EPA-650/2-74-051


July  1974
                          Environmental Protection Technology Series



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                                    EPA-650/2-74-051
   ASSESSMENT OF THE APPLICABILITY
  OF AUTOMOTIVE EMISSION CONTROL
TECHNOLOGY TO  STATIONARY ENGINES
                        by

         W. U. Roessler, A. Muraszew, and R. D. Kopa

                 Urban Programs Division
                The Aerospace Corporation
               El Segundo, California 90245
                  Grant No. R-802270
                  ROAP No. 21ADG-84
                Program Element No. 1AB014
             EPA Project Officer: John H. Wasser

                Control Systems Laboratory
            National Environmental Research Center
          Research Triangle Park, North Carolina 27711
                     Prepared for

           OFFICE OF RESEARCH AND DEVELOPMENT
          U.S. ENVIRONMENTAL PROTECTION AGENCY
                WASHINGTON, D.C.  20460

                      July 1974

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This report has been reviewed by the Environmental Protection Agency
and approved for publication.  Approval does not signify that the
contents necessarily reflect the views and policies of the Agency,
nor does mention of trade names or commercial products constitute
endorsement or recommendation for use.
                                  11

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                      ACKNOWLEDGMENTS
              Appreciation is acknowledged for the guidance and con-
tinued assistance provided by Mr. John H. Wasser of the Environmental
Protection Agency, Control Systems Laboratory, who  served as EPA

Project Office.
              The following technical personnel of The Aerospace
Corporation made valuable contributions to the study performed under
this grant.
                           A. Muraszew
                           R. D. Kopa
                                            U. ^bessler, Manager
                                   Stationary Engine Assessment Study
                                   The Aerospace Corporation
Approved by:
Merrill G. Hinton,  Director
Office of Mobile Source Pollution
The Aerospace Corporation
Toru lura, Associate Group
  Director
Environmental Programs Group
  Directorate
The Aerospace Corporation
Joseph ntireltzer, Group Elector
^Environmental Programs Group
  Directorate
The Aerospace Corporation
                                111

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                            CONTENTS


ABSTRACT	   xxi
1.     HIGHLIGHTS	   1-1
2.     INTRODUCTION	   2-1
3.     STATIONARY ENGINE CHARACTERISTICS	   3-1
      3. 1    Diesel Engines  	   3-2
             3.1.1   Engine  Description	   3-2
             3. 1. 2   Design  Considerations  	   3-7
             3. 1. 3   Applications	   3-8
             3.1.4   Emissions  	   3-11
             3. 1. 5   Fuel  Consumption	   3-41
      3.2    Spark-Ignition Engine Characteristics	   3-43
             3.2.1   Engine  Description	   3-43
             3.2. 2   Applications	   3-45
             3. 2. 3   Emissions  	   3-48
      3. 3    Gas Turbines  	   3-61
             3. 3. 1   Engine  Description	   3-61
             3. 3.2   Applications	   3-68
             3. 3. 3   Emissions  	   3-77
      References	   3-111
4.     AUTOMOTIVE EMISSION CONTROL
      TECHNOLOGY	   4-1
      4. 1    Diesel Engines  	   4-2
             4. 1. 1   Variations in Operating Conditions  ....   4-3
             4. 1.2   Component Modifications	   4-26
             4. 1. 3   Emission Control Devices	   4-36
             4. 1.4   Emission Control Systems	   4-59

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                      CONTENTS (Continued)
      4. 2    Spark Ignition Engines  	   4-70
             4. 2. 1   Modification of Engine Operating
                     Conditions  	   4-70
             4.2.2   Preventive Emission Control by
                     Engine Modification	   4-101
             4. 2. 3   Fuel Modification	   4-113
             4.2.4   Corrective Emission Control	   4-115
             4. 2. 5   Control of Emission from Blowby,
                     Carburetor, and Fuel Tank	   4-121
             4.2.6   Combined Emission Control Techniques.   4-123
      4.3    Gas Turbines  	   4-124
             4. 3. 1   Automotive Engine Emission
                     Control	   4-124
             4. 3. 2   Stationary Sources Emission
                     Control	   4-126
             4. 3. 3   Low Emission Combustors for
                     Stationary Gas Turbines	   4-128
             4.3.4   Water/Steam Injection	   4-143
             4. 3. 5   SOV Emission Control	   4-151
                        X
             4.3.6   Smoke, Particulates,  and Odor
                     Control	   4-153
      References	   4-155
5.     EMISSION CONTROL SYSTEM ASSESSMENT	   5-1
      5. 1    Diesel Engines  	   5-1
             5. 1. 1   Emission Control Techniques/
                     Devices	   5-2
             5. 1. 2   Emission Control Systems	   5-9
             5. 1.3   Economic Considerations	   5-9
             5. 1.4   Promising Emission Control
                     Systems	   5-13
                                 VI

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                      CONTENTS (Continued)
      5.2    Spark Ignition Engines  	   5-15
             5.2. 1   Emission Control Techniques/
                     Devices	   5-15
             5.2.2   Economic Considerations  	   5-22
      5. 3    Gas Turbines  	   5-25
             5. 3. 1   Economics of Emission Control	   5 -25
      References	   5-34
6.    RECOMMENDED PROGRAMS  	   6-1
      6. 1    Diesel Engines  	   6-1
      6.2    Spark Ignition Engines  	   6-2
      6. 3    Gas Turbines  	   6-3
      6.4    Emission Inventory  Data	   6-4
APPENDIX A 	  A-l
APPENDIX B.  VISITS AND CONTACTS    	  B-l
APPENDIX C.  UNITS OF MEASURE - CONVERSIONS	  C-l
GLOSSARY	  G-l
                                Vll

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                              FIGURES
3-1.   HC and CO emissions vs air-fuel ratio - four-
       stroke, naturally  aspirated, open-chamber
       diesels	  3-23

3-2.   NOX emissions and brake mean effective
       pressure vs air fuel ratio - four-stroke,
       naturally  aspirated, open-chamber diesel
       engines  	  3-23

3-3.   Specific mass emissions - four-stroke, naturally-
       aspirated,  open-chamber diesel (Engine No. 3)	  3-24

3-4.   Specific mass emissions - four-stroke turbocharged,
       open-chamber diesel (Engine No.  13)  	  3-24

3-5.   Specific mass emissions - four-stroke turbocharged,
       open-chamber large diesel (Engine No. 18)	  3-25

3-6.   Specific mass emissions - four-stroke, naturally
       aspirated divided-chamber  diesel (Engine No. 20) ....  3-25

3-7.   Specific mass emissions - four-stroke turbo-
       charged, divided chamber diesel (Engine  No. 23)  ....  3-26

3-8.   Specific mass emissions - two-stroke diesel
       (Engine No. 26)	  3-26

3-9.   Specific mass emissions - large two-stroke,
       turbocharged diesel (Engine No. 30)	  3-27

3-10.  Diesel engine part-load hydrocarbon  emissions -
       rated speed	  3-31

3-11.  Diesel engine part-load carbon monoxide
       emissions - rated speed	  3-32

3-12.  Diesel engine part-load oxides  of nitrogen
       emissions - rated speed	  3-32

3-13.  Exhaust smoke vs output power  for four open-
       chamber,  naturally aspirated diesel engines	  3-35
                                vui

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                       FIGURES (Continued)
3-14.  Performance of large naturally aspirated and
       turbocharged diesel engines	   3-42

3-15.  Emission map - automotive spark-ignition
       engine	   3-57

3-16.  Part load emissions of a heavy-duty spark-
       ignition engine	   3-59

3-17.  Effect of torque on engine performance - large
       two-stroke spark-gas engine (300 rpm)	   3-59

3-18.  Simple cycle gas  turbine	   3-62

3-19.  Regenerative cycle gas turbine	   3-63

3-20.  Combined cycle gas turbine and stream generator
       (STAG) system	   3-64

3-21.  Simple-cycle gas turbine performance	   3-66

3-22.  Regenerative-cycle gas turbine performance	   3-67

3-23.  Electric  Power Generation Schedule	   3-68

3-24.  Projected gas turbine generating capacity,
       megawatts,  in United  States 1966-1980	   3-72

3-25.  Trends in size of turbines sold for gas com-
       pression service	   3-74

3-26.  Theoretical effect of primary zone temperatures on
       NCL. emissions  	   3-83
          J\.

3-27.  Combined effect - residence time and  primary
       zone temperature	   3-83

3-28.  Operating characteristics at part and full-load
       of a 22 MW gas turbine	   3-86

3-29.  Nitric oxide emission ratio  and fuel-air
       ratio  versus load	   3-86
                                 IX

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                        FIGURES (Continued)



3-30.  NOX emissions — simple cycle oil-fired	   3-87

3-31.  Typical gas turbine NOX emission at base load	   3-87

3-32.  Emission data of pipeline gas turbines  —
       natural gas	   3-88

3-33.  NOX emissions , natural gas-fired, iso-conditions  ....   3-89

3-34.  CO versus load,  W-251 engine	   3-91

3-35.  Emissions for various gas turbine powerplants	   3-91

3-36.  CO versus NOX emission performance  of con-
       ventional gas  turbine engine  combustors  	   3-93

3-37.  GMA 100 gas  generator emissions  	   3-93

3-38.  HC emissions for various gas turbine
       powerplants  	   3-94

3-39.  SO2 versus load, engine W-251  	   3-95

3-40.  Smoke versus load,  engine W-251	   3-98

3-41.  Calculated particulate matter emission rate result-
       ing from black smoke particles versus  von Brand
       (reflectance) smoke  number	   3-98

3-42.  W-251 engine combustion contaminants (dry
       filter method) versus load	   3-100

3-43.  Particulate matter emission when burning crude
       and distillate oil  fuel	   3-100

3-44.  NEMA noise standards for industrial and
       residential centers	   3-103

3-45.  Sound level performance of heavy-duty  gas
       turbines compared to NEMA sound levels	   3-104

3-46.  Electric utility gas turbine fuel demand	   3-107

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                        FIGURES (Continued)
3-47.  Effect of fuel bound nitrogen on NOX formation
       at base load	   3-109

4-1.   Effect of speed and power output on emissions -
       Caterpillar four-stroke precombustion chamber
       diesel	   4-4

4-2.   Effect of intake-air temperature on NOX emis-
       sion of a turbocharged, open-chamber diesel
       engine at rated, speed	   4-5

4-3.   Effect of fuel injection on performance and
       emissions of a single  cylinder research engine	   4-10

4-4.   Effect of injection rate and timing on gaseous mass
       emissions, smoke, and performance of a naturally
       aspirated diesel engine	   4-12

4-5.   Effect of injection timing on  specific  fuel con-
       sumption peak opacity, and maximum power
       output of a naturally aspirated, open-chamber
       diesel engine — Engine No. 7	   4-14

4-6.   Effect of injection timing on  the emissions of a
       naturally aspirated,  open-chamber diesel
       engine - Engine No.  7	   4-15

4-7.   Effect of injection timing retard on the NOX
       emissions and specific fuel consumption of a
       large diesel engine	   4-19

4-8.   Effect of injection timing on  NOX emissions
       and specific fuel consumption of a two-stroke
       diesel engine	   4-19

4-9.   NOX reduction in diesel engines vs specific
       fuel consumption and timing  retard	   4-20

4-10.  HC,  CO and smoke variations  vs timing retard	   4-21

4-11.  Cetane number effects on the emissions of
       naturally aspirated and turbocharged diesels 	   4-23

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                       FIGURES (Continued)
4-12.  Measured effect of fuel cetane number on NCs
       vs BSFC tradeoff	  4-24

4-13.  Effect of bowl diameter and piston head clear-
       ance on diesel engine emissions and specific
       fuel consumption	  4-28

4-14.  Effect of compression ratio on NOX and HC
       emissions	 . .  .  4-30

4-15.  Effect of air swirl and turbocharging on smoke
       and NOX emissions of open-chamber diesel
       engines	  4-31

4-16.  Effect of injection rate on NOX emissions of a
       two-stroke diesel engine	  4-33

4-17.  Effect of EGR on NOX and smoke emissions of
       a naturally aspirated open-chamber diesel
       engine	  4-37

4-18.  Effect of EGR on naturally aspirated open-
       chamber diesel engine mass emission and
       performance (Engine No.  7)	  4-39

4-19.  Effect of EGR on NOX emissions of a turbo-
       charged, open-chamber diesel engine at
       rated speed	  4-40

4-20.  Effect of exhaust gas recirculation on NOX
       emission of an air-scavenged,  two-stroke
       diesel,  at rated  speed	  4-41

4-21.  Effect of EGR on performance and emissions
       of a single-cylinder naturally aspirated engine
       (80 psi BMEP, cooled EGR at 80°F)	  4-41

4-22.  Effect of EGR on diesel engine emissions and
       specific fuel consumption	  4-44

4-23.  Effect of inducted and emulsified water on the
       HC,  smoke, and NOX emissions of an open-
       chamber diesel engine	  4-46
                                XI1

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                       FIGURES (Continued)
4-24.  Reduction of NOX emission as a function of water
       induction — turbocharged, open-chamber diesel
       engine	  4-48

4-25.  Water induction versus NO emission — pre-
       chamber,  turbocharged diesel,  2200 rpm	  4-49

4-26.  NOX reduction versus water to fuel mass  ratio	  4-53

4-27.  Effect of catalysts on diesel engine emissions	  4-55

4-28.  Effect of turbocharging and intercooling on the
       emissions and specific fuel consumption of a
       family of open-chamber diesel engines 	  4-58

4-29.  Effect of turbocharging on specific fuel con-
       sumption and mass emissions (Engine No. 8)	  4-60

4-30.  Projected effect of emission control systems
       on emissions and specific fuel consumption	  4-67

4-31.  Effect of 10 percent EGR and  5° injection timing
       retard on specific fuel consumption and emis-
       sions (Engine No. 24)	  4-69

4-32.  Effect of air-fuel ratio on emission levels, gaso-
       line spark-ignition engine	  4-72

4-33.  Effect of air-fuel ratio on reactivity index and
       concentration measured by infrared analyzer	  4-73

4-34.  Effect of air-fuel ratio on exhaust hydrocarbon
       emission and fuel economy in car at 30 mph
       roadload	  4-74

4-35.  Effect of air flow rate on engine emissions and
       performance -  Cooper Bessemer GMVA-8
       2-stroke atmospheric spark-gas engine; 1080 bhp
       at 300 rpm, 82. 5 bmep,  base conditions  	  4-75

4-36.  Correlation between peak cycle temperature and
       NO concentration	  4-76
                                xm

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                        FIGURES (Continued)
4-37.  Effect of air-fuel ratio and of spark timing on
       oxides of nitrogen	   4-77

4-38.  Effect of spark timing on hydrocarbon compo-
       sition by class at rich air-fuel ratio	   4-78

4-39.  Effect of spark-timing on hydrocarbon compo-
       sition by class at lean air-fuel ratio	   4-78

4-40.  Effect of ignition timing on engine emissions
       and performance - Cooper Bessemer GMVA-8
       2-stroke atmospheric spark-gas engine  1080  bhp
       at 300 rpm, 82. 5 bmep, base conditions  	   4-80

4-41.  Effect of heating of inlet manifold on exhaust
       emissions  and mixture temperature	   4-81

4-42.  Effect of air manifold temperature on emissions
       and performance - Cooper Bessemer GMVA-8
       two-stroke atmospheric spark-gas engine 1080  bhp
       at 330 rpm, 82. 5 bmep, base conditions  	   4-83

4-43.  Effect of average combustion chamber surface
       temperature on  hydrocarbon emission	   4-84

4-44.  Effect of coolant temperature and spark advance
       upon  indicated specific NO, HC,  and CO
       emissions	   4-86

4-45.  Effect of engine speed on NOX concentration	   4-87

4-46.  Effect of engine speed on HC concentration	   4-88

4-47;  Effect of speed on emissions and performance -
       Cooper Bessemer GMVA-8 two-stroke atmospheric
       spark-gas engine, power output  1080 bhp,  base
       conditions	   4-90

4-48.  Effect of intake valve opening timing on HC and
       NO emissions	   4-92

4-49.  Effect of exhaust valve closing on HC and NO
       emissions	   4-92
                                xiv

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                       FIGURES (Continued)
4-50.  Nitric oxide vs charge dilution relationship for
       valve overlap,  recirculation and compression
       ratio tests 	  4-94

4-51.  Effect of cam advance and retard on hot cycle
       vehicle  emissions with the 1970 Federal test
       procedure	  4-94

4-52.  Effect of manifold air pressure on oxides of
       nitrogen	  4-96

4-53.  Effect of load at constant speed on emissions and
       performance -  large  two-stroke atmospheric
       spark-gas engine base conditions, speed
       300 rpm	  4-97

4-54.  Exhaust emissions, exhaust temperature,  and
       fuel-air ratio as functions of exhaust backpres-
       sure for three  absolute inlet manifold pressures,
       2000 rpm	  4-99

4-55.  Effect of deposit buildup on exhaust NO and HC
       concentrations	  4-100

4-56.  NO emission per unit output for different com-
       bustion  chamber shapes  and spark plug locations	  4-102

4-57.  Effect of combustion  chamber geometry on the
       surface-to-volume ratio	  4-103

4-58.  Composite values:  California chassis dyna-
       mometer  schedule	  4-103

4-59.  Texaco-Controlled Combustion  System	  4-107

4-60.  Effect of EGR on NO  reduction and specific
       fuel consumption	  4-109

4-61.  Test stand NOX emissions as a  function of A/F
       and recycle  rate  - 50 mph road load,  37-degree
       btc spark timing,  gasoline fuel	  4-111
                                 xv

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                        FIGURES (Continued)
4-62.  Effect of water injection on the emissions and
       specific fuel consumption of a CFR engine,
       5. 5 hp, 1200 rpm, 30° spark advance	   4-112

4-63.  Effect of water injection on NO and HC concen-
       tration, 900  rpm; spark advance 30° btdc	   4-113

4-64.  Effect of water injection on emissions and
       performance - Ingersoll-Rand PKVGR-12,  4-cycle
        naturally aspirated spark-gas engine	   4-114

4-65.  NOX versus SFC increase	   4-118

4-66.  Gas  turbine state-of-the-art emissions  (No.  2
       GT turbine oil:  15 percent 02)	   4-129

4-67.  Flame temperatures and equilibrium NO concen-
       trations as functions of air-fuel ratio for various
       inlet temperatures	   4-130

4-68.  Limits of flammability of a paraffin  hydrocarbon
       (CnH2n.f-2) showing the influence of inlet
       temperature	   4-131

4-69.  NOX variation with temperature rise for produc-
       tion  and modified  combustors	   4-133

4-70.  Effect of primary zone leaning  on NO emission —
       natural gas	   4-134

4-71.  Effect of cooled exhaust gas recirculation on NO
       emission — natural gas	   4-134

4-72.  High recirculation stabilized lean primary  zone
       combustor schematic and NO-, emission	   4-137

4-73.  External recirculation combustor	   4-138

4-74.  Effects of recirculation on emissions 	   4-139

4-75.  Schematic of a low-emission combustor concept —
       Ford externally vaporizing combustor (EVC) 	   4-140
                                 xvi

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                        FIGURES (Continued)


4-76.  Transpiration combustor	   4-142

4-77.  Schematic of water-injection system	   4-146

4-78.  Effect of water injection on NOX emis-
       sions from MS 5001 turbine — liquid fuel	   4-146

4-79.  NOX emissions from modified combustion sys-
       tem in MS 5001 gas turbine — gas  fuel	   4-147

4-80.  Effect of water injection on emissions — 5001K
       gas turbine engine  combustor with fuel oil	   4-148

4-81.  NOX reduction by water-injection, oil-fired
       Model 5000 engine, iso-conditions	   4-148

4-82.  NOX reduction by steam-injection, gas-fired
       MS-5001L engine,  site conditions	   4-149
                                    \
4-83.  Combustion laboratory NOX reduction with
       water injection	   4-150

4-84.  Total HC  versus  load,  W-251 engine  	   4-150

5-1.   Projected NOX abatement cost - turbocharged,
       open-chamber diesels (fuel cost only)	   5-12

5-2.   Projected average  NO  vs SFC  correlations	   5-12
                                xvli

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                             TABLES
3-1.   Estimated Installed Diesel Engine Horsepower
       for 1971	  3-9

3-2.   Emissions  from Four-Stroke, Naturally Aspi-
       rated Open Chamber Diesel Engines - Rated
       Conditions  	  3-17

3-3.   Emissions  from Four-Stroke, Turbocharged,
       Open Chamber Diesel Engines - Rated
       Conditions  	  3-17

3-4.   Emissions  from Four-Stroke Naturally Aspi-
       rated, Divided Chamber Diesel Engines -
       Rated Conditions	  3-18

3-5.   Emissions  from Four-Stroke, Turbocharged,
       Divided Chamber Diesel Engines - Rated
       Conditions  	  3-18

3-6.   Emissions  from Two-Stroke, Open Chamber,
       Blower Scavenged Diesel Engines  - Rated
       Conditions  	  3-19

3-7.   Emissions  from Large Two-Stroke, Open
       Chamber, Turbocharged Diesel Engines -
       Rated Conditions	  3-19

3-8.   Average  Diesel Engine Emissions at Rated
       Conditions  (Uncontrolled Engines)	  3-30

3-9.   Average  Steady-State Smoke Emission from
       Diesel Engines  ....-•-.	  3-36

3-10.  Average  Particulate Emissions from  Diesel
       Engines	  3-36

3-11.  Internal Combustion Engines —Number vs
       End  Use	  3-46

3-12.  Estimated Installed Horsepower of Spark-
       Ignition Gas Engines for 1971	  3-47
                               XVUJ.

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                        TABLES (Continued)
3-13.  Exhaust Gas Analysis Methods  and Instruments	   3-52

3-14.  Emissions from Four-Stroke, Naturally Aspi-
       rated Spark Ignition Heavy Duty Gasoline
       Engines	   3-55

3-15.  Emissions from Four- and Two-Stroke,  Aspi-
       rated and Turbocharged Spark-Ignition Gas
       Engines	   3-56

3-16.  Average  Spark-Ignition Engine  Emissions at
       Rated Conditions	   3-60

3-17.  Power Density Comparison	   3-70

3-18.  Gas  Turbine Power for Pipeline Use, 1958-1970	   3-73

3-19.  Stationary Turbines  -  U.S.  Manufacturers	   3-75

3-20.  Example of Emission Standards (Maximum
       Allowable) 	   3-78

3-21.  Gas  Turbine Emission Units (11,500 Btu/kwh;
       18, 500 Btu/lb fuel)	   3-79

3-22.  Total and Gas Turbine Stationary Power	   3-80

3-23.  Fuel Sulfur Content in Percent	   3-95

3-24.  Breakdown of Particulate Matter	   3-99

4-1.   Effect of Intake Air Cooling on  the Emissions and
       Specific Fuel Consumption of a Cooper Bessemer
       KSV-12 Diesel Engine	   4-8

4-2.   Single Cylinder Diesel Engine Emissions as a
       Function of Injection Timing Retard	   4-11

4-3.   Effect of 4° Injection  Timing Retard on Emissions
       and Fuel Consumption of a Large Stationary
       Diesel	   4-17
                                xix.

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                         TABLES (Continued)
 4-4.    Effect of Water Induction on the Emissions of an
        Open-Chamber Diesel Engine (13-Mode Cycle
        Data)	   4-47

 4-5.    Effect of Water Induction on Cooper Bessemer
        KSV-12 Diesel Engine Emissions and  Fuel
        Consumption	   4-51

 4-6.    Effect of Emission Control Systems on the Emis-
        sions of a Turbocharged, Open-Chamber Diesel
        Engine — 13  Mode Cycle	   4-62

 4-7.    Summary of Emissions and Fuel Consumption for
        Baseline and Combination of Parameters Tests
        for Diesel Engines	   4-63

 4-8.    Effect of Combined Emission Control Techniques
        on Diesel Engine Emission and Specific Fuel
        Consumption	   4-71

 4-9.    Thermal Reactor Summary	   4-117

 4-10.   Examples of Best Low Mileage Emission Mea-
        surements with Dual-Catalyst Systems on
        Experimental  1976 Vehicles	   4-121

 4-11.   Emissions as  Function of Mileage for Durability
        Tests on Dual-Catalyst Systems	   4-122

 4-12.   Effectiveness of Various Gas  Turbine Emission
        Controls	   4-151

 5-1.    Estimates of Sticker Prices for Emissions Hard-
        ware from 1966 Uncontrolled  Vehicles to 1976
        Dual-Catalyst Systems	   5-23

 5-2.    Water Injection Investment Cost (San Diego Gas
        and Electric)	   5-29

 5-3.    Water/Steam Injection Cost as  a Function of
        Power Plant Size	   5-29

A-l.    Estimated Installed Stationary Engine Horse-
        power - 1971	  A-2

                                 xx

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                            ABSTRACT
              This  report gives a review of the emission
characteristics  of uncontrolled stationary dies el,  spark ignition,  and
gas turbine engines,  including an analysis and evaluation of the appli-
cability of automotive emission control technology to stationary
engines.  Nitrogen oxides have been identified to be the principal
pollutant species emitted from these engines. In principle, the
emission  control techniques developed or evaluated for spark ignition,
diesel,  and gas  turbine engines are applicable to stationary engines.
However, in most cases, the emission reductions achieved with these
techniques are accompanied by sizeable losses in specific fuel con-
sumption  and uncertainties  relative to the effect of these techniques
on engine life and control system durability.

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                             SECTION 1

                            HIGHLIGHTS
              An examination and summarization was made of
available information pertaining to  (1) the exhaust emissions from
uncontrolled stationary diesel,  spark ignition,  and gas turbine engines
and (2) the automotive emission control techniques that  have been
developed to date and are potentially applicable to stationary engines.
These techniques include engine derating; fuel  injection and ignition
timing retard; exhaust gas recirculation; catalytic converters; water
injection; and engine component and operating condition modifications.
A considerable amount of technical  data relating to exhaust emissions,
fuel consumption characteristics, and cost and effectiveness  of various
control techniques was  obtained in the data acquisition process and is
presented in the main body of the report.  An analysis and evaluation
of these data resulted in the  following findings:
         1.   The oxides of nitrogen (NOX) are  the principal pollutant
             species emitted from  stationary diesel, spark ignition,
             and gas turbine engines.  Other pollutants emitted from
             these engines in various quantities include  hydrocarbons
             (HC), carbon monoxide (CO), smoke, particulates, oxides
             of sulfur (SOX), aldehydes, odor,  and noise.
                                  U-l

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2.  The HC and CO emissions in well-maintained diesel, gas
    turbine,  and gas-fueled spark ignition engines are much
    lower than those from automotive spark-ignition engines.
    However, stationary gasoline engines operating at fuel-
    rich mixtures have HC and CO emissions that are
    comparable to automotive engines.

3.  Smoke, particulate,  odor, and SOX emissions, while of
    some concern in diesels and  gas turbines, are generally
    very low in spark ignition engines.  SO  is directly
    related to the sulfur  content in the fuel and can be con-
    trolled by removing the sulfur during the refinery process.

4.  In principle, automotive emission control techniques
    developed or evaluated for spark ignition, diesel, and
    gas turbine engines are applicable to  their stationary
    counterparts.  However, the degree of emission control
    realized and the resultant effect on fuel consumption,
    engine life, and operating parameters in stationary appli-
    cations cannot be directly inferred from similar effects
    in automotive installations.   Each such control technique
    must be evaluated with regard to the specific  stationary
    engine type,  design,  and operating conditions.

5.  Derating of stationary engines as a  means of NOX control,
    although effective in  some cases, is not  considered to be
    economically attractive because of the attendant increase
    in the cost of the engine per unit horsepower output and
    the potential degradation in specific fuel consumption.
    For example, derating a turbocharged divided chamber
    diesel engine by 30 percent to obtain a 40 percent reduc-
    tion in NOX would increase the investment cost per unit
    horsepower of the engine by 30 percent.   This method
    may also result in increases  in HC  and CO emissions.
6.  Fuel injection timing or ignition timing retard represent
    the only potential NOX abatement techniques for diesel
    and spark ignition engines that require no hardware
    changes or additions. For this reason, these techniques
    are being considered by many manufacturers for applica-
    tion in these engines  until better methods can be devel-
    oped.  From a cost-effectiveness point of view, injection/
    ignition retard is the least desirable approach, resulting
    in substantially higher specific fuel consumption and
    exhaust gas temperatures,  (e.g. , up  to 12 percent
    increase in fuel consumption at 12 degrees injection
    retard for 50 percent reduction in NOX).   A limited
    amount of injection retard, combined with other tech-
    niques such as intake air cooling, might be feasible for
                        1-2

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     diesel engines to achieve a moderate NOX reduction
     (approximately 25 percent) with essentially no loss in
     fuel economy.
 7.   Both exhaust gas recirculation (EGR) and water injection
     into the engine intake are very effective NOX-abatement
     techniques for diesel engines,  permitting NOX reductions
     of up  to about 60 percent.  The small increase  in specific
     fuel consumption associated with EGR can be compen-
     sated for,  at least to some degree,  by incorporating
     intake air cooling.  In the case of water injection, the
     fuel consumption losses are generally negligible.

 8.   Water injection into the diesel engine cylinder, in the
     form  of fuel/water  emulsions, appears to be an attractive
     approach that merits further consideration primarily
     because of the greater  simplicity of this  technique rela-
     tive to water induction.  However, a number of potential
     problem areas related  to corrosion and wear of important
     engine components  would have to be resolved before incor-
     poration of these techniques into stationary diesel engines.
 9.   Turbocharging of naturally aspirated diesel engines,  com-
     bined with  retarded injection timing and intercooling,
     results in NOX reductions of up to 35 percent without any
     loss in specific  fuel economy relative to  the baseline con-
     figuration.  The  NOX reduction actually achieved in
     various engines is largely dependent upon the degree of
     intercooling employed.

10.   Variations in the fuel properties have very little effect on
     the NOX emissions  from diesel engines.  In general,  NOX
     and HC decrease slightly with  increasing cetane number
     while CO remains essentially constant.   However,  the
     fuel bound  nitrogen contained in the heavier fuels might
     seriously limit the  effectiveness of many potential NOX
     emission control techniques.
11.   Catalytic converters and thermal reactors are  considered
     to have very limited applicability as  HC and CO abatement
     devices in  diesel engines because of the already low HC
     and CO concentrations  and the contaminants contained in
     the diesel exhaust which might adversely affect catalyst
     durability.  Thus, the use of oxidation catalysts would  not
     be anticipated unless extremely low levels of HC and CO
     emissions  were required.  More significantly,  neither of
     these devices is  effective in reducing NOX because of the
     excess air contained in diesel  exhaust.
                         1-3

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12.   Achievement of diesel engine NOX emissions substantially
     lower than those obtainable through incorporation of the
     emission control techniques discussed above requires the
     development of new engines.  These would incorporate
     certain modifications in the combustion chamber geometry,
     injectors,  timing schedules,  turbocharger,  and inter-
     cooler.  The development of such an engine is a difficult
     task that would require several years.

13.   Although no Federal emission standards are currently in
     existence for  stationary gas turbines,  several states are
     enforcing the  Federal standards that have been promul-
     gated for  steam powerplants, and a number of counties
     and cities have formulated even more stringent regulations.
     With few exceptions, the current gas turbines are capable
     of meeting these limits by incorporation of modest  com-
     bustor modifications or  addition of water or steam  injec-
     tion.  However, these techniques are limited in their
     effectiveness  and are insufficient to achieve substantially
     lower emission standards.

14.   The modest combustor modifications required to meet
     current load regulations have a negligible effect on the
     investment and operating cost of gas turbines, particu-
     larly in the case of large units.  On the other hand,  water
     or steam injection can more than double the investment
     cost of small  units  (below  1 MW output) and increase the
     cost of large units (30 MW or above) by seven to  ten
     percent.

15.   Future emission control efforts in gas turbines are
     expected to concentrate  on the development  of low-
     emission  "dry" combustors utilizing the advanced  low
     NOX combustion technology which is  currently being
     developed for potential use in automotive gas turbines.
     The ultimate goal of these developments is a premixed,
     prevaporized,  well-stirred combustor.  It is estimated
     that the development time of such combustors would be
     between five and ten years.

16.   Externally mounted  gas  turbine combustors may  offer
     more flexibility and greater potential for emission  con-
     trol than conventional in-line combustors, particularly
     for heavy fuels and residuals.

17.   Catalytic combustors, having a potential of  low emissions,
     are of interest to automotive gas turbines but are not con-
     sidered applicable to most stationary gas turbines at this
     time because  of many unresolved potential problem areas
                          1.4

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     such as:  limited catalyst durability, uncertain specific
     heat release rate, ignition characteristics,  and high
     pressure drop.

18.  In principle, the emission control techniques developed to
     date for automotive  spark-ignition engines are applicable
     to stationary gasoline and gas-fueled spark ignition
     engines.  However,  in most cases the emission reductions
     achieved with these  techniques are accompanied by sizable
     losses in specific fuel consumption and  substantially
     increased operating cost.  For example, in one case utili-
     zation of water injection to reduce NOX  60 percent resulted
     in a six-percent increase in specific fuel consumption.
19-  On the basis of current state-of-the-art technology, the
     most cost effective NOX emission control technique pro-
     jected for use in stationary gasoline and gas  spark ignition
     engines appears to be a combined system consisting of
     optimum valve and port timing, intake charge cooling,
     intake air humidification, and operating  speed changes.
     However, a slight increase in the emission of hydrocarbon
     might occur in this case which then might further require
     the use of a thermal reactor or catalytic converter.

20.  It is difficult to evaluate the relative need for implement-
     ing any of the various control techniques shown to be
     applicable to stationary engines.  This difficulty arises
     principally because  there is no accurate inventory of
     emissions for the various engine classes and no perspec-
     tive as  to how these emissions relate  to emissions  from
     other sources and air quality.  In addition, no goals have
     been set for stationary engine emission reduction (aside
     from local regulations for gas turbines) and the future
     trends of utilization of the various stationary engine types
     is uncertain.
                          1-5

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                             SECTION 2

                           INTRODUCTION
              This report presents a compilation,  evaluation, and
assessment of available information pertaining to the applicability of
automotive emission-control technology to stationary diesel,  spark
ignition, and gas turbine engines, both in retrofitting current engines
and application to new engines.
              To fulfill the objectives of this  study, the work effort
was divided into two basic phases:  the first phase was concerned with
the compilation and review of applicable information acquired from:
(1)  an open literature survey and (2)  visits to engine manufacturers
and users for technical discussions of relevant engine and emission
control system performance characteristics and economic factors.  In
the second phase of the study,  a summarization and evaluation of all
data acquired in the first phase was made.
              The results of this study are presented in the following
order and .context:  Section 3 is concerned with stationary engine
characteristics and addresses the various engine design approaches,
applications, emissions, and specific fuel consumption.   Diesel  engines
                                 2-1

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are discussed in Subsection 3.1, spark ignition engines in
Subsection 3.2,  and gas turbines in Subsection 3.3.
              Section 4 discusses the effectiveness of all known
emission-control techniques/devices considered by the automotive
industry in their efforts to meet current and future emission control
standards.  Whenever possible,  data from stationary engines are
included and compared with corresponding automotive engine data,
both in terms of emission reduction and fuel consumption effects.  Sub-
section 4.1 is devoted to diesel engines, Subsection 4.2 to spark ignition
engines, and Subsection 4.3 to gas turbines.
              Section 5 presents an evaluation of the emission control
approaches  identified in Section 4 with respect to performance and
economics.
              Section 6 identifies  those areas where  further research
and development (R&D) efforts are needed to bridge existing data gaps
and provide the  technical  information required for a more compre-
hensive assessment of the cost effectiveness of various emission con-
trol approaches.
              Appendix A presents a  summary table  listing the
installed horsepower of all stationary diesel, spark ignition and gas
turbine  engines.  Appendix B lists those organizations which contributed
to this study,  either directly by providing useful engine test data, or
indirectly through general discussions of engine performance and
emission characteristics. Appendix  C presents metric system conver-
sion factors.
                                  2-2

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                             SECTION 3

             STATIONARY ENGINE CHARACTERISTICS
              In this section of the report,  brief discussions are
presented relative to the design and application of stationary diesel,
spark ignition, and gas turbine engines.  The major portion of the
section is concerned with the emissions aspects of these three engine
types,  both  from a rated-load and part-load point  of view. Although
the oxides of nitrogen have  been identified to be the principal pollutant
species from these engines, other emissions including hydrocarbons,
carbon monoxide, oxides of sulfur,  aldehydes,  smoke, particulates,
odor, and noise,  were given limited consideration.
              Since there is  a lack of emission data from stationary
engines, particularly with respect to large size/low speed designs,
applicable emission data from heavy-duty automotive engines are
included in the emission characterization of several of the engine cate-
gories considered in this section.  Because  of many similarities in the
design and operating parameters of heavy-duty automotive and large
stationary engines,  this approach appears to be justified.
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3.1           DIESEL ENGINES
3.1.1         Engine Description
              The diesel or compression ignition engine is a
reciprocating engine in which air is compressed in the cylinder and
the fuel is then injected into the hot air toward the end of the compres-
sion stroke.  Numerous 4-  and 2-stroke engine configurations have
been designed around the basic thermodynamic diesel cycle, and a
number of these are  currently being marketed by many manufacturers
throughout the world. The  various  designs can be grouped into two
categories:  open-chamber  (or direct-injection engines) and divided-
chamber (or indirect-injection engines).  Each of these categories can
then be subdivided into a number of engine classes, depending upon the
type of combustion chamber and  air induction system utilized.
              Many diesel  engines  are fitted with turbochargers to
increase their power output and to  improve the  specific fuel consump-
tion.  Frequently,  an aftercooler is added between the compressor
exhaust and the inlet manifold to reduce the air charge temperature,
thus raising the peak power output  of the engine.
              Because of their superior fuel consumption character-
istics, direct-injection engines are favored by many manufacturers for
use in stationary and heavy-duty truck applications. Conversely, the
divided-chamber  configurations are preferred by the makers of light-
duty automotive diesels and by some manufacturers of stationary and
heavy-duty truck engines.
              Briefly, in open-chamber diesel engines,  fuel is
injected several degrees before the piston reaches top dead center.  To
enhance fuel-air mixing and improve combustion,  a certain degree of
air swirl  is generated in the combustion chamber by means of specially
designed intake manifolds,  valves, and pistons.  Depending upon the
magnitude of the swirl in the combustion chamber,  the direct-injection
engines can be subdivided into  the following three classes:  (1) quiescent
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or low-swirl "Mexican hat" chambers, favored by many domestic
manufacturers,  (2) medium-swirl, deep-bowl chambers, used pri-
marily in Europe, and (3) high-swirl, spherical "M" combustion
chambers, used in special applications.
              The divided-chamber diesel engine employs two  com-
bustion chambers,  consisting of a main chamber and an antechamber
connected to the main  chamber through a communicating flow passage.
This engine category can be divided into three  classes:  (1) swirl
chamber or turbulence chamber--a configuration  favored by a number
of European and Japanese light-duty and heavy-duty diesel engine
manufacturers,  (2) precombustion chamber or prechamber, employed
in a number of European and domestic light-duty and heavy-duty diesel
engines,  and (3) air cell and energy-cell  combustion chamber.
              The important design features of these six engine
classes are briefly described in the following sections.
3.1.1.1       Open-Chamber, Low-Swirl Diesel
              The low-swirl, open-chamber diesel engine utilizes a
shallow dish (or a "Mexican hat") combustion chamber.  In the  absence
of an induced air swirl, mixing of the fuel and  the air is accomplished
by means of a sophisticated fuel injection system.  One or more injec-
tion nozzles are utilized, each having several fuel injection orifices
(Refs.  3-1 through 3-4).
              This engine class  is favored for  large low-speed appli-
cations (300-1200 rpm),  but is also used  in some  medium-size,
medium-speed (1200-2500  rpm) designs.   Normally the engine is
operated with  considerable excess  air; and a high combustion effi-
ciency is generally achieved even with low-grade  fuels.
              Because of the low-air swirl in  the chamber and the
lean air-fuel mixture operation,  the heat losses to the cylinder wall
and head are minimized, resulting in low specific fuel consumption,
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desirable starting characteristics,  and moderate cooling system
requirements.  Other features of this particular engine class include
long exhaust valve life (low exhaust gas temperatures) and residual
fuel capability.
3.1.1.2       Open-Chamber, Medium-Swirl Diesels
              Open-chamber, medium-swirl configurations are utilized
primarily in engines in the medium-power regime (less  than 100 hp per
cylinder) operating at speeds up  to about 3000 rpm.  In these engines,
the amount of fuel injected during each power stroke is smaller than in
the case of the previously discussed low-swirl engines.  Therefore,
fuel-air  mixing by means  of multiple fuel sprays is no longer possible,
and incorporation of some air swirl  is required to complete the mixing
process  within the short time span available.  The air swirl is gener-
ated by one  of several approaches,  including the use of directed intake
ports or masked  intake valves (Ref.  3-5).  Typically,  air-swirl ratios
(defined  as the rotational speed of the air swirl in the chamber just
prior to  ignition, divided by the  rotational  speed of the engine) between
3 and 6 are  employed in this engine class (Ref. 3-1).
              In general, medium-swirl,  open-chamber diesels
exhibit good fuel  economy combined  with acceptable  starting charac-
teristics and low heat losses to the  chamber wall.
3.1.1.3       Open Chamber, High  Swirl Diesels
              The high-swirl diesel, or "M" engine, utilizes air-
swirl ratios of the order  of 12 generated by means of a corkscrew-
type intake port and the squish action of the special piston (Ref. 3-6).
In general, this  engine utilizes a single injector nozzle which supplies
a coarsely atomized  spray into the spherical combustion chamber.
Most of the fuel  is carried to the wall by centrifugal forces where it is
deposited in the  form of a thin film.  The remainder of the fuel (approx-
imately five percent), which has been finely atomized during the injection
                                 3-4

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process, is then ignited.  As combustion proceeds, the flame front
progresses to the wall region of the chamber and ignites the vaporized
fraction of the fuel  initially deposited on the wall in liquid form.  The
burning rate of this fuel is controlled by the fuel vaporization process
and, as a result, relatively low maximum combustion pressures and
pressure rise rates are obtained.  Even at engine  speeds above
3000 rpm,  the pressure rise rates in this design are  comparable to
medium-swirl diesels.
              The  principal advantage of this engine  is its multifuel
capability.  Since combustion is vaporization-controlled, almost any
type of fuel can  be  burned, and the engine runs very quietly even on
gasoline.   The major drawback of the  engine is its poor  cold-start
capability which is  related to the high heat transfer into  the chamber
wall during compression.
3.1.1.4       Divided-Chamber,  Swirl-Chamber Diesel
              The  swirl chambers employed in these engines are
spherical or semi-spherical in shape and have a capacity of the order
of 50 percent of the total clearance volume (Ref. 3-7).  During the
compression stroke,  air is forced from the  main chamber into the
swirl chamber through the narrow passage connecting the two chambers
and a high  degree of air swirl  is then generated by the inrushing air.
The fuel is injected into the swirl  chamber and is deposited on the
chamber walls by the centrifugal forces set  up by the  air swirl.  Some
of the larger fuel droplets follow the air swirl along the  wall, while
the smaller ones start to  evaporate  immediately.  Ignition occurs in
the vicinity of the chamber throat and diffusion-flame-type combustion
spreads from there throughout the swirl chamber.  As the pressure in
the swirl chamber  rises,  the hot combustion gases are discharged into
the main chamber where sufficient excess air is available to complete
the combustion process.  Upon leaving the swirl chamber, a strong
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secondary air swirl is generated which enhances flow mixing in the
main chamber and assures high combustion efficiency.
              Relative to open-chamber  configurations,  swirl-chamber
engines have a number of inherent advantages.  These include higher
speed capability, lower fuel  sensitivity, lower maximum pressures,
lower pressure  rise rates, lower temperatures and lower exhaust
emissions.
              The principal disadvantage of the swirl-chamber diesel
is its higher fuel consumption, which is primarily due to the higher
surface-to-volume ratio of the combustion chamber and the  high swirl
velocities.  Furthermore, during a cold start, a sizable  fraction of
the total compression heat is transferred to the swirl-chamber wall  by
the high air  swirl,  and starting aids such as glow plugs are required
to minimize  engine cranking time.
3.1.1.5      Divided-Chamber,  Prechamber-Diesel
              Although there are significant design differences, the
operating characteristics of  the prechamber diesel engine are, in many
respects,  similar to the previously discussed swirl chamber.
              The prechamber is generally pear-shaped, rather than
spherical, and has a capacity of only 20 to 30 percent of the total clear-
ance volume. The passage connecting the prechamber and the  main
chamber is more restricted  than in the case of the swirl  chamber.
This increases air turbulence in the main chamber and tends to confine
the initial  shock occurring during combustion to the prechamber
(Ref. 3-8).
              In general terms, the advantages and disadvantages of
prechamber diesel engines relative to open-chamber designs follow
those of the swirl chamber.  The  specific fuel consumption of the pre-
chamber engine  is expected to be  slightly lower than for the swirl
chamber, because of the lower heat losses associated with the  lower
turbulence levels and physical  size of the prechamber.
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3.1.1.6       Divided-Chamber - Air-Cell Diesels
              The air-cell diesel engine employs a small antechamber
located in line with the fuel injection nozzle and connected to the main
chamber by a small orifice-type restriction.   Upon ignition in the main
chamber, additional air  is forced into the air  cell.  As the piston
descends, the air trapped in the air cell is gradually readmitted into
the main chamber generating considerable turbulence and improving
the combustion efficiency.  Since the advent of the swirl chamber,
interest in the air cell concept has declined.
              The energy cell or Lanova  cell concept is  a modified
version of the simple  air cell configuration and combines the  features
of the prechamber and the air cell.  The principal feature of the
energy-cell diesel is its smooth operation.  The cold-start charac-
teristics of the engine are favorable and the specific fuel consumption
is acceptable, but higher than that of equivalent open-chamber
configurations.
3.1.2         Design Considerations
              Generally, stationary engines  are designed for minimum
total operating cost (initial  cost, fuel consumption, and maintenance);
long life; good response  and overload capability; and low  noise and
vibration levels.  The useful life of large,  low-speed engines is of the
order of 30 years, with major overhauls every 20,000 hours.  Minor
overhauls (inspection, injector  cleaning, etc.) are scheduled  at
10,000-hour  intervals.
              For medium and small stationary diesels, the  time
between major overhauls has been quoted  by one manufacturer to be
about 5000 to 8000 hours.
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3.1.3         Applications
              Diesel engines are widely used by electric and natural
gas utilities, the petroleum industry, and many operators of small
electric power and pumping stations.
              Many electric utilities employ diesel engines as prime
movers of continuous- and peaking-power generators, standby power
installations and,  more recently, total energy systems.   Many of the
transmission line and process compressors utilized by the petroleum
industry are powered by diesel engines.  Also, diesel engines are
frequently used by the petroleum industry as drives for oil and gas
well drilling and pumping equipment, water pumps,  and electric gen-
erators.  The operators of small electric power  stations and  pumping
installations include municipalities and commercial firms which are
utilizing diesel engines to supply part of their electric power  needs
and to power total energy systems and water and sewage pumping
units.
              Generally, the large  (above 1000 hp), low-speed  diesel
engines marketed by a number of manufacturers  are designed for con-
tinuous  operation as frequently required in stationary applications.
Conversely,  with few exceptions, the medium (100 to 1000 hp) and
small (below 100 hp) stationary diesel engines are modified versions
of heavy duty on-highway and off-highway truck engines.  To increase
their  service life these engines are  then operated at derated power
and speed settings.  The modifications incorporated into the engines
might include a new head, governor, and fuel injection system.
3.1.3.1       Installed Power
              In 1971 the total estimated installed horsepower of
stationary reciprocating engines was 34.7 x 10  bhp (excluding gasoline
engines).  Of these, liquid-fueled diesel engines  contributed 11.8 x
10  bhp and gas diesels contributed  about  4.1x10  bhp with the
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remainder supplied by spark ignition engines. A breakdown of the
installed diesel engine horsepower for 1971 is presented in Table 3-1
(Ref.  3-9).
3.1.3.Z
Operating Modes
              A review of the available data on electric power capacity
and power generation from all sources in 1970 indicates that recipro-
cating engines, operated on liquid and gaseous fuels, represent about
1.2 percent of the total electric generating capacity in the United States
and about 0. 3 percent of the total power generated (Ref. 3-9).  The
capacity factor of these engines as a whole is only about 12 percent,
indicating that many engines are utilized only for short periods typical
of electric  peaking power and standby installations.  In this context,
      TABLE 3-1.  ESTIMATED INSTALLED DIESEL ENGINE
                    HORSEPOWER FOR 1971a (Ref. 3-9)
Application
Electric Power
Generation "
Oil and Gas
Pipelines
Oil and Gas
Exploration
Agricultural
Water and
Sewage
Total
Diesel Fuel
hp
1,570,000

830,000

1, 500,000
7,500,000

465, 000
11,865,000
Dual Fuel
hp
3,710,000

390,000

-
-

-
4, 100,000
Total Installed
Power, hp
5,280, 000

1,220,000

1,500,000
7,500,000

465,000
15,965,000
A summary table listing the installed horsepower of all stationary
engines is presented in Appendix A.
Estimated 1970 data.
                                  3-9

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the capacity factor is defined as the ratio of the total energy generated
per year by an engine  or engine class to their total energy generation
capacity.   The projected capacity factors for 1980 and 1990 are 8 per-
cent and 7 percent, respectively (Ref.  3-9).
               In continuous power installations, diesel engines are
typically operated at 80 to 90 percent of their rated power for about
6000 to 8000 hours per year. In peaking installations, the engine
operates normally for several hours per day near  full load, whereas
the standby units are generally run for 1 to 3 hours per week.
               In 1971  the total energy generated by diesel engines
used in oil and gas pipeline installations was 5000  x 10  bhp-hr for
liquid fueled engines and 2, 350 x 10  bhp-hr for diesels operated on
natural gas.  Based on the total installed power figures listed in
Table 3-1, the capacity factor of these engines was about 69 percent
which is in reasonable agreement with  data provided by one engine
manufacturer.
               As shown in Table  3-1,  the total diesel horsepower
installed by the oil and gas  exploration industry in 1971 was
1,500,000 bhp. These engines generated 1.94x 10  bhp-hr, result-
ing in a capacity  factor of about 15 percent (Ref. 3-9).
               The capacity factors  assumed by Shell (Ref.  3-9) for
agricultural pumping and municipal  water  and sewage pumping are
40 and 75 percent,  respectively.  Frequently, the  agricultural-pumping
units are operated near their intermittent  rating load levels for periods
extending from 1  to 10 days.
3.1.3.3        Fuel Requirements
               Liquid-fueled diesel engines are generally operated on
No. 2 diesel fuel, although  heavier fuels have been used with good suc-
cess in some of the large low-speed engines.  One manufacturer has
stated that attempts to run his engines  on heavy fuels proved to be
                                3-10

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unsuccessful because of excessive deposit buildup in the exhaust ports
and fuel injectors.  Other potential problem areas noted by several
manufacturers are  related to the high sulfur and metal content, nor-
mally found in heavy fuels,  which could be  detrimental to engine life.
In general,  fuels having a sulfur level of 0.5 percent or less are con-
sidered to be  acceptable, whereas higher concentrations are not
recommended because of the danger  of excessive engine wear and
buildup of harmful  acid compounds in the lubricating oil.
              Dual fuel diesel engines  are frequently utilized by the
electric power generation and petroleum industries.  In these engines
approximately 90 to 95 percent of the total  heat  input is supplied by
natural gas  and the remainder by distillate fuels such  as No.  1  or
No. 2 diesel fuel or fuel oil.  The  objective of the liquid fuel is to
initiate ignition in the  combustion chamber at the desired crank angle
position of the piston.
              The  principal advantages of gas diesels  relative  to oil
diesels include the  lower cost of gaseous fuels and the  longer life of
the lubricating oil. However, in view of the increasing shortage of
natural gas, it may not be possible to maintain the fuel cost differential
much longer.  In general, gaseous fuels result in  lower pressure rise
rates and peak pressures in the combustion chamber which is desirable
from an engine durability point of view.
3.1.4         Emissions
              Diesel engines,  like other heat engines, emit a number
of different  pollutant species,  including hydrocarbons,  carbon monoxide,
oxides of  nitrogen,  oxides of sulfur,  aldehydes,  particulates, smoke,
and odor. Except for  the oxides of sulfur and nitrogen these compounds
are the result of imperfect  or incomplete combustion.  To provide a
better  understanding of the  problems related to  diesel engine  emissions,
the gaseous pollutants and their formation in diesels are briefly
described in the following section.
                                 3-11

-------
3.1.4.1        Pollutant Formation
3.1.4.1.1     Hydrocarbons
               The hydrocarbons (HC) emitted from diesel engines are
considered to be the result of quenching of oxidation reactions occur-
ring in the wall region of the combustion chamber and in ultra-lean
zones  of the air-fuel mixture (Refs.  3-2 and 3-10 through 3-12).  Fac-
tors influencing the level of HC emissions include the degree of turbu-
lence inside the combustion chamber, and certain design details in the
injection and combustion system geometry affecting the formation of
the fuel spray (Ref. 3-2).   Relative to spark-ignition engines,  the
HC emissions from diesels at rated conditions are  low but tend to
increase with decreasing load.  This is further discussed in
Section 3. 1.4.2.
3.1.4.1.2     Carbon Monoxide
               Carbon monoxide is the  result of a deficiency in oxygen
during the combustion process.  While the overall air-fuel mixture
ratio in diesel  engines is lean,  oxygen deficient regions exist through-
out the combustion chamber. Although sufficient oxygen is ultimately
available to complete the CO reactions, this oxygen may not reach the
CO molecules before the temperature of the gases in the chamber
drops to a value too low for oxidation to proceed.  In general,  the
CO emissions from diesel engines are  rather low at full load and
decrease further as engine load is reduced (increasing air-fuel ratio).
3.1.4.1.3     Oxides  of Nitrogen
               The oxides  of nitrogen are the principal pollutant species
emitted from diesel engines. Their  formation during the combustion
process is kinetically  controlled and increases with increasing
                                3-12

-------
temperature,  oxygen concentration, and residence time of the gases
in the chamber (Ref. 3-2).
              As further discussed in Section 3. 1.4.2, the NO   emis-
sions from prechamber engines are generally lower than from existing
open chamber diesels.  In prechamber engines,  combustion begins
under fuel-rich conditions in the prechamber near the top dead center
(TDC) position of the piston.  Because of a lack  of oxygen, very little
NO is formed under these  conditions.   As combustion proceeds in the
prechamber, the burning charge is expanded into the main chamber
where an adequate amount of oxygen is available to complete the  reac-
tions.  Since the  air in the main chamber is relatively cool, the
NO formation reactions are quenched  rapidly, hence minimizing  NO
                                                                 }t
(Refs. 3-2 and 3-13).  In addition, the rapid  rate of  energy release in
divided-chamber engines provides for optional timing at a  more
retarded position compared to open-chamber engines.
              In open-chamber engines,  combustion  starts several
crank degrees before TDC.  Since the injection delay period is rela-
tively long in these engines,  there is an appreciable amount of vapor-
ized fuel available at the time  of ignition.  This  results in  very rapid
combustion followed by compression heating  leading to high local flame
temperatures and NO  formation rates (Refs. 3-13 and 3-14).
              Insufficient information is currently available regarding
the effects of fuel-bound nitrogen on the NO  emissions from diesel
                                          n
engines. However,  it is well known that in utility boilers up to
70 percent of the bound nitrogen is converted to  NO  (Ref.  3-15).
                                                 JC
3.1.4.1.4     Oxides of Sulfur
              The oxides  of sulfur emissions from diesel  engines are
directly  related to the sulfur content of the fuel.  In the absence of
adequately developed instrumentation,  it is generally assumed that all
                                 3.13

-------
 sulfur oxidizes to sulfur dioxide.  Thus, the mass emissions of SO
 are twice the mass of sulfur contained in the fuel (Ref. 3-16).
 3.1.4.1.5    Aldehydes
              Aldehyde emissions, which are the result of partial oxi-
 dation of hydrocarbons in the combustion process, may be a contribu-
 tor to diesel engine odor.   Test data indicate that many diesel engines
 emit  significant amounts of aldehydes.
 3.1.4.1.6    Particulates and Smoke
              Particulate matter found in the exhaust of diesel engines
 consists of impurities contained in the fuel and unburned- or partially-
 burned hydrocarbon molecules.  Frequently, carbon particles are
 emitted from diesel engines in the form of smoke which is formed in
 fuel-rich zones having  sufficiently high temperature to decompose the
 fuel.  In general,  smoke levels of well adjusted diesels are low except,
 perhaps,  at full load  where the  air-fuel ratio approaches  stoichiometric,
 Smoke and CO are formed under much the same conditions and their
 concentrations are generally proportional because both are  dependent
 upon the availability of oxygen (Ref.  3-2).
 3.1.4.1.7    Odor
              The odor from diesel  engine exhaust has long been
 recognized as undesirable.  Although the odor formation mechanism
has not yet been fully explained, there is evidence that the diesel odor
is the result of partial oxidation reactions taking place in the  lean
 regions of the combustion  chamber.
 3.1.4.2       Test Procedures and Instrumentation
              To simulate the operation of the engine in the field,  the
emission test procedures employed by most manufacturers of heavy-
duty diesel engines and by independent test laboratories are based on
                                 3-14

-------
measurements at steady-state operating conditions.  Emission maps
have been determined for many engines by operating the engine at a
number of discrete speed and load settings.  In particular, this pro-
cedure is utilized for engines used in both mobile and stationary instal-
lations.  Conversely, the emission data available for large stationary
engines are frequently limited to their  rated speed/load points.
              For  some engines, the available emission data are based
on the  13-mode procedure which has been developed to characterize the
HC,  CO, and NO  emissions of heavy-duty engines used in on-highway
                5C
truck applications (Ref.  3-17).  In this procedure the engine is operated
at 13 points (11 discrete sets of speed and load) and the specific mass
emissions over the cycle are then determined  by multiplying the mass
emissions obtained at each mode point  by  an appropriate weighting
factor  and dividing  the sum of these values by  the average power output
of the engine.
              The  instrumentation employed in diesel  engine  emission
test work includes nondispersive infrared (NDIR) analyzers for the
measurement of CO and NO concentrations and heated flame-ionization-
detectors (FID) for HC,  to prevent condensation of the  heavier
HC species  in the sampling lines.  More  recently, the  chemilumines-
cence analyzer has been added by some investigators to measure NO
and NO  separately.  Tests conducted by the Southwest Research
Institute on locomotive diesel engines indicate relatively small differ-
ences in the NDIR and chemiluminescence readings (Ref.  3-18).
              Aldehydes are normally determined by wet chemistry
and particulates are measured by gravimetric methods.
              Full-flow light extinction type analyzers, such  as the
PHS smoke  meter,  are preferred for diesel exhaust smoke measure-
ments  because of their good accuracy and fast response.  In cases
where  this technique would be difficult  to implement, other instruments
                                 3-15

-------
such as the Bosch and AVL smoke meters have been applied with
good success (Refs. 3-2 and 3-19).
3.1.4.3       Gaseous Emissions
               This section includes  a discussion of the HC,  CO, and
NO  emission characteristics of current uncontrolled heavy-duty
diesel engines.  The emissions at rated engine operating conditions
are presented in Section 3.1.4.3.1 and the part load emissions are
discussed in Section 3.1.4.3.2.  Although many of the engines con-
sidered in the study were originally designed for mobile applications,
their emission characteristics are considered to be applicable to
stationary engines  as well, because of many design and operational
similarities between the various diesel engine classes.
               The emission data and correlations presented in the
following sections are based on data  published in the open literature
or were  acquired directly from various  engine manufacturers.
3.1.4.3.1     Rated Conditions
              The emissions at rated conditions of many naturally
aspirated and turbo-charged, open-chamber and divided-chamber,
four- and two-stroke diesel engines are presented in Tables  3-2
through 3-7, in terms of pollutant specie concentrations and  specific
mass emissions.  Smoke, odor, air-fuel ratio, and specific  fuel con-
sumption data are also listed in these tables and these parameters
will be discussed in later sections of the report.  For  many engines,
the data  shown represent average values from two or more test runs.
              The emissions from eleven four-stroke, naturally
aspirated,  open-chamber diesel engines tested by a number of investi-
gators are listed in Table 3-2.  Ten  of these engines were  originally
designed for mobile (on- or off-highway) applications,  but several of
them are offered also for use in stationary installations.  Engine No. 11
                                3-16

-------
TABLE 3-2.
EMISSIONS FROM FOUR-STROKE,  NATURALLY ASPIRATED OPEN
CHAMBER DIESEL ENGINES - RATED CONDITIONS
Manufacturer
General Motors
Mack
Caterpillar

International
Inte mational
Caterpillar
Cummins
Perkins
-
-
Engine
Identi-
fication
No.
1
2
3
4
5
6
7
8
9
10
11
CID
478
678
573
-
407

-

354
-
Large
No. of
Cylin-
ders
6
6
6
-
6
6

-
6

-
Power, bhp
Rated
155
185
-
-
112
200
225
260
112
-
-
Measured
152
172
186
172
110
175
210
242
104
209
-
Speed, rpm
Rated
3200
2100

-
2400
3000
2800
2100
2800

low
Measured
3200
2100
3000
2100
2500
3000
2800
2100
2800
2100
-

HC
1000
630
420
680
930
470
200
71
230
210
-
CO
1400
1400
1 100
1000
1900
540
2075
1493
3300
2100
-
NOX
740
2550
1300
2250
1456
1030
1200
1231
1130
840

Specific
HC
2.2
1.3
O.S
1.3
1.9
1.0
0.4
0. 1
0.5
0.5
0.5
CO
6. 1
5.6
4.2
3.9
7. 5
2.2
7.3
5.3
14.6
8. 7
4.2
NO,
5.3
16.6
8.3
14.4
9.5
6.8
6.9
7.2
8.3
5.6
15.2
Smoke
Opacity,
%
16
2
5
-

7
14
19
20
-
-
Odor
Intensity,
DI Units8
4. 5
6.4
5.9
6.9
-
-

-
5.9
4.6
-
Specific
Fuel Con-
sumption,
Ib/bhp-hr
0. 395
0.421
0. 357
0.402
0.405
0.412
0.410
0.433
0.470
0.464

Air
Fuel
Ratio
24. 1
26.2
24.5
24.5
21.4
21.2
18.6
17. 8
20. 7
31. 7
-
Ref-
er-
ence
3-12
3-12
3-12
3-12
3-16
3-20
3-20
3-20
3-12
3-12
-
Diesel Intensity
    TABLE 3-3.
    EMISSIONS FROM FOUR-STROKE,  TURBOCHARGED, OPEN
    CHAMBER DIESEL ENGINES - RATED CONDITIONS
Manufacturer
Mack
Allis Chalmers
John Deere
Mack
Mack
Cooper
Cooper
Bes semer
_
.
_
.
aDiesel Intensity
Engine
Identi -
lication
No.
12
13
14
15
16
17
,7»

IB
18b
19
19b
CID
673
426
404
673
865
Large
Large

Large
Large
Large
Large
No. o(
Cylin-

-
-
6
-
-
12
12


-
.
-
Po

232
157
129
235
325
4300
4300

.


-
wer, bhp

232
148
136
228
290
4300
4300

-
-
-
-
Speed, rpm

2100
2200
2200
2100
-
600
600

Low
Low
Low
Low

2100
2200
2200
2100
2400
600
600


-
-
-
Con
HC
1140
137
844
423
331

.


-
-
-
centration,
ppm
CO
1000
643
530
720
518
-
.

-
-
-

NO,
1440
1510
1158
1601
1488
-
-

-
-
-

Specific
Mass Emission,
g/bhp-hr
HC
2.9
0.3
2. 1
1.0
1. 1
0. 1
5.2C

0.8
0.5
0.5
0.5
CO
5.2
2. 7
2.6
3.4
3. J
3.9
4.5

2.8
2.4
1. 1
1. 1
NOX
12. 2
10.6
9.3
12.5
15.6
II. 0
9.0

9.0
15.6
13. 1
11.0
Smoke
Opacity,
%
2
•
-
3. 5
2. 8

-


-
-
-
Odor
Intensity,
DI Units'
4.0
-
-
-
-
-
-


-
-
-
Specific
Fuel Con-
sumption,
Ib/bhp-hr
-
0.406
0.410
0.392

-
-

0. 395
6400d

-
Air
Fuel
Ratio
22.4
23.0
25. 8
26.7
29.6
-
.

31.3
34.0

-
Ref-
er-
ence
3-12
3-16
3-16
3-20
3-20
3-21
3-21

-
-
-


bCas - diesel
"•Not used in averaging
dBtu/bhp-hr

-------
            TABLE 3-4.
 EMISSIONS FROM FOUR-STROKE NATURALLY ASPIRATED,
 DIVIDED CHAMBER DIESEL ENGINES -  RATED CONDITIONS
Manufacture r
Daimler- Benz
Daimler-Benz
Engine
Identi-
fication
No.
20
21
CID
108
121
No. of
Cylin-
ders
^
4
Pou.er. bhp
Rated
29
60
Measured
28
57
Speed, rpm
Rated
2400
4200
Measured
2400
4200
ppm
HC
123
55
CO
2025
832
NOX
280
601
Specific
HC
0. 3
0. 1
CO
9. 1
3. 3
NO,
2. 1
4.0
Smoke
Opacity,
"o
-
Odor
Intensity,
DI Units"

Specific
Fuel Con-
sumption,
Ib/bhp-hr
0. 528
0. 526
Air
Ratio

Ref-
er-
ence
3-14
3-2Z
aDiesel Intensity
u>
I
CO
           TABLE 3-5.
EMISSIONS FROM FOUR-STROKE,  TURBOCHARGED, DIVIDED
CHAMBER DIESEL ENGINES - RATED CONDITIONS


Caterpillar
Caterpillar
Caterpillar
Cate rpillar
Engine
Identi-
fication
No.
22
23
24
25
CID
638
No. of
Cylin-
ders
6
Power, bhp
Rated
285
149
245
285
Measured
279
142
246
285
Speed, rpm
Rated
2200
1900
2200
2200
Measured
2200
1900
2200
2200

ppm
HC
80
23
36
CO
200
101
52
NO,
700
545
664
Specific
Mass Emission,
HC
0. 2
0. 1
0. 1
0. 2
CO
0. 9
0. 5
0. !
0. 7
NO,
5. 0
4. 3
5.0
4.3
Smoke
Opacity,
T.
2. 0
2.0
Odor
Intensity,
DI Units"
3. b
Specific
Fuel Con-
sumption ,
Ib/bhp-hr
0.425
0. 399
0. 399
Air
Ratio
24.7
26. 2
25. 1
Ref-
er-
ence
3-12
3-16
3-20
3-23
aDiesel Intensity

-------
TABLE 3-6.
EMISSIONS FROM TWO-STROKE,  OPEN CHAMBER, BLOWER
SCAVENGED DIESEL ENGINES - RATED CONDITIONS
Manufacturer
General Motors
Gene ral Moto ra
General Motors
Engine
Identi-
fication
N'o.
26
27
28
CID
426
284
-
No. of
Cylin-
ders
.
4
-
Power, bhp
Rated
206
Ml
-
Measured
208
1-41
-
Speed, rpm
Rated
2100
2100
-
Measured
2100
2100
2100
Concentration,
ppm
HC
130
170
199
CO
565
1000
1134
NOX
1675
1350
1392
Specific
Mass Emission,
HC
0.4
0.5
0.6
CO
3.6
5.6
7. 1
NO,
17. 1
12. 5
14.4
Smoke
Opacity,
%
.
-
1. 5
Odor
Intensity,
DI Unitsa
.
5.6
-
Specific
Fuel Con-
sumption,
Ib/bhp-hr
0.422
0.413
0.421
Air
Ratio
33.0
29.5
33. 2
Ref-
er-
ence
3-16
3-12
3-20
aDiesel Intensity
  TABLE 3-7.
  EMISSIONS FROM LARGE TWO-STROKE,  OPEN CHAMBER,
  TURBOCHARGED DIESEL ENGINES - RATED CONDITIONS
Manufacturer
General Motors
Engine
Identi-
fication
No.
29
30
31
'Diesel Intensity
"Dual fuel
Btu/bhp-hr
CIO
10,320
Large

No. of
Cylin-
ders
16

Power, bhp
Rated
3200

Measured
3208

Speed, rpm
Rated
900
Measured
900
Concentration,
ppm
HC
238

CO
567

NO,
945

Specific
Mass Emission,
g/bhp-hr
HC
0.6
0 »b
CO
3.0
1 6
1 o»
NO,
8.0
8 9
9 Ob
Smoke
Opacity.
Tt
-
Odor
Intensity,
DI Units"
;
Specific
Fuel Con-
sumption,
Ib/bhp-hr
0. 392
0. 379
6900C
Air
Fuel
Ratio
37.0
31.5
Ref-
er-
ence
3-18


-------
represents a large low-speed diesel engine which is utilized in stationary
installations only.  As indicated, the emission concentrations  and spe-
cific mass emissions show considerable variations.  On a percentage
basis, the largest variations are observed for HC, reflecting the pre-
viously noted emission sensitivity to injector  and combustion chamber
design details.  HC and CO of the large engine (No. 11) are consider-
ably lower than the average  emissions of this  engine  class, while NO
                                                                   X,
is substantially higher.  Attempts to correlate the emissions with dif-
ferent engine parameters  such as speed, air-fuel ratio, and brake
mean-effective-pressure (torque) proved to be unsuccessful, primarily
because of a lack of sufficient data from large engines.
               Table 3-3 presents the emissions  at rated conditions for
eight four-stroke,  turbocharged, open-chamber diesels.  Engines
No.  12  through  16 are utilized primarily in mobile applications whereas
Engines No.  17, 18,  and 19  are  large-size, low-speed stationary
engines which are operated either on No. 2 diesel fuel or natural gas.
The  average NO specific mass  emission of the  stationary engines is
                X.
about 12 g/bhp-hr compared to an average value  of about 11.5 g/
bhp-hr  for the  "mobile" engines.  This difference is  considerably
smaller than the observed engine to engine variability in NO .  The HC
and CO emissions of the large diesels  are  somewhat  lower than the
average levels obtained for the  smaller "mobile" units.  With  respect
to fuel effects on emissions, the data indicate very little variations in
HC,  CO,  and NO  when operating on No. 2 diesel fuel or natural gas.
Comparison with the  data in Table 3-2 indicates similar average
HC emissions for the naturally  aspirated and  turbocharged engines;
whereas CO of the turbocharged engines is considerably lower, reflect-
ing the higher air-fuel ratio used in this engine type.  Conversely, the
average NO   of turbocharged diesels is higher than for naturally
           X.
aspirated engines.  This is attributed to the higher intake manifold
                                3.20

-------
temperature  associated with turbocharging, which results in higher
compression and peak combustion temperatures,  hence higher
NO  formation rates.
   x
              Test data from two small four-stroke naturally aspirated
divided-chamber diesel engines are listed in Table 3-4.  As expected
from theoretical considerations, the NO  emissions of these engines
                                      J\.
are markedly lower than those of the open chamber engines shown in
Tables  3-2 and 3-3,  and HC  is extremely low as well.
              The emissions from four four-stroke, turbocharged,
divided-chamber diesel engines manufactured by Caterpillar Tractor
Company are listed in Table 3-5. Again the HC and CO emissions are
extremely low,  owing to the  staged combustion process employed in
divided-chamber diesels.  The NO  emissions of these engines are
                                 JC
considerably lower than in open-chamber engines but somewhat higher
than for the naturally aspirated divided-chamber engines listed in
Table 3-4.
              Two-stroke, open-chamber, blower-scavenged, and
turbocharged diesel engine data are presented in Tables 3-6 and 3-7,
respectively.  The blower-scavenged engines were manufactured by
General Motors and are used primarily in mobile applications, while
the turbocharged engines represent large, low-speed stationary designs.
Compared to the open-chamber, four-stroke engines, the medium-size
blower-scavenged, two-stroke diesels have lower HC,  comparable CO
but higher NO  emissions.  Conversely, the average emissions of the
             j\.
large turbocharged two-stroke diesels tend to be somewhat lower than
for the four-stroke designs.  However, when making comparisons
between the various diesel engine categories, the small data sample
currently available for  some of the engine categories must be  taken
into consideration.
                                3-21

-------
 3.1.4.3.2     Part Load Emissions
               HC, CO, and NO  exhaust emission concentrations of
                              X,
 several four-stroke, naturally aspirated,  open-chamber diesel engines
 are plotted in Figures 3-1 and 3-2 as a function of air-fuel ratio.  Also
 shown in Figure 3-2 are brake-mean-effective-pressure (load) data
 correlated with air-fuel ratio.  At full load the air-fuel ratio of four-
 stroke naturally aspirated diesels varies between about 19 and 24, but
 decreases rapidly with decreasing load.  As shown in Figure 3-1 the
 HC emission concentrations show considerable engine to engine varia-
 bility but  seem to be essentially independent of air-fuel ratio.  The CO
 exhaust concentrations have a maximum in the low  air-fuel ratio regime
                                          i
 corresponding  to high  engine loads and decrease rapidly,  as the mixture
 is leaned  out.  However,  increasing CO concentrations  are observed in
 some engines under very lean, light-load operating conditions which is
 attributed to quenching effects.   As expected,  the NO  concentrations
                                                   X.
 reach a maximum at rated engine conditions and decrease rapidly with
 decreasing load (increasing  air-fuel ratio), as a result  of a reduction
 in the quantity  of air involved in the combustion process.  Similar
 trends were obtained for other diesel engine classes.
               Typical part load HC, CO,  and NO  emission curves for
 the various diesel engine designs are presented in Figures 3-3 through
 3-9 in terms of specific mass emissions versus percent of maximum
 engine power.  Unlike  the data listed in Tables 3-2 through 3-7, the
 curves shown in these  figures represent individual test  runs.
               The specific mass emission characteristics of a typical
 four-stroke naturally aspirated,  open-chamber diesel engine are illus-
trated in Figure 3-3.  Design and operational details of this engine are
listed in Table  3-2. As indicated,   HC increases steadily with decreas-
 ing load and speed.  This trend follows from the constant HC concen-
trations shown  in Figure 3-2 and the higher exhaust flow rates per
                                3-22

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^V7

^ (8> A
O ©
°* *«b° %
D
Kl 1 1 1 1 1 1
N
"i
«gp

"
N 10 20 30 40 50 60 70 «
AIR/FUEL RATIO
        Figure 3-1.  HC and CO emissions
                     vs air-fuel ratio -
                     four-stroke,  naturally
                     aspirated, open-cham-
                     ber diesels
Figure 3-2.
NOX emissions and
brake  mean effec-
tive  pressure  vs
air fuel ratio - four-
stroke, naturally
aspirated,  open-
chamber diesel
engines

-------
            10
U)
i
          o
          (J
            12
           .


          "
                   20    40    60     80

                  PERCENTAGE OF MAXIMUM POWER
                                         100
                                                          1
                                              20     40     60    60
                                              PERCENTAGE OF MAXIMUM POWER
                                                                     100
       Figure 3-3.
Specific mass emissions -
four-stroke, naturally
aspirated, open-chamber
diesel (Engine No. 3)
Figure 3-4.  Specific mass emissions -
             four-stroke turbocharged,
             open-chamber diesel
             (Engine No. 13)

-------
                         A DIESEL FUEL
                         O NATURAL GAS
            20     40     60     80    100
            PERCENTAGE OF MAXIMUM POWER
Figure  3-5.   Specific mass emissions -
              four-stroke turbocharged,
              open-chamber large diesel
              (Engine No.  18)
                                                                         O 1400 rpm
                                                                         A 2400 rpm
           20    40     60    80
          PERCENTAGE OF MAXIMUM POWER
                                100
Figure  3-6.   Specific mass emissions -
              four-stroke, naturally
              aspirated divided-chamber
              diesel (Engine No. 20)

-------
to
I
                                                               40
                                                             * 30
                                                             O.
                                                             £
                                                             •
                                                             O 20
                                                               10
                                                                4
                                                              i.


                                                              I2
                                                                      I      I
                0     20    40     60    80    100
                   PERCENTAGE OF MAXIMUM POWER
         0     20    40     60    80    100
            PERCENTAGE OF MAXIMUM POWER
          Figure 3-7.  Specific mass  emissions -
                       four-stroke turbocharged,
                       divided chamber diesel
                       (Engine No.  23)
Figure 3-8.  Specific mass emissions
              two-stroke diesel
              (Engine No. 26)

-------
                         20   40    60    80
                       PERCENTAGE OF MAXIMUM POWER
                                            100
                 Figure 3-9.  Specific mass
                              emissions - large
                              two-stroke, turbo-
                              charged diesel
                              (Engine No. 30)
horsepower output obtained with decreasing load.  The other naturally
aspirated,  open-chamber diesels evaluated exhibit similar trends,
except for Engines  10 and  11 (Table 3-2) which show a small reduction
in HC down to about 60 percent of full load, followed by a rather sharp
rise.   The  CO specific mass emissions decrease with decreasing  load,
reaching a minimum at about 60 to 70 percent of full load and then
increase again as load is further reduced.  The NO  emissions of
Engine No.  3 increase steadily with decreasing load,  in spite of the
previously noted reduction in NO  concentration. Seven of the eleven
naturally aspirated, open-chamber diesel engines considered in this
study show  similar NO   trends.  Conversely, in the remaining engines
                     n
NO  decreases somewhat with decreasing load.  The  difference in the
NO  versus load relationship might be due to differences  in the fuel
injection timing schedules utilized on the various engines.  In some
engines, the timing  is retarded as load is reduced whereas  others
                                 3-27

-------
maintain a constant start of injection which in effect results in a timing
advance because the injection period becomes shorter with decreasing
fuel flow rate (load).  As further discussed in Section 4.1.1.3,  injection
timing retard reduces NOX whereas advanced timing increases NOX.
               Test data from two four-stroke turbocharged, open-
chamber diesel engines are shown in Figures 3-4 and 3-5.  Similar to
naturally aspirated diesels, the HC and CO specific mass emissions
follow the expected trends, although the rates of change appear  to be
lower with turbocharging.  At rated speed, the NOX specific mass
emissions  of Engine No.  13 remain essentially constant down to about
40 percent load but increase rapidly at lower loads.   Conversely,  at
the intermediate speed, NOX increases steadily with decreasing load.
Engine No.  18 (Table 3-3), when operated on No. 2 diesel fuel,  shows
rising NOX as the load  is reduced,  whereas a reduction in NOX is
realized with natural gas.  The difference in  the NOX trends is attri-
buted to  differences in the engine air-fuel ratio versus load and timing
versus load schedules utilized for the two fuels.
               Figure  3-6 presents the emission map of a small,
naturally aspirated, divided-chamber diesel engine operated at two
speeds.  The HC emissions remain essentially constant between full
load and 40-percent load but increase sharply at load levels below
40 percent. Again,  CO follows the trends of  the other diesel engine
categories while NOX tends to increase with decreasing load.  Engine
No.  21 shows  similar trends.
               The specific mass emissions of a typical four-stroke,
turbocharged,  divided-chamber diesel engine are presented in Fig-
ure 3-7, showing steadily increasing HC  as engine load is reduced.
As in open-chamber, turbocharged diesels, CO varies very little in
the 60- to 100-percent load regime but increases rapidly at lower
loads. NOX increases  steadily with decreasing load.
               Emission maps  of two typical two-stroke diesel engines
are presented in Figures  3-8 and 3-9.  One of these is a medium-size,
blower-scavenged engine  operated on No.  2 diesel  fuel while the other
                                3-28

-------
is a large, turbocharged unit which was run on diesel fuel and on natural
gas.  The emission characteristics of these engines are comparable to
many four-stroke engines.
              In general,  the HC and CO emissions of the two-stroke
engines are quite low, reflecting the high air-fuel ratio at rated con-
ditions.  Again, the different mass  emission trends obtained with the
two types of fuel are  attributed to differences in the air-fuel ratio
schedules used for the two fuels.
3.1.4.3.3    Selected Diesel Emission Values
              Based on the analysis of the available diesel  engine data
discussed Sections 3. 1. 4. 3. 1 and 3. 1.4. 3. 2  average HC, CO,  and NOX
emission concentrations and specific mass emissions at rated and part-
load operating conditions have been determined for the following uncon-
trolled engine categories.
         1.    Four-stroke, naturally aspirated, open-chamber
         2.    Four-stroke, turbocharged, open-chamber
         3.    Four-stroke, naturally-aspirated, divided-chamber
         4.    Four-stroke, turbocharged, divided-chamber
         5.    Two-stroke, blower-scavenged
         6.    Two-stroke, turbocharged
              The average pollutant specie concentrations and specific
mass emissions at rated engine conditions  are  presented in Table 3-8.
Where appropriate, the emissions are listed separately for the
medium-size engines, most of which were  originally designed for
mobile installations,  and the  large stationary engines.  Also listed in
this table are the observed emission variations.  As indicated in the
table, the emissions  from these uncontrolled engines vary over wide
ranges.  For example,  in naturally aspirated,  open-chamber engines,
the HC emissions vary between 0. 1 g/bhp-hr and 2. 2 g/bhp-hr.  In
view of the magnitude of these variations it is very difficult to establish
typical emissions for each engine category.

                                 3-29

-------
             TABLE 3-8.  AVERAGE DIESEL ENGINE EMISSIONS AT RATED CONDITIONS
                           (UNCONTROLLED ENGINES)

Engine Category


Four-Stroke ,
Naturally Aspirated,
Open Chamber
Four-Stroke,
Turbo -Charged,
Open Chamber
Four-Stroke,
Naturally Aspirated,
Divided Chamber
Four-Stroke ,
Turbo -Charged,
Divided Chamber
Two-Stroke,
Blower Scavenged
Two -Stroke,
Turbo -Charged3

Size


Medium
large
Medium
large3
Small,
light
duty

Medium

Medium
Large

Average
Concentrations ,
ppm
HC

485
575

90


46

167
-
CO

1630
680

1430


118

906
-
NO

1370
1440

440


636

1472
-
Average
Specific
Mass
Emission ,
g/bhp-hr

HC
1.0
0.5
1.5
0.6

0.3


0.2

0.5
0.4

CO
6.5
4.2
3.4
2.6

7.3


0.6

5.4
2. 1

NOX
8.9
15.2
12.0
11. 5

3. 1


4.7

14.7
8.6

Concentration Range,
ppm
HC

70 - 1000
130 - 1140

55 - 120


23 - 80

130 - 200
-
CO

540 - 3300
530 - 1000

830 - 2025


52 - 200

585 - 1 135
-
NO
X

740 - 2550
1150 - 1600

280 - 600


545 - 700

1350 - 1675
-

Specific Mass
Emission Range,
g/bhp-hr
HC

0.1 -2.2
0
0.3-2.9
0.1 -0.8

0. 1 -.0.3


0. 1 - 0. 2

0.4 - 0. 6
0. 3 - 0. 6
CO NO

2.2 - 14.6
ne engine o
2.6 - 5.2
1.1-4.5

3.3 - 9. 1


0.3 - 0.9

3.6 - 7.1
1.6 - 3.0
5.3 - 16. 6
nly
9.3 - 15.6
9.0 - 15.6

2.1-4.0


4.3 - 5.0

12. 5 - 17. 1
8.0 - 9. 0
3With intercooling
UJ
I
OJ
o

-------
               According to Table 3-8, the divided chamber diesels
show superior emission performance particularly with respect to HC
and NOx-  In general, the specific mass emissions of the large,  low-
speed stationary diesels tend to be  lower than for the  medium-size units.
               Part-load emission correlations for the six diesel engine
categories identified above are presented  in Figures 3-10 through 3-12,
showing HC,  CO,  NOx emission ratios at  rated speed as a function of
the ratio of engine power to rated power.  The emission ratios are
defined as the ratio  of the specific mass emissions at a given power
output point to the specific mass emissions at full load.  The  curves
represent arithmetic averages of all the available data in each engine
category. It should be noted that only a very limited number  of data
points are available for  several of the engine categories and therefore
the trends shown in  the figures may not be representative of all engines.
               As  illustrated in Figure 3-10, the average HC  specific
mass emission ratio of all diesel engine categories  increases steadily
with decreasing load. The naturally-aspirated diesels show a larger
rate of increase than the turbocharged and blower scavenged designs.
               The CO specific mass emission ratios  shown in Fig-
ure 3-11 indicate significant variations.  In most cases, CO decreases
rapidly as the load is reduced from  the rated value, but increases again
8       Four-stroke, naturally aspirated, op«n chamber
       Four-stroke, turbocharged. open chamber
      ~ A Four-stroke, naturally aspirated, divided chamber"
      O Four-stroke, turbocharged, divided chamber
      O Two-stroke, blower scavenged
      _ • Two-stroke, turbocharged
                                   Figure 3-10.
Diesel engine part-
load hydrocarbon
emissions - rated
speed
       30  40  SO  60   70  80
          PERCENTAGE OF RATED POWER
                                  3-31

-------
  3.5
  3.0
  2.5
  2.0
        OFour-stroke, naturally aspirated, open chamber
        D Four-stroke, turbocharged. open chamber
      - A Four-stroke, naturally aspirated, divided chamber-
        O Four-stroke, turbocharged, divided chamber
        O Two-stroke, blower scavenged
        • Two-stroke, turbocharged
  1.5
u
3
I
U

S  '•"
Q.
I/I
O
O  0.5
     0  30   40    50    60    TO   60

            PERCENTAGE OF RATED POWER
                                       90   100
  Figure 3-11.
                       Diesel engine part-
                       load carbon monox-
                       ide  emissions  -
                       rated  speed
^2.5
 o-  2'°
   1.5
   1.0
n-  0.5

id
a.
       O Four-stroke, naturally aspirated, open chamber
       D Four-stroke, turbocharged. open chamber
       A Four-stroke, naturally aspirated, divided chamber
       O Four-stroke, turbocharged, divided chamber
      - O Two-stroke, blower scavenged
       • Two-stroke, turbocharged
      0  30    40    50    60   70    80

             PERCENTAGE OF RATED POWER
                                        90  100
  Figure 3-12.
                       Diesel engine part-
                       load oxides of ni-
                       trogen emissions -
                       rated  speed
                       3.32

-------
at very low load levels.  This initial reduction of CO is related to the
associated increase in air-fuel ratio, whereas the increase at low loads
is attributed  to quenching effects.
              The NOX specific mass emission ratios at rated speed
are presented in Figure  3-12.  Except for the four-stroke, naturally
aspirated divided-chamber diesel engines,  the variations  in NOX in the
50 to 100 percent load regime remain within ±20 percent.   However,
it should be emphasized  again that individual  engines in some of the
engine categories  considered here show considerably larger variations.
For  example, at 40 percent load,  Engine 30 when operated on natural
gas shows a NOX emission ratio of 0. 52 instead of the  average value
of 0.7 plotted in the figure.
              Somewhat different part-load emission characteristics
are obtained  when engine speed and load are reduced simultaneously.
In this case,  the HC and CO  specific mass  emission factors at part-
load would be as much as 50 percent lower.  However,  for some
engines this reduction would be accompanied  by an increase in NOX
while for others there would be no change in NOX or even  some reduc-
tion.  Although there is insufficient data available to permit the  formu-
lation of a generally applicable speed versus  load correlation, it is
conceivable that some reduction in NOX could be achieved by properly
derating the engine.  This is further discussed in Section  4.1.1.1.
              The part-load emission factors presented in Figures  3-10
through 3-12 combined with the average specific mass emissions listed
in Table 3-8  for the rated engine  conditions provide all the information
necessary for the  computation of predicted specific mass  emissions at
part-load.
3.1.4.4      Smoke and Particulate Emissions
3.1.4.4.1     Smoke
              The smoke emitted from diesel engines can be grouped
into three categories:  white, blue,  and black.  White smoke, or cold
                                 3-33

-------
smoke, which usually appears in the exhaust during an engine cold
start,  consists primarily of raw fuel combined with other compounds
such as aldehydes.  It is formed in the wall region of the cylinders
where the temperatures are not high enough to ignite the  fuel.  Blue
smoke, emitted by some diesels, is the result of the combustion of
excessive amounts of lubricating  oil. Black smoke, or hot smoke,
consists of agglomerated soot particles containing very small (200-
300 Angstrom) carbon particles and some hydrogen.
              Black smoke is the most important specie and is
formed by vapor phase pyrolysis  of fuel molecules which occurs when
insufficient oxygen is available in the high temperature zones of the
combustion chamber (Ref.  3-Z4).  The soot formation process depends
on (1) the amount of incompletely mixed fuel and the fuel-air ratio of
the mixture at the time of combustion, and (Z) the associated tempera-
tures.  An increase in temperature generally increases  soot release.
Once formed, the soot particles  must find oxygen to burn and if this  is
not accomplished within the confines of the combustion chamber,
visible  smoke appears in the exhaust (Ref. 3-25).   In general,  inadequate
fuel atomization  and mixing with the air, overfueling, and maladjustment
in the fuel injector system  tend to increase  soot formation (Ref. 3-2).
Black  smoke is more likely to occur at high loads where less excess
air is available for combustion.  Normally, turbocharged diesels
operate at higher air-fuel ratios than naturally aspirated engines, and,
as a result they tend to emit less smoke.  Also, high-pressure  fuel
injection systems are believed to contribute to lower smoke levels
(Ref.  3-26).
              Smoke data from four naturally  aspirated, four-stroke
diesel engines are depicted in Figure 3-13 (Ref. 3-12).  The sharp
increase in the smoke level of Engines 1 and 9 (designated Engines B
and A in Ref.  3-12) at high  loads may be caused by poor injector
                                 3.34

-------
 5
 •o
x
ui
                                          Figure 3-13.  Exhaust smoke
                                                       vs output
                                                       power for four
                                                       open-chamber,
                                                       naturally
                                                       aspirated
                                                       diesel engines
                                                       (Ref. 3-12)
 Z5       50       75
PERCENTAGE OF RATED POWER
100
performance resulting in inadequate fuel atomization and fuel
impingement on the cylinder walls.  As shown in Tables 3-2 and 3-3,
the average smoke opacity of naturally  aspirated diesel engines at
rated load is about 13 percent,  while the turbocharged engines show
only about three-percent opacity.  There is insufficient data to
characterize the smoke  emissions of the other diesel engine categories.
However, it is believed  that little smoke should be emitted by the two-
stroke engines because of the relatively high air-fuel ratio utilized
in these engines under all operating  conditions.
              Average  steady-state smoke intensity data reported by
Southwest Research Institute for a number of naturally-aspirated and
turbocharged, open chamber and prechamber diesel engines are listed
in Table 3-9 (Ref.  3-16).  The  data show considerable engine-to-engine
variability. In all cases,  the highest smoke levels were obtained at
full load and intermediate speed.  Under these conditions the air-fuel
ratio is generally at its  minimum value.  The high, full-load smoke
levels of the two turbocharged open-chamber diesels at the inter-
mediate speed condition are attributed to operation near the stoichio-
metric air-fuel ratio.
                                 3-35

-------
TABLE 3-9.
AVERAGE STEADY-STATE SMOKE EMISSION
FROM DIESEL ENGINES (Ref. 3-16)

Engine

International D407
Perkins 4.236
Allis Chalmers 3500
John Deere 6404
Mercedes OM636
Onan DJBA
Caterpillar D6C
Detroit Diesel 6V-71

Chamber
Type

Open
Open
Open
Open
Divided
Divided
Divided
Open

Aspira-
tion

Natural
Natural
Turbo
Turbo
Natural
Natural
Turbo
Blower
Smoke Intensity in % Opacity at Condition
Low
Idle

1.2
1.0
0.5
2.0
1.0
0.5
2.3
0. 5
Load at Inter-
mediate Speed
0
1.0
1. 0
1.0
1.9
1. 5
0.8
2.0
0.8
Half
4.5
1.0
2.8
7.3
1.5
2.0
3.0
1.0
Full
20.0
10.3
31.5
25.5
8.5
2. 5
4. 5
1.0
Load at
Rate Speed
0
1.0
1. 7
1.2
2.2
2.0
1.0
3.2
1.0
Half
4.0
1.4
5.0
5. 5
1. 5
1.0
2.3
1.0
Full
9.2
7. 7
7.3
6.0
8.0
3. 0
2.7
1. 5
TABLE 3-10.  AVERAGE PARTICULATE EMISSIONS FROM
              DIESEL ENGINES (Ref.  3-16)

Engine

International D407
Perkins 4.236
Allis Chalmers 3500
John Deere 6404
Mercedes OM636
Onan DJBA
Caterpillar D6C
Detroit Diesel 6V-71

Chamber
Type

Open
Open
Open
Open
Divided
Divided
Divided
Open

Aspira-
tion

Natural
Natural
Turbo
Turbo
Natural
Natural
Turbo
Turbo
Particulate Concentration, mg/sfc
Low
Idle

4.5
1. 1
5.2
4.5
3. 1
6.4
2.4
0.9
Load at Inter-
mediate Speed
0
4.3
1.3
3.6
7.2
3.4
4.5
0.9
0.3
Half
8. 1
1.2
4.0
7.2
5.6
3.6
1.2
0.6
Full
20.0
11.6
17.4
22.4
9.4
5.2
1.6
0.6
Load at
Rate Speed
0
6. 1
9.8
2.6
3. 7
8.9
5. 5
0. 8
0.2
Half
7. 1
9.3
3.4
3. 8
7.9
8.6
2.0
0.6
Full
12.0
10.0
3.6
6.2
8. 7
8. 1
2.4
0.7
Standard cubic foot
                          3-36

-------
              Smoke (and particulate) emissions from diesel engines
can be minimized,  but not eliminated, by several means including
engine derating, proper fuel injection system design and maintenance,
and use of smoke suppressant fuel additives.  The latter two approaches
are further discussed in Sections 4.1.1.3 and 4.1.1.4.  Some reduction
in smoke has apparently been achieved by means of intake air heating
(Ref. 3-27).
3.1.4.4.2    Particulate s
              Particulate emission data from eight industrial diesel
engines tested by Southwest Research Institute are presented in
Table 3-10.  Comparison with the smoke data of Table 3-9 indicates
that the particulate levels of these engines correlate to  some degree
with visible smoke, particularly at high smoke levels.  However,  in
some cases a considerable amount of particulate matter was measured
under conditions where smoke was barely readable.  High smoke levels
were always associated with high particulate emissions  (Ref. 3-16).
Data from another  source indicate that a correlation might exist
between the opacity of black  smoke and the mass density of soot in the
exhaust gas (Ref. 3-28).  Tests conducted on a Mercedes Benz light-
duty 220D diesel engine indicate that the particulate emissions  of this
engine are of the order of 2 to 3 times higher than those obtained from
equivalent gasoline engines using leaded fuel and about 10 to 20 times
higher than for non-leaded fuel.  The material collected in these tests
was pitch black and very fine (Ref. 3-29).  Similar results were
reported by Daimler Benz (Ref.  3-30).
3.1.4.5       Odor
              Diesel exhaust odor has long  been recognized to be  a
very undesirable exhaust emission product.  Although a considerable
amount of research has been under way for  some time,  progress
toward determination of the cause of the odor has been very slow
because of the complexity of the heterogeneous combustion occurring
in diesels and the lack of sufficiently accurate instrumentation
                                 3.37

-------
(Ref. 3-10). The diesel odor is believed to be related to the mixing
and combustion process,  the chamber shape and, to a lesser degree,
fuel type and fuel composition.  Most likely, the odorants are formed
in the wall quench layer of the combustion chamber and in the hetero-
geneous pockets of unflammable fuel-air mixture.  (Ref. 3-31.)
               In general, diesel engine exhaust includes small
amounts of unburned fuel products of partial oxidation and pyrolysis
products (Ref.  3-31). Some of these compounds have strong odors,
even in very low concentrations, and mixtures of these  materials are
believed to be responsible for the formation of the typical diesel odor
(Ref. 3-2). Two major types of odor have been identified, one being
an "oily-kerosene" type apparently caused by  aromatic  hydrocarbons
and the other being of the  "smoky-burnt" type, and probably caused by
partially oxidized compounds.   Barnes (Ref. 3-32) has  concluded that
diesel odor is formed by partial oxidation reactions in ultra-lean
regions of the engine which are inevitably present in heterogeneous
combustion as  characteristic of diesel engines.
               In the absence of reliable instrumentation, odor has
been evaluated in the past by specially selected human panels trained
to recognize both odor quality and  intensity (Ref.  3-33 through 3-35).
The panel classifies the odor by comparing it  to twenty-eight different
odor qualities and intensities supplied from the bottles of the quality/
intensity (Q/I)  evaluation kit.  The overall  composite rating "D"
ranges from D-l to  D-12, with D-12 representing the strongest odor.
Burnt-smoky quality "B", aromatic quality "A",  oily quality "O", and
pungent quality "P",  have a range  between  1 and 4.
               More recently,  a diesel odor analysis instrument has
been developed by A. D. Little, Inc. on a Coordinating  Research
Council (CRC)  contract which is designed to permit direct measure-
ment of diesel  odor levels (Ref. 3-36).  Evaluation of the instrument
is in progress.
                                 3.38

-------
              Formulation of empirical correlations between odor
intensity and the concentrations of other pollutants has been attempted
by many investigators.  Although a fairly good correlation between
odor and total HC emissions has been found in some engines,  correla-
tions of this type have failed to hold for others (Ref. 3-12 and 3-37).
Attempts to correlate odor with CO and NOX emissions have failed
entirely (Ref. 3-37).  Tests  conducted by Vogh (Ref.  3-38) indicate
that light molecular weight aldehydes have  very  little effect on odor
formation in diesel engines,  and of the oxides  of sulfur and nitrogen,
only nitrogen dioxide represents a potential contributor to odor.
              The effect of engine operating condition on odor intensity
can vary markedly for different engines. In naturally aspirated engines
operated at rated speed, the minimum odor intensity seems to occur in
the mid-power range of the engine; whereas in turbocharged engines
the odor intensity remains essentially constant over the power range
(Ref.  3-12).  This trend is believed to be the result of a reduction in
the ignition delay time occurring  at higher  power levels due to an
increase in the intake manifold and compression temperatures.  Since
odorants are thought to be formed by partial oxidation of the fuel prior
to ignition, the constant odor intensity pattern would follow from this
effect (Ref. 3-12). Somewhat higher odor levels were reported by
Springer and Hare (Ref. 3-33) for a two-stroke engine-powered diesel
bus which was tested over a range of operating conditions.  The odor
was most severe during acceleration and idle.  This vehicle was also
tested with a catalytic converter installed in the exhaust which resulted
in some reduction in the odor intensity. This is further discussed in
Section 4.1.3.3.
              Tests conducted by Caterpillar indicate that the odor
level from prechamber diesels  can be  60 to 80 percent lower than for
open chamber engines (Ref.  3-25). Apparently, the odor notes from
both engine types are similar when the same fuel is used.
              Several approaches aimed at reducing diesel odor have
been considered by industry including injector modifications,  catalytic
                                 3-39

-------
mufflers and fuel additives.  Utilization of a needle valve injector with
a very small dribble volume  in place of a check valve type injector
has resulted in a substantial  reduction of the odor intensity arid alde-
hyde emissions from General Motors two-cycle GV-71 diesels (Ref. 3-35
and 3-39).  With the new injectors, the aldehyde emissions were
reduced by 50  to 85 percent over the operating range of the engine.
Since the  odor intensity varied  only slightly at these different operating
points,  it has been concluded by General Motors that aldehydes might
not be the  main contributors to diesel odor.
3.1.4.6       Noise
               Diesel engines are inherently noisier than gasoline
engines, particularly open chamber designs most frequently used in
heavy duty automotive and  stationary applications. The noise  is
especially pronounced at idle and during engine cold start.
               The principal forces contributing to diesel noise are
created by the combustion pressure transients, piston slap, and
timing gear impacts, with valve gear and fuel injection system impacts
being lesser sources (Ref. 3-40 and 3-41).  The major noise emitting
engine surfaces include the valve covers, intake manifold,  gear cover,
and oil pan (Ref. 3-26).
               Engine noise is a systems problem and reduction can
be attempted either at the source or by reducing the vibration  levels
of the external surfaces of the engine (Ref.  3-40).  Minimization of
the combustion pressure rise rate results in substantial noise abate-
ment, as evidenced by the favorable noise characteristics of pre-
chamber and swirl-chamber diesels.  The noise contributed by the
piston and timing gear  slap might be diminished to some degree by
reducing certain component clearances within the constraints dictated
by the design  and manufacturing techniques (Ref.  3-41).
               Further reduction in engine noise might be achieved  by
means of structural modifications on the engine.  These include:
(1) higher stiffness of the basic engine structure to reduce vibration,
                                 3-40

-------
(2) isolation of non-structural members, such as covers and
accessories, (3) damping of certain engine components to modify their
vibration characteristics and (4) application of acoustic enclosures
around the engine (Ref.  3-42).  Vibration isolation and stiffening have
resulted in a reduction of the diesel noise  by about 5 dB(A) while total
enclosure of the engine has reduced the noise level by as much as
20 dB(A) (Ref.  3-43).  Also, sheetmetal shields fitted with fiber glass
liners have been used  to cut diesel noise by as much as 4 dB(A)
(Ref.  3-26).
3.1.5         Fuel Consumption
              Typical performance curves for naturally-aspirated
and turbocharged, open-chamber diesel engines are presented in
Figure 3-14, showing engine torque, brake horsepower and specific
fuel consumption as a function of operating speed (Ref.  3-44).  The
data are  from a large stationary engine but the indicated performance
trends apply to smaller diesels and divided chamber configurations as
well.  As indicated in the figure, the specific fuel consumption of
naturally-aspirated diesels has a minimum in the mid-power/mid-speed
regime.  Conversely, turbocharged engines operate most economically
at or near full load, and fuel economy generally worsens with decreasing
speed.  Comparison of the data indicates that the specific fuel consump-
tion of turbocharged diesels is slightly lower (up to 5 percent) than for
naturally aspirated designs, particularly at rated conditions.
              In general,  the  specific  fuel consumption of divided-
chamber diesel engines is several  percent higher than that of equivalent
open chamber designs.   This difference is attributed to the higher sur-
face to volume ratio of the divided  chamber concept and the higher heat
losses through the chamber wall.   One manufacturer has stated that in
new engines there is  a 5 to 10 percent difference in specific fuel con-
sumption between the two engine types. However,  in the field the
difference is generally less because injection  system fouling tends to
occur less frequently in divided-chamber  engines compared to some
                                3-41

-------
            MODEL L57920S ENGINE ONLY
                TURBOCHARGED
     7000
     6000
     5000
     4000
        - CORRECTED TO 29.92 in. Hg, 60°F -
                                        X 5000
                                        2 4500
                                        B 4000
                                  MODEL L572D ENGINE ONLY
                                   NATURALLY ASPIRATED
                          B 3500
                          >- 3000
                                  1600
                                  1400 °>
                                  1200 £
1000 I
800 8
600
400
200
                                    m
                              - CORRECTED TO 29. 92 In. Hg, 60°F A
1100
1000
900 a,
800 >
700 n
600 g
   (/>
500 n
400 g
              600
                    800
                     rpm
                          1000
                                1200
                                                  600
                                          800
                                          rpm
                                                              1000
                                                                    1200
   CCO
      0.50
      0.45
      0.40
   ?« 0.35
1200 rpm-7
        - 800 rpm
                      1000 rpm
      u.iO.44

      §|j'j
      
-------
engines of Table  3-2 indicates slightly higher fuel consumption for the
two-stroke designs.  Conversely, the large two-stroke turbocharged
diesels listed in Table 3-7 have a slightly lower specific fuel consump-
tion than shown in Table 3-3 for one large four-stroke,  turbocharged
diesel engine.
3. 2            SPARK-IGNITION ENGINE CHARACTERISTICS
3. 2. 1          Engine Description
               The spark-ignition internal combustion engine is the
most-used powerplant in the world today, and many different designs
have been developed since this engine type was invented almost a cen-
tury ago.  In size, these engines range from small single-cylinder
units  producing less than one horsepower to large multicylinder units
with power output ratings of several thousand horsepower.  (Units
above several hundred horsepower are predominantly used in station-
ary power applications. )
               One of the basic advantages of reciprocating internal-
combustion engines is the  high maximum thermodynamic  cycle tempera-
ture resulting in high thermal efficiency.  Relative to current diesel
engines, spark-ignition engines operate at lower peak pressures.
Thus, the structural stresses are lower, permitting higher power-to-
weight and power-to-volume ratios  for spark-ignition engines.
               Conventional automotive spark-ignition engines  generally
operate on gasoline and in certain applications  on gaseous fuels such as
liquified petroleum gas (L.PG).  Conversely,  many of the  stationary
engines are gas fueled using natural gas, waste gas,  water gas,  and
methane.   Based on the type of work cycle employed, these engines
can be divided into two- and four-stroke designs.  These  engine classes
can be subdivided into different categories depending upon the type of
fuel supply system (carburetion, fuel  manifold injection,  direct cylin-
der injection) and air supply system,  (supercharged, turbocharged,
naturally aspirated, and blower-scavenged)  utilized.
                                3-43

-------
               More recently, the introduction of rotary engines
(Wankel engine) has added another type of spark-ignition engine
presently available  for use in automotive applications.  To date,  this
engine type has not  been introduced into the stationary engine market.
               Since the early 1960s, several organizations have been
involved in the development of stratified charge spark-ignition engines,
primarily  because of their low exhaust emission, specific fuel con-
sumption,  and multifuel capabilities.  The stratified charge engines
can be divided into two categories:  open-chamber (or direct-injection
engines) and divided-chamber (or prechamber engines).  These engine
configurations are discussed in Section 4. Z. 2. 4.
               In the selection of a stationary spark-ignition engine,
the emphasis is on investment cost,  economy of operation,  reliability,
and durability.  Most of the stationary engines operate at semi-
constant speed and load conditions.  This permits incorporation of
many design simplifications (simple carburetion, fuel injection,  spark
timing, and related  control systems),  as well as optimization of
important  engine operating and design parameters (valve timing and
porting, manifolding, etc. ).  Generally,  little or no weight  or space
limitations are imposed on stationary engines,  resulting in  greater
freedom in the selection of materials and in the design of certain
engine accessories  (e.g., cooling system).
               In regard to exhaust emission control,  stationary
engines have distinct advantages  over mobile  engines.  For instance,
because of the semisteady state  operation of these engines,  many
design and operating parameters could  be optimized for exhaust emis-
sions.   Furthermore, because of the freedom from space and weight
limitations,  various  exhaust gas  purification systems could be  applied,
which would be impractical for  use in mobile  installations.
                                3-44

-------
3.2.2         Applications
              Stationary spark-ignition,  internal-combustion engines
are used in a great variety of industrial,  municipal,  and urban applica-
tions.  Small gasoline engines (1 to 10 HP) are utilized to drive domes-
tic, agricultural, and commercial  power tools and equipment — such as
power  saws,  lawn mowers, and portable  compressor, pump, and elec-
tric generator units.  Medium-size gasoline engines (50 to 200 HP) are
mostly used for commercial and construction site compressors, pumps,
blowers, and electric power generator  units. Medium-large spark-
ignition engines (200  to 1000 HP) are generally operated on gaseous
fuels and are used for heavy-duty,  medium-speed applications.  Most
of them are of the four-stroke,  naturally • aspirated type and are used
to power gas compressors or stand-by  power generators.  Large
spark-ignition engines (1000 HP and up) are  always operated on gase-
ous fuels and are both four- and two-stroke, low-speed  (300 to
400 rpm) engines.  The application of these engines includes com-
pressor drives, gas-well recompression (in transmission lines), gas
plant compressors, refinery process compressors,  water pumping,
sewage pumping, and electric power generator drives for continuous
operation.  Table 3-11 presents a breakdown of the stationary spark-
ignition engines installed during the 1963-1970 time period (Ref. 3-9).
The total number of gasoline and gas-operated spark-ignition engines
in use  at the  end of 1970 outnumbered the diesel and gas-turbine
engines by almost two orders of magnitude.
3.2.2.1       Installed Power
              Table  3-12 presents the  breakdown of the estimated
installed horsepower  of spark-ignition gas engines at the end of 1971
(Ref.  3-9).  Because  of the great diversity of gasoline engine applica-
tions and sales to distributors rather than directly to the customers,
there is insufficient information available at this time to provide a
                                3-45

-------
         TABLE 3-11.  INTERNAL COMBUSTION ENGINES- NUMBER VS END USE (Ref.  3-9)
Year
Gasoline
Marine
Lawn & Garden i
Chain Saws I
Agriculture
Subtotal
Construction
Generator Sets
General Industrial
Subtotal
Total
Gaseous Fuels
Agriculture
Construction \
Generator Sets 1
General Industrial
Total
1970

61,663
8,013.961
169,977
8,245,601
152, 178
86,264
1,073. 564
1,312,006
9,557,607

2,987a
2,342a
1, 821
7, 150
1969

106,693
8,717, 864
187,437
9,011.994
175,605
90,760
1,249. 185
1,515,550
10, 527, 544

3,257
2,694
1,002
6.953
1968

84,624
8,236,693
193,380
8,514,697
104,638
67,798
1. 134,638
1,307,074
9,821,771

3,947
3, 547
1,028
8,522
1967

103,478
7,555.701
202. 167
7, 861,346
121,225
67,930
1,070, 887
1,260,042
9, 121.388

6, 873
4,799
1,260
12,932
1966

103,899
6,422,221
482, 194
509.543
7, 517, 857
132. Z14
76,678
1, 173.939
1,382.831
8.900,488

11.460
5, 539
1. 135
18. 134
1965

39, 937
5,766, 819
442,855
417, 507
6,667, 118
85,076
67,769
1,087.760
1,240.605
7.907,723

5,654
8,266
839
14, 759
1964

29.463
4.760, 683
435, 152
434, 175
5,659,473
66.052
59. 190
949.007
1,074.249
6.733, 722

5. 780
7,700
911
14,391
1963

28.005
5,084.262
373.263
430,362
5,915, 892
51. 136
43,542
851,068
945, 746
6, 825, 638

8. 788
5, 142
562
14,492
1963-70
Cumulative Total

557,762
56,291,668
2,544, 548
59, 393, 978
888, 124
559,931
8, 590,048
10,038, 103
69,432,081

48,746
40,029
8,558
97,333
Estimates based on distribution of gas engines in Agriculture and Construction/Generator Sets for previous three years
OJ
I

-------
     TABLE 3-12.   ESTIMATED INSTALLED HORSEPOWER OF
                    SPARK-IGNITION GAS ENGINES FOR 1971a
             Application
Installed Horsepower
     Electric Power Generation
     Oil and Gas Pipelines
     Natural Gas Processing Plants
     Oil and Gas Exploration
     Crude Oil Production
     Natural Gas Production
     Industrial Process
     Municipal Water and Sewage

              Total
         90,000'
     10, 990,000
      2,410, 000
        500,000
        852,000
      3,237,000
        230,000
        465,000

     18,774, 000
    A summary table listing the installed horsepower of all stationary
    engines is presented in Appendix A
   'Estimated 1970 data
breakdown of installed gasoline engine horsepower.  The total installed
horsepower of small gasoline engines (1 to 10 bhp) is estimated to be
about 180 x 10  hp and of the medium size gasoline engines (25 to
250 bhp) to about 600 x  10  hp.  This brings the grand total of all sta-
tionary spark-ignition engine (including gas  engines) installed horse-
power to about 800 million.
3.2.2.2       Operating Modes
              The duty cycles of large, low-speed spark-ignition
engines, driving electric power generators,  industrial, agricultural
and municipal water and sewage pumps, compressors, etc., are simi-
lar to the operating modes of diesels which are discussed in  Sec-
tion 3.1.3.2.  The duty cycle and the application of small- and
                                3-47

-------
medium-size gasoline spark-ignition engines varies widely.  Unlike
light-duty automotive gasoline engines, the majority of the small- and
medium-size stationary gasoline engines  operate at constant  or nearly
constant speed, with power output controlled by a governor.
3. 2. 2. 3       Fuel Requirements
               Stationary spark ignition engines can be operated on
various fuels including leaded,  low-lead,  and nonleaded gasoline;
white gas; gaseous fuels and various blends of other hydrocarbone
fuels.   In most stationary engines the fuel octane requirement is quite
low.
               Commercially available gasoline blends have a low  sul-
fur content  (0. 025 to 0. 048 percent by weight) and contain  insignificant
amounts of  other contaminants.  However, gum and carbon deposits
in the induction system, combustion chamber,  and crankcase have a
tendency to accumulate, particularly for those  duty cycles which
involve frequent engine starts and extended idle periods.
               In general, gaseous fuels cause  less deposit build-up in
the combustion chamber — which is attributed to lean air-fuel ratio
operation.  This inherent advantage  is reflected in better brake spe-
cific fuel  consumption and lower hydrocarbon and carbon monoxide
emissions.
3.2.3          Emissions
               Over the  past two decades,  extensive research has  been
conducted in the field of spark-ignition engine emissions.  In  general,
the research activities  have been concerned with automotive engines;
however,  many of the findings and conclusions  are applicable to sta-
tionary spark-ignition engines as well.
                                3-48

-------
3. 2. 3. 1       General
              Uncontrolled spark ignition engines emit air pollutants
from  four sources:  engine exhaust, crankcase blowby, carburetor,
and fuel tank. According to a  survey conducted on a large number of
uncontrolled automotive spark-ignition gasoline engines,  (Ref. 3-45),
the engine exhaust contributes  100 percent of the carbon monoxide  (CO)
and oxides of nitrogen (NO) emissions and about 65 percent of the total
unburned hydrocarbons (HC).   Another 25 percent of the HC is attrib-
uted to crankcase blowby,  with an additional five percent each  resulting
from  evaporation of gasoline from the carburetor and from the fuel
tank.   Furthermore, particulates constituting approximately 5 percent
by weight of the unburned hydrocarbons  are emitted from the engine
exhaust (Ref.  3-46).  These consist of lead compounds, carbon  particles,
motor oil,  and nonvolatile reaction products formed during the com-
bustion process  in the engine cylinder.  There are  literally hundreds
of reaction products emitted — most of them in trace quantities —
including organic acids, high molecular weight olefins, carbonyl com-
pounds, and sulfur  compounds  whose concentration depends on the  sul-
fur content of the gasoline.
              Statistical data  on the emissions from spark-ignition
engines operated on gaseous fuels (e.g. , natural gas)  are currently
not available. However,  in this case the evaporative emissions are
essentially zero  since the fuel  system is not vented.
3.2.3.2       Principles of Engine Combustion-Generated Emissions
              The  combustion process in spark-ignition engines
involves complex chemical reactions, whose  "kinetics" depend on the
chemical composition and the thermodynamic  state of the reactants
(fuel and air) and on the parameters of the reaction vessel (e. g. , the
engine combustion chamber) which affect the  combustion process.
Since  the nitrogen contained in the combustion air is heated to high
                                3-49

-------
temperatures during the combustion process,  some of it reacts with
oxygen to produce nitric oxide (Ref. 3-47).  As the combustion gases
are rapidly cooled during their expansion and exhaust, the nitric oxide
remains in a state of "frozen equilibrium", i.e. ,  it does not dissociate
to nitrogen and oxygen as predicted from chemical equilibrium con-
siderations (Ref.  3-48).  In gasoline engines combustion occurs fre-
quently in rich fuel-air mixture zones resulting in incomplete oxidation
of the fuel and production of significant quantities of CO (Ref. 3-49).
This  is further compounded by flame-quenching in the wall region of
the combustion chamber.  As a result, part of the fuel in the wall
boundary layer remains uncombusted and is then expelled during the
exhaust stroke (Ref.  3-50).
               The exhaust emissions  from spark-ignition engines are
primarily affected by the air-fuel ratio of the combustible mixture.
At air-fuel ratios below stoichiometric, little NO is produced because
of a lack of available oxygen.  However, under these conditions,  sub-
stantial quantities of CO are formed and a portion of the excess fuel is
converted to organic compounds, collectively called unburned
hydrocarbons.
               The concentration of nitric  oxide reaches its peak at air-
fuel ratios slightly leaner than stoichiometric.  Further leaning of the
mixture results in decreasing concentration of NO, because of the
attendant decrease in the peak combustion temperature (Ref. 3-51).
CO reaches the minimum concentration at air-fuel ratios  above about
15.5 and HC continues  to decrease with leaning of the mixture until
misfire sets in.  At that point,  the HC emissions increase again.
               Another variable which has a significant effect on the
concentration of the exhaust pollutants is the combustion temperature.
The theoretical concentration of nitric  oxide is an exponential function
of the  combustion temperature.   Conversely, the CO concentration is
                                3-50

-------
determined primarily by the air-fuel  ratio and is only slightly affected
by the combustion temperature (Ref.  3-52).  The HC emissions are
affected  by the combustion temperature only insofar as lowering the
combution temperature can extend the burning  time of the mixture
resulting in higher exhaust temperatures and more complete oxidation
of the HC species in the exhaust manifold (Ref.  3-53).
              In general, all spark-ignition engine parameters affect
the emissions inasmuch as they affect either the air-fuel ratio or the
combustion temperature, or both (a detailed discussion of these aspects
is presented in Section 4.2. 1).  In addition,  test data indicate that com-
bustion chamber wall effects and deposits might have some effect on
the emissions of NO  and HC (Ref. 3-54 through 3-57).
                   X
              When a spark-ignition  engine  is operated on gasoline at
substantially leaner than stoichiometric air-fuel ratio, a distinct alde-
hyde odor in the exhaust is sometimes noticeable.  This may result
from  partial oxidation of fuel hydrocarbons in ultra-lean zones  in the
combustion chamber.  As mentioned in Section 3. 2. 3. 1,  particulates,
consisting of lead compounds and carbon particles, are emitted from
the exhaust of spark-ignition engines  operated  on commercial gasolines.
However, with gaseous fuels (e.g., natural gas), the particulates  are
virtually eliminated and the exhaust odor is substantially improved
(Ref.  3-58).
3. 2. 3. 3       Test Procedures and Instrumentation
               The test procedures and instrumentation for determina-
tion of exhaust emissions from mobile and stationary internal combus-
tion engines are described in detail in numerous publications (Ref.  3-59
through 3-68).  Table 3-13 presents an overview of the principles,
methods,  and instruments generally used for the detection and analysis
of the various exhaust pollutants of spark-ignition engines.
                                3-51

-------
                 TABLE 3-13.  EXHAUST GAS ANALYSIS METHODS AND INSTRUMENTS
Specie
Detected
NO, NO2
NO, NO2
NO
NO
HC
HC
HC
CO
co2
Aldehydes
Particulates
Method
Phenoldisul-
fonic (PDS)
Saltzman
Nondispersive
Infrared (NDIR)
C hemilumine s -
cence (CL)
Nondispersive
Infrared (NDIR)
Flame lonization
Detector (FID)
Gaschromato-
graphic (GC)
Nondispersive
Infrared (NDIR)
Nondispersive
Infrared (NDIR)
Gaschromato-
graphic (GC)
Scanning electron
microscope (SEM)
Type of Method
Wet chemical
Wet chemical
Optical
C hemilumine s cence
Optical
Flame lonization
Chromato graphic
Optical
Optical
Chemical and
chromato graphic
Electron microscope
X-Ray spectrometer
Remarks
Grab samples
Analysis time 24 hr
Grab samples
Analysis time 1 hr
Continuous analysis
Sensitized to NO
Continuous analysis,
response 2 sec,
high sensitivity
Continuous analysis
Sensitized to n-hexane
Continuous analysis
Total HC detection
Grab samples
Individual HC
components
Continuous analysis
Sensitized to CO
Continuous analysis
Sensitized to CO2
Grab samples,
quantitative and
qualitative analysis
Grab samples
Resolution 200A
Ref.
3-59
3-60
3-68
3-61
3-62
3-63
3-64
3-62
3-62
3-65
3-66
CO
I

-------
              As the methods and instruments are perfected and the
understanding of the importance of particular pollutants in atmospheric
photochemistry is improved,  the test procedures are subject to pro-
gressive modifications or improvements.  The officially adopted test
procedures and the approved  instruments for exhaust analysis are
described in the Federal Register  (Ref. 3-68).  Moreover,  the test
procedures for  diesel engines,  presented in Section 3. 1.4.2, are appli-
cable to stationary spark-ignition engines as well.
3.2.3.4      Gaseous Emissions
               The characteristics of exhaust emissions of NO  , HC,
                                                            x     '
and CO from spark-ignition engines are strongly influenced by the type
of fuel utilized.  The influence of liquid fuels, such as the commercial
gasolines, is distinctly reflected in the composition and photochemical
reactivity of the exhaust hydrocarbons (see Section 4.2.1). The gaseous
fuels,  on the other hand,  are generally combusted with excess air, and
as a result the HC emissions (and their reactivity) are  quite low.  The
exhaust emissions are also adversely affected by fuel maldistribution
in the individual engine cylinders particularly in the case of liquid fuels.
               As previously noted, most (gasoline) engines are oper-
ated at rich or near-stoichiometric air-fuel ratios,  resulting in high
CO,  HC,  and NO  emissions.
   '     '        x
               In the case of gaseous fuel (e. g. , natural gas) the
engine is generally operated with excess air  resulting in relatively low
specific  mass emission of CO and HC.  These inherent differences in
the operating conditions  of liquid- and gas-fueled spark-ignition
engines are primarily responsible for the observed difference in the
emissions of these two classes.
                                3-53

-------
3.2.3.4. 1     Emissions at Rated Conditions
               The emissions from eight heavy-duty gasoline engines
and from seven medium large and large gas engines are presented in
Tables 3-14 and 3-15,  respectively.  The emissions of the gasoline
engines are presented for steady-state operating conditions (estimated
rated conditions for stationary applications) and in terms of composite
values over the 23-mode test cycle (Ref. 3-69).  The emissions from
large gas engines are presented for five engines at rated operating con-
ditions and for two engines as the composite  values of the Diesel Engine
Manufacturers Association (DEMA) test cycle.
               From comparison of the emission data, it is apparent
that the HC and CO specific mass emissions  from gasoline engines are
an order of magnitude higher than those of gas engines.  Conversely,
the NO  emissions are consistently lower.  The variation of the emis-
sions among the gasoline engines appears to  be lower than the variation
among the gas engines.  As indicated, the turbocharged  engines show
consistently higher NO  emissions.
                      A.
               The emission levels of the  heavy-duty gasoline engines
are consistent with the emission levels of pre-controlled automotive
engines.  However,  Tables 3-14 and 3-15 present only a small data
sample and,  depending on the age and mechanical  condition of the
engines, greater variations in the emission levels are expected to
exist among the stationary  engines currently in use.
3.2.3.4.2     Part Load Emissions
               Figure 3-15 (from Ref. 3-70)  shows the exhaust concen-
tration of CO,  HC, and NO from an automotive spark-ignition gasoline
engine in the form of engine performance maps.  At certain part-load
conditions,  minimum concentration of CO, HC, and NO results in the
exhaust gas; as expected,  the minimum  concentration of all three
pollutants occurs at distinctly different part-load operating conditions
of the engine.
                                3-54

-------
          TABLE 3-14.  EMISSIONS FROM FOUR-STROKE,  NATURALLY ASPIRATED
                         SPARK IGNITION HEAVY DUTY GASOLINE ENGINES


Type and
Maximum
bhp/rpm


Straight 6,
132/3500
Straight 6,
110/3800
V-8,
217/3200
V-8.
230/4000
V-8,
152/3600
V-8,
175/4000
V-8,
150/4000
V-8,
158/4000


Identi-
fication
No.


HD 1-1

HD 2-1

HD 1-2

HD 2-2

HD 1-3

HD 2-3

22

23



CID




-

.

-

_

-

318

345



No. of
Cylin-
ders


6

6

8

8

8

8

8

8


Continuous Duty
Rated
Powe r ,
bhp

69

58

122

125

82

95

73. 8

76. 7

Rated
Speed,
rpm

2300

2300

2300

2300

2300

2300

2300

2300

Specific Mass
Emissions ,
g/bhp-hr

HC
0. 37

1. 74

1.29

4.02

0. 63

2.76

2. 36

1.97


CO
7.6

25. 1

10. 5

49. 5

7. 8

35. 5

20.4

28.66


N0x
15. 8

13. 1

13. 3

8. 7

13. 1

10.6

11. 87

11.45

BSFC,a
Ib
bhp-hr

0. 531

0. 548

0. 509

0. 501

0. 578

0. 503

0. 498

0. 496

Air
Fuel
Ratio

16.0

14. 7

15.6

14.7

15.2

14.4

14.8

14. 3

b
Composite Values
Specific Mass
Emiss ions ,
g/bhp- hr

HC
9. 20

10. 05

12. 85

12.61

13. 10

9. 06

3.0

2. 41


CO
46. 2

47. 8

46. 8

89. 8

35. 8

30. 4

45. 2

31.0


NO
X
13. 9

11.9

12. 8

7.9

11.4

10. 7

10.3

9. 38

BSFC,a
Ib

bhp-hr
0. 677

0. 750

0.675

0.756

0.730

0.631

0. 577

0.643



Ref-
er-
ence


3-69

3-69

3-69

3-69

3-69

3-69

3-71

3-71

Brake Specific Fuel Consumption
23-mode test cycle
CO
I
Ul
Wl

-------
          TABLE 3-15.  EMISSIONS FROM FOUR- AND TWO-STROKE,  ASPIRATED AND
                        TURBOCHARGED SPARK-IGNITION GAS ENGINES

Engine Type


Four-Stroke,
Aspirated
Four-Stroke,
Turbocharged
Two-Stroke,
Turbocharged
Two-Stroke,
Aspirated
Four-Stroke,
Aspirated
Four-Stroke,
Aspirated
Four-Stroke,
As pi rated

Engine
Identi-
fication
No.


.



GMVH-
-8
GMVA-
-8


_

_


C1D


9896

6597

.



.

_

_


No. of
Cylin-
ders


12

8

8

8

.

_

_

Continuous Duty
Rated
Power,
bhp

1200

1 100

1600

2000

.

_

_

Rated
Speed,
rpm

900

900

330

250

_

_

_

Specific Mass
Emissions,
g/bhp-hr

HC
_



1. 58

1. 9

0. 25

1.25

0. 37


CO
_



0. 17

0. 32

0. 2

0. 6

0. 2


NOX
_

.

20. 1

15. 0

32.0

18. 0

26. 5

BSFC,3
Btu
bhp-hr

8400

7600

6632

7127

8000

7400

8100

Air
Fuel
Ratio

_

.



24. 8

21.5

_

21.0

Composite Values
Specific Mass
Emissions,
g/bhp-hr

HC
1. 32

7. 37





_

_

.


CO
5. 84

0. 65

.



_

_

_


NOX
14. 2

19.7

.



_

_

-

aBrake Specific Fuel Consumption
Composite emission values by DEMA test cycle
UJ
I
Ul

-------
           NITRIC OXIDE
                            18.1  MANIFOI.O
                                  PRESSURE
                            15.7
                                   xlO1
         CARBON MONOXIDE
                            70.5
                             ""V MANIFOLD
                                 PRESSURE
                                  XlO1
                                                       HYDROCARBON N-HEXANE
                                                            ENGINE SPEED
                 SPEED
Figure  3-15.   Emission map -  automotive spark-ignition engine
                 (Ref.  3-70)
                                    3-57

-------
              Figure 3-16 presents the relative specific mass emis-
sions of NO , HC,  and CO, specific fuel consumption, and air-fuel
           A.
ratio as a function  of the percent maximum power  of a heavy-duty gaso-
line engine (Ref. 3-71).  At 2300 rpm, the specific emission of NO
                                                                X
increases as the engine load is  decreased.  This is partly due to the
shift of air-fuel ratio setting from rich, at full load, to slightly lean
in the mid-range of engine loads.   At  low  load, the NO  mass emission
                                                    j"w
trend is reversed and NO   reaches a minimum at about 10 percent of
maximum power, mainly due to mixture enrichment.  The sharp
increase in NO  near idle  is attributed to  the low power output level of
the engine.  The CO mass emissions appear to be  directly related to
the air-fuel ratio distribution.  The high HC emission at about 30 per-
cent of  maximum power may be due to fuel maldistribution effects.
              For  comparison, test data  obtained from that engine at
reduced engine speed (1200 rpm) are plotted in broken lines in Fig-
ure 3-16.   Again, the emissions are referenced to the full load values
at 2300 rpm.  At 1200 rpm and  45  percent of maximum load, the spe-
cific NO  emissions are about 50  percent  lower than at the rated condi-
tions of the engine.  However, this improvement is accompanied by a
five-per cent loss in specific fuel consumption.
              Figure 3-17 (Ref. 3-9) shows the effect of engine load,
expressed in terms of brake-mean-effective-pressure (BMEP) on the
specific mass emissions of a large spark  ignition gas engine. As indi-
cated, NO decreases rapidly as BMEP is reduced from the base
          j£
condition while HC  increases.   As expected,  CO is essentially inde-
pendent of engine load.
3.2.3.5      Other  Polutants
              In addition  to  the  gaseous  pollutants discussed  above,
gasoline engines emit a large number of other reaction products, some
in trace quantities, such as organic acids, high molecular weight olefins,
                                3-58

-------
                      318 CIO, V-8 ENGINE
                      ISO Mm -4000rpm
                      —— 2300 rpm
                         (A) —f02.3 bhp
                      	1200 rpm
                         |B) —51.4 bhp
                         A |B) - WOT
                            b POWER
           0.2     0.4    0.6    O.B
         PERCENTAGE OF MAXIMUM POWER
                                   1.0
Figure 3-16.
Part load  emissions
of a heavy-duty
spark- ignition
engine
                                           ,J 600
                                           tx

                                           < 500
                                           a
                                           UJ
                                           | 400
                                           UJ

                                             300


                                            .?900




                                          = 2 5°°
                                            oc
                                            "" 300

                                           i 8000
                                           (D
                                           > 7500
                                           do

                                           y 7000
                                           a


                                              25
                                           .c
                                           I  20
                                           CD

                                           I  '5
                                           o>
                                           „•;  10

                                           o
                                           s  5
                                                                   FUEL CONSUMPTION


                                                                             I MASS EMISSIONS
                                               60
                                                          80    90
                                                          TORQUE BMEP
Figure 3-17.  Effect of torque
                on engine per-
                formance -
                large two-stroke
                spark-gas
                engine,  300 rpm
                (Ref.  3-9)

-------
and carbonyl  compounds.   some of these compounds are  believed
to be  carcinogenic while others contribute to engine exhaust odor.  The
sulfur contained in the fuel is emitted in the form of sulfur oxides and
sulfates.  Lead added to gasoline as an anti-knock compound is emitted
as lead oxide.  Under fuel-rich operating conditions,  considerable
quantities of carbon particles may be exhausted.  In some engines,
small amounts of lubricating oil entering the combustion chamber are
partially oxidized  during the combustion process and are exhausted as
a bluish smoke.
              In gas engines lead is nonexistent,  carbon particles are
virtually eliminated, and odor  is greatly reduced.
3.2.3.6      Average Emission Values
              In the previous  sections the emissions  at rated conditions
and at part loads were listed for several heavy duty gasoline engines
and large stationary gas engines. The average emissions for each
group are presented in  Table 3-16.  These values are representative
  TABLE 3-16.   AVERAGE SPARK-IGNITION ENGINE EMISSIONS
                AT RATED CONDITIONS

Spark Ignition
Engine Type


Heavy Duty
Gasoline
Medium-large
Gas
Large Gas
Specific Mass
Emissions
g/Bhp-hr

HC

1.9

4.3
1. 1

CO

23. 1

3.2
0.3

N0x

12.2

17. 5
22.3

BSFC
lb/
bhp-hr

0.52

—
—
Btu/
bhp-hr

—

8000
7450

A/F



14.9

—
22.4
                                3-60

-------
of engines in new or well-maintained  condition.  The statistical
emission data obtained on a large number of automotive engines indi-
cates that the emissions from spark ignition engines increase substan-
tially as  the mechanical condition of the  engine deteriorates  due to
wear in extended use or lack  of proper maintenance.
3.3           GAS TURBINES
3.3.1         Engine Description
              The stationary gas turbine derived its  application from
aircraft jet engines; and the first gas turbines installed for commercial
use were initially derated aircraft gas turbines.  Since 1965, gas tur-
bine use for  stationary application has greatly expanded as the confi-
dence in their operation grew and as larger units designed specifically
for stationary applications  became available.  In its simplest form, the
gas turbine engine consists of a compressor (usually multistage, axial)
in which the  air is compressed to 100 to 200 psig, followed by a com-
bustor in which  fuel (liquid or gaseous) is added for combustion which
occurs at overall  lean mixtures in the range of fuel-air ratio of  0. 01 to
0. 02.  The combustion gases are then expanded through a  turbine or
turbines to near atmospheric pressure, converting their energy into
power.   The design and performance  of the combustor  is critical not
only for the combustion efficiency but also  for the level of emissions
generated there.  This will be discussed in Section 4. 3. 3.  The magni-
tude and uniformity of the gas temperature at the turbine inlet are also
important parameters which, along with the compression ratio,  deter-
mine the engine efficiency.  The efficiency and the power output per
pound of air  increase with increasing turbine inlet temperature, but  a
limitation is imposed by the material and stresses in the turbine blades.
Typical turbine  inlet  temperatures for today's stationary gas turbines
are of the order of 1800 to  1900°F.   The turbine,  which can be  single
or multi-stage,  drives the  compressor.  In a single-shaft engine,
                                3-61

-------
excess power is generated by the turbine to drive an electric generator,
pump, or any other power absorbing machine.  Conversely in two-shaft
engines, a second separate turbine is provided downstream of the com-
pressor turbine for power extraction.
3.3. 1. 1
Simple Cycle
              A schematic of a simple cycle gas turbine is presented
in Figure 3-18. The typical simple cycle efficiency  is 25 to 30 percent.
Gas turbines used in today's modern electric utilities typically operate
at heat rates (heat input to output power level ratios) of 11, 000 to
12, 000 Btu/kwh, corresponding to thermal efficiencies of 3 1 and 28 per-
cent,  respectively.  However,  there are simple cycle gas turbines in
operation now which exceed these values, reaching thermal efficiencies
of 33 to 34 percent (Ref. 3-72).
3.3. 1.2
Regenerative Cycle
              The efficiency of simple cycle gas turbines can be
upgraded by further extraction of heat from the exhaust gases which
normally leave the turbine at 800 to 1100°F.  One of the ways of doing
so is with a regenerative cycle as shown in Figure 3-19 (Ref. 3-73).
       AIR IN
                     FUEL
                           EXHAUST
                       ^COMBUSTOR
          COM-
          PRESSOR
                       TURBINE
                  THERMAL  EFFICIENCY, 25 TO 30%
               Figure 3-18.  Simple cycle gas turbine
                             (Ref.  3-73)
                                3-62

-------
                                        EXHAUST
        AIR IN
                                              REGENERATOR
                  THERMAL EFFICIENCY, 35 TO 37%
            Figure 3-19.  Regenerative cycle gas turbine
                          (Ref.  3-73)
Air leaving the compressor is fed to the combustor via a heat exchanger
in which some of the heat of the  exhaust  gases  leaving the turbine is
transferred  to the pressurized air so that at fixed turbine inlet tempera-
ture less fuel  has to  be  added.   This  improves the thermal efficiency
to about 34  to 38 percent, corresponding  to heat rates of 10,000  to
9,000 Btu/kwh.
3.3.1.3
Combined Cycle
              Another way of utilizing the exhaust waste heat is in a
combined cycle  shown schematically in Figure 3-20 (Ref.  3-73).  In
this case, the exhaust gases are used  to produce steam in a steam
generator which can then be used to drive a steam turbine to produce
additional power.  The usual power generation split is:  60 percent gas
turbine;  40 percent steam turbine.  In some cases, for higher quality
steam, an afterheating device is added by burning additional fuel in the
exhaust (fired versus unfired cycle).  The combined cycle has the
                                 3-63

-------
                              EXHAUST GAS STEAM TURBINE
                HEAT RECOVERY
                STEAM GENERATOR
           AIR
          COM-
          PRESSOR
                                        BOILER
                                        FEED   CONDENSER
                 GAS TURBINE
                    THERMAL EFFICIENCY, 40 TO 42%
         Figure 3-20.  Combined cycle gas turbine and steam
                       generator (STAG) system (Ref. 3-73)
highest thermal efficiency, with some of the existing units operating at
40 to 42 percent, corresponding to 8530 to 8100 Btu/kwh heat rates.
3.3.1.4       Future Design Trends
               All types of gas turbine  cycles discussed above are in
operation.   Of course,  the simple cycle has the longer operating exper-
ience behind it while  the combined cycles started operation only
recently.  Before discussing the various applications,  a few comments
are presented on the  future design philosophy of stationary gas turbines.
               The future stationary gas turbines will be subjected to a
number of controlling parameters such as: a) rules regarding  exhaust
emissions,  b) requirement to operate on residual oils and possibly
coal, c) increase in plant size for intermediate load electric power
generation,  and d) higher performance  to offset rising fuel costs.  The
problem of emissions will be  discussed in detail in Section 3. 3. 3. 3;
suffice  it  to say here  that there  are opposing trends, with low NO
dictating lower combustion temperature while high efficiency implies
                                3-64

-------
higher temperatures and compression ratios.  Obviously, a
compromise between these two conflicting requirements has  to be found
and it will consist mainly in  improvements in the combustor,  in the pre-
mixing of the fuel with air, and in a reduction in the reaction tempera-
ture and residence  time.
              The  projected improvements  in thermal efficiency of
simple and regenerative cycle gas turbines as affected by higher tur-
bine inlet temperature and compression ratio are shown in Figures 3-21
and 3-22.  Turbine inlet temperatures of 2400°F and higher are pre-
dicted with improved materials and intercooled compressor air bleed
for turbine cooling, resulting in a simple cycle engine efficiency of up
to 40 percent.  For comparison, modern steam-electric plants have
design heat rates of 9500 Btu/kwh (36-percent efficiency) and future
projections are up to 8500  Btu/hr  (40-percent efficiency) for large
plants (1000  MW).
              The  combined cycle efficiency will increase as a result
of a  simple cycle efficiency increase and higher exhaust temperature
level, and in the next 10 years heat rates as low as 7000 Btu/kwh,  cor-
responding to an efficiency of approximately 50 percent are predicted
(Ref. 3-74).  This  is not unexpected since Curtiss-Wright presently
offers combined cycle MOD POD 75 units generating 99 MW at a thermal
efficiency of 44. 8 percent  (7650 Btu/kwh) (Ref. 3-74).
              With the gas turbine assuming not only peak loads but
also intermediate loads in electric power generation, its size will  be
increasing.  At present, the  largest gas turbine power plant  consisting
of two identical units has  a capacity of 480 MW (Ref. 3-75).  In future
plants such as the combined  cycle plant discussed in Reference 3-76 the
power level will be raised to 1000 MW.  Costs estimates for a plant
consisting of four units of  either simple cycle or regenerative cycle
configuration are presented in Ref.  3-77.  The projected cost varies
                                3-65

-------
   1  1970-DECADE TECHNOLOGY
  d  EARLY 1980'S TECHNOLOGY
   3  I ATE  )980'S TECHNOLOGY
                PRESENT-DAY GAS TURBINES
         	COMPRESSOR PRESSURE RATIO
         	  TURBINE INLET GA S,TE MPERTURE-F
     ,    50
    u
    2    45
    u
    Lk.
    LL.
    m    40
    ?   30
    ca
    LO
    <
    o
25

20
                                     COMPRESSOR BLEED AIR COOLED
COMPRESSOR BLEED
  AIR UNCOOLED
           0    50   100   150   200   250   300   J50  400 450
             SHAFT HORSEPOWER PER UNIT AIRFLOW - SHP/LB/SEC

                Figure 3-21.  Simple-cycle gas turbine
                             performance  (Ref. 3-77)

from $66.21/kW for  simple cycle units to $92.69/kW for regenerative
units (1970 dollars).
               With the looming shortage of light fuel distillates and
natural gas, it is to be expected that there will be a growing  require-
ment for gas turbines capable of burning heavy fuels or  fuels derived
directly from coal.
               A study of a combined cycle plant using coal (named
"Cogas") was made in Ref. 3-75.  Gasification of coal would generate
low Btu gas (150 to 200 Btu/scf) which could be produced at an invest-
ment cost  (over that of the cost of coal) of 22^/10  Btu.   Active  develop-
ment plans for building such a plant are underway, under sponsorship
                                 3-66

-------
                        FUEL-METHANE |HHV = IOOOBTU/FTJ)
                           AMBIENT aOF AND IOOOFT
  Z  50
  u
  oc
C  4*
Z
UJ
U
£  40
2  35
X
  I  30
     25
                    ?700f
                                 TURBINE INLET OAS TEMPERATURE=2tOOF
                                     2400F       6
                                                	•
                                                         10
           JOOOF
                                                       AIRSIDE
                                                       EFFECTIVENESS -
                                             -COMPRESSOR PRESSURE RATIO
            1970 DECADE TECHNOLOOY
                                       EARLY 1980'S TECHNOLOOY
  °    120  140   160   180  200   220  240   260   280  300  320   340
               SHAFT HORSEPOWER PER UNIT AIRFLOW-SHP/LB/SEC
             Figure 3-22.  Regenerative-cycle gas turbine
                           performance (Ref. 3-77)

of the Office for Coal Research with an objective of having a 135-MW
pilot plant in operation in mid-1976 (Ref.  3-78).  The plant would oper-
ate with a turbine inlet temperature of 2200°F and a cycle efficiency
(unfired boiler) of 47  percent.
               Another development which should be useful in permitting
emission-controlled combustion of heavy  and residual oils or coal is the
closed Brayton cycle  turbine under development by the Garrett Corpora-
tion under sponsorship of the U.S. Department  of Commerce.  The pre-
sent program is aimed at maritime applications and will culminate in a
closed air cycle turbine  developing power in the range of  40, 000 to
80, 000 hp (Ref.  3-79).  In this design the air is heated in an external
                                  3-67

-------
combustor thus permitting great flexibility in fuel selection and
emission control.
3. 3.2
   Applications
              The main applications of stationary gas turbines include
electric power generation for utilities and industrial use, pipeline ser-
vice,  and repowering.  These are discussed in the following sections.
3.3.2. 1
   Electric Power Generation
              An electric utility's demand grows daily from a nighttime
low to higher levels in the day,  with peaking at certain hours.  Fig-
ure 3-23 shows the three basic types of electric power generation: base
load,  which is continuous operation; intermediate load consisting, typi-
cally,  of 2000 to  6000 hours/year operation; and peaking load at about
1000 to 2000 hours/year (Ref. 3-80).  The peaking units have a reserve
load capability rating which is 10 to 15 percent above the nominal rating.
In addition,  there are standby emergency units which typically operate
less than 100 hours/year.
         Q
         3
UJ
Q.
U.
O
UJ
         u
         u
         tt
         UJ
         Q.
120
100
 80
 60
 40
 20
       I INSTALLED
       IRESERVE
        (15~-30%)  GENERATION
                                            INTERMEDIATE
                                            GENERATION H
                   JL
                           BASE
                           GENERATION
                           ,    ,\  ,
                     2000     4000     6000
                             HOURS PER YEAR
                                    8000    10,000
           Figure 3-23.  Electric power generation schedule
                        (Ref. 3-80)
                                 3-68

-------
              Gas turbines have been generally accepted by the
utilities in the last several years as peak load shavers, but it was only
in the last 2 to 3 years that they began to be installed for  intermediate
loads.  There are several reasons to explain the growing popularity of
gas turbines and Ref.  3-81 quotes about 20 of them.   The most impor-
tant are:  low initial cost; short delivery time; non-dependence on cool-
ing water for simple cycle and regenerative cycle machines; small
plant area; flexibility in operation on various fuels; and high thermal
efficiency of the regenerative and combined cycle.
              The investment cost of a simple cycle gas  turbine is
below $lOO/kw (for more details see Section 5. 3) which is less than
half of the investment  cost of a  steam plant.  The operating cost will
depend on the  percent  of gas turbine utilization and on fuel cost.
              The delivery time for a large gas turbine plant is
approximately two years  compared to four  years for a steam turbine
(Ref.  3-77).  This, combined with  a modular build-up in which most
of the assembly of the critical components  is done in the factory for
subsequent shipment to the site in the form of modules, reduces the
investment cost of gas turbines.
              The size of the gas turbine plant is considerably
smaller than that  of a  corresponding steam plant and Ref.  3-77 quotes
0. 006 acres/MW  for a regenerative cycle plant which amounts to
about one-tenth of the  requirement for a conventional steam  plant.
Table 3-17 presents a  comparison of the power density of different
type plants (Ref.  3-82).
              Most of the gas turbines in the above applications
operate with dual fuel  systems (No. 2 GT fuel and natural gas) and the
power turbine speed,  whether in a  single- or two-shaft engine, is
geared to 3600 rpm generators  (60 Hz current).  The typical Btu/kwh
data given before  correspond to fuel-to-air  ratios of approximately
                                3-69

-------
     TABLE 3-17.  POWER DENSITY COMPARISON (Ref. 3-82)
         Plant Type
Power Density (kw/ft ) (fuel to power
conversion equipment only)
  Reciprocating engines
  Fossil - steam turbine
  Nuclear - steam turbine
  Combined cycle (STAG)
  Simple  cycle gas turbine
                0. 01
                0. 07
                0.23
                0. 84
                3. 50
0. 018 to 0. 022 for simple and combined cycles operating at nominal
rating,  and 0. 012 to 0. 016 for regenerative configurations.
              Another feature of the gas turbine is its quick start
capability.  In a combined cycle,  the gas turbine can be at full load in
about 20 minutes with  the steam turbine following in another 20 min-
utes.  Ref.  3-75 reports  10 minutes minimum to 30 minutes normal
start to full load for a 240 MW gas turbine.  For comparison a con-
ventional steam-electric  plant requires about four hours.
3.3.2.2       Pipeline Service
              Gas turbines are used in pipeline service to drive
either compressors  (for gas) or pumps  (for oil).  The power range
is between 1000 to 12, 000 bhp and the average usage load factor is
high (90  percent). Because of the high load factor and the rising cost
of fuel,  high thermal efficiency is important and  regenerative cycles
are now  frequently used.
3.3.2.3       Re-powering
              Gas turbines are used for re-powering when  old, less
efficient steam plants  are being retired or  when there is a need for
power output increase. Ref. 3-83 lists some cases for re-powering
                                 3-70

-------
including:  (1) replacement of old boilers by a new set heated with gas
turbine waste heat, (2) replacement of steam turbines by a gas turbine
for an existing generator drive, and (3) gas turbines added to produce
power or to drive  forced draft fans of existing  boilers.  Detailed
examination of some cases has shown that repowering can not only
add power but also improve cycle  efficiency of old steam plants from
values less than 25 percent to better  than 35 percent.
3. 3. 2. 4       Installed and Projected Power
3.3.2.4.1     Electrical
              As  shown in Figure 3-24 (Ref. 3-81) the increase in
the installed  gas turbine power has proceeded very rapidly for  the
reasons discussed in Section 3. 3.2.1.  The trend toward larger units
is reported in Ref. 3-81 and this trend will continue as the gas turbine
assumes intermediate  and, in some cases, base load functions.  High
efficiency cycles will also be in a  growing demand and Ref.  3-74
reports that as of  December  1973  at least 15 combined and 15 regen-
erative cycle units were on order  by various utilities.
              The total  installed gas turbine power capacity in 1970
was 16, 500 MW compared to all generating power  sources of
331, 000 MW; that  is, approximately 5 percent  (Ref.  3-81).  The pro-
jected growth of turbine  power is shown in Figure  3-24, forecasting
approximately 69,  000  MW in 1980  (total U. S. capacity in  1980  is
estimated to  be 600, 000  MW).  The total U.  S.  electrical  power in
1986 was estimated by the Federal Power  Commission at  760, 000 MW
(Ref.  3-84) which  would  increase the gas turbine contribution to
approximately 9 percent. The average rate of gas turbine growth
between 1973 and  1980 is thus estimated at approximately 5300 MW /
year,  which compares with the average growth of 6000 to  7000  MW/
year in the last two years.  The next few years growth may be
tempered somewhat with the  growing shortage  of distillates and
                                 3-71

-------
                 2
                 z
(U, UUU


60, 000



50,000


40,000



30,000



20,000
10,000
n
1 i ' ' ' 	 P
/
/
/ -
/
/
/
/ -
/
/
/•
/ ~
/
/
/
/
/
/
/
- x
ff
^
r^*^

-------
     TABLE 3-18.  GAS TURBINE POWER FOR PIPELINE USE,
                    1958-1970 a (Ref. 3-81)
Horsepower
Range
,1000 - 4999
5000 - 9999
1000 and larger
TOTAL
Number of Gas Turbines
1260
470
260
1990
Total Installed •
hp x 103
1908
3416
3422
8746
aA summary table listing the installed horsepower of all stationary
engines is presented in Appendix A.
12.2 X 10  or 55.2 percent by reciprocating engines with the
remainder consisting mainly of electric power.  A  rather steady
growth in the rate of pipeline gas turbine power can be expected in
the future.  The development of the Alaskan pipeline will require
additional power as will the offshore drilling.  For these applica-
tions, the gas turbine provides the lowest cost and  weight per horse-
power unit.  For instance, the average cost of installed gas turbines
for a compressor  station is $250/hp versus about $400/hp for
reciprocating engines (Ref. 3-81).
              The operation of pipeline units is mainly continuous and
the gas turbines perform well at the high load factor showing smaller
initial, repair, and operating cost  than reciprocating engines
(Ref.  3-81).
3.3.2.4.3    Re-power ing
              In the era of growing fuel cost, the incentive for the
installation of gas turbines is increased thermal efficiency (reduced
fuel cost/kwhr) accomplished by combined cycle systems utilizing gas
                                3-73

-------
                  DC
                  2  30
                  zui 20
                  si
                     .0
  i      i      r
   lO.OOOhp AND LARGER-
1,000 -4,999hpv
                           V
                            I
                                    5,000 -9,999 hp
                      1965   1966
                                1967    1966
                                   YEAR
                                            1969
                                                 1970
           Figure 3-25.  Trends in size of turbines sold for
                         gas compression service
                         (Ref. 3-81)
turbine waste exhaust heat for steam production.  Ref. 3-86 quotes
approximately 800 MW re-powering with gas turbines by various
utilities in 1972.
3.3.2.4.4     Gas Turbine Manufacture and Operation
               The life of a stationary gas turbine is usually estimated
at 20 years.   However,  this figure  may vary somewhat depending upon
the duty cycle utilized.  The maintenance record of gas  turbines is
good and Ref. 3-87 quotes data from the operational experience of a
combined cycle, showing typical inspection intervals of 7500 hours
for the  combustor, 15, 000 hours for the hot gas section, and
30, 000  hours for major inspections.  The  steam turbine  generator is
inspected annually and overhauled every 10 years.
               The manufacturers of gas turbines  and important
engine characteristics are listed in Ref. 3-81 for  both domestic and
foreign manufacturers.   Table 3-19 summarizes most of the domestic
manufacturers of  stationary gas turbines.
                                 3-74

-------
TABLE 3-19.  STATIONARY TURBINES- U.S. MANUFACTURERS
Manufacturer
and Type
Allison-General
Motors
Model 404
Model 501-K

Model 91ZF

Avco- Lycoming
TF 12
TF 14
TF 25
TF 35
TF 40
Curtiss-Wright
Mod-Pod 20/25

Mod -Pod 30

Mod-Pod 40/50
Mod-Pod 60
Mod- Pod 75
TEC-150
TEC-350
General Electric
Model 1000
Model 3000




Model 5000


Model 7000


Model 9000

Power Range
(unit rating)


300 - 350 hp
4000 - 5000 hp

18, 000 - 22, 000
~20 MW

1 150 hp
1400 hp
2250 hp
2800 hp
3350 hp

24,000 - 31,000 hp
20 - 25 MW
35,000 hp
27 MW
40 - 50 MW
65 MW
70 MW
150 MW
350 MW

4000 - 5000 hp
8000 - 15,000 hp

5-15 MW


12,000 - 24,000 hp

11-26 MW
45 - 83 MW


80-91 MW

Application
(domestic only)


Pipeline
Pipeline
Electrical
Pipeline
Electrical

Pipeline
Pipeline
Pipeline
Pipeline
Pipeline

Pipeline
Electrical
Pipeline
Electrical
Electrical
Electrical
Electrical
Electrical
Electrical

Pipeline
Pipeline

Electrical


Pipeline and
process
Electrical
Electrical


Not for domestic
application
Cycle


Simple
Simple

Simple


Simple
Simple
Simple
Simple
Simple

Simple
Simple
Simple
Simple
Simple
Simple
Simple
Combined
Combined

Regenerative
Simple and
regenerative
Simple, com-
pound, regen-
erative
Simple and
regenerative
Simple
Simple, regen-
erative, com-
bined (STAG)
Combined
(STAG)
Fuel3


NG, DO
NG, DO, DF

NG, DO, DF


NG. DO
NG, DO
NG, DO
NG, DO
NG, DO

NG, DO, DF
NG, DO, DF
NG, DO. DF
NG, DO, DF
NG, DO, DF
NG, DO, DF
NG, DO, DF
NG, DO, DF
NG, DO, DF

NG
NG

NG, RO, DF


NG, DF

NG, DO. DF
NG, DO, DF


NG, DO, DF

NG - natural gas
RO - residual oil
DO - distillate oil
DF - dual fuel (oil/gas)
RG - refinery gas
                             3-75

-------
TABLE 3-19 - continued
Manufacturer
and Type
Solar
Spartan
Saturn

Centaur

Turbo Power and
Marine Systems
FT4A

FT4G

TP4-2 Twin-Pac
TP4-4 Twin-Pac
TP4-9 Twin-Pac
Turbo-Steam Pac
We sting house
Models: W-21,
W-31, 41, 52
Models: W-62,
W-72, 81, 82, 92
Models W-101,
W- 121, W-171
Model W-191
Model W-251
Model W-501
Model - Pace 20,
30
Model - Pace 260
Power Range
(unit rating)

225 kW
1 100 hp
750 - 800 kW
3000 - 3300 hp
2250 kW

20,000 - 30,000 hp
20 - 23 MW
35,000 - 42, 000 hp
25 - 32 MW
45 - 55 MW
90 - 110 MW
180 - 220 MW
75 - 125 MW

1800 - 5300 hp
5000 - 9500 hp
7 - 15 MW
14. 5 - 17 MW
18. 5 - 33 MW
37 - 65 MW
17. 5 - 30 MW
240 - 260 MW
Appl ication
(domestic only)

Electrical
Pipeline
Electrical
Pipeline
Electrical

Pipeline
Electrical
Pipeline
Electrical
Electrical
Electrical
Electrical
Electrical

Pipeline
Pipeline
Electrical
Electrical
Electrical
Electrical
Electrical
Electrical
Cycle

Simple
Simple
Simple
Simple
Simple,
regenerative

Simple
Simple
Simple
Simple
Simple
Simple
Simple
Combined

Simple,
regene rative
Simple,
regenerative
Simple, or with
heat recovery
Simple, or with
heat recovery
Simple
Simple
Combined
Combined
Fuel3

NG, DO, DF
NG, DO, DF
NG, DO, DF
NG, DO, DF
NG, DO, DF

NG, DO, DF
NG, DO, DF
NG. DO, DF
NG, DO, DF
NG, DO, DF
NG, DO, DF
NG, DO, DF
DO. RO

NG, RG
NG
NG, DO, DF
NG, DO
NG, DO, DF
NG, DO, DF
NC, DO
NG, DO. DF
NG - natural gas
RO - residual oil
DO - distillate oil
DF - dual fuel (oil/gas)
RG - refinery gas
                             3-76

-------
3.3.3         Emissions
              Before reviewing gas turbine  emissions,  some of the
existing or proposed emission standards are discussed to provide a
yardstick for the emission problem assessment.
3.3.3.1       Federal and State Standards
              The Federal Emission Standards for stationary power
plants  are defined by the Federal Register, Vol. 36, No. 247,
Part II of 23 December 1971.   These standards apply to new stationary
sources defined as "Fossil-Fuel Fired Steam Generators" and, there-
fore, are not applicable to gas turbines.  However, since standards
for gas turbines are now in the process of preparation,  it is of
interest to review these standards  since there  will probably be  some
similarity between these and the standards under consideration for
gas turbines.  Table 3-20 presents a partial list of currently appli-
cable emission standards.   (Ref.  3-88. )
              Table 3-21 presents conversion factors for the various
emission  units  computed for typical gas turbine fuel-air ratios.
              It should be noted that the emissions should always be
referenced to a  standard day condition and correlations  for the  con-
version from a current day to a standard day must be provided.  Tests
carried out indicate the dependence of the emission level on tempera-
ture and the following correction  formulae were suggested in
Ref. 3-89:

              CO . ,  ,   = CO     X 96
                 std day      meas

              HC . ,  .   = HC     X913
                 std day      meas

              NO   t , _,     NO        X i-
                 x std day =   x meas  04
                                 3-77

-------
            TABLE 3-20.   EXAMPLE OF EMISSION STANDARDS (MAXIMUM ALLOWABLE)
^"""---^^ Emission
^"^-—-^^^
Agency --— ^__^
Federal
Register



City of
Los Angeles Rule 67
San Diego Rule 50
County
Rule 52

Rule 62


Rule 08



Proposed EPA
(Ref. 3-90)







NO
X
0. 2 lb/106 Btu -
gas
0. 3 lb/10& Btu -
liquid
0. 7 lb/10° Btu -
solid
140 Ib/hr







125 ppm (3% O2) -
gas
225 ppm (3% O2) -
liquid
55 ppm (15% O2) -
gas
75 ppm (15% 02) -
liquid





CO


















90 ppm (1 5% O2)


or
2 1 5 ppm ( 1 5% O2 )
(plants < 50 X 10°
Btu/hr)


SO,
2
0. 8 lb/10° Btu -
liquid
1. 2 lb/106 Btu -
solid


200 Ib/hr




0. 5% per
weight of sulfur
in fuel




145 ppm (15% O2)


or
0. 8 lb/106 Btu




Smoke

20% opacity





20% opacity
Ringelmann #1









20 % opacity for
fired combined
cycle

10% opacity for
simple cycle and
unfi red
combined cycle

Particulates

0. 1 lb/10° Btu




10 Ib/hr


0 . 1 grains /
SCF
















Remarks

Applicable to
plants with
minimum of
250 X 106 Btu





After January
1974



Applicable to
plants with
50 X 106 Btu/hr
minimum









The ppm values are defined at specified percent O2 in the exhaust. The actual percent O^ '" Bas turbine exhaust at base load is
approximately 15%. The conversion to reference percent O2 is PPmactuai * [^ • ~ t^^2(ref)]/[^ ' ~ ^°O2 (actual)] = PPmref- Thus:
ppm (3% O2) = 3 X ppm (1 5% O2).
-J
00

-------
         TABLE 3-21.  GAS TURBINE EMISSION UNITS
             (11, 500 Btu/kwh;  18,500 Btu/lb fuel)

NO2
CO
SO2
HC (as C)
ppm
(F/A =
0.0166)
(measured)
100
100
100
100
ppm
(15% 02)
(reference)
112
112
112
112
lb/1000 Ib
fuel
9.5
5.8
13.2
2.5
lb/106 Btu
0.51
0.31
0.71
0.13
g/hphr
1.9
1.2
2.7
0.5
where:
                     T
                  "
                       inlet meas
                    T. ,  t  t ,  ,   '   inlet std day  =  519°R.
                      inlet std day               y
                To correct for  humidity the following correlation has
been proposed by Marchionna (Ref.  3-89).
                         NO
                           x meas
                      NO
                                      = e
                                         -19-H
                         x zero humid
where H - Ib of water/lb  of dry air.  Based on this equation a 17.3 per-
cent reduction  in NO  is realized  for each  1 percent of water vapor
increase in  the ambient air.
3. 3. 3. 2        Gaseous Emissions
3. 3. 3. 2. 1      Overview
               Before examining in more detail the stationary gas tur-
bine emissions, it would  be useful to get a perspective of them with
regard to the total emission problem in the U.S.  In particular, the
comparison will be made for NO  which is the predominant pollutant
                               X
for  stationary sources and probably most difficult to control.
                                3-79

-------
              In 1968,  stationary sources contributed approximately
60 percent to the total of 16 X 10° tons of NO  (calculated as NO?) emitted
                                           X                  LJ
in the U.S.  (Ref. 3-91).  These numbers are in reasonable agreement
with data published in Ref. 3-92  (10 X 10  tons of NO  for stationary
                     f                             x
sources and 20.6 X 10   ton total NO   emission).  Since the NO  emissions
                                  ^C                         jC
from gas turbines are substantially lower  than those of stationary recip-
rocating engines and  boilers,  the total gas turbine emissions in that
period were less than 1  percent of all stationary sources  (Ref.  3-85 and
3-92).
              The  comparison of the  growth in the total generated
power and in gas turbine power until 1980, which is needed for an
estimate of future gas turbine  emissions,  is  shown in Table 3-22.
              With regard to HC and  CO,  these emission species are
insignificant for gas turbines compared to other stationary  sources
(Ref. 3-85).  Moreover,  the HC  and CO contribution of all stationary
   TABLE 3-22.  TOTAL AND GAS TURBINE STATIONARY POWER

Total Installed Power, MW
Gas Turbine Power, MW
Gas Turbine Power, hpXlO
Electrical
1971
367,396
21,774
1972
399,606
27,918
1975
5l6,700a
42, 000b
1980
657,000a
69,000b
Other Stationary (pipeline, industrial)
6,313
6,628C
7,672C
9,792C
These data based on fossil and nuclear power plants already com-
mitted for this period (Ref. 3-84). Further power increase in the
1975-80 period probable with additional plant order (FPC estimates
electric generating power at 760,000 MW in 1980).
bFrom Ref. 3-81.
f*
5 percent/year growth assumed from 1971 value.
                                3-80

-------
combustors compared with all other emission sources is rather small,
amounting to only about 2 percent in 1968 (Refs.  3-93 and 3-94).
               The emission of particulates from hydrocarbon fuels
(mainly coal) in all stationary sources was 47 percent of the total par-
ticulate emission in the U.S.  (Ref.  3-95).  Again, the gas turbine  con-
tribution is insignificant because of operation with clean fuels.   These
aspects are further discussed in Section 3. 3. 3. 4.
               The emission of SO  from stationary sources is appre-
                                 X
ciable because of the high sulfur content in some of the fuels used,
particularly coal and oil crudes.  In 1968 the contribution of all
stationary sources amounted  to 75 percent of the total SO  emission
in the  U.S.  On the basis of Btu burned., stationary gas turbines would
contribute about 3 percent of  the total SO emitted by stationary units.
                                        X
In reality, their SO  emission will be much lower (about one percent)
because of the  low sulfur fuel (natural gas,  No. 2 GT  oil) used  in most
gas turbines.   The most economic control of SO  is related to fuel  -
5                                             x
desulfurization,  as discussed in Section 4. 3. 5.   Thus,  it appears that
while a control on stationary  gas turbine emissions is certainly bene-
ficial in preventing high local NO  or SO  concentrations, the  nation-
                                X       X
wide impact of such controls  will not be significant.
3. 3. 3. 2. 2     Emissions
               The gas turbine  combustors may be divided into the
primary or combustion zone and the post-flame or secondary zone.
The post-flame zone  can be subdivided  into  a thermal soaking  zone and
the final dilution zone in which  the excess air lowers the gas tempera-
ture to a value acceptable for use in the turbine rotating assembly
(Ref.  3-96). The principal pollutant species emitted from gas  turbines
are NO ,  followed by CO,  HC,  and SO?.  These  are  discussed  in the
       X                             C*
following  sections.
                                 3-81

-------
3.3.3.2.2.1   NQX
               The formation of NO  in gas turbines proceeds by three
                                  ji
known mechanisms, characterized as "hot air, " "prompt, " and "organic"
(Ref. 3-97).  The relative importance  of these three mechanisms on NO
is affected by the fuel composition,  fuel injector, flame temperature,
fuel-air  ratio, size  and shape of the combustor, and gas flow patterns.
               The "hot air" NO is produced in the hottest regions of the
combustor in accordance with the Zeldovich reactions:
               N2 + O ZT NO  + N
               N + O2 ZT NO  + O
               The studies on  the formation of NO led to three important
conclusions (Ref. 3-98):  The  NO formation rate is  exponentially depen-
dent on flame temperature while pressure and oxygen concentrations
have a secondary effect; NO is frozen at kinetically limited concentra-
tions when the combustion gases are quenched; the decomposition rate
of once-formed NO is so slow that it is not significant in a combustion
system.
               It appears,  therefore, that a key to low NO  is the
                                                        JC
achievement of a uniformly low flame temperature in the primary com-
bustor coupled with  a low residence time.  This is illustrated in
Figures  3-26 and 3-27.  The flame temperature in the primary  zone
can be lowered by operation with lean fuel-air mixtures with equivalence
ratios in the primary zone of 0.8 or less.  The experimental work with
advanced lean combustors in which the kerosene fuel was premixed  with
air and prevaporized (to reduce the  residence time), indicated that a
substantial decrease in NO  (with an attendent reduction in CO)  was
                          JC
indeed achieved (Ref.  3-99).
               As mentioned above,  the rate of formation of NO is
kinetically limited and, for instance, at 4025° F flame temperature it
would take over one second to  reach an equilibrium of 6000 ppm of NO
                                 3-82

-------
           JP-4FUEL
           70°F FUEL TEMPERATURE
           1 msec RESIDENCE TIME
           REQUIREMENT
                      O 0.02 100-1300
                      O 0.04 100-1300
                      D 0.06 100-1300
                      V 0.03    1000
        0 1000  1500  2000  2500  3000  3500 4000  4500
              PRIMARY ZONE TEMPERATURE, °F

      Figure 3-26.  Theoretical effect
                      of primary zone
                      temperatures on
                      NOX emissions
                      (Ref. 3-98)
      10
                     100 msec
1000     1500     2000     2500

PRIMARY ZONE TEMPERATURE, °F
                                           3000
Figure 3-27.  Combined effect  -  residence
               time and  primary zone
               temperature (Ref.  3-98)
                       3-83

-------
(Ref. 3-100).   Limiting the residence time and temperature will reduce
the NO appreciably, as shown in Figure 3-27.
               The second mechanism of "prompt"  NO formation is a
rapid but very short duration reaction involving highly reactive free
carbon or hydrocarbon radicals (Ref. 3-96).   The mechanism, which is
not yet completely understood, must occur during combustion and not in
the hot combustion products,  and measurements in laboratory flames
and estimates in combustion have shown "prompt" NO to be about one-
third of the amount produced  by the "hot air" mechanism  (Ref. 3-97).
               The third mechanism of "organic" NO formation is from
the chemically-bound nitrogen in fuel molecules.  Because the nitrogen-
carbon bond energy in the fuel molecules is so much lower than that of
molecular nitrogen, much of  the fuel nitrogen becomes oxidized.  These
reactions may proceed at lower temperatures than  needed for the two
former mechanisms, and organic NO is less responsive to the abatement
techniques now in use  and more to the fuel composition (Ref. 3-97).
               More data on fuel bound nitrogen are given in Sec-
tion  3. 3. 3. 6.   However the fraction of nitrogen converted to NO is not
firmly established.  Ref. 3-101 quotes for stationary boilers 30 percent
nitrogen conversion for fuel-oil,  and 50 percent for coal.   Ref. 3-102
gives the conversion factor as a function of total nitrogen content of fuel
oil.  Thus,  for instance,  fuel with 0.1 percent nitrogen should  contribute
approximately 30 ppm of NO  (based on  1 5 percent  O? in the exhaust) in
                           -?t                       C+
a typical gas turbine.
              A substantial amount of emission data,  particularly NO  ,
                                                                   i*L
was  obtained by the powerplant manufacturers and users.   A few of the
data are  presented below,  showing the effects of fuel  type (for  example,
fuel  oil versus natural gas) and operating  cycle (simple, regenerative,
control/overloaded).
              Based on the mechanisms of NO  formation,  as well as
                                             X*
analytical work by  various researchers,  it appears that NO  should be
                                3-84

-------
lower at part-load operation and when using gaseous fuels.  Gaseous
fuel operation approaches the ideal condition obtained in premixed,
vaporized, and well-stirred  combustors.  This is illustrated in
Figure 3-28 presenting measurements from a  20 to 22 MW simple cycle
gas turbine.  Over the whole load range the maximum NO  (no abatement
                                                       A
measures) was around 230 ppm with No. 2 fuel oil and about half of that
for gas  fuel (see Figure 3-29).  In these tests  the nitric oxide was
measured by the PDS (phenoldisulfonic acid) method (ASTM-D1608-60).
Similar results are shown in Ref. 3-103 from  another manufacturer's
35 MW gas turbine fired with No.  2 fuel oil.  The measurements of NO
       °                                                            x
were determined in this case by wet chemical  means with the PDS and
modified Saltzmann methods,  and by on-line gas analyzers. Figure 3-30
shows NO  emissions for two different size gas turbines,  26 and 66 MW.
Figure 3-31 was prepared from various data obtained by Aerospace and
illustrates:  a) increased NO  level for increased plant size because of
                           j£
increased residence time,  b) 30 to 50  percent  lower emission with fuel
gas than fuel oil,  and c) emission reduction by 20 to 30 percent obtained
with new combustor designs.  Advanced combustors are further dis-
cussed in Section 4. 3. 3. 3.
              Of interest are the emission data collected in Ref. 3-104
on gas turbines (and spark ignition engines) driving pipeline compres-
sors.   The data are for nine gas turbines varying in horsepower from
1000 to  20,000.  NO   was measured by the PDS and chemiluminescence
                  X
methods,  CO by NDIR (dispersive infrared), and HC (expressed as pure
carbon) by flame ionization.   The data of Ref.  3-104 are summarized
in Figure 3-32 in terms of ppm and g/hp-hr, the latter being more
characteristic for mechanical drive engines.  The data indicate that
the emissions of gas turbines are, on  the average,  an order of magni-
tude lower than those from uncontrolled spark ignition engines.
                                3-85

-------
          CAS FUEL
      is tuneinc it oii-
      HOHtltSS
      stu«n«t i on
      IS TUHKt I (I!
      ISIUUBKt t>MI
      »i iiKum it OK.
           ti «
-------
I-
Q.
 2
 m
 o
 2
 111
1.6
1.4
1.2
1.0
0.8
0.6
0.4
0.2
  0
                       MS-7001-B ENGINE
         MS-5001-N ENGINE
                                 _L
                                             220
                                        175  *;
                                        132
                                        88
                                                 O
           10    20    30   40    50
                    OUTPUT,  MW
                                   60   70
Figure 3-30.   NOX emissions — simple cycle oil-fired
              (Ref. 3-73)
     280

     240

  ^J 200
  ss
  £  160
  E
  O
  z
120

 80

 40

  0
        ^SIMPLE CYCLE
          NATURAL GAS
             I
              I
                          STANDARD COMBUSTOR
                        w OIL NO. 2
                        ® STANDARD COMBUSTOR
                          NATURAL GAS
                        A LOW NOv COMBUSTOR
                          OIL NO. 2
                        0 LOW NOX COMBUSTOR
                          NATURA! GAS
         10  20   30  40   50   60
                 BASE LOAD, MW
                                        70    80
  Figure 3-31.  Typical gas turbine NOX emission
                at base load
                         3-87

-------
f
o
8 ,

-

-

-
-1
; *
o *
i
»

*

•

rr i nrr. i

CE jp.too

f.t it i,, in

CE MI1IIR

P fc * HT-IOUT

P k ^* CCtC.4

9oUr S. turn T-IOOI I
, 1 1
1

]

1

1

1

")
5 *
VftJ
   ..... .	t..no.
f '

O

r
   Ct MUllt
              THC Emi.tion N
 - I 1.111 «.~rl. I-JOBI
                                                SI Frim> I






                                                ^g JP .IM~]






                                                CC M utti

                                                P fc W BT^QtCT  	I
                                              - CC rLm
                                             r- 1-	
                                                           CO Cont.ntr.l.OB.
                                                ^Hf >«tur* T- 1001
    Figure 3-32.   Emission data of pipeline gas turbines — natural gas

                    (Ref.  3-104)
                                       3-88

-------
               The NO  emissions of a regenerative cycle gas turbine
                      X
as compared to a simple cycle engine are subject to two opposing
effects.  The temperature of the air entering the combustor is higher
than in a simple cycle,  which would imply higher NOx formation rates.
On the other hand, with a fixed turbine inlet temperature, less fuel is
required in a regenerative  cycle (which accounts for higher efficiency)
and the primary combustion zone will run leaner with  lower flame tem-
perature which would imply a lower rate of NOx formation.  Thus,  the
regenerative and simple cycle engines should have similar NOx emis-
sions (Ref. 3-103 and 3-105).  However, this conclusion is not general
and is a function of the  combustor design, air flow pattern,  fuel type,
etc.   For instance, Figure 3-33 shows higher NOx emissions for a
natural gas-fired regenerative gas turbine than for  simple cycle engines.
               The combined  cycle with unfired exhaust will have similar
NO  emission  characteristics as that of a simple cycle because of the
   x
previously discussed low rate of NO decomposition.  The exhaust-fired
combined  cycle burns additional fuel but since the exhaust gases have  a
reduced oxygen content, the  flame temperature is lower even though the
initial temperature is higher.  Since the rate of formation of NO is highly
a °'7
z 0.6
o 0.5
*~ 0.4

3 0.3
                     uj 0.2
                     O
                                    SIMPLE CYCLE
                                     MS-500IM
                          REGENERATIVE  , ENGINE
                          M5-500I ENGINE /   \
                              \  /     V
                                           SIMPLE CYCLE
                                           MS-5001N ENGINE
                                  10
                                       15   20
                                    LOAD, MW
                                                    30
                    Figure 3-33.   NOX emissions,
                                   natural gas-fired,
                                   iso-conditions
                                   (Ref.  3-73)
                                   3-89

-------
temperature-dependent, less NO per megawatt is produced than in the
main combustor of the  gas turbine (Ref.  3-82).
3.3.3.2.2.2   CO
               Carbon monoxide is formed by incomplete combustion
(oxidation) of carbon and/or dissociation of CO?. After forming,  it can
be subsequently reduced by further oxidation to form carbon dioxide.
Similarly to the reaction rate between oxygen and nitrogen, the oxidation
rate of carbon monoxide is highly temperature-  and time-dependent. At
lower temperatures, the oxidation rate of CO is slower, and if the resi-
dence time is short more CO will be left in the combustion products.
               Thus, in general, conditions favoring low NO  will tend
                                                          X
to increase  the CO content and here lies  the difficulty in the design of
low emission combustors.  Fortunately,  most stationary gas turbines
operate at or  near base load and the combination of high temperatures
and high excess air  (typically 200 to 300  percent) results in relatively
low CO emission.
              In general,  concentrations of CO  are  higher at idle and
low power settings because of inadequate mixing, low temperature,  and
fuel quenching on the walls.  At higher power levels, some CO may be
formed by dissociation of CO.-, at high temperatures  in the primary zone
and be "frozen in" at levels higher than local equilibrium because of the
quenching effect from the dilution air (Ref. 3-106).
              Figure 3-34 shows the CO emissions  as a function of load.
The water injection  effect shown in the figure will be discussed in Sec-
tion 4. 3. 4.  The measurements of CO were obtained by an on-line
analyzer which responds to the concentration of  gas  by an electrochemi-
cal reaction through a. membrane sensor.
              Figure 3-35 presents CO emission data collected from
industry for fuel oil and natural gas-fired gas turbines.  One would
expect a lower CO emission with gas because of better mixing, but this
doesn't seem  to be the case, perhaps because of a lower primary  zone
                                 3-90

-------
 I
50
45
40
35
30
 O  25
 O  20
    15
    10
     5
     0
                    WITHOUT
                    WATER INJECTION
                      WITH WATER INJECTION
              10   15   20  25   30   35   40   45
                GENERATOR OUTPUT, MW
  Figure 3-34.  CO versus load,  W-251 engine
                (Ref. 3-103)
cuu
160
_

-------
temperature and the lower flammability limit obtained with gas as
compared to fuel oil (see Figure 3-28).  There is an increase of CO in
small combustors because of the attendent shorter residence times
available for further oxidation.
               The conflicting set of combustor  requirements for low
NO  concurrent with low CO is illustrated in Figure 3-36.  For example,
   X.
high CO at idle is accompanied by low NO .  The arrow in this figure
                                        X.
indicates the path for low emission combustor development. A similar
CO-NO  relationship is illustrated in Figure 3-37 (Ref.  3-107).
       X.
               The emission of CO is expected to be lower with a
regenerative cycle because of higher combustor inlet temperatures
(Ref. 3-108).
3.3.3.2.2.3  HC and Aldehydes
               The hydrocarbons are formed in a gas turbine combustor
in a  manner similar to that of carbon monoxide.  HC is the result of
incomplete combustion caused by poor mixing, inadequate fuel distribu-
tion  and atomization, and wall quenching.  Its formation is a function of
temperature,  and both  CO and HC will tend to behave in a similar
manner.
              Figure 3-38 shows HC data from  various  gas turbines.
Similarly to CO,  there  is no apparent relationship between  HC and fuel
type  (fuel oil or gas) but some increase in HC is noted for small engines.
              The data on aldehyde  emissions are rather limited.  In
general they are  a small proportion  of the total HC emissions which are
very low in most gas turbines.  Aldehydes are formed by partial  oxida-
tion  of hydrocarbon fuels combined with the formation of free radicals
which results  in a subsequent formation of the aldehyde group (CHO).
Typically, for fuel oil the aldehyde emissions amount to approximately
one-third of the total HC emissions (based on a colorimetric method of
measurement) (Ref. 3-109).   Like  HC the level of aldehydes declines
with  increasing load.
                                 3-92

-------
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                                     I  i i i i 111
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              NOX EMISSIONS, lb/1000 Ib FUEL
                                  100
 Figure 3-36.  CO versus NOX emission
               performance of conventional
               gas turbine engine combustors
               (Ref. 3-107)
         IDLE
         POWER
                         MAXIMUM
                         POWER
Figure 3-37.  GMA 100 gas generator emissions
              (Ref. 3-107)
                      3-93

-------
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0.12
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FUEL OIL o
• STANDARD COMBUSTOR
GAS
o
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0
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w
1 1 1 1 1 1 1
                        10    15    20    25
                              BASE LOAD, MW
30
35   40
               Figure 3-38.  HC emissions for various
                             gas turbine powerplants
3.3.3.2.2.4   SO,
              The oxides of sulfur emitted from gas turbines are
produced during the  oxidation process of the fuel which is the sulfur
donor.  Thus, the amount of sulfur oxide emissions are a direct
function of the fuel sulfur content and nothing can yet be done in a
practical sense to reduce it.  Thus, the Federal Emission Standard
for SO2 (0.8 lb/106  Btu for liquid fuels) led to an upper limit of
0.73 percent sulfur  in a liquid fuel.  This can be compared to presently
established sulfur limits for hydrocarbon  fuels, shown in Table 3-23.
              The combustion of sulfur results mainly in sulfur  dioxide,
although 3  to 4 percent are  further oxidized to sulfur trioxide which
readily dissolves in water to form  sulfuric acid (Ref. 3-105).
              Figure 3-39  shows SO- measurements as  a function of
load.   The change in ppm with load is caused by varying  the fuel-air
ratio.  In this case SO2 was measured by  grab  sampling  and by an
                                3-94

-------
  TABLE  3-23.  FUEL SULFUR CONTENT IN PERCENT (Ref.  3-73)
Fuel
Natural Gas
ASTM No. 1 GT
(kerosene)
ASTM No. 2 GT
(fuel oil)
ASTM No. 3 GT
(heavy oil distillates)
Minimum
-
0. 018
0. 070
0. 200
Average
-
0. 039
0.270
-
Maximum
0.010
0.098
0. 550
1.920
ASTM
Maximum
-
0. 5
0.5
-
           40
           35
           30
         E  25
         Q.
         t  20
        ^
        8  15
           10
            5
            0
               WITH WATER INJECTION
WITHOUT
WATER INJECTION
   I
                          10     15     20     25
                         GENERATOR OUTPUT, MW
            Figure 3-39.  SO2 versus load, engine W-251
                          (Ref. 3-103)

on-line analyzer, using membrane-type polarographic sensors.  The
small difference in SO  (10 ppm) with and without water injection is
probably due to different collection methods, i.e.,  "dry" versus wet.
              The emission of SO2 could become a problem in the  future
because of the continuing increase in total power and fuel consumption
                                3-95

-------
and a gradual change from no-sulfur natural gas to high sulfur crudes
and residuals (Ref.  3-109 and 3-110).
               Since the SO- emission is a function of fuel consumption,
all types of plants are equally affected.   The only plant-dependent
parameter would be its thermal efficiency and, of course, higher
efficiency would result in less SO? emission at a given  power output.
3. 3. 3. 2. 2. 5   Typical Stationary Gas Turbine Emissions
               From the information presented in Section 3. 3. 3. 2,
typical emission data for medium-size (30 MW) gas turbines with
state-of-the-art combustors are summarized as follows:
         NO   - (#2 oil)               150-220 ppm (15% O7)
           X                                            C*
              - (natural gas)           90-140 ppm (15% O2)
         CO  - (oil or gas)               5-100 ppm (15% O2)
         HC  (as C) - (oil  or gas)         1-20 ppm (15% O2)
               For smaller units, NO  decreases while CO and HC may
                                    JC
show an increase.  The opposite is true for larger  combustors.  Tech-
niques to reduce the emission levels (NO  in particular) to acceptable
values are discussed in Section 4. 3. 3.
3. 3. 3. 3        Smoke. Particulates and Odor
3. 3. 3. 3. 1      Smoke
               Gas turbine exhaust smoke which consists of small
carbon particles is a result of incomplete combustion in locally fuel-rich
zones  (Ref.  3-111).  The carbon particles forming the visible smoke are
0.01 to 1.0 microns in size.
               Several factors in the gas turbine combustor contribute  to
the generation of smoke.  These are:  coarse liquid fuel atomization,
fuel-rich pockets in the primary combustor,  low flame  temperatures,
insufficient fuelrair mixing and fuel composition.  In general,  heavy
fuels are greater smoke producers while  natural gas is almost
                                 3-96

-------
smoke-free.  In gas turbines,  the most favorable conditions for smoke
formation occur at idling when the atomization of liquid fuel is poor and
the temperatures are low.  Some smoke may also be generated at full
load when the fuel-air ratios are higher.
              Various methods for  smoke measurements have been
developed and some are  described in Refs. 3-112 and 3-113.  The
Bacharach smoke  number is obtained from the ASTM-D-2156  method
which specifies filtration of a standard volume of exhaust through a
Whatman paper filter (Ref.  3-103).  The  resulting smoke spot stain is
evaluated photometrically from zero (100-percent reflectance from an
unstained filter patch) to nine (10-percent reflectance from a heavily
stained patch).  The Ringelmann method consists of visual comparison
of smoke by trained observers against  four charts with numbers  from
one to  four, corresponding to increasing  smoke density.  The von Brand
smoke number is a measure of the reflected light from a  smoke sample
on filter paper taken under controlled conditions (Ref.  3-73).   The
reading of 100 corresponds to no smoke and zero to optically black
conditions.
              Typically the limit of smoke visibility in large  gas turbine
exhaust stacks is at Bacharach No.  5 (Ref. 3-114),  corresponding to
about 20 percent opacity which is defined as  a limit in  current Federal
Regulations.  Figure 3-40 shows typical smoke emissions from a 33 MW
gas turbine,  and Figure  3-41  shows smoke in von Brand numbers as a
function of particulates.   The reduction of smoke with  water injection
resulted from some of the combustor contaminants being  dissolved in
water.   Both units  shown in Figure  3-41 operate at  base load at
von Brand numbers of about 90 which is below the visibility threshold.
              The present day gas turbines  can,  in general, meet the
smoke limits imposed by various States (Ref. 3-88) without  difficulties,
but future operation on heavier fuels may require more development
work.
                                 3-97

-------
   o
   z
   o
                                       i      I
                                      ASTM-D-2156
                       WITHOUT WATER
                       INJECTION
                 WITH WATER INJECTION"
                     I
                        WITH SMOKE SUPPRESSANT
                        (w/o water Injection)
                        I	i	I	i
                     10    15    20    25    30
                      GENERATOR OUTPUT, MW
                                                35
    Figure 3-40.  Smoke versus load, engine W-251
                  (Ref. 3-103)
   §100
z1  90
u
O  80
to
2  70
u
   O
   >
      50-
          MS 7001B ENGINE
          SIMPLE CYCLE
          BASE LOAD
                                 MS 5001N ENGINE
                                 SIMPLE CYCLE
                                 BASE LOAD
                  0.005          0.010         0.015

         PARTICULATE EMISSIONS, lb/106 Btu INPUT
Figure 3-41.  Calculated particulate matter emission
              rate resulting from black smoke particles
              versus von Brand (reflectance) smoke
              number (Ref. 3-73)
                          3-98

-------
3.3.3.3.2
Particulates
              With regard to particulate matters, gas turbines produce
relatively low amounts because of the type of fuel burned and the high
efficiency of the combustion process (Ref.  3-115).  Light distillate fuel
oils burn relatively cleanly and contain very small amounts of ash and
trace materials,  and gaseous fuels are even better.   Carbonaceous and
sulfur-related particulate  species account for the majority of the  parti-
culate matter obtained with distillate fuels.  The former  can be con-
trolled by the combustor design and the latter by using low sulfur fuel.
When burning heavy, residual fuels,  the particulate emission attributed
to fuel ash and  fuel contaminants  can be appreciable (Ref. 3-115).  The
typical nature of particulate matter and its origin is shown in Table 3-24.
              Various methods for particulate emission measurement
exist and a few of those (EPA,  ASTM,  LAAPCD) are  described in
Ref. 3-98.  They can be divided into "wet" methods where a sample of
exhaust is discharged through a filter and wet impingers  which are sub-
sequently heated to remove the water,  leaving behind the residues; and
into "dry" methods where  the particulates are collected on dry, heated
TABLE 3-24. BREAKDOWN OF PARTICULATE MATTER (Ref. 3-115)
Particulate Matter
Smoke (carbonaceous)
Ash and trace metals
H2SO4, 2H2O, XSO4a
Organics
Ambient noncombustibles
Erosion products
Source
Fuel cracking

Fuel content and fuel additives
Fuel sulfur
Unburned fuel
Turbine air inlet
Material from gas flow path



surface
aConsidered particulate matter in only a few localities.
                                 3-99

-------
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      LIMIT RULE 67
AVERAGE EXHAUST CORRECTED FOR INLET -
                              SO, IN EXHAUST
                                ~
              10     15    20    25
                TURBINE LOAD, MW
                            30
35
 Figure 3-42.   W-251 engine combustion con-
               taminants  (dry filter method)
               versus load (Ref. 3-103)
U. f.\J
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) 0*0075 0.0100 0.0125 0.0150 0.0175 0.0200 0.0225
'FUEL/AIR RATIO
 Figure 3-43.  Particulate matter emission
               when burning crude and distil-
               late oil fuel (Ref. 3-103)
                    3-100

-------
filters.  According to Ref.  3-103,  the "wet" methods may show an
excessive particulate  content because of combined water, manufactured
sulfates and sulfated metal compounds.   This was shown by the use of
both "wet" and "dry" methods of measurement.
               Figure 3-42 shows the particulate emission (or combus-
tion contaminants) as  a function of load  using the "dry" filter methods
of measurement.  It can be seen that the limit of Rule 67 can be com-
fortably met.
               Figure 3-43 shows particulate emissions for distillate
and crude oil as a function of fuel-air ratio.  It is of interest to note
that while a 33 MW gas turbine consuming approximately 30,000 Ib/hr
of distillate oil would  meet the Rule  67 contaminant limit (10 Ib/hr),  it
would be much above the limit with crude oil.  This  confirms a comment
made earlier on heavy fuel operation.
3.3.3.3.3     Odor
               The odor from gas turbine exhaust is noticeable in high
traffic density  airports and some aspects of it have been investigated
by NASA (Ref.  3-116). It  is not  a problem in stationary gas turbines
but a few comments on the NASA investigations are appropriate.
               Exhaust odor intensity from gas turbines was evaluated
by a human panel and  graded progressively from the threshold level to
ranking No. 3.  The odor is created  by  fuel aromatics and by oxygenated
compounds such as alcohols,  aldehydes, and ketones (Ref. 3-116).  Test
data indicate that the odor increases with the concentration of oxygenates
as well as with combustion inefficiency.  There  was also dependence on
fuels in that high aromatic JP-5  had  the highest odor rating while natural
gas had the lowest.  The quality  of liquid fuel atomization had an effect
on odor by affecting combustion  efficiency, and air-assisted fuel injec-
tion was  especially effective in reducing the odor intensity.
                                 3-101

-------
3. 3. 3. 4       Noise
               Noise is a pollutant of some importance in stationary
gas turbine installations.  The advantage of the high-power density of
a gas turbine power plant and  the flexibility of its installation, which
is independent of water supply for simple and regenerative cycles,  also
makes possible its location at populated  load centers.  With this siting
advantage comes the necessity of quiet operation (Ref.  3-82).  What
other power generation plants gain in sound reduction by remoteness,
gas turbines must provide with silencing systems.
               The sources of the gas turbine noises include: the inlet;
the rotating machinery (compressor and turbine); and the exhaust.  The
noise generated by these three main contributors is attenuated by means
of a  suitable silencing system to meet the local sound criteria.   In addi-
tion to the gas turbine generated noise,  there is noise from the power
absorbing machinery such as electric generators and gas compressors
which also require silencing.  Most  of the engine manufacturers were
guided by the NEMA (National Electrical Manufacturers Association)
noise standards for industrial and residential centers.   These standards
are shown in Figure 3-44.  The standards applied to a given installation
will depend  on the plant location and usually the initial cost of a plant
includes noise silencing to a given NEMA level.  New standards are
being prepared by the American National Standards Institute (ANSI).
               Figure 3-45 presents typical noise emissions of  a station-
ary gas turbine compared against NEMA standards for industrial and
nearby residential installations.  The noise emitted falls between the
industrial and  residential limits and achievement of lower  values than
those shown requires some additional form of silencing.  The range
from the source to the listener is important in noise attenuation and if
the range is large compared to the length of the source, the noise is
reduced by  6 db every time the distance is doubled (Ref. 3-117).
                                3-102

-------
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                             RE:  NEMA STANDARDS
                                 PUB. NO. SM 33-1964
                                 4
                                 I
                                     6
                                     I
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                   63  125  250  500 1000  2000 4000 8000

                OCTAVE BAND CENTER FREQUENCY, Hz
                                                                               PEAKING
                                                                                         CONTINUOUS
                                                           TYPE OF AREA   DAYTIME  NIGHTTIME DAYTIME DAY &
        HEAVY INDUSTRY

        URBAN -
        NEARBY INDUSTRY

        URBAN —
        RESIDENTIAL

        SURBURBAN —
        RESIDENTIAL
                                                                       ONLY

                                                                        h


                                                                        9


                                                                        f
                                                          VERY QUIET —
                                                          SURBURBAN OR
                                                          RURAL RESIDENTIAL
                                                                                      ONLY

                                                                                       f


                                                                                       e
                                                                                     ONLY

                                                                                      g

                                                                                      f
                                                   NIGHT

                                                     e
             Figure 3-44.   NEMA noise standards for industrial and residential centers (Ref.  3-118)

-------
m  85
"^  80
ui  70
§  60
£  50
D
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ui
£  30
         O
         00
            20
                                               T
                   T
                                           HEAVY DUTY
                                           GAS TURBINE
                                           SOUND LEVEL
                  NEMA
                  LEVEL
              I
I
                                      I
                                                    I
                  32   63   125  250   500  1000  2000  4000  8000
                    OCTAVE BAND CENTER FREQUENCY, Hz
              Figure 3-45.  Sound level performance
                            of heavy-duty gas turbines
                            compared to NEMA sound
                            levels (Ref.  3-82)

              The "acceptable" versus "unacceptable" noise intensity
depends on many factors such as noise level (decibels), frequency,
duration, time of occurrence (day  versus night), background noise level,
etc.  A survey of community response was made in Ref.  3-118 and
compared against the proposed sound treatment levels by a gas turbine
manufacturer.
              The combined cycle and regenerative  gas turbines
provide their own noise attenuation.  In a combined cycle, the steam
generation chamber  attenuates noise either by the use of absorptive
materials or by  the reverberant room effect,  while the regenerator
absorbs noise by creating vortices in the flow passage through which
some of the acoustic energy is converted,  and by partial  scattering and
reflection.  This attenuation may become as high as  15 db at high
frequencies (Ref.  3-119).
             Various approaches to noise  legislation are described in
Ref.  3-120 and it appears that the technology is available to allow
                               3-104

-------
stationary gas turbine compliance with most of the existing noise
standards.
3. 3. 3. 5       Emissions Versus Gas Turbine Operating Conditions
              The stationary gas turbines operate, in most cases, at
nearly constant load level around the base or peak load design points.
Consequently, the emission control is believed to be somewhat easier
than for automotive gas turbines which have to  operate over the whole
load spectrum.
              The conflicting set of conditions leading to either low CO
and HC or low NO  are reflected in the gas turbine emission character-
istics showing high values of CO and HC at idle and peak loads,  with
much  lower values at design loads.  The opposite is true for NO
                                                              X
(Figures 3-36 and 3-37).  With high CO and HC at idle,  the appearance
of smoke becomes  more pronounced, and in a number of operational
units air blast atomization was added to improve the quality of fuel
atomization and to  reduce the smoke to an acceptable level.
              The trend in future gas turbines to improve cycle
efficiency and reduce fuel consumption will be in the direction of higher
compression ratios, higher turbine inlet temperature and in the con-
tinuing development of regenerative and combined cycles.   While this
trend  has a  beneficial effect on CO and HC, it will make the control of
NO more critical  because of the exponential relationship between the
rate of NO  formation and flame temperature.  However as shown
analytically in Ref. 3-121 and confirmed experimentally in Ref. 3-99,
the use of fuel-air  premixing combined with well-stirred combustion
and short residence times are very potent means in controlling NO .
                                                                X
              The regenerative  cycle gas turbines operate at lower
compression ratio  than simple cycles because of the compressed air
temperature increase in the regenerator.  It appears (Ref. 3-99) that
at a given turbine inlet temperature the emission of NO  may be similar
                                                     X
to or lower  than that of simple cycle engines.
                                3-1Q5

-------
               Combined cycle engines with unfired exhaust have
emission characteristics similar to those of simple cycles.  With fired
exhaust, additional emissions will be generated.  However per MW
output they should be lower than those of the main combustor.
3.3.3.6       Fuel Effects
               The fuels used most frequently in stationary gas turbines
include natural gas, kerosene, or naphtha (No.  1 GT) and fuel oil or
diesel fuel (No. 2 GT).  Future use will include heavy distillates or
crude  oil (No.  3 GT),  residual oils (No.  4 GT),  and products of coal
gasification.
               Natural gas is over 94-percent methane without any
metallic or sulfur contaminants.  Kerosene is a low carbon-hydrogen
ratio fuel with near zero contaminants,  low viscosity, and little varia-
bility in physical properties.  Fuel oil has higher viscosity than kero-
sene (Saybolt Seconds Universal 34 to 50 at 80°F),  a moderate carbon-
hydrogen ratio with near zero contaminants and little variability in
physical properties..  Heavy distillates or crude oils are characterized
by high carbon-hydrogen ratio, higher viscosity (Saybolt Seconds
Universal  180 to 2000 at 80° F) and  contaminants varying from zero to
about five ppm.  There is also a great variability in physical properties.
The  residual oils have still higher viscosity (of the order of 10,000 to
20,000 SSU), high carbon-hydrogen ratio, high contaminant level,  and
wide variability in physical properties (Ref.  3-122).  In many instances,
the high viscosity of Nos. 3  and 4 GT oils will require preheating of the
fuel to lower its viscosity to the 50 to 800 SSU level for  good atomization
(Ref.  3-123).   The product of coal gasification might be low Btu gas
(180  to 300 Btu/scf) consisting mainly of CO,  HZ, CH., CC",  and
about 50 percent N_.  Contaminants and sulfur might be removed in the
gasification process.
               The  current  interest in heavy fuels and coal is  stimulated
by the  growing shortage of natural gas and light distillates.  This is
                                3-106

-------
illustrated in Figure 3-46.  With the predicted rise and doubling of the
gas turbine fuel demand in the 1973-1979 time period,  there will be a
reduction in the use of natural gas and a rapid increase in the use  of
crude and residual oils.  The coal gas is not included in this  figure
since it will enter the picture in the post-1979 period.  However,  the
predictions shown in Figure 3-46 should be viewed with caution.  The
rapidly growing cost of light distillates may influence the refineries to
reduce the amount of residuals and new refineries  now being  built  may
not yield more than 2 to 3 percent  of the ingoing crude in the  form  of
residuals (Ref.  3-1Z2).  With low yield of residuals and a growing
demand for it in steam plants and gas turbines,  the price differential
                                                RESIDUAL
                                                CRUDE
                                                JET/
                                                KEROSENE
                                                DISTILLATE
                                                NATURAL GAS
                                                (thousand-
                                                barrel
                                                equivalent)
               (1972)   1973
1975
YEAR
1977
1979
            Figure 3-46.  Electric utility gas turbine fuel
                          demand (Ref.  3-109)
                                 3-107

-------
between No.  3 GT and No. 4 GT fuels may be narrowed considerably.
On the other hand,  there is a need for pretreatment of the fuel with
residual oils.  The residuals will contain almost all impurities of the
original crude in the form of trace elements such as  sodium,  potassium,
vanadium,  and lead (Ref. 3-124), as well as sulfur.
               Vanadium and sodium have been shown to cause turbine
bucket and nozzle corrosion and a limit of 0.5 ppm has been suggested
(Ref. 3-74).   Vanadium  and  sodium limits of 2 ppm are proposed in
Ref. 3-123.  Whichever limit is valid,  it is lower than the contaminants
present in the  residual oils requiring pretreatment of the fuel.  At
present this approach consists of washing out the alkali metal salts with
water and then separating the water by centrifuging or by electrostatic
means (Ref.  3-125).  Vanadium  can be inhibited by addition of magne-
sium in a weight ratio of three parts of magnesium to one part of
vanadium (Ref. 3-125).   The washing and inhibiting techniques report-
edly eliminate turbine corrosion as a major problem, but it increases
the total amount of combustion contaminants (Ref.  3-105).  These con-
taminants deposit in the turbine  and may be troublesome for runs over
ten hours duration.  When continued running is required,  the  turbine
can be cleaned by injecting a mild abrasive (spent refinery catalyst
or ground walnut shells) into the combustion system under load
(Ref. 3-74).  However,  this  will raise  the maintenance cost of the
plant, by 50 to 150 percent in comparison with GT  No. 2 fuel  operation.
The  pretreatment and contaminant removal add to the effective  cost of
the fuel.  According to a study conducted in 1971 (Ref. 3-123),  the cost
of residual oil pretreatment  would exceed the cost differential between
the No.  3 GT and No.  4  GT fuels.  However, this conclusion  may
require reappraisal in today's high priced fuel market.
               It is clear, however, that treatment of residual oils may
be another factor in limiting their use in heavy-duty gas turbines.  One
of the possibilities  of making the use of residual oils more attractive
                                3-108

-------
is to remove the impurities in the refining process.  Fuel desulfuriza-
tion, which at present could be done most effectively at the  refinery,
will also remove most of the trace metals,  hence eliminating or
reducing the  pretreatment requirements.
              The effect of fuel bound nitrogen on NO  was discussed
                                                   J\.
in Section 3. 3. 3. 2. 2. 1 and test data in Figure 3-47 show the exhaust
NO  concentration as a function of fuel-bound nitrogen. The data
emphasize the NO  problem with heavy and residual fuels.
                 Ji.
              The emission of SO2 is directly related to the sulfur
content in the fuel.  Some means of SO? removal are discussed  in
Section 4. 3. 5.
           100
            80
            60
            40
        -  20
in

I
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            10
             8
             6
                                            ENGINE
                                            5001-N
                               HEAVY
                               DISTILLATE,
              „  . NO. 2 DISTILLATE
                                    CRUDES
                                             RESIDUAL FUELS
                                                 l  l  I  I l 11
                 0.01
                      0.05   0.1
                    NITROGEN IN FUEL, %
0.5   1.0
         Figure 3-47.  Effect of fuel bound nitrogen on NOX
                       formation at base load (Ref. 3-126)
                                3-109

-------
               The smoke formation with heavy fuels and crudes does
not appear to be a problem and Refs. 3-123 and 3-124 report satisfac-
tory operation  with these fuels.  The smoke measurements show no
visible difference in comparison with No.  2 GT fuel due,  probably,  to
preheating of the heavy fuels for good atomization.
               Other means of smoke reduction were investigated in
the past in the  form of fuel additives and,  for instance, Ref. 3-111
reported "drastic smoke  reductions" of No. 2 GT fuel with 25 to 50 ppm
manganese addition.  However, the effect of this additive on turbine life
was not investigated and better techniques such as improved combustor
air flow distribution and atomization have since  been adopted for smoke
control.
               Extended gas turbine  operation on crude oils (1.3-million
hours) and on residual oils (3.5-million hours) performed by one gas
turbine manufacturer (Ref. 3-125) indicates no major technical problems.
The control of  emissions will be more critical with these fuels  requiring
careful attention to the combustor design and fuel atomization and, in
the case of residual oils,  more economical means of fuel pretreatment.
Operation with  low Btu coal gas should not present any  problem since
gas turbines  have  been operating successfully on blast furnace gas of
70 Btu/scf (Ref. 3-122).
                                3-110

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                          REFERENCES
3-1.    E.  F.  Obert,  "Internal Combusion Engines, "  Third Edition,
        International Textbook Company, Scranton,  Pennsylvania,
        (1968).

3-2.    R.  C.  Bascom,  L.  C. Broering, and D. E. Wulfhorst,
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3-3.    V.  S.  Yumlu and A. W. Carey,  "Exhaust Emission Charac-
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3-4.    "Cummins Power-Logging,  Construction and Mining, "
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3-5.    R. Mueller and L.  Lacey,  "New International Harvester Heavy
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3-6.    R.  Malcolm,  "International's New Motor Truck V-8 Engines, "
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3-7.    H.  Ricardo,  "The  High-Speed Internal Combustion Engine, "
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3-8.    E.  Eisele, "Daimler Benz Passenger Car Diesel Engines -
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        (January  1968).

3-9.    C.  R.  McGowin,  "Stationary Internal Combustion Engines in
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3-10.   P.  S.  Myers, O. A.  Uyehara and H.  K. Newhall, "The ABCs
        of Engine Exhaust Emissions, "  SAE Paper 710481 (1971).

3-11.   I.  M.  Khan,  G.  Greeves,  and C. H. T.  Wang,  "Factors
        Affecting Smoke and Gaseous Emissions from Direct Injection
        Engines and a Method of Calculation, "  SAE Paper 730169
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3-12.   W. F.  Marshall and R. D. Fleming,  "Diesel Emissions Rein-
        ventoried, " Bureau of Mines Report PB-201896 (July 1971).
                               3-111

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3-13.    C.  J.  Walder, "Reduction of Emissions from Diesel
         Engines," SAE Paper 730214 (January 1973).

3-14.    S. M.  Shahed, W. S. Chiu, and V. S. Yumlu,  "A Preliminary
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3-15.    G.  Blair  Martin and E.  E. Berkau, "An Investigation of the
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3-16.    C.  T.  Hare, and K. J. Springer,  "Exhaust Emissions  from
         Uncontrolled Vehicles and Related Equipment Using Internal
         Combustion Engines, " Final Report AR-898, Part  5, Heavy
         Duty Farm,  Construction and Industrial Engines; Southwest
         Research Institute (October  1973).

3-17.    Control of Air Pollution from New Motor Vehicles  and  New
         Motor Vehicle Engines;  Environmental Protection Agency,
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3-18.    "Exhaust Emissions from Uncontrolled Vehicles and
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         Counterparts; Southwest Research Institute (October 1972).

3-19.    F.  J.  Hills,  T.  O. Wagner,  and D.  K. Lawrence,  "CRC
         Correlation of Diesel Smoke Meter Measurement, " SAE
         Paper 690493 (May 1969).

3-20.    "Characterization and Control of Emissions from Heavy Duty
         Diesel and Gasoline Fueled Engines, " Bureau of Mines,
         Bartlesville,  Oklahoma  (December 1972).

3-21.    F.  S.  Schaub and K.  V.  Beightol,  "NOX Emission Reduction
         Methods for  Large Bore Diesel and Natural Gas  Engines. "
         ASME Paper 71-WA/DGP-2, November 28 - December 2,
         1971.

3-22.    K. J.  Springer, "Emissions from a  Gasoline and Diesel
         Powered  Mercedes 220 Passenger Car," AR-813, Southwest
         Research Institute, San Antonio, Texas (June 1971).
                                3-112

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3-23.   R.  E.  Bosecker and D. F. Webster, "Precombustion
        Chamber Diesel Engine Emissions - A Progress Report,"
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3-24.   I.  M. Khan,  C. H. T. Wang and B. E. Langridge, "Effect of
        Air Swirl on Smoke and Gaseous Emissions from Direct
        Injection Diesel Engines, " SAE Paper 720102 (10-14 January
        1972).

3-25.   J. M.  Perez and E.  W. Landen,  "Exhaust Emission Charac-
        teristics of Precombustion Chamber Engines, " SAE Paper
        680421 (20-24 May 1968).

3-26.   J. L. Butler, J.  H.  GarrettandJ.  L.  Hoch, "Cummins
        K-Series Engines, " SAE Paper 740036 (25 February -
        1 March 1974).

3-27.   N.  A.  Henein and J.  A. Bolt, "The Effect of Some Fuel and
        Engine Factors on Diesel  Smoke, " SAE Paper 690557
        (11-14 August 1969).

3-28.   S. R. Krause and G. L. Green,  "Effect of Intake Air Humidity
        and Temperature on  Diesel Emissions with Correlation
        Studies, " Work performed under  contract to the Engine Manu-
        facturers Association (28  September 1972).

3-29.   H.  A.  Ashby, "Final Report, Exhaust Emissions from a
        Mercedes Benz Diesel Sedan, " Test and Evaluation Branch,
        EPA, Ann Arbor, Michigan (July 1972).

3-30.   Statement by Mercedes Benz of North America; Hearings
        Before the Subcommittee on Air and Water Pollution of the
        Committee on Public Works, United States Senate (18 May  1973).

3-31.   C.  W.  Savery, R. A. Matula,  and T. Asmus,  "Progress in
        Diesel Odor Research," SAE Paper  740213 (February 1974).

3-32.   G. J. Barnes,  "Relation of Lean  Combustion Limits in Diesel
        Engines to Exhaust Odor Intensity, " SAE Paper 680445
        (May 1968).

3-33.   K. J. Springer and C.  T.  Hare, "Four Years of Diesel Odor
        and Smoke Control Technology Evaluations - A Summary, "
        ASME  Paper 69-WA/APC-3 (November  1969).

3-34.   K. J. Springer and R.  C.  Stahman,  "Control of Diesel
        Exhaust Odors, " Conference on Odors,  The New York Aca-
        demy of Sciences; New York, New York, Paper No.  26
        1-3 October 1973.

                               3-113

-------
 3-35.    C. T. Hare, K. J.  Springer,  J. H.  Somers, and T. A.  Huls,
         "Public Opinion of Diesel Odor, " SAE Paper 740214
         (February 25 - March 1,  1974).

 3-36.    "Chemical Analysis of Odor Components in Diesel Exhaust, "
         Arthur D. Little Report No. ADL 74744-5  (September 1973).

 3-37.    K. J. Springer, "An Investigation of Diesel Powered Vehicle
         Odor and Smoke - Part III, " Southwest Research Report
         No.  AR-695 (October 1969).

 3-38.    J. W. Vogh, "Nature of Odor  Components  in Diesel Exhaust, "
         Journal of the Air Pollution Control  Association, 19, (10)
         (October 1969).

 3-39.    D. F. Merrion, "Effect of Design Revisions on  Two Stroke
         Cycle Diesel Engine Exhaust," SAE  Paper  680422 (May 20-24
         1968).

 3-40.    M. F. Russell,  "Reduction of Noise  Emissions  from Diesel
         Engine Surfaces," SAE Paper 720135 (January 1972).

 3-41.    M. F. Russell,  "Automotive  Diesel  Engine Noise and Its
         Control, " SAE Paper  730243  (8-12 January,  1973).

 3-42.    D. D. Tiede and D.  F. Kabele, "Diesel Engine  Noise Reduc-
         tions by Combustion and Structural Modifications, " SAE Paper
         730245 (January 1973).

 3-43.    G. E. Thien, "The  Use of Specially  Designed Covers and
         Shields to Reduce Diesel  Engine Noise, " SAE Paper 730244
         (January 1973).

 3-44.    Waukesha V12 Diesel  Power Units, Waukesha Motor Company
         Bulletin 5124.

 3-45.    E. P. Grant, "Auto Emissions, " Motor Veh. Poll.  Cont.
         Board Bulletin. Vol. 6, No. 4, p. 3, (1967).

 3-46.    H. C. McKee and McMahon, Jr. , "Automobile Exhaust
         Particulates - Source and Variation,  "  53rd Annual Meeting
         of APCA, Cincinnati,  Ohio, May I960.

3-47.    N.  Gilbert and F. Daniels, "Fixation of Atmospheric Nitrogen
         in a Gas Heated Furnance, " Ind. and Eng.  Chem. , 40,  1719,
         (1948).	
                                3-114

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3-48.    H.  K. Newhall and E. S.  Starkman, "Direct Spectroscopic
         Determination of Nitric Oxide in Reciprocating Engine
         Cylinders, " SAE Paper No.  670122, SAE Automotive Engr.
         Congress, Detroit, Michigan (January 1967).

3-49.    L.  C. Broering, Jr.  "An Evaluation of Techniques for
         Measuring Air-Fuel Ratio, " SAE Paper No. 660118, SAE
         Congress, Detroit, Michigan, January 1966.

3-50.    W.  A. Daniel, "Flame Quenching at the Walls of an Internal
         Combustion Engine, " Sixth Symposium (International)  on
         Combustion, New York, Reinhold Publishing Co. ,  1957.

3-51.    R.  J.  Steffensen, J.  L. Agnew and R. A. Olsen,  "Combustion
         of Hydrocarbons Property Table, " Engineering Bulletin of
         Purdue University, Engineering  Extension Series 122,
         Lafayette, Indiana, (May 1963).

3-52.    E.  S.  Starkman and H. K. Newhall, "Characteristics  of the
         Expansion of Reactive Gas Mixtures as Occurring in Internal
         Combustion Engine Cycles, " SAE Paper No.  650509, SAE
         Transactions, _75_ (1967).

3-53.    J.  B.  Edwards and D. M.  Teague, "Unraveling the Chemical
         Phenomena Occurring in Spark Ignition Engines, " SAE Paper
         No. 700489.

3-54.    W.  A. Daniel and J.  T. Wentworth, "Exhaust Gas Hydro-
         carbons - Gensis and Exodus, " SAE Paper  No. 486B;  SAE
         Technical Progress Series,  Vehicle Emissions,  6^,  192,  (1964).

3-55.    C.  E. Scheffler, "Combustion Chamber Surface Area,  A Key
         To  Exhaust Hydrocarbons, "  SAE Paper No.  660111, SAE
         Progress in Technology,  Vehicle Emissions, Part II,  12, 60,
         (1966).                                             ~

3-56.    J.  C.  Gagliardi, "The Effect of Fuel Anti-knock Compounds
         and Deposits on Exhaust Emissions, "  SAE Paper No.  670128,
         SAE Automotive Engineering Congress, Detroit, Michigan,
         (January 1967).

3-57.    T.  A. Huls and H.  A. Nichol, "Influence of Engine Variables
         on  Exhaust Oxides of Nitrogen Concentrations from a Multi-
         Cylinder Engine, " SAE Paper No. 670482,  SAE Mid-Year
         Meeting,  Chicago,  Illinois (May  1967).

3-58.    R.  W. McJones and R. J.  Carbeil,  "Natural  Gas  Fueled
         Vehicles Exhaust Emissions and Operational  Characteristics, "
         SAE Paper No.  700078, SAE Congress, Detroit, Michigan,
         January 1970.

                                3-115

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3-59.    "Oxides of Nitrogen in Gaseous Combustion Products
         (Phenoldisulfonic Procedure), " ASTM Standard D-1608-60.

3-60.    B.  E. Saltzman, "Modified Nitrogen Dioxide Reagent for
         Recording Air Analyzers, "Analytical Chemistry,  32,  135
         (I960).

3-61.    H.  Niki, A.  Warmick, and R.  R. Lord, "An Ozone-NO
         Chemiluminescence Method for NO Analysis in Piston and
         Turbine Engines, " SAE Paper No. 710072, SAE Congress,
         Detroit, Michigan, January 1971.

3-62.    B.  M. Sturgis et al,  "The Application of Continuous Infrared
         Instruments  to the  Analysis of Exhaust Gas, "SAE Paper
         No. 11B, SAE Annual Meeting,  January 1958,  SAE Technical
         Progress Series, Vehicle Emissions, £>, p.  81 (1964).

3-63.    L.  B. Graiff, C. E.  Legate and I. C. H.  Robinson, "A Fast-
         Response Flame lonization Detector  for Exhaust Hydrocarbons, "
         SAE Paper No.  660117,  SAE Congress,  Detroit, Michigan,
         January 1966.

3-64.    L.  J.  Papa,  "Gas Chromatography - Measuring Exhaust
         Hydrocarbons Down to Parts Per Billion, " SAE Paper
         No. 670494,  SAE Progress in Technology,  14,  43,  (1971).

3-65.    P.  E. Oberdorfer, "The  Determination  of Aldehydes in Auto-
         mobile Exhaust Gas," SAE Paper No. 670123, SAE Progress
         in Technology, Vehicle Emissions,  14,  32, (1971).

3-66.    L.  Settlemeyer,  "Application of the Scanning Electron Micro-
         scope /X-Ray Spectrometer to Automobile  Exhaust Particulates, "
         SAE Paper No.  710637,  SAE Joint Meeting, Midland, Michigan,
         October 1970.

3-67.    Analytical Chemistry, 4j_,  (5),  4R,  (April 1969).

3-68.    "Control of Air Pollution From New  Motor Vehicles and New
         Motor Vehicle Engines, " Part in, EPA, Federal Register,
         June  28, 1973, 3Q, (124),  and August 7, 1973,  W, (151).

3-69.    K.  J.  Springer,  "Baseline Characterization and Emissions
         Control  Technology Assessment of HD Gasoline Engines, "
         Final Report to EPA,  Contract  EHS 70-110, Southwest
         Research Institute  (November 1972).

3-70.    K.  T. Matsumoto,  T. Toda and H. Nohira, "Oxides of Nitrogen
         from Smaller Gasoline Engine, " SAE Paper No. 700145, SAE
         Congress,  Detroit, Michigan, January  1970.

                                3-116

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3-71.   "Characterization and Control of Emissions From Heavy
        Duty Diesel and Gasoline Fueled  Engines, " Final Report,
        No. EPA-LAG-0129 (D), Prepared by Fuels Combustion
        Research Group, Bureau of Mines (December 1972).

3-72.   Gas Turbine World,  February - March  1973.

3-73.   N. R.  Dibelius and E. W. Zeltmann,  "Gas Turbine Environ-
        mental Impact Using Natural Gas and Distillate Fuels, "
        General Electric 73-GTD-6, (February 1973).

3-74.   J. Papamarcus,  "Gas Turbine Faces Challenge," Power
        Engineering  (1  December 1973).

3_75.   D. Bruce, "Installation and Operation of Two 240 Megawatt
        Peaking Plants, ASME, 73-GT-39 (April  1973).

3-76.   F. L.  Robson and A.  J.  Giaramonti;  "The Use of Combined
        Cycle  Power Systems in Non-Polluting Central Stations. "

3-77.   F. R.  Biancardi and G. T. Peters, "Advanced Non-Polluting
        Gas Turbines for Utility Applications in Urban Environments, "
        ASME, 72-GT-64 (March 1972).

3-78.   Gas Turbine World,  May 1973.

3-79.   "MarAd 9000 Hp Closed Cycle Program, " Gas Turbine World,
        (August - September  1973).

3-80.   H. L.  Smith and R. J. Budenholzer,  "Cyclic Energy Demands
        Supplied Economically with  Gas  Turbines and Combined
        Cycle  Plants, " ASME, 71-GT-71 (April 1971).

3-81.   J. W.  Sawyer,  "Gas  Turbines in Utility Power Generation, "
        Sawyer's Gas Turbine Catalogue  (1973).

3-82.   W. J.  Ahner, "Environmental Performance, " General Electric
        Report GER-2480 (1971).

3-83.   E. J.  Willson,  "Re-powering.  An Economic Alternate
        Meeting Mid-Range Power Requirements, " ASME,  73-GT-35
        (April 1973).

3-84.   "New Generating Capacity, " Power Engineering (April 1973).

3-85.   C. R.  McGowin, "Stationary Internal Combustion Engines in
        the United States, " Shell Report EPA-R2-73-21 0 (April 1973).
                                3-117

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3-86.    "Annual Plant Design Report, " Power  (November  1973).

3-87.    H.  W. Carlson,  "The STAG Cycle," General Electric Report
         USDA-4-72, (September 1972).

3-88.    N.  R. Dibelius and R.  J. Ketterer,  "Status of State Air
         Emission Regulations Affecting Gas  Turbines, " ASME,
         73-WA/GT-8 (November 1973).

3-89.    ASME, Gas Turbine  Division,  letter to EPA (1 May 1973).

3-90.    "Response  to Preliminary (draft) Proposed Standards for
         Control  of Air Pollution from Stationary Gas Turbines, "
         General Motors  (March 1973).

3-91.    W. Bartok  and A.  Shiepp, "Control of  U. S. NOX Emissions
         from Stationary  Sources, " Chemical Engineering Progress.
         6J7, (February 1971).

3-92.    Air Pollution Control Administration,  "Control Techniques
         for Nitrogen Oxide Emissions  from Stationary Sources, "
         U. S.  Department of Health,  Education and Welfare
         (March 1970).

3-93.    Air Pollution Control Administration,  "Control Techniques
         for Hydrocarbons and Organic Solvent  Emissions from
         Stationary Sources," U.S. Department of  Health, Education
         and Welfare, (March 1970).

3-94.    Air Pollution Control Administration,  "Control Techniques
         for Carbon Monoxide Emissions from Stationary Sources, "
         U.S.  Department of Health,  Education and Welfare (March 1970).

3-95.    J.  P.  Tomay et al, "A Survey of Nitrogen-Oxides  Control
         Technology and the Development of a Low NOX Emission
         Combustor, " Journal of Engineering Power (July 1971).

3-96.    C.  R. Fenimore, M.  B. Hilt and R.  H. Johnson, "Formation
         and Measurements of Nitrogen Oxides  in Gas Turbines, "
         Gas Turbine International. July-August 1971.

3-97.    M.  B. Hilt and R.  H.  Johnson, "Nitric Oxide Abatement in
         Heavy Duty Gas Turbine Combustors by Means of Aerodynamics
         and Water Injection," ASME, 72-GT-53 (March 1972).

3-98.    T.  F. Nagey, P. M.  Kolenki and M. E. Nayler, "The  Low
         Emission Gas Turbine Car," ASME, 73-GT-49 (April 1973).
                                3-118

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3-99.   N.  A.  Azelborn et al,  "Low Emissions Combustion for the
        Regenerative Gas Turbine,  "ASME, 73-GT-1 2 (April 1973).

3-100.  W.  Cornelius and W. R. Wade,  "The Formation  and Control
        of Nitric Oxide in a Regenerative Gas  Turbine Burner, "
        SAE paper 700709 (September 1970).

3-101.  M.  R. Beychok,  "NOX Emission from Fuel Combustion
        Controller, " The Oil and Gas Journal  (February  26, 1973).

3-102.  E.  E.  Berkau and D. J. Lachapelle,  "Status of EPA1 s Com-
        bustion Program for Control of  Nitrogen Oxide Emissions
        from Stationary Sources, " EPA Report (September 19,  1972).

3-103.  M.  J.  Ambrose and E.  S. Obidinski, "Recent Field Tests
        for Control of Exhaust Emissions from a 35 MW  Gas Turbine, "
        ASME, 72-JPG-GT-2 (September 1972).

3-104.  H.  E.  Dietzmann and K. J. Springer, "Exhaust Emission
        from Piston and Gas Turbine  Engines used in Natural Gas
        Transmission,  " AR-923, Southwest Research Institute,
        (January 1974).

3-105.   R.  J.  Johnson  et al,  "Gas Turbine  Environmental Factors -
         1973," GE Report (1972).

3-106.  F.  W. Lippert, "Correlation of Gas Turbine Emission Data, "
        ASME, 72-GT-60 (March 1972).

3-107.  J.  N.  Barney and F.J. Verkamp,  "Aircraft Gas Turbine
         Engine High  Altitude Cruise Emissions, "  Detroit Diesel
        Allison Report, (1 August 1973).

3-108.  C.  A. Amann et al,  "Some Factors Affecting Gas Turbine
        Passenger Car Emissions," SAE paper 720237.

3-109.  V.  De Biasi, "Double Standard" on Fuel Oils Would Favor
        Steam over Gas Turbine Plants, "Gas Turbine World
         (September 1973).

3-110.  F.  S.  Olds,  "SOX and NOX" Power Engineering (August 1973).

3-111.  S.  M. DeCorso et al,  "Smokeless  Combustion in Oil-Burning
        Gas Turbines," ASME,  67-PWR-5 (September 1967).

3-112.  F.  J.  Hills et al,  "CRC Correlation of Diesel Smokemeter
        Measurements," SAE paper 690493 (May 1969).
                               3-119

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3-113.   "Diesel Engine Smoke Measurement," SAE Technical Report
         J255.

3-114.   M. J.  Ambrose and J.  Bott,  "Zero Exhaust Visibility - A
         Goal Attained for Peaking Gas Turbines, " ASME,  70-PWR-17
         (September 1973).

3-115.   M. B.  Hilt and D.  V.  Giovanni,  "Particulate Matter Emission
         Measurements for Stationary Gas Turbines," ASME, 73-PWR-
         17,  (September 1973).

3-116.   H. F.  Butze and D. A. Kendall,  "Odor Intensity and Charac-
         terization Studies of Exhaust from a Turbojet Engine Com-
         bustor, "  NASA Technical Memorandum, NASA-TMX-71429
         (November 1973).

3-117.   R. B.  Tatge, "Noise Control of Gas Turbine Power Plants,"
         Sound  and Vibration (June 1973).

3-118.   Turbo Power and Marine Systems,  "Technical Talk — Noise
         Control. "

3-119.   M. Weiss,  "Acoustic Behavior of Exhaust Heat Recovery
         Systems for Gas Turbines, " ASME,  73-GT-33 (April 1973).

3-120.   W. A.  Smith, "Noise Control Legislation, " Journal of the
         Air Pollution Control Association, 23, (April 1973).

3-121.   W. R.  Wade et al, "Low Emission Combustion for the
         Regenerative Gas Turbine," ASME,  73-GT-ll (April 1973).

3-122.   "Fuels for Gas Turbines," The Westinghouse Gas  Turbine.

3-123.   G. Vermes, "Heavy Oil or Residual Oil -  New Opportunity
         for the Utility Gas Turbine," ASME, 71-GT-81 (April 1971).

3-124.   S. M.  DeCorso,  "Crude Oil Firing in the Utility Gas Turbine, "
         ASME, 71-WA/GT-ll (December 1971).

3-125.   A. O.  White, "20 Years Experience Burning Heavy Oils  in
         Heavy Duty Turbines, " ASME, 73-GT-22 (April  1973).

3-126.   GE Gas Turbine Department letter to EPA, "Comments on
         the Preliminary Draft" (1 March 1973).
                               3-120

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                             SECTION 4

         AUTOMOTIVE EMISSION CONTROL TECHNOLOGY
              This section reviews the various emission control tech-
niques, devices, and systems which have been or still are under investi-
gation by the automotive industry.
              The presently known emission control approaches appli-
cable to  reciprocating and gas turbine engines can be divided into two
categories: preventive and corrective.  The preventive control methods,
which include modifications of the operating conditions of the engine as
well as certain engine design changes,  are aimed at minimizing the
formation of pollutants in the  combustion chamber. Conversely, the
corrective control methods include certain add-on devices, which are
designed to convert the pollutants emitted from the engine to harmless
compounds before admission into the atmosphere.
              Potential emission control techniques for diesel engines
are discussed in Subsection 4.1.  Gasoline engines are treated in Sub-
section 4. 2 and  gas turbines are treated in Subsection  4. 3.
                                 4-1

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4.1            DIESEL ENGINES
               This section of the report deals with the work performed
by many investigators to reduce the exhaust emissions from diesel
engines and presents the most significant results achieved to date.
Although a limited amount of pertinent test data have been reported for
large stationary diesel engines, most of the research and development
work concerned with emission control of diesels was performed on
engines designed for use primarily in on-highway applications.  Emis-
sion standards for this engine  category have been in existence for some
time, and more stringent standards might be promulgated by the EPA
for future model year heavy-duty atuomotive engines.
               The emission control techniques considered to date for
use in diesel engines fall into one of the following four categories:
         1.     Variation in engine operating conditions
         2.     Engine component modifications
         3.     Incorporation of emission control devices
         4.     Combined systems.
               The variations in engine operating conditions include
engine load and speed;  intake air temperature and pressure; fuel injec-
tion timing; and fuel characteristics.   The effect of these parameters
on emissions and  fuel consumption is  discussed in Section 4.1.1.  The
engine component modifications are described in Section 4.1.2, includ-
ing changes in  the combustion  chamber geometry,  compression ratio,
valve timing, air  swirl, and fuel injection system.  The emission con-
trol devices evaluated to date include  exhaust gas re circulation; water
induction or injection; catalytic converters; thermal reactors; exhaust
gas scrubbers;  as well as incorporation  of a turbocharger.  These
particular devices are examined in Section 4.1.3.  Finally, Sec-
tion 4. 1.4 presents data from  a number of emission control systems,
consisting of combinations of various  control devices/techniques.
                                 4-2

-------
 4.1.1          Variations in Operating Conditions
 4.1.1.1       Engine Load and Speed
               The effects of varying engine load and speed on the HC,
 CO, and NO  mass emissions of a four-stroke,  turbocharged, divided-
 chamber diesel engine are illustrated in Figure  4-1 (Ref. 4-1).  As
 indicated in the figure, a reduction in engine load at a given speed tends
 to increase the  specific mass emissions of HC and NO ,  while reducing
                                                    JC
 CO.  Conversely, at constant load, a  reduction in engine speed is
 accompanied by a marked improvement in the specific mass emissions
 of HC and NO , and a substantial rise in CO, except for the very low-
 load regime where the trends are reversed.  Derating the engine from
 its rated operating point (285 bhp at 2200 rpm) down to 200 bhp and
 1500 rpm results in a reduction in NO  from about 4. 3 g/bhp-hr to
                                    H
 about 2. 5 g/bhp-hr.  This is accompanied  by a 60-percent reduction in
 HC and a 150  percent increase in CO.   As  discussed in Section 3.1.4,
 similar trends have been observed on other diesels, while  some
 engines indicate inverse relationships.  These differences  in trends are
 attributed to variations in the component design details and fuel injec-
 tion versus  engine load/speed schedules.
              Although engine derating appears  to be beneficial from a
 NO  reduction point of view, at least for some engines,  the related
   Jt
 investment cost increase in terms of dollars per maximum brake
 horsepower and the associated changes in  specific fuel consumption
 would have to be taken into account when considering this approach.
 4.1.1.2       Intake Air Temperature and Pressure
              Variations in the intake air temperature and pressure
have been shown to have some effect on the emissions and specific fuel
 consumption of diesel engines.  In general,  NO  is expected to decrease
with decreasing charge temperature.  This trend is attributed to the
 associated reduction  in the compression and local reaction temperatures
                                 4-3

-------
                       320
                       280
                       240
                       200
                     t 160
                       120
                       80
                       40
                                        i     r
                             NO, g/hr
                                     200
                         0 V400   800    1200   1600   2000   2400
                                   ENGINE SPEED, rpm
320
280
240
200
160
120
 80
 40
      I      I

      HC, g/hr
                  i      r
  0 V400    800    1200   1600   2000   2400
             ENGINE SPEED, rpm
                                           320
                                           280
                                           240
                                           200
                                         I
                                           160
                                           120
                                           80
                                           40
    I      1

    CO, g/hr
                                                                     300
                                                            300
0 "400    800    1200   1600   2000   2400
            ENGINE SPEED, rpm
     Figure  4-1.   Effect of speed and power output on emissions -
                     Caterpillar four-stroke precombustion chamber
                     diesel (Ref.  4-1)
                                        4-4

-------
and NO  formation rates.  Moreover, a reduction in charge temperature
       j£
tends to increase the ignition delay in the combustion chamber, which
can further reduce NO  .
                     H
              The primary effects related to intake air throttling are

mixture enrichment and reduction in the air mass flow rate. Based on

theoretical considerations and the available test data (Ref. 4-2),  it
appears that  little benefit could be derived  from this particular technique.

4.1.1.2.1     Intake Air Temperature

              The effect of intake air temperature  on the NO emissions
                                                           JC
of a turbocharged, open-chamber diesel engine, rated at 230 bhp, is
shown in Figure 4-2 (Ref. 4-2).  In these tests, the engine was oper-
ated at its  rated speed of 2100 rpm and the intake air temperature was
controlled  by means of a water-cooled heat exchanger installed between
the turbocharger and the engine intake manifold.  As expected, NO
                                                                jC
decreased  substantially with decreasing manifold temperature, and the

rate of reduction was essentially independent of engine load. Apparently,
                      i         i
                  O NO COOL ING WATER
                   (maximum air temperature 265 F)
                  x85°F COOL ING WATER
                  • 50 F COOLING WATER
                              100       150
                            POWER OUTPUT, bhp
         Figure 4-2.  Effect of intake-air temperature on NOX
                     emission of a turbocharged, open-cham-
                     ber diesel  engine at rated speed (Ref. 4-2)
                                 4-5

-------
cooling of the intake air had no detrimental effect on the CO, HC, and
smoke emissions  of this engine.  Also, fuel economy and odor remained
essentially unchanged.  With 50 F cooling water, the NO  emissions of
the engine were reduced by about 40 percent.  Although cooling to 50 F
might not be practical for most applications, it is evident that  this
approach is quite  effective in reducing NO  .  Particularly in stationary
                                        J\*
engines, where the volume constraints are less  severe,  incorporation
of a high effectiveness aftercooler might prove to be a cost effective
NO  abatement technique.
               Test data from a turbocharged, divided-chamber diesel
engine (No. 24 in  Table  3-5) indicate that changing the air temperature
from Z50 to 150 F resulted in a 20 percent reduction in NO , accom-
panied by a four-percent improvement  in specific fuel consumption.
The HC, CO, and smoke emissions had a tendency to increase  slightly
with decreasing intake air temperature.  At half load,  the data showed
considerable scatter,  but there is some indication that the effectiveness
of charge cooling  with respect to  NO  reduction diminishes in the part-
                                   Ji
load regime (Ref. 4-3).
               Similar NO  reduction rates were obtained by the Bureau
                         jC
of Mines on a turbocharged, open-chamber diesel engine (No.  16 in
Table 3-3) which was tested with  air intake temperatures of 250 and
150  F.  However, in this case  the indicated improvement in specific
fuel consumption was only about one percent.  Data published by
Bascom,  et al. (Ref. 4-4) indicate a reduction in NO  of about  30 per-
                                                  X.
cent by lowering the intake air temperature from 250 to  150 F.  Some-
what smaller improvements were reported by Wilson,  et al. (Ref. 4-5),
for single-cylinder, open-chamber,  and pre chamber diesel engines.
Reduction in the intake air temperature from 200 F to 100  F resulted
in a 10 to 15 percent improvement in NO ,  accompanied by a slight gain
                                       X.
in specific fuel consumption. At the same time, the smoke increased
                                 4-6

-------
somewhat in the prechamber engine, but remained unchanged in the
open-chamber engine.
              Data from a large turbocharged diesel engine, rated at
4300 bhp, are presented in Table 4-1 (Ref. 4-6). This particular
engine was operated at its design conditions (600 rpm and ZOO psi
BMEP) using both No. 2 diesel fuel and natural gas.  With diesel fuel,
reduction in the manifold temperature from 130  F (standard setting) to
100 F  resulted in a five-percent reduction in NO and a concomitant
1.5-percent improvement in  specific fuel consumption.  In the case of
natural gas, NO  was  reduced by ZO percent, whereas specific fuel
               n
consumption showed a 0.5-percent  increase.  Only small variations
were observed in the emissions of  CO and HC.
4.1.1.2.2     Projected Benefits
              Although intake air cooling is not a very powerful NO
                                                                 X.
reduction technique, it can be implemented rather easily, particularly
in the case of  stationary engines where  size considerations are  fre-
quently less significant.  Since specific fuel consumption tends to
decrease with intake cooling, this technique might prove to be particu-
larly attractive in conjunction with other control methods,  such as fuel
injection timing retard and EGR, to compensate for the loss in engine
efficiency generally encountered with these approaches.  The perfor-
mance characteristics of combined techniques are futher discussed in
Section 4. 1.4.
              It is concluded that a 0. 15 to 0.3  percent reduction in
NO can  be achieved per degree reduction in intake air temperature.
   A.
This is accompanied by a 0 to 0.04 percent improvement in fuel econ-
omy.   The effect of manifold temperature on HC, CO,  smoke,  and odor
appears to be  rather insignificant.
                                 4-7

-------
           TABLE 4-1.  EFFECT OF  INTAKE AIR COOLING ON THE EMISSIONS AND
                        SPECIFIC FUEL CONSUMPTION OF A COOPER BESSEMER
                        KSV-12 DIESEL ENGINE (Ref. 4-6)
Fuel
No. 2
Diesel
No. 2
Diesel
No. 2
Diesel
No. 2
Diesel
Natural
Gas
Natural
Gas
Natural
Gas
Natural
Gas
rpm
600
600
600
600
600
600
600
600
Load
Full
Full
Full
Full
Full
Full
Full
Full
Mani-
fold
Temp,
°F
130
100
130
100
130
100
130
100
Injection
Timing
Standard
Standard
-4°
-4°
Standard
Standard
-4°
-4°
Emissions, g/bhp-hr
HC
0. 13
0. 17
0. 21
0.20
5. 16
5. 28
3.22
3. 20
CO
3. 85
3.82
4.45
4.09
4. 50
4.25
7. 21
6.45
NOX
10.99
10. 54
9.31
8.71
8.96
7.27
8.41
6.63
BSFC,a
Btu
bhp-hr
6677
6583
6732
6636
6340
6377
6410
6444
Emission Ratios
HC
HC0
-
1. 3
-
0. 95
-
1.02
-
0.99
CO
C00
-
0.99
-
0. 92
-
0. 94
-
0. 89
NOX
NOXQ
-
0.96
-
0. 94
-
0. 81
-
0. 79
BSFCa
BSFCQ
-
0. 986
-
0. 986
-
1 . 006
-
1. 005
Brake Specific Fuel Consumption
Subscript zero refers to baseline conditions
oo

-------
4.1.1.3       Fuel Injection Timing
              Injection timing is probably the single most important
parameter affecting the emissions from diesel engines.  Early injection
timing tends to produce high combustion pressure rise rates and peak
temperatures, particularly when combined with high injection rates.
As a result, the formation of NO  in the combustion chamber increases.
                                A.
Conversely, retarded timing minimizes the  formation of NO . Gener-
ally, a timing retard has a small impact on other pollutant species
(except for  very late  timing) but degrades fuel economy (Refs. 4-1
and 4-4).
4.1.1.3.1     Single Cylinder Engine Data
              The effect of fuel injection timing variations on diesel
engine emissions and performance was evaluated by Pischinger and
Cartellieri,  using a single-cylinder, open-chamber research engine
with a  compression ratio of 16.2 and an inlet swirl ratio of 2. 6
(Ref. 4-7).   As indicated in Figure 4-3, timing retard resulted in
decreasing  NO  emissions at the expense  of some increase in CO,
              Jv
smoke, and specific fuel  consumption. Conversely, timing advance
caused sharply higher NO  and CO.  The smoke intensity decreased
somewhat with advanced timing, while HC remained essentially con-
stant.  With respect to the optimum setting, five-degree timing retard
resulted in  a 44-percent reduction in NO  , an 86-percent increase in
                                       JC
CO,  a 50-percent increase  in smoke,  and a  three-percent increase in
specific fuel consumption.  For three degrees  retard, NO  decreased
by 31 percent, while  SFC increased by 1. 1 percent, CO by 45 percent,
and smoke by 33 percent.  This indicates that  the first few degrees of
timing retard are particularly  effective from an SFC versus  NO
                                                             X
tradeoff point of view.  The magnitude of the effect is dependent upon
the baseline timing.
                                 4-9

-------
                   20
                 •f 15
                 Q.
                 O>
                 «/» 10
                 z
                 g
                 to
                 to
                 UJ
             r- t 0.45
       O I  o
       50  2
       ^ o  iL ^
         S  1L £0.40
         m
               u
               li.
               v>
               CD
                 0.35
                             2600 rpm             ,
                             FUEL DELIVERY 75 mrrT/CYCLE
                              EQUIVALENCE  RATIO 0.69
                                        NO,
                                         BMEP
                                   I
I
I
        120
        100
        80 !

        60
        40
                  m
                  '
                     30      25      20      15     10      0
                      DYNAMIC INJECTION TIMING, deg BTDC
   Figure 4-3. Effect of fuel injection on performance and emissions
               of a single cylinder research engine (Ref.  4-7)

              Test data reported by Shahed et al. (Ref. 4-8), Wilson
et al. (Ref. 4-5),  and Valdmanis and Wulfhorst (Ref. 4-9)  for single-
cylinder,  open-chamber diesel engines are summarized in Table 4-2,
indicating somewhat lower NO  reduction rates.
                                4-10

-------
    TABLE 4-2.  SINGLE CYLINDER DIESEL ENGINE EMISSIONS
                 AS A FUNCTION OF INJECTION TIMING
                 RETARD
Timing,
Degrees
BTDC
20
17
15
12
20
17
15
25
20
15
Timing
Retard,
Degrees
0
3
5
8
0
3
5
0
5
10
NO *
X
NO
X0
1.0
0.76
0.67
0.57
1.0
0.82
0.69
1.0
0.81
0.61
HC
HCo
-
-
-
-
-
-
-
1.0
0.60
0.23
CO
coo
-
-
-
-
-
-
-
1.0
0.85
0.71
Smoke
Smoke
0
-
-
-
-
-
-
-
1.0
1.27
1.33
SFC
SFC0
0
1.007
1.012
1.030
-
-
-
-
-
-
Reference
4-8



4-5


4-9


*
Subscript zero refers to standard timing.
4.1.1.3.2     Multicylinder Four Stroke Naturally
              Aspirated Diesels
              The effect of injection timing on the emissions of
medium-speed,  four-stroke naturally-aspirated,  open-chamber diesel
engines was investigated by Kahn et al. (Ref. 4-10)  and by the Bureau
of Mines (Ref. 4-3).  Figure 4-4 reproduces the data published by Kahn
for a 245 CID 4-cylinder engine. In these tests, the engine was oper-
ated at 2000 rpm (60 percent of rated  speed) and full load, and the injec-
tion timing was varied from the standard setting of  20 degrees BTDC
down to 0 degrees, using four discrete injection period settings  (Sys-
tems 1 through 4). As indicated in the figure, the NO  emissions and
                                 4-11

-------
to
        1.0
        0.8
     I  0-6
     •t
     in
     X

     1  0.4
     10
        0.2

                 I
                              \
                                    \
                                           \
                 20     15     10     5     0
                   INJECTION TIMING,  deg BTDC
                             a
                                                           120
                                100
                                 80
                                                            60
                                                      Oss  120
                           croo
                                                           100
                                                        f
                                                        o
                                                        u
                                                            20
                                                            10
                                      • SYSTEM 1, PERIOD - 20°

                                      O SYSTEM 2, PERIOD - 18°

                                      A SYSTEMS, PERIOD - 17°

                                      D SYSTEM 4, PERIOD - 15.5°
                                                                   20     15     10     5     0

                                                                     INJECTION TIMING, deg BTDC
             Figure 4-4.
Effect of injection rate and timing on gaseous mass emissions,
smoke,  and performance of a naturally  aspirated diesel

engine (Ref.  4-10)

-------
the power output capability of the engine decreased markedly with
increasing timing retard,  while smoke and specific fuel consumption
increased.  Not shown in the figure is HC, which remained essentially
constant over the range of timing and injection period settings.  Reduc-
ing the  injection period resulted in higher NO  emissions and lower
                                           ^C
SFC, CO,  and smoke levels.  At the standard injection period of
20 degrees, retarding the injection timing from  20 degrees BTDC to
five degrees caused a change  in NO  from 12 g/bhp-hr to 4 g/bhp-hr,
                                  jC
a reduction of 66 percent. However, this improvement was accom-
panied by an 18-percent loss in engine power (BMEP) and SFC  and a
substantial increase in CO and smoke.   Conversely, with the shortest
injection period of 15.5 degree,  retarding the  timing from 15 degree
(which is optimum for this setting as far as fuel economy is concerned)
to zero degree resulted in a reduction in NO  from 14.5 g/bhp-hr to
about 4.4 g/bhp-hr.  In this case,  the loss in SFC and power output is
only about  12  percent relative to the  standard  setting (System 1), and
the smoke  and CO emissions are reduced  also.
               Curves computed from test data published by the Bureau
of Mines (Ref. 4-3) are presented in Figures 4-5 and 4-6, showing
engine performance and mass emission parameters of a four-stroke,
naturally aspirated, open-chamber diesel as a function of injection
timing and load/speed settings.  The standard timing of this  engine  is
34 degree  BTDC and the data obtained at this point were utilized to
normalize  the parameters plotted in the graphs.
               Except for the  intermediate speed/full-load condition,
the specific fuel consumption is a minimum at the standard injection
timing.  For a given operating condition,  the exhaust smoke  intensity
varied very little in the retarded regime, but  increased  rapidly as
timing was advanced from the standard setting.  As expected, smoke
intensity increased with increasing load.  Also shown in Figure 4-5 is
the maximum power output capability of the engine as affected by timing,
                                4-13

-------
              O
              L.
              V)
              m

              O
              Q.
              i/>
              CQ
                 1.08
                 1.04
1.00
                 0.96LA-
                  40
             t   30
             O
             Q.
             O
             UJ
             Q.
 20

 10

  0

I.Or-
                 0.95
                 0.90
           Figure 4-5.
               O
             Q.
             I
        n 2200 rpm, FULL LOAD
        A 2200 rpm, HALF LOAD
        * 3000 rpm, FULL LOAD
        O 3000 rpm, HALF LOAD
       25      30       35       40

                INJECTION TIMING, deg
                                                        45
       Effect of injection timing on specific
       fuel consumption, peak opacity,  and
       maximum power  output of a naturally
       aspirated,  open-chamber diesel
       engine — Engine No. 7 (Ref.  4-3)
Retarding or advancing the timing degraded engine power somewhat.

For instance, five degrees retard caused a three-percent loss in power.

At full load, the HC emissions tend to decrease with increasing retard

and show a  substantial rise for advanced timing.  At half load, these

trends are reversed.   CO varied moderately  in the  retarded regime,
                                 4-14

-------
                        a 2200 rpm, FULL LOAD
                        A 2200 rpm, HALF LOAD
                        A oftAA	FULL LOAD
A 2200 rpm
* 3000 rpm, rwui. •.«««
03000 rpm, HALF LOAD
                              30       35       40
                          INJECTION TIMING, deg

          Figure 4-6.  Effect of injection timing on the emis-
                      sions of a naturally aspirated, open-
                      chamber diesel engine  - Engine No. 7
                      (Ref. 4-3)
but increased rapidly for advanced timing.  NO  shows a linear

relationship with timing. Five degrees  retard resulted in a 35-

percent reduction in NO .
                                4-15

-------
              NO  emission and performance data from another
                 jC
naturally aspirated engine tested by the Bureau of Mines (Engine No.  7
in Table 3-2) show a more rapid reduction in NO   with increasing  tim-
                                              jC
ing retard.   For example, 5 degree retard caused a  50-percent reduc-
tion in NO ,  accompanied by a 7.6-percent increase in SFC and a  7-
          Jt
percent loss in engine power output capability.  The HC and CO emis-
sions of this engine  showed considerable  data scatter,  but HC decreased
with increasing timing retard, whereas CO showed a tendency to increase .
These trends are in general agreement with the curves in Figure 4-6.
4.1.1.3.3    Multicylinder,  Four Stroke, Turbocharged Diesels
              A medium-speed, turbocharged, open-chamber die sel
engine, rated at 325 bhp-hr (Engine No. 16 in Table  3-3) was tested by
the Bureau of Mines with standard injection timing and with 3 degree
retard (Ref.  4-3). At  the retarded setting, the NO emissions at full
load and at half load were 25 percent lower than for zero retard.  This
reduction was accompanied by a 2. 7-percent increase in specific fuel
consumption and a one-percent  loss in maximum engine power.  At  the
same time,  the peak smoke opacity of the  engine increased from 8. 6 per-
cent to 12. 2 percent, whereas HCand CO showed only moderate variations .
              One manufacturer of medium-speed die sel engines has
provided test data showing a 30-percent reduction in NO  and a
four-percent loss in specific fuel consumption when operating with
5 degree retard.
              Data published by Cooper-Bessemer for a large-bore
die sel engine tested  with both No.  2 die sel fuel and natural gas are
listed in Table 4-3 (Ref. 4-6).  As indicated, retarding the timing  by
4 degrees resulted in a 15-percent reduction in NO for No. 2 die sel
                                                 Jt
fuel and a six-percent  reduction for  natural gas.  In both cases, the
specific fuel  consumption of the engine increased  by about one percent,
while the observed changes in HC and  CO were relatively small.
                                4-16

-------
             TABLE 4-3.  EFFECT OF 4U INJECTION TIMING RETARD ON EMISSIONS AND

                          FUEL CONSUMPTION OF A LARGE STATIONARY DIESEL

                          (Ref. 4-6)
Fuel
No. 2
Diesel
No. 2
Diesel
Natural
Gas
Natural
Gas
rpm
600
600
600
600
Load
Full
Full
Full
Full
Injection
Timing
Standard
-4°
Standard
-4°
Mass Emissions,
g/bhp-hr
HC
0. 13
0. 21
5. 16
3. 22
CO
3.85
4.45
4. 50
7.21
NOX
10.99
9.31
8. 96
8. 41
BSFC,a
Btu
bhp-hr
6677
6732
6340
6410
Emission Ratios
HC
HC0b
1. 61
0. 62
CO
C00
1. 16
0. 94
NOV
A
N°xO
0. 85
0. 94
BSFCa
BSFC0
1.01
1. 01
aBrake Specific Fuel Consumption
Subscript zero refers to standard timing
*>.
I

-------
               Test data from another large-bore, turbocharged, open-
chamber diesel engine are plotted in Figure 4-7 as a function of injec-
tion timing retard.  Comparison of these data with Table 4-3 indicates
that varying injection timing on this engine had a larger effect on NO
                                                                  Jw
than in the case of the Cooper-Bessemer engine.  For example, retard-
ing the timing by four degrees from the standard  setting caused a change
in NO  from 12.5 g/bhp-hr to 8.5 g/bhp-hr,  a reduction of 32 per-
     X,
cent.  However, the NO  improvement was associated with a
4.2-percent loss in specific fuel consumption.
4.1.1.3.4     Two Stroke Diesels
               The effect of injection timing variations on NO , CO,
                                                           Jt
and BSFC of a General Motors 6-71N diesel engine is illustrated in
Figure 4-8 (Ref. 4-11).  Since engine flow rate remained nearly con-
stant at each speed, the mass emissions were proportional to the con-
centrations  shown in the figure.
               For both  speeds, NO  decreased  with increasing timing
                                  J\.
retard, but increased with engine load due to the higher temperature
levels  in the combustion chamber.  At peak torque speed, NO  reached
                                                           Jx
a maximum at about 90-percent load and then decreased again,  indi-
cating  a deficiency in available oxygen.  At 2100 rpm,  NO  decreased
by about eight percent per degree injection retard, and at 1200 rpm,
the rate of reduction was about seven percent per degree.  Within the
range of acceptable smoke levels, the effects of timing changes on HC
and CO were not very significant. However,  when timing was retarded
beyond these levels, CO increased sharply at full loads,  while HC in-
creased at light loads, indicating incomplete combustion.  As shown in
Figure 4-7,  timing retard has a detrimental effect on SFC.  For exam-
ple,  five degrees retard resulted in a 3. 5-percent loss in SFC.  These
rates of change are in reasonable agreement with the previously dis-
cussed four-stroke engines.
                               4-18

-------
                      2    4    6    8    10   12

                       INJECTION TIMING RETARD, deg
                                             14
              Figure 4- 7.
Effect of injection
timing retard on the
NOX emissions and
specific fuel con-
sumption of a large
diesel engine
       INJECTOR TIMING
STANDARD   3.4* RETARD   5.0* RETARD
                2100 rpm
 ISOOrpm
                                Figure 4-8.
20  40 60  80 100 120 140 160 180 ZOO 220
      BRAKE HORSEPOWER
                   Effect of injection tim-
                   ing on NOX emissions
                   and specific fuel con-
                   sumption of a two-
                   stroke diesel engine
                   (Ref.  4-11)
                             4-19

-------
4.1.1.3.5     Proposed NOx Reduction Correlations
               The NO  reductions determined from the engine tests
                      Jt
discussed in the preceding sections are summarized in Figure 4-9 as a
function of timing retard and  specific fuel consumption degradation.
The medium-speed,  multicylinder engines are distributed almost uni-
formly over the NO  versus SFC data band, whereas the single-
                   X.
cylinder engines are concentrated in the upper half of the band and the
large, low-speed engines in the lower half.  Conversely, in the NO
versus timing  chart, most of the single-cylinder engine data fall within
the lower half  of the  band, while the majority of the medium-speed,
multicylinder engines is included in the upper half.
               The available test data relating the effect of injection
timing retard and the variations in HC,  CO,  and smoke are shown in
Figure 4-10.  In some engines, HC and CO increase markedly with
increasing timing retard, but remain constant or decrease in others.
   eo
   60-
 § 40-
 Q
 ui 20
 (X
         \     I    I     I
          A SINGLE-CYLINDER ENGINES
          O MEDIUM-SPEED ENGINES
          a LOW-SPEED ENGINES
                       CORRELATION
                       AVERAGE
             466
             INCREASE IN BSFC, %
                           10
                                12
             4    6     8    10
             TIMING RETARD, deg
                                12
Figure 4-9.  NO  reduction in die-
             sel engines vs specific
             fuel consumption and
             timing retard
                                 4-20

-------
               A SINGLE CYLINDER ENGINES
               o MEDIUM-SPEED ENGINES
               n LOW-SPEED ENGINES
                 468
               TIMING RETARD, deg
                                    Figure 4-10.
                                   HC, CO and smoke
                                   variations vs tim-
                                   ing retard
                                 10
Smoke tends to increase as timing is retarded, but the magnitude of
the increase seems to vary widely.
4.1.1.4
Fuel Effects
               Diesel engine emission test work conducted to date,
using different distillate type fuels, has indicated that variations in the
fuel can produce small changes in the emissions.  However,  the poten-
tial improvements related to variations in fuel properties cannot always
be predicted and are often affected by the  design of the engine,  as well
as its operating conditions (Ref. 4-12).  In some cases, changes in cer-
tain fuel properties affect the exhaust emissions by altering the engine
operating characteristics rather than by changing  the combustion proc-
ess directly.  Thus, a change in engine design or  its mode of operation
might achieve the desired result more easily than would a revision of
fuel quality.
                                 4-21

-------
               On the other hand, utilization of heavy fuels might cause
higher NO  emissions as a result of the conversion of the fuel-bound
nitrogen to NO.  Test data indicate  that up to 70 percent of the fuel-
bound nitrogen is converted in steam boilers (Ref. 4-13), but informa-
tion is lacking regarding the conversion in diesel engines.
4.1.1.4.1     Cetane Number and Composition
               The effect of fuel cetane number on the HC, CO, and
NO  emissions of a number of open-chamber,  naturally aspirated and
   Jt
turbocharged diesel engines is shown in Figure 4-11 (Ref. 4-14).  As
indicated, the HC, CO,  and NO  emissions decrease  moderately with
                              3C
increasing fuel cetane number, particularly in the case of the two
naturally aspirated engines.
              Similar results were  reported by Cummins for a single-
cylinder experimental diesel engine which was operated with three dif-
ferent fuels having cetane  ratings of 40, 52, and 56, respectively
(Ref. 4-8). As shown in Figure 4-12, the NO   specific mass emissions
decrease with increasing cetane number over the whole range of fuel
injection timing settings investigated.  Increasing the cetane number
of the fuel resulted in a  more favorable tradeoff between fuel consump-
tion and NO  emissions.
           x
              The influence of fuel  composition on the emissions of
several four-stroke,  naturally aspirated and turbocharged open-
chamber, and two-stroke, open-chamber,  air-scavenged diesel engines
has been  studied  by the Bureau of Mines (Ref.  4-2).   The properties of
the seven test fuels used in the program covered wide ranges:  specific
gravity  varied between 0.81 and 0.87, the aromatic content varied
between 14 and 44 percent, the cetane number varied between 50 and
42, and the sulfur content  varied between 0.04 and 0.4 percent.  The
emission variations obtained with these different fuels were quite small,
except for CO which,  at  full-load conditions, increased markedly with
                                4-22

-------
12
10
8
6
4
2
0
i i i i i I
v "*>L ^**"^>V-8 TURBO-
XJ ° >• 4 ° CHARGED
\ A ^A
S^. *. 1-6 NATURALLY
°£j>«^D^ 0 ASPIRATED
~~T) V-8 NATURALLY
D ASPIRATED
* ^. ^ ^ Q PROTOTYPE 1 -6
O "~ " ~~~
A 1 1 1 1 1 1
is
12
10
»• 8
X 6
4
2

v 30 35 40 45 50 55 60 0
CETANE NUN
14
12
10
r
0 6
O
4

2
1 1 1 1 1 1
„ _
6 ./PROTOTYPE 1-6
\
\\ /V-8 NATURALLY
vy ASPIRATED
V-8 TURBO\ » O 1-6 NATURALLY
.CHARGED \0$^0 v /ASPIRATED -
A 1 1 ~~l ~ ™"l 1 1
v 30 35 40 45 50 55 6
ilBER CETANE NUMBER
II II
-
r Y
K X%w 0 V-8 NATURALLY
"PC ^^ ASPIRATED
A ^^^/ ^^^ ^^D
^^ — — "S 1-6 NATURALLY
D ASPIRATED
^"•OO- -». _ ___ o PROTOTYPE 1 -6
V-8 TURBOCHARGED
A 1 1 I 1 1 1







                30
                     35   40   45   50
                       CETANE NUMBER
                                      35   60
Figure 4-11.  Cetane number effects on the emissions
              of naturally aspirated and turbocharged
              diesels (Ref.  4-14)
                         4-23

-------
          10
        1*
        5
        o

        t 6
        O
        lil
        Q.
        
-------
Conversely,  an increase in fuel volatility has resulted in reduced smoke
intensity (Ref.  4-16).
4. 1.1. 4. 2     Smoke Suppressant Additives
               A number of smoke  suppressant fuel additives have been
evaluated by several investigators.  Basically, these additives can be
divided into two groups. One group is of the detergent type and is
designed to maintain a clean fuel injection system.  The second group
consists of metal-based materials which reduce the ignition tempera-
ture of the carbon and promote its oxidation.
               Barium-based fuel additives have been used successfully
by a number  of investigators to reduce diesel smoke intensity (Refs.  4-4
and 4-17 through 4-19).  Test data reported by Bascom et al. (Ref. 4-4)
for  a four-stroke, naturally aspirated engine operated with a commerci-
ally available barium-based fuel additive indicate a reduction in smoke
                           o                    o
intensity from  about 9 mg/ft  to less than 1  mg/ft when 0.5 percent
(by  volume) of  the additive was used.  However, after  30 hours of test-
ing  with the treated fuel, the HC emissions had increased considerably,
whereas CO and NO  remained unchanged.  The increase in HC was not
due to  an increase in deposit buildup in the engine, converse to what
was  observed by General Motors (Ref. 4-20).
               An extensive study of the effects of fuel additives on
diesel  smoke was performed by Saito and Nabetani (Ref. 4-18),  using
two- and four-stroke open-chamber  and prechamber engines.  The
The  barium-containing additives proved to be most effective,  resulting
in a reduction in smoke opacity between 35 and 60 percent.  In general,
the  effectiveness of the additive increased with increasing barium con-
tent.  In these  tests, the additives had essentially no effect on HC and
showed a tendency to increase CO and to decrease NO  .
                                                    Jt
               Tests conducted earlier by the Bureau of Mines support
these findings. However,  after 100 hours of testing,  the smoke level
                                4-25

-------
 on one of the engines used in this program started to increase and the
 level obtained with the untreated fuel was approached after 100 addi-
 tional hours of testing.  This increase was attributed to deposit
 buildup in the  combustion chamber and the injection nozzles (Ref.  4-19).
               Approximately 70 percent of the barium added to the
 fuel is exhausted from the engine in the form of harmless barium sul-
 fate, whereas part of the  remainder consists of soluble compounds of
 which some  are toxic.  Although the barium-containing additives are
 quite effective in reducing diesel smoke levels,  they are not recom-
 mended by most engine manufacturers because of adverse effects on
 engine durability.
 4.1.1.4.3     Odor Effects
               Although fuel composition often is suggested to have  a
 prominent influence on diesel odor,  the supporting experimental evi-
 dence  is not strong.  This conclusion is supported by tests conducted
 by the Bureau of Mines (Ref. 4-19).  In these tests, a two-stroke diesel
 engine was operated with seven different diesel fuels which had differ-
 ent sulfur  and aromatic contents.  Although some differences in exhaust
 odor level were noted,  the variations could not be reliably related to
 any one fuel  property.
 4.1.2          Component Modifications
 4. 1. Z. 1        Combustion Chamber
               Modifications in the design of  the combustion chamber
 are known to influence the fuel-air mixing and combustion processes
occurring in diesel engines,  and, as a result, the emission and spe-
 cific fuel consumption characteristics of these engines  are altered.
The design parameters evaluated by a number of engine manufacturers
include the geometry of the combustion chamber,  compression ratio,
                                4-26

-------
valve timing,  and air swirl.  These factors are briefly discussed in the
following paragraphs.
4.1.2.1.1     Chamber Geometry
              The effect of variations in the ratio of combustion
chamber bowl diameter to cylinder bore on the emissions of a single-
cylinder, open-chamber research engine was evaluated by Pischinger
and Cartellieri (Ref. 4-7).  Increasing the  bowl-diameter-to-bore ratio
from 0.51 to 0.62 resulted in moderate reductions in NO   and HC with-
                                                      x
out affecting specific fuel consumption. However, the exhaust smoke
level of the  engine was  substantially higher for the larger bowl diam-
eter.  The observed  changes are attributed to better fuel and tempera-
ture distributions in the combustion chamber achieved with the larger
bowl design.
              Results  from a combustion chamber shape study pro-
vided by one engine manufacturer are presented in Figure  4-13,  show-
ing specific fuel consumption and NO   and HC emissions over the
13-mode test cycle.  Again, NO decreases with  increasing bowl size.
Also, NO  and HC emissions and fuel consumption vary slightly with
piston head  clearance (Ref. 4-14).
              Based on the available  data, it is concluded that varia-
tions in the  bowl-diameter-to-bore ratio  can have a  small  but distinct
effect on NO , HC, and engine efficiency.  Although the magnitude of
            ?c
the effect is expected to vary from engine to engine, optimization of
this parameter can be beneficial as far as the tradeoff between emis-
sions and fuel consumption is concerned.
              In prechamber engines, the  ratio of prechamber volume
to total cylinder volume, the  size of the orifice or orifices  connecting
the two chambers,  and  the  internal shape of the prechamber have been
found to have some effect on the NO  and HC emissions (Ref. 4-1).
                                  j£.
Test data indicate that  the NO  and smoke emissions from these  engines
                                 4-27

-------
  4.8
  4.6
  4.4
  4.2
1 4.0
  3.6
  3.4 -
    o  VCU
                   n     i      i     r
                      NOX , g/bhp-hr (13-MODE)
             J	I	I	I
                                                I      I     I      I
i   0.08  0.10   0.12   0,
 PISTON-HEADr in.
      4.8
j	I  LA_I	i	L
.14   0.160 VQ.040706OTC
                                                                J	I
    08   0.10   0.12   0.14   0.16
PISTON-HEAD, in.
                     4.6
                     4.4
                   oT 4.2
                   < 4.0
                   o
                   § 3-8
                     3.6
                     3.4
                                                  i      r
                                 J	i	i	i
                        0  V0.04  0.06   0.08  0.10   0.12   0.14   0.16
                                   PISTON-HEAD, in.
  Figure  4-13.   Effect of bowl diameter and  piston head clearance
                   on diesel engine emissions and specific fuel con-
                   sumption (Ref.  4-14)
                                     4-28

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increase with increasing prechamber size, whereas specific fuel
consumption shows some  improvement (Ref.  4-5).  These trends are
plausible, because more air is burned  in the  prechamber as  its size
increases, thus reaching  higher flame  temperatures and NO  levels.
                                                         Jt
4.1.2.1.2     Compression Ratio
              The effects of engine compression ratio on the NO  and
                                                              jC
HC emissions of an open-chamber  diesel engine are shown in Fig-
ure 4-14.  The HC emissions declined  significantly with increasing
compression ratio, while NO  decreased slightly up to a compression
                           Jt
ratio of about 16:1  and remained almost constant beyond that point.
Since high compression ratios result in higher mechanical engine loads,
the practical limit  in compression  ratio is dependent upon the BMEP
and the engine structure but is about 18:1 for large diesel engines.
Also, the small clearance volumes associated with high compression
ratios inhibit the fuel-air mixing process, hence causing a loss in
specific fuel consumption.
              These trends are in agreement with the data reported by
Wilson, et al. ,  for a single-cylinder prechamber diesel engine
(Ref. 4-5).  However, Wilson's test data for  an open-chamber  engine
indicate increasing NO  emissions as the compression ratio  of the
engine was raised from 14 to  17.
4.1.2.1.3     Valve  Timing
              Valve  timing, which affects cylinder scavenging and
combustion  efficiency, has  been shown to have  some influence  on the
emissions from diesel engines.  In the past,  the engines were  designed
for minimum fuel consumption and thermal loading of internal  parts.
However, in future engines, a lower scavenging efficiency might be
selected  to provide some  internal exhaust gas recirculation (EGR) for
the purpose of reducing NO (Ref.  4-1).  The effectiveness of this
                                4-29

-------
                   S.6
                   4.8
                   4.0
                  I 1.1
                  Z
                  o
                   2.4
                   1.6
                   0.8
                         I     I    I    i
      'J-VSir
                                           NO
                              1
                                  1
                                      1
                                          1
                                              1
                             13   14    15    16   17   18
                              COMPRESSION RATIO
                 Figure 4-14.  Effect of compression
                               ratio on NOX and HC
                               emissions
                               (Ref.  4-14)
approach would have to be compared with other potential emission
control techniques,  such as  timing modifications, EGR,  or water
injection, before its  application would be warranted.
4.1.2.1.4
Air Swirl
               Air motion in the combustion chamber has important
effects on the combustion process occurring in diesel engines and hence,
emissions.  In general, increasing the air swirl reduces the exhaust
smoke intensity, but tends  to increase NO .  This trend is attributed
                                         X.
to the associated improvement in fuel-air mixing which increases the
reaction rates in the combustion chamber, causing a rise in the local
temperatures and NO  formation processes (Refs. 4-4 and 4-21).
                     5v       ~ ~
               Air swirl is particularly important in naturally aspi-
rated engines, which often are operating at the smoke-limited
                                 4-30

-------
air-fuel ratio.  Conversely, turbocharged diesels appear to be less
sensitive to air motion,  probably because of the high excess air  ratio
used in these engines.  This is  illustrated in Figure 4-15, showing an
inverse relationship between smoke and NO .   The open-chamber en-
                                          X
gine used in these tests was modified to reduce the smoke level by (1)
increasing the air swirl,  (2) varying injection timing, and (3) turbo-
charging (Ref.  4-4).  At a given smoke density, the lowest NO  emis-
                                                            X
sions were achieved with turbocharging.   Conversely, increasing air
swirl resulted in unfavorable NO  versus smoke  tradeoffs.
              Similar results were obtained by Khan et al. (Refs. 4-10
and 4-21) and Wilson et al. (Ref. 4-5) on single-cylinder, open-chamber
diesels.  Increasing the swirl ratio resulted in lower specific fuel con-
sumption,  smoke, and NO  levels and made the smoke vs. timing
curve less sensitive to timing changes.  Since  higher swirl accelerates
combustion,  it might be desirable to operate the  engine with retarded
timing, which could result in a net reduction in NO  without sacrificing
                                                 X.
fuel economy (Ref. 4-21).
   20
   18
   '2
 o
               HIGH SWIRL
         246
           SMOKE DENSITY, mg/cu ft
Figure 4-15.  Effect of air swirl
              and turbocharging on
              smoke and NOX emis-
              sions of open-chamber
              diesel engines
              (Ref. 4-4)
                               10
                                 4-31

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4.1.Z.Z       Injection System
4.1.2.2.1     Injection Characteristics
               The design and operation of the fuel injection system has
an important effect on the emissions from diesel engines — in particular,
HC and smoke (Ref. 4-4).  To minimize these emission species, the
injection system must be capable of providing a uniform injection pulse
and a properly distributed fuel spray without  impingement on the piston
or cylinder walls.  Injector cup and  spray geometry and injection
duration are important design variables, and  their effects on emissions
are intimately related to the overall design of the combustion chamber
(Ref. 4-4).
4.1.2.2,2     Sac Volume
               Tests conducted by General Motors on a number of two-
stroke diesel engines indicate that the sac volume at the tip of the
needle valve was responsible for most of the  relatively high HC emis-
sion obtained with one particular fuel injection system (Refs. 4-11 and
4-22).  In this design, the fuel contained in  the sac volume is boiled off
after completion of the main injection event and is then discharged into
the combustion chamber too late  in the cycle  to be fully oxidized.
Reduction of the sac volume from the original 3.5 cubic millimeters to
0.5 cubic millimeters resulted in a reduction in HC from about
350 ppmC to about 90 ppmC. Further reduction to about 40 ppmC was
achieved by means of an experimental needle  valve design which covers
the tip of the injector orifice in the closed-valve position.  The afore-
mentioned reduction in HC was obtained without sacrificing fuel econ-
omy,  NO , and CO (Ref. 4-11).   Similar improvements were realized
         X.
by the Electromotive Division of General Motors on large locomotive
diesels.
                                 4-32

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4.1.2.2.3
Injection Rate
               The rate of injection has been shown to have a marked
effect on the NO  emissions of diesel engines.  This  is illustrated in
Figure 4-16,  showing the NO   concentration in the exhaust of a two-
stroke diesel, for three fuel injection rate settings (Ref. 4-11).  In the
high-load regime, a 20-percent reduction in NO  was realized by
                                               JC
increasing the rate  of injection from 5.7 to 8. 3 cubic millimeters per
degree, and this effect was essentially independent of injection timing.
At the  same time, CO was reduced slightly,  indicating improved com-
bustion efficiency, while HC remained constant.  Similar results were
reported by Khan et al. (Ref.  4-10).
               The principal drawback related to the  use of shorter
injection periods  is the requirement of higher injection pressures,
causing higher pushrod and cam loadings.  As a result,  engine dura-
bility might be adversely affected (Ref. 4-11).
                                RATE OF INJECTION
                     5. 7 mm /deg    7. 2 mmVdeg     8. 3 mm /deg
                 1200
                El 000
                &
                 800
                  600
                £400
200
 0
                       I
                              2100 rpm
             (W.O.T. END OF INJECTION® 0° T.D.C.)'
            I   I   I   I   I   I   I   I   I  I
                    0 20  40  60  60 100 120 140 160 180 200 220 240
                               BRAKE HORSEPOWER
                Figure 4-16.  Effect of injection rate on
                              NOX emissions of a two-
                              stroke diesel engine
                              (Ref.  4-11)
                                 4-33

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4.1.2.2.4     Orifice Size and Spray Angle
               Injection nozzle  size affects the formation of the fuel
spray and the fuel-air mixing process in the combustion chamber, and
can have a significant impact on the emissions and performance of
diesel engines. Test data from one engine indicate that a 20 to 40 per-
cent reduction in NO was achieved by increasing the orifice diameter
                    Jt
from 0.0055 inch to 0.0065 inch (without reducing the number of ori-
fices).  However, this improvement was accompanied by higher spe-
cific fuel consumption (three percent), CO emissions (70 to 100 percent),
and smoke.  These trends indicate that the observed reduction  in NO
was due primarily to lower combustion efficiency (Ref. 4-11).  Reduc-
ing the number of injection orifices from eight to  six,  without changing
the total orifice area, resulted in some reduction of NO  at the expense
                                                      Jt
of slightly higher fuel consumption, smoke, and CO emissions.  Again,
this is attributed to lower combustion efficiency caused by incomplete
fuel-air mixing in the cylinder.
               Similar effects were reported by Khan et al. (Ref. 4-10)
and Wilson et al. (Ref. 4-5). The available data show that an optimum
number and size of injection holes might exist for each particular
engine design, hence permitting a tradeoff between the emissions and
fuel economy.
               Correlations  showing the effects of injection hole num-
ber and spray  angle on diesel engine emissions and fuel consumption
•were provided by one manufacturer (Ref.  4-14).  Although the effects
are not very large, optimization of these design parameters is  desir-
able  to minimize engine emissions.  Apparently, the  spray angle giv-
ing minimum fuel  consumption and HC causes NO to reach  a
maximum.
                                4-34

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4.1.2.2.5     Fumigation
              Adding fuel to the intake air is known as fumigation.
This technique has been used in the past to reduce the pressure rise
rate in the combustion chamber and the noise of the engine.  In the tests
reported by Bascom et  al. (Ref. 4-4), fumigation with kerosene
increased the CO, HC,  and smoke emissions of a naturally aspirated,
open-chamber diesel engine while lowering NO  slightly.   With propane
                                             jC
fumigation, the NO  reduction effectiveness was improved, particularly
                  .X
at part loads.
4.1.2.2.6     Combined Effects
              A number of the injector modifications discussed above
were incorporated into  a single injector which was then tested  in a
General Motors two-stroke diesel engine.  The modifications included
valve-covered orifices,  higher injection rates, larger tip orifices plus
retarded timing.   Over the 13-mode test cycle,  the following improve-
ments in emissions were achieved relative to the standard engine set-
tings.  HC was reduced  from 0. 6 g/bhp-hr to 0. 2 g/bhp-hr, NOx
was reduced from 9.0 g/bhp-hr to 5.0 g/bhp-hr, while CO increased
from 3.6 g/bhp-hr to 4.8 g/bhp-hr.  These changes were accom-
panied by a 2. 5-percent loss in fuel economy and an increase in the
exhaust smoke level by about one Bosch number.
              The effect of injection system modifications on the emis-
sions of a two-stroke, naturally aspirated, open-chamber  diesel, rated
at 280 bhp (engine No. 28 in Table 3-6), was investigated by  the Bureau
of Mines (Ref. 4-3).  The  engine  was operated at  its  rated  speed
(2100 rpm), using the standard injectors  and a modified injection sys-
tem incorporating low sac volume, high injection rates,  and  a  special
timing schedule that provided for variable beginning and constant end-
ing of injection.  The test  results indicated substantial reductions in
HC and NO  with the new injector, particularly in the low power regime.
                                4-35

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 These improvements were accomplished without compromising SFC,
 CO, and smoke.   For example, near rated load, the specific mass
 emissions of NO  and HC were reduced by about 25 and 60 percent,
 respectively.   At half load, the corresponding improvements were
 50 and 80 percent.
 4.1.3         Emission Control Devices
 4.1.3.1        Exhaust Gas Recirculation
 4.1.3.1.1     General
               Tests conducted by a number of investigators have indi-
 cated that incorporation of exhaust gas recirculation is an effective
 means to lower the NO  emissions from open-chamber as well as
                      jt
 divided-chamber diesel engines.  Generally, this reduction is  associ-
 ated with some loss in engine performance and an increase in  smoke.
               The mechanism involved in reducing NO  is related to
                                                     Jt
 the intake charge dilution obtained with EGR, resulting in lower  com-
 bustion temperatures and lower oxygen concentrations.  Both effects
 have a tendency to inhibit the formation of NO  during the combustion
                                            X,
 process,  particularly at the lower air-fuel ratios associated with full-
 load operation.   Best results, in terms of NO  reduction and associated
                                            j£.
 engine performance loss, have been achieved when the exhaust gases
 were cooled before recycling. Tests with hot EGR have resulted in
 substantially higher fuel consumption and smoke emissions.
 4. 1.3. 1. Z    Experimental Programs
              The effects of EGR  on the NO  and smoke emissions of
                                          x
 a naturally aspirated diesel engine are illustrated in Figure 4-17
 (Ref. 4-4).  In these tests,  the EGR flow was cooled to the temperature
 of the intake air.  As indicated in the figure,  substantial reductions in
 NO  were achieved with moderate  EGR flow rates.  For instance,
utilization of ten  percent EGR resulted in  a reduction in NO  of this
                                4-36

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                                        SMOKE
                               I
                                                    35
                                                     25
                                                     15
                                                       o
                               10      15
                              EGR FLOW, %
20
25
               Figure 4-17.  Effect of EGR on NOX and
                             smoke emissions of a
                             naturally aspirated open-
                             chamber diesel engine
                             (Ref. 4-4)
engine from about 7 g/bhp-hr to about 4 g/bhp-hr.  However,  the
smoke density (soot) of the engine exhaust increased in this case from
             o                 o
about 8 mg/ft  to about 22 mg/ft .   The increase in smoke is the
direct result of a reduction in oxygen availability in the  combustion
chamber, which also is responsible for the observed loss in engine
power output capability and the increase in CO emissions.  These
problems can be alleviated by adding a turbocharger.
              Pischinger (Ref.  4-7) conducted EGR tests on a 610 CID
multicylinder, four-stroke, open-chamber, low-swirl diesel engine
which was operated at an intermediate speed of 1400 rpm and at part-
load conditions (52 psi BMEP).  NO  decreased almost linearly with
increasing EGR flow rate, at the expense of some increase in smoke
and fuel consumption.  HC was rather independent of EGR, while CO
showed no change up to 30 percent EGR, but increased rapidly for
higher EGR rates.   The use of 10 percent EGR resulted in a 20 percent
reduction in NO at practically zero loss in fuel  consumption. With
20 percent EGR, NO was diminished by about 30 percent,  while the
                                4-37

-------
specific fuel consumption suffered a loss of two percent.  No information
is available regarding the temperature of the EGR flow used  in these
tests.  Since the reported effect of EGR was less than observed by other
investigators using cooled exhaust,  it is conceivable that uncooled EGR
was employed in this particular investigation.
              Emission versus EGR curves computed from the data
published by the Bureau of Mines (Ref. 4-3) for a naturally aspirated,
open-chamber diesel (engine No.  7  in Table 3-2) are presented in Fig-
ure 4-18.  In this  case, the EGR  flow was extracted near the muffler
and was cooled before induction into the intake manifold  of the engine.
As indicated, the NO  emissions  decreased rapidly with increasing
EGR, particularly at low EGR flows.  As expected,  EGR was more
effective under high load conditions because of the higher combustion
temperatures prevailing at the higher load points and the lower oxygen
concentration in the exhaust gases.  Also, EGR appears to be more
effective at higher engine speeds.  As indicated, utilization of ten per-
cent EGR  resulted in a reduction  in NO  of about 60  percent at full load
and 45 percent at half load.  Under these conditions, the specific fuel
consumption increased by about three percent and the peak smoke
intensity increased by about 80 percent.  At half load, the CO emissions
remained essentially constant, whereas a substantial increase in CO
was observed at full load for EGR flow rates above about eight percent.
The HC emissions show some  reduction with increasing  EGR flow rate,
particularly under full-load operating conditions of the engine.  Con-
versely, the peak smoke intensity increased markedly with EGR,  while
the specific fuel consumption showed only a small increase.  The  power
output capability of the engine  deteriorated mildly with increasing EGR
flow.
              EGR test data from a four-stroke, turbocharged, open-
chamber engine rated at 230 bhp are presented in Figure 4-19 (Ref. 4-2),
As shown, ten percent EGR resulted in a 35-percent reduction of NO  .
                                4-38

-------
tr.
O
UJ
 . 0.5
O

*x

z
  0.2
                     D 1700 rpm, FULL LOAD'
                     A 1700 rpm, HALF LOAD
                     » 2800 rpm, FULL LOAD
                     O 2800 rpm, HALF LOAD
                       A
                      10

                    EGR,
                                         20
 Figure 4-18.  Effect of EGR on naturally
               aspirated open-chamber
               diesel engine mass emis-
               sion and performance,
               Engine No.  7 (Ref. 4-3)
                   4-39

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  _  6
   "
   8
   o
   z
           I      I     I      I
        O NO EXHAUST GAS RECIRCULATION
        D 10% ECR
        A 15% EGR
           50    100    150
              POWER OUTPUT, bhp
                           200
                                250
Figure 4-19.  Effect of EGR on
              NOX emissions of a
              turbocharged, open-
              chamber diesel
              engine at rated speed
              (Ref. 4-2)
 Apparently, there was little effect on the fuel consumption and power
 output capability of the engine.  Further reduction in NO  was achieved
 with higher EGR rates, but the  performance  of the engine suffered
 greatly, particularly in terms of low-speed power capability.  Accord-
 ing to Figure 4-19, EGR is more effective  at the higher power levels
 because the oxygen content in the recirculated gas decreases with
 increasing load.
               NOx emission data from a divided-chamber,  turbo-
 charged diesel engine indicate a nearly linear reduction of NO  with
 increasing EGR  flow rate.  With ten percent EGR, NO  was  reduced  by
                                                     JC
 about 40 percent, and with 15 percent EGR, the  average reduction was
 about 65 percent (Ref. 4-1).
               The effectiveness of EGR in  an air-scavenged, two-
 stroke diesel engine, rated at 148 hp  and 2100 rpm, is shown in Fig-
 ure 4-20.  At 80 hp output,  the use of 25 and  40 percent EGR resulted
 in NOx reductions of about 70 percent and 90 percent, respectively.
 However,  the engine had to be derated by as much as 40 percent to
 stay within the smoke limits of the non-modified engine.  Also, HC
 and CO emissions increased substantially with this high EGR rate  and
this required incorporation of a  catalytic converter in the  exhaust  sys-
tem (Ref. 4-2).
                                 4-40

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     	1	
     O NO ECR
     O 25% ECR
   - • 40% EOR
       25   50    75    100
            POWER OUTPUT, bhp
                           125
                                ISO
       Figure 4-20.  Effect of exhaust gas
                     re circulation on NOx
                     emission of an air-
                     scavenged, two-
                     stroke diesel,  at
                     rated speed
                     (Ref.  4-2)
               Single-cylinder,  naturally aspirated diesel engine data
are presented in Figure 4-21  (Ref. 4-23). In these tests, the EGR flow

was cooled to 80° F before induction into the engine intake,  and the fuel
flow rate was held constant, resulting in an air-fuel ratio of 24.8 at
zero EGR and 19.9 at 20 percent EGR.  Again,  NO  decreased approxi-
                                                  3t
mately linearly with increasing EGR flow rate,  whereas smoke  shows

an inverse relationship.   Up to about 12 percent EGR, the specific fuel
consumption of the engine remained constant, and HC was not affected

at all by the use of EGR.  Utilization of ten percent cooled EGR  resulted
in a 40-percent reduction in NO .
                               x
                  £ 0.330
                  f 0.320
                  £
                  ~ 0.310
                  u
                  So 0.300
                      10
                   v»
                   i  s
                   a >.
                      -
                   S
ISFC
                            4    8    12    16   20
                          EXHAUST IN INTAKE CHARGE, %
                  25

                  20 _
                     U)
                  15 ?

                  10

                  5

                  .0
                                                 Is
           Figure 4-21.  Effect of EGR on performance and
                          emissions of a single-cylinder
                          naturally aspirated engine (80 psi
                          bmep, cooled EGR at 80°F)
                          (Ref. 4-23)
                                 4-41

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              Experimental results  from a single-cylinder,
open-chamber diesel engine, operated at 2100  rpm and 143 psi BMEP,
indicate a 25 percent reduction in NO  when using 10 percent EGR
(Ref.  4-8).  With 20 percent EGR,  the observed NO  reduction was
                                                 J\.
about 45  percent.
              Exhaust gas  recirculation tests  over the 13-mode cycle
were  conducted by Ricardo  and Company, using an 855 CID, divided-
chamber diesel engine (Ref. 4-24).  In these tests,  the EGR flow was
cooled to 122 F,  and the rate of EGR admission was varied over the
cycle, employing 20 percent EGR at loads up to 50 percent, ten percent
EGR at 75 percent, and zero EGR at full load.  Under these conditions,
the NO  emissions  over the cycle were reduced from 7. 1 g/bhp-hr to
       ?t
5. 3 g/bhp-hr, an improvement of about  25 percent.
              One  manufacturer of large stationary diesel engines has
reported that a 45-percent reduction in NO  was  achieved in an explora-
tory test program,  using  15 percent EGR.  In these tests,  the  specific
fuel consumption showed no change with EGR,  but the smoke intensity
increased by about  100 percent.  The observed NO  reduction is in
reasonable agreement with the test data discussed above.   Another
manufacturer feels that a 20-percent reduction in NO might be
achieved without a loss in specific fuel consumption. However,  a
reduction in NO  by 40 percent would be accompanied by a 10-percent
loss in specific fuel consumption.
4.1.3.1.3     Potential Problem Areas
              Although basically feasible, most  manufacturers of
heavy-duty diesel engines feel that  a number of potential problem areas
would have to be resolved before incorporation of EGR  into stationary
engines could become a reality.  These problems include corrosion and
deposit buildup in the EGR circuit,  and in the case of turbocharged
engines,  fouling of  the compressor and intercooler.  Furthermore, the
                                4-42

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long-term effect of EGR on engine wear and lube oil contamination is
currently not known, and systematic tests would be required to deter-
mine the magnitude  of these effects.  In any case, one manufacturer
feels that engine maintenance costs would definitely increase with the
application of EGR,  and another expressed concern that the problems
with EGR in diesels might be more severe  than in gasoline engines
because of the higher sulfur content in diesel fuel relative to gasoline.
4.1.3.1.4    Projected EGR Characteristics
              The available test data indicate  significant variations in
the observed effects of EGR on  the  emission and performance  charac-
teristics of diesel engines.  This is illustrated in Figure 4-22, showing
data bands containing all the available test data discussed in Sec-
tion 4. 1. 3. 1.2.   When using these curves, consideration must be given
to the fact that the effects  of EGR depend on many factors including
engine type, operating speed and load, and EGR temperature.   Cur-
rently, there are insufficient data available to derive accurate corre-
lations for all these parameters.  Cummins has developed a fairly good
correlation for medium  speed diesels, using the oxygen concentration
in the intake manifold as the correlating parameter (Ref. 4-14).
              Based on the data presented here,  it is concluded that
EGR is an effective technique to reduce NO  emissions from diesel
engines.  However, the full benefits of EGR cannot be realized without
some sacrifice  in engine performance and some increase  in CO and
smoke emissions.
4.1.3.2       Water Addition
4.1.3.2.1    General
              Like EGR,  water added to the intake air or injected into
the cylinder of diesel engines,  either separately or in the form of water-
fuel emulsions,  acts as a charge diluent, resulting in lower compression
                                 4-43

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   1.08
              5        10       15
                EGR FLOW RATE, %
                                      20
Figure 4-22.  Effect of EGR on diesel
              engine emissions and
              specific fuel consumption
                 4-44

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and combustion temperatures and oxygen concentrations, hence lower
NO  formation rates.
   x
               Water inducted with the intake air is superheated during
the compression stroke of the piston and distributed fairly homogene-
ously throughout the combustion chamber.  In this case, the tempera-
ture at the start of combustion is not  significantly different from the
no-water  case,  since the mass  of water inducted is generally small,
relative to the mass of air.   As a result, the  ignition characteristic
of the charge  does not change much relative to operation without water.
               Conversely, when water is injected directly into the
engine cylinder as part of a fuel-water emulsion, vaporization of the
water and mixing with the air takes place locally. Since water  has  a
high latent heat, vaporization of the water causes a significant reduc-
tion in the local temperature, resulting in an increase in the ignition
delay. Therefore, to  achieve satisfactory operation  of the engine,  the
fuel injection  timing must then be advanced, which negates the expected
benefits of the water (Refs. 4-4  and 4-9).
4.1.3.2.2     Water Induction
               The effect of water induction on diesel engine emissions
has been studied by a number of investigators using open-chamber  and
divided-chamber, naturally-aspirated, and turbocharged engines.
               Results obtained by Cummins (Ref. 4-9) on a single-
cylinder,  open-chamber diesel engine are presented in Figure  4-23,
showing NO and HC concentration and smoke measurements as  a func-
tion of the water flow rate for two different engine speeds. Also shown
in this figure  are data obtained with emulsified fuel,  which are  further
discussed  in Section 4.1.3.2.3. As indicated, the  NO   and smoke
                                                     Jt
emissions decreased markedly with increasing water flow rates,  and
the observed effects were nearly independent of engine speed.  For
example,  for  a water flow rate  of 40 percent (by volume) of the com-
bined water plus fuel flow, NO   was reduced by about 25 percent and
                                 4-45

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    •o 4
                  800   2600
                 rpm   rpm
                  O     •   EMULSIFIED FUEL
                  A     A   INDUCTED WATER
                 OPT. TIMING
                 F/A = 0.052
                 20    30    40

                  WATER ADDED,
Figure 4-23.  Effect of inducted and emulsified water
              on the HC,  smoke, and NOX emissions
              of an open-chamber diesel engine
              (Ref. 4-9)
                        4-46

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the smoke level was reduced by about 20 percent.  However, under
these conditions, HC increased by about 150 percent, whereas little
change was  observed in CO.  The indicated mean effective pressure
decreased slightly with increasing water flow rate.  It should be noted
that the air-fuel ratio was held constant during these tests and the
injection timing was set at the optimum value for each test point.
              Tests with water induction, conducted by Ricardo and
Company (Ref. 4-24) on a multicylinder, open-chamber diesel engine,
resulted in  substantial NO  reductions,  with a concomitant small
                         jc
increase in  CO and HC,  and a slight loss in power.  In these tests,  fuel
injection timing was set for optimum performance and the engine was
operated over the 13-mode test cycle.  Emission measurements were
performed for water-to-fuel-flow ratios of 0.5 and 1.0.  The test data,
listed in Table 4-4,  indicate that a 50-percent reduction in NO  was
achieved for a water-to-fuel-flow ratio of 1.0.  Again,  the NO  reduc-
                                                            .X
tion was approximately proportional to the amount of water added.
              The effect of water injection into the intake system of a
four-stroke, turbocharged, open-chamber diesel engine is illustrated
in Figure  4-24,  showing NO  concentration as a function of water vapor
                           Jw
content in the intake air (Ref.  4-2).  As indicated, the NO reduction
        TABLE 4-4.
EFFECT OF WATER INDUCTION ON THE
EMISSIONS OF AN OPEN-CHAMBER DIE-
SEL ENGINE (13-MODE CYCLE DATA)
Water
Fuel
Ratio
0. 1
0. 5
1.0
Emissions, g/bhp-hr
CO
3. 5
4. 3
4. 7
HC
0. 5
0. 5
0.6
N0x
15. 5
11.4
7.8
Power
Loss ,
%
_
2
4
NOX
Reduction,
%
_
26
50
                                4-47

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     0      1.2       3      4       5       6
      WATER VAPOR CONTENT OF INTAKE AIR, % by mass
Figure 4-24.  Reduction of NOX emission as a function
             of water induction— turbocharged, open-
             chamber diesel engine  (Ref. 4-2)
                       4-48

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observed on this engine is approximately proportional to the water flow
rate.  Utilization of four percent water (by mass) resulted in a NO
                                                                Jt
reduction of about 50 percent at 2100 rpm and about 30 percent at
1200 rpm.  Under full-load conditions, the water flow rate used in
these tests was approximately equal to the fuel flow rate.  Apparently,
water addition resulted in no significant detrimental effects on CO, HC,
and  specific fuel consumption.
               Figure 4-25 presents NO  emission data for a turbo-
                                       5C
charged, divided-chamber engine which was operated at maximum
torque and 60-percent torque, using water-to-fuel-flow ratios between
0 and 2.5 (Ref. 4-1).  For a water-to-fuel-flow ratio of  1.0, the
observed reduction in NO was about 70 percent  at the  high torque con-
                         Jt
dition and about 65 percent at the intermediate torque point.  As  shown
in the figure,  water induction is particularly effective for water-to-
fuel-flow ratios up to about 1.0.  Beyond that point, the  gains diminish,
while the probability of water condensation in the intake  system
increases.  In these tests, the water was added ahead of the turbo-
charger.  Subsequent analyses  conducted by Caterpillar  indicate,
  1400
  1200
                                    Figure 4-25.
        0.5   1.0   1.5   2.0   2.5
           WATER/FUEL MASS RATIO
                               3.0
Water induction ver-
sus NO emission —
prechamber, turbo-
charged diesel,
2200 rpm (Ref. 4-1)
                                 4-49

-------
however,  that adding  the water just  ahead of the intake  valve  or
directly into the cylinder itself would  alleviate potential problems re-
lated to water condensation and "settling out"  of the water in the in-
take system (Ref. 4-1).  Steam injection into the engine has been con-
sidered as an alternate method which  might provide a better and more
effective distribution of the water throughout the  combustion chamber.
However, the heat of vaporization of the water would  not be available
in this case and this would tend to lower the effectiveness  of this
technique.
              The effects of water  induction on NO , soot, and indi-
cated specific fuel consumption of a single-cylinder,  open-chamber and
a divided-chamber diesel engine were determined by Wilson et al.
(Ref. 4-5).  Again, NO  decreased  with increasing water flow rate,
                      JC
particularly for water-to-fuel-flow  rates up to about  1.0,  while SFC
showed very little change.  However,  with increasing water induction,
soot formation showed a tendency to increase somewhat.  For both
engine types, operation at water-to-fuel-flow rates of 0.5 and 1.0
resulted in  NO  reductions  of 35 percent and 60 percent, respectively.
              jf.
The data indicate some variations with engine speed and air-fuel ratio
(load), but the observed trends are  not consistent.
              Water induction data reported by Cooper-Bessemer for
a large stationary diesel engine are listed in Table 4-5 (Ref.  4-6).  In
these tests, the water was injected  into the intake air manifold of each
cylinder and the  engine was  operated  at rated speed and load conditions,
using either No.  2 diesel fuel or natural gas.   In the tests with diesel
fuel, a water flow rate of 0.5 gallons  per minute was used, which
corresponds to a water-to-fuel-flow rate of about 0.17. This  resulted
in a seven-percent reduction in NO  .  Assuming a linear relationship
                                 X.
between NO  reduction and water addition, a 40-percent reduction in
NO  is predicted for  this engine when operated at a water-to-fuel-flow
ratio of about 1.0,  compared with about 50 to 60 percent for the
                                4-50

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         TABLE 4-5.  EFFECT OF WATER INDUCTION ON COOPER BESSEMER KSV-12
                      DIESEL ENGINE EMISSIONS AND FUEL CONSUMPTION
Fuel
No. 2
Diesel
No. 2
Diesel
Natural
Gas
Natural
Gas
rpm
600
600
600
600
Load
Full
Full
Full
Full
Water
Rate,
gpm
0
0. 5
0
0.3
Mass Emissions,
g/bhp-hr
HC
0. 13
0. 16
5. 16
6.48
CO
3.85
4.25
4. 50
3.39
NOX
10. 99
10.20
8.96
8. 35
BSFC,a
Btu
hp-hr
6677
6643
6340
6374
HCb
HCQ
-
1.23
-
1.26
CO
co0
-
1. 10
-
0.75
NOXQ
-
0. 93
-
0. 93
BSFCa
BSFCQ
-
0. 995
-
1.005
aBrake Specific Fuel Consumption
Subscript zero refers to operation without water
01

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smaller, higher-speed engines discussed above.   The data listed in
Table 4-5 show relatively small variations in HC, CO, and SFC, and
this might be due to data scatter.
4.1.3.2.3      Emulsions
               Single-cylinder engine tests were  conducted by Cummins
to determine the effect of emulsified fuel on diesel engine emissions
(Ref.  4-9).  These tests showed that the emulsified fuel had a marked
effect on the combustion process, and ignition timing advance was
required to  compensate for the increasingly larger ignition delays
obtained with increasing water content in the emulsion.  As a result,
the NO  emissions increased with increasing water content,  as shown
      ji,
in Figure 4-23. Also, the HC emissions increased substantially with
increasing water content.
4.1.3.2.4      Projected Characteristics
               The available data on water injection into the intake sys-
tem of diesel engines  are summarized in Figure  4-26,  showing the
percentage reduction in NO  as a function of the water-to-fuel mass
                          X.
ratio.  As indicated, there is a large engine-to-engine variability in
the data which  might be due to variations in certain engine design and
operating parameters. However, in most cases,  there is insufficient
information available  to establish correlations between the NO  reduc-
tion and water  flow rate for the different engine  categories.  Moreover,
there is a lack of data relating water induction rate and HC,  CO, smoke,
and specific fuel consumption.
4.1.3.2.5      Potential Problem Areas
               Although water injection appears to be an effective tech-
nique to reduce the NO  emissions from diesel engines, most manu-
                      X
facturers are concerned about potential problem  areas directly related
to the use of water. These include  corrosion and wear of intake system
                                4-52

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   90
   80
   70
   60
 O 50
 S
 Q
 Ul
 a
 if
S 40
   30
   20
   10
       ® VALDMANIS (A/F = 19.2), open chamber
       O RICARDO (13-mode), open chamber
       O MARSHALL, turbo, open chamber
       A BOSECKER, turbo, divided chamber
       7 WILSON, single cylinder
       • COOPER BESSEMER, turbo, open chamber
        0.2
            0.4   0.6  0.8   1,0
            WATER/FUEL MASS RATIO
                              1.2
                                      Figure 4-26.
                                                   NOX reduction ver-
                                                   sus water to fuel
                                                   mass ratio
                                   1.4
components,  such as valves  and manifolds, as well as fouling of the
water injection nozzles.   To minimize corrosion and deposit buildup,
one manufacturer suggests the use of distilled water.  Other difficulties
include the requirement of a water flow control system to meter the
flow in proportion to the  intake  air flow rate, and a water pressuriza-
tion system to overcome the high intake manifold pressure in turbo-
charged engines.  Also,  the  water must be protected from freezing,
preferably without the addition of alcohol, which would probably
increase the HC emissions of the engine (Ref. 4-9).  Since a portion of
the inducted water may find its  way into the engine crankcase, more
frequent oil changes might be required if water injection would be
incorporated.  Currently, there is insufficient information available to
allow a comprehensive assessment of the long-term effects of water
injection on the performance and durability characteristics  of diesel
engines.
                                  4-53

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4.1.3.3       Catalytic Converters
4.1.3.3.1     General
               During the past few years, considerable efforts have
been directed by the automotive industry and other organizations toward
the development of catalytic converters for the control of exhaust emis-
sions from spark ignition engine powered automobiles. Although cata-
lytic performance degradation with mileage accumulation has been a
serious problem area in the past,  a number of catalytic systems have
now reached an advanced state of development and have been shown to
be quite effective in reducing the HC and CO exhaust emission levels
from automobiles. However, progress has been rather slow  in the
development of practical NO  reduction catalysts.  The NO  catalysts
                           A,                             X,
currently considered by the automobile  manufacturers are dependent
upon the availability of adequate  amounts of reducing  species  in the
engine exhaust, such as CO or H?  to convert NO to N~ + CO-,  or to
N? -I- H_O.  Since CO is practically nonexistent in diesel exhaust, the
currently known reduction catalysts are not considered to be effective
for diesels.
               It has been  suggested to produce the required amounts of
CO and/or H2 by means of a methane  or propane burner installed in the
exhaust system or by adding ammonia to the exhaust gases (Ref. 4-1).
Although theoretically feasible, there is insufficient data available at
this  time to permit a meaningful assessment of these  approaches.  In
any event, the  added engine operating cost and complexity resulting
from the use of these techniques and the requirement  of subsequent
oxidation of any excess CO in the exhaust would have to be weighed
against other potential NO  abatement methods.
                         Ji
4.1.3.3.2     Experimental Results
               To date, a limited amount of exploratory catalyst test
work has been  performed on diesel engines.  For example, Cummins
                                4-54

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has run a commercially available catalyst, which was specifically
designed for diesel applications, in one of its naturally aspirated, open-
chamber diesel engines (Ref. 4-4).  In these tests,  CO was  reduced by
about 90 percent and HC was reduced by about 50 percent, while no
change was  observed in NO .
              The HC and CO reductions reported by Caterpillar
(Ref. 4-1) for a catalyst installed in the  exhaust system of a prechamber
diesel engine are presented in Figure  4-27. At full load,  the HC and
CO conversion efficiencies are over 90 and 80 percent,  respectively.
However, with decreasing load, the effectiveness of the catalyst
decreases as a result of the  lower exhaust gas temperatures and the
associated reduction in the HC and CO oxidation reaction rates.  Also
shown in Figure 4-Z7 are test  data obtained by the Bureau of Mines
(Ref. 4-3) from a  naturally aspirated, open-chamber diesel engine
which was fitted with two  Universal Oil Products (UOP)  catalysts.  As
indicated, the trends in HC and CO reduction versus engine  load are in
reasonable agreement with the Caterpillar data, but the magnitude of
                                     BUREAU OF MINES   ~1
                                     CATERPILLAR
                               	 HC
                               	CO
                     I
I
                     20       40       60       80
                    PERCENTAGE OF MAXIMUM POWER
        100
     Figure 4-27.  Effect of catalysts on diesel engine emissions
                                4-55

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the observed reductions are higher in the case of the UOP catalysts.
According to Ref. 4-3, incorporation of the catalysts resulted in a sub-
stantial reduction in aldehyde  emissions and exhaust odor intensity
while the smoke level remained essentially unchanged.  With catalysts,
the engine back-pressure increased by  a factor of 2, relative to the
standard  configuration,  and this caused a small loss in the power out-
put capability of the engine.
               Similar results were obtained by the  Bureau of Mines
on another naturally aspirated, open-chamber engine (Engine No. 7 in
Table  3-2), equipped with two Engelhard noble metal catalysts,  and by
Pischinger and Cartellieri (Ref. 4-7).  Again, incorporation of the
catalysts had no effect on NO  and smoke, but resulted in some  reduc-
                            X.
tion of the exhaust odor  intensity.
               A number of base metal  and noble metal catalysts were
tested by Southwest  Research Institute in  a city bus  which was operated
over its normal duty cycle (Ref. 4-25).  Some of the catalysts were
experimental and some were "off the shelf" but intended for different
applications.  In several cases, installation of the catalyst resulted in
lower emissions of HC,  CO, odor,  and  certain partially oxygenated
materials.  No beneficial effect on  smoke was detected in any of these
tests.
               It should be noted that the results presented above for a
number of catalyst installations were achieved with  fresh catalysts and
do not include potential catalyst degradation effects  due to accumulation
of test time.
4.1.3.4        Thermal Reactors
               A thermal reactor is a chamber replacing the conven-
tional exhaust manifold system of the engine and is sized and configured
to increase the residence time of the exhaust  gases  and permit further
chemical  reactions,  thus lowering the HC and CO emissions. Devices
                                4-56

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of this type have been successfully applied in automotive spark
ignition engines, but are not considered to be very effective in diesel
engines.  At most operating conditions, the temperature of the diesel
exhaust and the concentration of HC and CO are believed to be too low
to sustain the HC and CO oxidation reactions.   More importantly, NO  ,
which represents the major pollutant specie in  diesels, is  not sub-
stantially altered in thermal reactors.
4.1.3.5        Exhaust Scrubbers
               Exhaust scrubbers, which have been used in mining
vehicles,  are not considered feasible  for application in stationary
diesel engines.   Testing of scrubbers conducted by one engine manu-
facturer indicates some reduction in HC and particulate emissions.
However, the most toxic species — CO and NO  — remained unchanged
                                            Jt
in these tests.
4.1.3.6        Turbocharging
               The power output of naturally aspirated diesels is fre-
quently limited by excessive amounts  of smoke  which is formed when
operating at low  air-fuel ratios.  Incorporation of a turbocharger
increases the air flow rate of the  engine and permits operation at
higher air-fuel ratios.  As a result,  the smoke-limited power cutoff
point of the engine moves to higher levels.  With turbocharging, the
manifold temperature increases,  causing higher combustion tempera-
tures and NO  formation rates.  These trends can be counteracted to
some degree by means  of an inter cooler installed between  the com-
pressor of the turbocharger and the engine intake manifold.
              The effect of turbocharging on a family  of 404 CID four-
stroke medium-swirl, open-chamber  diesel engines is illustrated in
Figure 4-28,  showing NO  , brake specific fuel consumption, and smoke
                        X.
levels at rated speed as a function of brake mean effective pressure
(power output) (Ref. 4-23).  Incorporation of a  turbocharger increases
                                4-57

-------
        1.
       ^•0.450
       .a
         0.350
       ffl
       X
       Ul
       o
       z
20
           15
       g Q.
       2| 10
       LJ
        X
       O
              I      I       I       I      I
               RETARDED TIMING EFFECT
            A 8° ON INTERCOOLED ENGINE
            O 6° ON TURBOCHARGED ENGINE
     NATURALLY
     ASPIRATED
          TURBO
          CHARGED
                                                         HO
                           1 **
                         8  ? °
                             m
                         n  -x
                         0  Q X
                 60
             80
100
120
140
160
                       RATED SPEED BMEP, Ib/in.'
        Figure 4-28.  Effect of turbocharging and inter cooling
                      on the emissions and specific fuel con-
                      sumption of a family of open-chamber
                      diesel engines (Ref. 4-23)
the power output capability of the engine and results in better specific
fuel consumption and lower exhaust smoke levels.  With an intercooler,
NO  can be  further reduced to the levels of the naturally aspirated
   X.
engine while maintaining lower SFC and exhaust smoke.  As  shown in
Figure 4-28, retarding the injection timing of the turbocharged engine
by six degrees increased the specific fuel  consumption to the level of
the naturally aspirated engine, but reduced NO  from about 12 g/bhp
                                             3t
to 7 g/bhp-hr.   This  represents  a reduction of about 12 percent  rela-
tive to the naturally aspirated  engine.  With intercooling and eight
degrees retard, the NO  emissions were further reduced to about
                                4-58

-------
5. 3 g/bhp-hr, while maintaining the specific fuel consumption of the
naturally aspirated engine.  This corresponds to a 35-percent reduction
in NO  relative to the naturally aspirated engine. In general, timing
     j£
retard tends to increase the smoke emissions, but the rate of change
observed on this engine family was rather small.
              Similar results are presented in Figure 4-29,  showing
curves computed from test data published by the Bureau of Mines
(Engine No. 8 in Table  3-2) (Ref. 4-3).  The emission and SFC values
plotted are referenced to the corresponding values obtained without
turbocharging (subscript 0).   As indicated,  turbocharging results in a
reduction in SFC and CO, particularly in the high power regime, but
causes higher emissions of NO   and HC. The effect of engine speed on
SFC,  HC, and CO is negligible,  but speed does affect NO  to some
                                                      X,
degree.  Although not shown in the  figure, the addition of the turbo-
charger  resulted in a reduction in the peak smoke opacity of the engine
from 20  to  about 2 percent.
              Tests conducted by Pischinger on open-chamber and
prechamber engines (Ref. 4-7) indicate  a 20-percent increase in NO
with turbocharging.  However, incorporation of an after cooler brought
the NO   emissions of the turbocharged engine back to the level obtained
      H
without turbocharging.
              Based on these data, it is concluded that incorporation
of a turbocharger  and intercooler,  combined with injection timing
retard, represents an effective technique to reduce  smoke and NO
                                                               3C
emissions without incurring a loss  in specific fuel consumption relative
to operation without a turbocharger.
4.1.4         Emission Control Systems
              The emission and fuel consumption characteristics of
several combinations of the emission control devices/techniques
described in Sections 4. 1. 1 through 4. 1. 3 were evaluated by  a number
                                4-59

-------
O
u.
   1.1
   1.0
   0.9
          O 1550 rpm
          A 2100 rpm
     4r
   1.5

8  1.0
o
<->  0.5
   i.8
  o
ST..4
O
z
1.0
   4

  0
               TT
                                       O A
               50       100      150
                     POWER OUTPUT, bhp
200
250
  Figure 4-29.   Effect of turbocharging on specific fuel
                consumption and mass emissions
                (Engine No.  8) (Ref.  4-3)
                        4-60

-------
of organizations including the Bureau of Mines (Refs. 4-2 and 4-3),
Cooper-Bessemer (Ref. 4-6),  and Cummins Engine Company
(Ref. 4-26).  The majority of the  Bureau of Mines tests was  conducted
over the 13-mode, heavy-duty engine cycle, with the remainder con-
sisting of constant-speed data  covering a wide range of engine load
conditions.  The  test data reported by Cooper-Bessemer are from one
of its large production engines operated at  rated speed and load on both
No.  2 diesel fuel and natural gas.  The Cummins data represent
13-mode projections for three  different emission control systems
applied to present production engines.  The data are briefly discussed
in the following sections.
4.1.4.1        Thirteen-mode Data
               Specific mass emissions of HC, CO, and NO  from a
                                                         JW
230-bhp, four-stroke, turbocharged, open-chamber diesel engine oper-
ated over the 13-mode cycle are listed in Table 4-6.  The engine was
equipped with several emission control systems  consisting of combina-
tions of intake  air cooling, EGR,  and water injection.  Minimum NO
                                                                 n
mass emissions of 6.4 g/bhp-hr (52-percent reduction) were achieved
on this engine with 15 percent EGR plus intake air cooling and with
water induction at a  rate of six percent (by mass of total intake flow)
plus intake air cooling.  The variations in CO and HC obtained in these
tests follow the trends observed with the  individual control techniques
discussed in Sections 4.1.1 through 4.1.3.
               Table 4-7 summarizes the emission and specific fuel
consumption results obtained by the Bureau of Mines for a number of
production engines (Ref. 4-3).   These engines,  identified in Tables 3-2
through 3-7, were operated over the 13-mode cycle and were equipped
with experimental emission control  systems consisting of various com-
binations of injection timing modifications, EGR, intake  cooling, modi-
fied  injectors,  and oxidation catalysts.
                                4-61

-------
  TABLE 4-6.
EFFECT OF EMISSION CONTROL SYSTEMS ON THE

EMISSIONS OF A TURBOCHARGED, OPEN-

CHAMBER DIESEL ENGINE- 13 MODE CYCLE

(Ref.  4-2)


Configuration


Standard
Intake Air
Aftercooled
50 °F
185°F
Intake Air
Aftercooled and
Cooled Exhaust
EGR (185°F)
10% EGR rate
15% EGR rate
Intake Air
Aftercooled
(185°F) and
Water Injection
1% water by
mass
2% water by
mass
6% water by
mass

Carbon
Monoxide,
g/bhp-hr

4. 5


3. 4
3. 1




3.8
6.3




2.2

2. 3

3. 1

Hydro-
carbon
Calculated
as CH2,
g/bhp-hr
1.9


2.0
1.8




1.2
1.2




2.0

2. 1

1.4

Nitrogen
Oxides
Calculated
as NC>2,
g/bhp-hr
13. 3


9.9
12. 1




8.7
6.4




11.5

10.6

6.4


Percentage Change


CO

-


-25
-31




-15
+40




-51

-49

-31


HC

-


+5
-5




-37
-37




+5

+ 10

-26


NOX

-


-25
-9




-35
-52




-13

-20

-52

aReferenced to standard engine
              Results from Engine No. 7 show that NO  can be
                                                    JC

reduced from the 8.3 g/bhp-hr level of the standard engine to about 4.2


g/bhp-hr by combining three degrees retarded injection timing and


ten percent EGR.  This corresponds to a 50 percent reduction in NO .
                                                                Jt

However,  a seven percent loss in fuel consumption was incurred under


these conditions. Using only five percent EGR  resulted in a 40 percent
                               4-62

-------
TABLE 4-7.
SUMMARY OF EMISSIONS AND FUEL CONSUMPTION FOR BASELINE
AND COMBINATION OF PARAMETERS TESTS FOR DIESEL ENGINES

Control
System

1


2


3



1
2
3
4

1
2

Description of Engine
Adjustment and
Accessory Hardware
Baseline, Engine No. 7d
Standard Timing, 10% ECR,
Englehard Catalyst,
(ECR cut off at full load)
3' Retarded Timing, 5% ECR,
Cnglehard Catalysts,
(EGR cut off at full load)
3' Retarded Timing, 10% ECR,
Cnglehard Catalysts,
(ECR cut off at full load)
Baseline, Engine No. 24d
Standard Timing, 10% EGR,
No Aftercooling
Standard Timing, 10% EGR,
150'F Aflercooling
5' Retarded Timing, 10% EGR,
No Aftercooling
5* Retarded Timing, 10% EGR,
150'F Aftercooling
Baseline, Engine No. 28
Experimental Injectors,
Standard Timing
Experimental Injectors,
3.4* Retarded Timing
13-Mode Cycle
Emissions,
g/bhp-hr
CO
5.93

1.73


1.89


1.93

0.83
1.20
1.02
1. 58
1.68
7.94
5.45
9.78
HC
2.90

1.22


1.78


1.79

0.26
0.28
0.20
0.29
0.64
1.62
0.72
0.88
NO2
8.27

6.24


4.97


4. 18

4.93
3. 13
2.98
2.78
2.44
15.9
10.5
7.91
BSFC.a
Ib
bhp-hr
0.445

0.454


0.453


0.476

0.422
0.421
0.414
0.443
0.430
0.476
0.492
0.491
Aldehydes,
g/bhp-hr
0.23

0. 16


0.29


0.20

0.082
0. 028
0.021
0.058
0. 107
0. 088
0.059
0. 125
Odor
Intensity
Dl UnitsO
4. 3

3.6


3.8


4.0

4.6
1.9
2.8
4.5
5.9
4.2
3.4
4.3

Lug -Down
Smoke,
% Opacity
13.3

16.6


17. 5


14.0

6.9
17. 5
20.9
10.0
6.5
2.8
4.5
18.2
Percentage Change
CO
-

-71


-68


-67

_
+45
+23
+90
+ 102
_
-31
+23
HC
-

-58


-39


-38

_
+8
-23
+ 11
+ 146
.
-55
-45
N0x
-

-25


-40


-49

_
-36
-39
-44
-50
.
-34
-50
BSFC3
-

+2.0


+ 1.8


+7.0

_
-0.2
-1.9
+5.0
+ 1.9
.
+ 3.4
+3.2
Aldehydes
-

-30


+26


- 13

_
-66
-74
-29
+30
_
-33
+42
Odor
-

-12


-7


-2

_
-58
-39
-2
+28
_
-19
+5
Smoke
-

+25


+31


+5

.
+ 154
+204
+45
-6
_
+61
+550
aBrake Specific Fuel Consumption
bQdor intensity in diesel intensity (DI) units determined with samples diluted 400:1 for engines 15 and 19, and 100:1 for all other engines
cReferenced to baseline values
dSee Tables 3-2 through 3-7

-------
TABLE 4-7 (continued)

Control
System

1
2
3
4
5
6



1


2


3



Description of Engine
Adjustment and

Baseline, Engine No. 6
Standard Timing. 10% ECR
(ECR cut off at full load)
Standard Timing, 15% EOR
(ECR cut off at full load)
2* Retarded Timing, 10% ECR
(EGR cut off at full load)
2" Retarded Timing, 15% ECR
(EGR cut off at full load)
Standard Timing, UOP Catalysts
Standard Timing, 10% EGR,
UOP Catalysts,
(EGR cut off at full load)
Baseline, Engine No. 16
3° Retarded Timing,
Special Pump and Nozzle Kit,
200"F Aftercooling
3° Retarded Timing,
Special Pump and Nozzle Kit,
150°F Aftercooling
3" Retarded Timing,
Standard Pump and Nozzles,
150°F Aftercooling
13-Mode Cycle
Emi ssions ,
g/bhp-hr
CO
5.93
5.71
6.01
5.68
6.05
0.68

0.68

4.48

3.70


3.45


3.76

HC
4.41
3.22
2. 85
4. 12
3.79
0.53

0. 55

2.62

1.87


1.76


2.60

NOZ
7.41
6.41
6.00
5.95
5.30
8.47

6. 80

15. 1

10.4


10.3


8.82

BSFC.a
Ib
bhp-hr
0.504
0.498
0.512
0.510
0. 520
0.514

0.520

0.490

0.502


0.507


0.495

Aldehydes,
g/bhp-hr
0.23
0. 17
0.23
0.33
0.33
0.046

0.065

0. 14

0. 12


0. 13


0. 18

Odor
Intensity,
DI Unitsb
4. 5
4.8
5. 1
4. 7
5.4
2. 6

2.3

5.9

5. 1


5.0


7.4


Lug-Down
Smoke,
% Opacity
12. 5
11. 7
1 1. 7
10. 1
10. 1
14.0

13.4

7.6

7. 7


6.6


7.4

Percentage Changec
CO

-4
+ 1
-4
+2
-89

-89



-17


-23


-16

HC
-
-27
-35
-7
-14
-88

-88

_

-28


-33


-1

NOX
-
-13
-19
-20
-28
+ 14

-8

_

-31


-32


-41

BSFCa
-
-1.2
+ 1. 6
+ 1.2
+3. 2
+2.0

+3.2

_

+2.4


+3. 5


+ 1.0

Aldehydes
-
-26
0
+43
+43
-80

-72

_

-14


-7


+29

Odor
-
+ 7
+ 13
+5
+20
-42

-49

_

-13


-15


+25

Smoke
-
-6
-6
-19
-19
+ 1Z

+7

.

+ 1


-13


-2

Brake Specific Fuel Consumption
Odor intensity in diesel intensity (DI) units determined with samples diluted 400:1 for engines 15 and 19. and 100:1 for all other engines
dSee Tables 3-2 through 3-7

-------
reduction in NO ,  accompanied by only a two percent penalty in fuel
consumption.  A pair of Engelhard noble metal catalysts incorporated
in the engine exhaust reduced CO by about 70 percent and HC by 38 to
58 percent. The observed variations in aldehydes, odor intensity, and
peak smoke levels are quite moderate.  Control System No. 2, which
consisted of three degrees timing retard,  five percent EGR,  and Engel-
hard catalysts,  appeared to be the most cost-effective configuration
for this particular engine.  It reduced NO  by 40 percent at the ex-
pense of only a  1.8 percent loss in fuel consumption.
              Test data from Engine No. 24 indicate that the Emission
Control Systems No.  2 and No.  4 are most effective from a NO  emis-
sion reduction and SFC point of view.  In the case of System No. 2, NO
                                                                    Ji.
was reduced by  about 40 percent, with a concomitant loss in SFC of
about two percent.  System No.  4 resulted in a 50-percent reduction in
NO , but the specific  fuel consumption increased by about two percent.
It is estimated that a 45-percent reduction in NO  could be achieved
without any loss in fuel economy by means of an emission control sys-
tem consisting of ten percent EGR, intake air cooling to  150  F, and
three degrees timing retard (instead of the five degrees applied in
System No. 4).  In this case, aldehyde and odor emissions would not
change very much from the levels  of the unmodified engine, but smoke
would increase, perhaps by as much as  100 percent.
              Incorporation of  emission control System  No.  2  into
Engine No. 28 (3.4-degrees timing retard and experimental injectors)
resulted in a 50-percent improvement in NO , accompanied by  a
                                          Jt
3.2-percent loss in fuel economy,  and substantial  increases in alde-
hyde and smoke  emissions.  However,  it should be noted that the lug-
down smoke of the baseline engine was lower than  for the other engines
tested  in this program.
              The NO  reductions achieved with the various  emission
control systems on Engine No.  6 are lower than for the other engines.
                                4-65

-------
The largest reduction (28 percent) was realized with Emission Control
System No. 4, utilizing two degrees timing retard and 15 percent EGR,
at the expense of a 3.2-percent increase in specific fuel consumption.
The use of a pair of UOP noble metal catalysts caused reductions in
CO and HC of about 88 percent.  However,  it should be noted that these
were fresh catalysts and durability testing would be required to verify
the performance of these catalysts  over  a long time period.  With
catalysts, the aldehyde and odor emissions were reduced substantially,
relative to the baseline levels, while the smoke intensity remained
essentially unchanged.
               Engine No.  16 was tested with three different emission
control systems incorporating a special pump and nozzle kit, after-
cooling, and injection timing modifications.  The  special pump was
equipped with a built-in puff limiter for acceleration smoke control and
an experimental idle feature,  and the special nozzles had a wider spray
angle than the standard nozzles.  Minimum NO  emissions were
                                             jC
obtained with Control System No. 3, utilizing three degrees timing
retard, intake air cooling to  150 F, and the standard pump and nozzles.
In this case, NO was reduced by 41 percent, while specific fuel con-
                ji
sumption increased by only about one percent. Aldehydes, odor, and
smoke were not affected much by the different emission control systems.
               Emission versus fuel consumption  correlations projected
by Cummins for future diesel engines are shown in Figure  4-30
(Ref. 4-26).  As indicated, the emissions of the current production
engines might be reduced to the 5 g/bhp-hr (HC +  NO  ) level  specified
                                                   j*t
by California for 1977 model year trucks by incorporating three  sets of
engine modifications, designated Phase I, Phase II, and Phase III.
Phase I consists of smaller production tolerances, retarded fuel injec-
tion timing, higher engine compression ratio, a new turbocharger, and
intercooling in the lower BMEP range.  Phase II includes the Phase I
modifications plus a larger camshaft and variable injection timing.
                                 4-66

-------
o 12
**
o>
o 10
        ^  6
        I
        o
                T    I     I     I     r
                                                PRESENT
                                                PRODUCTION
                                                DESIGN
                                                PHASE I
                                          O—-O PHASE II
                                          G—D PHASE III
                              I
               0.35 0.36  0.37 0.38  0.39 0.40 0.41  0.42 0.43
                   BRAKE SPECIFIC FUEL CONSUMPTION, Ib/bhp-hr
          Figure 4-30.  Projected effect of emission control
                        systems on emissions and specific
                        fuel consumption (Ref. 4-26)

Finally,  in Phase HI, a new injector will be utilized in conjunction with
the Phase II modifications.
              Referring to Figure 4-30,  the fuel economy of the pres-
ent production engines can be achieved at the 5 g/bhp-hr emission
level by incorporating the Phase II or Phase III modifications.  Con-
versely,  at the 10 to  12 g/bhp-hr emission level typical of many cur-
rent open-chamber diesels, the fuel consumption of the Phase III
engine can be reduced by about ten percent.  In these systems, reduc-
tion  in NO is accomplished by the timing retard and improved after-
cooling,  while the lower HC levels are primarily due to the  higher
engine compression ratio and variable timing.  The design of these
engine modifications  has not yet been finalized.  Based on some prob-
lems experienced during testing of the Phase I hardware, Cummins is
very concerned about rapidly meeting acceptable durability  and
                                4-67

-------
reliability standards for these new engines.  However, with normal
development time, Cummins is confident that these problems can be
resolved.
               Similar design modifications are being  considered by
other diesel engine manufacturers to meet future emission standards.
For example, General Motors intends to meet the 1975 California
standards for heavy-duty vehicles by incorporating a new turbocharger,
an air-to-water aftercooler,  and retarded  injection timing.  No informa-
tion is available relative to the fuel consumption of these engines.
4.1.4.2       Steady-state Tests
               The combined effects of injection timing and air-swirl
modifications were evaluated by Khan et al. (Ref.  4-10),  using a  single-
cylinder, open-chamber, naturally aspirated diesel.   Increasing the air
swirl makes the smoke versus timing curve less sensitive to timing
changes and retards the timing for optimum efficiency.  In addition, by
using the optimum combination of air  swirl and timing, NO  can be
reduced without sacrificing fuel economy.  For example, changing the
swirl ratio  from 2 to 8 and the injection timing from 17 BTDC to about
6°BTDC resulted  in a reduction in NO  from 1350 ppm to 1000 ppm, an
                                    n
improvement of about 25 percent, with no loss in fuel  consumption.
               Emission and  specific fuel consumption factors of a
turbocharged,  divided-chamber diesel engine  tested by the  Bureau of
Mines (Ref.  4-3), without  and with incorporation of an emission control
system, are presented in Figure 4-31.  The particular emission control
system  used on this engine consisted of ten percent EGR, combined with
five-degree injection timing retard.  At rated speed and full load,  NO
was  reduced by about  50 percent, at the expense of a six percent loss in
specific fuel consumption.  At the intermediate speed  of 1600 rpm, the
NO  reduction at full load was about 60 percent, while the loss in spe-
   Jt
cific fuel consumption was over ten percent.   The  NO  effectiveness of
                                4-68

-------
                          o 1600 rpm
                          A 2200 rpm
                           10% EGR +
                           3° RETARD
                           50    100    150    200
                             POWER OUTPUT, bhp
250
        Figure 4-31.  Effect of 10 percent EGR and 5° injection
                      timing retard on specific fuel consump-
                      tion and emissions (Engine No. 24)

the emission control system decreased with decreasing load, but the
loss in specific fuel consumption decreased also, for both operating
speeds.  The effect of the control system on HC was essentially inde-
pendent of speed and varied by about ±50 percent over the load range
evaluated in this program.  With emission control, CO is higher,
particularly in the high load regime.
              Data published by Cooper-Bessemer for its KSV-12
diesel engine operated on both No. 2 diesel fuel and on natural gas and
                                4-69

-------
fitted with several different combinations of emission control devices/
techniques are listed in Table 4-8 (Ref. 4-6).  The control systems
evaluated by Cooper-Bessemer include various combinations of timing
retard, increased air flow,  lower intake air temperature, and water
injection into the intake manifold. In the  case  of diesel fuel,  lowering
the manifold temperature from 130  F to  100  F, combined with four
degrees timing retard, resulted in a Zl-percent reduction in NO  and a
                                                              3t
slight improvement  in specific fuel  consumption.  Increasing the air
flow rate had no effect on NO  , but  provided further improvement in the
                            jt
specific fuel consumption of the engine.  Adding water injection caused
an additional six percent reduction in NO  at the expense  of an increase
in specific fuel consumption back to the level of the uncontrolled engine.
When operating with natural gas,  these emission control measures were
more effective in terms of  reducing NO , but the specific fuel consump-
                                      Jt
tion increased by as much as 3.6 percent. In this case, the observed
variations  in HC and CO were not very large.  As expected, CO
increased at the lower air intake  temperature, but decreased again as
more air was added.
4. 2           SPARK IGNITION ENGINES
4. 2. 1         Modification of Engine Operating Conditions
              A number of important engine  operating parameters
have been identified which have a strong  effect on the NO  ,  HC, and
                                                      X
CO emissions from  spark ignition engines.  These include air-fuel
ratio; ignition timing; compression  ratio; mixture and coolant tem-
peratures; engine  speed and load; valve timing, exhaust backpressure;
and combustion chamber deposits.  These parameters are discussed
in the following subsections.
4. 2. 1. 1       Air-Fuel Ratio
              As shown in Figure 4-32,  the air-fuel ratio of the com-
bustible mixture has a very pronounced effect on the NO ,  HC,  and
                                                      X
                                 4-70

-------
TABLE 4-8.  EFFECT OF COMBINED EMISSION CONTROL TECHNIQUES
             ON DIESEL ENGINE EMISSION AND SPECIFIC FUEL
             CONSUMPTION (Ref.  4-6)
Fuel
No. 2 Diesel
No. 2 Diesel
No. 2 Diesel
No. 2 Diesel
Natural Gas
Natural Gas
Natural Gas
Natural Gas
Injection
Timing
Standard
-4°
-4°
-4°
Standard
-4°
-4°
-4°
Air Intake
Tempera-
ture, °F
130
100
100
100
130
100
100
100
Air Flow
Rate
Standard
Standard
+6%
+6%
Standard
Standard
+ 10%
+ 10%
Water
Injection,
gpm
0
0
0
0.5
0
0
0
0. 3
Mass Emissions
g/bhp-hr
HC
0. 13
0.20
0.20
0.20
5. 16
3.20
7.39
7. 76
CO
3.85
4. 09
3.39
3.67
4. 50
6.45
3. 44
3. 17
NOX
10.99
8.71.
8.66
7.98
8.96
6.63
5.29
5.27
BSFC, a
Btu/bhp-
hr
6677
6636
6583
6664
6340
6440
6530
6568
Emission Ratios
-H£
HC0b
1.0
1. 54
1. 54
1.54
1.0
0.62
1.43
1. 50
CO
co0
1.0
1. 06
0. 88
0.95
1.0
1.43
0. 76
0. 70
NOX
NOXQ
1.0
0. 79
0.79
0.73
1.0
0.74
0. 59
0. 59
BSFCa
BSFC0
1.0
0. 994
0.986
0.998
1.0
1.016
1. 030
1.036
Brake Specific Fuel Consumption
Subscript zero refers to baseline values.

-------
                                   Figure 4-32.
Effect of air-fuel
ratio  on emission
levels, gasoline
spark-ignition
engine
CO exhaust concentrations emitted  from spark ignition engines.
Adjustment of the air-fuel ratio to a fuel rich mixture results in low
NOX emission, at the expense of an increase in CO and HC,  as well
as fuel consumption.  This approach is usually taken in combination
with corrective emission control measures  such as thermal reactors
and/or catalytic converters.
               In conventional gasoline engines, adjustment  of the air-
fuel ratio to a fuel lean mixture is limited to slightly leaner than stoi-
chiometric  (approximately 15.5 to 16).  At this setting,  the  HC and
CO concentrations approach a minimum whereas, the concentration of
NO  reaches a maximum.  Substantial further leaning of the air-fuel
mixture,  which would result in the lowest concentration of all three
contaminants,  appears feasible only by means of mixture stratification
or improved carburation and  fuel injection techniques.
               It  should be noted that the smog-forming potential of the
exhaust gases depends not only on the quantity of HC  and NO  emitted
                                                          ji
but also on the reactivity of the various HC compounds (Refs. 4-27,
4-28,  and 4-29).   Among the  many HC types emitted, the olefines have
                                4-72

-------
the highest reactivity while the paraffines have the lowest reactivity.
The effect of air-fuel ratio on reactivity is illustrated in Figure 4-33
(from Ref.  4-28)  showing chromatograph data from a single-cylinder
spark-ignition gasoline engine.  As  indicated, the total hydrocarbon
concentration decreased rapidly with leaning of the mixture and reached
a minimum at  air-fuel ratio of about 16. Conversely, the total HC re-
activity index  (smog-forming potential) showed much less variation
over the range of air-fuel  ratios investigated.  The apparent benefit of
leaning the air-fuel mixture (up to A/F = 17) for reduction of unburned
hydrocarbons is considerably diminished by the relatively modest
reduction of the total reactivity index of the exhaust.  In addition to
air-fuel ratio the total reactivity index of the exhaust depends on the
engine design and the type of fuel used.  The effect of air-fuel  ratio on
the fuel economy  of an automobile equipped with a V-8 spark ignition
engine and cruising at 30 mph is shown in Figure 4-34 (Ref.  4-30).  In
this case the best fuel economy is achieved at an air-fuel ratio of about
16, which coincides  with minimum  HC.
    8000

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                           Figure 4-33.  Effect of air-fuel
                                         ratio on reactivity
                                         index and concentra-
                                         tion  measured by in-
                                         frared analyzer
                                         (Ref. 4-28)
     10   1?    14    16
             AIR-Flfl RATIO
                                 4-73

-------
                                 Figure 4-34.  Effect of air-fuel ratio
                                               on exhaust hydrocarbon
                                               emission and fuel econ-
                                               omy in car at 30 mph
                                               roadload (Ref.  4-30)
               The effect of air-fuel ratio on the emissions of a Cooper
Bessemer GMVA-8,  two-stroke  stationary spark ignition gas engine,
rated at  1080 bhp,  300 rpm and 82. 5 bmep is illustrated in Figure 4-35
(Ref. 4-31).  In these tests the air-fuel ratio was varied over a narrow
range (A/F = 22 to 26) by changing the scavenging air flow rate.   Again,
NO  decreased with increasing air-fuel ratio while CO  remained
   X
unchanged.   However, unlike automotive engines,  HC showed no varia-
tion over the range of air-fuel ratios evaluated by Cooper  Bessemer.
The specific fuel consumption has  a minimum at the baseline operating
conditions indicated in the figure,  while the firing pressure and the
spark plug gasket and cylinder exhaust temperatures decline steadily
with increasing air-fuel ratio.
4. 2. 1. 2
Ignition Timing
              Ignition timing has a strong effect on the HC and NO
emissions of spark ignition engines.  In general these species decrease
substantially with increasing ignition timing retard.
                                4-74

-------
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                       CYLINDER EXHAUST TEMP.
                          V        	"-•
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                          i
                   FIRING PRESSURE
MASS EMISSIONS

   NO,
   HCT
                  I

                            CO
                           	I
             I
         120      140      160      180      200     220

          AIR FLOW RATE, % DISPLACEMENT AT 29" Hg, 80"F



Figure 4-35.  Effect of air flow rate on engine

               emissions and  performance  - Cooper

               Bessemer GMVA-8 2-stroke atmospheric

               spark-gas engine;  1080 bhp at 300 rpm,

               82.5 bmep, base conditions  - (Ref. 4-31)
                         4-75

-------
              From  several experimental and theoretical studies
(Refs. 4-32 and 4-33) it can be concluded that the nitric oxide concen-
tration in the exhaust correlates reasonably well with the peak combus
tion temperature in the engine.   This is illustrated in Figure 4-36
(Ref.  4-32) showing a plot of measured peak cycle temperature and
NO versus spark advance (air-fuel ratio 11:1).  The general trends
shown in the figure support the view that the reduction of nitric oxide
by means of spark retard is due  to the attendant reduction of the peak
combustion temperature.  The combined effect of spark  retard and
air-fuel ratio on the NO  concentration in the exhaust of a CRF engine
                       A.
is shown in Figure 4-37 (Ref. 4-34).  As indicated,  the effect of igni-
tion timing is very pronounced for lean mixtures (air-fuel ratio >15)
and rather modest for rich mixtures (air-fuel ratio <14).  Similar
results were reported by Huls, et al (Ref.  4-35).
                     I    I    I    I    I    I    I    I
                    PEAK CYCLE FLAME TEMPERATURES
                    FOR VARIOUS SPARK ADVANCES
                    1250rpm, A/F 11:1
                    35  30  25  20  15  10   5
                       SPARK ADVANCE, deg BTDC
            Figure 4-36.  Correlation between peak cycle
                          temperature and NO concentration
                          (Ref.  4-32)
                                4-76

-------
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            JOTAL PARAFFIN

                      ~
                                         Figure 4-38.
                                                           Effect of  spark timing
                                                           on hydrocarbon com-
                                                           position by class  at
                                                           rich air-fuel  ratio
                                                           (Ref. 4-28)
       0           10          20
        SPARK TIMING 'RETARDED FROM MBT
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  1400
  1200
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         SINGLE-CYLINDER ENGINE
         CONSTANT POWER
         AIR-FUtL RATIO 13.0
         FUEL 1
       TOTAL HYDROCARBON
              TOTAL PARAF_n_N_
               TOTAL AROMATIC
            ACETYLENE
      J	L
                                         Figure 4-39.  Effect of spark-timing
                                                          on hydrocarbon com-
                                                          position by class at
                                                          lean  air-fuel ratio
                                                          (Ref.  4-28)
       0           10          20
        SPARK TIMING 'RETARDED FROM MBT
                                        4-78

-------
               Figure 4-40 presents the effect of ignition timing on the
exhaust emissions of a Cooper Bessemer GMVA-8, two-stroke sta-
tionary engine using gaseous fuel (Ref. 4-31).  Retarding the ignition
timing from the standard setting of ten degrees BTDC to four degrees,
results in a 15 percent reduction in NO  , accompanied by very small
                                     Jt
variations in HC and CO.  However,  the specific fuel consumption of
the engine increases substantially, from about  7070  Btu/bhp-hr to
about 7500 Btu/bhp-hr.  At the same time,  the firing pressure
decreases, whereas the cylinder exhaust gas temperature increases.
4.2.1.3        Mixture Temperature
               The temperature  of the air-fuel  mixture  supplied to the
engine has a  noticeable effect on the  concentration of the contaminants
in the exhaust.  Several factors  determine the temperature  of the mix-
ture prior to its admission into the engine cylinder including the  atmo-
spheric conditions of the intake air (temperature,  density and humidity),
the heat capacity and heat of vaporization of the fuel, and the amount of
heat transferred to the mixture in  the engine intake manifold.
               The effect of atmospheric conditions is relatively  small,
but important enough to require  numerical correction factors to the
measured emission  data of nitric oxide, in order to make possible an
unbiased comparison of test data,  taken under varying conditions of
atmospheric  humidity (Ref. 4-33).
               The heat transfer and the fuel evaporation process in
the engine inlet manifold affects  the mixture temperature more sub-
stantially.  Figure 4-41 (Ref. 4-36) presents the effect  of inlet mani-
fold heating on mixture temperature and on the  resulting emissions of
NO ,  HC, and CO.  The tests were performed  on a car equipped with
a V-8 spark ignition engine.  The emissions were  measured at car
cruising speeds up to 70 mph, with the inlet manifold heated in a con-
ventional way by exhaust gases passing through the manifold cross-over
                                4-79

-------
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                         CYLINDER EXHAUST TEMP.
                         SPARK PLUG GASKET TEMP.
              FIRING  PRESSURE
                     MASS EMISSIONS

                       NOV
              FUEL CONSUMPTION
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Figure 4-40.   Effect of ignition timing  on engine emissions and

               performance - Cooper Bessemer GMVA-8
               2-stroke atmospheric spark-gas engine,  1080  bhp at
               300 rpm,  82.5 bmep, base conditions  (Ref. 4-31)
                               4-80

-------
   400O
   3000
           O HEATED INLET MANIFOLD
           A COLO INLET MANIFOLD
      160
            ...J	J	
           O HEATED INLET MANIFOLD
           A COLO INLET MANIFOLD
              I
                   20    30    40

                      CAR SPEED. MPH
Figure 4-41.   Effect of heating of inlet manifold
                on exhaust emissions and  mixture
                temperature (Ref.  4-36)
                      4-81

-------
passage.  For  the  tests with cold inlet  manifold,  the  cross-over
passage was blocked off on both sides of  the manifold.  As indicated in
Figure 4-41  a mixture temperature drop of approximately 20°F (with
blocked-off heating passage) resulted in a substantial reduction (20 to
40 percent) of NO .  However, hydrocarbons increased slightly (at low
                 Ji
speeds only) and carbon monoxide increased substantially throughout
the test range.  The observed reduction in NO  is attributed to the
lower peak combustion temperature which results from the  lower intake
temperature.
              On the other hand the  higher HC and CO emissions are
considered to be due to fuel maldistribution and wall quench effects
(Ref. 4-37).  The magnitude of these effects depends to some degree
on the engine inlet manifold and fuel  system design and the type of fuel
used.  In general, cooling of the mixture decreases the tendency to
engine knock (Ref.  4-38).  No information is  currently available on  the
effects of mixture temperature on the reactivity index  of the exhaust
gas.
              Figure 4-42 (Ref. 4-31) shows the effect of inlet air
temperature on the exhaust emissions of a Cooper Bessemer GMVA-8,
two-stroke stationary gas  engine.  In general terms, the observed
trends are quite similar to those of automotive spark ignition engines,
but the magnitude of the changes varies considerably.  For  example,
a reduction of the manifold temperature from the standard setting of
130 to 100°F caused a decrease in NO  from about 15 g/bhp-hr to
                                    x
11 g/bhp-hr, a  reduction of about 27 percent.   This was accompanied
by a slight increase in air-fuel ratio and  specific fuel consumption.
The  HC and CO emissions  remained essentially unchanged over the
range of temperatures investigated by Cooper Bessemer.
4.2.1.4       Coolant Temperature
              For  satisfactory engine operation the temperature of
the engine must be maintained within predetermined limits.   If the
                                4-82

-------
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                  CYLINDER  EXHAUST TEMPERATURE
             o——o	
          L  O
                  SPARK  PLUG GASKET TEMPERATURE
                TRAPPED AIR/FUEL RATIO
                 FUEL CONSUMPTION
                                              I
             80       100      120      140      160
                AIR MANIFOLD TEMPERATURE, °F

Figure 4-42.  Effect of air manifold temperature on
              emissions and performance - Cooper
              Bessemer GMVA-8 two-stroke atmo-
              spheric spark-gas  engine;  1080 bhp at
              330  rpm, 82. 5 bmep, base conditions
              (Ref. 4-31)
                        4-83

-------
engine becomes excessively hot then preignition and "engine knock"
may develop and adversely affect its performance.  Varying the engine
temperature within the permissible limits changes the degree of pre-
heating of the combustible mixture prior to ignition, hence affecting
the exhaust emissions, as previously discussed in Section 4. 2. 1. 3.
Furthermore,  the combustion chamber wall temperature affects the
thickness of the quench layer adjacent to the chamber walls.  A portion
of this quench layer which contains unburned fuel is  exhausted during
the exhaust stroke (Ref. 4-39).
               The relationship between the  combustion chamber sur-
face temperature and the HC  emissions  has  been the subject of inten-
sive studies.  Figure 4-43 presents the effect of the average combustion
chamber surface temperature on the concentration of exhaust hydro-
carbons (Ref. 4-40).   The data shown are for a production 6-cylinder,
230 cu in.  spark ignition engine having a compression ratio of 8. 5.
The engine was operated on isooctane fuel using a constant air-fuel
ratio of 14, constant inlet  air temperature (123°F) and constant exhaust
back pressure.  As indicated, HC decreased markedly with increasing
surface temperature.  Similar trends were observed at other engine
speed and load settings.

120

110
100

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SPARK TIMING

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                                 Figure 4-43.
Effect of average com-
bustion  chamber sur-
face temperature on
hydrocarbon emission
(Ref. 4-40)
                                4-84

-------
               Test data published by Daniel (Ref. 4-41) for a
single-cylinder engine indicate that the effect of coolant temperature
variations on the concentrations of the various HC species is  very
similar.  As a result the exhaust reactivity index is believed  to be
proportional to the total HC concentration in the exhaust.
               The effects of  coolant temperature on the NO , HC, and
CO emissions, as determined on a single cylinder gasoline engine,  are
shown in Figure 4-44 (Ref. 4-42). In addition, the figure shows the
effects of spark timing and air-fuel ratio.  As expected,  the  specific
mass emission of NO  (g/ihp-hr) increased noticeably with increasing
coolant temperature.  At the  stoichiometric fuel-air ratio, changing
the coolant temperature from 370 to 212°F resulted in a 20-percent
reduction in  NO , accompanied by a 30-percent  increase in HC and
               X
some reductions in CO.   In these tests the indicated mean effective
pressure  increased somewhat with decreasing coolant temperature,
while the  specific fuel consumption showed a tendency to decrease,
particularly  for retarded spark settings.  Unfortunately this investiga-
tion was limited mostly to tests with richer than stoichiometric air-
fuel ratios and on the basis of the data published  it  cannot be ascer-
tained whether similar trends would apply also to lean air-fuel ratios.
4.2.1.5       Engine  Speed
              Relatively little information has been published on the
effects of engine speed on exhaust emissions. Figure 4-45 presents
the concentration of nitric oxide as a function of engine speed
(Ref. 4-34).  The tests were  carried out on a single-cylinder experi-
mental engine and the  mixture temperature,  manifold  pressure, igni-
tion timing and compression ratio were  kept constant.  With rich
mixtures,  NO  increased with increasing speed.  Conversely, in the
             X
lean regime  NO  decreased sharply with increasing speed.
               X
                                 4-85

-------
                   COOLANT TEMPERATURE
                                         370V
       95X CONFIDENCE LIMITS ON MEAN
       80COMPRESSION RATIO
         tO 0-10-20-30   10 0-10-20-30    10 0-10-20-30
                   RELATIVE SPARK ADVANCE
                 DEGREES ADVANCE FROM MPST
                                                    350 j-


                                                    300


                                                    2 SO


                                                    200-
                                                       .

                                                    ISO
            COOLANT TEMPERATURE
     I49°F          212"F

95X CONFIDENCE LIMITS ON MEAN
80 COMPRESSION RATIO
                                                                                          370'F
                                                  y  100
                                                  C3

                                                  "  SO


                                                      0
                                                          -.1 .L._l  11   . J  I   I	I	1	1.  _1  L  I  1
                                                           10 0-10-20-30  10  0  -10-20-30   10  O-tO-20-30
                                                                      RELATIVE  SPARK ADVANCE
                                                                    DEGREES  ADVANCE FROM MPST
-10
  •
   9SX CONFIDENCE LIMITS ON MEAN
   «0 COMPRESSION RATIO
                   COOLANT TEMPERATURE
          I49'F           212*F
                                       370°F
      --J—I—I	I._J	L. I_L	
       10  0-10-20-30   10  0-10-20-30   10  0-10-20-30
                  RELATIVE SPARK ADVANCE
                DEGREES ADVANCE FROM MPST
     Figure 4-44.   Effect  of coolant  temperature and spark advance
                         upon  indicated specific  NO,  HC,  and CO
                         emissions  (Ref.  4-4Z)
                                              4-86

-------
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Figure 4-45.  Effect of engine
              speed on NOX con-
              centration
              (Ref. 4-34)
          10     20    30     40
             ENGINE SPEED - rpm t 100
              In a more recent study the effects of engine speed and
load on nitric oxide emissions were investigated with a single-cylinder
engine operated on a  rich indolene 30/air mixture (83 percent of
stoichiometric) and with optimized ignition timing (Ref. 4-43).  Again
NO  increased with increasing engine speed.  This increase is attributed
by the author to the variation in engine speed and load which caused a
change in the degree  of charge dilution (by residual gas).   Furthermore,
the author concluded  that the  effect of speed on NO was greatest at low
                                                 5C
engine load and the effect of load was greatest at low engine speed.
Also, the speed and load effects appeared to be  more pronounced in the
case of larger valve  overlap.
              The effect of engine speed on HC  concentration is  shown
in Figure 4-46 for three automotive spark ignition engines (144 CID,
292 CID,  and 352  CID) operated at three different air-fuel ratios
(Ref.  4-44).  In  all cases the concentration of exhaust hydrocarbons
decreased with increasing rpm.  This trend is even more apparent
                                4-87

-------
     500
   2 400
   Q.
   a
   z
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   a)
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   u
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   a:
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     100
           ENGINE A- 144 C I 0

            i500 RPM 12 OBHP
            2500 RPM 22 OBHP
                          OBSERVED BRAKE HORSEPOWER
               ENGINE B- 292 Cl 0

                ISOO RPM 17 OBHP
                2500 RPM 250BHP
12 I A/F"
ENGINE C- 352 C.I 0

 IJOO RPM 20 OBHP
 2500 RPM S3 OBHP
          1500    2000    2500
                             1500   2000    2500
                               ENGINE SPEED (RPM)
                                                 1500
                                                       2000
                                             2500
       Figure 4-46.  Effect of engine speed on HC concentration
                     (Ref. 4-44)
when the emissions are expressed in terms  of g/bhp-hr.  Similar
results were  obtained by Daniel (Ref. 4-41).
               While the decrease of exhaust hydrocarbons concentra-
tion with increasing engine speed is  well known, it has been one of the
most difficult phenomena to explain.  In a  study of the surface phenom-
ena associated with exhaust hydrocarbon emissions (Ref. 4-45), a hypo-
thesis is presented which relates the HC emissions to the  piston top
land clearance volume.  The observed decrease in HC with increasing
engine  speed is attributed to  an increase in  the average cylinder
                                 4-88

-------
surface temperature resulting  in  thermal expansion of the piston
crown,  hence a reduction of the top-land volume and the amount of
ejected unburned hydrocarbons.
               No effect of engine speed on exhaust carbon monoxide
concentration has been reported.
               Test results from a larger two-stroke stationary gas
engine are presented in Figure 4-47 (Ref. 4-31).   As indicated NO
                                                                jf.
follows a trend similar to automotive spark-ignition engines.  Rough
engine operation was noted at 400 rpm,  which might explain the
sharp increase in HC.  Increasing the engine speed from 300 rpm
(standard speed) to 330 rpm resulted in a reduction in NO  from about
15 g/bhp-hr to about 7  g/bhp-hr.   This effect which is accompanied
by a small increase in  specific fuel consumption, is the direct result
of a reduction in the brake mean effective pressure of the engine and
a substantial increase in air-fuel ratio.
4.2.1.6       Valve  Timing
               Traditionally,  the major objective of cam design,  for
the timed actuation of the engine valve  system,  is to provide the highest
volumetric efficiency and the least charge dilution (by residual gases
in the engine cylinder)  over the entire operating range of the engine.
Fulfillment of this objective results in  good engine performance and
fuel economy.
               However, a change  in valve timing from the optimal
setting  has a significant effect on exhaust emissions,  particularly NO .
                                                                   Jt
Generally any change in timing of the inlet or exhaust valve, will result
in an increase of the charge dilution by a larger amount of residual
gases trapped in the engine cylinder.  As a result the combustion
temperature  and the formation of NO  are  reduced.
                                4-89

-------
                 I   CYLINDER EXHAUST
                      TEMPERATURE
                          D.
                   SPARK PLUG GASKET
                     TEMPERATURE
                          TRAPPED AIR/FUEL RATIO
         — BASE CONDITIONS:
        275
300
325      350
 SPEED, rpm
375
400
Figure 4-47.  Effect of speed on emissions and per-
              formance - Cooper Bessemer GMVA-8
              two-stroke atmospheric spark-
              gas engine, power output 1080 bhp,
              base conditions (Ref. 4-31)
                       4-90

-------
              The magnitude of the reduction depends also on the
temperature of the internally recycled exhaust gas.  In the case of
early inlet valve opening the exhaust gases are partially discharged
into the inlet manifold (intake reversed flow) where they are effectively
cooled in contact with the relatively cold walls of the inlet manifold.
During the piston intake stroke, the air-fuel mixture flow brings the
cooled exhaust gases back into the engine cylinder, thus increasing
the charge dilution (Ref. 4-46).
              Since early inlet valve opening causes the exhaust sys-
tem to be  in communication with the  low pressure intake system for a
longer period of time, exhaust gases are returned to the cylinder
through the exhaust valve.
              The exhaust gas backflow originates from the end of
the exhaust stroke and thus is expected to have larger hydrocarbon con-
centrations (Ref. 4-39).  Retention of these hydrocarbon-rich quench
gases is believed to  be  the main reason  for the observed reduction in
hydrocarbon emission,  as indicated in Figure 4-48..  The data which
were  obtained on a V-8 experimental engine converted to single-
cylinder operation showed a considerable reduction of HC and NO
(g/ihp-hr)  when the intake valve opening was advanced from standard
timing. Retarding the intake valve opening slightly increased the NO
                                                                  jt
emission but had no  effect on hydrocarbons.  For the particular test
conditions  shown intake valve opening timing had no significant effect
on CO or specific  fuel consumption.
              In another  series of tests the effect of  exhaust valve
closing timing was investigated.   Figure 4-49 presents the results.
Advancing  the closing time of the  exhaust valve by about 50 degrees
resulted in a 25-percent reduction of HC and  a 45-percent reduction
of NO  .  Retarding the  closing time by 35 degrees resulted in HC
      Jt
and NO reductions of about 18 and 50 percent,  respectively.
                                4-91

-------


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5™- SOpsilMEP
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ill i i i
30 10 -10 -30 -50
INTAKE VALVE OPENING, deg BTDC

_


—



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Figure  4-48.
             Effect of intake valve open-
             ing timing on HC and NO
             emissions (Ref. 4-39)
BC
1
i
TC
1
rr

INT.
BC

             EARLY
                        LATE
              ,
                      MBT       \
                      14.5:1 A/F   A\
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                        STD.

                         I     .
    -50   -30    -10   10    30    50

        EXHAUST VALVE CLOSURE, deg ATDC
                                   70
                                       O
Figure 4-49.  Effect of exhaust valve
               closing on HC and NO
               emissions (Ref. 4-39)
                 4-92

-------
Apparently early closing of the exhaust valve prevented the discharge
of hydrocarbon-rich quench gases (Ref.  4-39) and was therefore
more effective in reducing the emission of hydrocarbons than late
closing.
               On the other hand, in the case of NO  abatement late
                                                 Jt
closing of the exhaust valve was  more effective because of higher
cooling of the residual gases in the inlet manifold.  Again exhaust
valve closing timing had no significant effect on CO and specific fuel
consumption.
               Valve tinning, and other parameters  such as exhaust
back pressure, air-fuel mixture temperature,  engine load, engine
speed, air-fuel ratio,  spark retard and external exhaust gas recircu-
lation (EGR) singly or in combination  have pronounced effects on NO  .
This is illustrated in Figure 4-50 (Ref. 4-47).  As indicated  NO
                                                             X
depends only on the amount of charge  dilution and is independent  of
the dilution method used.  However, the relationship between spark
retard and exhaust back pressure and NO  is more  complex.  On the
basis of this  comparison, it was  concluded that not  all engine variables
exhibited a consistent relationship between NO emission and  charge
dilution and that charge dilution,  although quite important, is definitely
not the only factor affecting NO emission.
              Emission test results obtained on a car equipped with  a
350 CID engine with 9:1 compression ratio are  depicted in Figure 4-51,
showing the  effects of cam advance and retard on NO  (1970 Federal
                                                   Jt
test procedure).  Both advancing and retarding the camshaft  by
40 degrees reduced NO  emission by about 40 to 50 percent.   In addi-
                      Jt
tion, a small reduction in HC was observed  (Ref. 4-48).  The idle
quality was good, but there was a significant loss in power resulting
in poor performance at part throttle.  However, at  speeds above
1200 rpm  driveability was  acceptable.
                                4-93

-------
   40
   32
   24
i •
MBT
16.0 A/F
1.2 bhp
                    V
            TEST VARIABLES

         • VALVE OVERLAP
         o EXHAUST RECIRCULATION
         • COMPRESSION RATIO
       0.04      0.08     0.12      0.16
             CHARGE DILUTION FRACTION
            Figure 4-50.  Nitric oxide vs
                           charge dilution re-
                           lationship for valve
                           overlap, recircula-
                           tion and compres-
                           sion ratio tests
                           (Ref. 4-47)
 GM/MI
HC 50
40
30
20
10
CO 5
4
-
-


3

\

I
1
2
1
0
NOx
-
-
-




1
1
    Figure 4-51.  Effect of cam advance and  retard on hot cycle
                    vehicle emissions with the 1970 Federal test
                    procedure (Ref.  4-48)
                                  4-94

-------
              In another test utilizing variable valve overlap NO  was
reduced to about 1.2 g/mile,  (1970 Federal test procedure) compared
to 3.5 g/mile with the standard camshaft.  However, HC was higher
mainly as a result of the richer idle mixtures required to reduce
misfire.
              Complementary information on variable cam timing as
a tool for emission control is presented in References 4-49 and 4-50.
              Tests conducted by Freeman et al (Ref.  4-48) indicate
that the fuel consumption increased slightly as  the cam was advanced
and decreased slightly as the cam was retarded.  Advancing the cam
causes opening of the exhaust valves before the end of the expansion
stroke.  Thus the cylinder pressure is  relieved prematurely resulting
in some loss of power and fuel economy.
4.2.1.7       Engine Load
              Generally,  the concentration of NO  increases markedly
                                               JC
with increasing inlet manifold air pressure  (increasing engine load),
as shown in Figure 4-52 (Ref. 4-34).  The increase of NO  with increas-
                                                       JC
ing engine load is due to the attendant rise of the combustion tempera-
ture and the reduction of the  mixture dilution by residual exhaust  gases
in the engine cylinder (Ref.  4-49).
              Engine load also has an effect on the average surface
temperature  (AST)  of the engine cylinder walls and consequently will
influence the "wall  quenching" phenomenon and the emission of unburned
hydrocarbons.  Test data indicate that, on the average, only about
43 percent of the decrease in exhaust hydrocarbon, which occurred
when indicated power was increased from 2 to  13 hp, can be attributed
to increasing AST (Ref. 4-40).  Other  factors most likely to affect HC
as power output is increased are quench zone gas temperature, pres-
sure,  turbulence, and after-reactions.
                                4-95

-------
            6000
            5000
- COMPRESSION  6.7
  RATIO SPARK  30° BTC
  SPEED        lOOOrpm
         E  4000
         a
         a.
          *
          x
         i  3000
            2000
            1000
                                                  40-in. MAP
                       /   20-in.
                       /   MAP
                             •£."^'	10-in. MA^-'	
                           12     14     16      18
                               AIR/FUEL RATIO
            Figure 4-52.  Effect of manifold air pressure
                          on oxides of nitrogen (Ref. 4-34)
               Since the average surface temperature has been
identified as a variable affecting exhaust hydrocarbon concentration,
methods of increasing this temperature represent an interesting
HC abatement technique (Ref. 4-40).  Unfortunately, any increase in
AST adversely affects engine octane requirement and volumetric
efficiency.  The effect of engine load on CO appears to be very small.
               Test data from a large stationary two-stroke gas engine
are shown in Figure 4-53 (Ref.  4-31).  Generally,  the same trends  of
NO ,  HC, and CO, as in the case of automotive spark ignition engines,
   X
are apparent.  The decline in NO  as engine load is increased above
                                X
95 psi bmep is believed to be due to progressive enrichment of the
mixture.
                                4-96

-------
    700
 u_
 0

 u;  600
 a:
 D

 <  500
 cc
 LLJ

 |  400
 LU
 >—

    300



  .?900


0^700
£ of.


£ $ 500
        cc
        a.
          300
       a 8000

       CD
       > 7500
      U 7000
      00
      CO
     25


     20
      a.
      x
      00
       0)
       *
      00
      z
      o
     15


     10


      5


      0
                 CYLINDER EXHAUST TEMP
                                I
                 SPARK PLUG GASKET TEMP.
            FUEL CONSUMPTION
                                 MASS EMISSIONS
                                    CO
            60       70       80       90
                            TORQUE BMEP
100
                                               110
Figure 4-53.  Effect of load at constant speed on emissions
               and performance - large two-stroke atmo-
               spheric  spark-gas  engine,  base conditions,
               300 rpm (Ref. 4-31)
                             4-97

-------
4.2.1.8       Exhaust Backpressure
              Test data published by several investigators indicate
that exhaust backpressure can affect the exhaust emissions from spark
ignition engines (Ref.  4-46).  For example,  Eltinge et al (Ref. 4-51)
concluded that raising the exhaust backpressure lowers the average
hydrocarbon concentration of the exhaust gases,  partly because it
inhibits the exhaust of the last part of the charge, which has the
highest hydrocarbon concentration (Ref. 4-39).  Furthermore,
increasing backpressure in the exhaust system increases the oxida-
tion reaction potential in the exhaust manifold, particularly when the
increased backpressure is accompanied by higher exhaust gas tem-
perature.  To minimize fuel consumption losses,  a sophisticated
exhaust backpressure  control system might be employed, which would
raise the exhaust backpressure only at  certain engine operating
conditions.
              Test data from a V-8, 351 CID gasoline engine are pre-
sented in Figure 4-54  (Ref. 4-5Z), showing the effect of  exhaust back-
pressure on emissions.  Increasing  the exhaust pressure from 30 in Hg
absolute  up to 44 in Hg resulted in a moderate reduction of the NO  and
                                                              jt
HC exhaust concentration.  Since the exhaust backpressure adversely
affects the engine power output, the  reduction in HC and NO  on a
                                                         X
specific mass basis (g/bhp-hr) would be even less.  However, there
are indications that elevated exhaust pressure might be more effective
in conjunction with increased valve overlap (Ref. 4-46).   The adverse
effects of exhaust backpressure on engine power and fuel economy
might be counteracted  by expanding the  exhaust  pressure through a
turbocharger.  In Ref. 4-46,  it was  concluded that the relative effect
                                4-98

-------
              Z  1000


                   0
                 600
                 500
                 4OO
                 30O
                 2OO
                 100
                 z  0
                 °  4
                 2  3


                 o  0
                 I5OO
                 1400
                 I3OO
                 0.08
                 °07
                 006
                            1971 FORD 351-W ENGINE
                                   2000 rpm
1
• 17 In Hg T
x 21 in Hg \ Manifold Pressure
o 26 in HgJ   ,   |     |
                    12   16  2O  24  28  32  36  40  44  48
                    EXHAUST BACK PRESSURE. Inches of Mercury Absolute

       Figure 4-54.  Exhaust emissions, exhaust temperature,
                      and fuel-air ratio as functions of exhaust
                      backpressure for three absolute inlet
                      manifold pressures, 2000  rpm  (Ref. 4-52)
of increased exhaust backpressure on exhaust emissions was only slight
at standard valve timing, but quite noticeable in the case of increased
valve overlap.

4. 2. 1. 9       Combustion Chamber Deposits

               Extensive laboratory and field test results show that the
combustion chamber deposits exhibit a significant effect on the exhaust
                                  4-99

-------
emissions from spark-ignition internal combustion engines.  As an
example,  Figure 4-55 (Ref.  4-53) presents the effect of deposit buildup
on the walls of the engine combustion chamber on HC and NOx>  Before
the start of the test,  the combustion chamber deposits were removed
from all cylinders.  As indicated, the emissions of NO and HC in-
creased steadily with the buildup of deposite during the engine opera-
tion.  After 142 hours of operation,  the deposits were removed and
the exhaust concentration of nitric oxide dropped nearly 25 percent and
the concentration of exhaust hydrocarbons dropped about 66 percent.
Similar results were observed in a fleet test program (Ref. 4-54).
                z
                a
                a.
                g
                i-
                <
                o
                z
                o
                UJ
                9
                o
(T
I-
Z
                                   17
                                                   16
                                                   15 « j£
                                                      O
                                                   100
                                                   50
                        0  25  50  75  100  125  150     S
                      OPERATION TIMEON INDOLENE 30(HOURS)     "

          Figure 4-55.   Effect of deposit buildup on exhaust
                         NO and HC concentrations  (Ref. 4-53)
                                 4-100

-------
              The effect of leaded and unleaded gasolines on exhaust
emissions as influenced by combustion chamber deposits is presented
in Ref.  4-55.  On the basis of an intensive state-of-the-art  review of
numerous test data submitted by 18 different companies, it  was con-
cluded that (1) lead content in gasoline has a significant effect on the
formation of combustion chamber deposits,  (2) the equilibrium deposit
level  relative to hydrocarbon emission is 7 to 20 percent higher when
leaded instead of unleaded gasoline is used,  (3) the presence of lead in
C'asoline has no effect on carbon monoxide emission, and (4) no lead
content effects on nitric oxide emission have been observed.
              An investigation  conducted with a number of  gasoline
additives has indicated  that a reduction in HC by approximately 50 per-
cent was possible. This effect  is attributed to an attendant modification
of the deposits in the engine  (Ref. 4-56).
4. 2. 2         Preventive Emission Control by Engine Modification
4. 2. 2. 1       Combustion Chamber Modification
              The design of the combustion chamber of a spark ignition
engine has an important influence on its performance,  fuel  economy,
and NO  and HC  emissions (Refs. 4-38 and 4-57).   The shape of the
combustion chamber  and the relative location of the valves  affect the
flow pattern and  the degree of turbulence, and consequently the speed
of flame propagation  and the thickness of the wall boundary layer.
Intensive turbulence tends to decrease the thickness of the wall boundary
layer  containing  unburned fuel and promotes  its post-flame oxidation.
However, intensive turbulence increases the heat transfer to the  cylinder
walls  and tends to increase the  combustion pressure rise rate, the peak
combustion temperature, and the  formation of NO .  On the other hand,
                                                ji
intensive turbulence tends to decrease the selective retention of hydro-
carbon rich boundary layer gases in the combustion chamber during the
exhaust cycle, and for this reason may in some cases  cause high
emission of exhaust hydrocarbons.

                                4-101

-------
              The effect of combustion chamber shape and spark plug
location on NO  is shown in Figure 4-56 (Ref. 4-58).   As indicated,
              n
the chamber having the highest surface-to-volume ratio (configuration
No. 3) shows the lowest NO  emission.  Conversely,  configuration
                          X.
No. 4 is the most compact chamber, which results in the highest
NO  emission.
   x
              The influence of other factors  on surface-to-volume
(S/V) ratio is shown in Figure 4-57.  For example, as the combustion
chamber becomes smaller, S/V increases.  The effect of S/V on HC
is shown in  Figure 4-58,  depicting test  results obtained on a fleet of
cars equipped with similar engines (Ref. 4-50).  The S/V ratio can  be
varied by changing the  compression ratio  of the engine (Ref. 4-59).
Based on  steady-state dynamometer and vehicle tests, the specific
mass emissions of NO  have been shown to be independent of compres-
sion ratio as long as the engine is operated on rich mixtures. However,
                        (e) NO UNDER CONSTANT EGR RATE
                Figure 4-56.  NO emission per unit
                              output for different
                              combustion chamber
                              shapes and spark plug
                              locations (Ref.  4-58)
                                4-102

-------
  10
o
I-
<
3
e!
I
           8
     COMPRESSION RATIO
  10
u
3
in

        PRODUCTION CARBURETION
        SPARK NOT PORTED
                        CALIFORNIA LIMIT
CALIFORNIA LIMIT.
              567
            SURFACE/VOLUME RATIO
                                         Figure  4-58.
                                  Composite values:
                                  California chassis
                                  dynamometer
                                  schedule (Ref. 4-50)
                                     4-103

-------
for near-stoichiometric and lean mixtures,  NO  reaches a maximum
                                              x
at compression ratios between 8:1 and 9:1 and  decreases for lower
and higher compression ratios (Ref. 4-59).
4. 2. 2. 2       Fuel System Modifications
               Generally,  automotive spark ignition engines are
designed for operation on  gasoline which is  supplied by means  of
carburetion or fuel injection.  However, more  recently, a consider-
able number of light-duty  vehicles and trucks have been converted to
gaseous fuels (LPG,  natural gas).
               Since the close of the last century, thousands of patents
have been granted for numerous carburetion and fuel injection concepts.
The  general goal of these  inventions was to  improve the engine per-
formance and fuel economy and, in some cases,  extend the engine
operation to heavy or less volitile  fuels. Fuel  injection is favored by
some automobile manufacturers in order to improve the mixture uni-
formity.  As a result, an  extension of the lean  limit of engine  operation
can be achieved with a concomitant improvement in fuel economy and
emissions.  However, there are indications that similar gains might
be attained with more sophisticated carburetion systems.  For example,
a mixture-optimizer  system for use in conjunction with a conventional
carburetor or fuel-injection system has been recently developed
(Ref.  4-60).  It includes a feedback-type electronic control device
which automatically selects the air-fuel ratio,  yielding minimum fuel
consumption at all operating points.  Apparently,  the minimum fuel
consumption occurs for mixture ratios close to the misfire limit.
Although no test data are available from this system, the emissions
are expected to be low.
               The lean limit of engine operation can be extended by
means of charge homogenation, which might be accomplished by pass-
ing the air-fuel mixture  generated in a conventional carburetor through
                                4-104

-------
a vaporization tank.  For example, utilization of a steam-jacketed tank
designed to completely vaporize the liquid fuel and thoroughly mix the
fuel vapor and inlet air before induction into the engine permitted an
extension of the lean  limit from an air-fuel  ratio of about  17 to about 21,
This improvement, which is attributed to improvements in the fuel
distribution and cyclic cylinder pressure  variations,  resulted in sub-
stantial reductions in HC and CO (Ref. 4-37).
              In a more recent investigation, charge homogenation
was accomplished by vaporization of the fuel and subsequent mixing
of the vapor with unheated inlet air (Ref.  4-61).  In this case, NO
was reduced considerably,  accompanied by  some  reduction in HC
and CO.   Another fuel atomizing carburetion system  including EGR
was tested in  an automobile,  and substantial HC,  CO, and NO  reduc-
                                                            X
tions relative to the standard carburetor were achieved (Ref.  4-36).
4. 2. 2. 3       Inlet Manifold Optimization
              The size, shape, and length of the inlet manifold have
an important effect on the air-fuel mixture formation, fuel distribution
to the individual cylinders, and the resulting flow pattern  of the charge
in the cylinders.   This in turn has an important effect on the engine
combustion process (Refs.  4-38, 4-57,  4-62, and 4-63).
              In general, the smaller the cross-sectional area of the
inlet manifold, the higher the flow velocity in the  manifold,  resulting
in high turbulence and improved mixing of the charge.  The attendant
loss in volumetric efficiency can be alleviated by  means of a dual inlet
manifold  (Ref. 4-64). In this device, a manifold having a small cross
section was fed from a small-bore carburetor while  a large manifold
was coupled to a conventional two-barrel  carburetor. At low engine
load,  the  small  manifold supplied the charge, while the large manifold
took over at high loads.   Although a considerable  reduction in HC and
CO was obtained with this device, the driveability of the automobile
was not satisfactory.
                                4-105

-------
4.2.2.4       Stratified Charge Concepts
               The principal objective of charge stratification in
spark-ignition engines  is the achievement of an extension of the lean
limit of engine operation.  In this case, the flame front originates in
a region of high-energy, near-stoichiometric air-fuel mixture, and
propagates from there  into lean zones of the chamber.
               The charge stratification can be accomplished either
by localized fuel injection into a single combustion chamber (open-
chamber stratified charge engines) or by supplying a rich mixture to
a prechamber, and a lean mixture or pure air to a secondary (main)
chamber,  which is connected to the prechamber by means of a com-
municating passage  (divided-chamber stratified charge engines).
               In the open-chamber configuration, exemplified by the
Texaco TCCS (Ref.  4-65) and Ford PROCO (Ref. 4-66) engines,  a
single combustion chamber is  employed similar to that of conventional
spark-ignition  engines.  During engine operation,  an air swirl is set
up in the cylinder by means of directional intake  porting, combined
with special piston cup designs. Fuel is injected into each cylinder
toward the end of the compression stroke.   Upon ignition of the swirl-
ing,  rich mixture surrounding the  spark plug, the burning charge
expands into the  lower  regions  of the  combustion chamber where the
combustion process is  then completed in an oxygen-rich environment.
This is shown  schematically in Figure 4-59.  Attempts are currently
being made to  replace the  more expensive fuel injection systems
employed in these engines by conventional carburetors.
               The divided-chamber stratified charge engines or pre-
chamber engines, exemplified by Honda's CVCC  engine concept
(Ref. 4-67), employ two interconnected combustion  chambers per
cylinder.   During the compression stroke of the piston,  a fuel-rich
mixture is inducted into the generally small prechamber while the
main chamber  is charged with a lean mixture or  even pure  air.  Upon
                                4-106

-------
                           NOZZLE
                                            DIRECTION OF
                                            AIR SWIRL
              SPARK
              PLUG
               1. FUEL SPRAY
               2. FUEL - AIR MIXING ZONE
               3. FLAME FRONT AREA
               4. COMBUSTION PRODUCTS
        Figure 4-59.  Texaco Controlled Combustion System
                      (Ref.  4-65)
ignition in the prechamber, hot gases expand into the main chamber
where combustion is then carried to completion.  The principal
advantage of prechamber engines over conventional engines is their
ability to operate with very lean overall air-fuel mixtures, resulting
in low emissions,  particularly NO .  However,  because of the less
                                 A
favorable combustion chamber surface-to-volume ratio combined
with high turbulence, the heat losses of this engine tend to be higher
in conventional designs.  The benefits in terms  of emission reduction
and fuel  economy improvement that might be realized in a particular
design depend upon the tradeoffs between the heat losses and the
inherently higher thermodynamic cycle efficiency obtained with opera-
tion in the lean air-fuel mixture regime.
                                4-107

-------
               Without incorporation of emission control systems, the
HC and NO  emissions from vehicles adjusted for minimum fuel con-
sumption and  equipped with open-chamber stratified charge engines
are comparable to those of conventional spark ignition engines, while
CO emissions and fuel consumption are substantially lower.  Con-
versely, with emission control, consisting of EGR, oxidation catalysts,
and intake air throttling at low power levels,  the vehicles meet the
statutory 1976 Federal emission standards at low mileage and show
equal or slightly better fuel economy than the average 1973 certifica-
tion vehicles tested.
               Both open-chamber and divided-chamber stratified
charge  engines have demonstrated a lower sensitivity to fuel octane
number.  Because of the late fuel injection combined with immediate
spark ignition, the Texaco TCCS engine can operate on a wide range
of diesel fuels.   The  potential benefits that might  be derived from this
combustion  process include good fuel economy and low NO   emissions.
The stratified charge  engines  developed to date, notably by Ford,
Texaco, and Honda, have achieved an advanced state of development.
4. Z. 2. 5       Exhaust Gas Recirculation
               Exhaust gas recirculation (EGR) as a method of NO
                                                                J\.
control in internal combustion engines has been under evaluation  since
the early 1960s (Ref.  4-68).  Briefly, the  effect of EGR on  NO  is
                                                            X
directly related to the attendant reduction  in combustion temperature.
Since the amount of nitric  oxide produced in the engine cylinder is an
exponential  function of the combustion temperature, even a moderate
reduction in the combustion temperature results in a significant
decrease in the formation  of NO .
                               x
               Experimental and theoretical data relating NO  reduction
                                                           X
to exhaust gas recycle rate are presented  in Figure 4-60 for engines
operating at conventional air-fuel ratios (Ref. 4-69).  The agreement
between prediction and test data is good.   For low recycle  rates,
                               4-108

-------
   BO
  §
  o
  o 60
   c 6
   §
   I '
   s
                  KtWHWL
                	«RCO
                O  AT iO raph
                •  II 10 npM
                	ESSO
                &  PUUOUTH
                4  CHEVROLET
                                 Figure 4-60.  Effect of EGR on NO
                                               reduction and specific
                                               fuel consumption
                                               (Ref. 4-69)
    0   i    10   15   20   25   10
             PERCENT RECYCLE
the reduction of NO  is nearly proportional to the amount of exhaust
                   X
gas recycled.  For higher quantities of recycle,  the effect diminishes.
Substantial (approximately 40 to 80 percent) NO  reductions are
                                              J\.
achievable at 10 to 20 percent recycle rates in the conventional air-
fuel ratio range.   However,  because of the dilution of the charge and
reduced peak combustion temperature, a reduction in power output
occurs (at the same spark advance setting) which effectively translates
into a fuel economy loss, as shown in  Figure 4-60.
               Because of the interrelationship of spark timing,  cycle
temperature, and power output,  it is  possible to advance spark timing
to avoid or minimize the effects of EGR on power and SFC.  In tests
performed by Esso  (Ref. 4-70),  EGR was  shown to have a much lower
fuel economy penalty than spark retard for the same NO  reduction.
                                                      Jt
It was found  possible to operate  with both recycle and some spark
                                4-109

-------
advance and obtain some NO  reduction with a slight improvement
(approximately 2 percent) in SFC in one case.
              For any given engine, then, the fuel consumption penalty
would  be strongly influenced by the baseline engine air-fuel ratio and
NO  emission characteristics,  the amount of NO  reduction reqiired
   X                                           X
to meet a given  standard, and the  potential for optimizing spark timing
and recycle rate within these constraints.
              More recently, a number of EGR systems have been
subjected to extensive testing and  evaluation (Refs. 4-71 to 4-76).
Typical test data from one design  are presented in Figure  4-61
(Ref.  4-75),  showing NO  (in g/mile) and fuel economy obtained on
a 318 CID, V-8  Plymouth engine,  as a function of air-fuel  ratio and
EGR flow rate.  It is of  interest to note that in this case, a small
improvement in fuel economy was  realized by operating the engine
with 4 to 9 percent EGR and an air-fuel ratio between  16 and  18.  Test
data from another study indicate that some improvement in fuel economy
might  be achieved with EGR, particularly when spark  timing is advanced
to maximum power (Ref. 4-77).
              There are indications that the use  of EGR might result
in a reduction of piston ring wear (Ref.  4-78).  In one  engine, incor-
poration of 12 percent EGR resulted in near-zero wear rate after six
hours of steady-state engine operation.
              The major concern  arising from the use of EGR is
related to the occurrence of unstable engine operation, commonly
called  "power surging. "   This phenomenon is attributed to an
increase in the cyclic pressure variation occurring in  the engine
cylinder during operation with EGR and the attendant reduction of the
lean limit of engine operation.  Mixture homogenation  appears to be
a feasible approach to circumvent  this  problem.
                                4-110

-------
                                       0% Recycle Lin*
                                             4% Recycle Line
                167- indicates
                recycle rate, etc.
                                        97. Recycle Line
                                        13% Recycle Line

                                         Region of Poor Combustion
           10
MODES:
                                 18
                                                 20
22
                                                                24
                                   A/F RATIO
                               STOICH.
                             C        B
4.2.2.6
         Figure 4-61.  Test stand NOX emissions as a function
                        of A/F and recycle rate  - 50 mph road
                        load,  37-degree btdc  spark timing,
                        gasoline fuel (Ref. 4-75)
       Water  Injection
               Water injection has been used  since the early years of

kerosene engines as a means of suppressing "engine knock."  More

recently, water  injection has been  considered to reduce the NO

emissions from  internal combustion engines.  The observed reduction

in NO   is due partly to the latent  heat of water and partly due to  a

change  in the specific  heat of the  mixture.  Test data from a single-

cylinder CFR engine operated at constant speed and load are presented
                                 4-111

-------
in Figure 4-62 (Ref.  4-79).  In these tests, the loss  of power due to
water injection was compensated for by increasing the air-fuel ratio
of the charge.  As indicated, water injection at a rate of one pound of
water per one pound  of fuel decreased NO  by about  80 percent, and
                                         X
improved the specific fuel consumption by about two  percent.
               Based on another study (Ref. 4-80), it appears that
water injection may be accompanied by an increase in the volumetric
efficiency of the engine due to evaporative cooling of the  inlet charge
during the induction.
               Water injection directly into the  cylinder has been shown
to increase  the effectiveness of the water, particularly when the water
is injected during the intake stroke.  This is illustrated in Figure 4-63,
showing that only  0. 6 pound of water per  one pound of fuel was required
to obtain an 80 percent reduction in NO  (Ref.  4-81). This improve-
                                      X
ment is accompanied by a slight increase in HC.  Similar results have
been achieved over a wide range  of air-fuel ratios.
              In the past, attempts have been made  to simplify the
water injection procedure by using water-fuel emulsions.  Limited
test data indicate  that the reduction in NO  is about the same  as in
                                         x
the case of water  injection into the inlet manifold.
 yt 60 -
                 O SALTZMAN
                 • PHENOLDISULFONIC
       0.2
0.4  0.6  0.8  1.0  1.2
WATER INJECTION RATE
                         Figure 4-62.  Effect of water in-
                                       jection on the emis-
                                       sions and specific
                                       fuel consumption of a
                                       CFR engine;  5. 5 hp,
                                       1200 rpm,  30° spark
                                       advance (Ref. 4-79)
                                4-112

-------
          Injection Timing
           OIO* BTDC (Companion Stroke)
           6 25* BTDC (Compreiiion Strokt)
           O 45* BTDC (Companion Strokt)
           OJ40-8TDC (Intokt Strokt)
                          3D-
                                          1.5-
                                          10
                                         0.5
         Q3    0.6     09     1.2    1.5
          Watir-Fu*l Ratio, Ibm HjO/ltxn fuel
                                hjecton Timing
                                 O 10* BTDC (Comprewon Stroke)
                                 A 35* BTDC (Comprtuion Stroke)
                                 O 63* BTOC (Comprtuion Strokt)
                                 O340-BTOC {Intokt Strokt)
                                03     06    09    1.2     IS
                                 Wattr-Futl Ratio. Ibm HjO / I Dm fuel
         Figure 4-63.   Effect of water injection on NO and HC
                         concentration,  900 rpm; spark advance
                         30° btdc (Ref. 4-81)
                Water injection test data from an Inger soil-Rand
PKVGR-12, four-stroke, naturally aspirated stationary gas engine
are presented in Figure 4-64  (Ref. 4-82).  The trends of NO  and
                                                                X
HC mass emissions,  as affected by water injection rate, are similar
to the emission trends  obtained in automotive spark ignition engines.
However,  the specific fuel  consumption increases  rapidly with
increasing water flow rate.  The  increase of CO shown in the figure
may be due partly to  the reduction of the  air-fuel ratio and partly to
flame qienching caused by  non-uniformly distributed water spray.
4.2. 3
Fuel Modification
                The prospect of spark ignition automotive engine exhaust
emission control by means of fuel modifications has been the target of
                                   4-113

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                                   Figure 4-64.  Effect of water in-
                                                 jection on emissions
                                                 and performance -
                                                 Inge rs oil-Rand
                                                 PKVGR-12, 4-cycle
                                                 naturally aspirated
                                                 spark-gas  engine
                                                 (Ref.  4-82)
           0,6    1.0   1.5    2.0
            WATER INJECTION RATE. I
numerous investigations for over two decades  (Ref. 4-83).  Most of the
earlier research was aimed at the investigation of the smog-forming
potential of commercial gasoline blends.  More recently, attention has
been directed to the effects of fuel volatility on engine emissions
(Ref.  4-84).  In addition to the effect on evaporative losses from the
carburetor and fuel tank, the volatility of gasoline blends has an effect
on the reactivity index of the exhaust hydrocarbons.  However,  no
significant effect on nitric oxide and carbon monoxide emission has
been found.
              Exhaust emission of polynuclear aromatic hydrocarbons
(PNA) and of phenols has been studied with a variety of test fuels, using
cyclic tests in five vehicles including one without emission control,  two
                                4-114

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with engine modification control, and two with  experimental
very-low-emission systems (Ref.  4-85).  The tests revealed that the
fuel composition influenced emissions both directly and through buildup
of engine deposits.
               NO  and CO emitted from spark ignition engines can be
                 X
influenced  to some extent by fuel composition (Refs. 4-86 and 4-87).
Carbon-to-hydrogen ratio of the fuel appears to be equally influential
as the energy content of the fuel in determining the relative  production
of nitric oxide and carbon monoxide (Refs. 4-87 and 4-88).
               Partial or complete substitution of gasoline by uncon-
ventional fuels (e.g.,  methanol, ethanol,  methane, propane, hydrogen,
ammonia,  hydrazine,  etc. ) is currently being investigated from the
point  of view of exhaust emission reduction as well as economic feasi-
bility (Ref. 4-89).  Some fuels  (e. g. , methane and hydrogen) offer the
possibility of engine operation essentially free of exhaust hydrocarbons
and strongly reduced nitric oxide emission.
               Relatively little  information has been published on the
effects of fuel additives  on  the exhaust emissions  from spark ignition
engines.  Reference 4-90 presents limited experimental  data of the
effect of non-metallic combustible additives and metallic compounds
on NO production.  Another investigation was concerned  with gasoline
additives which form a low surface-tension coating in the engine induc-
tion system,  As a result,  an improved fuel distribution  has  been
achieved,  accompanied by improvements in driveability,  fuel economy
(3 to 4 percent) and HC emissions (15 to 25 percent) (Ref. 4-91).
4. 2. 4          Corrective Emission Control
               A considerable amount of work has been conducted to
date on corrective emission control approaches,  including thermal
reactors and catalytic converters.  These are  briefly discussed in the
following sections.
                                4-115

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4.2.4.1        Thermal Reactors (Lean and Rich)
               A thermal  reactor is  a chamber (replacing  the
conventional engine exhaust manifold) into which the hot exhaust gases
from the engine are passed.   The chamber is sized and configured to
increase the residence time of the gases and permit further chemical
reactions, thus reducing HC and CO concentrations.  In general, the
thermal reactor embodies a double-walled  and insulated (between walls)
configuration, with port liners to direct the exhaust gases to  its inner
core  section.  In some instances,  baffles and/or swirl plates are used
to further promote mixing.
               There are two different types of thermal reactors in
research  and development by several  companies:  the rich thermal
reactor (RTR)  and the lean thermal reactor (LTR).  The RTR is
designed for fuel-rich engine operation.  As a result of the chemically-
reducing atmosphere and lower combustion temperatures in the engine
combustion chamber,  the amount of NO  formed in the engine is
reduced.  If the engine is  run sufficiently rich (air-fuel ratio approxi-
mately 11-12),  it is possible to limit the formation of NO  to less than
2 g/mi; however,  fuel economy penalties at these  rich mixtures are
as high as 20 percent.  As the exhaust from the  cylinders contains
large quantities of HC and CO,  secondary air supplied by a pump is
injected into the reactor to permit further oxidation of these  species.
               The thermal reactor should  be designed for minimum
thermal capacity to minimize cold-start emissions. Since relatively
high temperatures  (1700 - 2000°F) are achieved in the RTR,  high-
temperature materials (e.g., Inconel 601 containing 60-percent Ni,
23-percent Cr,  14-percent Fe,  1-1/2-percent Al) are required for
the inner core,  baffles, and port liners. At these high temperatures,
engine misfiring,  which produces high HC levels,  could lead to exces-
sive local temperatures and material  burnout conditions  in the RTR;
                               4-116

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therefore,  temperature control devices are necessary to protect it.
Ceramic materials, which could  be more tolerant to overtemperature
conditions than metals, have not  demonstrated to date the necessary
thermal and mechanical shock properties.  A summary description of
proposed experimental designs for RTR types is presented in Table 4-9
(Ref. 4-70).
              As mentioned previously, a carburetor calibration change
of three air-fuel units (15 to 12) to minimize NO  to less than 2 gm/mi
                                              X
may incur a fuel economy penalty of 1 5 to 20 percent.  According to
Figure 4-65, if an exhaust gas recirculation system is added to the RTR
to further control NO  to levels below approximately 1 g/mi, fuel
economy penalties are as high as  20 to 30 percent (Ref. 4-69).
      TABLE 4-9.  THERMAL REACTOR SUMMARY (Ref.  4-70)
Reactor
Type
Rich Reactors
Du Pont Type V
Du Pont Type VIl'b)
Esso Synchrothermal
Esso Modified RAM
IIEC/Ford Type H
IIEC/Toyo Kogyo
IIEC7 Nissan
British Small Engine
Lean Reactor
Ethyl Lean Reactor
Induction
Air-
Fuel
Ratio
14
11.5- 12. 5
12.2
11-13
(a)
(a)
(a)
10 - 14
17 - 19
Reactor
Operating
Temperature
(a)
(a)
1600 - 1900(c)
1600 - 1750
1600 - 1850(cf)
1600 - 1800
-------
                S  3
                6
                i
                                 SYSTEM AND SOURCE
LTR.EGRIETHYL PLYMOUTH)
LTRt ECR (ETHYL PLYMOUTH)
LTH. ECR (ETHYL PONTIACI
LTRtEGR (ETHYL PONTIACI
RTR f EGR (OUPONTCHEV I
RTR,EGR (RECENT DUPONT SYSTEM]
RTR.EGR(ESSORAM)
RTR,EGR(ESSORAM|
RTR,EGR» HC/CO CAT CONV
(FORD "MAXIMUM EFFORT" VEH I
RTR, EGR, HC/CO CAT CONV
(FORO "MAXIMUM EFFORT" VEH I
RTR , EGR , HC/CO CAT. CONV
(FORD"MODIFIED MAX EFFORT"VEH)
RTRtEGR(CHRYSLER)
HC/CO CAT CONV t EGR
(FORD PACK "B' I
DUAL CAT CONV t EGR
(FORD PACK "C"l
                                                   DRIVING SCHEDULE
                                                   CITY
                                                   CITY-EXPRESSWAY
                                                   CITY
                                                   CITY-EXPRESSWAY
                                                   CARB CAR POOL
                                                   NOT SPECIFIED
                                                   TURNPIKE
                                                   CITY
                                                   CITY-SUBURBAN

                                                   CVS CHASSIS OYNA

                                                   CVS CHASSIS DYNA
                                                   NOT SPECIFIED
                                                   CVS CHASSIS OYNA
CVS CHASSIS OYNA
                                        GENERAL CORRELATION
                                        ESTIMATED FOR ADDITION OF NO,
                                        CATALYST BED AT 75 PERCENT EFFICIENCY
                    0      5      10     15     20     25     30     35
                          PERCENT SFC INCREASE (CVER UNCONTROLLED VEHICLE )

                 Figure 4-65.  NOX versus SFC increase
                                  (Ref.  4-69)
                 The LTR is used in conjunction with an  engine  operated
on the lean side of stoichiometric mixtures; i.e.,  with excess  air.
Currently LTR systems are limited to air-fuel ratios of approximately
19.   In this  case,  the HC  and CO emissions  are much lower than in the
case of the RTR (but NO   levels are somewhat higher).  Therefore,
                           j£
little chemical heat is generated in the reactor and its temperature
is governed to a large extent by the sensible heat in the exhaust gas.
This means that the oxidation of HC and  CO is accomplished  within
the  LTR at  lower temperatures  than  for the RTR, and without the
requirement for additional air  (i.e. , no air pump is needed).  Because
of the  lower  operating temperatures,  the durability requirement can be
met by less expensive materials for the construction of the reactor core
and baffles; however,  careful attention must be given to minimizing
heat losses  or the conversion is limited  by low reaction rates.   On the
                                      4-118

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other hand, more stringent requirements  exist for engine air-fuel
mixture control and cylinder-to-cylinder fuel distribution.  This may
require utilization of an advanced carburetor or electronic fuel injection.
EGR is generally added for additional NO  control. Although little or
no fuel economy penalty is chargeable to the LTR itself, with EGR an
approximate 10-percent decrease in fuel economy is realized for NO
                                                                  X
levels  of approximately  1. 5 g/mi.  Peak power loss due to lean
operation causes a small loss in vehicle performance.
              The Ethyl Corporation lean reactor is the only known
design of a lean operating system for which specific details of configu-
ration and  performance are available.  It is  designed  for operation at
air-fuel ratios of between 17 and 19.  As shown in Table 4-9, its
operating temperature is 1400 to 1600°F,  or  200 to 300 degrees lower
than those  for rich reactor systems.  The reactor is cylindrical and
consists of an open-tube liner made of 310 stainless steel, surrounded
by a layer  of insulation which in turn is enclosed by an  outer casing of
310 or 430 sheet stainless steel (Ref. 4-69).
4. 2. 4. 2       Catalytic Converters (Oxidizing  and Reducing)
              The oxidation of the HC and CO in the exhaust gas can be
accomplished by means  of a catalyst at temperatures  lower than those
in a thermal reactor.  A catalytic converter can be placed farther from
the  engine  than a thermal reactor and can maintain its effectiveness
without rich mixture engine operation (normally required to  maintain
the  necessary chemical  energy level of the exhaust gases in  a thermal
reactor).   As a consequence,  the fuel economy penalty  is lower.
Generally, secondary air for  oxidation has to be supplied similarly,
as for the thermal reactors; however,  the control of the secondary
air  flow rate is critical,  because overheating due to excessive oxida-
tion reactions in the converter can destroy the  effectiveness of the
catalyst.  For the same reason,  the carburetor of the spark ignition
engine is required to maintain accurately the predetermined mixture
air-fuel ratio, under all engine operating  conditions (Ref. 4-92).

                                4-119

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               The need for more close control of the exhaust-gas
composition in the catalytic converters led to a closed-loop system
of fuel metering.  In such a system, an oxygen sensor is used to detect
the level of oxygen in the exhaust stream and to supply an error feed-
back signal either to an electronic fuel-injection control module or to
a specially constructed carburetor  (Ref.  4-93).  This signal causes
adjustment in the fuel or air supply, thereby maintaining close control
of air-fuel ratio.
               The automotive emission control systems  in develop-
ment to meet the statutory 1976 emission standards utilize two cata-
lyst beds to reduce NO  as well as  HC and CO emissions.  The bed
 1                    x
closest to the engine is used to remove NO  and is operated in a
reducing environment.   Secondary air is then added to the exhaust
stream between the catalyst beds (downstream of the NO   reducing
                                                      X
catalyst bed), and the remaining HC and CO are removed in the
second catalyst, the  oxidation bed.  At least one catalyst manufacturer
has been working on  the development of a tricomponent catalyst  (single
bed) which, under carefully controlled operating conditions, simultane-
ously promotes the oxidation of HC  and CO, and the reduction of NO .
Examples of best low-mileage emissions obtained with dual catalyst
systems are presented in Table 4-10 (Ref. 4-94).
               Lead additives, as well as sulfur and phosphorus content,
are toxic to catalyst  materials,  resulting in a degradation of its effec-
tiveness (Ref.  4-69).  This is illustrated in Table 4-11, which summa-
rizes the results of the most promising durability tests conducted on
some of the dual catalyst vehicles (Ref.  4-94).  The causes of the
rapid deterioration in NO  catalyst  efficiency are not yet quantitatively
                         5C
understood.  However, it is clear that poisoning of the active  catalyst
material by contaminants in the fuel (nonleaded gasolines) is the cause
of part of the observed deterioration.
                                4-120

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   TABLE 4-10.  EXAMPLES OF BEST LOW MILEAGE EMISSION
                 MEASUREMENTS WITH DUAL-CATALYST
                 SYSTEMS ON EXPERIMENTAL 1976 VEHICLES
                 (Ref.  4-94)
Company
American Motors
Chrysler


General Motors




Ford



Vehicle
Weight,
Ib




4500
4500
4500
4500
4500
4000
5000
4000
5000
Engine
Size,
CID
258
360
360
360
350
350
350
350
350
250
351
250
351
EGR

No
No
Yes
Yes
Yes
Yes
Yes
Yes
No
No
No
Yes
Emissions, g/mia
HC
0.27
0. 17
0.23
0. 34
0.24
0.42
0. 17
0.21
0.37
0.45
0.43
0. 52
0. 48
CO
5. 7
3.0
2. 5
3.9
1. 7
3. 1
1.0
1.0
1.8
2.9
2.4
3.7
3.3
NOX
0. 55
0. 5
0. 52
0. 44
0. 15
0.21
0. 19
0.22
0.27
0. 38
0.27
0. 39
0. 39
Catalyst Data
(1) |f§ (2) NOX

(1)
(2) Noble, monolith
(1)
(2) Base
(1)
(2) Noble, pellets
(1) UOP, platinum, pellet
(2) Gulf, monolith
(1) Air products, pellet
(2) Gulf, pellet
(1) UOP, platinum, pellet
(2) Johnson-Matthey, monolith
(1) UOP, platinum, pellet
(2) Johnson-Matthey, monolith
(1) UOP, platinum, pellet
(2) CM, pellet
(1) Engelhard, monolith
(2) Promoted Base, pellet
(1) Engelhard, monolith
(2) Gould, GEM
(1) Engelhard, monolith
(2) Id. pellet
(1) Engelhard, monolith
(2) ICI, pellet
a!975 CVS-CH test procedure. Data were usually averages of several tests and were received up to
November 1972.
Emission-control systems include a manifold thermal reactor before the NOx-reduction catalyst
4.2. 5
Control of Emission from Blowby, Carburetor,
and Fuel Tank
              Positive crankcase ventilation systems have been used

in  automotive  engines  since  1962,  and are designed to prevent the
                               4-121

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TABLE 4-11.
EMISSIONS AS FUNCTION OF MILEAGE
FOR DURABILITY TESTS ON DUAL-
CATALYST SYSTEMS (Ref. 4-94)
Manufacturer,
Vehicle4
General Motors
4500 Ib
350 CID
Chevrolet
EGR
4500 Ib
350 CID
Chevrolet
EGR
4500 Ib
350 CID
Chevrolet
EGR
Ford
5000 Ib
351 CID
Ford
EGR



5000 Ib
351 CID
Ford
EGR

4000 Ib
250 CID
Ford
No EGR
Catalysts
(1) HC (2) NO
CO x

(1) UOP, noble,
pellet
(2) Gulf, noble,
pellet
(1) UOP, noble,
pellet
(2) Johnson-Matthey,
noble, monolith
(1) UOP, noble,
pellet
General Motors
Research, pellet

(1) Engelhard, mono-
lith
(2) monolith

8 - 10 g of
platinum,6 dual-
bed converter
(1) pellet
(2) pellet

Dual-bed
converter
(1) -
(2) Id, pellet


Mileage

0
1, 000
7,000
13,000
0
7,000


0
4,000



Low
3,000
6, 000
9,000
12,000
16,000
20,000
Low
1, 000
2,000
6,000

Low
4, 000


Emissions, g/mi
HC


0.32
0.39
0.52
0.21
0. 47


0. 36
0. 57



0. 3
0. 33
0.48
0.72
0.66
0.66
0. 82
0. 35
0.61
0.59
0.68

0.52
0. 65


CO


1.7
3.0
4.8
1.0
1.8


1.8
4. 1



1. 5
1. 5
2.6
1.9
3.6
5.4
3.8
3.8
3.3
3.6
4.2

3.7
5.2


NOX

0.22
0. 42
0.45
0. 73
0.21
0. 59


0.28
0. 51



0. 56
0. 49
0. 70
0. 89
0. 75
1. 3
1. 5
0.68
0.99
1.25
1.72*

0.39
0. 48


NOX
Catalyst
Efficiency, c
%

78d
58d
55d
27d
79d
41d


72d
49d



78
80
71
63
64
46
37
70
-
_
25

89
86


aEmission Control System includes engine modifications, air pump, NOX catalytic
converter, oxidation catalytic converter, and EGR and manifold reactor where
noted.
1975 CVS-CH test procedure. Data received up to November 1972.
NOX catalyst efficiency is percent NOX removed in catalytic converter.
NOX catalyst efficiency estimated from approximate engine NOX emission of 1 g/mi.
G
Catalyst judged by vendor not to be available in commercial quantities.
EGR system failure.
                      4-122

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escape of blowby gases from the crankcase  to  the  atmosphere  by
recycling them back to the inlet manifold (Ref.  4-95).  Although these
systems offer almost  100 percent control of pollutants from the crank-
case, they affect the engine exhaust emissions  and  the engine perfor-
mance, primarily during engine idling and off-idle  operation.  The
essential part of these systems is a spring-loaded control valve which
is subject to clogging  and malfunctioning,  which may adversely affect
the engine performance and exhaust emission (Ref. 4-96).  Resulting
deposits in the carburetor and increased crankcase corrosion can be
minimized by detergent fuels, high-quality lubricants, and frequent
servicing of the system.
              The design and location of carburetor vents and the
volatility of the fuel have a pronounced effect on the evaporative
HC emissions from carburetors  (Ref. 4-97).  The reactivity or "smog-
forming  potential"  of the  evaporative HC emission is strongly affected
by the composition of the gasoline and decreases with increasing vola-
tility of the initial fuel (Ref.  4-98).  In current automobiles, evapora-
tive losses from  the carburetor and fuel tank are controlled by  means
of an adsorption-desorption device which is  periodically stripped  by
purging the canister with air aspirated into the inlet manifold during
predetermined operating  modes of the engine (Ref.  4-99).
4. 2. 6        Combined Emission Control Techniques
              Based on the above review of preventive and corrective
emission control devices/techniques,  it is apparent that a considerable
number of effective methods and devices have been developed to date
for use in automotive spark ignition engines. It is  conceivable  that
even  more effective  systems  might be possible by optimizing the
design and operation of the various system components.  For example,
a system consisting of EGR, lean-mixture carburetion (about 25 per-
cent excess  of air), and thermal reactor may prove to be a very cost
effective NO   abatement approach.
                                4-123

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4.3            GAS TURBINES
               The fundamental mechanism of the formation of the
various emission species discussed in Section  3. 3. 3. 2. 2 is the same
irrespective of the type of combustor and/or cycle.   The emission
quantities produced and their proportions will vary as a function of the
combustor design, operating conditions, fuel composition,  etc.
               In the  following sections a number of emission control
devices intended for use in automotive engines and steam boilers will
be briefly reviewed and the possible application of these to stationary
gas turbines will be discussed in detail.
4. 3. 1          Automotive Engine Emission Control
               The development  of a low emission automotive power
plant under EPA guidance has proceeded in two directions (Refs.  4-100
and 4-101):  one is the emission improvement in the standard spark-
ignition engine which now dominates and will probably dominate the
automotive field for the next decade; and the other is the work on
Advanced Automotive  Power Systems (AAPS) for the  post-1980  era.
The  latter is currently centered around the Brayton cycle  and the
Rankine  cycle.
               The automotive application of the Brayton cycle consists
typically of a regenerative gas turbine with a free power turbine as a
drive unit.  The AAPS work on gas turbine emissions consists mainly
on the development of an improved  combustor by means of gas recircu-
lation  (Solar, jet-induced circulation) and precise and fine liquid atom-
ization and fuel-air mixing (Aerojet platelet injector); and by two sur-
face combustion concepts represented by porous plate (GE) and
catalytic combustor (EPA/NASA program) concepts (Ref.  4-102).
               Briefly in the jet-induced circulation concept, impinge-
ment of the primary air and fuel  streams generates a recirculation
pattern of partially burned gas in the primary zone which permits
                                4-124

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operation at lean equivalence ratios (< 0. 5), combined with good fuel
atomization by air assistance and lower temperature rise in the
primary zone.  All these factors will have a beneficial effect on the
NO  formation.
   x
               The platelet injector permits formation of a pre-mixed
fuel-air charge with a short residence time in the primary combustion
zone which was shown to be an important factor in NO  reduction.
               Surface combustion involves the promotion of gas-phase
oxidation and reduction reactions between fuel and air in close proxim-
ity to a solid surface (Ref.  4-103).    Low emission characteristics of
this concept, in particular  NO ,  result from  the combustion process
occurring at reduced temperatures.   The surface combustion can be
either catalytic or noncatalytic.  In the noncatalytic  surface  combustor,
such as the  porous combustor, a fraction of the heat of combustion is
immediately transferred from the  flame  layer to the adjacent solid
surface from which it is extracted by cooling means resulting in com-
bustion temperatures below the adiabatic flame  temperature. In the
catalytic surface combustor,  reduced flame temperatures are achieved
by operation with very lean fuel-air mixtures.   The  catalyst serves the
function of promoting chemical reactions which,  under these particular
operating conditions, would occur  too slowly for efficient low-emission
burning.  In the surface combustor,  the fuel must be prevaporized and
premixed with  air before it is fed through the porous or  catalytic
surface, and that limits the application of surface combuetor devices
to fuel gas or light distillates (Ref. 4-103).
               All the emission control devices  described above are  in
various phases of development, those for spark ignition  engines being
most  advanced. Important design and operational details of these  con-
cepts are presented in Section 4. 3. 3. 3.
                                4-125

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4. 3. 2          Stationary Sources Emission Control
               The coal fired utility boilers have been found to be the
top ranking sources of NO   emissions,  accounting for 30 percent of
                         X*
the total stationary sources  (Ref. 4-104).  Gas and oil fired utility
boilers contributed 5 and 4 percent,  respectively.  Consequently, the
combustor in utility boilers,  primarily  coal burners, became the
priority subject of EPA-directed programs initiated in 1970 for the
development of emission controls.  The program consisted of four
major components:  field testing and surveys; process research and
development; fuels research and development; and fundamental com-
bustion research (Ref. 4-104).
               Two parts of  this program are of particular  interest to
stationary gas  turbines:  the fuel research and combustion research.
               Part of the fuel research work is focused on the develop-
ment of burner-combustor combinations which would result in low
emissions with fuel oil and gas  (Rocketdyne,  Dynamic Science).  The
combustion research is grouped into the chemistry of pollutant forma-
tion (Esso,  Rocketdyne, BMI); the aerodynamics and physical factors
affecting pollutant formation (MIT, United Aircraft, JPL); and math-
ematical modeling of pollutant formation (Dynamic Science, UARL,
Princeton University). This work will include a  study of the chemical
mechanics of fuel-bound nitrogen conversion to NO  and a study on the
effects  of flame quenching and recirculation zones on NO  reduction
                                                       .X
(Ref. 4-104).
               A number  of emission control techniques have been
successfully incorporated into several stationary boilers.   For example,
operation at low excess air (LEA) has resulted in NO  reductions of
                                                   ji.
up to 35 percent.  In a horizontally fired boiler the NO  was reduced
                                                     3C
from approximately 700 ppm to 450 ppm by reducing the oxygen level
in the fuel gas  (decrease in excess air)  from 3. 5 percent to just over
two percent (Ref.  4-105).
                                4-126

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              Staged combustion or two-stage combustion consists of
operating the first stage at fuel-rich mixtures and injecting secondary
air into the second stage to complete the combustion process.  In the
first stage, the  NO formation is limited by the unavailability of oxygen.
Removal of heat between stages reduces the combustion temperature in
the second stage thereby kinetically limiting formation of the  NO
(Ref.  4-105).  Two-stage combustion offers about a 50-percent reduc-
tion in NO  and when combined with LEA reductions up to 90  percent
were observed.
              Flue gas recirculation to the combustion zone has the
principal effect  of lowering the  peak flame temperature.  The oxygen
concentration is also lowered and both effects favor the reduction of
NO .   Ref.. 4-104  reports a 70-percent reduction of NO  with gas and
50 percent with  fuel oil.
              Water or steam injection are other potential methods for
NO  reduction.  These techniques are not considered  to be satisfactory
for steam boilers  because of the attendent loss in thermal efficiency
(5 to 6 percent)  and the cost increase associated with steam or water
injection (Refs.  4-104 and 4-105).
              Flue gas treatment,  an alternate method of NO  reduc-
                                                            3C
tion, consists of removal of nitrogen oxides as well as sulfur oxides by
either  catalytic  decomposition or reduction, or adsorption by solids
such as metal oxides.  Principal deterrents to the use  of this method
are the cost and longevity of the various  catalysts under the high
temperature conditions (Ref.  4-105).
              Surface combustion has been considered for possible
application in large utility boilers (Ref. 4-103).  However, the pros-
pects for this concept did not appear bright, even if suitable long life
catalysts could  be developed,  because of the complexity in the design
and the large surface areas required for combustion.
                                4-127

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4. 3. 3         Low Emission Combustors for Stationary Gas Turbines
               The discussion of the gas turbine emissions presented
in Section 3. 3. 3  led to the conclusion that the emissions which are
mainly influenced by the combustor design are NO  and CO.  The
                                                 j£
formation of HC  is closely  related to the formation of CO.  Smoke com-
posed primarily  of unburnt carbon particles is related to HC, and the
same applies largely to particulates.  Thus,  a low CO combustor is
generally expected to have  low HC and smoke emissions.  The SO2
content is a function of the  fuel composition and is almost independent
of the combustor design.   Consequently, the  following discussion shall
be limited to  NO  and CO with the tacit understanding that a reduction
                Jt
of those species  will automatically result in a reduction of all the other
pollutants (except SO£).
               The emission  data presented in Section 3. 3. 3 and the
various emission limits,  existing or proposed, are  summarized in
Figure 4-66,  in terms of specie concentration.  It can be seen that  only
a very small  fraction of the current  gas turbines could meet the  pro-
posed EPA limits for CO and NO ,  or the Rule 68 limits for NOV at
                                X                             X
part load without additional emission control devices such as, for
instance, water injection.  The  situation is even worse for the Rule 67
limits which become progressively more  stringent (when expressed in
NOX   ppm)  for larger plants. Thus,  there is indeed a need for the
development of emission  control devices for  stationary gas turbines and
they will be discussed below.
4. 3. 3. 1      Design Approach
              Since the NOX  and CO are  formed during and immediately
following the period of the combustion process taking place in the prim-
ary zone of  the gas turbine  combustor,  it is the design of this zone  and
the fuel preparation on which the attention must be focused.   To  lower
the temperature  of the primary zone, the combustion must take place at
                                4-128

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  1000
i
a
O
   100
   10
         EXISTING CAS TURBINE EMISSION RANGE -
       PROPOSED EPA LIMITS
       {plant size - SO x 106 Btu)
               10        100
                NO , ppm
Figure 4-66.  Gas  turbine state-
              of-the-art emissions
              (No. 2 GT turbine
              oil;  15 percent 0,,)
                                1000
fuel-air equivalence ratios either well above or below unity.  Operating
at fuel-rich conditions would require subsequent cooling of the  gases to
avoid high temperatures in the "afterburning" period.  It would also be
conducive to formation of CO and it would require greater combustor
volume.  Consequently,  the leaner than stoichiometric fuel-air ratio
approach in the primary zone is  usually selected by most researchers,
although the opposite is true for  low emission steam boilers.
               Figure 4-67 illustrates the reduction of NOX as  a func-
tion of combustor inlet temperature (T^n) and fuel-air ratio which
jointly determine (for homogeneous mixtures) the flame temperature.
The range of T±n = 500 to 800° F  will correspond to a simple cycle,
while T.  = 800 to  1500°F will correspond to a regenerative cycle.
Fortunately,  in the actual combustion process the quantity of NO is
kinetically limited because of the short residence time of the combustion
gases,  as  illustrated in Figure 3-27.  Extrapolation of the data shows
that if the  flame temperature could be maintained below 3000°F and the
residence  time  below 10 milliseconds,  the NO   concentration would be
                                             X.
less than 25 ppm assuming homogeneous fuel-air mixture burning  in
the combustor.   The data exclude "prompt" NO or fuel bound NO effects
                                 4-129

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  6000
   5000
 K
 D
   4000
 a. 3000
   2000
 < 1000
I
 • 10,000
   1000
o
u
o
z
a
o
3
O
UJ
   100
    10
         0.05
                            1500  F
              0.033   0.025
              AIR/FUEL RATIO
                          0.02
                               0.017
Figure 4-67.  Flame temperatures
              and equilibrium NO
              concentrations as
              functions of air-fuel
              ratio for various
              inlet  temperatures
              (Ref. 4-106)
which are not greatly temperature-dependent.   Nevertheless, the
indication is clear that primary combustors  operating at fuel to air
equivalence ratios of 0. 5 or less,  homogeneous fuel-air mixtures and
short fuel combustion times will form the base for the first two cri-
teria leading to a low emission combustor.   In order  to approach this
goal, liquid fuel must be well atomized for quick vaporization and mix-
ing with the air before ignition.
               This leads to the next two criteria for  low emission com-
bustion, i.e.,  fine atomization and premixing.   With  gaseous fuels
there is,  of course,  no need for atomization but there is  still the need
for intimate  mixing with air before combustion.  Compliance with the
criteria indicated above will lead,  for instance,  to the design of a com-
pact, prevaporized,  premixed, well-stirred combustor as discussed in
                                  4-130

-------
various references (for instances, 4-107) for automotive,  aircraft, and
stationary gas turbines.
               A design and a successful development of such a com-
bustor is a complex task.   In addition to the consideration of emissions,
there is  also the overall cycle efficiency problem related  to the addi-
tional pressure drop in the combustor required for air blast atomization,
mixing and recirculation.   Also,  there is a problem  of prevention of
combustion flameouts by operating the primary zone within certain
fuel-air  ratio limits as shown,  for instance, in Figure 4-68.  Further-
more,  in order to prevent preignition in the premix chamber the
mixture  residence time in the premix chamber must be longer than the
evaporation time of the fuel droplets but shorter than the  mixture igni-
tion delay (Ref. 4-106). Since the evaporation time of small droplets
is typically of the order of  5 ms,  and the fuel-air mixture ignition delay
is about  20 ms (at 1500° F),  the residence time in the premix chamber
must be  balanced between the two values leaving as great  a time  margin
as possible to prevent preignition (Ref. 4-106).  All  this points to a
  2000
                                    Figure 4-68.
          0.05
0.025   O.OIT
AIR/FUEL RATIO
                           0.012
                                0.01
                                   Limits of flamma-
                                   bility of a paraffin
                                   hydrocarbon
                                   (CnH2n+2) showing
                                   the influence of in-
                                   let temperature
                                   (Ref. 4-106)
                                 4-131

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very careful flow balance which,  for automotive gas turbines operating
at a variety of loads,  may not be possible without some variable control
of primary or  secondary air flow, and which for stationary units
designed for long life will require extensive development and testing.
               To comply in the interim with the emission limitations,
the stationary  gas turbine manufacturers have proceeded in two direc-
tions:  (1) introduction of modest modifications in the  combustor to
reduce the emissions as  much as possible without prolonged develop-
ment work, and (2) water or steam injection into the primary zone of
the combustor.  The effect of water injection  will be dealt with in
Section 4. 3.4 while the effects  of interim combustor modifications and
the farm term development  plans are discussed below.
4. 3. 3. 2       Interim Combustor Modification
               Of the modest combustor modifications aimed at emission
reduction,  the following few can be quoted as  more  significant than
others:  air-blast or air assist atomization; leaning of the primary
zone; early flame quench; and flow recirculation (reverse flow  in the
primary zone).  Application of these modifications to  various com-
bustors produced different results due probably to different flow pat-
terns,  different primary combustor designs and different degrees of
atomization.  Air-blast atomization has been shown to be helpful in
smoke reduction, and this effect was generally confirmed by various
stationary  gas turbine  manufacturers and users (Ref.  4-108).   The
effect of the air-blast or air assist approach is related to improving
the quality of liquid fuel atomization by reducing the mean Sauter diam-
eter from about lOOfi for pressure injection to less  than 40(0. for air
atomization (Ref. 4-107).  This results in a reduction of the evaporation
and combustion time,  and leans the locally  rich hot  fuel pockets which
are smoke and carbon  producers.   On the other hand, Ref. 4-108
stated that changes that decreased NOX increased the CO emission and
vice versa.
                                4-132

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               Ref.  4-109  reports  test  results  with reduced  flame
residence time accomplished by moving the upstream air dilution holes,
in order to direct more air and partially burned recirculation gas into
the primary  zone.  The effects of these changes are illustrated in Fig-
ure 4-69, showing NO  for the standard combustor, with 25-percent
reduction in  flame residence time (Curve II),  with primary zone  leaned
to equivalence ratios  of about 0. 5 (Curve III) and 0. 3 (Curve IV), as a
function of the temperature rise in the combustor.  It can  be seen that
the NO  reduction is more pronounced in the high temperature rise
      X
regime  (above 1000°F).  Similar tests performed in scaled combustors
operated at various combustor exit temperatures  (operational tempera-
ture is  1800°F) show  more moderate  reductions in NOX.  According to
Figure  4-70, showing laboratory model data,  progressively leaner
primary zones (Mod 1, Mod 2) resulted in NO reductions of the
                                              2C
order of 10 to 20 percent.  Similar reductions were obtained with fuel
oil.  In these tests some reduction in CO and  HC was  observed with
the leaner  primary  zone designs.   The effect  of primary zone gas
recirculation is shown in Figure 4-71 for natural  gas,  showing that the
NOX reduction becomes more pronounced at higher temperatures.
  240
  220
  200
  180
  160
 E MO
 *'»
 ° 100
   80
   60
   40
   20
 I
 II
 III
. IV
I    I     I    I    I    I
W25I-AA PRODUCTION COMBUSTOR
REDUCED RESIDENCE TIME
LEANED PRIMARY ZONE WITH REDUCED
RESIDENCE TIME
VERY LEAN PRIMARY ZONE WITH  ./I
REDUCED RESIDENCE TIME
       200  400   600   BOO  1000  1200 1400
        COMBUSTOR TEMPERATURE RISE, °F
                                 Figure 4-69.
                                            NOX variation with
                                            temperature  rise
                                            for production and
                                            modified combustors
                                            (Ref.  4-109)
                                 4-133

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   120
2
I
 »
UJ
o
x
o
o
Of.
    80
   60
   40
    20
       CONDITIONS
       AIR FLOW: 1.95 Ib/sec
     _ AIR INLET: 560°F
       COMBUSTOR PRESSURE: 60 psla
       FUEL: NATURAL GAS
       NOZZLE: D . 30 . 5 . 6
  CURVE

-   I

    II
COMBUSTOR

6 in, STANDARD

6 in. MOD 1

6 in. MOD 2
          800   1000  1200  1400  1600   1800  2000

           COMBUSTOR EXIT TEMPERATURE, °F
 Figure 4-70.
               Effect of primary zone lean-
               ing on NO emission — natural
               gas  (Ref.  4-109)
   100



    80



    60



 9  40

 S
 O
 u  20
w
s

TJ

E
a.
a.
        I     I      I     I
  CONDITIONS

  COMBUSTOR: 6 In. STANDARD
  FUEL: NATURAL GAS
  AIR FLOW: 2 Ib/sec
  AIR INLET TEMP.: 575°F
  COMBUSTOR PRESSURE: 60psla
  CURVE I: NO RECIRCULATION
  CURVE II: RECIRCULATION
          (02 = 19.2%)
              I
            800   1000  1200   1400  1600   1800

            COMBUSTOR EXIT TEMPERATURE, °F
                                            2000
Figure 4-71.
               Effect of cooled exhaust gas  re-
               circulation on NO emission —
               natural  gas (Ref. 4-109)
                     4-134

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Again, similar trends were observed  with fuel oil.   The effect of
recirculation on CO and HC was negligible (Ref. 4-109).
               Thus, summarizing the possibilities of interim com-
bustor modifications,  it appears that  the  modest modifications of
standard combustors, while effective in reducing NO  (while slightly
increasing or not affecting  the CO emissions) will not be  sufficient to
meet any of the regulations quoted in Section 4. 3. 3. 1 at base or near
base load conditions.  Figure 4-66 indicated that NOX reduction of
over 70 percent is necessary for the average state-of-the-art gas
turbine to meet Rule 68  or  the proposed EPA limit.   However,  approx-
imately half of the required improvement might be achieved by modifica-
tion of the current combustors (air atomization, lean primary zone
and gas recirculation).  This applies both to natural gas and fuel oil
since tighter limits for NO  are applied to gas.
                          .X
              Of course,  emission reduction is desirable,  and
incorporation of these combustor modifications would permit,  for
instance,  a  lower  rate of water injection to bring the NOX emissions
within the specified  limits.
4. 3. 3. 3        Low Emission Advanced Combustors
               Most of the work on advanced gas turbine combustors is
being performed on  aircraft and automotive gas turbines, and the previ-
ous discussion on the  design approaches to advanced  low  emission com-
bustors was  based on  work on these two types of gas  turbines.  The
stationary gas  turbines will directly benefit from this work, but it
should not be confined to it since the restrictions on weight and size
always associated with mobile applications of gas  turbines do not apply
to stationary power  plants.  Most of the domestic  stationary gas
turbines descend from their aircraft predecessors and thus inherited
their in-line compactness  which, while indispensable in an aircraft or
automobile,   is of secondary importance in a stationary power plant.   It
                                4-135

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is quite possible, for instance, that an externally (i. e. , not in-line
between the compressor and turbine) mounted combustor could possess
more flexibility in the incorporation of emission reducing features than
a combustor squeezed between the compressor and turbine whose length
is kept as  short as possible to preserve the required stiffness of the
shaft and housing.  Thus, while the results presented here are obtained
either for  aircraft or automotive  gas turbines, it should be kept in mind
that similar or better results are being obtained, although they are not
yet published,  on externally mounted combustors.  Such combustors
are being built by Brown-Boveri  for use in Turbodyne Corporation
power plants and, reportedly,  such plants will meet Rule  68 standards
with "dry" (no water injection) operation.
4. 3. 3. 3. 1      Solar  Combustor
               One of the AAPS advanced low emission combustors is
being developed by Solar in the form of its jet induced circulation (JIC)
combustor (Ref. 4-102 and Ref. 4-107).  The schematic of the primary
zone of this combustor is shown in Figure 4-72.  It  contains all the
elements of an ideal  well-stirred reactor. Air blast atomization of
liquid fuel  assures small droplet  sizes which are vaporized and mixed
with air in a premix  chamber.  The fuel-air mixture emerging from
the premix chamber  acts as an ejector inducing a high degree  of recircu-
lation required for stability at lean equivalence ratios (<0. 35 in the
primary combustor. ) Through this entrainment  of hot products, the
fuel-air mixture in the jet is effectively ignited.  In its alternate
design Solar utilizes  eight primary air and fuel jets  which are inclined
by 45° in the upstream direction.  The jets impinge  in the center of the
combustor thus providing a motive jet for the  recirculation flow.
               The emissions from this combustor,  operated with
kerosene fuel, were  typically less than 1.0 Ib NO /1000 Ib fuel
                                                !X
(<10 ppm), less than 0.5 Ib CO/1000 Ib fuel (<10 ppm) and less than
                                 4-136

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               1.4.-
               1.2 -
             I
           ALLOWABLE LIMIT

        COMBUSTOR INLET TEMPERATURE 1000°F
        COMBUSTOR INLET PRESSURE 30 psig
               1.0 -
01 '
en
m°-8
o
V)
i/>
— n 6
3 u. o
UJ
z 0. 4
0.2
0
MAIN FUEL
INLET . EXHAUST
1 , T
f| I ^ RECIRCULATING FLOW
1 ^ ^_ .S
— 1 Lb ^ ^-* y
^^ \
FUEL * \
MIXING \
ZONE VLAME
ZONE
1 1 • 1







                0.016       0.018        0.020       0.022
                      AIR/FUEL RATIO (from C02 measurement)

           Figure 4-72.   High  recirculation stabilized lean
                          primary zone combustor  schema-
                          tic and NO2 emission (Ref. 4-107)
0. 5 Ib HC/1000 Ib fuel (<20 ppm).  If such a combustor could be
developed for stationary gas turbines, it would meet all of the cur-
rent regulations with "dry" operation.
4.3.3.3.2
GM Combustor
               Another advanced combustor (External Recirculation
Combustor) under development at GM for automotive engines is
described in Ref. 4-110 and is shown in Figure 4-73.  This  combustor
incorporates all the criteria discussed before: lean primary zone with
flame temperature  not exceeding 2400° F; short residence time; fuel va-
porization and premixing upstream of the combustor.  The flame stabil-
ity and further emission control are  enhanced by hot gas recirculation.
                                 4-137

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             6
              fir
L
\J
                                                 VARIABLE
                                                GEOMETRY
 PRIMARY
COMBUSTION
  ZONE
     Figure 4-73.  External recirculation combustor (Ref. 4-110)

The effects  of this recirculation on the NO  and CO emissions are
                                         .X
shown in Figure 4-74.  The level of NO  is drastically reduced  and
                                      JC
while the emission of CO increases with increasing recirculation,  the
maximum level is still very low (less than 20 ppm). Above one-half
load, the NO level is less than 2 lb/1000 Ib fuel (<20 ppm), CO is less
than 4 lb/1000 Ib fuel (<70 ppm), and HC is less than 0.5 lb/1000 Ib
fuel.   It should be noted that this combustor,  as well as Solar's  com-
bustor, require air flow controls (variable geometry) to meet the
varying load demand of automotive engines.
               The emissions of the GM-DDA (Detroit Diesel Allison)
advanced combustor  and their comparison with today's  conventional gas
turbines was illustrated in Figure  3-36.  Of course, the advanced  com-
bustor data are only  rig test results and there is a long term develop-
ment work ahead before they become characteristic of an operational
gas turbine.
               Extensive  investigation of low emission gas turbine com-
bustors was undertaken by DDA for an U. S. Army application in a
light duty helicopter  (Ref. 4-111).   The work consisted of an interim
modification of the combustor along the lines discussed in Sec-
tion  4. 3. 3. 2 and of a preliminary evaluation of certain design features
                                4-138

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                  NOTES:
                  CYCLE POINT 3 (1308°Fbit, 22 psia)
                  PZ TEMP. * 1820 - 1840°F
                  F/A OVERALL =0.0058

                  HYDROCARBON LEVELS WERE LESS
                  THAN 0.4 ppm AT ALL  RECIRCULATIONS
                     35
40   45   50   55   60
   RECIRCULATION, %
70
               Figure 4-74.  Effects of recirculation on
                             emissions (Ref.  4-110)
for an advanced combustor.  The  interim combustor modifications
resulted in approximately 50 percent emission reduction.  The long
range design changes included introduction of a pre-chamber in which
the fuel was prevaporized and premixed.   This was similar to the
approach taken by the other investigators discussed in this section.

4. 3. 3. 3. 3      Ford Combustor

               Comprehensive analyses of advanced  low emission
combustors were performed by the Ford  Motor Company and test data
                                4-139

-------
confirmed the soundness of the analytical approach (Refs. 4-106 and
4-112).  Figure 4-75  shows a  schematic of Ford's  advanced low
emission combustor (Externally Vaporizing Combustor,  EVC) which
incorporated all the previously discussed features.   The steady state
CO and NO emissions of this  combustor were less than  1 lb/1000 Ib
           X
fuel at temperatures of 3000°F (Ref.  4-112).  The level of HC was
insignificant.
4.3.3.3.4    Aerojet Combustor
              Another approach to achieve quick  fuel vaporization and
fuel-air mixing is the Aerojet Liquid  Rocket Company platelet gas
turbine (PGT) combustor (Ref.  4-102).  In this concept, fuel is uniformly
injected from many small orifices into the throats of atomizing Venturis
through which the primary zone air is passed.   The  resulting finely
atomized spray  is rapidly vaporized and mixed with the primary air
before combustion at low equivalence ratios with subsequent rapid
quenching of the flame by the secondary  air.   The combustor emissions
FUEL  ^
FLOW
 TOTAL
 PRIMARY
 ZONE
 AIRFLOW
                                               DILUTION
                                               AIR
                                   PRIMARY
                                  -STABILIZATION
                                   ZONE
                                                  O
                                                  O
                                 AIR	 /  	
                                 FUEL MIXTURE,
              Figure 4-75.  Schematic of a low-emission
                            combustor concept — Ford
                            externally vaporizing com-
                            bustor (EVC) (Ref. 4-106)
                                 4-140

-------
achieved to date show NO  from 0. 5 to 2 pounds/1000 pounds fuel and
CO from 4.5 to 1.7 pounds/1000 pounds fuel, for primary zone fuel-air
ratios of 0. 035 to 0. 045,  respectively.
4.3.3.3.5     GE Combustor
              A different type of  advanced low emission concept is the
surface combustor  consisting  of  either porous plate or  catalytic
surfaces (Ref.  4-103).
              The porous plate combustor is under development by
GE. As previously noted, this combustor is  aimed at operation below
the adiabatic flame temperature.  A schematic of a porous plate com-
bustor is shown in Figure 4-76.  In this design, the fuel is gaseous or
vaporized before being  percolated through the porous layer.  Flame
flashback was one of the problems occurring  in this combustor. This
problem was alleviated by lowering the equivalence ratio from 0. 9  to
0. 7 and employing more effective  cooling of the porous  bed.  The
measured NOX emissions were between 0. 1 to 1 pound/1000 pounds
fuel, and CO between 10 to 40 pounds/I 000 pounds fuel,  the larger
values corresponding to surface air velocities above 20 m/sec.  This
combustor  is applicable only in conjunction with premixed and pre-
vaporized fuels.
4. 3. 3. 3. 6     EPA/NASA Combustor
              Catalytic surface combustors currently under develop-
ment by EPA/NASA operate at very lean fuel-air  ratios and low flame
temperatures as determined by the catalyst bed material.  For instance,
•y-alumina is limited  to 1750°F (Ref.  4-103) while some  of the new
ceramic materials of monolithic structure have a temperature capability
up to 2400° F.  The ideal  range of operation of catalytic combustors is
between 2000 to 2700° F since a rapid decrease in combustion efficiency
occurs at catalyst temperatures below 2000° F (Ref.  4-108).  The
emission of "hot gas" NOX at the  low temperatures (<2500°F) will be
                               4-141

-------
         Figure 4-76.  Transpiration combustor (Ref.  4-102)

significant, but the formation of "prompt"  and fuel-bound  NO   in
catalytic combustors is open to investigation.  Formation of CO and
HC is likely to occur under low temperature conditions.
               The development work on surface combustion is in the
early phase but catalyst durability, long life efficiency and mechanical
integrity are emerging as main problem areas.
                                4-142

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4.3.4         Water/Steam Injection
              Water (or steam) injection in stationary gas  turbines is
primarily used as a means of NO  reduction to permit meeting either
the Federal or local regulations. Its effect on  CO, HC, and  smoke
varies depending on the particular combustor configuration and the
mode of injection.  Water  injection detracts from the siting flexibility
of both  simple cycle and regenerative cycle gas  turbines in that supply
of water has to be assured in quantities about equal to the liquid fuel
supply.  Furthermore, to  avoid detrimental effects on the turbine
durability, the water has to  be  purified to a maximum of 5  ppm
(preferably 1 to 2 ppm) of dissolved solids.  This additional cost of
installation of the water injection equipment, the water treatment plant,
and the water itself has to be considered in the economics of  stationary
gas turbines.
              For these reasons, water injection has to be regarded
as an interim step in stationary gas  turbine emission control until such
time when advanced "dry"  combustors are fully developed.
              The kinetics of water  injection on NOX formation is
treated in Ref. 4-113.  The  main effect of water is the reduction in the
flame temperature by the  heat required to heat and vaporize the water
(which amounts to approximately 30  Kcal/mole).  The dissociation
energy  (^O — H + OH) constitutes only approximately 2 percent of the
heat absorbed (Ref. 4-113).
              The chemical effect of water injection is minimal.
Early in the combustion stage there  is a temporary excess of O, OH
and H radicals above their equilibrium concentrations and the small
amount of dissociated water (
-------
               As previously stated, the "hot gas" NO  formation is
controlled by the Zeldovich reactions,  particularly the reaction
N2 +O —NO + N, which is highly temperature-dependent.  For instance,
reduction of the flame temperature by  200° F reduces the rate of NOX
formation by a factor of three (Ref. 4-114).  From the analytical work
performed in Ref.  4-113,  a  flame temperature drop of approximately
350° F was calculated for a water-fuel  flow ratio of unity and a combus-
tion  zone  equivalence ratio of 0. 8.  Thus, theoretically even greater
NO  reduction rates might be possible.  In reality, the reduction is
less than  theoretical because the  water which is atomized  into droplets
of varying size is not uniformly vaporized and mixed with  the gas. The
rate of water droplet vaporization is important.  Too slow vaporization
may produce peak temperatures similar to dry temperatures  which
would enhance NOX formation,  while very fast vaporization might
cause an initial reduction of the temperature which would  enhance the
formation of CO and HC.  The baseline case  considered in Ref. 4-113
was  with 60|j. diameter water droplets which were vaporized in
6 x 10"  seconds.
               Another reason for lower than theoretically predicted
reduction in NOX with water  injection is in the fuel bound  nitrogen and
"prompt"  NOX which may be less temperature dependent than the "hot
gas"  NOX.  Thus, one would  expect that natural gas,  which is almost
nitrogen-free, should show greater NOX reduction with water injection
than fuel oil which,  in turn,  would look better than heavy oils and
residuals.
               Since the heat of vaporization is about one-third of the
total heat absorbed by water, steam injection requires greater flow
rates than water to achieve the same reduction in flame temperature
but the better mixing achieved with steam tends to reduce  the difference
in the flow rates.   This  will  be shown from some of the industry data.
The amount of heat expended on water  vaporization is  not  recovered  in
                                4-144

-------
the cycle, hence reducing  the useful heat input into  the turbine.
Consequently, a drop in thermal efficiency is expected with water
injection while steam injection, which increases the mass flow of the
gas turbine without additional compressor work,  should improve the
thermal efficiency (assuming the steam is provided "free").
              Ref. 4-115 quotes a drop in the efficiency of  approxi-
mately  one percent for a water injection rate of one  percent of the air
flow (one  percent of the  air mass flow rate is equivalent to  about
60 percent of the fuel flow rate), but the power available increases by
3. 5 percent for every  one percent of water because of the higher mass
flow.  Ref.  4-116 quotes a larger loss in efficiency with water injection,
three percent per one  percent of water, and a gain in efficiency with
steam of similar magnitude.
              As stated before,  the  effect of water injection depends
greatly  on the manner in which it is atomized and injected into a com-
bustor.   The effectiveness of water injection can be  assessed by a
water effectiveness factor suggested in Ref. 4-117.

                  ,,   ,.         (NO wet/NO dry) calculated
           water effectiveness = U.^	 JMAJ  V	3—
                                (NO wet/NOdry) measured
It was determined in Ref. 4-117 that the water  effectiveness factor was
67 percent for gas fuel and  51 percent for fuel oil, probably due to
bound nitrogen in the latter as well as better  mixing in the former.  As
an example,  Figure 4-77 shows a system where the  water is injected
shower-like from a toroidal ring around the fuel nozzle.  Figures 4-78
and 4-79  show the reduction in NO (expressed as NO ) for a gas tur-
                                                   jC
bine with 1 percent water injection burning fuel oil and natural gas,
respectively.  The upper curves indicate the NOX emission with an
original unmodified combustor, which was  subsequently modified by
leaning  out the primary zone and by  reducing the residence time. This
was insufficient to meet the Rule 68  NO  limits.  However,  by adding
                                      ji
water injection NO  was reduced by  38 percent for fuel oil and by
                               4-145

-------
  WATER
  INJECTION
  RING
                 - VORTEX GENERATOR

                 FUEL NOZZLE
                 COMBUSTOR CAP
                                  COMBUSTION
                          7
                       VORTEX
                       GENERATOR
                               COMBUSTION
                               LINER
Figure 4-77.
                  Schematic of water-
                  injection system
                  (Ref.  4-117)
  200


   ISO



  100

   60
 o
   40
   20
                             OIL-DRY
          ORIGINAL COMBUSTOR
     MODIFIED COMBUSTOR

          RULE 68 LIMIT
                  MODIFIED COMBUSTOR + t% H2O
        I    I   I    I
                       I   I    I   I
     02   46   8   10   12  14   16  18  20
                  LOAD, MW
Figure 4-78.  Effect of water injection
                on NOX  emissions from
                MS 5001 turbine - liquid
                fuel (Ref. 4-117)
                  4-146

-------
 100
  80

  60
a

if"
  20
  10 -
       \   1   I   I
                           GAS-DRY
  ORIGINAL COMBUSTOR


MODIFIED COMBUSTOR

     RULE 6B LIMIT
             - MODIFIED COMBUSTOR + 1% H20
                                     Figure 4-79.
NOX emissions from
modified combustion
system in MS 5001
gas turbine — gas
fuel (Ref. 4-117)
             6   8   10  12  14  16  18 20
                LOAD, MW
45 percent for gas fuel and the limits were met.  CO, HC,  and smoke
emissions were essentially unaffected by water injection and remained
at a very low level.
               Ref. 4-118 reported results  with various  water injection
schemes.  Injecting water upstream of the  fuel injector  was successful
in reducing NO but CO increased by about 200 percent.  Injection
of fuel and water in a concentric injector was  successful not only in
reducing NOx, with CO being essentially unaffected, but  also in requir-
ing only approximately one-third of the water  flow  rate of previous
schemes.  The emission data with this system are  shown in Figure  4-80.
As indicated, at a water-fuel ratio of 0.8, the NO  emissions  were
reduced by approximately 80 percent.
               Figures 4-81 and 4-82 show NO  reductions  achieved
with water and steam for oil and gas  fired gas turbines.  In the case of
                                4-147

-------
             0.2     0.4    0.6     0.8
                  WATER/FUEL RATIO
                         1.0
Figure 4-80.
Effect of water injection on
emissions — 5001K gas turbine
engine combustor with fuel oil
(Ref. 4-118)
     a
     z
       i.o
       0.4
     §0.2
               10
                           STANDARD SYSTEM
                           1% WATER
                           IN REACTION ZONE
      20     30
      OUTPUT, MW
                                  40
                                         50
Figure 4-81.  NOX  reduction by water-injection,
               oil-fired Model 5000 Engine,
               iso-conditions (Ref.  4-116)
                   4-148

-------
Q. ft fl
« u« o
ffl
*o
S
~ 0.4
1
V)
!2 0.2
u
o" o
I I I
/DRY
/
/
/

1









_
/ 2.0% STEAM IN COMPRESSOR
/ /DISCHARGE



X^ .X 4-°* STEAM IN COMPRESSOR
^^ DISCHARGE
**l 1 1
Z 0 10 20 30

1
40


SO
                                 OUTPUT, MW
               Figure 4-82.  NOX reduction by steam-
                             injection, gas-fired MS-
                             5001L engine,  site con-
                             ditions (Ref. 4-116)
fuel oil a 50 percent reduction of NOX was obtained with one percent
water while the same  reduction required about two percent steam.
According to Ref. 4-109, six percent steam is required to achieve
55 percent NOX reduction while 2. 5 percent steam resulted in a NOX
reduction of about 20 percent. In this case,  12 percent steam increased
CO by approximately 200 percent, with a similar increase in HC.
               Different results were obtained in Ref.  4-119 where
water injection reduced all  emissions.   The effect on NO  is shown in
Figure 4-83.  Figure 3-34 shown  previously, illustrates the reduction
of CO; Figure  4-84 shows the effect of HC, and Figure 3-40 shows the
smoke reduction.  These latter data  illustrate the effectiveness of
water injection in a combustor in which good matching with water
injection rate and mode was obtained.
               Another way of introducing water is emulsifying it in
liquid fuel,  but reportedly low water effectiveness was observed with
that scheme.
               The use of water injection is now applied in many utili-
ties since it is at present the only known means to enable the state-of-
the-art gas turbine combustors to meet the various NOX standards.
                                4-149

-------
 iS '-°
 *« °'9
 5 * 0.8
 55 0.7
 I I-
     0.6
= i  0.5
 xQK 0."
if  0.3
o g  0.2
5 tj  o.i
  T   0
                      I  I  I I  T  I  II
         I I  I  I  I  I I  I  I  I  I  I I  I  I  I
                                       Figure 4-83.
       0   0.2  0.4  0.6  0.8  1.0  1.2  1.4  1.6
             WATER/FUEL FLOW RATIO
Combustion
laboratory NO
reduction with
water injection
(Ref. 4-119)
0.90
E 0.80
^0.70
zO.60
xO.50
Ul
<0.30
I0.20
0.10
0
	 WITHOUT WATER INJECTION^/
^\
WITH WATER INJECTION -1
-
i I 1 I 1 I 1
            10   IS   20   25   30
             GENERATOR OUTPUT, MW
                              35
                                    40
                                        Figure 4-84.   Total HC ver-
                                                       sus  load, W-251
                                                       engine
                                                       (Ref. 4-119)
Up-to-date experience does not indicate any deleterious effect of water
on turbine life provided the dissolved solids are kept below 5 ppm.  If
steam injection is used,  care must be taken to maintain it in a super-
heated condition up  to the injection point to prevent condensation and
slugs of water entering the combustor.  At least one  case of turbine
bucket damage resulting  from water condensation has been  reported.
               In summary, in the case of fuel oil,  water injection with
rates less than  or  equal to the fuel  flow rate can  reduce the NO
emissions by 35 to 50 percent in low emission combustors and 50 to
75 percent in a standard  combustor.  Greater reduction (60 to 90 per-
cent) can  be  expected with natural gas (Ref. 4-113).  The level of CO
                                4-150

-------
and HC can either be reduced, remain the  same, or increase
depending on the way the water is injected.  If steam injection is used,
higher (by  a factor of 2) flow rates than water are required and  an
improvement in the thermal efficiency of the cycle can be expected.  A
summary of the effectiveness of various emission control techniques
is presented in  Table 4-12.
4.3. 5
SO  Emission Control
  x
              It was stated before that sulfur oxides are formed from
the sulfur in the fuel.   To comply with the proposed regulations the
amount of fuel sulfur for liquid fuels should be <0.8 percent by weight.
Estimates by fuel specialists indicate that this limit can be met by all
of the available distillates produced in this country, by 75 percent of
         TABLE 4-12.  EFFECTIVENESS OF VARIOUS GAS
                        TURBINE EMISSION CONTROLS
Design
Modifications
Inte rim
Primary Zone
Leaning
Combustor Gas
Recirculation
Water/Steam
Injection
Advanced Combustion
Premixed,
Prevaporized,
Well-Stirred,
External
Combustors
Emissions
NO
X

10 - 30%
reduction
~30%
reduction
50 - 75%
reduction
(oil)
60 - 90%
reduction
(gas)

20 ppm
achievable
CO

Small
reduction
Negligible
effect
Some
reduction or
increase

40 ppm
achievable
HC

Small
reduction
Negligible
effect
Small
reduction or
small
increase

5 ppm
achievable
Smoke

Reduction
Negligible
effect
Small
increase
possible

Invisible
                                4-151

-------
the crudes, 50 percent  of the blends, and about  15  percent of  the
residuals (Ref. 4-120).  If more stringent limits for SO  are contem-
                                                     JC
plated,  or if more of the crudes,  blends and residuals are to be used
for stationary gas turbines,  then two ways are open for SOX  reduction:
one is improvements in  the refinery process which would reduce the
fuel sulfur to an acceptable level; and the other is  flue gas treatment to
remove the SOX.   The refinery technology is available to achieve sulfur
reduction (Ref. 4-121) but at an increase in fuel costs.  Various flue gas
treatment methods are now in  development for steam boilers which
will substantially increase the investment and operating costs of the
plant.  Application of these techniques to stationary gas turbines would
be even more  costly because of the  higher gas turbine flow rates and
exhaust temperatures  relative to steam power plants.  A brief review
of the flue gas treatment method is  given below since in principle it is
applicable to stationary  gas turbine  exhaust.
              In principle there are two  different  ways of SOX removal.
In one approach,  SC>2 is recovered  to form a useful product (such as
sulfuric acid),  and in another,  a solid waste is  formed.  Both, the
recovery  and throwaway methods  can  be carried out in either wet or
dry systems (Ref. 4-122).
              There are more than 50 individual processes known for
SOX removal but only eight have found some degree of acceptability by
utility companies.  Most of these  processes are wet processes consist-
ing of wet lime, alkali or magnesium  oxide scrubbing.  One process
consists of acidified water scrubbing and a dry  catalytic oxidation
system  (Ref. 4-122).  Most of the processes produce solid waste which
poses a disposal problem.  An exception  to this is the magnesia process
in which the products, magnesium sulfite and sulfate, are used for
regeneration of the magnesium and  formation of sulfuric acid.  The
acidified water process  produces  gypsum, and the  catalytic oxidation
                                4-152

-------
procedure produces sulfuric acid.   Currently the reliability of the
various processes  is low (maximum 33 percent), mainly because  of
corrosion, scaling, and plugging problems (Ref.  4-122).
               The investment cost of SOX  scrubbing equipment is
$80/kw for a 550 MW steam unit and for  future larger units the cost is
estimated at $50/kw which adds about 25 percent to the investment cost
of a steam plant.   The operating costs are estimated at 0. 7 mills/kwh
including sludge disposal with additional  three percent of energy con-
sumption (Ref. 4-123).  As noted previously,  the cost figures for gas
turbine installations would be higher than for  steam power plants.
               Summarizing,  it appears that for liquid fuels it would be
more economical to remove any excess sulfur during the refining
process.   If this is not possible (for coal, for instance) the flue gas
treatment would have to be employed.  However, for steam plants this
process will not be developed until the 1975 to 1977 time period
(Ref. 4-120) and more time would be required for gas turbines.   In any
case, SOX control  causes an increase in the investment and/or operat-
ing cost of the plant.
4.3.6          Smoke, Particulates, and  Odor Control
               As discussed in Section 3. 3. 3. 3 smoke (which consists
of small carbon particles),  particulates (consisting of carbonaceous
materials  as well as burnt fuel particles),  and odor formation depend
largely on the  combustion efficiency.  An efficient combustor, which
is characterized by low  CO and low HC,  will also have inherently
low smoke and particulates, and in most cases, low odor (also a
function of fuel aromatics).  Thus,  the development of low emission
combustors,  discussed in Section 4. 3. 3,  will not only maintain low
levels of NO  and CO but will reduce the  other emission species as
            X
well, with the exception of the fuel-dependent SO .
                                4-153

-------
               The interim,  near-term combustor modifications,  as
discussed in Section 4. 3. 3. 2, were,  in general, successful in reducing
smoke and particulates to  an acceptable level (Ref. 4-124).  Conse-
quently,  the use of manganese in the form of methylcyclopentadienyl
manganese, which was found very successful in smoke reduction
(Refs. 4-124 and 4-125), is usually not required.
               There are no  reports  on the effect of the interim modi-
fications on odor which,  in itself, is not a problem in stationary gas
turbines,  but judging from Ref. 4-126, any reduction in HC should have
a beneficial effect on the odor.
                                4-154

-------
                           REFERENCES
4-1.      R. E.  Bosecker and D. F. Webster,  "Precombustion
          Chamber Diesel Engine Emissions — A Progress Report, "
          SAE paper  710672 (August 1971).

4-2.      W. F.  Marshall and R. D. Fleming,  "Diesel Emissions as
          Related to Engine Variables and Fuel Characteristics, "
          SAE paper  710836 (October 1971).

4-3.      "Characterization and Control of Emissions from Heavy
          Duty Diesel and Gasoline Fueled Engines, "  Bureau of
          Mines,  Bartlesville,  Oklahoma (December 1972).

4-4.      R. C.  Bascom,  L. C.  Broering, and D. E. Wulfhorst,
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          paper No. 710484.

4-5.      R. P.  Wilson,  Jr.,  E. B. Muir, and F. A. Pellicciotti,
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          SAE paper  740123, (25 February - 1 March 1974).

4-6.      F. S. Schaub and K. V. Beightol, "NOX Emission Reduc-
          tion Methods for Large  Bore Diesel and Natural Gas
          Engines, "ASME paper  71-WA/DEP-2 (28 November-
          2 December 1971).

4-7.      R. Pischinger and W. Cartellieri, "Combustion System
          Parameters and Their Effect Upon Diesel Engine Exhaust
          Emissions," SAE paper 720756 (11-14 September 1972.

4-8.      S. M.  Shahed,  W. S. Chiu, and V. S. Yumlu, "A Preliminary
          Model for the Formation of Nitric Oxide  in Direct Injection
          Diesel Engines and  the Application in Parametric Studies, "
          SAE paper  730083 (January 1973).

4-9.      E. Valdmanis  and D. E. Wulfhorst,  "The Effects of
          Emulsified Fuels and Water Induction on Diesel Combus-
          tion, " SAE paper 700736 (14-17 September 1970).

4-10.     I. M. Khan, G.  Greeves, and C. H. T. Wang, "Factors
          Affecting Smoke and Gaseous Emissions  from Direct
          Injection Engines and a  Method of Calculation, "  SAE
          paper 730169 (January 1973).
                               4-155

-------
4-11.     R. J. Hames, D. F. Merrion, and H. S.  Ford,  "Some Effects
          on Fuel Injection System Parameters on Diesel Exhaust
          Emissions, " SAE paper 710671 (16-19 August 1971).

4-12.     G. McConnell and  H. E. Howells, "Diesel Fuel Properties
          and Exhaust Gas — Distant Relations, " SAE paper 670091
          (9-13 January 1967).

4-13.     G. Blair Martin and E. E. Berkau,  "An Investigation of the
          Conversion of Various Fuel Nitrogen  Compounds to Nitrogen
          Oxides in Oil Combustion, " AIChE National Meeting,
          30 August 1971.

4-14.     Personal Communication with Cummins Engine Company,
          Columbus,  Indiana.

4-15.     D.W. Golothan,  "Diesel Engine Exhaust Smoke, The
          Influence of Fuel Properties and the Effects of Using
          Barium — Containing Fuel Additive, "  SAE paper 670092
          (January 1967).

4-16.     N. A. Henein and J. A.  Bolt, "The Effect of Some Fuel and
          Engine Factors on Diesel Smoke, " SAE paper 690557
          (11-14 August 1969).

4-17.     W.F. Marshall and R.D. Fleming,  "Diesel Emissions
          Reinventoried, " Bureau of Mines Report PB-201896
          (July 1971).

4-18.     T. Saito and M.  Nabetani,  "Surveying Tests of Diesel
          Smoke  Suppression with Fuel Additives, " SAE paper 730170
          (January 1973).

4-19.     W.F. Marshall and R.W. Hum,  "Factors Affecting Diesel
          Emissions, " SAE paper 680528 (12-15 August 1968).

4-20.     J. G. Brandes,  "Diesel Fuel Specification and Smoke
          Suppressant Additive Evaluations, " SAE paper 700522
          (18-22  May 1970).

4-21.     I. M. Khan,  C.H. T. Wang and B. E.  Langridge,  "Effects
          of Air Swirl on  Smoke  and Gaseous  Emissions from Direct
          Injection Diesel Engines, "  SAE paper No. 720102 (10-14 Jan-
          uary 1972).
                               4-156

-------
4-22.     H. S. Ford, D. F. Merrion,  and R. J.  Hames,  "Reducing
          Hydrocarbons and Odor in Diesel Exhaust by Fuel Injector
          Design," SAE paper 700734 (14-17 September 1970).

4-23.     R.F.  Parker and J.W.  Walker, "Exhaust Emission Control
          in Medium Swirl Rate Direct Injection Diesel Engines, " SAE
          paper 720755 (11-14 September 1972).

4-24.     C.J. Walder, "Reduction of Emissions from Diesel Engines, "
          SAE paper 730214 (January 1973).

4-25.     R. J. Springer and C. T. Hare, "Four Years of Diesel Odor
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4-26.     Presentation to the  California Air Resources Board Staff,
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4-27.     P.S. Myers, D.A.  Uyehara and H. K.  Newhall,  "The ABCs
          of Engine  Exhaust Emissions," SAE paper 710481 (1971).

4-28.     M. W. Jackson, "Effect of Some Engine Variables and
          Control Systems on Composition and Reactivity of Exhaust
          Hydrocarbons, "  SAE Transactions, Vol. 75 (1967),  also
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4-29.     J. D. Caplan, "Smog Chemistry Points the Way to Rational
          Vehicle Emission Control, " SAE  Transactions,  74 (1966).

4-30.     M. W. Jackson, W. M.  Wiese, and J. T. Wentworth, "The
          Influence of Air-Fuel Ratio, Spark Timing,  and Combustion
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          paper No.  486A, SAE National Automobile Week,  Detroit
          (March  1962); SAE  Technical Progress Series, Vol. 6,
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4-31.     C. R. McGowin,  F. S. Schaub, and R. L. Hubbard, "Emission
          Control of a Stationary  Two-stroke  Spark-gas Engine by
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4-32.     P.S. Myers,  "Automobile Emissions -- A Study in Environ-
          mental Benefits versus  Technological Cost, " SAE paper
          No. 700182,  SAE Progress in Technology, Vehicle Emissions,
          Part III, Vol. 14 (1971).
                               4-157

-------
 4-33.     S. R.  Krause, "Effect of Engine Intake-air Humidity,
          Temperature, and Pressure on Exhaust Emissions, "
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 4-34.     G. J.  Nebel,  and M. W. Jackson, "Some Factors Affecting
          the Concentration of Oxides of Nitrogen in Exhaust Gases
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 4-35.     T. A. Huls,  P. S. Myers,  and O. A. Uyehara,  "Spark Igni-
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 4-36.     R. D. Kopa, "Control of Automotive Exhaust Emission by
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 4-37.     J. A.  Robison and W. M. Brehob, "The Influence of Improved
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 4-38.     C. F. Taylor, The Internal Combustion Engine in Theory
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4-39.     W.A. Daniel and J. T. Wentworth, "Exhaust Gas Hydro-
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4-40.     J. T.  Wentworth,  "Effect of Combustion Chamber Surface
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4-41.     W.A. Daniel,  "Engine Variables Effects on Exhaust Hydro-
          carbon Composition, "  SAE paper No. 670124.
                               4-158

-------
4-42.     R. C.  Lee,  "Effect of Compression Ratio, Mixture Strength,
          Spark Timing, and Coolant Temperature Upon Exhaust
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4-43.     R. W.  Aiman, "Engine Speed and Load Effects on Charge
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4-44.     D. F.  Hagen, and G. W. Holiday,  "The Effect of Engine
          Operating and Design Variables on Exhaust  Emissions, "
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4-45.     W. W. Haskell,  and C. E.  Legate,  "Exhaust Hydrocarbon
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4-46.     R. M.  Siewert,  "How Individual Valve  Timing Events Affect
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4-47.     J. D.  Benson and R. F. Stebar,  "Effects of Charge Dilution
          on Nitric Oxide Emission from a Single-Cylinder Engine, "
          SAE paper  710008 (January  1971).

4-48.     M. A.  Freeman and R. C.  Nicholson,  "Valve Timing for
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4-49.     G. B.  Meacham Kirby, "Variable Cam Timing as an Emission
          Control Tool, "  SAE paper No.  700673, SAE Meeting, Los
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4-50.     C. E.  Scheffler,  "Combustion Chamber Surface Area, A Key
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4-51.     L. Eltinge, F. J. Marsee, and A. J. Warren, "The Potenti-
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          tions, " SAE paper No, 680123, SAE Congress, Detroit,
          Michigan (January 1968).
                               4-159

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4-52.     J. A. Bolt, S. P. Bergin, and F. J. Vesper,  "The Influence
          of the Exhaust Backpressure of a. Piston Engine on Air Con-
          sumption, Performance, and Emission, " SAE paper
          No. 730195, SAE Congress, Detroit,  Michigan (January 1973).

4-53.     T. A. Huls and H. A. Nickol,  "Influence of Engine  Variables
          on Exhaust Oxides of Nitrogen Concentrations from a Multi-
          cylinder Engine, "  SAE paper No. 670482, SAE Mid-year
          Meeting, Chicago (May  1967).

4-54.     J. C. Gagliardi,  "The Effect of Fuel Anti-knock Compounds
          and Deposits on Exhaust Emissions, " SAE paper No.  670128,
          SAE Automotive Engineering Congress, Detroit (January 1967).

4-55.     H. E. Leikkanen and E. W.  Beckman,  "The Effect  of Leaded
          and Unleaded Gasolines on Exhaust Emissions as Influenced
          by Combustion Chamber Deposits, " SAE paper No. 710843,
          SAE Meeting,  St. Louis, Missouri (October 1971).

4-56.     R. P. Doelling, et al, "Additives Can Control Combustion
          Chamber Deposit Induced Hydrocarbon Emissions, " SAE
          paper No. 720500, SAE Meeting, Detroit, Michigan
          (May 1972).

4-57.     H. R. Ricardo and J. G. G. Hempson, "The High-speed
          Internal  Combustion Engine, " Fifth Edition,  Blackie & Son
          Ltd.,  Glasgow (1972).

4-58.     Y. Sakai, H. Miyazaki, and K. Mukai,  "The Effect of Com-
          bustion Chamber Shape  on Nitrogen Oxides, " SAE paper
          No. 730154, SAE Congress, Detroit,  Michigan (January 1973).

4-59.     A. E. Felt and S. R. Krause,  "Effects of Compression Ratio
          Changes  on Exhaust Emissions, " SAE paper  No. 710831,
          SAE Meeting,  St. Louis, Missouri (October 1971).

4-60.     P. H. Schweitzer,  "Control of Exhaust Pollution Through a
          Mixture-Optimizer, " SAE paper No. 720254, SAE Congress,
          Detroit,  Michigan (January 1972).

4-61.     R. Lindsay, A.  Thomas,  J. A.  Woodworth,  and E. G.
          Zeshmann, "Influence of Homogeneous Charge  on  the Exhaust
          Emissions of Hydrocarbons,  Carbon Monoxide,  and Nitric
          Oxide from a Multicylinder Engine, " SAE paper No. 710588,
          SAE Meeting,  Montreal, Canada (June 1971).
                               4-160

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4-62.     D. A.  Trayser and F. A.  Creswick, "Effect of Induction
          System Design on Automotive Engine Emissions, " paper
          presented at the ASME Annual Meeting, November 1969,
          Los Angeles, California.

4-63.     G. F.  Leydorf,  R. G. Minty,  and M. Fingeroot,  "Design
          Refinement of Induction and Exhaust Systems Using Steady-
          state Flow Bench Techniques,"  SAE paper No. 720214, SAE
          Congress, Detroit,  Michigan (January 1972).

4-64.     E. Bartholomew, "Potentialities of Emission Reduction by
          Design of Induction System, "  SAE paper No. 660109,
          Vehicle Emissions,  Part II,  Progress in Technology, 12,
          (192), Society of Automotive Engineers,  Inc., New York.

4-65.     E. M. Mitchell,  M.  Alperstein,  et al,  "A Stratified Charge
          Multifuel Military Engine — A Progress Report, " SAE paper
          No. 720051  (January 1972).

4-66.     A. Simko, M. A. Choma, and L. L. Repko,  "Exhaust
          Emission Control by the Ford Programmed Combustion
          Process - PROCO, " SAE paper No. 720052 (January 1972).

4-67.     J. L. Bascunana, "Divided Combustion Chamber Gasoline
          Engines — A Review for Emissions and Efficiency, " EPA
          paper for presentation at the 66th Annual Meeting of the
          APCA, Chicago, Illinois (June 1973).

4-68.     R. D.  Kopa and H.  Kimura, "Exhaust Gas Recirculation as
          a Method of Nitrogen Oxides Control in an Internal Com-
          bustion Engine, " paper presented at APCA  53rd Annual
          Meeting, Cincinnati, Ohio  (May 1960).

4-69.     M. G. Hinton,  Jr.,  T.  lura,  J.  Meltzer,  and J. H.  Somers,
          "Gasoline Lead  Additive and Cost Effects of Potential 1975-
          1976 Emission Control Systems, " SAE paper No.  730014,
          SAE Congress,   Detroit,  Michigan (January 1973).

4-70.     "An Assessment of the Effects of Lead Additives in Gasoline
          on Emission Control Systems Which Might be Used to Meet
          the  1975-76 Motor Vehicle Emission Standards, The Aero-
          space Corporation,  Report No.  TOR-0172(2787)-2,
          15 November 1971.

4-71.     J. C. Chipman,   et al, "Field  Test of an Exhaust Gas Recir-
          culation System  for the Control of Automotive Oxides of
          Nitrogen," SAE  paper  No.  720511.
                               4-161

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4-72.     W.  Glass,  et al,  "Evaluation of Exhaust Recirculation for
          Control of Nitrogen Oxides Emissions, " SAE paper
          No.  700146, SAE Congress,  Detroit, Michigan (January 1970).

4-73.     G. S. Musser, et al, "Effectiveness of Exhaust Gas Recircula-
          tion with Extended Use, " SAE paper No.  710013,  SAE
          Congress,  Detroit, Michigan (January 1971).

4-74.     A. L. Thompson, "Buick's 1972 Exhaust Gas Recirculation
          System,"  SAE paper No. 720519,  Detroit, Michigan
          (May 1972).

4-75.     J. G. Hansel,  "Low NOX Emissions from Automotive Engine
          Combustion," SAE paper No. 720509 (May 1972).

4-76.     E. N. Cantwell,  et al,  "A System Approach to  Vehicle
          Emission Control, "  SAE paper No.  720510,  SAE Meeting,
          Detroit,  Michigan (May  1972).

4-77.     H. K. Newhall, "Control of Nitrogen Oxides  by Exhaust
          Recirculation — A Preliminary Theoretical Study, "  SAE
          paper No.  670495, SAE  Transactions, 1820-1836 (1967)
          (Discussion by R. D. Kopa).

4-78.     R. D. Kopa, R. G. Jewell, and R. V. Spangler, "Effect of
          Exhaust Gas Recirculation on Automotive Ring Wear, " SAE
          paper No.  S-321, SAE Southern California Section
          (March 1962).

4-79.     R. D. Kopa, "Possible Method of Controlling the Oxides of
          Nitrogen Content of Auto Exhaust, " First Technical APCA
          Meeting, (UCLA — Dept. of Engineering,  Report No. 56-26,
          Automobile Exhaust) Los Angeles (March 1957).

4-80.     J. E. Nicholls, I. A.  El-Messiri, and H. K. Newhall, "Inlet
          Manifold Water Injection for Control of Nitrogen Oxides —
          Theory and Experiment, " SAE paper No. 690018, SAE
          Congress,  Detroit, Michigan (January 1969).

4-81.     S. S. Lestz, W. E. Meyer, and C. M. Colony,  "Emissions
          from a Direct Cylinder Water Injected Spark Ignition Engine, "
          SAE paper No. 720113,  SAE Congress, Detroit, Michigan
          (January 1972).

4-82.     C. R. McGowin, "Stationary  Internal Combustion Engines in
          the  United  States, " Contract No. EHSD 71-45; prepared for
          EPA, Washington, D. C.  (April 1973).
                               4-162

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4-83.     E.E.  Wigg,  R. J. Campion, and Wm. L. Petersen,  "The
          Effect of Fuel Hydrocarbon Composition on Exhaust Hydro-
          carbon and Oxygenate Emissions, " SAE paper No.  720251,
          SAE Congress, Detroit,  Michigan (January 1972).

4-84.     B.H.  Eccleston,  B. F. Noble, and R. W. Hum, "Influence
          of Volatile Fuel  Components  on Vehicle Emissions, "
          RI 7291,  U.S. Bureau of Mines (February  1970).

4-85.     G. P.  Gross, "The Effect of Fuel and Vehicle Variables on
          Polynuclear Aromatic Hydrocarbon and Phenol Emissions, "
          SAE paper No. 720210,  SAE  Congress,  Detroit,  Michigan
          (January 1972).

4-86.     R. C.  Carr,  E. S. Starkman,  and R. F.  Sawyer,  "The
          Influence of Fuel Composition on Emissions of Carbon
          Monoxide and Oxides  of Nitrogen, " SAE paper No.  700470,
          SAE Meeting, Detroit, Michigan (May 1970).

4-87.     E. S. Starkman,  et al, "Alternative Fuels for Control of
          Engine Emission," APCA Journal,  Vol. 20 (2) 87-92 (1970).

4-88.     J. A. Harrington and R. C.  Shishu,  "A Single Cylinder
          Engine Study of the Effects of Fuel Type,  Fuel Stoichiometry,
          and Hydrogen-to-Carbon Ratio on CO, NO,  and HC Exhaust
          Emissions," SAE paper No.  730476,  SAE Meeting,  Detroit,
          Michigan (May 1973).

4-89.     "Current Status  of Advanced  Alternative Automotive Power
          Systems and Fuels,"  Vol. Ill — Alternative  Nonpetroleum-
          based Automotive Fuels, Aerospace Report No. ATR-74
          (7325)-2,  Vol. Ill, The  Aerospace Corporation, El Segundo,
          California (March 1974).

4-90.     R. E.  Taylor and R. M.  Campau,  "The IIEC — A Cooperative
          Research Program for Automotive Emission Control, "
          paper No. 17-63 presented at the 34th Midyear Meeting of
          the  American Petroleum Institute,  Chicago, Illinois (May 1969).

4-91.     A. A.  Zimmerman,  L. E.  Furlong,  and H. F. Shannon,
          "Improved Fuel  Distribution  — A New Role for Gasoline
          Additives, "  SAE paper No. 720082,  SAE Congress,  Detroit,
          Michigan (January 1972).
                               4-163

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4-92.     "Report by the Committee on Motor Vehicle Emissions,"
          The Environmental Protection Agency and the National
          Academy of Sciences,  NAS, Washington,  D.C.
          (February  1973).

4-93.     J.N.  Reddy, "Closed-loop Emissions  Control for Automotive
          Engines, " Bendix Technical Journal (Spring 1973).

4-94.     "Automotive Spark Ignition Engine Emission Control Systems
          to Meet the Requirements of the 1970 Clean Air Amendments, "
          Report of the Emission Control System Panel to the Com-
          mittee on Motor Vehicle Emissions, National Academy of
          Sciences (May  1973).

4-95.     P. A. Bennett,  C.K. Murphy, M.W. Jackson, and R. A.
          Randall, "Reduction of Air Pollution by Control of Emission
          from Automotive Crankcases," SAE Paper No.  142A,  SAE
          Annual Meeting,  January I960; SAE Technical Progress
          Series,  Vol. 6, Vehicle  Emissions, 1964.

4-96.     G. D.  Ebersole and G. E. Holman,  "Lubricant Closed PCV
          System  Relationships Influence Exhaust Emissions, " SAE
          paper No. 680113, Automotive Engineering Congress,
          Detroit  (January 1968).

4-97.     J. T.  Wentworth, "Carburetor Evaporation Losses, " SAE
          paper No. 12B,  SAE Annual Meeting (January 1958); SAE
          Technical Progress Series, 6,  Vehicle Emissions,  p.  146
          (1964).

4-98.     G. D.  Ebersole and L. L. McReynolds, "An Evaluation of
          Automobile  Total Hydrocarbon Emissions, " SAE Transac-
          tions, Vol.  75 (1967); SAE Progress in Technology,  Vol.~12,
          Vehicle  Emissions, Part II,  p. 413 (1966).

4-99.     P. J.  Clarke, J. E. Gerrard, C. W. Skarstron, J. Vardi, and
          D. T.  Wade,  "An Adsorption-Regeneration Approach to the
          Problem of  Evaporative  Control,"  SAE paper No.  670127,
          SAE Automotive Engineering Congress, Detroit (January 1967).

4-100.    The Aerospace Corporation,  "Final Report.  Status  of
          Industry Progress Toward Achievement of the 1975  Federal
          Emission Standards for Light-Duty Vehicles, " Aerospace
          Report ATR-73 (7322)-!  (July 1972).
                              4-164

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4-101.    Environmental Protection Agency, "Automobile Emission
          Control - The State of the Art as of December  1972, " EPA
          Division of Emission Control Technology (February 1973).

4-102.    "Fifth Summary Report AAPS Contractors' Coordination
          Meeting," U.S. Environmental Protection Agency (June 1973).

4-103.    W.U. Roessler, et al, "Investigation of Surface Combustion
          Concepts for NOX  Control in Utility Boilers and Stationary
          Gas Turbines, " The Aerospace Corporation Report
          No. ATR-73 (7286)-2 (August 1973).

4-104.    E. E.  Berkau and D. J. Lachapelle, "Status of EPA1 s Com-
          bustion  Program for Control of Nitrogen Oxide Emissions
          from Stationary Sources, " EPA Report (19 September 1972).

4-105.    W. Bartok and A.  Shiepp, "Control of U. S.  NOX Emissions
          from Stationary Sources, " Chemical Engineering Progress
          67 (2) (February 1971).

4-106.    W. R. Wade,  et al,  "Low Emission Combustion for the
          Regenerative Gas Turbine," ASME,  73-GT-ll  (April 1973).

4-107.    D. G.  White,  et al,  "Low Emission Variable Area Combustor
          for Vehicular Gas  Turbines," ASME,  73-GT-19 (April  1973).

4-108.    J. N. Barney and F. J. Verkamp, "Aircraft Gas Turbine
          Engine High Altitude Cruise Emissions, "  Detroit Diesel
          Allison  Report (1 August 1973).

4-109.    P.P.  Singh, et  al, "Formation and Control of Oxides of
          Nitrogen Emissions from Gas Turbine Combustion Systems,"
          Journal of Engineering and Power (October 1972).

4-110.    T.F.  Nagey, P.M. Kolents,  and M. E. Nayler, "The Low
          Emission  Gas Turbine Car," ASME,  73-GT-49 (April 1973).

4-111.    D.L.  Troth,  et al,  "Investigation of Aircraft Gas Turbine
          Combustor having  Low Mass Emissions," Report 73-6,  U.S.
          Army Air Mobility Reserve  and Development Laboratory
          (April 1973).

4-112.    N. A.  Azelborn,  et al, "Low Emission Combustion for the
          Regenerative Gas Turbine," ASME 73-GT-49 (April 1973).
                               4-165

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4-113.    R. Kollrack and L.D. Aceto, "The Effects of Liquid Water
          Addition in Gas Turbine Combustors, " Journal of the Air
          Pollution Association,  23  (2) (February 1973).

4-114.    R. J. Johnson, et al, "Gas Turbine Environmental Factors —
          1973," GE Report (1972).

4-115.    H.W. Carlson, "The STAG Cycle, " GE Report USDA-4-72
          (September 1972).

4-116.    N. R. Dibelius and E. W.  Zeltmann, "Gas  Turbine  Environ-
          mental Impact Using Natural Gas and Distillate Fuels,"
          73-GTD-6, General Electric (February 1973).

4-117.    M.B. Hilt and R.H. Johnson,  "Nitric Oxide Abatement in
          Heavy Duty Gas Turbine Combustion by Means of Aero-
          dynamics and Water Injection," ASME,  72-GT-53
          (March  1972).

4-118.    "Response  to Preliminary (draft) Proposed Standards for
          Control of Air Pollution from Stationary Gas Turbines, "
          General Motors (March 1973).

4-119.    M. J. Ambrose and E.S.  Obidinski, "Recent Field Tests for
          Control of Exhaust Emissions from a 35 MW Gas Turbine,"
          ASME, 72-JPG-GT-2 (September  1972).

4-120.    V. De Biasi,  "Double Standard on Fuel  Oils Would Favor
          Steam Over Gas Turbine Plants, " Gas  Turbine World
          (September 1973).

4-121.    C.A. Robinson,  "NASA Plans Award on Engine Emissions,"
          Aviation Week and Space Technology (8  April 1974).

4-122.    Battelle- Columbus,  "SO2 Control:  Low-Sulfur Coal Still
          the Best Way, " Power Engineering (8 April 1974).

4-123.    F.S. Olds, "SO   and NO , Power Engineering (August 1973).
                        X        X
4-124.    S.M. De Corso, et al, "Smokeless Combustion in Oil-
          Burning Gas Turbines," ASME,  67-PWR-5  (September 1967).
                              4-166

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4-125.    L. Plonsker, et al, "Reduction of Gas Turbine Smoke and
          Particulate Emissions by a Manganese Fuel Additive, "
          Ethyl Corporation Report, presented at Central Section/The
          Combustion Institute (March 1974).

4-126.    H.F.  Butze and D. A.  Kendall,  "Odor Intensity and Charac-
          terization Studies of Exhaust from a Turbojet Engine Com-
          bustor," NASA Technical Memorandum, NASA-TMX-71429
          (November 1973).
                               4-167

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                             SECTION 5

             EMISSION CONTROL SYSTEM ASSESSMENT
              This section of the report presents an assessment of
the emission control techniques/systems identified in Section 4 relative
to their applicability to new and existing  stationary engine installations.
In addition, preliminary information on the economics of emission con-
trol in stationary engines is included in this section.
              Subsection 5.1 is concerned with diesel  engines, and
Subsections 5.2 and 5.3 are devoted to spark ignition engines and gas
turbines,  respectively.
5. 1           DIESEL ENGINES
              The following discussion is concerned with an evaluation
of techniques/devices  potentially applicable to diesel engines.  Since
NO is the predominant exhaust pollutant from these engines, the dis-
cussion will be concerned primarily with that particular species.   The
most promising NO  abatement techniques are identified and system
                  Jk
cost data are presented whenever possible.
                                 5-1

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5.1.1         Emission Control Techniques/Devices
5. 1. 1. 1       Engine Derating
              In general,  reducing the engine load at rated speed
increases the specific mass emissions of NO  and HC of naturally
aspirated,  open-chamber and turbocharged divided-chamber diesel
engines.   Conversely, slight reductions in NO  have been observed in
the  case  of turbocharged,  open-chamber diesels.  In some engines,
simultaneous reduction of  engine speed and load might result in sub-
stantially lower NO  ,  CO, and  smoke emission levels.
                  X*
              Although insufficient information is currently available
to permit a quantitative evaluation of this approach,  it appears that
limited engine derating might be practical for some  engines, particu-
larly in conjunction  with other emission control techniques.  Basically,
engine derating could be implemented in both new and existing engines,
but  a change  in gears would be generally required to compensate for the
lower engine speed.   However,  engine derating raises the  engine
investment cost per  horsepower and has some impact on the specific
fuel consumption, maintenance  requirements, and life of the engine.
Obviously, all  these parameters would have to be considered in a mean-
ingful cost effectiveness evaluation of this technique, relative to other
emission control approaches.
5. 1. 1.2       Intake Manifold Temperature
               Reduction of the  air intake temperature has a beneficial
effect on the  NO  emissions from diesel engines.  Although the  effect
                X.
is rather small (20  percent reduction per  100°F temperature drop),  it
is accompanied by a slight improvement in specific  fuel consumption
and essentially no change in the HC, CO and  smoke  emissions.
               Intake charge cooling,  which would be particularly effec-
tive at high ambient temperature conditions requires the installation of
a heat exchanger upstream of the engine.  In turbocharged diesels,
                                 5-2

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intercooling is a proven technique to increase the power output
capability of the engine and to lower its  specific fuel consumption.
Additional improvements in engine performance and NO might be
accomplished by increasing the size and/or effectiveness of the  cooler
and by employing air or city water as coolants.  This approach merits
further investigation,  particularly for stationary engines where  the
volume constraints of the air cooler would be  less severe than in mo-
bile installations and where city water is available.  It is applicable to
both new and existing engines and might be particularly attractive in
conjunction with other emission control  techniques,  such as injection
timing retard or EGR.
5. 1. 1.3       Fuel Injection Timing Retard
               The effect of fuel injection timing retard on diesel emis-
sions has been determined experimentally by many investigators and the
available test data are in reasonable agreement for all diesel engine
classes. Particularly the first few degrees retard are  very  effective
in reducing NO  at the expense  of moderate increases in CO,  smoke,
and specific fuel consumption and a small loss in output power.  HC
either increases,  decreases, or remains constant with  increasing timing
retard, but the variations are generally less  than ±70 percent.  For
most engines,  injection  retard is  probably limited to about six degrees or
less because the specific fuel consumption deteriorates rapidly beyond
that point.  Typically, six degrees injection retard lowers  NO by about
40 percent while CO and exhaust smoke  increase by about 50 percent or
less.   Since CO and  smoke emissions are generally low in well main-
tained diesel engines, this increase is probably tolerable in most cases.
               If stringent emission standards would be promulgated
for  stationary diesel  engines, it appears that most  manufacturers might
incorporate  some  degree of injection retard.   This technique  requires
no hardware  changes and can be easily incorporated into new  and exist-
ing  engines.  The  associated loss in specific  fuel consumption might be
                                 5-3

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alleviated to some degree by simultaneously optimizing the timing
retard and injection period schedules.   Potential problem areas related
to injection timing retard include exhaust valve and turbocharger dura-
bility which might be adversely affected by the higher exhaust gas tem-
peratures obtained with retarded timing.
5. 1. 1.4       Combustion Chamber Modifications
               Optimization of the combustion chamber geometry in
terms of bowl diameter to bore ratio,  piston to head  clearance,  com-
pression ratio,  and air  swirl can result in lower NO   emission and
specific fuel consumption of diesel engines.  However, the  combustion
cycle is presumably optimized for specific fuel consumption, and any
reduction in NO  would  be accompanied by an increase in fuel consump-
               X.
tion relative to the optimum condition.  Although the  observed effects
are rather small, optimization of these parameters is desirable and
could be fairly easily implemented in new engine designs at no increase
in hardware cost.  Conversely,  such modifications are not  considered
to be cost effective for retrofit applications.
5. 1. 1.5       Injection System Modifications
               A number of fuel injection system parameters, includ-
ing injection rate, orifice size,  spray  angle, and sac volume are
important design parameters, impacting the emission characteristics
of diesel engines.  On one engine, optimization of the fuel injection
rate has resulted in a 20-percent reduction in NO  with no attendent loss
in fuel economy.  However, incorporation of this approach  might
require a new fuel injection system which would be capable of with-
standing the higher stress levels associated with shorter injection
periods.  Optimization of the fuel injector spray angle and  orifice size
is expected to cause moderate reductions in NO  and HC.
               Basically, these techniques are considered to be  appli-
cable to new engines and retrofit installations alike.  However,  a
                                5-4

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meaningful evaluation of the cost effectiveness of these modifications,
as applied to the various engine categories, is not possible at this time
because of a lack of applicable test data.
5. 1. 1. 6       Fuel Effects
              Variations in the fuel composition and cetane number
have some effect on diesel engine emissions.  In medium speed
naturally  aspirated and turbocharged diesels,  NO ,  HC, and CO tend to
decrease  with increasing fuel cetane number.  However, it appears
that the observed reductions (about 10 percent) are not sufficiently
large to warrant the manufacture of higher cetane fuels for use in these
engines.   Conversely, a reduction in cetane number substantially below
current levels could result in markedly higher HC, CO, and NO
emissions.
              Fuel bound nitrogen has been identified to be a significant
contributor  to the total NO  emissions  from oil-  and coal-fired boilers.
                          x
Although similar effects might be obtained in diesel engines when oper-
ating on heavy distillate and residual fuels, there is insufficient data
available  at this  time to permit a meaningful assessment of this
phenomena.
              Smoke suppressant fuel  additives, although  quite effec-
tive in reducing smoke,  are not recommended by most manufacturers
because of potential adverse effects on engine durability and human
health.  Work on odor suppressant additives has  not been successful.
5. 1. 1.7       Fumigation
              Based on a very limited data sample it appears that
small reductions in NO  might be achieved in diesel engines by means
of fumigation (injection of a small amount of fuel into the intake system).
This method could  be incorporated into both new and existing engines.
However,  additional development programs would be required to provide
the data needed  for a meaningful assessment of this particular technique.
                                 5-5

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5. 1. 1. 8       Exhaust Gas Recirculation
               Exhaust gas recirculation has been shown to be a very
effective NO  abatement technique for all types of diesel engines, par-
ticularly in those cases where the EGR  flow is cooled before admission
into the intake manifold.  However, incorporation of EGR generally
causes some increase in CO, smoke, and specific fuel consumption.
accompanied by moderate  reductions in the HC emissions and engine
power output capability.  At rated engine load incorporation of about
ten percent cooled EGR typically results in a 50-percent reduction in
NO  , accompanied by a 100- to 150-percent increase in CO and smoke,
   X
a 1.5-percent  increase in SFC,  and some loss in  engine power output.
EGR is generally less effective at part load conditions because of the
attendant increase of the oxygen concentration in the engine exhaust.
               Most manufacturers feel that the application of EGR
would  create a number of potential problem areas in the engine.   These
include corrosion and deposit buildup in the EGR circuit,  particularly
the EGR cooler, as well as excessive engine wear and fuel oil contami-
nation resulting from the sulfur and metallic compounds contained in
the fuel. Additional difficulties might arise in turbocharged and after-
cooled engines, due to deposit buildup in the compressor and heat
exchanger.  The deposit-related problems might  be alleviated by incor-
porating a filter system which would be serviced periodically.
               In principle,  EGR could be added to both new and exist-
ing engines.  However, considerably more testing would be required
to permit a  meaningful assessment of its long term effects on engine
durability,  performance and emission control effectiveness.
5.1.1.9       Water Injection
               Although the available test data show  considerable  scat-
ter,  there is no doubt that water injection into the intake system (induc-
tion) is an effective method to reduce the NO  emissions from diesel
                                           X.
                                 5-6

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engines.  The principal advantage of water injection relative to many
other techniques is the fact that the improvement in NO  can be accom-
plished with little or  no loss in specific fuel consumption.
              Typically, water induction at a rate about equal to the
fuel injection rate reduces the NO   emissions by about 50 percent.
Based on a very limited data sample it appears that HC increases
slightly with increasing water injection.  Conversely,  smoke  shows a
tendency to decrease with increasing water flow rates, whereas CO
and SFC show very little  change with water flow rate.
              Water induction could be added fairly easily to new and
existing engines, but a number of potential problem areas would have
to be resolved before this technique could be seriously considered for
use in stationary engines. These include corrosion and wear of the
intake system, intake valves and water injection nozzles, as well as
degradation of the lube oil.   Although these problems might be allevi-
ated by using distilled or demineralized water, the related cost increase
would have to  be considered  when comparing this approach to other
potential control techniques.
              Although water injection in the form of  emulsified fuel
has not been successful in one engine,  this technique merits further
consideration  primarily because the previously noted corrosion and
wear problems in the intake  system would be eliminated.  However,
wear and  corrosion in the fuel system are considered  to be potential
problem areas.   Furthermore, only a single set of injectors would be
required which would be beneficial from a cost point of view.
5. 1. 1. 10      Catalysts
              When fresh,  catalytic converters are quite effective in
the control of  HC, CO, and odor from diesel engines.  However, the
durability of these catalysts  has not been demonstrated, particularly in
conjunction with low grade fuels.  More importantly, there are no
known NO  reduction catalysts that would be  effective  in the oxidizing
                                 5-7

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atmosphere typical of diesel exhaust.   Although theoretically possible,
generation of sufficient amounts  of reducing species such as CO and H_
by means of afterburning is not considered to be economically feasible
at this time.
              Since the HC and CO emissions are generally low in
diesel engines, it appears that the use of oxidation catalysts would be
limited to those cases where excessive amounts of HC and CO would  be
generated as a result of the application of NO   abatement systems.
                                           Ji.
5. 1. 1. 11      Thermal Reactors
              Because of the low HC  and CO concentrations in the
exhaust of diesel engines and the relatively low exhaust gas tempera-
tures, thermal reactors would probably be rather  ineffective in terms
of HC and CO abatement.  Furthermore,  thermal reactors have no
effect on NO ,  the principal pollutant  species emitted from diesel
            X
engines.
5. 1. 1. 12      Turbocharging
              In general, incorporation of a turbocharger into naturally
aspirated diesel engines increases the specific mass emissions by as
much as 70  percent while lowering CO, smoke, and  SFC.  However,  by
combining turbocharging  with retarded injection timing and intercooling,
NO  reductions up to 35 percent  have  been demonstrated without occur-
   X,
ring any loss in specific fuel consumption relative to equivalent naturally
aspirated engines.
              Basically, turbocharging/intercooling is applicable to
both new and existing diesel engines and is being seriously considered
by many manufacturers of automotive diesels to meet future emission
control standards. For durability reasons,  the compression ratio of
existing engines might have to be lowered to compensate for the higher
combustion  processes obtained with turbocharging.
                                 5-8

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5.1.2         Emission Control Systems
              To date, only a very limited amount of research work
has been conducted to characterize and optimize the effectiveness of
promising  emission control systems for potential use in both new diesel
engines and retrofit installations.  In general, the emission reductions
achieved with these  control systems  are in reasonable agreement with
projections,  as determined from the available test data of the individual
devices/techniques.
              Best  results, in terms of NO  reduction and attendent
SFC loss,  have been reported for a control system consisting often per-
cent EGR and intake air cooling.  In this  case,  NO  was reduced by
about 40 percent  while the specific fuel consumption improved  by almost
two percent.   Basically, a control system of this type  could be incor-
porated fairly easily into new and existing engines.  However,  it should
be emphasized again that these results are for new control systems
and do not  include any allowance to account for potential performance
degradation of the system.
              According to one diesel engine manufacturer,  even more
favorable NO versus specific fuel consumption tradeoffs might be
realized by means of extensive engine modifications and optimization
of certain engine  components.
5. 1.3         Economic Considerations
5. 1.3. 1       Emission Control Equipment and Maintenance Cost
              In view of the very limited emission control system work
conducted to  date, it is impossible at this time to perform an accurate
assessment of the initial cost and maintenance requirements of the
various potential emission control devices and systems.
              With  respect to EGR,  fairly reliable cost figures  are
available for automotive systems which have achieved  an advanced
state of development.  However, for a number of reasons these data
                                 5-9

-------
are not considered to be applicable for liquid-fueled diesel engines.
First, the EGR flow used in diesels would have to be cooled to achieve
acceptable effectiveness and this would add  substantially to the com-
plexity and cost of the system.  In addition, because of the higher cor-
rosiveness of the  diesel exhaust gases relative to spark ignition
engines, there is  considerable uncertainty at this time regarding the
degree of EGR filtering required to minimize deposit buildup in the
EGR  cooler and wear of certain engine parts.  Also,  incorporation of
EGR  tends to degrade the output power capability of the engine, thus
further raising the engine  cost per  horsepower output.
               One manufacturer has  stated that incorporation of a
turbocharger increases the cost of a medium size diesel engine by
about 10 percent,  or  $2. 50 to $3.00 per horsepower.  In large station-
ary diesels (typical cost about 100 $/hp),  the percentage increase is
somewhat  lower.  Addition of an inter cooler would contribute another
30 to 50  cent per horsepower to the overall  cost of the engine.  How-
ever,  since intercooling improves the specific fuel consumption  of the
engine and its power  output capability the cost of the intercooler would
be recovered within a relatively short time.
               No initial cost and maintenance figures are currently
available for water injection systems. It is estimated that the initial
cost of such  a system would be  comparable  to the cost of the fuel injec-
tion system.   Depending on the  required  purity of the water, additional
cost penalties would be due to the manufacture of the purified or distilled
water and  the cost of construction  of the  purification plant.  Pertinent
water cost data are presented in Section  5.3 for  gas turbines.
               One engine  manufacturer has estimated that incorpora-
tion of a typical emission control system would increase the initial cost
of the engine by five to ten percent, or $1.25 to $3.00 per horsepower.
In addition, the maintenance cost of the engine would increase by about
10 to  15  percent.  These estimates are corroborated by another
                                5-10

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manufacturer  who indicated that the initial cost of emission control
hardware would not greatly affect the sales price of the engine.  How-
ever, any increase in specific fuel consumption as a result of emission
control would  have a significant impact on the operating cost  of the
engine.  This  becomes apparent when considering the operating sched-
ule of engines employed in continuous power installations.  Based  on
the assumption of a fuel cost of six cents per pound,  and  an engine
operating time of 7000 hours per year,  the total fuel cost per horse-
power  is about $170 per year, or $5100 for 30  years, which represents
the average life of large stationary diesel engines.  Hence, in this case,
each percentage point loss in specific fuel consumption is equivalent to
about $50 per  horsepower over the life of the engine.  Of course,  the
importance of fuel cost decreases with decreasing operating time of
the engine.
              Although fuel cost appears to be  the controlling param-
eter for many installations, the effect of the emission control system
on engine life,  maintenance requirements,  and  parts replacement
would have to  be determined before an accurate economic analysis of
diesel  engine emission control would be  possible.
5.1.3.2       Operating Cost
              The estimated cost of NO  control due to changes in
engine specific fuel consumption in terms of dollars  per  ton NO  re-
                                                            X
moved versus NO emission level for turbocharged open-chamber
diesel  engines is presented in Figure 5-1.  Maintenance  cost differ-
entials and water cost are not included in the figure.
              The curves plotted in Figure 5-1 are based on the SFC
vs NO  reduction data shown in Figure 5-2 and  were computed from
the following equation, using a fuel cost of $0.06 per pound,  a base-
line NO  level (NO   ) of 12 g/bhp-hr and a baseline specific fuel
       x          x, o       °    r
consumption (BSFC  ) of 0. 380 Ib/bhp-hr.
                                5-11

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     600
     500
     400
  O  300
     200
  o
  2  100
  I
  cc
  8
  o
    .too
    -200
    -300
®
®
®
®
®
©
0
        INTAKE COOLING (0-100° F)
        TIMING RETARD
        INCREASED INJECTION RATE + RETARD
        EGR (0 — 15*1
        WATER INDUCTION (W/F = 0-1.0)
        CUMMINS PHASE III (13 mode data)
        DIVIDED CHAMBER
        10% EGR + 5° RETARD + COOLING
        10% EGR + COOLING
•V"-
       0     2
                   468
                      NOX, g/bhp-hr
                                      10
                                            12
Figure  5-1.
         Projected NOX abate-
         ment cost - turbo-
         charged,  open - cham-
         ber diesels (fuel cost
         only)
                                                                 +20
                                                                 4-16
                                                                 +12
                                                                Q.
                                                                2
                                                                y  o
                                                                 -12
                                                                 -16
INTAKE COOLING (0-100°F|
TIMING RETARD (0-10°FI
HIGHER CETANE N0.(40-50)
HIGHER INJECTION RATE +
TIMING RETARD
EGR (0-15%)
WATER INDUCTION (W/F = 0-1.0)
CUMMINS PHASE I (13 mode)
CUMMINS PHASE III (13 mode]
DIVIDED CHAMBER VS
OPEN CHAMBER
                                                                      A SWIRL • TIMING RETARD
                                                                      V TURBOCHARGING + AFTERCOOLING
                                                                      O 10% EGR + 5° RETARD + COOLING
                                                                      O 10% EGR + INTAKE COOLING
                                                                       I	I	I	|	I	
                                                            10    20    30    40    50
                                                                    NOX REDUCTION, %
                                                                                                     60
                                                                                                          70
                                                 Figure 5-2.
       Projected average NOX
       vs SFC  correlations

-------
                                   , BSFC
                       =  0.908 x 10  T^	-x$/LBfuel
              Ton NO      "<7VU " ±v  NO
                    X                   X, O
                             BSFC  _ A
The curves presented in Figure 5-2 represent averages for the various
techniques discussed in Section 4.1.
              Based on the above equation, the fuel related NO  abate-
ment cost is proportional to the specific fuel consumption and inversely
proportional to the NO  emission level of the  uncontrolled engine.
Hence,  the NO  abatement cost for other diesel engines (naturally
              Ji
aspirated and  divided chamber) can be easily  determined by ratioing
the cost data plotted in Figure  5-1 by the baseline BSFC and NO levels
                                                             Jk
of these engines.
              As  indicated in this figure,  timing retard is the least
cost effective  NO  abatement technique.  Conversely, the most cost
effective approaches considered to date are the  Phase III modification
under development by Cummins and a  system consisting of ten percent
EGR and intake air cooling.
5.1.4         Promising Emission Control Systems
              Based on the evidence presented  in the previous  subsec-
tions and in Section 4.1,  the following approaches offer the best com-
promise between NO  emission control and specific fuel consumption
                   Jt
degradation.
5.1.4.1       Near Term Controls
              Of all known emission control techniques for diesel
engines, fuel injection timing retard is the only potential NO  abate-
ment method that involves no hardware changes.  However, because of
                               5-13

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substantial losses in specific fuel consumption, timing retard is also
the least cost-effective technique considered to date for use in diesel
engines.  Some improvement in specific fuel consumption might be
achieved by employing optimized timing retard and injection period
schedules.  This combined approach, which might require new fuel
injectors, is considered to be applicable to both new and existing
engines.
               Intake air  cooling, although not extremely effective by
itself, appears to be very attractive for use in conjunction with other
control techniques,  such as timing retard, EGR, or water injection.  In
this case, the  increase in fuel consumption due to timing retard and/or
EGR would be  compensated for, at least in part, by the reduction in
SFC obtained with charge  cooling.   Manifold  cooling of the air by means
of a water-to-air or air-to-air heat exchanger would be particularly
effective for turbocharged diesels because of  the relatively high tem-
perature rise in the compressor.  In some large diesel engines in
which intercooling has been employed for some time,  further NO
                                                               a.
reduction might be achieved by improving  the effectiveness of the heat
exchanger.
               Water induction into  the  intake manifold of the engine or
injection into the cylinder either directly or in the form of emulsified
fuel are other  potential near-term NO  abatement techniques for use in
new and existing engines.  These methods  merit further investigation,
particularly with regard to the effects on engine life and  specific fuel
consumption.   The NO reduction effectiveness of water induction has
been demonstrated in  limited test work, but consistent data on specific
fuel consumption are lacking  as are  data on emulsified fuels.
5.1.4.2        Far Term Controls
               As indicated in Figure 5-1,  a new engine design cur-
rently being developed by one manufacturer represents the most cost-
effective NO  abatement approach projected for diesel engines to date.
                                 5-14

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In this design,  NO   control is accomplished by means of limited
                 Ji
timing retard combined with many engine  component modifications
optimized for low emission operation.  Further NO  reduction, possibly
                                                 X.
at the expense  of a  slight increase in specific fuel consumption and
some of the other pollutants,  might be achieved by adding intake air
cooling, EGR, or water injection.  However, as previously stated, the
long term effects of EGR and water injection would have to be deter-
mined experimentally, before these particular  techniques could be
safely incorporated into  stationary engines.
               Because of the inherently lower  NO  emissions  of pre-
chamber diesel engines,  relative to open chamber configurations, it
appears that those  engine design and operating  parameters having the
greatest impact on  the formation of NO and the specific fuel consump-
                                      X.
tion of prechamber engines should be  optimized.  Again, intake air
cooling, combined  with EGR or water injection, might further  improve
the cost effectiveness of NO  abatement and these techniques merit
additional research and development.
5. 2            SPARK IGNITION ENGINES
5. 2. 1          Emission Control Techniques/Devices
5.2.1.1       Air-Fuel  Ratio
               One  of the most important engine operating parameters
relative to exhaust  emission control is the air-fuel ratio of the com-
bustible mixture.   Relative to automotive  engines, substantially lower
HC, CO, and NO emissions and lower specific fuel consumption can be
                X
achieved in stationary engines by operating these engines with  excess
air.
               Very low  HC and CO emissions have been demonstrated
in stationary gas fueled spark ignition  engines operating with 30 percent
to 50 percent excess air.  However,  the observed NO  emission is very
                                5-15

-------
high in these engines, probably because of high local combustion tem-
peratures resulting during the combustion process in large engines and
the relatively long residence time of the reacting species in the high-
temperature post-flame zone.  In this case, further leaning of the  air-
fuel mixture might result in lower NO  at the  expense of higher HC
                                    3t
and fuel consumption.
               Stationary gasoline engines operate generally in the rich
or  near stoichiometric air-fuel ratio regime.  As a  result, HC and CO
are higher than in gas engines but NO  is  lower.   In order to  minimize
                                    }t
these pollutants, gasoline engines should  be operated with at least
30  percent excess air.   However,  achievement of stable engine opera-
tion under these conditions would require  uniform air-fuel mixture.
5.2. 1.2       Ignition Timing
               Spark timing  retard from the maximum brake torque
setting results in  moderate reductions of  NO  and HC, accompanied by
                                           X
a loss in engine power  and fuel economy.   Furthermore,  at high  engine
loads, overheating and burn-out of exhaust valves may occur  in the
case of retarded timing.  For these reasons,  this approach is not
considered to be a viable emission control technique for  stationary
spark ignition engines.
5. 2. 1.3        Mixture Temperature
               Lowering the  mixture temperature results in a decrease
of the NO  specific mass emission from spark-ignition engines.   In this
         Jt
case, there is  no effect on CO but a slight increase in HC might occur
in gasoline fueled  spark-ignition engines.  In general,  lowering the mix-
ture temperature has a beneficial effect on fuel economy, volumetric
efficiency, and fuel octane requirement of spark ignition engines.
               Reduction of the mixture temperature in gasoline engines
can be accomplished by eliminating inlet manifold heating and in  gas
engines by passing the inlet air through an evaporative cooler. In  this
                                5-16

-------
case, NO  would be even further reduced because of the increased
         X
moisture content of the inlet air.
5.2.1.4       Coolant Temperature
               Engine tests have indicated that increasing the engine
coolant temperature results in considerably lower HC emissions.  How-
ever, NO  might increase slightly,  particularly at lean mixture opera-
         A.
tion.  Thus,  this technique appears to be feasible only in combination
with other  control methods.
               Conversely, lowering the engine coolant temperature
might be feasible to achieve moderate NO  emission control in gas
engines where a slight increase in HC could be tolerated, because of
the low HC emission level typical of these engines.
5.2. 1.5       Engine Speeci
               In general, increasing the engine speed at constant
torque or constant  power is  accompanied by some reduction in HC.
However, the degree of improvement in HC depends on several factors,
including the design of the piston rings,  the piston ring condition, and
the blow-by flow rate.
               Increasing the engine speed  at constant power results in
substantial reduction in NO  .  In this case engine load decreases,
resulting in higher  charge dilution in the engine  cylinder (internal EGR).
Although this  technique might be applicable to stationary engines,
further evaluation of the attendant affects on engine life would be
required before application to stationary engines could be seriously
considered.
5.2.1.6       Valve Timing
               In current spark ignition engines  the valve timing  is set
for maximum engine power output.  This requires high volumetric
efficiency, combined with minimum charge dilution by residual gases.
                                5-17

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Changing the valve timing from the factory setting increases the amount
of residual gases resulting in lower HC and NO  emissions.  These
improvements might be achieved with little or no penalty in specific
fuel consumption.
5.2.1.7       Engine  Load
              In most spark ignition engines NO  decreases with
decreasing engine load, while HC increases as a result of the lower
average surface temperature of the cylinder walls and the attendant
increase of "wall  quenching."  In principle, this technique is applicable
to stationary engines,  but it would result in a proportionate reduction
in the power output of  the engine, resulting in a substantial  increase in
engine cost per unit horsepower  output.
5.2.1.8       Exhaust Backpressure
              In some engines,  increasing the exhaust backpressure
might result in some reduction in NO   , especially in combination with
                                   Ji
a change in valve timing.   However, because of the adverse effects of
increasing exhaust backpressure on engine power and fuel economy,
this technique would probably not be attractive  from a cost/benefit
point of view.
5.2.1.9       Combustion Chamber
              Combustion chamber modifications including changes in
the engine head, the location of the valves and spark plug,  the piston
top shape,  and the  stroke-to-bore  ratio have a  direct influence on the
exhaust emissions from spark ignition engines.  Although the majority
of these modifications could not be incorporated into existing engines,
these parameters  warrant further  consideration for potential applica-
tion to new engine designs.
                                5-18

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5.2.1.10      Fuel System
              Although considerable research and development has
been devoted to the improvement of carburetors, fuel maldistribution
to the individual engine cylinders  remains an unsolved problem area.
The electronic fuel injection systems utilized  in some automobiles are
claimed to improve the fuel distribution and, to some extent, the
exhaust emissions and fuel economy of these engines.  However,
statistical data on durability, maintenance requirements, and emission
control deterioration of these systems  are presently lacking, and no
information is available  relative to the potential effectiveness of these
systems in stationary engines.
5.2. 1. 11      Homogenation of Air-Fuel Mixture
              Experimental evidence clearly  shows a considerable
potential for emission reduction by operating spark-ignition engines
with homogenized air-fuel mixtures.  In the absence of size  limitations,
this emission control technique appears feasible for stationary  engine
applications and additional development efforts should be considered in
that area.
5.2. 1. 12      Stratification of Air-Fuel Mixture
              Stratification of the air-fuel mixture, particularly the
open-chamber concept, offers the possibility of substantial emission
reductions as well as gains in fuel economy.  However,  application of
this technique would require redesign of the engine and incorporation
of a high pressure fuel injection system.  To date this engine concept
has not been evaluated for potential use in stationary engines.
              Recent developments in prechamber type stratified
charge engines indicate potential advantages of this engine concept in
terms  of improved exhaust emissions and specified fuel economy rela-
tive to conventional engines.  It  is conceivable  that the prechamber
concept might be applicable also to new and existing stationary engines.
                                5-19

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 5.2. 1. 13      Exhaust Gas Recirculation
               Exhaust gas recirculation (EGR) is a simple and effec-
 tive method of NO  control in spark-ignition engines.  Over 80-percent
                  X
 reduction of NO   might be achieved by recycling 15 to 20 percent of the
               ji
 exhaust gas.  However,  this  technique results  in a considerable fuel
 economy loss and sometimes an increase in HC.  If only five  to ten per-
 cent of the exhaust is recycled, the fuel economy loss may be compen-
 sated for by advancing the spark timing or leaning the mixture.  In this
 case, NO  might be reduced by as much as 40  percent.  However, a
 number of potential problem areas related to the long term effect of
 EGR on engine life would have to be resolved before incorporation into
 stationary engines could be seriously considered.
 5.2. 1. 14      Water Injection
               Water injection is a very effective  NO  abatement tech-
 nique for  spark ignition  engines.  If water is ingested by the engine in
 finely atomized form, cooling of the intake charge would improve the
 volumetric efficiency of the engine and would result in lower  NO  emis-
                                                              X.
 sions.   For example, an 80-percent reduction  might be  achieved by
injecting water at a rate equal to the fuel flow.   This improvement in
NO  might be accompanied by small increases  in HC and specific fuel
   jf.
consumption.
5.2.1.15      Fuel Modifications
              Conversion from liquid fuels such as the gasoline to
gaseous fuels (LPG,  natural gas,  etc. ) is relatively simple and results
in a  substantial reduction in the exhaust emissions.  This approach
appears to be an attractive one for use in stationary engines.
5.2.1.16      Thermal Reactors
              Thermal reactors have been shown to be quite  effective
in reducing the HC and CO emissions from automotive spark-ignition
                                5-20

-------
engines, particularly in conjunction with other emission control tech-
niques such as EGR and spark retard.  Although NO  is not altered in
                                                  X
thermal reactors, these devices might be useful for those  stationary
engines that have high HC and CO emissions.
5.2. 1. 17       Catalytic Converters
               Like  thermal reactors, catalytic converters, when fresh,
are very effective in reducing the HC and CO emissions from gasoline
and gas-fueled spark-ignition engines.   However,  a number of potential
problem areas related to catalyst performance degradation would have
to be resolved before these devices could be seriously considered for
use in stationary engines.   At this time,  there are no reducing catalysts
available for NO  control, that would be  effective  in the oxidizing
                X.
atmosphere typical of  many stationary  spark-ignition engines.
5.2. 1. 18       Control of Emission From Blow-by,
               Carburetor and Fuel Tank
               Positive crankcase ventilation and vapor recovery sys-
tems have  become a part of the emission control  systems used in cur-
rent model year automobiles.  These systems, which are designed to
eliminate HC emissions from these sources have little impact on the
operation of the engine and  are directly applicable to stationary engines.
5. 2. 1. 19       Combined Emission Controls
               Several combinations  of the previously discussed emis-
sion control devices are applicable to stationary spark-ignition engines.
The  system selection depends  on many factors including engine applica-
tion,  engine size,  type of fuel  used,  duty cycle,  environmental factors,
desired emission reduction, and permissable loss in specific fuel
consumption.
               Determination of the optimum emission control system
for the various spark ignition engine classes requires a comprehensive
optimization study considering many factors such  as fuel consumption,
maintenance requirements and effects on other engine components.
                                5-21

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 5.2.2         Economic Considerations
 5.2.2. 1       Emission Control Equipment Cost
               A number of studies have been conducted to establish the
 economics  of emission control equipment installed either  as  retrofit
 systems in existing automobiles or as accessories in new engines
 (Refs.  5-1  through 5-4).  Although these devices were designed for use
 in automotive spark ignition gasoline engines,  the cost figures as well
 as the cost-effectiveness index can serve as guideline values in estimat-
 ing the installation and operational costs of similar equipment on sta-
 tionary gasoline engines.
               Table 5-1 (Ref.  5-5) presents the estimated sticker
 prices for emission control hardware for automobiles ranging from
 positive crankcase ventilation systems to dual-catalyst systems.
 Table 5-1  shows the  cost of individual components  and of total systems
 installed in particular  model year vehciles.  The first column presents
 either value added or hardware  cost including material, labor, over-
 head,  and G&A.  These cost values provide a base for estimating the
installation cost of similar devices on stationary gasoline  engines.
Some of the listed emission control components are available from the
shelf;  others may require modification for use in stationary engines.
Of course, the cost of the modified devices  might be substantially higher
than for the shelf units listed in  Table 5-1.
              A meaningful assessment of the cost (per unit horsepower
output) of emission control systems or devices for large stationary  gas
engines is not possible  at this time.  However, it is conceivable that
the cost of  these systems might be many times higher than for equiva-
lent systems used in automotive  gasoline engines.
                                5-22

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TABLE 5-1.
ESTIMATES OF STICKER PRICES FOR EMISSIONS
HARDWARE FROM 1966 UNCONTROLLED VEHCILES
TO 1976 DUAL-CATALYST SYSTEMS (Ref. 5-5)
Model
Year
1966
1968
1970




1971-
1972




1973






Configuration
PCV-Crank Case
Fuel Evaporation
System
Carburetor Air/Fuel Ratio
Compression Ratio
Ignition Timing
Transmission Control
System
Total 1970
Anti-Dieseling
Solenoid
Thermo Air Valve
Choke Heat By-Pass
Assembly Line Tests,
Calif (1/10 vol)
Total 1971-72
OSAC (Spark Advance
Control)
Transmission Changes
(some models)
Induction Hardened Valve
Seats (4 and 6 cyl)
ECR (11 - 14%)
Exhaust Recirculation
Air Pump — Air
Injection System
Quality Audit, Assembly
Line (1/10 vol)
Total 1973
Typical Hardware
Value
Added
1.90
9.07
0.61
1.24
0. 61
2.49

3.07
2.49
2.74
0. 18

0.48
0.63
0.72
5.48
27. 16
0. 23

List
Price
2.85
14. 25
0.95
1. 90
0.95
3.80

4.75
3.80
4. 18
0. 57

0. 95
0.95
1.90.
9. 50
43. 32
0.38

Excise
Tax
0. 15
0.75
0.05
0. 10
0. 05
0. 20

0.25
0.20
0. 22
0. 03

0. 05
0. 05
0. 10
0.50
2. 28
0. 02

Sticker
Price
3. 00
15. 00
1. 00
2.00
1. 00
4.00
8.00
5.00
4.00
4.40
0. 60
14.00
1. 00
1. 00
2.00
10.00
45.60
0.40
60.00
                          5-23

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TABLE 5-1  (continued)
Model
Year
1974





1975














1976


Configuration
Induction Hardened
Valve Seat V-8
Some Proportional EGR
(1/10 vol at $52)
Precision Cams, Bores,
and Pistons
Pretest Engines —
Emissions
Calif. Catalytic Converter
System (1/10 vol at $64)
Total 1974
Proportional EGR
(acceleration-
decelc ration)
New Design Carburetor
with Altitude
Compensation
Hot Spot Intake Manifold
Electric Choke (element)
Electronic Distributor
(pointless)
New Timing Control
Catalytic — Oxidizing-
Converter
Pellet Charge (6 Ib at
$2/lb)
Cooling System Changes
Underhood Temperature
Materials
Body Revisions
Welding Presses
Assembly Line Changes
End of Line Test
Go/No-Go
Quality Emission Test
Total 1975
2 NO Catalytic Converters3
x a
Electronic Control
Sensors
Total 1976
Typical Hardware
Value
Added
0.72
3. 21
2. 44
1. 80
4. 02

20. 07
7. 52
2.87
2.67
4. 35
1. 40
18.86
12. 00
1. 17
0. 63
0. 67
0. 13
1.85
1.22

22. 00
28.00
3. 00

List
Price
1. 90
4. 94
3. 80
2.85
6. 08

30. 02
14. 25
4.75
4.75
9. 50
2. 85
34. 20
20. 52
1.90
0.95
J.90
0. 95
2.85
1.90

37. 05
47.50
5.70

Excise
Tax
0. 10
0. 26
0. 20
0. 15
0.32

1. 58
0.75
0. 25
0. 25
0. 50
0. 15
1. 80
1. 08
0. 10
0. 05
0. 10
0. 05
0. 15
0. 10

1.95
2. 50
0. 30

Sticker
Price
2. 00
5. 20
4. 00
3. 00
6. 40
20. 60
31. 60
15.00
5. 00
5. 00
10. 00
3.00
36.00
21. 60
2. 00
1. 00
2.00
1.00
3. 00
2. 00
138. 20
39. 00
50. 00
6. 00
134. 00
1976 most common configuration
                              5-24

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5.2.2.2       Operating Cost
              The general correlation trend of fuel penalty (increase
in SFC) as a function of the reduction of NO emission for  several auto-
                                          Jf,
motive emission control systems has been presented in Figure 4-26.
Although the data on which this correlation is based are results from
extensive testing and evaluation of the emission control systems by
several companies, an interpretation of this correlation for estimates
of operating costs of stationary gasoline engines equipped with similar
systems is  subject to the following considerations.  The emission test
data were obtained on the basis of a typical car driving schedule or
standard CVS test cycle.   Therefore, the  relative emission reduction
at rated engine condition may be different.  To satisfy the  requirement
of good car driveability, the automotive emission control systems  are
in most cases designed for rich air-fuel mixture operation.  In sta-
tionary engines operating mostly at constant speed, leaner mixtures
can be tolerated and therefore the fuel penalty associated with the
incorporation of emission  control systems might be lower  than in the
case of automotive engines.  At this  time,  there is insufficient data
available to permit a meaningful evaluation of these factors.
5.3           GAS TURBINES
5.3.1          Economics of Emission Control
               Based on the discussions presented in Section 4.3, it is
apparent that incorporation of emission control devices into stationary
gas turbines would involve some additional costs either in the form of
special R&D efforts aimed at emission reduction or for special hard-
ware serving that purpose, or both.  These additional costs, which
include investment and operating costs, are expected to vary as a func-
tion of the design of the gas turbine, the locally established emission
regulations,  and the available resources such as water, for instance.
                                5-25

-------
The cost data shown below should be considered to be average values
rather than representative of a specific design.
              In order to obtain reference levels for the emission con-
trol costs, investment and operating costs of stationary gas turbines
operating without emission controls (except when indicated) are pre-
sented in the following subsection.
5.3. 1. 1       Investment Cost (Base Power Unit)
              The investment cost (installation and  equipment cost) of
gas turbines utilized in electric power generating plants varies  as a
function of plant size,  emission and noise levels, and location.  For
simple cycle peak load shaver installation in the continental United
States, typical 1973 cost data are as follows:  Output - $88.50/kw for
20 MW units (Ref.  5-6); $ 101. 30/kw for 26 MW units using water injec-
tion for NO  abatement (Ref.  5-7).  The investment  cost of combined
           ^C
cycle units for intermediate loads  (500-560 MW output) varies between
$132 and $l68/kw (Ref.  5-8).  The latter value is for a Puerto-Rico
installation and comparison with other data shows that the  continental
United States installations would cost approximately $30/kw less.  This
is in agreement with Ref. 5-9 which quotes an average price for com-
bined cycle units of $125/kw.  Ref. 5-10 quotes a combined cycle cost
of $130-$l60/kw and recent reports for a 700 MW combined cycle plant
indicate a cost of approximately $H6/kw.  According to Ref.  5-10,  the
cost of a regenerative gas turbine is higher than that of a  simple cycle
configuration.
5.3.1.2       Operational Cost (Base Power Unit)
              The cost per kwhr  energy produced is strongly affected
by (assuming fixed rate of interest, depreciation, taxes, and insurance)
the fuel cost which contributes approximately 50 percent for the peak-
ing unit and approximately 80 percent for the intermediate load, and on
                                5-26

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the annual usage rate.  Typically, an 80 MW simple cycle gas turbine
plant operated as a peaking unit produces energy at a  cost of 20 to
24 mils/kwhr.  In an intermediate load installation, the cost is about
13 to 15 mils/kwhr (fuel cost 80^/10  Btu,  1972 prices).
              For comparison, steam electric plants in the  1000 MW
size category require investment costs of the  order of $300/kw.  In
this case, the energy cost for continuous operation is about 8 mils/
kwhr, based on a fuel cost of 35^/10  Btu (Ref. 5-11).  Thus, it appears
that gas turbines operating at intermediate  load levels would approach
the operating cost of steam power plants  if  the  gas turbine  could oper-
ate with cheaper fuels or in conjunction with more efficient cycles such
as the combined or regenerative cycles.
5.3. 1.3       Near Term Control Cost
              In the absence of Federal regulations,  various  emission
limits have been established for gas turbines by many states,  counties
and cities (Ref.  5-12).  The interim emission controls used in these
installations  consist of moderate modifications to the existing com-
bustors, supplemented by water or steam injection, if required.
5.3.1.3.1     Combustor Modification Costs
              In several instances the combustor modifications were
implemented over a period  of several years and smoke reduction was
one of the early benefits. Combustor changes require careful
evaluation  with respect to their effect on the life of the hot end com-
ponents, since long time intervals between overhaul and  long engine
life are principal design requirements for stationary  gas turbines.
Consequently, evolution of combustors for these turbines is a slow
process, although  a  competitive  market and the national goal of
tighter emission controls as now evidenced in the automotive field,
will maintain the pressure for continuous  improvements.  The cost
                                5-27

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 associated with even moderate combustor improvements has to be
 amortized in the cost of units subsequently sold, and the cost increase
 (discounting inflation and rising labor cost) will vary from about one
 percent for gas turbines in the 30 MW class to over 10 percent for
 smaller units (3 MW or less) (Ref.  5-13).
 5.3.1.3.2     Water/Steam Injection Costs
               Water or steam injection are other methods  that have
 been adopted by various manufacturers (GE,  GM, TP&M, Westinghouse)
 to meet various NO  emission limits.  In the evaluation of these con-
                   x
 cepts,  cost data provided by  several gas turbine manufacturers and
 utilities were utilized.
               As previously stated in Section 5.3.1,  the baseline
 investment cost of uncontrolled simple cycle gas turbines is about
     to  100/kw and the operational cost (capital cost plus maintenance
 plus fuel), is 20 to 24 mils/kwhr for intermediate load turbines (6000
 hours per year; fuel cost at 80^/10  Btu).  The cost increases due to
 water or steam injection were computed from the above baseline data.
 As an example,  the investment cost breakdown for a water injection
 system consisting of combustor/injector modifications, water pump,
 valving, piping, storage, and water treatment equipment, as provided
 by San Diego Gas and Electric is shown in  Table 5-2.
               The investment cost of water injection can vary from
 approximately ten percent of the baseline cost for a 2-MW plant to
 approximately six percent for 49- and  81-MW plants.   The investment
 and operating costs of steam injection  systems are generally higher
 unless superheated steam is  available  from other plant sources.  This
 is illustrated in Table 5-3, which presents average data from other
 sources.
              As shown in Table 5-3,  the water injection investment
and operational cost is prohibitive for small gas turbines.  However,
                                5-28

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TABLE 5-2.  WATER INJECTION INVESTMENT COST
            (SAN DIEGO GAS AND ELECTRIC)
Control System
Combustor Modifications
including Water Injection
Nozzles
Water Injection Pumps
and Water Regulation
System
Associated Piping and
Water Storage Facilities
Water Treatment
Equipment
General Expenses includ-
ing Engineering, Adminis-
tration, Testing Taxes
TOTAL
Gas Turbine Size
20 mW

$1.00/kw
$3.54/kw
$1.72/kw
$0.90/kw
$ 1 . 1 5 / kw
$8.31/kw
49 mW

$0.86/kw
$2.88/kw
$1.05/kw
$0.47/kw
$0.82/kw
$6.07/kw
81 mW

$1.04/kw
$3.10/kw
$0.87/kw
$0.47/kw
$0.57/kw
$6.05/kw
 TABLE 5-3.  WATER/STEAM INJECTION COST AS A
             FUNCTION OF POWER PLANT SIZE



mW Output

0.26 (350 hp)
2. 90 (3900 hp)
30.00
33.00
65.00
:|i
For peaking gas

Investment
Cost,
Percent
Baseline
Water
100.0
18.0
10.0
7.3
7.3
Steam
150.0
Z4.0
12.0
10.6
10.6

Operational
Cost,
Percent
Baseline
Water
55.0
6.5
6.0
5.7
5.7
Steam
165
32
—
—
"
turbine, 1 000 hour /year
                       5-29

-------
 since the NO  emissions are  inherently lower for smaller  gas turbine
 units (because of the reduced residence time of the combustion products
 in the chamber), water injection may not be required for this engine
 category, particularly if less stringent emission standards would be
 imposed on the  smaller  engines.
               The operational cost of water injection decreases with
 increasing gas turbine usage  and for intermediate loads (6000 hour/
 year) it might approximate 2. 5 percent of the basic operating cost of
 the plant.
               No operational steam injection costs are available for
 the larger power plants.  However, it appears that the higher gas tur-
 bine efficiency achieved with  steam injection combined with the  higher
 power output  capability of the engine would compensate for the cost of
 steam generation.
               The cost  of water injection and its effect on NO   abate-
                                                           A.
 ment cost is further illustrated in  the following example.   In this case,
 a simple cycle 30 MW plant is considered having uncontrolled emissions
 of 220 ppm of NO  (Figure 3-31), which corresponds to approximately
                 X
 420 pound/hour  of NO .   As indicated in Figure  4-80, the use of water
 injection at a  rate of about 80 percent of the fuel flow rate would reduce
 the NO  by 75 percent, or 315 pound/hour.
              For a peak shaving turbine  (1000 hours/year), the
 operating cost would be 5.7 percent of, say 20 mils/kwhr, or $34/hour.
 Thus, one pound reduction of  NO  is achieved at a cost of approximately
                               j£>
 11 cents,  which corresponds to $216 for every ton of NO  removed.
                                                      X.
              For the same gas  turbine operating at intermediate loads
 (6000 hours/year), the water  injection cost would be only $10. 50 per
hour or approximately 33 mils per pound NO reduction, which  corre-
 sponds to an added operational expense of $67 per ton of NO  removed.
                                                         j"C
 Thus, NO  control becomes more cost effective  with increasing load
         J\.
factors.
                                5-30

-------
              Water and steam injection is currently being used by
some utilities (for instance,  San Diego Gas and Electric).  To date,
deleterious effects have been observed on the combustors, turbine
buckets, and nozzles.  However, several years of operation are nec-
essary to obtain meaningful, long-life data.  Some gas turbine manu-
facturers feel that about two years of research work costing several
millions of dollars would be  required for the development of a proven
water injection system.
5.3.1.4       Far Term Control Costs
              As discussed  in Section  4.3.3.3, the development  of
advanced low emission combustors and improved fuel delivery systems
is required to meet future emission goals with "dry operation."  The
time and cost of such a development program will be substantial.  It is
estimated that the basic development would require approximately three
to six years with an  additional  two to four years  required for field
testing.  According to Ref.  5-13, over four million dollars would be
required for the development of a small 3 mW gas turbine combustor.
The development of a low emission combustor for a large aircraft gas
turbine is of the order of $100  million (Ref.  5-14).  Stationary gas  tur-
bine combustor developments will probably be close to the latter figure.
              The progress reportedly achieved with externally
mounted combustors  (Section 4.3.3.3) cannot be overlooked.  If indeed
such combustors can meet the  emission limits of San Diego County (see
Table 3-20),  an advanced,  low-emission combustor may appear on the
scene sooner than expected,  since the highly competitive gas turbine
market will not allow other manufacturers to lag far behind.
5.3. 1.5       Emission Controls Evaluation
              The discussion and data presented in Sections 4.3 and
5.3 permit a general evaluation of the various  emission control  devices
for near and far term application.
                                5-31

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               The interim devices include minor combustor modifica-
tion and water/steam injection.  The far term controls include exten-
sive changes in the in-line combustor, closed  cycle gas turbines (see
Section 3.3. 1.4), and externally-mounted  combustor gas turbines.
5.3. 1.5. 1     Interim Emission Control Devices
               The modest combustor modifications consisting of lean-
ing the primary zone, internal gas  recirculation, and air-assisted fuel
injection are, in many cases, already in operation.  These devices are
generally not  sufficient to meet  the more stringent emission limits,
except smoke (see Table 3-20),  but they are undoubtedly steps in the
right direction.  The incremental cost of such changes is small and
the main attention must be paid  to the long-term effect on the life of
the hot end components.
               Water or steam injection is very effective in the reduc-
tion of NO  .  It appears possible to obtain this reduction without any
          n
increase in HC or CO by tailoring the water  or steam injection system
to the particular combustor design. The investment cost increment for
water injection equipment is acceptable, at least for units above 20 MW
output,  amounting to approximately ten percent of the basic unit cost.
The operational cost of water injection adds  only a few percent to the
basic operating cost.
               The effect of water or steam injection on long term engine
life is not yet known,  but operation up to date did not encounter any
major problems.
               One of the disadvantages of  water injection is that it
limits the siting flexibility of simple cycle and regenerative gas turbines
which normally require no water supply.   This is of particular impor-
tance for areas with limited water  sources or  subfreezing ambient
temperatures.  Another disadvantage is related to  an attendant reduc-
tion in engine  efficiency which gains importance  with the rising cost of
fuels.
                                 5-32

-------
              Therefore, water or steam injection should be considered
to be useful stop-gap arrangements which permit compliance with
most of the existing or planned regulations,  until a "dry" advanced
combustor becomes available.
5. 3. 1. 5. 2    Far Term Emission Control Devices
              The advanced in-line combustor with prevaporized,  pre-
mixed fuel-air charge, short residence time,  and lean primary com-
bustor, when developed, will be able to meet very stringent emission
regulations.  However, such a combustor might not become available
until the 1980s.
              The closed Brayton cycle gas turbine,  if successful,  will
probably be limited to smaller units (less than 20 MW) because  of higher
investment cost and greater  complexity.  The main advantage of this
engine type is its flexibility in burning a variety of fuels,  and the
applicability of low emission combustion system technology developed
for both steam boilers and gas turbines.
              Similar comments apply to external combustors  pro-
jected for use in open cycle gas turbines.  In this design, the volu-
metric constraints of in-line  combustors characteristic of mobile gas
turbines are alleviated,  thus permitting more effective utilization of
more conventional emission control techniques, including staged com-
bustion and flue gas recirculation.  Although gas turbines with external
combustors are now in production, no design and performance details
are currently available from these engines.  These configurations
which are projected to be fully developed by the late 1970s may pro-
vide an early answer to the low emission "dry" combustor.
                                5-33

-------
                           REFERENCES
5-1.     Bascunana, J. L. and Webb,  M. J. ,  "Effectiveness and Cost
         of Retrofit Emission Control Systems for Used Motor Vehicles, "
         SAE Paper No.  720938,  SAE Meeting in Tulsa, Oklahome,
         November 1972.

5-2.     "Medium Duty Vehicle Emission Control Cost Effectiveness
         Comparisons, "  Volumes I and II, Aerospace Report
         No.  ATR-73(7327)-l, Urban Programs Division,  The
         Aerospace Corporation, El Segundo, California.

5-3.     "Lead Cost-Benefit Study, " Interim Report No. TOR-0172
         (2787)-!, The Aerospace Corporation, El Segundo,
         California,  September 1971.

5-4.     "An Assessment of the Effects of L/ead Additives in Gasoline
         on Emission Control Systems Which Might  Be Used to Meet
         the 1975-76 Motor Vehicle Emission Standards, "  Final Report
         No.  TOR-0172(2787)-2,  The Aerospace Corporation,  El
         Segundo,  California, November 1971.

5-5.     "Report by the Committee on Motor Vehicle Emissions, "  The
         Environmental Protection Agency and the National Academy
         of Sciences,  NAS, Washington,  D.C.,  February 1973.

5-6.     Gas Turbine World,  February-March 1973.

5-7.     Gas Turbine World,  May 1973.

5-8.     Gas Turbine World,  November  1973.

5-9.     Sawyer, J.  W. ,  "Gas Turbines in Utility Power Generation, "
         Sawyer's Gas Turbine Catalogue (1973).

5-10.    Carlson,  H.  W. ,  "The STAG  Cycle, " General Electric
         Report USDA-4-72 (September 1972).

5-11.    "18th Steam Station Survey,"  Electrical World, November 1,
         1973.

5-12.    Gas Turbine World,  February-March 1973.

5-13.    "Response to  Preliminary (draft) Proposed Standards for  Con-
         trol of Air Pollution from Stationary Gas Turbines, " General
         Motors (March 1973).
                                5-34

-------
5-14.   Robinson, C. A. ,  "NASA Plans Award on Engine Emissions, "
        Aviation Week and Space Technology (8 April 1974).
                                5-35

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                             SECTION 6

                    RECOMMENDED PROGRAMS


              Based on an evaluation of the available test data, a
number of emission control approaches have been identified which
might offer  substantial improvements in the emissions of stationary
engines without harmful side effects  relative to specific fuel consump-
tion and engine life.  However,  significant data gaps exist in most  cases
and these would have to be  resolved before a final assessment of the
applicability of these techniques to stationary engines would be possible.
The research programs outlined below are aimed at providing part of
the required information.
6.  1           DIESEL ENGINES
              Because of the long service life of most stationary
diesels,  it appears that retrofitting of existing engines might be
required to  achieve any near-term impact on the emissions from
these engines. Two potentially attractive techniques, water injection
and EGR, have been identified which might be applicable to both new
and existing engines.
                                 6-1

-------
         1.    Although exploratory testing of emulsified fuels has not
              been successful in one engine,  it is suggested that this
              technique be further evaluated because of certain advan-
              tages in terms of design simplicity, effectiveness, and
              reliability inherent in this technique relative  to water
              injection into the intake manifold.  The first phase of the
              proposed experimental program would be conducted on a
              single-cylinder prechamber or open-chamber engine and
              would be concerned with a parametric evaluation  of NO ,
              SFC, and water-to-fuel flow ratio in the emulsion.  In
              these tests,  different fuels containing varying amounts of
              fuel-bound nitrogen would be utilized to determine the
              NO conversion factor under various engine operating con-
              ditions.   Upon successful conclusion of this program
              phase,  durability testing would be performed using emul-
              sified fuels to establish the  long-term effects of the water
              on important engine components.  Another objective of
              this particular program phase would be the determination
              of the required water purity, which has a significant
              effect on emission control economics.  If these single
              cylinder  tests produce satisfactory results, follow-on
              work with multicylinder engines should be undertaken.
         2.    EGR has been shown to be a rather cost-effective NOX
              abatement technique for diesel engines, particularly in
              conjunction with intake air cooling.  However, there is a
              lack of  information relative to the long-term  effects of
              EGR on the engine and the required cleanliness of the
              recycle  gas.  It is recommended that these parameters
              be evaluated experimentally on a single-cylinder  engine.
              Based on these data, an optimum EGR system configura-
              tion would be identified with respect to EGR temperature,
              filter system, and fuel quality requirements.   Upon
              successful completion of this initial program phase,  test-
              ing of the optimum  system would be performed in a suit-
              able multicylinder engine.   Results of this program
              could be  also extrapolated to the  case of the stationary
              spark-ignition engine.

6.2      .     SPARK IGNITION ENGINES
         1.    The specific mass emission of NOX from large stationary
              gas  engines is substantially higher than from stationary
              gasoline engines, in spite of the fact that gas engines
              generally operate with very lean fuel-air mixtures.  Lean
              operation in automotive engines normally results in
              reduced NO  formation.  Several causes such as high
                                 6-2

-------
              local temperatures in the large combustion chambers,
              combined with relatively long residence times of the
              reacting species in the high temperature environment
              and stratification of the mixture in the combustion cham-
              ber, might contribute to the high NOx emission from
              these engines.  To resolve these potential problem areas,
              it is recommended that an experimental program be con-
              ducted on a typical low-speed gas-fueled stationary
              engine to determine the effects  of all engine variables  on
              NOX , including  mixture homogenation, residual gas con-
              tent,  wall temperature, and species residence time.
              These results would enable determination of changes in
              design and/or operation which would be required to
              reduce NO .
                        x

6.3           GAS TURBINES

         1.    The interim emission control techniques currently used
              in some  stationary gas turbines  require more basic
              research and  development. In particular, the limited
              data available on water and steam injection show incon-
              sistencies relative to HC and  CO.  It is recommended,
              therefore, that an experimental  program utilizing a
              small gas turbine be conducted  to determine the effect
              of water atomization,  degree  of  fuel-water (steam)  mix-
              ing,  and water-fuel ratio on the  emissions produced.
              Other operational parameters of the gas turbine would
              be monitored  during the program to obtain the compara-
              tive (dry  and wet) performance  data required for a com-
              prehensive assessment of water and steam injection.

         2.    Because of the extremely low NO  emission potential of
              prevaporized, premixed combustors,  (as evidenced  by
              analysis and limited experimental modeling) and the lack
              of applicable test data, it is recommended that a research
              program  be performed which  is  aimed at the develop-
              ment of such a combustor.  The initial development work
              should be performed in the 100  to 200 hp equivalent
              category.  The principal objectives of the proposed
              program  would include the demonstration of concept
              feasibility, emission performance,  combustion effi-
              ciency, and operational safety.

         3.    A third desirable program, which would be both analyti-
              cal and experimental in nature,  would be directed to a
              study of an externally-mounted,  low-emission com-
              bustor.  The goal in this program would be the demon-
              stration of the benefits realized  by the utilization of the
                                 6-3

-------
               greater volume and design freedom associated with the
               external combustor.  The design goals of the combustor
               proper are concomitant emission  reduction,  high com-
               bustion efficiency, and uniform outlet temperature
               distribution.

         4.     A further area for basic research is in the catalytic
               combustor.  A program in this area is required to
               resolve known potential problems before their appli-
               cation to stationary gas turbines can be further seriously
               considered.  This  research program should  explore the
               effects and characteristics in the area of catalyst dura-
               bility, specific heat release rate,  ignition characteristics,
               and pressure drop. Initial evaluations in small scale
               would be adequate.  Larger scale experiments would not
               be recommended until a satisfactory demonstration of
               potential in the small  scale experiments.

6.4            EMISSION INVENTORY DATA

               The relative importance of reducing stationary engine

emissions,  and hence the relative importance of implementing control
techniques for  stationary engines, is largely dependent upon the total

emission inventory in the region of interest. An accurate tabulation of

the contribution of the various stationary  engine classes is  not available
today.   In order  to evaluate the relative need to implement  research
programs in the  stationary engine emission control area, it will be

necessary to first compile a more accurate inventory of stationary

engine emissions and evaluate their  impact  relative to other sources
in the air quality control  region.  It  is recommended that a  study to
develop this inventory comparison be implemented to enable more

comprehensive planning for stationary engine emission control

research.
                                 6-4

-------
                            APPENDIX A
               The following summary table lists the total estimated
installed horsepower of all stationary engines in the United States in
1971.  The data listed in this table should not be used to estimate total
emission contributions or fuel usage, since these factors do not reflect
actual engine duty cycles.
                                 A-l

-------
TABLE A-l.  ESTIMATED INSTALLED STATIONARY ENGINE HORSEPOWER - 1971
Application
Electric Power
Generation
Oil and Gas
Pipelines
Natural Gas
Processing
Oil and Gas
Exploration
Crude Oil
Production
Natural Gas
Production
Agricultural
Industrial
Process
Municipal Water
and Sewage
TOTAL
Diesel Engine hp
Diesel
Fuel
1,570, 000a
830,000
1,500,000
-
7,500,000
—
465,000
1 1,865,000
Dual
Fuel
3,710,000a
390,000
—
—
—
—
—
4,100,000
Spark Ignition Engine hp
Gas
90,000a
10,990,000
2,410,000
500,000
852,000
3,237,000
—
230,000
465,000
18,774,000
Gasoline
i



800, 000, 000b
For all
Applications




800,000




,ooob
Gas
Turbine
hp
29,000,000
8,746,000°
~*
—
—
—
-
37,746,000
aEstimated 1970 data
See Section 3. 2. 2. 1 for basis of estimates
CInstalled in the 1958-1970 time period

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                            APPENDIX B


                      VISITS AND CONTACTS
              During the data gathering phase of the study the following
organizations were visited or contacted by telephone.

        Organization                           Primary Contact(s)

Allis Chalmers Corp.                        Mr. E.G. Goodpaster
Harvey,  Illinois

Avco Lycoming                               Dr.  M. Bentele
Stratford, Connecticut                        Mr. H. Grady
                                             Mr. G. Opdyke
                                             Mr. N. Marchionna

Caterpillar  Tractor Company                 Mr. R.E. Bosecker
Peoria, Illinois                               Mr. K.J.  Fleck
                                             Mr. R.D. Henderson

Chrysler Corporation                        Mr. L.K. Haddock
Detroit, Michigan                            Mr. C.M. Heinen
                                             Mr. J. Hurst

Colt Industries                               Mr. W.A. Brill
Beloit, Wisconsin                            Mr. C.L. Newton
                                B-l

-------
        Organization
  Primary Contact(s)
Cooper Bessemer Corporation
Mount Vernon, Ohio

Cummins Engine Corporation
Columbus, Indiana
Curtiss Wright Corp.
Woodridge, New Jersey

Diesel Engine Manufacturers
  Association
Beloit,  Wisconsin

Engine Manufactures Association
Chicago, Illinois

General Electric
Schenectady, New York
General Motors
Warren, Michigan

Ingersoll Rand Corp.
Painted Post, New  York

International Harvester Company
Melrose Park, Illinois

Murphy Diesel Corp.
Milwaukee,  Wisconsin

Perkins Engine Company
Farmington, Michigan

San Diego Gas  & Electric Company
San Diego, California
Mr. J.W. Holmes
Mr. K. W. Stevenson

Mr. C.T.J. Ahlers
Mr. R.C. Bascom
Mr. A.W.  Carey, Jr.
Mr. P.R. Kahlenbeck
Dr.  W.T. Lyn

Mr. S.  Lombardo
Mr. F. Spindler

Mr. C.L. Newton
Mr. T.C. Young
Mr.  N.R. Dibelius
Dr. H.L. Hamilton
Mr.  M. Jarvis
Mr.  R.H. Johnson
Dr. E.W. Zeltmann

Mr.  G.P. Hanley
Dr. C.K. Powell
Mr.  W. Hutchins

Mr.  J.C. Basiletti
Mr. N.G. Beck

Mr.  L.D. Evans
Mr.  N. Hartwell
Mr.  H. Arfman
Mr.  R. Miller
                                B-2

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         Organization

Solar, Division of International
 Harvester
San Diego, California

Turbopower and Marine Systems,
 Subsidiary of United Aircraft
 Corp.
Farmington,  Connecticut

Waukesha Motor Corp.
Waukesha, Wisconsin
Westinghouse Electric
 Corporation
Lester, Pennsylvania

White Motor  Corporation
Springfield, Ohio

Worthington - CEI, Inc.
Buffalo, New York
 Primary Contact(s)

Mr.  R. Kress
Mr.  P.A. Pitt
Mr.  W. Brazel
Mr.  F. Cartona
Mr.  J. Sleeper
Mr.  G. Stebbins

Mr.  N. Cox
Mr.  M. Groenewold
Mr.  W. Snyder

Mr.  M. Ambrose
Mr.  J. Farrow
Mr.  R.E. Strong

Mr.  S. Lamb
Mr.  J.M. Buckley

Mr.  L. Atwood
Mr.  E.L. Case
                                B-3

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                            APPENDIX C

               UNITS OF MEASURE - CONVERSIONS
              Environmental Protection Agency policy is to express all
measurements in Agency documents in metric units.  With a few excep-
tions, this report uses British units. For conversion to the.metric sys-
tem, use the following conversions:
     To convert from               to               Multiply by
       °F                     °C                   5/9 (°F-32)
       ft                      meters               0.304
       ft2                     meters               0.0929
        3                            3
       ft                      meters               0.0283
       in                      cm                  2.54
       •  2                       2                 ,  .-
       in                      cm                  6.45
       BTU                   kcal                 0.252
       BTU/lb                cal/g                0.556
       kW/m3                kW/m3               35.3
       hp                     kW                  0.746
       lb/106BTU            g/106cal            1.80
       lb/in2                 mm Hg               51.71
       Ib/hr                  g/hr                 453.6

                                C-l

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                            GLOSSARY
AAPS         Advanced Automotive Power Systems




A/F          air-fuel ratio




ANSI         American National Standards Institute




ATC          after top center




Bhp          Brake horsepower




BMEP        brake mean effective pressure




BSFC         brake specific fuel consumption




BTDC         before top dead center




CID          cubic inch displacement




CO           carbon monoxide




CRC          Coordinating Research Council




CVS-CH      constant volume sampling-cold, hot




DBA          Detroit Diesel Allison




DEMA        Diesel Engine  Manufacturers Association




EGR          exhaust gas recirculation







                                G-l

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EPA           Environmental Protection Agency



ERG           external recirculation combustor



EVC           externally vaporizing combustor



FID           flame ionization detector



GE            General Electric Company



GM            General Motors  Corporation



HC            hydrocarbons



ICI            Imperial Chemical Industries



JPL           Jet Propulsion Laboratory



LAAPCD      Los Angeles Air Pollution Control District



LPG           liquified petroleum gas



LTR           Lean thermal reactor



MIT           Massachusetts Institute of Technology



MW           megawatts



NASA          National Aeronautics and Space Administration



NDIR          nondispersive infrared



NEMA         National Electrical Manufacturers Association



NO            nitric oxide



NO            oxides of nitrogen
   x                        °


PDS           phenoldisulfonic acid



ppm           parts  per million



Q/I            quality/intensity



S/V           surface-to-volume ratio



R&D           research and development
                                G-2

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RTR           rich thermal reactor



scf            standard cubic foot



SFC           specific fuel consumption



SO             oxides of sulfur
   x


TDC           top dead center



TPM           Turbo Power and Marine Systems



UOP           Universal Oil Products



WOT           wide open throttle
                                G-3

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                                 TECHNICAL REPOHT DATA
                          (Please read luitrin'lions on llic /rivnr hcjiirc
 I. HbCOHT NO.
 EPA-650/2-74-051
                                                       3. RECIPIENT'S ACCESSION- NO.
•1. TITLE ANDSUOTITLE
 Assessment of the Applicability of Automotive
  Emission Control Technology to Stationary Engines
            b. REPORT DATE
            July 1974
            G. PERFORMING ORGANIZATION CODE
7. AUTHOR(S)
                                                       8. PERFORMING ORGANIZATION RLPORT NO
 W.U.  Roessler, A.  Muraszew, and R. D.  Kopa
            ATR-74 (7421)-!
9. PERFORMING ORGANIZATION NAME AND ADDRESS
Urban Programs Division
The Aerospace Corporation
ElSegundo, California 90245
            10. PROGRAM ELEMENT NO.
            1AB014; ROAP 21ADG-084
            11. CONTRACT/GRANT NO.

            Grant R-802270
 12. SPONSORING AGENCY NAME AND ADDRESS
 EPA, Office of Research and Development
 NERC-RTP, Control Systems Laboratory
 Research Triangle Park, NC 27711
            13. TYPE OF REPORT AND PERIOD COVERED
            Final
            14. SPONSORING AGENCY CODE
15. SUPPLEMENTARY NOTES
16. ABSTRACT
The report gives a review of the emission characteristics of uncontrolled stationary
diesel, spark ignition, and gas turbine engines, including an analysis and evaluation
of the applicability of automotive emission control technology to stationary engines.
Nitrogen oxides have been identified to be the principal pollutant species emitted
from these engines.  In principle, the emission control techniques developed or
evaluated for spark ignition, diesel, and gas turbine engines are  applicable to
stationary engines.  However, in most  cases, the emission reductions achieved with
these techniques are accompanied by sizeable losses in specific fuel consumption
and uncertainties relative to the effect  of these  techniques on engine life and control
system durability.
 7.
                              KEY WORDS AND DOCUMENT ANALYSIS
                 DESCRIPTORS
                                           Ib.lDENTIFIERS/OPEN ENDED TERMS
                         c. COSATI Held/Group
Air Pollution
Stationary Engines
Diesel Engines
Spark Ignition Engines
Gas Turbine Engines
Nitrogen Oxides
Fuel Consumption 	
A.ir Pollution Control
Stationary Sources
Emission Characteristics
\utomotive Emission
  Controls
Engine Life
13B
21G
21E
07B
21D_
18. DISTRIHUTION STATEMENT
                                           19. SECURITY CLASS (Tliii keport)
                                           Unclassified
Unlimited
                                                                    21
                            NO. OF PAGES
                                362
20. SECURITY CLASS (Thispage)
Unclassified
                                                                    22. PRICE
EPA Form 2220-1 (9-73)

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