated steady state operation but they could not meet the nitric oxide (NO) emission
levels nor could they operate over the wide fuel flow range  and rapid transients required
for the automotive Rankine cycle engine.

        Two approaches are open for emission control from automobile engines —
limit NO formation during combustion or eliminate NO after it is formed. Spark
ignition engines can generate reducing exhaust gases and therefore eliminate NO by
catalytic reduction with CO.  Rankine  cycles,  gas turbines, and  diesels cannot run
fuel  rich because of smoke production, inefficiency and overtemperature, and there-
fore cannot generate the necessary reducing exhaust gases. Thus the limit NO for-
mation approach must be employed in  the combustion zone.

        The Solar approach has been  to obtain low NO emissions through air-fuel
ratio-control in combustion  zones.  A schematic shows the air-fuel ratio  control
and  residence time for the primary, secondary, and tertiary zones of the combustor
which were demonstrated to perform well within the established  1980 AAPS emission
goals.

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              872 LB/HR AIR
             436 LB/HR AIR
SOLAR-LOW EMISSION-CONTROLLED ZONE-COMBUSTOR
              654 LB/HR   872 LB/HR
                AIR         AIR
                ^
        FUEL
       DROPLET
       PATTERN
       109 LB/HR
         FUEL
V V
SECONDARY
ZONE
18 A/F ^
TOTAL W
1962 LB AIR
109 LB FUEL
V. J

If
TERTIARY
ZONE
26 A/F ^L
TOTAL V
2834 LB AIR
109 LB FUEL

> , i 1
                         0.0      0.002    0.004     0.006    0.008
                            RESIDENCE TIME (SEC) - 109 LB/HR FUEL
                                         0.01
The rotating cup atomization system yields precision fuel droplet size and pattern
independent of fuel flow.  Precision air and fuel metering values allows scheduling
of the desired air-fuel ratios over the 100 to 1 heat release estimated as a maximum
range necessary to cover all engine systems.

         The volume of the demonstrated  combustor would be 1.1 cubic feet for the
optimum design which included an optimum fan, demonstration controls and combus-
tor.  This corresponds with an initial goal of 1.3 cubic feet.  The parasitic power is
computed at 1.25 horsepower without the  vaporizer and can be substantially reduced
at part load with minor modifications of the demonstrated control system. Fan and
combustor noise was  above acceptable limits during many test conditions. An
analysis of the noise indicated that noise could be reduced to acceptable limits by use
of special design approaches for both the  fan and combustor.

         A fully modulated air-fuel control system has been designed and was demon-
strated  in integrated combustor emission tests. All major components of the system
except the fan were specifically designed  to obtain accurate control of air-fuel ratios.
Design emphasis has  also been placed upon high response capabilities necessary for
low emission city traffic operations  of automotive vehicles.  Stop and start duty cycles
require an almost continually fluctuating heat release rate from the combustor system.
On-off control modes, although simple, were rejected because of their potential for
higher emissions during the frequent start-up and shutdown transients required in
automotive Rankine engines.  The system developed basically consists of a variable
area air valve mechanically linked to a variable area fuel valve.  Integrated with the

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basic area control is the air supply (electric fan) and the pressure control system (air
by-pass valve).  A fuel supply (electric pump) and fuel pressure regulator subsystem
are incorporated in a manner that allows for compensation of fan motor voltage re-
ductions or other changes that tend to vary the desired air-fuel ratio.  High response
to power level demands is provided, while a prescheduled air-fuel ratio is maintained
to minimize emissions.

        In both the air and fuel metering systems, a constant (or proportionally
controlled) pressure differential is maintained across the variable area valves.  Thus
the 100 to 1 flow range is obtained by providing a 100 to 1 area ratio in each valve.
A low fuel  supply pressure,  10 psig, has been selected to minimize the system's
sensitivity to contamination and allow the use of inexpensive pumps.  A fuel pressure
regulator is integrated into the system to allow the air-fuel ratio to be essentially
independent of fan motor voltage (or fan efficiency, leakage, brush wear,  etc.).  An
important feature of incorporation of this fuel pressure control scheme is that part
load parasitic power demands of the fan motor can be reduced by as  much as 65 per-
cent by scheduled voltage reductions while still maintaining the air-fuel ratio  indep-
endent of the inertial lags of the fan motor.

        A series of demonstration tests at various steady state and  transient power
levels were performed on the integrated combustion system.   Emissions monitored
during transients and indicated that the peak levels of emissions would remain below
the limits at a transient rate  of 50 percent power level change per second (54  Ibs/hr
per sec).   In most transient ranges the rate of change of power level could go as  high
as 150 percent per second without significant emission peaks.   However, decreasing
power transients below 30 Ibs/hr at rates of slightly above 50 percent per second
were observed to produce emission peaks of CO above the limits for three or  four
seconds.  Startup from cold to maximum firing rate in three seconds or less appeared
to present no significant emission peaks.   Steady state emission levels were in
general significantly below the 1980 AAPS goals.  Measured values across the entire
heat release range were on an average 61 percent below the goal for CO, 90 percent
below the goal for HC and 37  percent lower in the  case of NO.   The only condition that
caused an emission level above  the goal was at 1 pound per hour fuel flow.  At this
low flow condition CO levels were above the limit.  It should be noted that 1 pound per
hour fuel flow is an extreme boundary normally  outside the range of  typical engine
systems.

        Basically the  component and system demonstration tests have indicated the
system can provide the necessary degree of air  and fuel regulation for a low emission
combustion system that has imposed on it a wide heat release  range  and frequent
high response power level changes.  Construction, size, weight,  and reliability
features of the system have the  potential to be incorporated into inexpensive mass
production automotive  engines.

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                                     3
                                CONCLUSIONS
        This program has demonstrated that a Rankine cycle combustor with a two
million BTU per hour heat release in a 1.33 cubic foot volume operating on JP-5 fuel
can meet the 1980 AAPS emission level goals over a 100 to 1 heat release range at
steady state or during rapid transients.  The transients include startup to full power
in three seconds and  a 50 percent power rate change per second.

        The program has further demonstrated that the package size and power
requirements of combustors need not be excessive.  The components used are not
complex and are capable of being mass produced at low cost. All of the components
employed proven concepts that appear capable of a high degree of reliability.

        The novel fuel atomization system assures that the fuels can be rapidly
ignited in even the coldest weather and does not need warm up to maintain low emission
as do the current spark ignition engines.  Further the fuels used need no special
additives for combustion control and a wide variety of currently available fuels  can  be
used (however, leaded gasoline is not  acceptable).

        The novel rotating cup has been shown to be essential for low emission control
over the 100 to 1 heat release range.   The precision air and fuel metering valves and
controls have also been shown to be essential to maintain the low emissions  over the
100 to 1 heat release range.

        Fan and combustor noise was shown to be a problem,  thus  special attention
must be given  to its elimination in future development programs.

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                                     4
                             RE COMME NDATIONS
        The demonstrations were conducted to show that given certain boundary
conditions, such as heat release range, fuel and emission levels, a Rankine cycle
combustor could perform satisfactorily and provide incentive for further Rankine
cycle engine development. All major components except fan and cup drive motor
were specially designed and developed to conduct this investigation.  Further develop-
ment is necessary, especially in regard to integration with other engine components
because they effect both the shape and final performance of the combustor. The
principal component having the greatest effect on emission levels will be the vaporizer.
The temperature boundary conditions on the combustor will be altered by the addition
of the vaporizer and thus  need special development.  Further the emission levels will
be altered because of boundary conditions, heat up times required,  response time, and
stability of the vaporizer. All of these factors must be  considered when coupling the
vaporizer with the low emission combustor.  The work conducted on the program has
demonstrated that these problems are  solvable and that  a combustor with emissions
well under the goals can be expected.

        Minimum fan power must also be demonstrated for the optimum design
Rankine cycle combustor. Fan power,  system volume and uniform air distribution
(air-fuel ratio control)  require the design of a fan that is aerodynamically and geo-
metrically integrated with both the air valve and combustor.  The fan motor repres-
ents one of the largest inherent cost factors in the Rankine cycle combustor control
system.  Minimum fan power becomes a prominent factor in a continued development
program.  Elimination of a separate cup motor by using an extension shaft on the fan
motor will also be necessary to  minimize cost and volume.  Additional development
testing is  necessary to match the cup and fan speeds  to obtain satisfactory atomization
and fan  aerodynamic design at the same shaft speed.   The combustor pressure drop
must be held at a minimum.  However, if the pressure drop is too low emissions and
temperature distribution  into the vaporizer and exhaust  system will become a problem
and can cause hot spots and low  reliability of the vaporizer.  Further test analysis,
using a  vaporizer and optimum fan should be conducted to  establish engineering require-
ments necessary for minimum emissions with a reliable automotive vapor generator.

        Vaporizer air side pressure drop must be held  at minimum to conserve fan
power.  However, the weight and response time for the  vaporizer becomes excessive

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if the pressure drop is too low.  An optimum balance will require an integrated
development program.

        Transient performance of the combustor-vaporizer system must be further
demonstrated.  One of the critical items is the start and warm up period to roadload
power of the complete Rankine cycle engine.  Time must be maintained at a minimum
and therefore will require optimum performance of each component during this period,
while still maintaining minimum parasitic power.  Further development and  integrated
system tests are essential to demonstrate transient and warm up characteristics with
a vaporizer.

        A final but important consideration is the development and demonstration of
an integrated vapor generator and fuel control system.  Response, temperature
distribution, heat flux, wall temperature, flame radiation, thermal inertia,  feed
control, fluid hot spots, and stability must be considered in a high response  control
system.  It is  necessary to demonstrate that fuel and air flows can be made to respond
rapidly and accurately enough to regulate the vapor pressure and temperature within
acceptable limits as vapor flow rates respond to automotive duty cycles.  A low cost
control system must be developed without sacrificing the inherent low emission
characteristics of the combustion system.
                                      10

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                                     5
                     SYSTEM REQUIREMENT DISCUSSION

        The purpose of this program was to design,  fabricate and test a full scale
combustion system suitable,  with minor modifications, for installation in a Rankine
cycle engine powered vehicle.  Included in the system was the combustor,  fan,
controls, ignition system and fuel pump.

5.1  SYSTEM PERFORMANCE GOALS

        The performance goals of the combustion system are as follows:

           •Maximum heat release of 2,000, 000 BTU/hr (which corresponds to
            109 pounds per hour fuel flow) and a minimum heat release of 20, 000
            BTU/hr which results in a 100 to 1 turndown ratio. The maximum
            heat release was  stated as a time averaged rate, thus permitting use
            of an on-off or modulating system or a combination of both.

            •Steady state emissions - 1980 AAPS goals

                                                        Corresponding gram
                                       gm Pollutant       pollutant per mile
                   Pollutant          Kilogram of Fuel   (assuming 10 mpg)
             Carbon monoxide              16.25                4.7
             Unburned hydrocarbons
             reported as CI^ 85             0.48                0.14

             Oxides of nitrogen
             reported as NO2                1.38                0.4
             Participates                   0.10                0.03

         In addition no visible smoke is allowed at any operating condition.  Meeting
 these goals results in high combustion efficiency,  which was also listed as a
 goal.
                                      11

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           • The steady state emission goals are to be met within 2 seconds after
            any  change in fuel flow.  There is to be no severe degradation of
            emission levels during a transient.

           •Startups and shutdowns are to be as clean as possible since they are
            included in the  latest automotive emission test procedure.

        The combustion  system is to  have the following constraints:

           •Maximum combustor volume - 1.33 cubic ft.
           • Maximum parasitic power - 2. 0 HP (without the vaporizer)
           •Test fuel - Kerosene, #1 Diesel, Jet A
           • Test conditions - ambient pressure  and temperature

5. 2  DESIGN APPROACH

        Reviewing the requirements the major problem was identified as maintaining
low emission levels over a wide range of heat release  rates or fuel flows.  To obtain
low emission levels of CO, HC  and NO simultaneously involves a careful trade-off of
combustor parameters  such  as  local and overall air-fuel ratios, peak temperatures,
residence time and velocity.  In general me factors which contribute to low NO promote
high CO and HC and vice  versa.  This will be discussed in greater detail in Section 7.  One
conclusion established was that me air-fuel ratio must be closely controlled over the
entire operating  range.  The air control valve and fuel control valve are  linked to-
gether in such a  way as to provide the  desired overall  air-fuel ratio at each fuel flow.
The local  air-fuel ratios  are determined by location, size and number  of the combus-
tor air entry holes.

        Based on published  data and our own experience that startups  and shutdowns,
i.e., turning the  combustor  on and off, would result in substantial emissions, it was
concluded that the system must be completely modulating over the entire range. In
order to accomplish full modulation a fuel atomizing device which would atomize the
fuel adequately at flow rates  from 1 to 109 pounds per  hour was required.  Not only is
the wide turndown ratio a problem but  so is adequate atomization of a fuel flow as low
as 1 pound per hour. Our investigation indicated mat the rotating cup fuel atomizer
was the most likely device to fulfill the requirement.

        The major components of the  demonstration system are the air supply system,
air control valve, fuel control valve, fuel pressure regulator valve,  and  the combustor
with its rotating  cup fuel  atomizer.  The selection and/or development of these com-
ponents is discussed in the following sections.
                                      12

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                                      6
                               CONTROL SYSTEM
6.1  AIR AND FUEL CONTROL SYSTEM DESIGN

        Low emissions were the main design criteria used to synthesize the control
system.  Analysis and experience determined that a fully modulated control of both
air and fuel across the full range of heat release rates would  offer the optimum low
emissions approach.  Inherent with this  design concept was the  necessity to maintain
the ideal air fuel  ratio at any heat release rate from one percent to 100 percent.  In
addition to  the wide range of heat release rates necessary in automotive systems, the
very frequent and high response demands for power level changes dictates that air-
fuel ratio control must be free of fan motor  inertia time lags. A fully modulated
control system independent of inertial lags of fan  motor was designed with simple
hydromechanical  control components that have the potential of being incorporated
into low cost automotive type engine systems.

6.1.1  System Analysis

        Experience  and analysis has shown  that the optimum air flow control and
combustion air delivery system requires a high degree of symetry to provide uniform
flow (and consequent correct air-fuel ratios) into  the  combustor.  Figure 1 schematic-
ally describes the overall air and fuel control system.  Air mass flow is controlled
by varing the flow area of 12 orifice ports leading directly into the combustor outer
casing.  To ensure that fuel flow is directly  proportional to air  flow, a mechanical
coupling syncronizes the control of fuel flow area to air flow area. Once the area
ratios have been fixed, it is only necessary to regulate pressure drops to maintain
the desired weight flow ratios.

        For low  pressure rise, air compression can be neglected and the weight
flow relationships can be written as:
                                       13

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                   BYPASS VALVE
                                                                              . J
ilX.' -V., lii^CiiU.          BAFFLE
                                          POWER   ,
                                          LEVER   V
      ±1.0"
FUEL METERING
VALVE 1.0 to 109 LB/HR
                                          ^  -g_.
                          .\\\\ A\\\\\\\\\*A\\Wp]
               FIGURE 1.   CONTROL SYSTEM SCHEMATIC

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                                                                          (2)


where       W  =  mass air flow into combustor
              a
            A™ =  air metering valve area

             PI =  upstream pressure (fan discharge)

             P9 =  downstream pressure
              u
             T  =  air temperature at metering valve

             Wf =  fuel mass flow

            APf =  pressure drop across fuel valve

             AT =  fuel metering valve area

If the inlet pressure (altitude) and temperature are constant, a simple relationship
can be written for the air-fuel ratio:
                                                                          (3)
        To ensure the air fuel ratio remains a function of flow area only, the pressure
drop across the fuel valve is controlled proportional to the pressure drop across the
air metering ports.  A APr regulator valve performs this function by a force balance
across diaphragms.  Operation of this component is illustrated in Figure 1.  P^
pressure is connected (through a density compensator) to the bottom side of diaphragm
A^.  Downstream pressure (?2) is connected to the opposite side of the diaphragm.
Fuel system pressure is regulated by a flapper valve that by-passes excess fuel to the
tank.  A small diaphragm on the fuel side balances the pressure difference across the
air metering ports Aj^ against the fuel pressure.  If the  air pressure drop increases
(causing Wa to increase), the force  across  the system becomes unbalanced and the
fuel pressure regulator moves up; reducing the by-pass flow,  and thus increasing the
fuel pressure and its mass flow across the  fuel metering valve.  Since the force
balance must be maintained across the 4Pf regulator, we have:

             ZForces = Ag (4Pf - PZ> =  \(P±-'?2)

where P.J = P, adjusted for inlet air density and where dP* is large compared to  P2
we have      AP  = -  (P ' - P )                                          (4)
               I    A    1    2
                                      15

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thus equation (3) can be rewritten as
                            - P2> A2
                                   " =  constant
        Fan speed changes due to voltage or load variations, fan efficiency reductions
due to fouling or wear, and by-pass valve leakage variations are automatically com-
pensated by maintaining the APf as a function of the air valves pressure differential.
        It can be seen from equation (1) that the air flow is a function of
              T!

Thus, an altitude and inlet air temperature correction can be incorporated by adding a
PI/TJ compensator (Fig. 1).  Fan discharge pressure (P,) is admitted to the AP
regulator by a contoured needle valve.  This valve is positioned by a force balance
across a diaphragm having P , T^ on one side and cavity sealed with air at standard
conditions.  Thus, as the air temperature  increases, the sealed air will expand and
move the valve upwards. This action will  increase the pressure drop across the con-
tour valve and thus reduce P,  to a corrected value of PI on the bottom side of
diaphragm A.  As P^ decreases,  the fuel pressure will drop producing the desired
fuel flow reduction as the ambient air temperature increases.

6.1.2  Air Metering Valve Analysis

        Air to the combustor is metered across a series of twelve ports cut into two
plates (Fig.  2).  An input lever rotates one plate with respect to the stationary backup
plate.  The metering port areas are caused to open or close as the input power lever
rotates matched  ports in each plate.  Figure 3 shows one of the  12  port configurations.
The area is established  by maximum mass  air flow requirements of the system:

            Maximum Power =  2 x 106 BTU/hr.

Let         LHV of fuel =  18, 350 BTU/lb for JP-5
             .'. fuel flow =  109 Ibs/hr or 1.82 Ibs/min.

            Air fuel ratio at maximum rated heat release  = 26 (See Sec. 7)
            .'. Maximum Air Flow =  (26) (1.82) =  47. 3 Ibs/min  =  630 SCFM
                                      16

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             FIGURE 2.  AIR METERING VALVE
  OPENING IN
METERING PLATE
     ROTATION
     TO OPEN
                                             OPENING IN
                                           BACKUP PLATE
    FIGURE 3.  AIR METERING VALVE PORT SHOWN IN THE
              60 PERCENT POWER POSITION
                           17

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The area to flow this quantity of air depends upon the selection of a design pressure
drop across the metering valve.  It is desirable to maintain this  loss as low as possible
to reduce parasitic power requirements.  However,  very small pressure differences
require larger and heavier valves.  A tradeoff analysis indicated that a AP of 2 inches
of H2O results in a valve size that is  within the dimensional envelope of the combustor:

        Flow relationship =
where
Allow

Assume
For
AM =  in.   (area of orifice)
  y =  lb/ft3 (density at standard conditions)

Wa =  Ibs/min.

  P =  in. ofH2O

AP =  2 in.  of H2O

  K =  0. 6 discharge coefficient

  y =  0. 075 and a maximum flow rate condition of 47.3 Ibs/min.
           47.3  = 7.617(0.6) AM  ^(0.075) 2
            AM  = 26.8 in.

Since a constant AP will be maintained, the area will be varied proportional to the
power lever position, X, (1 to 100%) modified to allow leaning of air-fuel ratio at low
heat release rates (Fig. 3).   (Combustor tests had indicated leaning out was necessary
for minimum emissions.)

Delta-P, Voltage Compensation System

        The pressure differential across the fuel metering valve can be regulated to
maintain the correct air-fuel mass flow ratio regardless of inlet air temperature or
absolute pressure (altitude).  Initially the system was designed to respond to relatively
small variations of voltage ±10 percent (equivalent to ±10% flow variation from the fan).
A regulator capable of compensating for voltage changes of ±10 percent was considered
adequate for a simple constant voltage  motor concept. As component testing progressed
it became evident that modifications to the regulator could extend its range sufficiently
to accommodate a two speed fan.   By incorporating this into the  system, a 65 percent
reduction of parasitic power at normal driving ranges can be achieved along with a
substantial noise  level reduction.  This modification was incorporated and tested at a
                                       18

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component level in the delta-P regulator.  However,  since a two speed fan system
requires a redesigned by-pass valve,  program schedules required that the emission
demonstration tests be conducted with the single speed system.

        An analytical estimate of the part load power reduction presently possible can
be made from component test results.  Development  tests on the regulator valve have
shown it capable of accurately regulating the fuel pressure directly proportional to
pressure differential at the air metering valve across the range of 1. 0 to 2. 5 inches of
H2O.  This wide range control allows the fan motor voltage to be reduced as a function
of power lever position.  If the air by-pass valve is maintained closed for the power
lever position range from 100  to 70 percent, the air flow into the combustor can be con-
trolled by proportionately lowering the voltage to the  fan while the regulator compensates
for the reduced pressure differential across the  air metering valve.  Since  a mechanical
linkage  maintains a known ratio of air valve area to fuel valve area, the regulator will
automatically maintain the air-fuel ratio as the power is reduced.  A limitation  on this
system  is the lowest pressure differential signal that can be utilized as an actuator input
to the simple diaphragm regulator.  Analysis and test results have shown that 1.0 inch
of H2O is a reasonable lower limit.   Using this value, the part load power demands of
the combustor system can be established.   Although the combustor requirements are
moderate (1.25 HP for the optimum configuration and 2.3 HP for the  demonstration
system), addition of a boiler can make parasitic power  losses at part loads  unacceptable.
If a boiler had a high delta-P (as recommended by some designers*), an additional 2. 4
horsepower would be required. Total parasitic power levels would then be  as high as
3. 65 horsepower.  If the system were to require this high power level from full power
down to idle condition,  a highly undesirable condition would exist.  Lowering the input
voltage  across a voltage range that allows accurate AP compensation  by the  regulator
will eliminate the motor inertial speed lags since the air-fuel ratio changes are con-
tinually maintained by the regulator.   Power reduction by operating at 1. 0 inch  of H9O
                                                                              ^
can be estimated by simple calculations  based on fan  laws for a series wound motor.
For voltage changes of approximately 2  to 1, the following relationships give reasonably
accurate results.

        For a given voltage change V^ to V2

             flowQ:       Qj/Qg  =  V1/V2

             pressure p:
             fan BHP is:   BHP,/BHP9  =  (V,/V9)3
                              1     £•      ±6

        By maintaining the by-pass valve closed and reducing the voltage to the motor,
the flow through the metering valve will drop as a function of both the motor voltage
and the area change of the metering valve.  Pressure will drop as a function of the square
*  "Condensers and Boilers for Steam-Powered Cars: A Parametric Analysis of Their
   Size, Weight, and Required Fan Power", by William C. Strack,  NASA TN D-5813,
   May 1970.
                                       19

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of flow.  By using 1.0-inch as the lower limit of adequate control,  the by-pass valve
can be maintained in a closed position until the voltage is reduced to  Jl/2 or O.TOTV^.
Since power is approximately a cubic function, we will obtain (0. 707)  (BHP), or a
65 percent reduction in power.  With the optimum system and a relatively high air
side boiler pressure drop,  a parasitic power loss reduction of approximately 2.4 horse-
power occurs.  Total power with this high pressure drop boiler would be  approximately
1.25 horsepower  at vehicle power demands below 35 percent.  One problem with this
approach is that it is more difficult to make the system linear.  However, since this
is not a driver input command, linearity is of secondary importance if the correct
air-fuel ratios can be maintained.

6.1.3  Fan Selection

        Fan pressure requirements are established by the selected metering AP of 2
inches of F^O added to the combustor maximum rated pressure drop of 8  inches of
H^O. A number of fan designs were investigated but program schedules required  the
use of an off-the-shelf unit.  A Joy P/N X702-93 was selected based upon its immediate
availability and performance.  To match the fan's output characteristics to the com-
bustor flow and pressure requirements, the fan was operated above its design voltage
level.  In Figure  4 the characteristics at both design (27V) and the  increased voltage
(33V) used to match the fan indicates a significant electrical  input power increase.
Final electrical power inputs were approximately 2.4 HP at the less efficient 33V
overload operation point.  This power includes all losses associated with  distribution,
air turning, air valve friction and combustor.  This power can be reduced by a large
percentage by use of a fan specially designed and matched to the aerodynamics of the
air valve and combustor.  It should be noted that in an integrated system, the
pressure drop across the vapor generator and its associated ducting would also have
to be added to these pressure rise requirements.  A general characteristic of fans
is to increase flow as the output pressure decreases.  In order to meter the air from
100 to 1 percent at a high response rate independent of motor inertia! effects, a by-
pass valve has been incorporated.  It allows excess  fan air at low power settings to
be returned to the atmosphere  thereby establishing the correct pressure character-
istics across the  air metering valve and combustor.

6.1.4  By-Pass Valve Analysis

        The by-pass valve is designed to regulate the delta-P across the  air
metering valve at a constant 2  inches of HgO regardless of power lever  settings (1 to
100% range). As  the power is  reduced from 100 percent,  the required combustion
air flow must be reduced proportionately.  The corresponding pressure drop across
the combustor varies as a function of flow  squared as shown in Figure 5.  Since the
by-pass valve is designed to maintain a constant air delta-P of 2 inches of ^O
across A^j, the pressure drop across the by-pass valve is equal to the combustor
                                      20

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                                       ELECTRIC INPUT (HP)
 1.4
                     200
                               400
 600       800
VOLUME IN CFM
                                                           1000
1200
1400
       FIGURE 4.   FAN CHARACTERISTICS (ALL CURVES AT 27V INPUT
                    EXCEPT AS NOTED) - JOY, MODEL NUMBER
                    P/N 702-93

pressure drop plus 2 inches of H-O (shown in Fig. 5).  As the back pressure on the
fan is reduced, its output flow increases (Fig. 4) and must be accommodated in the
design of the by-pass valve ports.  This increasing flow characteristic causes a
significant increase in the dynamic head in the fan annular flow passage (Fig. 6).
A baffle mounted immediately downstream of  the fan discharge is designed to reduce
the effects of the dynamic head by turning and directing the flow into the larger
plenum-like space containing the valve elements.

        By  means of this baffle arrangement, it was possible to establish the by-pass
valve size as a function of static pressure rise.
                                      21

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                                            10 Inches
                                              4P - (P1 - Pg) - 2 Inches
                                         AP
                                           COMBUSTOR
          20
FIGURE 5.
  30  40  50  60  70   80   90  100
  X  POWER LEVER POSITION,%
COMBUSTOR AND BY-PASS VALVE PRESSURE DROP
VERSUS POWER DEMAND
    Q
    u
    X
    U
    i
    z
    B
       600
                     800
                                          1100
                                                 1200
           FIGURE 6.
               800      1000
            VOLUME FLOWRATE, CFM

           DYNAMIC HEAD INANNULUS
                                                       1300
                              22

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        Air flow across the by-pass valve is equal to the difference between fan flow
and the combustor flow (Fig.  7).  A plot of air flow through the by-pass valve (Fig. 8)
is obtained by subtracting the two curves plotted in Figure 7.  By-pass valve area
versus power lever position can now be calculated from the known pressure drop  and
required weight flow. Figure 9 shows the resultant area requirements.  It can be
seen that the area deviates only slightly (~5 percent) from a linear valve flow area
across the entire 100 to 1 flow range.  At the maximum power conditions (X = 100%),
the valve is fully  closed and progressively opens to its full area of 52 square inches at
the 1 percent position.   This  action is exactly opposite the air metering valve sequence,
thus both are mechanically coupled to ensure correct high response matching.

6.1.5  Fuel Pressure Regulator Analysis

        The fuel pressure regulator serves the basic function of  maintaining the
correct air fuel ratio in conjunction with  the fuel and air flow area control valves.
Its incorporation  into the design allows the system to operate across the 100 to 1  fuel
flow range with sufficient accuracy to minimize emissions.  Inertial effects of fan
motor leakage, blade fouling, motor performance degradation, air cleaner blockage,
and pressure gradients across the fan inlet can be compensated as can scheduled
voltage reductions to lower parasitic power demands at reduced loads.  The demon-
stration system incorporates  the basic delta-P  regulation system that allows the valve
to regulate the fuel pressure  in proportion to the volumetric air flow.  This feature
allows compensation for the factors described  above. Although a secondary to the
primary goal of the program, a design has been completed to provide an additional
compensation for altitude and inlet air temperature.  This added compensation has
not been incorporated into the demonstration system, but analysis indicates a simple
mechanical system will perform the task.

Delta-P Compensation System

        Delta-P  compensation can be obtained with a single diaphragm actuated
regulator valve.  Static  pressure upstream (P,) and downstream of the air  metering
valve is sensed and transmitted to either side of the delta-P diaphragm actuator
(Fig. 1).  Neglecting the small correction between P^ and P^' (discussed in the next
section) we can analyze  the action of the delta-P regulator by letting P^'  =  P,.
The differential pressure P^  - P£ across the large diaphragm produces a force
which is opposed  by the  differential pressure across a smaller diaphragm,  one side
of which is subjected to  fuel pressure upstream of the fuel metering valve,  and the
other,  to P2.  Neglecting the  effect of the springs which function only to provide
stability and centering action  of the movable diaphragm assembly, the forces will be
in equilibrium if:

             (Pfuel - P2> A2  '  Al  '
                                       23

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100
 90 h
                      :iV/
                       if

-------
     90 -
            10   20  30  40   50   60   70  80   90  100
               X  POWER LEVER POSITION. %
FIGURE 8.  BY-PASS FLOW AS A FUNCTION OF POWER REQUIRED
                            25

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            100

             90

             80

             70
           ge
           <60
           w
           «
           < 50
           £
           J 40
           UH

             30

             20

             10
                   100% = 52 In.
                      =  CONSTANT  2 INCHES OF HO
                0   10   20  30   40   50   60   70   80
                          X POWER LEVER POSITION %
                                       90   100
         FIGURE 9.   BY-PASS VALVE AREA VERSUS POWER REQUIRED


The system is so designed that fuel delta-P across the metering valve will always
remain relatively high (approximately 10 psid).  On the other hand, fuel pressure
downstream of the metering valve enters the burner through the rotating cup at a
pressure substantially equal  to combustor pressure, Pc.  Thus, the pressure drop
across the metering valve equals:

            APt = Pc  . -  (P.    + P )
               f     fuel     loss    c
where
              loss
= fractional flow loss between fuel valve and combustor.
Since P,    varies from 0 to 0. 5 psi and PC differs from P by approximately 0. 3 psi
(for a low loss boiler),  the total error is 0. 8 psi or ±0. 4 psi.  Since pressure
variation about 10 psig  is ±0.4 psi if bias springs are incorporated flow error is thus
equal to:
                          =  ±2% flow accuracy
                   10
                                      26

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Thus, for purposes of analysis, we can set the equilibrium to:



            4pf = 7T    •
                     Lf

        Fuel entering the small diaphragm cavity leaves through a flapper-type valve
which opens or closes as the diaphragm moves in response to changes in the sensed
pressures, and by passes the fixed-displacement pump output back to the tank,
thereby controlling the fuel pressure so that force equilibrium is always maintained.
To a high degree of accuracy,  therefore,  the fuel delta-P is controlled proportional to
the applied air delta-P,  according to the above equation.  The large diaphragm (Aj)
is 7.15 inches in diameter and has an effective area of 40 square inches. Small
diaphragm (A2) is  0.6 inch diameter  and has an area which is only about 0.28 square
inch.  Thus fuel delta-P will be 140 times as great as the air delta-P.  The air throttle
valve delta-P, P^ - P2, is nominally 2 inches HgO.  With voltage reductions of as
great as 30 percent, the normal range of operation will be 1 to 2 inches of HgO with
corresponding fuel pressures of 5 to  10 psig.

        Caution has been taken in the design to eliminate sources of friction whenever
 possible, even when this is at slight sacrifice in theoretical accuracy.  A balanced
design of the by-pass flapper valve was considered and was rejected because it would
have introduced possible hysteresis.   The design shown i8 not balanced, but the forces
due to the imbalance contribute a flow error of less than 1/2 percent.

Air Density Compensator

        Although not incorporated into the demonstration system, a design  analysis
has been completed that indicates a simple mechanical system could be used to obtain
inlet air density compensation.  The  variable air pressure, P1 applied to the delta-P
control, is modulated by the air density compensator in such a way that the  following
relationship is obtained:
            pi'  - "2  ' K'  r   

A small amount of air at pressure, P-p  sensed upstream of the air throttle valve, flows
through a variable orifice, controlled by movement of a density- responsive diaphragm,
and into the delta-P control valve diaphragm cavity at the pressure P', finally leaving
through a fixed orifice, where it connects to the P2 pressure sense line downstream of
the air metering valve.  This small amount of air which effectively by-passes the AM
is a negligible fraction of the burner  flow, even at minimum power levels.
                                       27

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        For low air delta-P's,  the air flow equations for the first and second orifices
(areas An and AO, see Fig. 1) and steady state continuity of mass required that
                      -T (P  -P')  =   K A   --  ,/(P ' -PJ
This can be solved for P  ' - P  in terms of P  - P  as follows:
P! - P
1 2
1
/A ^
. -^
£ P '
f =1 + iJ

                            A  /  P
                             n'    1
On the basis of absolute pressures,  P..'  is never more than one part in 200 different
from unity, thus we can simplify to:
P ' - P —
Pl P2
1
/A N2
L / o \ . J
(P1 ' P2>
                                    + 1
Therefore, the necessary density correction can be achieved simply by contouring
the variable orifice needle so that
                             = K'
                      + 1
Now, the density compensator consists, essentially,  of a sealed cavity, closed at one
side with a flexible air-tight member, within which is trapped a predetermined weight
of air (neglecting the effect of the spring which serves only to apply a gently loading
action to eliminate diaphragm slack), the pressure inside the sealed cavity will always
be substantially equal to P., applied to the opposite of the diaphragm.  Similarly, for
steady state conditions,  considering the variations in temperature are due almost
entirely to changes in day temperature, the temperature of air inside the sealed cavity
will be substantially the  same as the temperature, T^, of the air flowing across the
opposite side of the diaphragm and surrounding the control.  It will  be required that the
temperature of the air flowing across the diaphragm  be representative of the actual air
                                        28

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temperature at the metering valve.  Although the air is being bled directly from up-
stream of the valve (P, cavity) it may not be of sufficient mass flow to rapidly compen-
sate for hot soak conditions under the hood.  By insulating the sealed cavity and an
additional bleed of air from P^ around the bottom of this cavity, a much faster adjust-
ment to inlet temperature variations could be made without undue error  resulting from
underhood temperature gradients.  The volume of the expandable cavity will be
where R is the gas constant,  and U)is the weight of entrapped air.  Since the volume is
proportional to depth of the cavity, Y, which varies with diaphragm position, we have
             v -


The desired contour for the variable needle is, therefore,  given by

             A
            _ n
             A~
              o
Since density varies only by about 20 percent over the complete range of operation,
the variable orifice configuration can be selected to operate in a fairly linear range.
Once the appropriate contour has been established, any required periodic readjust-
ment of the density compensator is easily achieved by removing the compensator
assembly so as to expose the end of the contoured  needle protruding through the
variable orifice.  Then, while the plug in the sealed cavity is opened, the needle is
held in a prescribed position, depending on day temperature and altitude, and the
cavity is then resealed. With the correct adjustment, the net differential pressure
applied  to the delta-P control valve will be properly compensated for air density so
that fuel delta-P becomes:


                    Ai     pi
            4p(  "T  KT2  

then the fuel-air ratio to the burner will be constant from equations (1) and (2)
                                        29

-------
w
a
Wf
A K
a a A
A(Kf \
/P1P1-P2
' P
1 Pf
AaKa
AfKf
                                                                    - P2>K'

Thus, the optimum air-fuel ratio will be maintained as a function of the area ratios
selected for the air and fuel valves.

6.1.6  Fuel Metering Valve

        Fuel must be metered from 109 to 1. 0 pounds per hour. At low power
settings (1 pound per hour), the flow is approximately two drops per second.  Standard
valves do not have sufficient range to accommodate these severe requirements with
sufficient accuracy for a low emission combustor.  Additionally, the valve should not
be sensitive to  temperature induced fuel viscosity changes.  A new  approach to this
problem has been taken by the application of a dual  slotted  shear valve (Fig. 10).  Two
flat (ground  and lapped) plates with matched contour slots 90 degrees to each other.
At the intersection of the two slots, a square orifice is formed whose area is a
function of the relative position of the top movable plate.  The square shape (and thus
the discharge coefficient)  can be maintained constant throughout the entire 100 to 1
area ratio.  Since  the plates are in contact,  fuel will flow only through the slot in each
plate  and not between the plates ,  thereby reducing the clearance leakage path to the
microfinish  of the  contacting surfaces.  A drain groove is provided between the up-
stream pressure and  the metered outlet fuel passage, thereby reducing the leakage
pressure potential to  the level of  the frictional flow loss to the rotating cup.  Since
this is normally less  than 0. 5  psid, resulting leakage of metered fuel into the drain
system will  be  negligible.

        The size of the orifice slots used is a trade-off between four factors.

            •Fabrication capabilities - requires large dimensions

            •Contamination - requires large dimensions

            • Backpressure sensitivity - requires high pressure  and thus, small
            sizes

            •Temperature sensitivity - it is desired that changes in fuel tempera-
            ture have little effect on the coefficient of discharge.   This factor
            requires a high Reynolds number and thereby small orifice dimension.

                                       30

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                                                             DRAIN TO TANK
      MOVABLE PLATE
                                          FLOW AREA SHADED
      DJLET FROM FUE1
     PRESSURE REGULATOR
                                                SPRING
                                                                POWER LEVER
                                                                POSITION INPUT
         HARDENED AND LAPPED
         SHEAR PLATE SURFACE!
                                    ~0 PSIG
                                                   TO ROTATING CUP
     FIGURE 10.   FUEL METERING VALVE CONCEPT FOR 100:1 TURNDOWN


        Figures 11 and 12 show the effect of orifice size on Reynolds number and
its consequent effect upon the discharge coefficient. These curves indicate the
dimension of the orifice should be approximately 0. 005 inches on a side at the one
pound per hour flow rate to minimize viscosity effects.

        A pressure difference of 10 psi will produce the desired minimum flow in a
0.0057 inch  square orifice and has been selected as the design point. Each slot varies
in width from 0.0057 to 0.057 inch according to a square root contour for flow linear-
ization with  power lever position.

6.2  CONTROL SYSTEM COMPONENT TESTS

        Each  of the control component designs was evaluated in a series of tests.
Development of the units and final calibration was accomplished at a component level
prior to integration into the  complete combustor system for low emission demon-
strations.  Final fuel metering valve performance at 10 psid is 61 pounds per hour
per inch stroke with a linearity of ±2. 5 percent across a range of 3 to 115 pounds per
                                       31

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             0.020

             0.018

             0.016

           3 °.°14
           X
           g 0.012

           g °'010
           oo
           U 0.008

           5 0.006
           o
             0.004

             0.002
               0
                      4000    8000   12.000  16.000   20.000  24,000  28,000
                                  REYNOLDS NUMBER
FIGURE 11.   REYNOLDS NUMBER AS A FUNCTION OF TEMPERATURE AND
              ORIFICE SIZE AT 1 LB/HR FOR KEROSENE THROUGH A SQUARE
              ORIFICE
             0.66 r
             0.64
           H
           u
             0.62
           8
           u
           Ct
           X
           U
           J3
           0
             0.60
             0.58
             0.56
                                 T • 130
                                                      T • 50* F
                                                  - O'F
                     0.004   0.008   0.012   0.016

                           ORIFICE SIZE, INCHES
0.020
          FIGURE 12.   DISCHARGE COEFFICIENT AS A FUNCTION OF
                        ORIFICE SIZE FOR KEROSENE
                                        32

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hour.  Good repeatability and flow control to as low as 0. 5 pounds per hour has
demonstrated the valve has a dynamic range greater than 200 to 1.  Fuel pressure
regulator test results show a useful input actuator range from 1. 0 to 2.5 inches of
H2O while regulating fuel flow within ±4 percent.  Air metering valve performance
calibrations have shown a problem with aerodynamic matching to  the fan discharge
air.  Swirl, dynamic head and turning losses required the incorporation of baffles
and flow straightners.  Although these devices have given  reasonably linear control
down to 2 percent of flow, they produced an undesirably high pressure loss.   For the
final demonstration tests, aerodynamic matching between  the fan  and air valve was
simplified by the use of a plenum created by extending the ducting 36 inches.  Although
this provided good control of air flow down to 1 percent, it also indicated the  import-
ance of correct aerodynamic matching that would be necessary for a minimum volume
system.  An optimum configuration of fan  to air valve design has  been established
from this experience and is discussed in Section 8.

6.2.1   Fuel Metering Valve Calibrations

         In order to obtain linearity and reproducibility  over the unprecedented 100
to 1 dynamic range requirements, a new approach to fuel metering was taken.  To
minimize the effects of air valve  to fuel valve linkage distortions  (which directly
change air fuel ratio) due to actuation load, backlash, or thermal expansion,  the
valve gain was made as low as practical by making the stroke long.  Rigidity  was
also emphasized for both its mounting flange and actuation rod. These features
can be seen in Figures 2  and 13.  The rigid mounting flange with five bolt holes and
the large diameter actuator are shown in this photograph.  Rigidity and freedom from
distortion are essential with a valve that must regulate from 1 percent to 100 percent
flow, since a linkage blacklash or distortion of 1 percent at the low flow end can
change the air fuel ratio by 100 percent. It requires a change of 0. 018 inch for a 1
percent flow variation.  The air valve lever is connected directly to the fuel valve
actuator rod, permitting overall allowances in backlash  to be kept below ±0. 002-inch
(±0. 001 at each pinned joint) or approximately ±10 percent flow at the 1 percent
valve position.  Design analysis indicated  that standard metering  spool,  needle, or
flapper configurations could  not be expected to provide repeatable flow regulation  as
a function of stroke from 109 down to one pound per hour.

         Calibrations were made by a total weight versus elapsed  time to obtain a
true  weight flow measurement independent of viscosity and accuracy errors associated
with  flow meters.  A precision laboratory balance determined the mass of fuel metered
by the test valve with 10 psi  pressure differential across the valve housing. Tests were
                                       33

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                FIGURE 13.  FUEL METERING VALVE ASSEMBLY


conducted with all major components of the fuel system connected as on the combustor
system.  Fuel was pumped by the system's electric driven gear pump to 12 psig.
Using shop air supply a constant 2.15 inches of H^O was maintained across the delta-
P regulator simulating the pressure drop across the air control valves metering
plates.  Differential air pressure across the actuator controls the fuel pressure by-
pass valves position to maintain 10 psid  across the fuel valve at all flows.  A slight
increase in air pressure ( 0.15 inches of I^O  above the nominal 2)  was necessary to
maintain the fuel pressure at a constant  10 psid from 1 to 115 pounds per hour.  All
fuel not being metered by the test valve was bypassed by the regulator back to the
fuel tank. From the discharge side of the valve the fuel passed  either through the
flow meter or directly into a container on the scale.

        Data obtained as a final calibration of me value is listed in Table I.  At 21
different valve stroke positions the weight of fuel that flowed into the container was
determined to the nearest  thousandth of a pound.  A digital timer recorded the
elapsed time within 1  second with the lowest time span being 300 seconds.  The
greatest source of error was associated with the valve position measurement method.
                                       34

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                                   TABL.K 1

            FUEL METERING VALVE WEIGHT FLOW CALIBRATIONS
Test Point
1
1A
2
3
4
5
5A
6
7
8
9
10
10A
11
12
13A
13
14
15
15A
16A
16
Weight
(pounds)
0.267
0.222
1.032
1.695
2.450
2.634
1.426
3.832
6.853
10.688
7.182
8.882
10.518
11.132
6.114
6.934
6.980
7.775
8.556
8.600
9.458
9.488
Elapsed
Time
(sec)
1800
1800
1800
1500
1500
1200
600
1500
1200
1200
600
600
600
600
300
300
300
300
300
300
300
300
Calculated
Flow Rate
(pph)
0.53
0.44
2.06
4.06
5.88
7.90
8.55
9.20
20.56
32.06
43.09
53.29
61.30
66.79
72.10
83.10
83.80
93.30
102.80
103.2
113.6
113.6
Upstream
Pressure
(psig)
10.8
10.8
10.8
10.8
10.8
10.8
10.8
10.8
10.8
10.8
10.7
10.62
10.60
10.62
10.65
10.75
10.70
10.70
10.70
10.75
10.80
10.85
Downstream
Pressure
(psig)
0.80
0.80
0.80
0.79
0.80
0.62
0.80
0.80
0.80
0.80
0.72
0.68
0.62
0.62
0.62
0.75
0.75
0.70
0.70
0.75
0.80
0.85
Temp.
Fuel
oF
76

75
75
78
78

78
77
77
78
77

77
77

78
78
78


78
Valve
Stroke
(inch)
0.006
0.0065
0.031
0.063
0.0915
0.120
0.131
0.141
0.324
0.521
0.704
0.862
1.021
1.096
1.210
1.382
1.397
1.538
1.729
1.746
1.898
1.900
Date
12/15/70
12/11/70
12/15/70
12/15/70
12/15/70
12/15/70
12/11/70
12/15/70
12/15/70
12/15/70
12/15/70
12/15/70
12/11/70
12/15/70
12/15/70
12/11/70
12/15/70
12/15/70
12/15/70
12/11/70
12/11/70
12/11/70
A two inch stroke dial indicator was rigidly clamped to the valve body with its probe
zeroed against a rigid metal plate bolted to the input rod on the valve (see Fig.  14).
A simple screw jack arrangement locked the valve stem  into each test position.
Accuracy of the dial indicator should be well within ±0. 0005 inch but a total hysteresis
of 0. 0015 inch was observed. Backlash within the valve  actuator shaft and in the dial
indicator caused the hysteresis.  This small amount of backlash has no significant
effect on valve performance until flow values of less than two pounds per hour are
being controlled.

        Fuel backpressure was approximately 0. 8 psig throughout the test due  to
the height (approximately 25 inches) that the fuel was required to be raised into the
measurement container.  Inlet pressure was compensated to account for  this head
effect.  Additionally both the inlet and pressure gages were located at the same height as
the fuel valve to eliminate pressure head errors.
                                      35

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         FIGURE 14.   FUEL METERING VALVE STROKE MEASUREMENT
                      ARRANGEMENT
        Data from Table I is plotted in Figure 15 for a range of flows from 0.5 to
115 pounds per hour.  A good degree of linearity is exhibited over this very wide
range.  At high flow the deviation is approximately 2.5 percent below a true linear
operating line.  An expanded scale plot of the data from 0.5 to 10 pounds per hour
(Fig.  16) shows an excellent degree  of linearity across this low flow range.  These
data points determine a linear operation with a slope of 64.2  pounds per hour per
inch of stroke.  The overall slope across the entire operating range is 61 pph/in. as
indicated by the dashed line in Figure 16.  By establishing 0. Oil inch stoke equal
to one pound per hour flow, we can define the valves performance as linear within
±2.5 percent from 3 to 115 pounds per hour with a gain of 61.0 pounds per hour per
inch.  Below 3 pounds per hour the absolute deviation is a maximum of 0.2 pounds
per hour at the lower  range limit (1  pph).  Although small in  absolute terms, it is a
large percentage  of the relative flow. However, since it was repeatable, slight
modifications in the flow area of the air valve allowed correct air-fuel ratios to be
adjusted at low fuel rates.

6.2.2  Air Metering Valve Development Tests

        The air metering valve required more extensive development testing than any
of the other control components.  Figure 17 functionally describes the configuration
of the integrated air valve and fan used for development tests.  Air is pumped axially
                                      36

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               110 -
CO
PRESSURE DIFFERENTIAL:  10 PS1D
FUEL:  JP-S
TEMPERATURE. 76'F
                                                                                                                 ODATA RECORDED 12/11/70
                                                                                                                 • DATA RECORDED 12/15/70
                        0.1     U. 2
                                               0.4     0.5
                                                                     0.7      0.0     u.a      1.0
                                                                                                                          1.4     Li      1.6      I.T
                                                                              FUEL VALVE POSITION (INCHES)

                                FIGURE 15.    FUEL  METERING VALVE FINAL WEIGHT FLOW CALIBRATION

-------
             10.0 -
              9.0 -
                                                    SLOPE • 61 PPH/IN.
                                                PRESSURE DIFFERENTIAL: 10 PSIO
                                                FUEL:  JP-5
                                                TEMPERATURE: 76'F
                                                  O DATA RECORDED 12/11/70
                                                  • DATA RECORDED 12/15/70
                         LINEARIZATION REFERENCE POINT (1.0 PPH AT 0.011 INCHES)
                      0.02    0.04    0.06    0.08    0.1    0.12    0.14   0.16
                                 FUEL VALVE POSITION (INCHES)

         FIGURE 16.   FUEL METERING VALVE PERFORMANCE FROM 0.5
                       TO 10 POUNDS PER HOUR (FINAL WEIGHT FLOW
                       CALIBRATION)


across the fan motor case by a single stage four blade fan operating at 13,000 rpm.
Stationary vanes  were incorporated immediately downstream of the fan blades to
recover the rotational energy in the fan discharge.  A cavity formed by the fan on one
side and the air valve chamber on the other provides a plenum for  air distribution.
As testing progressed, data  indicated that special attention to uniform distribution
of static pressure and velocity gradients  would be necessary to obtain uniform air
flow circumferentially around the combustor.   Flow straightners and baffles were
incorporated at the component test level to reduce swirl and provide uniform static
and dynamic pressure fields across each of the 12 symetrical parts.
                                        38

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                                                           rANHOTM
                                                           coourarM
                                                                             •ICT
      FIGURE 17.  Am VALVE AND FAN FLOW TEST MOCK-UP SCHEMATIC
        A variable geometry sheet metal mock-up of the air flow configuration (Fig.
18) of the valve was tested to establish performance characteristics.  The sheet
metal test valve had all of the functional flow features that were incorporated when the
final design unit was fabricated.  It was fabricated of light gage sheet metal to facilitate
rapid rework for the comparison of different port, baffle, and instrumentation config-
urations.  From the photograph, it can be seen that the valve is in approximately the
10 percent flow metering position since the by-pass ports are approximately 90 percent
open.  Metering is accomplished by the plates immediately downstream of the by-pass
ports. The 12-inch diameter section in the center represents the combustor interface
with the air valve. Figure 19 is the schematic arrangement of the flow test rig.
Actual test component arrangement and instrumentation is shown in Fig. 20. Air
is pumped by the fan into the air valve cavity.  Because of wide variations in the
dynamic .head, three static pressure probes located on  the face of the air metering
valve proved to give the most consistent results.  The average of these three mano-
meter readings was reported at P^.  In a similar manner, P2  readings were picked
up by three flush mounted static probes located in the wall of the 12-inch diameter
cavity immediately on the discharge side of the air metering orifice ports.  P3 and
PC measurements were made by a total of 5 additional static probes (see Fig. 19),
designed to determine flow pressure loss in flow duct and across the calibrated
metering orifice plate.  The metering orifice plate is designed to have 60 (1/2-inch
diameter) holes.  This arrangment allows rapid matching of the calibration orifice
pressure drops to the characteristic pressure drop of the combustor.  Initial design
point at 100 per cent flow requires the back pressure on the metering valve  to be 8
inches of H2O (design combustor pressure drop).  The  area of the calibration orifice
                                      39

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FIGURE 18.  AIR VALVE FLOW TEST MOCK-UP
                                                    uun
                                           I FOOT KCTMHDM PflB
                                           ••on WLoerrr C
FIGURE 19.  AIR VALVE FLOW TEST SCHEMATIC
                     40

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                      FIGURE 20.  Am FLOW TEST BENCH
plate was adjusted by covering ports in conjunction with cross calibrations against a
venturi meter until the desired 8 inches of H2O was obtained at a flow of 47 pounds per
minute.

        Procedures for all tests were similar.  The valve would be positioned to the
100 percent power lever position and the fan motor operated at a voltage necessary to
obtain 10 inches static pressure at P,. The power lever position would be then adjusted
to lower power setting to determine the flow characteristics from 100 to 10 percent
flow.  At each lever position setting, the pressure distribution and  me flow in the test
rig were measured to determine the characteristic performance of  me valve.  When
indicated by test results, design changes were incorporated to adjust the performance
toward the  established goals. Results from concurrent emission monitoring tests
being performed on the combustor  rig revised air fuel ratios during these tests.  These
tests indicated that the valve should be designed to produce a lean condition at lower
power settings rather than the rich air-fuel ratios incorporated in the original port
configuration.  As a consequence,  the nonlinear cutout in the orifice ports  were re-
moved and  the valve was designed to produce a constant air-fuel ratio of 25.5 across
the range from 100 to 10 percent.

        A  rich condition (small flow areas)  was  intentionally maintained to facilitate
optimization of the valve air-fuel ratio during combined combustor air yalve emission
tests.  Leaning of the ratio required only metal removal from the orifice plates by
means of simple hand file operations in the range of 10 to 100 percent.  Below 10 per-
cent, the design allows the backup  plate to be rotated with respect to the housing to
provide a wide adjustment of air-fuel at the low heat release rates.

                                       41

-------
        Initial tests were performed with a single conical baffle on the outlet of the
fan.  Correct flow,  pressure levels, and pressure distributions were obtained down to
a 50 percent power  lever position with electrical input power at 2. 02 HP  (29.2 volts
and 51. 5 amps).  Unsatisfactory results were obtained below the 50 percent flow position.
Investigation isolated the problem to swirl resulting from the air  valve-fan combination.
Static pressure readings in the system were severely affected and biased toward the high
side by the swirl  conditions.  A venturi meter initially used for flow measurement was
particularly sensitivie to swirl induced errors.  Since swirl of any significant magni-
tude can cause serious maldistribution of air in the combustor (in addition to the instru-
mentation problem), a series of configuration modifications were made and tested  to
reduce this aerodynamic problem.  The edge configuration of the  rotating and stationary
plates were found to be a prime factor in the swirl condition as was the apparent high
and unstable dynamic pressures in the air valve cavity.   Metering edges  were made as
sharp as possible and a number of baffle configurations were tested.  Although not
optimum, the present configuration has appeared to reduce swirl  to acceptable levels.
Baffle arrangements are shown in Figure 17.  Three radial baffles were  found to be
necessary to reduce the high dynamic head (4 inches  of HgO pressure)  to allow linear
valve control.   Although these baffles permitted linear operation of  the air valve, the
pressure difference across the metering ports was not a  constant 2  inches of HgO as is
required for accurate operation of the compensation system.  Swirl was reduced by
use of 12 axial baffles that compartmented the air valve cavity into 12 separate flow
passages supplying  air to each of the variable orifice ports.  Although  the valve func-
tioned well in flow bench tests with this configuration, an additional pressure loss
(approximately 3  inches of HgO) has been required.  As a consequence, the fan must be
matched to the system by operating at 33 volts (14,200 rpm and 2. 3 HP).

Flow Metering Test Results

        Component level test results with a calibrated orifice are listed  in Table n.
Power lever position has been used as the main parameter for these tests and has been
established by dividing the total stroke between 10  and 100 percent into 9  equal divisions.
An 8-foot long, 8.5-inch diameter,  extension to the calibration orifice duct was added
to obtain an accurate flow calibration below 10 percent by means of  a smoke velocity
calibration. A puff of smoke was introduced at a port 12 feet from the end of the duct.
From the elapsed time it takes to traverse the duct length (37 seconds  for 1%), accurate
flow measurements were obtained.  This was necessary since at 10 percent the back
pressure on the air valve must be maintained at 8(0.1)2 or 0. 08 inch of H^O. At the 1
percent flow, the back pressure must be maintained at 0.0008 inch of H^O.   Thus, the
smoke velocity method appears to be the best approach for accurate flow  calibrations
down to the 1 percent level.  Below the 10 percent position of the power lever, smoke
velocity calibrations were made to determine the 7, 5, and 2 percent flows.   These
positions were recorded and compared  with the 10  to 100 percent  valve gain  on a linear
basis to determine accuracy.
                                       42

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                                   TABLE II

     AIR CONTROL VALVE TEST RESULTS WITH 2. 25 INCH WIDE BY-PASS
            VALVE (0.5-INCH BAND ON ORIFICE SIDE OF BY-PASS)
Power Lever Position
PI (In. of H90)
P0 (In. of H00)
- -


P - P., (In. of H^O)
L/P1 ~P2
V 2 " 1
L
P.J (In. of H.,0)
p (In. of I100)
'' (Main Man.)
p (In. of H00)
'' (6° Man.)
Measured Flow
/ Posit ion \
\J Flow )l°
1007r *


'•
,_ Deviation From
I)" , .
Linear
Overall Air Fuel Ratio, dP •
Metering Accuracy
100
9.60
8.27
1.33
-18.0%
7.91
8.02

100.0
0
30.0
90
8.37
6.75
1.62
-10.0
6.66
6.73

91.5
•1.8
28.4
80
6.85
5.25
1.60
-11.0
5.08
5.11

79.5
-0.5
28.2
70
5.70
4.02
1.G8
-8.0
3.84
3.85

69.2
-1.1
27.3
60
4.70
2.97
1.73
-7.0
2.78
2.77

58.6
-2.4
26.7
50
3.92
2.09
1.83
-4.0
1.90
1.80
1.76
46.8
-6.5
24.8
40
3.43
1.45
1.9S
-5.0
1.24
1.14
1 . 145
37 . 8
-5.8
24.2
30
3. 1.3
.90
2.23
.5.0
0.74
0.7
O.B7
28.8
-4.2
23.3
20
2.95
. 52
2.43
•11.0
0.40
0. .15 •
0.30
19.3
-.1.5
22 3
 * dP Compensation
  Effect on Fuel
        A plot of Table II results and smoke velocity calibrations is presented in
Figure 21.  The 100 percent position was made to be correct by adjusting motor
voltage and valve  by-pass port configurations to simulate combustor pressure drop
at P2.  It can be seen that the valve has a high degree of linearity across a wide flow
range.  Cross plotted with valve position is the main parameter which is to be con-
trolled, air fuel ratio as affected by the air metering valve performance. From 100
percent flow down to 6.5 percent,  the air-fuel ratio remains within the initial design
established limits of ±10 percent.   Below this flow setting, the air flow is below the
limits, indicating larger flow areas are necessary.  This final flow adjustment has
been accomplished in combined valve and combustor emission tests.  A comparison
of preliminary emission test results shows that to meet the optimum air-fuel ratio in
the low power setting, the valve required an increased flow area to lean it down to
34:1 air fuel at 1 percent. Since the total area open at 1 percent is 0.2 inch2, increas-
ing flow area at low power settings was accomplished by rotation of the backup plate by
approximately 0. 010 inch.  A pressure drop across metering valve ports proportional
to flow is important for proper functioning of the fuel pressure compensation system.
If fan voltage inadvertently changes,  or if it is intentionally reduced at part loads to
                                      43

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                         OOMI ONET AIR V*L VI
                         PERF< RMAf CE
                       BE ADJXJi TED 1 4 CO* BINEI
                     /AI.V* CQ1T lUSTQ J EMI SON
                         CALC JLATBD
                         ACTUIL
                                               TljBT WTjH 2.2J WIDE
                                               BY-PA88 VALVE
                                               (BAND Oil ORfftCE
                               4	L.	,
                                         CALIF RATIO N8
                    10   20   30   40   SO   60   70   80   90   100
                             POWER LEVEL POSITION (PLP) %

     FIGURE 21.  FLOW CONTROL PERFORMANCE OF AIR METERING VALVE
lower power losses, the fuel pressure regulator is designed to automatically reduce
flow proportional to air flow.  To achieve accurate air-fuel ratio control, it is necessary
that a pressure difference across the metering ports be maintained at 2 inches of H9O.
Figure 22 shows flie pressure differential as a function of valve position.  At high
flows,  the PI - P2 is less than calculated and at low flows, it is approximately the same
amount (~0.5 inch of H2O) above the 2 inches of H2O design point.  Assuming perfect
fuel pressure regulator and fuel metering valve, the effect on fuel flow and  air-fuel
ratio are also plotted in Figure 22.  It is seen that the air-fuel ratio exceeds the 10
                                       44

-------
          hi
             30
             28
             26
             24
             20
             18
                            B' '-PAS VAL'
                               SID
                                  or
                                                                      LEA*
                                                                     '10%
                                                                     -10%

                                                                      RICH
g
u *•
|S •«
If'
J i
| oT-io
*
-9n









N



X




^-,



'^s



' 	 H



^~^








\

1
•1
LI
i
i

k
»
IH
r

           e.
           i
            H
           P.
             10
           71
           u
           X
           H
           X
                                                      /
                    10   20   30   40   SO   60   70   80   90   100  %

                            POWER LEVER POSITION. %
FIGURE 22.   AIR VALVE AP TEST RESULTS WITH MOCK-UP AIR VALVE
                                       45

-------
percent lean band at high flows and the 10 per cent rich band at low flows.  It appears
that the cause of this large difference between calculated and actual pressure differences
at the metering port is associated with the high dynamic head in the small diameter fan
discharge annulus.  High velocity jets must change their direction as a function of geo-
metry and the amount of flow required.  Several approaches to the problem exist. The
optimum method would be to specially design a fan to match the large diameter geometry
of the combustor and air valve and thereby have lower speeds and dynamic turning
losses.  Schedule and scope limitations of the program have not allowed the design and
incorporation  of an optimum fan.   An extension duct  to allow a larger plenum has been
used in  most of the final emission demonstration tests as a simulation of the  aero-
dynamics expected in a matched fan-combustor system.

        An even greater effect upon the pressure distribution across the metering
port can be obtained by baffle and flow straightener configuration changes.  Another
significant factor is the location of the static pressure probes for P^.  An apparent
pressure gradient across the face of the valve plate has been observed and found to
vary with power lever position.  By optimizing by-pass valve geometry, baffle
configurations and static probe location, significant reductions in variations of  P, - P2
with power lever setting were  obtained.   A major portion of the valve development
tests were directed toward this, since a stable P^ -  P£ not only can be used to  compen-
sate for operating variations in fan voltage but can form the basis for major flow
changes accompanying voltage reductions  to save on  parasitic losses.

        From these test results  the final configuration of the valve was fabricated.
Figures 2 and 23 show the final design prior to installation into the demonstration
system. Mechanical functioning is very simple as the air metering port moveable parts
are mechanically fastened to the by-passport fingers.  Air pressure and springs force
the sliding valve elements against the valve housing and backup plate.  Capability to
change the low flow rate air-fuel  ratio is provided by allowing the backup plate  to be
rotationally indexed with respect  to the power lever position and by an adjustable
linkage  between the power lever and  the fuel valve (See Fig. 2 and 23). Adjustments
to the final air-fuel ratio were made in integrated systems tests to minimize emissions
from the combustor (See Section 7).

Fan Noise
        Due to the high dynamic turning and straightening losses, the fan has been
operated at 33 volts to match it to the system requirements.  As a consequence, the
noise level was above the 80 db at fan rated  conditions.  An analytical and experimental
investigation of the problem and its potential solutions has been initiated.

        Sound pressure levels of the Rankine burner fan section operating separately
from  combustor were taken with and without inlet guide vanes.  The results were 102 db
                                       46

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              FIGURE 23.  DEMONSTRATION SYSTEM AIR VALVE


(ref. 20ptN/m^) overall sound pressure level with inlet guide vanes and 94 db overall
sound pressure with the inlet guide vanes removed.  A 1/10 octave frequency analysis
of noise with guide vanes  removed, is given in Figure 24.  Test details are given
below:

            • Blower Power Setting:  31 Volts, 53 amps

            • Blower Speed:  14,000 rpm

            • Number of Blades:  4

            • Microphone  Location and Orientation:  Five feet off blower axis from
            blade location.  Set 45 degrees from vertical toward blower and perpen-
            dicular to blower  axis (see Fig. 25).

            •Surroundings: Semi-reverberant (see  Fig. 25 for details)

            • Equipment  General Radio Type 1564A Sound and Vibration Analyzer
                        General Radio Type 1521 B Graphic Level Recorder
                        General Radio Type 1560-P4 Preamplifier
                        General Radio Type 1560-2131 Microphone
                        General Radio Type 1562 Sound-Level Calibrator
                                      •17

-------
00
            M    MM
                                                          FREQUENCY (Hi)
                      FIGURE 24.   SOUND PRESSURE LEVEL VERSUS FREQUENCY FOR BLOWER
                                   WITHOUT INLET FLOW STRAIGHTENER

-------
      POURED CONCRETE
K
U
ff\
2
&.-»
SMOOTH





i
J

WOOD CEILING

CONCRETE FLOOR

- f
5'
t
(X
U
H
OT
3
*A
SMOOTH



x> 45°

-j-*7!
^c


        WOOD PANELING
    FIGURE 25.   NOISE MEASUREMENT ROOM GEOMETRY AND MATERIALS


           •Background Noise:  Background noise was observed to be more than
            10 db below peak levels occurring at 250 cps, 980 cps, 2000 cps,
            and 3000 cps.

            •Background levels are at least 5 db below  blower levels at all other
            frequencies except in the ranges 25-220 cps and 350-400 cps where
            blower and background levels are equivalent.

        An approximation of overall noise level  of the blower in a free acoustic field
can be made using the following relation:
                                                     a2
            Reduction in Noise Level (db)  = 10 log „   —
                                                 10   a
where
™ is the new sound absorption coefficient which is one for this case by
  definition of a free field
            a  is the room sound-absorption coefficient determined by relative room
               wall areas multiplied by their respective material absorption coefficients.

        If S is the total room wall surface area, then for a predominant level of 980 cps

                  2S
            a   = —  (Absorption coefficient of concrete)
             1     «3

                  2S
                + —  (Absorption coefficient of wood paneling)
                   5
                                       49

-------
                  s
                + — (Absorption coefficient of smooth plaster)
                =  s|- (0.0175) +  - (0.06) + i (0.035)
                    5             55

                =  0.034S

Then Reduction in Noise Level is = 10 Iog10 1. 00/0. 034 =  14.7 db.

        This implies that the blower operating in a free field without inlet guide vanes
would register approximately 80 db overall sound  pressure level assuming the same
relative microphone position.

        Figure 24 shows four distinct frequency peaks of

            1.  92 db at 980 cps (this was also noted as the predominant noise level
                frequency with vanes.)

            2.  90 db at 2000 cps

            3.  83 db at 2900 cps

            4.  81 db at 245 cps

        Blade passage frequency is

            14,000 Rev/Min     Cycles
              60 Sec/Min    4  isT   =

        The first three measured peaks have frequencies corresponding to the first,
second and third order of blade  passage frequencies, respectively. A fourth peak is
coincident with the blade support rotational frequency.  It can be deduced  that bearing
noise and or single blade  passage frequency is the cause of this peak.  The exact cause
is not of immediate interest  because it is considerably lower than  the fundamental and
first harmonic peaks caused by  blade passage excitations.   Appendix A discusses
several design approaches that could be incorporated into an optimum fan design that
can reduce blade passing frequency tones in an optimum configuration.
                                      50

-------
6.2. 3  Delta-P Fuel Pressure Regulator Calibrations

        In order to maintain the desired air-fuel ratio independent of fan voltage, fan
or motor efficiency reductions,  or blade fouling a delta-P regulator was developed to
control fuel pressure directly proportional to the pressure differential across the air
metering valve.  Figure 1 functionally describes the operation of the valve.  In order to
obtain as low an operating range as practical a large diameter actuator was used (Fig. 26) to
convert the pressure drop (P^ - P2) across the air metering valve into an effective
control force. Development tests on the regulator valve were performed with the fuel
metering valve as the flow load  component.  Supply pressure was provided by the
Rankine combustor's system fuel pump.  Air for the actuator input was regulated from
the shop air system as the main test parameter variable.  At a differential actuator
input of 1.0,  2.0 and 2.5 inches of H2O differential, the  metering valve was placed In
the 100, 50 and 1 percent flow positions.  The fuel bypass adjustment was then set to
give the best regulation across the range of fuel flows  and input air pressures.  Through-
out the remainder of the test it was locked in position while both air pressure and fuel
flow were varied.  Regulated fuel pressure was recorded at each of the three flow
conditions at every 0.1 inch from 1.0 to 2. 5 inches of  HgO input pressure.

        Several different configurations of springs and diaphragms were tested. A
PVC molded diaphragm and a thin flat rubber diaphragm proved to be the best.  A
plot (Fig.  27) of the regulated pressure versus actuator input signal for the PVC
                                     ^
            FIGURE 26.   FUEL PRESSURE REGULATOR INSTALLED
                          ON P  - P  ACTUATOR
                                       51

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01
to
                                                                                                  •  100% FLOW

                                                                                                  A  50% FLOW

                                                                                                  .   1% FLOW
            1.0
                            1.2
                                     1.3
                                                                                                                   2.2
                                                                                                                            2.3
                                                                                                                                    2.4
                                                      1.5       1.6      1.7       1.8       1.9      2.0       2.1

                                                      CONTROL ACTUATOR INPUT PRESSURE DIFFERENTIAL (INCHES OF WATER)


                 FIGURE 27.   FUEL PRESSURE REGULATOR WITH MOLDED 0.012 INCH THICK PVC DIAPHRAGM
                                                                                                                                            2. 5

-------
diaphragm shows that it has reasonably good performance across the extended oper-
ating range.  The dashed line plotted in the figure represents exact (zero error) com-
pensation.  An upper and lower 5 percent error in fuel flow is also included.  Only
at the low end of the input pressure range does the regulated pressure output cause a
flow error of more than 5 percent (7.5% rich). This diaphragm configuration used
0.012 PVC with a molded convolution between the outside ring seal and the center
plate support. It is a rugged diaphragm but has corresponding high stiffness.
Irregularities in the convolution and the high stiffness caused the nonlinearities in the
performance curve.

        A flat diaphragm made  of thin resiliant rubber (0. 008 inch thick) had virtually
zero stiffness. Results of test with this material showed that it regulated pressure
accurately enough to hold the flow of fuel within ±4 percent of the ideal across the entire
range of input pressures (Fig. 28).  Final calibration tests results are given in
Table III.   Normal operation of the system would be in the  range of 1. 0 to 2. 0 inches
(if voltage  to the motor is reduced to reduce part load parasitic  losses).   Within this
range the valve regulates pressure accurately enough to hold the flow within ±3 per-
cent of the ideal.

6. 3  TRANSIENT RESPONSE EMISSIONS TESTS

6.3.1   Description of  Demonstration Combustor  System

         A fully integrated combustion system was tested for emissions. Figure 29
shows a cross section  of the complete  system. Air is drawn through the fan and
delivered  into the air valve chamber.  From this chamber  the air either is  metered
into the combustor section through  the  air metering ports or it is bypassed  through
the ports located  on the outer diameter of the valve housing.  Figure 29 is only one
of several  configurations tested  in the fully integrated system emission tests.  Most
of the tests were  performed with a  36-inch long extension between the air valve and
fan to provide a better aerodynamic matching between these components.

        Figure 30 illustrates the system with the fan mounted inside of the  extension.
By mounting the fan within the duct it was possible to vary  the axial distance between
the fan discharge and air valve.   Instrumentation, control and a self containpd fuel
system  including  tanks were mounted in a mobile test stand for maximum test
flexibility (Fig. 30 and 31). All steady state power levels and transient emission
tests were performed with the use of the power lever position as the only variable.
Regulation of air  and fuel flows was automatically maintained by the previously
described components  integrated with the combustor.   Fuel flow rates during a
particular  test run were determined by prior calibrations of the fuel metering valve
and measurement of the pressure differential across the metering ports.  Two variable
area flow  meters were connected through manual valves in series  with the flow meter-
ing valve.  These were used to verify the calibration of the flow metering valve
                                       53

-------
                                                                                •  100% FLOW
                                                                                A   504 FLOW
                                                                                •   1% FLOW
1.0     1.1
                               1.4      1.5      1.6      1.7      1.8      l.B      2.0      2.1

                                   CONTROL ACTUATOR INPUT PRESSURE DIFFERENTIAL (INCHES OF WATER)
                                                                                                       2.3
                                                                                                                        2. 5
 FIGURE 28.   FUEL PRESSURE REGULATOR PERFORMANCE WITH 0.008 INCH FLAT RUBBER DIAPHRAGM

-------
               TABLE III

PRESSURE REGULATOR CALIBRATION WITH
   0.008 INCH FLAT RUBBER DIAPHRAGM


Test
Points
1A
2A
3 A
4A
5A
fiA
TA
-A
!iA
HIA
11A
12A
13A
14A
ISA
IfiA
50*
IB
2B
3B
40
.">B
6B
7B
SB
9B
10B
11B
12B
13B
14B
15B
1KB
1C
2C
3C
4C
5C
6C
7C
8C
9C
IOC
11C
12C
13C
14C
15C
16C

Flow Motor
Readings (pph)
(Uncorrected)
.0
.0
.05
. 1
. 1
.2
O.'J
0.8
0.7
0.7
0.7
0.7
0.7
O.fi
0.6
0.6

41. 1
nil n
.SI. 5
62.5
53. 5
:"). 0
48.0
47.0
45.5
44.0
43.0
41.8
40. 10
39.0
37.8
36.0
96.0
98.0
99.5
100.2
104.0
106.2
93.5
91.0
88.0
85.5
83.2
81.0
78.0
76.0
73.5
70.2
Regulated
Fuel Valve
Inlet Pressure
(psig)
9.9
10.5
10,8
11.4
11.9
12.5
3.6
y.os
8. 5fi
8.0
7.S
7.25
6.70
6.30
5.85
5.40

9.!l
10.30
10.90
11.30
11.70
12.20
9.55
9. 10
8.50
8. 15
7.70
7.20
6.75
6.30
5.90
5.35
9.60
9.90
10.35
10.80
11.20
11.60
9.20
8.70
8.20
7.75
7.30
7.00
6.45
6. 15
5. 70
5.25
Pump
Discharge
Press
(psig)
10.8
11.3
11.8
12.2
12.8
13.3
10.4
10.0
9.4
9.0
S.8
8.1
7.5
7. 1
6.8
6.2

10.8
11. 1
11.8
12.2
12.6
13.1
10.4
10.0
9.4
9.0
8.6
8.1
7.7
7. 1
6.8
6.3
10.5
10.8
11.2
11.6
12.1
12.5
10.0
9.6
9.2
8.8
8.2
7.9
7.5
7.0
6.5
6.2

Actuator
Sensor Input
(in. of HjO)
2.0
2. 1
2.2
2.3
2.4
2.5
1.9
1.8
1.7
1.6
1.5
1.4
1.3
1.2
1. 1
1.0

2.0
2. 1
2.2
2.3
2.4
2.5
.9
.8
.7
.6
.5
.4
.3
.2
.1
.0
0.95
1.00
1.05
1.10
1. 15
1.20
0.92
0.88
0.80
0.75
0.70
0.65
0.60
0.55
0.50
0.45

Regulator
Back Pressure
(psig)
0.32
0.35
0.35
0.37
0.38
0.38
0.38
0.38
0.38
0.38
0.38
0.38
0.38
0.38
0.38
0.38

0.05
0. 10
0. 10
0.12
1.4
1.6
0.08
0.05
0.02
0.01
0
0
0
0
0
0
2.0
2.1
2.2
2.3
2.4
2.5
.9
.8
.7
.6
.5
1.4
1.3
1.2
1.1
1.0

Valve
St roke
(Inches)
0.006































0.866






























0.8915






























                  55

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01
          r i r
                                    FIGURE 29.   DEMONSTRATION SYSTEM

-------
 FIGURE 30.  DEMONSTRATION SYSTEM WITH LONG MIXING DUCT


FIGURE 31.   INTEGRATED SYSTEM DEMONSTRATION TEST STAND
            ARRANGEMENT
                            57

-------
periodically throughout the test phase.  They could not be incorporated in the system
continuously since they have a flow sensitive pressure restriction that would modify
the fuel pressure regulator's control of air-fuel ratio.  Two flow meters are necessary
since a 10 to 1 flow ratio is the normal range of these instruments.  Fuel delta-P
across the metering valve was recorded on both a precision laboratory gauge with
divisions of 0.1  psi and an electric transducer displayed on a strip recorder during
transient tests.  An electric position transducer connected directly to the power lever
position recorded stroke or power level directly in percent of full power on a strip
recorder.  Its basic function  was to establish power rate input changes during trans-
ients.  Air-fuel  ratio was obtained by an averaging probe CC>2 measurements in the
exhaust and cross checked against combustor air flow calibrations made with standard
orifice flow meters.

6. 3. 2   Startup and Shutdown  Transients

        The design goal for this demonstration was to be able to start the combustor
and to be up to full power in three  seconds.  Experimental analysis indicated that
startup was optimum at the 30 percent power lever position.  Figure 32 shows the
results of a startup from 30 percent and followed by an immediate power increase  to
100 percent in three seconds.  In this start the blower was at operating speed prior to
ignition excitation and  opening of the fuel valve solenoid.  Fuel was circulating through
the regulator valve and being bypassed to the tank at the time the solenoid valve opens
to connect fuel to the rotating cup feed tube. Transient gas emissions were obtained by
recording the  output of each of the four gas analyzer systems on a strip chart.  Response
delays of each of the four measurement systems is  from 2 to 5 seconds with the relatively
short exhaust  gas sampling lines used in this test (less than 20 feet).  It is not possible
to directly plot the emissions in  gm per Kgm of fuel since the fuel flow and air flow
cannot be exactly established during a transient.  Additionally the transient  response
of present state-of-the-art gas analyzers is not sufficiently fast to provide for exact
analysis of a transient load change.  However, a reasonably good quantative estimate
of transient performance can be  made by an analysis of the direct volumetric emission
data obtained as  a function of time on the Beckman strip chart recorder.   Design goal
limits  are indicated on the exhaust gas concentration scales.  These limits are based
on calculations that relate the mass emission goals to the volumetric flowrate and  the
air-fuel ratio  as obtained from the CC>2 concentrations.  At the startup transient none
of the emission levels  exceeded the volumetric design goals.  Hydrocarbon (measured
as C,) came relatively close to the limit, 20 ppm compared to a  limit of 35 ppm.
After one minute of operation the system was subjected to decrease power transient
from 100 percent to 5 percent in two seconds and  then an increase to 100 percent in
2. 7 seconds.  After another twenty seconds at 100 percent power the combustor was
shutdown by closing the systems fuel cutoff solenoid valve. A transient peak of 700
ppm in HC  emissions was recorded. Cause of this high peak appears to be a small
amount of fuel leaking  into the combustor after the initial closing of the solenoid.
                                       58

-------
1000-
900-
800-

700-
E 600-
o.
Q.
¥• 500-
400-
300-

200-
100-
0-
16-
14-

12-

10-
o
(J Q_
6-

4-

2-
0-
           1000

            900

            800

            700
           O.   '
           ex
           Ssoo-
             400-

             300J

             200-

             100-

               0-
 150-

 135

 120

 105-

E 90-
OL
O.
O
Z75-

  60-

  45-

  30>

  15-

   0-
•HC
                                  40
                                        60    80    100
                                         TIME (SECONDS)
                                                          120
                                                                140
                                                                      160  ISO
               FIGURE 32.   STARTUP AND SHUTDOWN EMISSIONS


In the demonstration configuration the solenoid valve was located approximately 12
inches from the rotating cup. In a specially designed installation the shutoff valve could
be located directly in the fuel supply tube to eliminate any possibility of after drip.
An alternative would be to utilize a solenoid valve with sufficient "negative" displace-
ment to extract a small fuel volume from the fuel supply tube as the valve closes.

6. 3. 3   Power Level Transient Emissions

        A  major performance goal of the program was to be able to rapidly modulate
power from any one power level to another without severe increases in emissions.  A
response rate of 50 percent  of full power per second was established as a preliminary
goal. Table TV lists emission level transients in ppm obtained from strip chart
records for power changes.  For each transient test the demonstration system's
power lever was manually positioned from  a high heat release (A) down to a lower
heat release (B).   At level (B) the fuel flow was stabilized with a dwell period of
approximately 0. 5 second.   Immediately after the dwell the lever was moved back to
the initial heat release position of (A).   Movement of the lever was  controlled in each
test  to  obtain a rate of change of greater than 50 percent per second.  Fuel flow
corresponding to the lower lever position is listed in the first two columns of Table IV.
                                       59

-------
O5
o
                                                        TABLE IV
                        POWER LEVEL TRANSIENT EMISSIONS (COMBUSTOR CONFIGURATION "D")
Transient
From A and B
to A
A(pph) B(pph)
109 5
109 10
109 30
109 50
109 70
50 5
50 20
80 30
Transient Time
(sec)
4 12 13
1.70 0.6 1.85
1.40 O.G 1.65
1.70 0.4 0.9
0.7 0.7 0.6
0.3G 0.46 0.30
0.65 0.7 0.75
0.26 0.58 0.23
0.79 0.58 0.6
Emissions Measured
CO (ppm)
Limit
534
534
534
591
534
591
591
591
Steady
State
45
45
40
15
45
15
15
15
Peak
8
285
40
223
45
940
30
60
NO (ppm)
Limit
28
28
28
31
28
31
31
31
Steady
State
42
4G
49
43
47
46
46
52
Peak
45
48
50
39
47
46
46
52
HC (ppm)
Limit
32
32
32
35
32
35
35
35
Steady
State
14
13
15
13.5
15
12.5
18
16
Peak
14
15
15
25
15
23
18
16
C02 (7o)
Steady
State
G.5
6.5
6.65
7.2
6.5
7.2
7.2
7.2
Peak
8.3
9.7
8.05
G.3
6. 65
9.5
7.7
8.05
* Combustor Configuration "D"
                            FUEL FLOW TRANSIENT TEST SEQUENCE
                                     (MANUAL OPERATION)
                           STEADY STATE A
                                              DWELL
                                                                              EMISSION TRANSIENT
                                                                                CHARACTERISTIC
                                                                                 PEAK VALUE
                                                                         STATE
      STATE
                                                                                  3-10
                                                                                    sec
:*i

-------
Transient time periods between lever positions is tabulated in the next columns. A
comparison between the steady state and peak value of emissions is used to determine
the transient  performance characteristic of the system.  From the results it can be
seen that only CO has a characteristic transient peak significantly higher than the
steady state emission levels.  In two transient tests  the HC also exhibited a significant
peaking tendency.

        NO emission levels showed essentially no deviations from steady state levels
during these  transient tests.  Combustion configuration  "D" (see Section 7) was used
for transient  tests and  thus the steady state NO levels were higher than  the design
goals.  However, the basic measured response characteristic of the NO to power
level transients is not expected to vary significantly  with the lower NO configuration
"A".  The  configuration differences were that "D" had 25 percent more  air in the
primary zone than "A".  "D" also had less heat loss from the flame due to the use of
radiation shielding. An analysis of the HC and CO peaks against the design goal
limits indicated that only in one test was the limit exceeded.  A detailed analysis
of this test (50 to 5 pounds per hour) was made by comparing the results strip chart
records of the time dependent parameters.  Figure 33 compares the high peak
emission results with a typical low peak  emission tests. In the top graph a high CO
peak is recorded well above the limit (indicated by an arrowhead on the  scale).  A
peak in the HC level is relatively high but well within the limit.  For comparison a
typical response characteristic (Fig.  33B)  indicates  virtually no HC peak and a CO
peak well below its limit. A  detail analysis of the lever position versus time (dotted
lines) indicates that one major difference between the two tests is  rate of change of
power. From this data,  and other  similar traces it appears that the emission peaks
on CO and  HC increase significantly as the rate of change in lever position (fuel  flow)
increases. It also appears that the sensitivity to rate change is only during power
decrease transients below 30 pounds per hour fuel flow.  It appears transient at  rates
several times the goal of 50 percent per  second can be achieved without significant
emission peaking at all conditions except decreasing from 30 pounds per hour.  The
rate of decrease that appears to have  caused  the high CO peaking was 140 percent
per second.

        Although well below  limits even a rate of 68 percent per second causes
significant peaking in CO.  It should be noted the HC emissions have virtually no peaks
at the same 68 percent per second  rate.

        In summary, it can be stated that the system has good transient emission
characteristics and can meet the emission levels  at the  transient rates established
as goals for  interfacing with  the vapor generator. However, since some vapor
generators may require very high combustor system response (depending upon margin
of safety between fluid  operating and decomposition temperatures)  additional consid-
eration should be given to improvement of the air-fuel ratio control during transients
at the low power levels.
                                       61

-------
 150
 133
 118
 102
t 71
 42

 27

 14

  0
    1000
    865
     740
     620
     510
a-410- >Q50
    0320
     235
 150
  75

   0
              DEMISSION COAL LMT (IN PPM
               AT TEST CONDITION A/F RATIO)
          0L
                 2   34   56   7   8   9  10  11   12  13  14   18  16  17
                 TIME FROM START OF  TRANSIENT  (SEC)
        A.   HIGH CO AND HC PEAK RESPONSE (POWER LEVER RATE
             APPROXIMATELY 150%/SEC)
150
133
118
1O2
87
I*
S56
42
27
14
0
C 1000
865
740
• 620
_ 510
0. 410
O
235
* 150
75
0
1UU
90

^ 80
UJ

r
z60
^
*50
S?
a 40
UJ
UJ
IJ3O

ct

10



: I POWER
: ".
I i t :— t!B%«
l"1 '"" -; • -INDICATES SLOPE
; ; PERCENT / secoo
': - : 	 ^EMISSION GOAL LIMIT (IN PPM
. ; .: X^ ^x Al TEST CONDITION A/F RATIO)
'-: : / ^^^
\ _^f ^^^^_ CO2



1 ,^57*;^
\ / 	 	 "NO

: ' / ^^
Ueev.* • / \
'•• • / \ — -
f- 	 \. HC
\j y ^-^_^L_

1 2 34 56 7 8 9 10 11 12 13 14 18 16 17 1t
TIME FROM START OF TRANSIENT (SEC)
1O.U 1
13.5 •

11.2 •
90
fe
72^.
UJ
55*
tf
40 •
O
O
2B •


4 Q
1.8
8 -
i
i

DU
45

40
35

305
i
25"
I
20"


15



5



           B.   TYPICAL CHARACTERISTICS (POWER LEVER RATE
                APPROXIMATELY 50%/SEC)

      FIGURE 33.  POWER LEVEL TRANSIENT EMISSIONS (COMBUSTOR
                   CONFIGURATION "D", SEE FIGURE 81)
                                     62

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                                      7
                           COMBUSTOR DISCUSSION


7.1  REACTION KINETIC STUDY BY COMPUTER MODELLING

7.1.1  Summary

          The computer model showed that, given adequate combustion volume,  CO
and NO emissions can be kept at a low level at high fuel rates.  At low fuel flows NO
emissions would be unacceptable but could be  reduced by introducing heat losses to
the vaporizer from the flame.

7.1.2  Emissions Analysis

        This section describes the methods used to calculate the emissions from the
combustor.

        The method used are incorporated into three computer programs:

            1.  Steady-State Combustion Program (SCP)
            2.  Chemical Equilibrium Program (ODE)

            3.  Generalized Kinetics Program (GKP) - Dynamic Science Report
                No. TR-C70-227-I.

        The first calculates the liquid phase heat release profile in terms of the
vaporization rate of the liquid droplets.  Liquid properties and injection parameters
are required input,  and the output consists of the vaporization rate of the liquid spray
in the combustion field.

        The Chemical Equilibrium Program (ODE) is used to calculate initial equilib-
rium concentrations in species in the primary zone to initiate the kinetics calculations.

        The Generalized Kinetics Program is used for calculation of emissions, and
solves the one-dimensional nonequilibrium reacting gas flow equations for a pressure
or area defined stream tube for any input defined gaseous chemical reactions.  Reaction
rates are input as an Arrhenius reaction expression and all competing reaction for a
                                      63

-------
particular species are included.  The ability of the program to consider arbitrary mass,
momentum, and energy addition,  coupled with the  steady state combustion program,
gives GKP the added capability of handling two phase flow problems (liquid droplet
combustion).

         The chemical species and the kinetic reaction system are defined by the input
of the symbolic species names (N, NO,  NO2, O, etc.) and the reaction set in symbolic
form (NO+ M = N+ O + M), where M is an arbitrary third body (all other species).
Program output consists of the fluid dynamic variables (residence time, velocity,
temperature, area ratio, and density) and species  concentrations as a function of
normalized distance.  Plotting capability is available for fluid dynamic properties,
chemical species concentration, total derivative with respect to distance, partial
derivatives of these total derivatives with -respect to any other variable and the net
production rate of any reaction.

         The fluid dynamic and chemical relaxation equations are integrated numerically
using an implicit procedure developed by Dynamic  Science personnel.  The  advantage
of the implicit technique is that it allows chemical  systems near equilibrium to be
analyzed in a practical manner. This is important in emission analysis as  the thermo-
dynamic properties of the reacting mixture are controlled by  the near equilibrium
concentrations of the major species.

         The program was developed for the kinetic analysis of rocket engines,  and
was selected by the Interagency Chemical Rocket Propulsion Group as the reference
program for the Aerospace industry.  It has been modified to analyze generalized
chemical flow problems rather than the more limited rocket nozzle analysis.

         The flow chart in Figure 34 gives a brief description of the input and output
of GKP.

7.1.3  Initial Calculation

         The initial effort involved setting up the Steady-State Spray Combustion
Program to provide the liquid heat release rate profile.  Figure 35 shows the
resulting calculation of the evaporation rate for Jet A fuel injected into a 200 feet per
second air stream at low velocity (5 feet per second velocity component along the air
stream).  Initial mass median drop size was estimated as 50  micron and a relatively
low standard deviation of the spray (nearly uniform size distribution) was used as
these are the characteristics of the rotating cup atomizer being employed.  (200  Feet
per second being me air stream velocity that will occur at full heat release  in the
preliminary combustor design.)
                                      64

-------
                                     PROGRAM INPUT
1
FLUID DYNAMIC MODEL INPUT
a) pressure or area profile
b) initial conditions (P,V,T)
c) mass addition profile
d) momentum addition profile
e) energy addition profile


CHEMICAL MODEL DATA
a) chemical reaction set e.g.
NO-t-M = N-KHM
b) initial concentrations of the
species in the reaction set

1
THERMO-CHEMICAL DATA
A library of thermal chemical
data on over 400 species Is
available.
                                          GKP
                                     The computer program
                                     solves the 1-D non-
                                     equilibriom reaction gas
                                     flow equations considering
                                     mass, momentum and energy
                                     transfer using an implicit
                                     numerical integration scheme.
                                    PROGRAM OUTPUT
  FLUID DYNAMIC PROPERTIES

  Pressure, temperature, density,
  velocity,  specific heat, and enthalpy
  are printed out for specific print
  stations
CHEMICAL SYSTEM PROPERTIES

Species concentrations,  net pro-
duction rates, and influence co-
efficients are printed out for
specific print stations         ,
         FIGURE 34.  FLOWCHART,  GENERALIZED KINETICS PROGRAM


         From Figure 35, it is seen that vaporization is nearly completed two inches
from the injection point.  Also shown are vaporization histories for three of the typical
drop groups comprising the spray.  This short lifetime of the liquid at the full heat
release condition of the combustor provides the logical basis to initiate the emission
calculations based on gas-gas equilibrium.  Droplet lifetimes at lower heat release
rates are shown in Figure 36.


         To provide a comparison base between the non-equilibrium emission estimates
calculated using the kinetics program and those expected at chemical equilibrium, a
matrix of air-fuel ratios was run.  Figure 37 shows the calculated flame temperatures
and  Figure 38 shows equilibrium concentrations of carbon monoxide and nitric oxide
(defined  as CO and NO,  respectively).  A second series of  runs on the equilibrium
program was run with NO suppressed (program did not consider NO species)  to provide
start conditions for the non-equilibrium kinetics calculations.  Kinetics  runs were
                                         65

-------
                  30 MICRON        50 MICRON
               INITIAL DROP SIZE  INITIAL DROP SIZE
                                                                       TOTAL SPRAY
                                                                          70 MICRON
                                                                       INFTIAL DROP SIZE
  .1     .3     .5      .7      .9      1.1     1.3      1.5
DISTANCE ALONG COMBUSTOR FROM POINT OF INJECTION - INCHES
                                                           1.7     1.9      2.1
                                                             JET A FUEL AT 80'F
    FIGURE 35.  FUEL DROPLET LIFETIME AT MAXIMUM HEAT RELEASE
                 RATE (2 x 106 BTU/HR)
                                   1.09
                                          54.5
                                                   109
                                         AIR FUEL OF 26 TO 1
                               0.2  0.4  0.6  0.8  0.9  1.0
                            DISTANCE ALONG COMBUSTOR
                            FROM POINT OF INJECTION, INCHES
FIGURE 36.  FUEL DROPLET LIFETIME,  50 MICRON DROP SIZE, AT 109, 45.5,
             AND 1.09 LB/HR OF FUEL
                                       66

-------
4000
iiOOO
2000
                         FUEL IS JET A AT 80° F
                         LOWER HEAT VALUE 18,400 BTU/LB.
                         AIR AT  80° F
                              AIR/FUEL RATIO
                        10                   20                   30
     FIGURE 37.  EQUILIBRIUM FLAME TEMPERATURE AS A FUNCTION
                  OF AIR/FUEL RATIO
                                     67

-------
        10
                                           2CT
                              AIR/FUEL RATIO
FIGURE 38.
EQUILIBRIUM CONCENTRATIONS BY VOLUME OF CARBON
MONOXIDE AND NITRIC OXIDE AS A FUNCTION OF AIR/
FUEL RATIO
                                   68

-------
    A/F = 4 . 0
        Mole Fraction

           ,«6]i957

           l.ir-k-2«
C02

CO
t

\?0
H?C
:w         4. ?."<£-14
           c ,
.240527

0.

o.a

.000021
A/F

N2
02
H20
H2
OM
0
H
ARGON-
NO
N
N02
C02
CO
C
N20
\M3
= 6.5
Mole Fraction
.53»73l
2.46E-13
.032206
,165783
5.S7E-8
1.8AE-12
,000007
,006907
0.
3,V?E-14
6.51E-1B
0,021520
•1V4641
0,
5.1.1E-14
,m)C004
A/F

N2
02
H20
H2
OM
0
H
A3GON
NO
\
N02
C02
CO
C
N?0
NM3
= 9.2
Mole Fraction
0.671527
5.45E-8
.114005
,079833
0.01)0034
8.7F-B
.000274
.007968
0.
2,3<5t-10
2.8F-12
.044272
.13?084
0.
1.35E-1C
4.1F-7
              FIGURE 39.  PRIMARY EQUILIBRIUM COMPOSITION

initiated with zero initial NO as the time to vaporize liquid was calculated as small
compared with NO formation time. Also, the primary flame zone is fuel-rich and
little NO is formed even at equilibrium (infinite dwell time) for a fuel-rich system.
Figure 39 shows the equilibrium start conditions used for primary zones at air-fuel
of 4.0, 6.5, and 9.2.  The equilibrium results are for a kerosene type fuel (C-^^22
molecular weight used) with heating value of 18,400 BTU/lb and thus are not general
results.  Other fuels must be examined on an individual basis.

        The third area of initial calculations involved gathering input data for the
kinetics program.  The program starts a kinetic chemistry calculation at a prescribed
input chemistry (calculated from equilibrium) representative  of the primary zone.
Mass,  momentum, and energy are then inputs to the primary zone at prescribed rates
based on mixing processes occurring within the combustor.  The basic input then con-
                                       69

-------
sists of tables of these quantities as a function of distance through the combustor.  At
constant pressure the gas dynamic relations within the program (hen calculate the
residence time (velocity) as a function of these quantities, i.e., residence time is not
a constant throughout the combustor.  Addition rate tables used for analysis of the
combustor configurations will be presented in the following section.

        The kinetic reaction mechanism used for emission estimates is shown in
Figure 40.   The low molecular weight hydrocarbon reactions were necessitated by
the addition  of the methane species at an air-fuel of 4 (see Fig. 39).

7.1.4   Combustor Design by Computer  Modelling

        The preliminary combustor design used is based on typical criteria used in
the design of gas turbine combustors modified to limit the peak temperatures in the
flame (Fig.  41). This has been  shown by the reaction studies to reduce the formation
of nitric oxide.  The control of flame temperature is achieved  by control of the amount,
position, mode and  speed of air  injection into the combustor.

        The combustor is divided into three zones as follows:

        1.   A primary flame zone consisting of about one-third of the combustor
             volume and lying immediately downstream of the  point of fuel injection.
             This is where ignition and  combustion commences.

        2.   A secondary flame  zone consisting of about one-third of the combustor
             volume and lying immediately downstream of the  primary flame zone.
             In this region,  most of the  combustion  reactions approach completion.

        3.   A tertiary flame zone consisting of  the remaining combustor volume and
             lying immediately downstream of the secondary flame zone.   In this
             region, air is added and mixed with the combustion products  to bring
             them to a safe level of temperature for use in the engine boiler.

             Air is  injected into these three regions either as  high velocity jets or as
             low velocity films flowing along the combustor walls. The later films
             serve  in the office of a control on wall  temperature by minimizing the
             amount of hot combustion productions that can impinge on the walls of
             the combustor.

        To  calculate emissions, a knowledge of  the method of air injection and mixing
in the combustor is  required.  Two different methods  of air injection were devised
and designated as designs A and  B (ref.  Fig. 41 and 45, respectively).  Small
modifications in the methods  of air injection were done with both designs and the
resultant emissions estimated.
                                       70

-------
REACTIONS
N2
02
H2
OH
HJ>0
M02
N20
C02
CO
CH4
CH3
CH?
CH
END
CO?
CO
H2
OH
W20
OH
W20
MQ
N!2
NQ
N2
CO
CO
CO
CO
CO
CO
C
C
C
CO
CH3
C
C
CH4
CH3
CH4
CH3
H2
M
M20
H20
OH
0
OH
CH3
LAST
• N * N , A«l , OElfl ,
=0*0, Aoi . 9Elft ,
= H f H , AS7.5E1*.
s H * (J, Ae.J,e>£*l8. N
= H * OH, Asi.~l7fc.17,
- 0 * NO, Atl.6El5,
s 0 * N2. As i. OE18.
= CO * 0, As 5.1E-H5, N»0
s C * 0 i A = 3,nEl6, N=!],5» 8*0
= fH3 + H , is 3.QE16. ^'= 0
- CH? * H , AS J.CF16, \= C
s CH * H , 4= 3. OF. ift, \s o
' C * H , AS 3.0E16, >.'* 0
TUK REAX
+ H = CO + CH , AS 5.6E.11,
*
*
*
»
*
*
+
+
CO
02
h =
H s
0 s
0 =
= CO?
s OH
H2 + 0
OH * H2
H * 0?
OH + OH
0 s 02 * N ,
0 - k
0 * !>' ,
«• C?s N02* 0
*•
*
*
*
+
*
*
+
*
*
*
*
*
*
*
*
•f
*
*
*
+
+
*
*
*
*
02 =
02 *
H =
.'M :
0 = C
H? '
N20 * 0 ,
C02 «• 0,
C * OH,
C * MO,
* 02,
CH +OH ,
H s CH +0 »
H2
H2Q
OH
OH
H2 s
CH3
CH2
C
CH
CH
CH3
CH
CH3
CH
CH2
CH
CH3
CH3 •
H20
s CH
= CH
= CH
s CH
CH4 * H,
s P.H
* CH
s CH3
: CH?
s CH3
* CH2
= CH?
• CH2
= CH2
= CH3
» CH2
• CH2
CH4 * 0.
* C
* OH
, Ac
A*
As
, As
Asl
A»l
A=l
'As 3
A=1,9E
, A = 2
, A»2
1 , 7 4 F. 1 3 ,
2 , 1 9 F 1 3 ,
2 . Ofc!14 .
•5.75^12,
.8E8, N
.3E13. -N
Nil , 0. B*0,
Uo 0 1 5 1 8s 0 1
=liQi B« Oi
• J » 0 . 0 1 B * 0 ,
.'N8l,.n. fa«0,0
,« R«3.58.
,0>
,5, B=0,0,
,5, 8=0,0.
,5, B*0,0.
,5, 8=0,0,
fJiC.O, B»l,
.11E16,
.70E16,
N s o , n ,
'I»C , 0»
N«0,0i
NaO.Oi
•-l.S.Bs
•0,0. 8=
,8El3, N«0,0. Bs
.OE*l3i
*13, NB0
ASJ..2E + J4, N«0
A= 1 . 2E*
16, ;i=i,
N« 0,0,
, B>54,
, B« 25
0 B«0,,
N =
N =
P« 9
B= 5
R»17
BCQ ,
5,94
0.0,
1,05
B>
15.
,83,

0,
0,
n,
0,
,79,
,
080,
1,0,
1,0,
,45,
,15,
,0,
78,
1

1
26.8,



As ?,4E!+l3» NE0, b°l,9')7,
t* 3 ,
Ao 3.
* M
* OH
* C
* 0?
f>13, n*
F*13. N=
A«5



A« 1.5E*
* CH2
* Cw
+ CH
* CH2
* CH2
+ CH<
* H
* H2
* OH
+ OH
* 0
* OH












AS 1
A«5
Acl
L4. Nt
AB1
A«l
A* 1
A'l
A»l
A»l
A'l
A'l
Ael
AE1
A'l
A'l
Ai 1,E*13, N«0
• CH4 * OH. *• 5
C , B»0.
0 « B"0.
•30E11,
.05E11,
,3EH,
•75E10,
0 , . Bil4
.05E11.
.05E11,
,05E11.
.05E11.
.05E11.
.05E11.
.05E11.
.05L11,
.05F.11,
,05Ell.
.05E11.
,05E11.
,. B» 7,
,E*14, N«0,, B«9
,
,
N =
N«
N«
N>
i
N»
Nc
N«
N>
N«
Nc
N«
No
N«
N«
N>
N*
3,
,9.


-,•5,
-,5,
-,5,
-,5,

-.5,
-,5,
-.5,
-,5,
',5,
-,5,
-,5,
-,5.
-,5,
-,5,
-,5,
-.5,


DDK PACE 7-6
DDK PACE 7.6
ODK PARE 7-6
14
LEEDS 2 NOV 6«, P-31
TR^A69-103 P-17, JAN 69
16
ODK. P-7-6
1D-2P PROG, P 7-49
1D-2P PROG, P 7-49
1D-2P PROC, P 7-49
1D-2P PROC, P 7-49
LEEDS 1 MAY 6«, P-4
B»0,0i 1D-2P PROC 7-bCi.
8»53,, 1D-2P PROC 7-51,
LEEDS ? NOV 68, P-l
LEEDS ? NOV 6B, P-V
TS-AA9-103 P-21, JAN 69
LEEDS 2 NOV 6fl, P-20
TR-A69-103 P-18, JAN 69
TR»A69«103 P-18. JAN (SV
TR.A69-103 P-18. JAN 6V
28 50
163
103
ioa
107
42
43
B«2,24,1D-2P PROC 7-5l,
B«2,24,iD-2P PROG 7-5i,
B»2,24,1D.2P PROC 7-51.
8-28,2ilD-2P PROC 7-5li
174
B«2,24,10.2P PROC 7-52.
B«2,24,1D-2P PROC 7-52.
B«2,24,1D«2P PROC 7-b2.
B«2,74,io.2P PROC 7-52.
B«2,74,1D«2P PROC 7«52i
B-2.82.1D-2P PROC 7-52.
B«2,74,1D-2P PROC 7-52.
8«2.8«,1D.2P PRCC 7-52.
B«2,74,iD«2P PROC 7-53.
B«3,19,1D-2P PROC 7-53.
8"2,74ilD.2P PROC 7-53.
B«2,83,10-2P PROC 7-33.
172
176
33
43














49
50
54
59

63
64
66
6'
70
71
80
81
86
8*
100
101


CARD
                         FIGURE 40.  REACTION SET
                                      71

-------
Design A, Configuration No. 1 (Fig.  41b)

        Twenty-five percent of the total air required for combustion is injected over
the rotating cup and mixes instantaneously with the fuel ejected from the lip of the
cup.  Another twenty-five percent of the total air is injected as film cooling along the
combustor wall.  The remaining air is injected at the juncture of the primary and
secondary flame zones.  The assumed rates of mixing of the various quantities of air
is shown in Figure 41a, the resultant air-fuel  ratios, gas temperatures and emissions
are shown in Figures 41d, e and f, respectively, for the maximum heat release rate
of 2 x 106 BTU/hr and an exhaust temperature of 2500° F.

        The results show carbon monoxide on the high side but low values of nitric
oxide.

Design A, Configuration No. 2 (Fig.  42b)

        Identical arrangment of air admission as in Configuration No.  1, but with
modifications to the method of the air injection at the juncture of the primary and
secondary flame zones so that half this air recirculates into the primary zone rather
than,  as before, into the secondary zone.

        The rates of mixing of the air,  resultant air-fuel ratios, gas temperatures
and emissions are shown in Figures 42 c, d, e and f, respectively.

        The results again show carbon monoxide on the high  side and also a consid-
erable increase in nitric oxide.  The large increase in  nitric  oxide is caused by the
higher primary zone flame temperature which  results from the recirculation into it
of additional  air.

        Therefore, for minimum nitric oxide, the recirculation of air must be mini-
mized and the primary flame zone  maintained with minimum air.  Some increase in
secondary flame zone volume must be provided to minimize carbon monoxide.

Design A, Configuration No. 3 (Fig.  43b)

        The large quantity of wall cooling used in Configurations 1 and 2,  while pro-
viding low wall temperatures and hence long combustor life, will tend to reduce effec-
tive combustor volume due to quenching of the  flame reaction at the cool wall.   The
quantity of film cooling was therefore drastically reduced as it appears that combustor
volume was low.

        The results show that emissions are very similar to, but slightly higher than,
the results obtained with Configuration No. 1,  and have low nitric oxide with high
carbon monoxide.

                                      72

-------
Design A, Configuration No. 4 (Fig.  44b)

        Identical arrangement as in Configuration No. 3,  but with modifications to the
method of air injection at the juncture of the primary and secondary flame /ones so that
half this air recirculates into the primary zone rather than into the secondary zone,  as
was in Configuration No.  3.

        Again, as with Configuration No. 2,  a large increase  in nitrix oxide but small
change in carbon monoxide is achieved.

        The analysis gives results which would not, in practice,  he expected; i.e.,
the increased wall temperatures would provide a larger effective volume in Configura-
tions  3 and 4 and reduce carbon monoxide.  Therefore, testing of these  configurations
is needed for better understanding of the processes at work.

Design B, Configuration No. 1 (Fig.  45b)

        Modifications were made to the design of the combustor as in Figure 45b, and
air admissions as in Figure 45c used.  The essential difference between Design B  and
A is  that a much slower rate of mixing of the air admitted  at the juncture of the primary
and secondary zones is used in Design B.

        Calculations at full heat release were made of the flame temperature and
resultant emissions of carbon  monoxide and nitric  oxide are as shown in Figures 45c
and f.

        In comparison with Design A, Configuration No.  1,  a  large increase in nitric
oxide occurs,  though within required values.  Some increase in carbon monoxide also
occurs.

        The prime reason for the increased nitric oxide lies in the reduction of the
rate  of admission of the secondary air.  This permits combustion to proceed at higher
temperatures for a longer period of time than compared to previous configurations.
Increased admission of secondary air and earlier injection of tertiary air would serve
to reduce both NO and  CO considerably.

        Because the combustor has to operate at any heat release rate from 100 per-
cent  to 1 percent, calculations of temperature and emissions were made at heat releases
of 50  percent and 1 percent also.  These results are shown superimposed on the results
for 100 percent heat release in Figures  45e and f.

        A very large increase in NO  emissions occurs as heat release; is reduced;
such values lying outside specification limits.  The reason for  this is the large increase
                                       73

-------
in time spent at high temperature when combustor heat output is reduced as likewise,
more time is available for complete reaction to carbon dioxide to occur.

Design B, Configuration No.  2 (Fig. 46b)

         Identical arrangement of air admission as in Configuration No. 1 but with
modifications to the method of the air  injection at the juncture of the primary and
secondary flame zones so that 100 percent of this air recirculates into the primary
zone rather than, as above,  into the secondary zone.

         As before, the rates of mixing of the air, resultant air-fuel ratios, gas tem-
peratures, and emissions were computed and are shown in Figure 46c, d, e, and f,
respectively, for 100 percent heat release and also for 50 percent and 1 percent of
full heat release.

         The results are similar to those obtained with Design B, Configuration No. 1,
save that some increase in nitric oxide occurs.  This is due to the increased time at
high temperature caused by introduction of additional air into the primary zone by
recirculation from the  secondary zone.  A similar result was obtained with Design A,
Configuration No. 2, and for  the same reason.

Heat Losses

         At low levels of heat release,  it is expected that substantial reductions in
flame temperature could occur due to  the large masses of cool metals surrounding
the flame.  This heat loss would depend considerably on the design and location of ihe
boiler.  Calculations were done of the effects of small heat losses on the emissions
from Design B in both Configuration No. 1 and 2.

         A substantial reduction  in volume of nitric oxide emissions is obtained with
only small heat losses,  reference Figures 47 and 48 with both configurations.  Effects
on carbon monoxide formation are small.

         As nitric oxide emissions are still excessive at low heat release rates,  it is
necessary for considerable reductions in flame temperature to be made.

Conclusions

         Emissions of nitric oxide are, at low heat releases, considerably outside
required limits and well within limits  at high heat releases.  Carbon monoxide is well
within limits at low heat releases, and marginally in limit at high heat releases.
Vaporizer (boiler)  design could have a major effect on emissions as heat losses from
the flame can be high.
                                       74

-------
 10,000_
 1,000
a
a
i
E-

K
H
Z
W
O
z
o
o
   100
    10
                  CO


                  NO
                                                           5% HEAT

                                                             LOSS
                                                        NO HEAT

                                                          LOSS
                               6        8

                         DISTANCE - INCHES
                                                10
                                                        12
FIGURE 47.
              EMISSIONS AT ONE PERCENT HEAT RELEASE WITH

              AND WITHOUT A FIVE PERCENT HEAT LOSS FOR

              DESIGN A,  CONFIGURATION NO. 1
                               87

-------
 10,000.-
  1,000 -
I
o
H
U
u
                                                        5% HEAT
                                                         LOSS
                     468
                       DISTANCE - INCHES
                                                    12
  FIGURE 48. EMISSION AT ONE PERCENT HEAT RELEASE WITH
             AND WITHOUT A FIVE PERCENT HEAT LOSS FOR
             DESIGN A, CONFIGURATION NO. 2
                             88

-------
Discussion

        Results show that for minimum emission of nitric oxide, the rate of admission
of air must be carefully controlled.  In the primary zone, the flame must be kept rich,
with minimum air; and in the secondary zone excess air must be added as rapidly as
possible.  If this can be done over the entire operating range  of the combustor from
1 percent to 100 percent heat release, then the formation of nitric oxide can be elimi-
nated almost completely, especially if heat losses to the vaporizer  can be used to limit
flame temperature.  Carbon monoxide should not present a problem with such an
arrangement of air admission, save perhaps at or near maximum heat release.  If
this is so, an increase in combustor volume may be required  to hold emissions within
limits.  By providing sufficient volume, the carbon monoxide  formation would be
reduced to an extremely low level. Heat losses to the vaporizer would, however,
tend to raise CO emissions.

        In essence,  the distriction of CO and the formation of NO is a process that
is controlled by the mixing residence times and heat losses of the reaction gas within
the combustor. The  mixing process can be controlled by hardware configuration and
it is possible,  therefore, to design a combustor that will minimize  both NO and CO
emissions over a wide  range of turndown and operating conditions.

7. 2   COMBUSTOR AND TEST RIG DESIGN

7.2.1  Combustor Design

        A combustor was designed and fabricated. These parts are shown in
Figures 29,  49, 50, 51, 52, and  53.  The distribution of the air admission into the
combustor was set up initially per Design B, Configuration No.  1.

        The heart of the system is the rotating cup fuel injector which makes the
100 to 1 turndown ratio possible.   The rotating cup has the following advantages:

            •Essentially no fuel pressure.  This allows the use  of a small low
            pressure fuel pump and  avoids fuel contamination because there
            are no small fuel metering orifices.

            •Degree  of atomization is not dependent on fuel flow or viscosity.

            •Minimum cost.  The cup and its drive motor are inexpensive as is
            its auxiliary equipment (fuel pump and ignition).

            •Is a proven design.   It has been used for many years in combustors.
                                       89

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                     COMBUSTOR CASE
               EXHAUST STACK & LOCATION
                    FOR BOILER

                                     I

                                .


                     '

                                  '
                                                  •
                                                    .--'

      FIGURE 49.  SIDE VIEW OF ROTATING CUP COMBUSTOR ASSEMBLY

           •Controlled spray - the angle is predictable and hence ignition reliability
            is high.

           •No power requirements.  For test flexibility the cup is separately
            driven.  In practice it would be mounted onto the fan motor shaft and
            the additional power required would be negligible.

        A weight breakdown for the experimental and production versions of the
combustor assembly, together with the materials used is shown below.
            Current Weight
               Pounds

                28.78
                38.00
                 0.69
                 3.00
                 4.28

                 1.25
                76.00
Combustor
Combustor Case
Fuel Injector Cup
Injector Drive  Motor
Combustor Swirler and
Dome Assembly
Support Pins and Ignitor
          Total Weight
Production Version Weight
	Pounds	
          6.0
          4.75
          0.23
          *
          1.25

          0.75
         12.78
        * Integrated with fan motor.
                                     90

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              COMBUSTOR CASE
                                              ••flHB	

     FIGURE 50.   FRONT VIEW OF ROTATING CUP COMBUSTOR AND CASE
Combustor materials are:
            Combustor Case and Flanges
            Combustor Liner
            Combustor Swirler and Dome
            Assembly
            Fuel Injector Cup
            Support Pins
Mild Steel (0.062 in. gauge)
Hastelloy X (0.062 in. gauge)
321 Stainless Steel

Mild Steel
Mild Steel
In practice, tests indicate that all materials used could be low cost mild steel
or cast iron save for the liner which would be a 300 series stainless steel.
                                      91

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                                  FIGURE 51.

                                  FRONT VIEW OF ROTATING CUP
                                  COMBUSTOR AND CASE WITH ROTATING
                                  CUP REMOVED
           COMBl'STOn DOME
COMDUSTOrf DOME
                    PRIMARY DILUTION TUBES
                                         FIGURE 52.

                                         REAR VIEW OF ROTATING CUP
                                         COMBUSTOR AND CASE
                               92


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                                                  ROTATING
                                                  CUP MOTOR
              FIGURE 53.  ROTATING CUP AND MOTOR ASSEMBLY


7.2.2  Combustor Test Rig Design

Air Supplies

        Prior to tests with a fan and air metering valve, tests were done with a
slave air supplied by a remote air compressor.  The very wide range of fuel flows
to be investigated required a corresponding wide range of air flows. For
tests to have significance, the air flow must be measured accurately.  This requires
metering orifices having pressure losses far in excess of the capability of the chosen
fan and therefore air to the combustor had to be provided from a higher pressure
source.  This air was supplied from a centrifugal compressor that supplied air, via
a system of metering orifices, through a ten inch diameter pipe, directly mounted
onto the combustor case, Figure 54 and 56.  This air was manually controlled to any
required  flow by means of a throttling valve.  Throughout the test,  the air supply was
                                      93

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                                           MANOMETERS
        EMISSION
EXHAUST  PICKUP
         PROBE
                                                                FLOW STRAIGHTENER
                                                                              TC
                                                                                               AIR INLET
                                                                              3.0 ORIFICE
                                                                          /   SHARP EDGED    \
                                                                               AIR FLOW
                                                                          \ MEASURING ORIFICES/
                                                                                                   10.0 DIA.

                                                                                        FLOW STKAIGHTNER
BOILER     COMBUSTOR
       - «4»-   —*f
SECTION      SECTION
                                                                                        FLOW STRAIGHTENER
                                                                           AIR MEASUREMENT

                                                                              SECTION
        FIGURE 54.   SCHEMATIC OF  COMBUSTOR TEST  RIG

-------
FIGURE 55.  REAR VIEW OF FAN AND CONTROL VALVE ASSEMBLY
           SHOWING ANTI-SWIRL PLATES INSTALLED
  FIGURE 56.  COMBUSTOR RIG SHOWING ARRANGEMENT OF AIR
             METERING ORIFICES
                           95

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maintained at levels of flow, temperature, and pressure appropriate to the fan
capabilities (2830 Ib/hr,  80° F,  and 12 inches water), but is capable of a much larger
range, with up to 7000 pounds per hour of flow,  at temperatures from -40° F to +1200° F
and pressures to 250 psig.

        There is an aerodynamic influence on the combustor caused by the particular
method of control of the air from the fan  to the combustor.  Air is passed through a
metering valve consisting of twelve variable area openings into the combustor at a
velocity of 95.5 feet per second as dictated by the pressure drop across this area
(Fig.  55).  Because of the interaction of the aerodynamics of  the fan and the air by-
passed overboard, and because of the turning losses in the metering chamber, this
air has a profile of velocity in both circumferential and radial directions, and a tan-
gential velocity component that varies considerably over the operating range of the
combustor.

        At maximum fuel flow,  the air exiting from the metering valve into the com-
bustor annulus,  is accelerated and passes into the combustor at a velocity of 195 feet
per second. At  this flow, the effects of radial and swirl velocity profiles on combustor
air distribution are assumed negligible.  However, at minimum fuel flow the air
velocities into the combustor are small at 1.95 feet per second, and as the same 95. 5
feet per second velocity is maintained at  the control area,  the effects of velocity pro-
files could be considerable. It is not possible to simulate such aerodynamic influences
except by  use of the fan and this, of course,  destroys the object of the test rig.  There-
fore,  final combustor development must be carried out with the fan.  In order to
minimize  differences  between fan and rig aerodynamics, the combustor was designed
such as to minimize the effects of entry velocity variations.   On the rig, any velocity
profiles from the upstream metering section are eliminated by means of a flow
straightener.

7.2.3  Instrumentation

Flow  Measurements - Air

        Air flow is measured by means of any of three orifice runs that are mounted
in parallel upstream of the combustor (Fig. 56 and 54).  Because of the wide range
of air flows, three different orifices of 3/4,  1-1/2, and 3 inches diameter are used.
Each  orifice is sized for optimum Reynolds number at air flows corresponding to
respective fuel flows midway in the ranges of 109 to 22.8,  22.8 to 4.77,  and 4. 77 to
1 pounds per hour.  The orifices are sharp edged, with corner pressure taps,  and
have upstream flow straighteners to minimize aerodynamic influence on the metering
section from the upstream on-off valves.   These valves are manually operated so that
air flow can be diverted to the metering orifice appropriate to the  air flow  being used.
Because of the cumbersome nature of this instrumentation,  it is not possible to simulate,
except approximately, the transients that will in practice occur as fuel flow is varied.

                                       96

-------
Flow Measurements - Fuel

        Fuel flow is measured by three Fisher-Porter flow meters mounted in-series
and covering a range of flows from 0. 2 to 200 pounds per hour.  The flow meters are
calibrated against a master flow meter which is also used to calibrate the flow meters
used in tests on the fuel metering valve.  This minimizes the possibility of cumulative
errors.

Temperature - Measurement

        A Solar design and fabricated Pt-Rh thermocouple (Fig. 57 and 58) was used
to measure combustor outlet temperatures.  Triple radiation shielding is provided to
a 0. 04 inch diameter bare bead Pt-10%Rh versus Pt thermocouple.  Three concentric
Pt-Rh radiation tubes are arranged to reduce radiation losses to low levels in this
unit.  High convective heat inputs (to offset radiation heat losses) are provided by a
high velocity aspiration system that draws the  combustor discharge across the thermo-
couple bead.  Good results were obtained at temperatures up to 3030° F and velocities
of 0. 6 Mach number.  The thermocouple could be mounted in forty  different locations
to provide an accurate description of both circumferential and radial temperature
profiles at the exit of the combustor as shown in Figure 59.

Pressure Measurement
        Pressure measurements are taken by means of Bourdon gauges, water and
mercury manometers.  All critical pressures are duplicated to minimize errors.
Because at fuel flows much below 20  pounds per hour, the air flow is so small as to
cause negligible combustor pressure loss, it is not possible nor is it very meaningful
to accurately record the actual loss at such flows.  Fuel pressure is not measured, it
can be considered with negligible error, as being at the prevailing combustor pressure
and which is, without a vaporizer, ambient pressure.

Fuel Supplies

        Fuel is supplied to the combustor via the flow meters at a pressure slightly
in excess of the  combustor pressure  and is throttled  to the required level by means of
a needle valve.  The standard  fuel used in the tests is JP-5.  Such a fuel has a more
precisely controlled specification than the fuels required to be used  (#1 Diesel, Jet
A, and kerosene), and has combustion characteristics typical but  slightly worse than
these fuels (i.e., is likely to have slightly higher CO and H/C emissions), Appendix
B.  Other fuels available and briefly used are JP-4 and #2 Diesel, and which represent
fuels  having combustor characteristics outside the best and worst likely limits expected
of the specific fuels.
                                       97

-------
                       THERMOCOUPLE LEADS
                                                           TO ASPIRATOR
                        JUNCTION
                       -1ST RADIATION SHIELD
                         2ND RADIATION SHIELD

                          3RD RADIATION SHIELD

FIGURE 57.  END VIEW - HIGH TEMPERATURE PROBE WITH TRIPLE RADIATION
            SHIELD AND HIGH VELOCITY ASPIRATION SYSTEM
                                                              '
   FIGURE 58.  INSTALLATION OF A HIGH TEMPERATURE THERMOCOUPLE
                                   98

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 FIGURE 59.   CIRCUMFERENTIAL AND RADIAL POSITIONS OF THERMOCOUPLE
              AT THE EXIT OF THE COMBUSTOR


        Emissions are powerfully influenced by air-fuel ratio which is influenced by,
among other things, fuel volatility.  Therefore,  a range of volatilities was deemed
necessary in the tests in order to obtain as much understanding as possible of the
processes involved in emissions (especially emissions of NO2)

Emission Sampling Probes

        Various types of sampling probes were investigated and the final type used is
as shown in Figure 60. The probes consisted of 29 pickup points located at points of
equal area across the combustor to provide  a representative average sample.

        In the case of CO,  NO and CO2 the  probe was cooled to room temperature by
means of an air cooling jacket so as to ensure no further reactions inside the probes
and to allow removal of water vapors.

        A separate probe,  of identical design,  was used to sample  for unburnt hydro-
carbons. To  ensure no reactions inside the probe, but also to prevent condensation,
the probe was cooled to a temperature of 350 to 375° F.  The entire  HC sampling line
connecting the probe to the FID  was electrically heated to 350 to 375° F for the same
reason.
                                      99

-------
                              PROBI TEMPERATURE
                               CONTROL AIR FLOW
                                         • WELDED TOT BMP. CONTROL TUBE
                                                 21 PORT
                                                 EQUAL
                                                 ARIA
                                                SAMPLING
                                                PROBBS (I)
                                                ONI roR HC
                                              ONC FOR NO * CO
                                               (10- TO EACH
                                                 OTHER)
                                                                EXHAUST
                                                          12 INCHC8
                                                       SAMPLE TO FID FOR HC PROBI
                                                     SAMPLE TO NMR FOR NO 4 CO PROBI
                                 COOLINC
                                   AB
/ {   ELICTRIC HIATIR
     FOR HC PROBE
"JAMBIDTT AIR TIMPEHATURI
    FOR CO It NO PROBI
             FIGURE 60.  SCHEMATIC OF EMISSION PICKUP PROBE

7.3  COMBUSTOR DEVELOPMENT

7.3.1  Summary

        Initial combustor development was done on a rig without use of a fan and
metering valve and emissions were at reasonable levels.  Tests with a fan and
metering valve required more development to get acceptable emissions.  When a
vaporizer was installed at the combustor exit emissions levels increased to unaccept-
able levels and indicates some radiation shielding may be necessary to get acceptable
emissions.

7. 3.2  Preliminary Combustor Rig Tests

        The combustor was run over a range of fuel flows from 5 to 109 Ib/hr using
JP-5 fuel,  and with the air-fuel ratio maintained at various values between 22 and 30.
From these results it was possible to find the optimum air-fuel at any fuel flow for
minimum emissions.  This optimum air-fuel is shown in Figure 61.  It can be seen
that only at maximum fuel flow is the design value of 26 air-fuels achieved.  With
reducing fuel flows a small reduction in air-fuel is required.  However, at values of
fuel flow less than about 50 Ib/hr a considerable excess of air is required to obtain an
acceptable  flame.
                                      100

-------
          a
               100
               90
                80
                70
                GO
             o  .r>0
                40
                10
                                 I
                                       I
                                            I
                                 24    2(i    28   30
                                  AIR-FUEL RATIO
34
    FIGURE 61.   Affi-FUEL RATIO FOR MINIMUM EMISSIONS AS A FUNCTION
                 OF COMBUSTOR AIR FLOW


        The actual emissions resulting from maintenance of this optimum air-fuel
relationship with fuel flow are shown in Figure 62 and 63 together with the increased
emissions resulting from a 10 percent error in air-fuel (it is not to be expected that
control of air-fuel can, in practice, be maintained much closer than this).

7. 3. 3  Combustor Pressure Loss Reduction

        Up to this date, combustor pressure loss had been 12. 7 inches HgO.  Ic was
necessary to reduce it to 8 inches I^O in order to keep  the power consumption of the
fan within reasonable limits.  This was done and there was a considerable increase  in
flame length and emissions, notably hydrocarbons and carbon monoxide.  Changes to
the hole pattern of the combustor were made though the  original staging of air-fuel
ratios within the primary, secondary,  and tertiary flame zones was maintained.
                                     101

-------
              17
                                           Wa
                                                 = 2835 LB/HR AIR
                                                     ±10% FROM OPTIMUM
                                                         AIR-FUEL
                                                       OPTIMUM AIR-FUEL
                                                       100
                                      '100%
FIGURE 62.   EMISSIONS OF CARBON MONOXIDE AS A FUNCTION OF COMBUSTOR
              AIR FLOW, AT OPTIMUM AIR FUEL FOR MINIMUM EMISSIONS AND
              ALSO WITH A ±10% DEVIATION OF AIR  FUEL FROM OPTIMUM


        The effect of reduced pressure drop was  that penetration of air into the flame
was reduced.  Consequently,  the poor mixing of air and fuel resulted in lean and rich
pockets that caused lengthening flame due to, in both instances,  a reduction in the
rate of combustion.

        Modifications were made to the hole size and pattern to improve mixing with
beneficial results, though it was not possible to reduce  flame  length to the values
                                      102

-------
             1.6 -
             1.4 ^
          u
          3
          ^  1.2
          S
§
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to
S
             1.0
             0.8
          O  0.6

          €
             0.4
             0.2
                                           NOX LIMIT
10% FROM OPTIMUM
   AIR-FUEL

   OPTIMUM
   AIR-FUEL
                                     Wa    --- 2835 LB/HR OF AIR
                    10
                        20
                            30
                                 40
                                     50
                                 AIR FLOW,
                                          60   70
                                           Wa
                                                  80
                                                       90  100
                                          Wa
                                            100%
FIGURE 63.  EMISSIONS OF NO2 AS A FUNCTION OF COMBUSTOR AIR FLOW,
             AT OPTIMUM AIR-FUEL FOR MINIMUM EMISSIONS AND ALSO
             WITH A ±10% DEVIATION OF AIR-FUEL FROM OPTIMUM


recorded with the high loss combustor. Results are shown in Figures 64.  By
maintenance of an air-fuel of 22, it was possible to achieve the required  goals in
emissions.  Small errors in air-fuel would, however,  result in emissions outside
limits.

        In general, the  results are  in agreement with theory.  Figure 64a shows the
time dependence of NO formation as with decreasing fuel flow (inversely  proportional
to combustion time), emissions increase.

        Figure 64b shows the time dependence of CO formation in that increasing
fuel flows result in increased emissions of CO.  Minimum emissions are achieved
with  reduced air-fuels as a result of increased reaction rate  of the higher temperatures.
It would appear that an optimum combustor, operating at the  design point of 26 air-
fuels, would have NO emissions considerably below limits but would require increased
combustor length to reduce CO.

        Figure 64c is more difficult to interpret as large variations in hydrocarbon
emissions are seen.  This might be  explained by the high background emissions and
below which it is possible to maintain the flame when air-fuel is optimum.
                                      103

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                    LOOT 1.38
0.9

0.8

0.7

0.6

0.5
H
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£ 0.3

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 £  0.6
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                                  22.1 A/F
       JPS FUEL
  8 INCH COMBUSTOR LOSS
        10  20  30  40  SO  60   70  80  90  100  110
                FUEL FLOW LB/HR
   B.  CARBON MONOXIDE
     JPS FUEL
8 INCH COMBUSTOR LOBS
                        28.5 A/F
                                    A.  NO2
                                        REDUCED PRESSURE LOSS
                                                                          LIMIT 1«.2B
                                                                    JPS FUEL
                                                                 INCH COMBUSTOR LOSS
                                       10  20  30  40  SO  60  70 80  90 100  ISO
                                                FUEL FLOW LB/HB
                                              C.   HYDROCARBONS
        10 20  30  40  SO  60  70  80  90  100 ISO
                 FUEL FLOW LB/HR
  FIGURE 64.   EFFECT OF FUEL FLOW ON EMISSIONS AT DIFFERING AIR-FUEL
                 RATIOS USING RIG AIR SUPPLIES
                                          104

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          FIGURE 65.   COMBUSTOR RIG WITH VAPORIZER INSTALLED


7. 3.4  Simulated Vaporizer Tests

        A vaporizer was designed to simulate an actual installation and assembled
onto the end of the combustor (Fig. 65).  The vaporizer consisted of 250 feet of 0.7
inch inside diameter stainless tube. In order to obtain sufficient heat transfer it was
necessary to use a gas side pressure drop of 1.75 inch of water.  It was not therefore
possible to use with the fan because of the increased pressure losses.

        Emission measurements,  Figure 66, showed a considerably reduction  in
flame performance.  Unburned hydrocarbon emissions increased by up to fifty times,
carbon monoxide increased slightly and NG>2  emissions decreased about fifty
percent.

        This flame performance is not acceptable and is a direct result of the heat
losses from the flame to the vaporizer caused, principally,  by flame radiation.

        Therefore, because of the major influence of the vaporizer on flame perfor-
mance, it is essential to develop the combustor as an integrated unit with the vaporizer.

        It is noteworthy that nitrogen oxide emissions were reduced due to heat
losses, as is to be expected.  This would allow a control on nitrogen oxide emissions
somewhat independent of air-fuel ratio provide sufficient volume was provided as to
allow  reaction of hydrocarbons and carbon monoxide. By shielding the vaporizer
from the flame, the emissions as previously reported could be maintained.
                                      105

-------
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         LIMIT- 1.38
                                N02 EMISSIONS
                                   \    1
                                                       I	I
                          10  20  30   40   50  60  70  80  90
                           LIMIT - 21.3
                                CARBON MONOXIDE EMISSIONS
                          j	i	i	i	i
                          10  20  30   40   50  60   70   80   90



                          HYDROCARBON EMISSIONS

                          LIMIT- 0.48
                          10   20  30  40  50   60   70

                                 FUEL FLOW LB/HR
                                    80  90
FIGURE 66.   EFFECT OF VARYING FUEL FLOW ON EMISSIONS WHEN AIR-FUEL

              IS MAINTAINED AT A CONSTANT VALUE (26/1) AND USING A RIG

              AIR SUPPLY AND BOILER
                                       106

-------
        The low levels of nitrogen oxides that have been reported in domestic furnaces
are probably as a result of radiation heat losses.  Heat losses can be readily controlled
in such an installation because of the single fueling rate used. However, modulation
of heat rate requires on-off control of fuel which results in unacceptable warm-up
emissions of carbon monoxide and hydrocarbons.

7. 3. 5  Combustor Tests With Fan

        The object of reducing combustor pressure loss was to  permit testing with a
fan whose power requirement was within the limits set as a goal for parasitic power.
The combustor and  fan were therefore assembled together, with the air metering
valve, for further combustor tests (Fig.  29).  The fuel and air metering valves were,
however, manually  operated.

        Tests were done over a wide range of operating conditions varying from 109
to 5 Ib/hr.  It was obvious  that considerable differences in flame performance from
that seen using the test rig air supplies.  Two effects were seen.

        1.   At high fuel flows considerable circumferential variations in air flow
             occurred. Consequently, large differences, circumferentially, in flame
             length resulted.  Emissions were high, especially of hydrocarbons and
             temperature distribution was unacceptable.

        2.   At low fuel flows the flame  appeared uniform but air-fuels in  the
             various zones were  different from those occurring  on the combustor
             test rig.

        The circumferential air  maldistributions were investigated by flow tests.
The basic check was to coat the flow surfaces with oil and then inject talcum into the
fan.  These checks  indicated that the flow was unstable, resulting in separations in
air at different circumferential locations.  The nature of such flow differences are
shown schematically in Figure 67a and b.

        The only method of effectively controlling such air maldistributions without
use of excessive pressure loss or combustor volume is to design the fan and combustor
as an integral unit with low flow velocities and the minimum of diffusion, as in
Figure 67c.  As the location of the vaporizer can also have an effect on aerodynamics,
it also must be integrated into the aerodynamic design.

        The axial air maldistributions are explained by the presence of swirl  in the
fan.  The swirl angle was too low to measure, but the tangential velocity component
was high in relationship to  the combustion velocities.  The resultant swirl  caused
large axial and radial variations in the hole discharge coefficients.
                                       107

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                                       LATS SEPARATION OF AIR RESULTING IN HIGH VELOCITY A IB
                                       ENTERING COMBUSTOR CAUSING A COLO SPOT.
                                      EARLY SEPARATION OF AIR RESULTING IN LOW VELOCITY AIR
                                      ENTERING COMBUSTOR CAUSING A HOT SPOT.
                                                                J
                                          PREVENT SEPARATION BY USING A LARGER FAN OF LOWER VELOCITY
                                          THAT DOES NOT REQUIRE DIFFUSION.
                CUBaCNT SOLUTION IS TO IMTKKFOU ALONG M1XJNO DUCT BtTWCEN FAN AND
                COMBUSTOR TO ALLOW WAKtt TO DIMPATI TO A LOWEB VILOC1TT.


FIGURE 67.    AIR MALDISTRIBUTIONS  DUE TO  UNSTABLE DIFFUSION
                                            108

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        To control such a phenomena means design of a fan with low air velocities
and with a minimum of swirl.  It is an advantage in this respect to run the fan at low
speeds when fuel flows are  low, and to have the highest possible combustor pressure
loss.  Metering plate pressure loss should be kept low.

        An optimum fan design was not feasible at this stage involving, as it did,
unknowns in vaporizer  configuration, cost, and time outside the program scope.  There-
fore,  as a temporary expedient,  a large mixing duct,  36 inches in length, was placed
between the fan and combustor, Figure 67d.  An immediate and radical improvement in
flame performance was noted though it was still apparent at the low fuel flows that
all air swirl had not been eliminated.

Temperature Traverses

        It is essential that rig and fan tests duplicate each other.  Otherwise, any
demonstration of a working package by use of a rig becomes worthless.  Experience
indicates that at least half the  problems of emission control lies in proper design of
a fan  and its associated fuel and air control system.  Such a system of control,
covering as it does an extreme range  of flow variations, is considerably beyond current
state-of-the-art.   A means of  checking simulation precisely and in a simple manner is
by temperature traverses of both rig and fan package.  Ordinarily, at the flame tem-
peratures involved (up  to 3200° F), thermocouples are inaccurate and have a very
short life.  The alternate approach of emission traverses is time consuming, laborious,
and expensive.  Solar has developed a high temperature thermocouple that has a long
life and is accurate (to 1%)  at the temperatures involved.  Thus a simple tool is avail-
able that enables precise definition of aerodynamic simulation and hence  combustion
reactions by comparing temperature profiles on the rig and actual hardware.

        A thermocouple was therefore installed for such a traverse.  Traversing
was done by locating the thermocouple in forty different locations as shown in
Figure 59 and the results and conclusions noted as follows.

        Figures 68 and 69  show the average radial profile at combustor  exit at a
fuel flow of 88 IbAr when using rig air and fan air supplies, respectively.  A com-
parison of the radial profiles is shown in Figure 70.

        It is apparent  that good repeatability in radial profile is achieved,  indicative
of good simulation, aerodynamically (and hence of combustion reactions), between
rig and fan. Allowing for the small difference in air flow between the two tests,  the
difference in temperatures  recorded is one percent, and so the  traverses can be
assumed accurate  and representative.  The large difference in temperature  that exists
between the center and outside of the combustor is as  a result of inadequate  mixing of
fuel and air.  Consequently, there exists large radial  variations in air-fuel.
                                      109

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                                                                 HOT SPOT
              14(111    Minn   IMOO    2000    2200    2400    2fiOO
                                COMBUSTOR OUTLET TEMPERATURE.
                                                              2*00
          AVERAGE TEMPERATURE   2384" F
          HOT SPOT TEMPERATURE   2920°F
          COLD SPOT TEMPERATURE  1458' F
                                   TEMPERATURE SPREAD
                                   FUEL FLOW
                                   COMBUSTOR PRESSURE LOSS
   3000

  1462"F
  88 LB/HR
  5.7 INCHES WATER
FIGUBE 68.   AVERAGE RADIAL PROFILE OF TEMPERATURE OUT OF COMBUSTOR
               USING RIG AIR SUPPLY
          5. O
          4.0
O
 _
en"
s
a
          3.0
          2.0
        u
        j
          1.0
          0.0
             COLD SPOT
                                                                   HOT SPOT
               1300 1400    1600    1800    2000    2200    2400    2600
                                   COMBUSTOR OUTLET TEMPERATURE, *F
        AVERAGE TEMPERATURE   2335'F
        HOT SPOT TEMPERATURE   2380' F
        COLD SPOT TEMPERATURE  1315" F
                                 TEMPERATURE SPREAD
                                 FUEL FLOW
                                 COMBUSTOR PRESSURE LOSS
                                                                   2800
1565' F
88 LB/HR
5.8 INCHES WATER
FIGURE 69.  AVERAGE RADIAL PROFILE OF TEMPERATURE OUT OF COMBUSTOR
              USING FAN AIR SUPPLY
                                         110

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       5.0
        0.0
           1400   KiOO     1800     2000     2200    2400    2600     2800

                      COMBUSTOR OUTLET TEMPERATURE, °F

FIGURE 70.  REPEATABILITY OF RADIAL PROFILE OF TEMPERATURE OUT OF
             COMBUSTOR USING BOTH FAN AND RIG AIR SUPPLIES
Theoretically, either a lean or rich mixture will cause an increase in emissions and
so there's some room for considerable improvements to be made.

        In addition to the radial variations in temperature (and air-fuel)  noted, it is
seen from Figures 71  and 72 that circumferential variations in radial profile exist.
This is shown more clearly in Figures 73 and 74 which show the circumferential
temperature distributions at various radial position for  rig and  fan air systems,
respectively.

        The variations in temperature distribution between rig and fan are significant,
indicating reasonable, but not exact, aerodynamic simulation.  The major source of
temperature maldistributions lies, however,  in inadequate mixing that, as shown
above, is simulated exactly on rig and fan systems.

        Radial traverses were then done at differing fuel  flows, holding air-fuel
constant, Figure 75.   Changes in profile occurred, indicative of changes  in mixing as
a function of fuel flow.  In general, however, the  basic characteristics of cold  center,
hot wall profile was maintained over a wide range  of conditions.
                                      Ill

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                RADIAL POSITION #1
               j	i	I	l
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  2400

  1900

  1400

   900
RADIAL POSITION #2
                RADIAL POSITION #3
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                RADIAL POSITION #5
                                        2900
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             RADIAL POSITION #6
                                        2900
                                        2400
                                        1900
                                        1400
                                         900
                                                  RADIAL POSITION #8
                                                         I
                          345          012345
                                   RADIUS (INCHES)
           FUEL FLOW 88 LB/HR        COMBUSTOR PRESSURE DROP 5.7 INCH WATER
FIGURE 71.   RADIAL PROFILE OF  TEMPERATURE OUT OF COMBUSTOR AT
              SEVERAL DIFFERENT CIRCUMFERENTIAL LOCATIONS AND USING
              THE RIG AIR SUPPLY
                                        112

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                RADIAL POSITION #1
2900




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1400




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             RADIAL POSITION #2

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                RADIAL POSITION #7
                                         2900
                                         2400
                                         1900
                                         1400
                                          900
          RADIAL POSITION #8
            012345          012345

                                    RADIUS (INCHES)

            FUEL FLOW 88 LB/HR       COMBUSTOR PRESSURE DROP 5.8 INCH.WATER


FIGURE 72.  RADIAL PROFILE OF TEMPERATURE OUT OF THE COMBUSTOR  AT

              SEVERAL DIFFERENT CIRCUMFERENTIAL LOCATIONS AND USING

              THE FAN AIR SUPPLY
                                        113

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   3000

   2800

   2600

   2400
                      AVERAGE
                   TEMPERATURE
                         °F
                        2662
                                                     (4.75 R)
   3000  r
                                                                         2743
                                                                         2283
                                                                         1857
   1600
   2000 r
                                                     (2.00 R)
1800

1000
1400
" ^^ ^^ T
^*» ,-— •— 	 * ^
i i i i i i _j
12 3456 78
CIRCUMFERENTIAL POSITION (1.25 R)
                                                                         1704
      FUEL FLOW. 88 LB/HR
      COMBUSTOR PRESSURE LOSS, 5. 7 INCHES WATER
      AVERAGE TEMPERATURE.  2384 "F
HOT SPOT TEMPERATURE, 2920° F
COLD SPOT TEMPERATURE, 1458'F
FIGURE 73.   CIRCUMFERENTIAL VARIATION OF COMBUSTOR OUTLET TEMPER-
              ATURE AT DIFFERENT RADII USING RIG AIR SUPPLIES
                                       114

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  2800



  2600


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    2000
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    1800


    1600


    1400


    1800


    1600


    1400


    1200
        12        34        56

                      CIRCUMFERENTIAL POSITION


        FUEL FLOW, 88 LB/HR

        COMBUSTOR PRESSURE LOSS, 5.8 INCH WATER

        AVERAGE TEMPERATURE, 2335'F
                                                        (4.75 R)
                                                        (4.00 R)
                                                        (3.00 R)
                                                      (2.00 R)
                                                                            AVERAGE
                                                                         TEMPERATURE
                                                                               OF

                                                                             2700
                                                                             2725
                                                                           2190
                                                                           1750
                                                                           1555
                                                      (1.25 R)


                                                   HOT SPOT TEMPERATURE, 2880° F

                                                   COLD SPOT TEMPERATURE,  1315°F
FIGURE 74.  CIRCUMFERENTIAL VARIATION OF COMBUSTOR OUTLET TEMPER-

              ATURE AT DIFFERENT RADII USING FAN AIR SUPPLY
                                          115

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X
z

D
5
   4.0
   2.0
f
CQ
O  1.0
u
     o
                                                         27 LB/HR
                                                         50 LB/HR
                                                        109 LB/HR
                                                         88 LB/HR
       -V:
                                                                               j
     0    1400    1GOO     1800    2000    2200     2400
                                COMBUSTOR OUTLET
                                                             2600
2800   3000
Fuel Flow
Lb/Hr
27
50
88
109
Combustor Loss
Inch Water
0.515
2.05
5.7
8.4
Temperature Spread
8 F
1130
1510
1210
1060
       FIGURE 75.   RADIAL PROFILE AT VARIOUS FUEL FLOWS, USING RIG
                    AIR SUPPLY


 7.3.6  Final Combustor Tests

         Because of the large radial temperature profile adjustments were made in
 the position and number of dilution holes.  Considerable modifications to the shape of
 the radial profile could be made and the final result is shown in Figure 76.  A reduction
 in radial profile spread from about 1300° F to 850° F was made.  In practice, by more
 extensive experimentation, it should be possible to obtain  a flat temperature profile,
 especially if vaporizer inlet temperature retirements can be kept low.
                                       116

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             5.0
             4.0
          u
             3.0
K 2.0
O
w
pa
S 1.0
o
o
FIGUBE 76.
                                                                  \


                                HEF. FIGURE 68
                                PRELIMINARY DESIGN
                                                         FINAL
                                                     DEMONSTRATION
                                                     ARRANGEMENT
                                                               2(JOO
                                                           2800
  1400    1600     1800    2000      2200     2400
         COMBUSTOR OUTLET TEMPERATURE (° F)
RADIAL PROFILE OF TEMPERATURE (FINAL DEMONSTRATION COM-
PARED TO PRELIMINARY)
        Based on previous emission tests the fuel and air metering valves were
adjusted to provide for as optimum as possible air-fuel ratio.  The resultant combus-
tor outlet temperature as a function of fuel flow is shown in Figure 77.  Because the
dilution air represents about 20 percent of the total air flow through the combustor this
temperature could be increased by a similar amount (to about 3000° F maximum) by its
deletion from  the combustor.  Conversely the temperature could be reduced by any
amount by addition of more dilution air.  Theoretically either addition or subtraction
of dilution should not affect emissions.  However,  if temperatures of  3000° F are re-
quired to  minimize vaporizer size, temperature distributions will be  dependent on
mixing in the flame zone and there would be no excess dilution air available for final
trim of the radial profile.

        Emission tests were then run  and the results are shown in Figures 78, 79,
and 80, and also in Figure 81.

        All emissions were substantially below the required limits except for carbon
monoxide  which,  at the one pound an hour fuel flow, was above limits. This was
caused by excessive  air addition (Fig.  77) and, by  small modifications to the fuel and
air controls it should be possible  to achieve the required emission goals.
                                       117

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       2300
       2100
       1*00
       1700
       1SOO
       1300
        900
        700
                  20
                            40   50    60
                            FUEL FLOW (LB/MRI
                                                   90
                                                       100   109
FIGURE 77.  FINAL AIR-FUEL CONTROL SYSTEM. DISCHARGE TEMPERATURE
           VARIATION WITH FUEL FLOW
  UJ
  o>
 1.20

 1.00


0.80

0.60


0.40


 0.20
                            138 NOg  LIMIT
         0    10   20   30   40   50  60   70   80   90  100  109
                         POUNDS PER HOUR  FUEL  FLOW

       FIGURE 78.  EMISSION OF NO2 AS A FUNCTION OF FUEL FLOW
                                118

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           271
         16

         14

         12


         10
        e8
        |e
        i
        8
16.25 CO LIMIT
           0   10   20  30  40   50  60   70   80  90  100  109
                       POUNDS PER HOUR FUEL FLOW

FIGURE 79.  EMISSIONS OF CO AS A FUNCTION OF FUEL FLOW (TEST A)
         0.40
         0.30
         Q20
          0,10
                             0.48  HC LIMIT
            0   10   20   30  40  50  60  70  80  90   100 109
                         POUNDS  PER HOUR FUEL  FLOW

FIGURE 80.  EMISSIONS OF HC AS  A FUNCTION OF FUEL FLOW (TEST A)
                               119

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Fuel Flow pph .
CO
Test A HC
•NOz
CO
Test B HC
NOj
CO
Ted C HC
NOj
(HC not adjusted
for background)
CO
Test D HC
NO,

109
10.4
0.053
0.616
6.17
0.400
1.818
1.41
0,168
1.452
0.322
2 iS
0 053
1.702
EMISSION AS A FUNCTION OF FUEL FLO V IN GM/KGM
100
4.74
0.044
0.754
5.94
0.364
2.100
1.13
0.089
1.542
0.239
2 60
0 048
1.530
90
2.74
0.039
0.978
7.29
0.250
1.921
1.07
0.129
1.528
0.272
0.69
0.048
1.930
80
2.69
0.036
0.872
3.82
0.241
1.840
0.55
0.056
1.5CO
0.197
0.13
0.055
1.702
70
2.80
0.055
0.851
2.54
0.241
1.928
0.14
0.044
0.430
0.191
0 00
0 060
1.708
60
2.75
0.054
0.835
1.79
0.234
1.912
0.07
0.029
0.330
0.172
0 00
0 055
1.640
50
1.34
0.042
0.854
1.21
0.234
1.913
0.07
0.014
1.196
0.156
0.00
0.059
1.540
40
0.99
0.036
0.874
1.02
0.233
1.910
0.07
0.014
1.065
0.158
0.00
0.062
1.511
30
0.91
0.032
0.795
0.09
0.248
1.840
0.00
0.014
1.060
0.157
0.00
0.055
1.388
20
0.86
0.029
0.672
0.10
0.263
2.010
0.08
0.000
1.093
0.123
0 00
0.061
1.351
15
0.79
0.025
0.734
0.09
0.253
1.836
0.09
0.000
1.158
0.112
0 00
0.063
1.270
10
1.07
0.019
0.661
0.10
0.261
1.808
0.10
0.000
1.118
0.116
0 00
0 070
1.172
5
4.33
0.026
1.105
0.14
0.308
2.350
0.20
0.000
1.208
0.184
0 93
0 110
1.831
2
7.80
0.030
1.140
0.21
0.362
2.910
0.27
0.000
1.468
0.218
5 81
0.202
1.108
1
27.10
0.191
1.300
3.04
0.283
2.650
0.48
0.000
1.970
0.330
3 66
0.2S2
1.430
Fuel
JV 5
JP 5
JP 4
JP 5
Test A - Standard configuration, JP 5 fuel maximum heat losses to ambient. • Measured as NO and Reported NOj
Test B - A« with Test A but with minimum possible heat losses to ambient. (Appendix D)
Test C - Primary flame zone leaned out 25%, JP 4 fuel, minimum possible heat losses.
Test D - As with Test C but with JP S fuel.
FIGURE 81.  EMISSIONS AT VARIOUS FUEL FLOWS, WITH DIFFERENT FUELS
             AND AT TWO AIR-FUEL RATIOS, WITH AND WITHOUT HEAT LOSSES


        These tests were done by taking emission samples immediately aft of the
combustor exit.  Heat losses were substantial, especially at low fuel flows, and would
approximate to the heat losses that would occur if a vaporizer had been installed at
this emission sampling point.

        A long mixing duct was installed aft of the combustor, of ten inches length.
This hot duct reduced the view and hence radient losses of the flame to ambient. The
resultant emissions are shown in Figure  81 Test B.  There was a substantial increase
(approximately double) in NOg emissions and  a reduction in CO emissions as would be
expected.  The HC emissions increased and this latter phenomena was unexpected
and unexplainable. Further investigations are warranted.

        Keeping the same  overall air-fuel as previously, modifications were made to
the primary flame zone such as  to increase its air-fuel by 25 percent.   The radiation
shielding was retained and  the results shown in Figure 81, Test D.  The results
were a very substantial reduction in CO and HC emissions such that, for most of the
operational range that would ordinarily be used in city driving, these emissions are
negligibly small.  Emissions of  NO2 reduced  slightly,  and thus indicates that a more
optimum arrangement of air-fuels in the various zones of combustion exists which
could further reduce emissions.

        Maintaining the same lean primary flame zone and radiation shielding the test
was repeated using JP-4 and the results are in Figure 81,  Test C.  Changes in
emissions were noted, of a small amount. At some fuel flows emissions were higher,
at other flows, less than when JP-5 fuel was used, but in general, emissions with
                                     120

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jp_4 were slightly less than with JP-5.  Also shown are the HC emissions not corrected
for background.  (At the test site background HC emissions were high due to the
proximity of the municipal airport.)  The HC emission levels recorded indicate that the
combustion process serves to lower these emissions to less than background.

        The results indicate that emissions can be kept below the goals set, but that
the design of the vaporizer must be carefully integrated with the combustor in order to
avoid excessive heat loss and resultant high emissions.

7.3.7  Combustor Noise

        Certain operating conditions of combined Rankine Cycle fan and combustor
result in an acoustic resonance. A combustor modification solved the resonance
problem,  but unfortunately had  adverse effects on emission levels.  Sound level
frequency analysis of "before-and-after-fix" runs are shown in Figure 82.  The
modification reduced maximum combustor frequency peak from 107 db (sound
pressure level, Ref. 20nN/nO  to 94 db at 109 Ib/hr fuel flow (full power).  Figure 83
distinguishes noise components  of the combustor at an off-exit location.

        The test conditions were:

            •Fan inlet guide vanes:  installed

            • Microphone position and orientation:  (1)  on fan axis three feet in front
             of inlet at 45 degrees to axis for runs shown in Figure 82.  (2)  Data in
             Figure 83 is taken 12.5 feet radially from combustor exit, five feet from
             ground,  45 degrees orientation. (See Fig. 84).

            •Surroundings:  Outside, ten feet from brick wall (See Fig. 84).

            • Equipment  Same  as described for fan noise  tests (Section 6).

           •Background noise:   Figure  82 indicates noise levels with  only fan power
            generator operating.  Figure 83 includes operational fan  as part of
            background noise so that combustor components can be separated.

Table V summarizes the significant frequency peaks of the test series.

Summary  of Results

        1.  Before modification, combustor resonance shows a frequency shift
            from 240 cps to 310 cps with a change  of fuel flow from 20 Ib/hr to
            109.5 Ib/hr.  The  sound level remains approximately the same  at 106
            to 107 db.
                                      121

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  110
  100
J  90
u
s
o
8  80
BURNER RIG OPERATIONAL
 BEFORE MODIFICATION
   SO
                                                  BACKGBOUND (INCLUDES MOTOR GENERATOR SET)
                                         _L
                                               J	L
                                      _L
J	1	1	1	1	1
     80   TO  80   90  100

  nor
  100
                                ISO
                                        200    250   300
                                                              400    SOO   600   TOO  800 900 1000
                        FREQUENCY (opt)
                         BURNER RIG OPERATIONAL
                           AFTER MODIFICATION
   60
   sol	1	L
                                              BACKGROUND
                                      (INCLUDES MOTOR GENERATOR SET)
    60   TO   80  90 100
                                ISO      200     ISO   300
                                           FREQUENCY (cp*)
                                     400    500    600  TOO 800 900 1000
FIGURE 82.   BEFORE AND AFTER MODIFICATION BURNER NOISE AT LOCATION
                "A"
                                            122

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                                  COMBUCTOR. FAN AND BACKGROUND
to
OS
                                           \ BLOWER AND BACKGROUND
               SOt
                250   300      400     500
                                          600   700  800  BOO 1000        1500

                                                            FREQUENCY (cpa)
2000    1500   3000     4000   5000   6000 TOM 8000900010.000
                 7   8   9  10
                                                                                                             150       ZOO    150
                             FIGURE 83.   IDENTIFICATION OF COMBUSTOR NOISE AT LOCATION "B"

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                      ONE STORY BRICK STRUCTURE
         ASPHALT GROUND COVER
 FAN LOCATION BEFORE
COMBUSTOH MODIFICATION
               EXHAUST
                         RANKINE CYCLE
                            BURNER
                           FAN LOCATION AFTER  V
                          COMBUSTOR MODIFICATION
                                                   LOCATION "A"
                                                   MOTOR GENERATOR
                                                   SET
                        12.5'
  OUTSIDE LOCATION
                             LOCATION "B"
                               WOOD CABINET
FIGURE 84.  DIAGRAM OF COMBUSTOR ACOUSTIC TEST LOCATION
                             TABLE V

         SIGNIFICANT NOISE FREQUENCY PEAK LEVELS
Frequency
(cps)
90
240
310
350
900
1750
2700
Before Combustor Modification
After Modification
Inlet Axis Location "A"
20 Ib/hr Fuel Flow
Off Exit Location "B"
109. 5 Ib/hr Fuel Flow
Fan and Combustor
None
106 db
No Peak
94 db
100 db
90 db
None
None
95.5 db
107 db
94 db
101 db
90 db
98 db
94 db
93 db
91 db
94 db
104 db
88 db
94 db
89 db
None
92 db
93 db
88 db
89 db
None
Fan Only
None
None
None
None
88 db
87 db
78 db
Source
Combustor
Fan
and
Harmonics
                                124

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        2.  The combustor modification eliminates the combustor as the dominant
            noise source as measured at location "A". A new 94 db peak at 90 cps
            occurs after the change and  the 107 db peak at 310 cps is significantly
            reduced to 91 db.

        3.  After modification with fan and combustor operating a measurement
            to be  99 db at location "B".  (Measured in the all frequency pass mode,
            flat frequency response mode-weighting contours not used.)

            (It should be noted that the change in the fundamental fan peak level
            at 900 cps from 101 db to 104 db before and after combustor modification
            is due to a relocation of fan  position.  See Figure 84 for details.)

7.3.8  Ignition Tests

        A prime source of hydrocarbon emissions is cold light off.  Ignition must
be instantaneous otherwise resultant emissions are unacceptable.  Throughout all
the tests ignition procedures were monitored and, at all times, good ignition occurred
and no failures to light were seen.  Tests were done, simulating cold day starts at
-40° F with kerosene,by using a heavy, less volatile fuel of 17 centistokes viscosity
(the highest viscosity likely).  Good  ignition could be obtained and the flame perfor-
mance at maximum fueling rate appeared similar to that obtained with JP-5 (no
emission data  was  taken). However, ignition reliability was only 50 percent and it is
suspected that a reduction in fuel flow at  ignition (nominally 10 Ib per hour of fuel) due
to viscous effects on the fuel control system was the cause of this.  At high fuel flows
this viscous effect  is reduced (it is a function of Reynolds number) and fuel flow was
satisfactory.   Some further investigations are indicated.

7. 3. 9  Aldehydes and Smoke

        No instrumentation was available to measure either aldehydes or smoke.
Throughout the tests  reliance had  therefore to be placed on sight and smell to assess
these emissions. In  steady state operation  the flame was odorless and smoke free.
Indicating that aldehydes and smoke  were at a very low  level or perhaps nonexistant.
During transient decelerative operations, at low  levels  of fuel flows it was seen that
some smoke was emitted though emissions of CO, NC>2  and HC were maintained low.
The level of smoke was a direct function  of  deceleration and  it was therefore a likely
result of differing response characteristics of the fuel and air flow.  The problem was
minimized by controlling the deceleration rate in transient tests as a function of fuel
flow.  At high  flow rates deceleration was maintained at a higher rate than at low flows.
This fits in well with the response requirements  of the  system in that vaporizer response
at high fuel rates must be quicker than at low fuel rates.  Future work must involve
complete package testing if minimum smoke is to be obtained and synchronization of
the air and fuel metering systems must be improved.


                                      125

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                                     8
              OPTIMUM DESIGN APPROACH FOR RANKINE CYCLE
                             COMBUSTION SYSTEM
        Based upon the results of the analysis,  design and test program conducted on
the demonstration system a number of specific conclusions may be formulated:

           • The 1980 AAPS emission levels for automotive vehicles are feasible.

           • A relatively small combustion system incorporating hydromechanical
            control components can  operate at high response across a fully mod-
            ulated heat release rate of 100 to 1 (2 x 106 to 2 x 104 BTU/hr).

           •Emissions during the  required high frequency automotive startup,
            power level transients and shutdown cycles can be maintained below
            emission level goals (start up in 3  seconds, combustor only, transient
            response 50%/sec).

           •Air flow changes resulting from voltage, leakage and efficiency changes
            can be compensated to maintain correct air-fuel ratios for low emissions.

           •Heavy fuels such as Diesel No. 1, Jet A, Kerosene or JP-5 can be
            burned across the full flow range.

           •Parasitic power can be minimized at part loads by the fuel delta-P
            compensator.

        A set of design factors determined to be of major importance in the develop-
ment of a low emission combustion  system have  also been identified.  Listed below
are some of the major design guides determined by analysis and test results.

           •Precise control of combustion in three axial zones with rapid time
            transitions between primary and  secondary zones is  necessary.

           •Mixing must be rapid and uniform.
                                      127

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           • Overall air-fuel ratio must be precisely controlled across the entire
             heat release range.

           • Uniform air distribution into the air valve and combustor is essential
             for low emission combustion.

           • Flame radiation losses are important factors in emission levels.

           •High response of both air and fuel flow control is necessary throughout
             transients.  Fan speed control is probably inadequate due to inertia!
             lags.

            .Air valve appears to require flow  symetry to obtain 100 to 1 flow control.

            • Fan design  and configuration'has a major effect upon the air valve design
             and combustor performance.

            .-Vaporizer effects on aerodynamics and  radiation flame radiation heat
             loss may also have significant effects upon the design and emission
             performances of the combustion system.

        An optimum system  can be synthesized based upon results of the demonstration
system tests.  A major factor in determining the package configuration of the demon-
stration system  was component availability and the need for design flexibility in this
highly developmental program. Although the best fan found to be readily available
has low volume (0.2 cubic feet) and low weight (13 pounds), it is not optimum. Initially,
it was designed for  aircraft applications.  Weight and volume  are emphasized in these
designs with noise only as a secondary consideration. Thus,  the unit is a small diameter,
high speed fan having a high dynamic head. In order to obtain accurate air regulation
across the 100 to 1  range,  the adverse aerodynamic effects of the dynamic head must
be eliminated.  To do this, a relatively large volume cavity at the discharge side of
the fan 10  needed to turn  the flow and reduce its velocity. Another design feature that
would also be eliminated  in a production type unit would be a separate  drive motor for
1he rotating cup  atomization system.  Its use in the demonstration unit is necessary
due to rotational speed requirements and need for initial independent optimization of
the atomization system.  It also added to the length (~4  inches) and  overall volume of
the complete system.

        The package design depends critically on the overall  pressure drop through
the system.  The use of a low boiler pressure drop results in a different system than
with a high boiler pressure drop.  The effect on the combustor and  control design
philosophy is slight and it appears likely that the best system  with either high or low
loss boiler will incorporate a fan of larger diameter than currently used and which
                                       128

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will result in considerable improvements in performance and substantial reductions
in length.  The presence of a vapor generator could modify flame performance
depending upon its design.  By proper integration of the vapor generator, flame per-
formance may not be effected and may even be improved.

        Several significant areas for improvement can be identified by matching fan
size and speed to the geometry and flow requirements of the air valve and combustion
system.  Optimizing the system will reduce volume, power and noise significantly.
An optimum configuration of this type is  compared to the demonstration system in
Figure 85.

        The optimum system considered is based on slight modifications to the
demonstration system design.  Basically a better aerodynamic interface is the key
element in the optimum configuration. Lower turning and metering losses are
expected together with a corresponding reduction of parasitic power.   A small
reduction in combustor pressure drop also appears quite feasible by increasing the
outside diameter of  the basic combustor.  As a result of use of a large diameter fan
and a combined fan motor rotating cup drive, the overall length and volume can be
reduced to as low as 14 inches.  Parasitic power loss at full load could be as low as
1. 25 HP with part load parasitic power drain well under 0. 5 HP if voltage is reduced
at lower power demands.
                                      129

-------

f
.1





\
/
l.ll! FT'1
COMBUSTOR
PLl'S CASE








V CUP
' MOTOR

12.
[*- T ^

1 \
FAN
MOTOR
». . Ili FT'1
AIR VALVE
1 /

1
Jr
1.17 F
1
FAN
                                                                        • FAN T1LADE
                                                                          FAN STATOH
                                                                          -CONTROL SECTION
                                                          PANCAKE
                                                          DC MOTOR
                                                         FUEL INLET
                 TOTAL VOLUME 1. R9 FT3
                 DEMONSTRATION SYSTEM
                                                                       Cl'T>
                                                                                        13.11- DIA
14.2.1
                                                                      OPTIMUM SYSTEM

3
Total Volume (ft )
Length (Inch)
Diameter (Inch)
3
Combustor Volume (ft )
Combustor Diameter (Inch)
Horsepower
Combustor Loss (Inch Water)
Diffuser Loss (Inch Water)
Metering Loss (Inch Water)
Overall Pressure Loss (Inch Water)
Fan Efficiency (%)
Motor Efficiency (%)
Overall Efficiency (%)
Motor Speed (rpm)
Demonstration
System
Design
1.89
31.0
13.00
0.687
11.00
2.30
8.00
3.00
2.00
13.00
75.0
75.0
56.2
14000
Optimum
System
Design
1.09
14.25
13. 00
0.687
12.00
1.25
6.0
0.5
1.5
8.0
85.0
75.0
68.0
7000
Spec
1.33



2.00








Note:  Ignition,  Fuel and Air Regulators not included.
            FIGURE 85.   TWO FAN SKETCHES - PRESENT AND OPTIMUM
                                           130

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APPENDIX A

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                                 APPENDIX A

            EMMISSION MONITORING EQUIPMENT AND PROCEDURES
        Emission measurements were taken using three Beckman Model 315A infrared
analyzers,  Figure A-l,  and a Beckman Model 402 hydrocarbon analyzer, Figure A-2.
These instruments provided continuous and automatic determination of the exhaust
components.

        The infrared analysis system is based on a differential measurement of the
absorption  of infrared energy.  An  infrared radiation source is transmitted through two
long cells one containing the exhaust sample and the other a reference gas.  In oper-
ation, the presence of the infrared  absorbing component of interest in the sample
stream causes a difference in the radiation absorption levels between the sample
and reference sides of the system.  Due to this difference the gas in the reference cell
is heated more,  thus raising the  pressure and causing a metal diaphragm to distend.
This metal  diaphragm is part of a capacitor circuit, as the IR source beams  are
alternately  blocked and unblocked,  it pulses, thus causing a cyclic change in the
detector capacitance  (Luft Principle).  The resultant signal is then  routed to the am-
plifier control section, and finally to the recorder.

        The hydrocarbon sensor is a burner where a regulated flow of sample gas
passes through a flame sustained by regulated flow of a fuel gas and air.  Within the
flame, the  hydrocarbon components of the sample stream undergo a complex ionization
that produces electrons and positive ions. Polarized electrodes collect these  ions,
causing current to flow through measuring circuitry located in the electronics unit.
The 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
original sample.

        The following conditions were employed for all analysis:

                 Gas          Detection             Range
                 NO         IR-41" cell      0-150 ppm
                 CO         IR-10" cell      0-1000 ppm
                 CO2        IR-0.25" cell    0-16%
            Hydrocarbons   FID            0-500 ppm, 0-10 ppm
                                      133

-------
                                  * • '  J « > t i t » »""
                                                          trie Oxid
                                                          Inch Cell
                                Electronic Units
Carbon MonoxideBCarbon Dioxide
   2.5 Inch CeUl»).25 Inch Cell

  10.0 Inch Cell
  FIGURE A-l.  BECKMAN MODEL 315A INFRARED ANALYZER
                             134

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cc
01
                          FIGURE A-2.   BECKMAN MODEL 402 HYDROCARBON ANALYZER

-------
        Each instrument was calibrated and maintained per manufacturer's recomm-
endations.  Calibration curves supplied by Beckman were cross checked each day with
known gases to determine the validity of these curves. The NDIR unit was kept on at
all times in order to permit maximum stabilization.  No  difficulties or large variances
were encountered in maintaining day-to-day gain settings, and zero settings.

        Certified calibration gases as received from the vendor were within ±0.5
percent of the stated values with the exception of the nitric oxide standards.  Nitric
oxide calibration gases have  an analytical accuracy of ±5.0 percent.  Each NO stand-
ard gas was analyzed by the modified Saltzman* method employing the evacuated
bottle sampling technique.

        Table A-I shows the nitric oxide values obtained by the Saltzman method as
compared to the vendor's analysis.  (Average of a minimum of 5 determinations.)

                                  TABLE A-I

                        NDIR COMPARED TO SALTZMAN
Stated Value
140
105
53
Infrared
	
100 ppm
45 ppm
Saltzman
144 ppm
118 ppm
47 ppm
        The flame ionization hydrocarbon analyzer was calibrated immediately prior
to each run and set up according to the manufacturer's specifications.  Eight ranges
are available on the instrument varying from 0-10 ppm to 0-50, 000 ppm.  All deter-
minations required the lower two ranges. The heated sample line was maintained at
350° F and the detector at 400° F to avoid condensation problems.

        Intensive investigations were performed at the start of the program to
evaluate the sampling technique.  The following facts were observed:

            •Each analyzer unit required its own probe.  Attempts to tee off from
            one probe showed adverse flow effects.
* Cornelius, W.,  and Wade, W. R., "The Formation and Control of NO in a Regen-
erative Gas Turbine Burner", SAE Report No.  700708.
                                     136

-------
           • A filter (glass-wool)-dryer (Drierite-nonindicating) was used between
            the NDIR and the probe.

           • Sample line length (20-200 ft) had little if any effects on the measure-
            ments, other than flowrate.

           • The use of the pre-drier, plus the refrigerator and additional NO drier
            were more than adequate to remove all the water.  (Aquasorb was used
            just prior to the NO cell).

           • An air-quenched averaging probe was determined suitable for the
            combustor tail pipe.

        The accuracy of the NDffi is dependent upon the accuracy of the calibration
gases used.  Each of the calibration gases were analyzed by the  supplier using "gold"
standards.  These primary standards were prepared using National Bureau of Stand-
ards weights.

        Reproducibility of the instrument is within one percent.  Reproducibility
tests were performed on separate days and under varying  sampling conditions.  The
results showed  that the measurements were reproducible.

                                  TABLE A-II

                           REPRODUCIBILITY TEST
Power
Setting
10-28-29
50%










75%
1
<
2|
f
>%
10-29-70
50%
'1
>%
25%
ppm NO

56
61
60
60
56
59
64
55
47
47

59
47
62
56
49
47
ppm CO

128
120
120
128
120
120
282
305
53
53

120
143
295
335
45
38
% C02

7.9
7.9
7.8
7.9
7.7
7.7
7.7
7.9
6.8
6.5

7.5
7.9
7.8
7.9
6.0
6.5
3
Flow ft /m.

4
2
1
4
2
3
3
3
3
3

3
3
3
3
3
3
Remarks

Probe Preheated



~200' Sample Line
Probe Not Heated
Probe Heated
*200' Sample Line

Probe Heated
Probe Unheated

Probe Heated
Probe Unheated
Probe Heated
Probe Unheated
Probe Heated
Probe Unheated
                                      137

-------
        Because of the very low emission levels obtained with this combustor (com-
pared to Otto cycle engines), special attention was required to calibrate the measure-
ment system at the low ranges.  Calibration gases in the actual ranges being recorded
were obtained to establish the validity of the instrument deflection  to volumetric
concentration supplied by Beckman. Gases used for calibration were certified by
the span gas suppliers to be within  the following concentrations:

        Nitric oxide
             Concentration               Certified Accuracy

             10.5 ± 0.5 ppm              ± 5. 0% of component
            103.0±2.1ppm              ± 2. 0% of component
            140.0 ±2.8 ppm          .    ± 2.0% of component

        Carbon dioxide

             Concentration               Certified Accuracy
             14.31 ± 0.29%               ± 2% of component
               4.31 ± 0. 09%               ± 2% of component
               2.10 ± 0. 04%               ± 2% of component
               1.12 ± 0. 02%               ± 2% of component

        Carbon monoxide
             Concentration               Certified Accuracy

             115 ppm ± 2 ppm             ± 2% of component
             225 ppm ± 4 ppm             ± 2% of component
             850 ppm ± 17 ppm            ± 2% of component

        (Recorder deflection error is ± 0. 5% full scale.)

        Results of calibration tests are shown plotted in Figures A-3, A-4,  and
A-5.  The basic curve was supplied by Beckman. Data points used by Beckman to
verify the basic curve slope are shown as open circles.  Solar calibration points
including the zero span gas point are shown  with the closed circles. Good correlation
was obtained throughout the range required to measure emissions with the Rankine
cycle combustor.
                                     138

-------
    10(1  r
     so
   z
   O
   p GO
   u
   u

   u.
   u
   c
            •  SOLAR CALIBRATION POINTS

            O  IJKCKMAN SUPPLIED CALIBRA-

               TION POINTS

            Q  ZERO GAS POINT
     4(1
                    .10          100


               PPM NO IN N BY VOLUME
  FIGURE A-3.   NO CALffiRATION RESULTS
100
 80
       •  SOLAR CALIBRATION POINTS

       O  BECKMAN SUPPLIED CALIBRATION POINTS

       D  7.ERO GAS POINT
 60
U
U
J
u.
u
D

C
U
c
s.
 40
            50
                      100       150       200


                    PPM CO IN N0 BY VOLUME
                                                  250
  FIGURE A-4.   CO CALIBRATION RESULTS
                         139

-------
mo r
                 • SOLAR CALIBRATION POINTS
                 O BECKMAN SUPPLIED CALIBRATION POINTS
                 g ZERO CAS POINT
            4.0      8.0       12.0      l(j.O
           PERCENT CO, IN NO BY VOLUME

FIGURE A-5.   CO2 CALIBRATION RESULTS
                      140

-------
APPENDIX B

-------
                                 APPENDIX B
                         TEST FUEL SPECIFICATIONS
        The fuels specified to be used for combustion tests are #1 Diesel, kerosene
and Jet A.  These fuels can vary widely  in their combustion characteristics because
their specifications are loose.  Refineries in differing locations can produce different
fuels,  meeting the same spec, because their production is guided by economics dictated
by local conditions of demand and the type of crude that is available.

        Solar is  using JP-5, which is a more rigorously controlled fuel than those
specified by EPA.  It is a fuel that typically will be more difficult to burn than  those
specified.  In addition, JP-4 and #2 diesel will be used to provide information with
fuels that are respectively considerably  more and less difficult to burn than those
specified (Fig. B-l).

        The Government has recognized this problem of fuel variability in regard to
the testing  of diesel engines* and has proposed a fuel spec for #1 diesel as follows:
                                            Specification
         Fuel Property
Cetane
Distillation ° F
  Initial Boiling Point
          10%
          50%
          90%
        End  Point
Gravity °API
Total Sulfur %
Aromatics %
Paraffins, Napthenes, Olefins
Flash Point  °F (min.)
Viscosity  (Centistokes), at 100° F
ASTM D-975
 Grade 1-D
   40 min.
  550 max.
  0. 5 max.
 100 or legal
   1.4-2.5
    P roposed
Government Spec

     48-54

    330-390
    370-430
    410-480
    460-520
    500-560
     40-44
     .05-. 2
      8-15
   Remainder
      120
    1.6-2.0
MILrT-5624
Grade JP-5
                                   400 min.
                                   550 max.
 140
 * Federal Register,  Volume 35, Number 136,  Part II.
                                      143

-------
   700
   HOO
   .100
H
PS
D
H
U
Pk
w
H
   400
   200
   100
                           OF
                     SPl
               SPECIFIEI
               JET A Al
BAD IN DISTILLATION
FUELS (KEIOSENE,
 #1 DIESEL)
                20         40        60
                       PERCENT EVAPORATED
                                 80
                   100
  FIGURE B-l.
VARIATION OF ASTM DISTILLATION TEMPERATURES
FOR THE TEST FUELS (AVERAGE VALUES)
                             144

-------
The proposed fuel is more rigorously controlled than current ASTM specs, agrees
closely to the JP-5 fuel used by Solar and is representative of a typical grade of kero-
sene that could ordinarily be available.  As the combustion characteristics of a diesel
engine are different from those of an atmospheric combustor, it is likely that the
specification above should be modified so as to highlight any combustion difficulties
inherent in an  atmospheric flame.  Items that warrant more specific control are noted
below with the reasons for such control

            •Flash Point, °F, 140°F  minimum, 150°F maximum
            A high flash point makes ignition  more difficult and therefore should
            be controlled to the high end of typical practice.  Current specs for
            fuels used have a maximum of 150° F but no bottom limit.

            •End Point, °F, 550° F minimum,  600° F maximum
            A high end point tends to carbon build and smoke in exhaust.  Current
            specs allow 575 ° F maximum but a bottom limit is not specified.

            •Aromatics, 20  to 25%
            High aromatics tend to cause smoke and carbon  build.  No specs
            currently apply and typically are  lower than shown.

            •Viscosity, 2.0  to 2.5 centistokes
            High viscosity makes control of fuel more difficult and makes for
            problems of fuel  atomization with certain  types of combustion systems.
            Typical values  are less than shown and the range is 1.4 to 2.5.

            •Olefins and Diolefins, 5  to 10%
            High diolefins tend to cause deterioration  of fuel with resultant gum
            formation.  This causes unreliability in fuel metering and fuel injection.
            No specs exist  and typical values are  below 5%.

        Tests have indicated that most of the problems seen with burning a heavy
grade of fuel such as JP-5 will tend to diminish with the lighter grades of kerosene.

        The current specifications of the fuels to be used  and of JP-5 are as
follows:
                                      145

-------


Flash Point
10% Point
50% Point
End Point
Jet A
(ASTM D1655)
110-150°F
400°F Max.
450°F Max.
550° F Max.
Kerosene
(VV-K-211)
115° F Min.


572°F Max.
Kerosene
(VV-K-220)
125°F Max.


510°F Max.
JP5
(MIL-T-5624)
140° F Min.
400°F Min.

550°F Min.
Of these fuels, Jet A and JP5 are the most tightly controlled.  JP5 will inevitably be
heavier than Jet A as flash point and 10 per cent point tend to be higher.

        The specifications for the fuels to be used to monitor the limits of typical
fuels are as follows:

                                                       JP4
              #2 Diesel (ASTM D975)               (MIL-T-5624)
            Flash Point 125°F or Legal         20% Point 290°F Max.
        90% Point 540°F min., 640°F max.      50% Point 370°F Max.
                                              90% Point 470°F Max.
                                       146

-------
APPENDIX C

-------
                                  APPENDIX C

                       FAN NOISE REDUCTION METHODS

        The National Air Pollution Control Administration Division of Motor Vehicle
Research and Development has recently established a goal of 77 dbA as the maximum
noise generated by the vehicle.  Either the combustor or compressor  fan are likely to
be the critical noise source in the Rankine cycle vehicle.  Thus,  a detail analysis of
the fan and its noise generation mechanisms is necessary.  Results of noise measure-
ments on the fan currently being used (because of availability) are analyzed in Sections
6 and 7.  These results in general indicate that significant noise  reduction would be
necessary in a production system.   An optimum low noise fan could be designed also
to provide better aerodynamic and mechanical matching to the air valve  and combus-
tor.  A design analysis indicates that this could be accomplished by incorporating the
following changes:

           •  Increased diameter (up to 13 inches O. D.)

           •  Decreased rotational speed

           •  Decreased pressure rise

           •  Larger number of blades

           •  High vane to rotor blade ratios

           •  Increased spacing between blades and rotors

        An investigation into the mechanism and relative noise improvements that
can be expected has been initiated.   Recent advances in aerospace acoustic analysis
technology  have made the optimum  design approach method relatively  clear.  Up to
the recent work in acoustics on turbo fan engines, the best empirical method to pre-
dict noise of fans was:

             PWL =  90+10  log HP + 10 log Ps                                 (1)

or           PWL -  55 + 10  log Q + 20 log Ps                                   (2)

or           PWL = 125 + 20  log HP - 10 log Q                                   (3)

where       PWL - db - overall sound power level re 10

              HP - hp - rated motor horsepower

                Q - cfm - fan discharge flow
                                      149

-------
             P -in. HO - fan total to static pressure rise *
              S      £

        These equations indicate the importance of maintaining as low a pressure rise
     as possible.  A more comprehensive approach to axial vane fans has recently been
published by M. J. Benzakein and S.B. Kazin**.  The following important design infor-
mation has been abstracted from this work.

        A study of various fan/compressor noise reduction methods is presented.
The analytical treatment of the basic mechanisms of fan/compressor noise generation
is described.   The results are presented in parametric form  and indicate the effects
of fan/compressor design, number of blades,  vane/blade ratio,  aerodynamic para-
meters, and blade row spacing on pure tone noise reduction.  These results  are based
on non-steady aerodynamic treatment of wake and potential interaction effects and
theoretical extensions of spinning mode theories.

        A listener in  the vicinity of an axial-flow fan may distinguish two distinct
sound components known as "discrete-frequency noise" and "broadband noise".  The
first component consists of a number of pure, or nearly pure, tones which combine to
form a high frequency noise, best described as a whine.  The second component is a
background hissing noise,  caused by a superposition of sounds over a continuous band
of frequencies from the lower audible range to the higher, and without pronounced
peaks at any particular frequencies.

        The relative importance of the two noise components depends on the type of
fan or compressor. Noise from a many-bladed fan, working  at subsonic tip  speed  in
an unobstructed airflow, has broadband characteristics.  Noise from a high speed
propeller has predominantly discrete-frequency characteristics.  From a compressor
in which the rotor interacts with stators, or a fan with bearing support struts or other
obstacles near the rotor face, the noise is  a mixture of the two components.   General-
ly, however,  the discrete-frequency noise  dominates the frequency spectrum.  The
object of this  paper is to present different methods by which the blade passing freq-
uency tones can be reduced.

BASIC NOISE GENERATION MECHANISMS

        The complexity of the fan/compressor noise generation phenomena lead many
researchers to a largely empirical approach to the problem.  Some broad understand-
ing of the various sound sources has been derived from experimental data in the last
*  Beranek, L. L., "Noise Reduction",  McGraw Hill, New York, 1960.

** ASME 69-GT-9 Fan Compressor Noise Reduction, March 1969, M. J. Benzakein and
   S. B. Kazin.

                                      150

-------
ten years.  It is felt, however, that a basic knowledge of the different noise sources
is indispensable if fan/compressor noise is to be reduced at the source.

        Different mechanisms are involved in the pure tone generation in fans and
compressors.  These different mechanisms vary in importance from configuration to
configuration, and in a design,  from speed to speed.  Each particular mechanism can
become the main noise contributor for a particular fan design at a particular speed.
The major mechanisms are defined later.

Rotor Alone Noise

        "Rotor alone noise" arises from the pressure field that surrounds each blade
as a consequence of  its motion.  In a moving blade, the pressure distribution on each
section along the blade span produces force fluctuations on the surrounding air.  The
force produced on the air by each blade is equal and opposite to the force produced on
each blade by the air, and the latter force can be resolved into lift and drag com-
ponents of the force  on the blade along its aerodynamic axis.

        The rotor alone noise is similar in nature to the propeller noise which has
been extensively studied by Gutin,  Garrick, and Watkins, and other investigators.
The propeller noise theories are based,  primarily, on three mechanisms:  (a)  thick-
ness noise,  (b) lift noise, and (c) vortex noise.   The experimental results on fan/
compressor noise showed however, that the lift  (blade loading) portion is the primary
contributor of the rotor alone noise. Attention has therefore been directed towards a
prediction of rotor noise due to steady aerodynamic loading.   The analysis consists
of an extension of Gutin's work that includes the effect of a many-bladed rotor and the
presence of the duct. It was assumed in this work that the blades Jo not interact.
That is, each blade carries its own discrete pressure profile,  and this periodic dis-
turbance (in an absolute frame of reference) is mathematically described as an im-
pulse occurring at the blade passing frequency.

Wake Interaction Noise

        Turbomachinery aerodynamicists have long recognized that the airfoils
arranged in rotating  and stationary cascades of axial flow machinery, do not operate
in steady flow. A distinction should be made, however, between the unsteadiness in
the flow field created by  the presence of an adjacent blade row which is characterized
by the generation of pure tones at the blade passing frequency, and the unsteadiness
due to the flow turbulence along the airfoils and in their wakes which represents the
primary source of broadband noise in the machine. The broadband noise generated,
is in general,  of a lower order of intensity.  A brief look will first be taken at the
noise created  by the  interaction of the wakes shed by a stationary or rotating cascade
with the following blade row.   The wakes leaving a rotor or  stator row necessarily
impinge on the adjacent blade row as they pass downstream.   Within the wake there
is a reduction in velocity, and the primary effect of this velocity defect is to cause a
fluctuating incidence at the downstream blade.  This fluctuating incidence gives rise

                                       151

-------
to a fluctuating force which results in sound radiation.

PARAMETRIC STUDY OF ROTOR ALONE NOISE

        Sound pressure levels generated by a rotor alone are a direct function of the
circulation around the blade row. The larger the circulation the higher  the pure tone
levels.  This is consistent with aircraft propeller experience. A study has been made
to investigate these effects,  parametrically, from a turbomachinery standpoint.  Some
of the results are shown in Figures C-l and C-2.

           •  Figure C-l shows that if the tip speed is held constant and  the pressure
            ratio is increased,  the fundamental frequency sound power levels
            increase.

           •  Figure C-2 shows that as the number of rotor blades is increased,
            the sound power levels generated decrease.  This seems to indicate
            that a high blade  design is favorable.

PARAMETRIC STUDY OF INTERACTION NOISE

        A parametric study was carried out to determine the functional relationship
between the pure tone noise generated in the fan by the wake interactions and the pri-
mary turbomachinery aerodynamic and geometric parameters.  This study was re-
stricted to fans and compressors without inlet guide vanes which, at the present time,
tend to produce lower noise levels than comparable machines with inlet guide vanes.
The study was also directed towards designs incorporating large blade row spacings,
where the wake and not the potential interaction is the major source of noise.

Pressure  Ratio Effect
        In an initial study, the tip speed of the machine and the fan geometry (num-
ber of blades and vanes,  spacing/chord,  and so on) were kept constant.  Different
designs with pressure ratios varying from 1.2 to 1.4 were investigated.  When the
pressure ratio is increased at a particular speed, more turning has to be done in the
blade row and the loading goes up.  The terms V1/V2 abd sin 8 in the expression for
the coefficient of unsteady upwash subsequently increase; this is translated into an
increase of pure tone levels.  It can be seen from C-3 that when the fan design pressure
ratio is increased from 1.2 to 1.4,  the fundamental blade passing frequency power
levels increase about 8 db for a fan of constant size, and 4 db for  a fan of constant
thrust.  These results indicate that pressure ratio is an important parameter that
cannot be neglected.  The use of simple correlation formulas that do not take this
effect into account may lead to erroneous results.
                                      152

-------
         14
         12
a
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u

u

o:

£

s
Q
       2
       u
         10
            1.2
                               CONSTANT SPEED

                               CONSTANT NUMBER OF VANES
                               i.3


                           PRESSURE RATIO
                                                   1.4
   FIGURE C-l.   EFFECT OF PRESSURE RATIO ON ROTOR ALONE

                 BLADE PASSING FREQUENCY NOISE
tn
•a
u


i
Q

3
O
in



I
u
K
   2
  -2
   -8
                                       CONSTANT PRESSURE RATIO

                                       CONSTANT TIP SPEED
     20
            30             40


                    NUMBER OF BLADES
                                                50
                                                              60
  FIGURE C-2.
         EFFECT OF NUMBER OF BLADES ON ROTOR

         ALONE BLADE PASSING FREQUENCY NOISE
                              153

-------
 CQ
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 >
 u
 cc
 u

 I
 Q
 2
 I
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                                         ROTOR/OGV INTERACTION
                                       CONSTANT SIZE
                               CONSTANT THRUST
                                        CONSTANT PRESSURE RATIO
                                        CONSTANT SIZE
                                        CONSTANT TIP SPEED
                                        CONSTANT SPACING/CHORD
      1.20
                  1.25
     1.30
PRESSURE RATIO
            1.35
                   1.40
       FIGURE C-3.
                  EFFECT OF PRESSURE RATIO ON INTERACTION
                  NOISE GENERATED AT THE BLADE PASSING
                  FREQUENCY
n
•o
J
u
>
u
j
K
U

I
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C/3
U

H

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K
 -5
-10  —

CONSTAN

IT PRESSUR
^
E RATIO
CONSTANT VANE/BLADES
CONSTANT TIP SPEED
CONSTANT SIZE
CONSTANT SPACING/CHORD
i
\


ROTOR/
^>

'OGV INTER
X

ACTION
^

     10
              20
                    30
40
50
60
70
80
                             NUMBER OF BI.ADES
       FIGURE C-4.
                  EFFECT OF NUMBER OF BLADES ON INTERACTION
                  NOISE GENERATED AT THE BLADE PASSING
                  FREQUENCY
                                 154

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Number of Rotor Blades Effect

        The number of rotor blades is an important parameter in the noise generation
and can easily be modified (within certain vibration and aerodynamic bounds) to suit
the acoustic designer.  Should fan designs,  therefore, be oriented toward a high or a
low number  of blades? The  final answer depends upon the fan size.  The number of
blades and the fan rpm will determine the pure tone frequency, which should be kept
out of the  critical area of the NOY curve.  Therefore, the size, as well as the blade
tip speed of  the fan, must be known  before a final selection of the number of blades
can be made.  An investigation can be done, however, of the effect of the number of
blades on  the sound power levels generated at the blade passing frequency.  In the
following study, the pressure ratio, tip speed, blade row spacing,  and the vane/blade
ratio were kept constant and the number of blades were varied from 20 to 80.   The
results are  shown in Figure  C-4. It can be seen that a design incorporating a high
number of blades will reduce the fundamental frequency tones.  The fan size and rpm
will determine the optimum  configuration from a PNdb viewpoint.

Vane/Blade  Ratio Effect

        The analysis of the  sound generation shows that the  number of interaction
diametral modes (or in other terms, the vane/blade combination) has a major effect
on the blade passing frequency tones.   Figure C-5 shows  that a high vane/blade ratio
can be beneficial, not only from a sound transmission, but from  a sound generation
viewpoint, as well.

Blade Row Spacing Effect

        It is well known that increasing the spacing between blade  rows will decrease
the interaction noise.  Several experimental investigations have determined the reduc-
tion in pure  tone levels that can be obtained when the blade row spacing is increased.
Some researchers  show a 2 db reduction in sound pressure levels per doubling of the
axial separation, while others prescribe 4 or even 6 db SPL.  This inconsistency can
be explained first by the fact that several other parameters besides axial spacing are
involved in the interaction process,  and second, because  the concept of "per doubling"
is not strictly correct.

        In the present study, a particular rotor/OGV configuration was chosen and
the spacing varied  from 0.1  to 2. 0 spacing/chord ratio.   The results are shown in
Figure C-6.   It can be seen that an appreciable reduction in pure tone power levels
can be obtained with large blade row spacings.
                                      155

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                     CONSTANT SIZE
                     CONSTANT TIP SPEED
                     CONSTANT SPACING/CHORD
    1.0
         1.4         1.6

         VANE/BLADE RATIO
                                               1.8
                                                         2.0
  FIGURE C-5.   EFFECT OF VANE/BLADE RATIO ON INTER-
                ACTION NOISE GENERATED AT THE BLADE
                PASSING FREQUENCY
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                  .5            1.0           1.5
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                                                         2.0
   FIGURE C-6.
EFFECT OF SPACING ON INTERACTION NOISE
GENERATED AT THE BLADE PASSING
FREQUENCY
                              156

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SUMMARY AND CONCLUSIONS

        Some theoretical methods have been developed to predict fan and compressor
sound power levels generated at the blade passing frequency.  These prediction tech-
niques have been used to study the effects of different aerodynamic and geometric
parameters on fan and compressor noise.  The results indicate that low sound power
levels will be obtained for:

           • Low pressure ratios

           • Low rotor blade loadings

           • Low rotor diffusion factors
           • High number of rotor blades
           • High vane/blade ratios

           • Large blade row spacings

        An appreciable reduction in pure tone levels can be obtained by judicious
selection of design parameters.  This reduction of noise at the source can be quite
attractive and should be incorporated in low noise fan and compressor designs.

        By incorporation of this latest state-of-the-art design data, an optimum fan
could be designed to interface with the Rankine combustor.
                                      157

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APPENDIX D

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                                 APPENDIX D

                          EMISSION DATA REDUCTION
        The raw emissions data are read from the strip chart on the Beckman
instrumentation cart and converted to observed values in parts per million (ppm) by
volume or percent by volume from the previously obtained calibration curves for the
instruments.

        These observed values for carbon monoxide (CO), nitric oxide (NO), carbon
dioxide (CO2), and unburned hydrocarbons (H/C) together with the other pertinent
rig operational parameters are utilized as input data for the data reduction program
carried out on an IBM 360 Model 40 computer.

        The program performs several correctional operations to the volumetric
concentrations before conversion to mass concentrations.

            • The observed CO and NO readings are corrected for other gaseous
            component interferences which produce higher readings. The  inter-
            ference correction factors are obtained from the emissions instrumen-
            tation manufacturer.

            • A correction is made to the  nitric oxide value for ambient humidity.
            This factor is contained in Section 1201.86 of the February 27,  1971
            Federal Register (Volume 35, Number 40), and is normalized  at a
            humidity corresponding to a water content of  75 grains water per pound
            of dry air.

            • An  exhaust water vapor correction is applied to put  the volumetric
            concentrations on a wet rather than a dry basis.  The wet volume con-
            centrations are used for conversion to mass concentrations and flow
            rates of the individual species.  The correction factor is calculated as
            a function of test point air-fuel ratio and fuel ultimate analysis.

            • The volume concentrations are converted to unity equivalence ration,  i.e.,
            are quoted at a stoichiometric air-fuel ratio. This  correction is thus a
            ratio of the actual test air-fuel ratio to the stoichiometric  air-fuel ratio,
                                      161

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            the latter being calculated from the fuel ultimate analysis.  The volume
            concentrations are normally quoted as ppm or percent dry, corrected to
            stoichiometric (and to standard humidity conditions in the case of NO).

        As noted, the wet volume  concentrations are used to convert to the mass
concentration using the test air-fuel ratio and the appropriate specie molecular
weight.  The molecular weight of the exhaust gas is assumed to be that for air.

        A calculated value of CO^  is obtained from a knowledge of the test air-fuel
ratio and the fuel compositions assuming complete  combustion.  This calculated CO2
value is compared to the measured value and in this manner any spurious results due
to irregularities in the air, fuel, or emissions measurements can be eliminated.

        It can be seen from the attached sample sheet of program output that the
measured nitric oxide volumetric concentration is expressed on a weight basis in
terms of the oxidized product,  nitrogen dioxide  (NC^).  This assumes that all the
nitric oxide ultimately  reacts in the atmosphere to  nitrogen dioxide.

        The unburned hydrocarbons are quoted on  a weight basis as an "average"
hydrocarbon, CH.. „_,  as contained in the Federal  Register requirements for light
duty vehicles.   This pseudo hydrocarbon was proposed for gasoline fueled power
plants and is unlikely to represent  the "average" unburned hydrocarbons while
operating on JP-5 but its use does  form  a basis for comparison.

        The horsepower figure used in the  input data and in the brake  specific
emissions output is actually the test point fuel flow for convenience and has no other
significance.
                                      162

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DATA POINT
                           FNGINE SPEED
                                            0.0
FUFL % CARBON
T HYDROGEN
LHV 3TU/LB
HUMITITY GN/LB
                   85.09999
                   14.90000
                    18400.0
                    75.0000
                          CO 2 %
                          HORSEPOWER
                          HA LB/SEC
                          HF LR/HR
  6.80000
109.00300
  0.941 10
109.00000
-*-*_ «-*-*
WATER VAPOUR
EQUIVALENCE
HJMIDITY FED
HUMIDITY CAL
               CORRECTION
                  0.93305
                  2.08027
                  I.00000
                  0.99981

                         NO
                       FACTORS
PPM 3RSERVED
PPM 30RR  INT
PPM <
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