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.
3-1
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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
3-2
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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,
3-3
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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
3-5
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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.
3-6
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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.
3.7
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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
3-8
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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.
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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
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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
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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
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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
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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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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
-------
100
UJ
u.
£ 10
I 1
111
o
o
0.1
DETRIOT DIESEL
ALLISON ADVANCED
COMBUSTOR
I T I I II
'CONVENTIONAL:
COMBUSTORS :
O MODEL 250-
D MODEL 501
AGMA 100
0 P&W JT9D
GMA300
EMISSION 1
TECHNOLOGY
ADVANCEMENTS
i i i i i 111
MAXIMUM ~
POWER
I i i i i 111
0.1 1 10
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
-------
U.14
0.12
"o
£0.16
§
^0.08
0.04
0
II I I i l i
0 STANDARD COMBUSTOR
FUEL OIL o
• STANDARD COMBUSTOR
GAS
o
D
0
•• o -
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
-------
10
JE 9
^8
IT
!•
H5
i4
13
O
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
jg
| 15
•,
§ 10
2
_i
§"<
or
<
Q.
1 1 1 1 1 1
-
o__ DISTILLATE OIL
-o
- A 1 1 1 1 1 1
) 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
-------
LO
I
O
00
.0
•o
QJ
tu
90
80
70
X 60
L.
I 50
. 40
O p
z i 30
O
If*
20
10
0
RE: NEMA STANDARDS
PUB. NO. SM 33-1964
4
I
6
I
8
J
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
«/> 40
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
o*
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
-------
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,
"Design Factors That Affect Diesel Emissions, " SAE Paper
710484.
3-3. V. S. Yumlu and A. W. Carey, "Exhaust Emission Charac-
teristics of Four-Stroke, Direct-Injection, Compression
Ignition Engines, " SAE Paper 680420.
3-4. "Cummins Power-Logging, Construction and Mining, "
Cummins Engine Company Bulletin No. 952790 (February
1972).
3-5. R. Mueller and L. Lacey, "New International Harvester Heavy
Duty Diesel Engines, " SAE Paper 993 A (January 1965).
3-6. R. Malcolm, "International's New Motor Truck V-8 Engines, "
SAE Paper 660076 (January 1966).
3-7. H. Ricardo, "The High-Speed Internal Combustion Engine, "
Blackie and Syn Ltd. , London (1953).
3-8. E. Eisele, "Daimler Benz Passenger Car Diesel Engines -
Highlights of 30 Years of Development, " SAE Paper 680089
(January 1968).
3-9. C. R. McGowin, "Stationary Internal Combustion Engines in
the United States, " EPA R2-73-210, Shell Development
Company (April 1973).
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
(January 1973);
3-12. W. F. Marshall and R. D. Fleming, "Diesel Emissions Rein-
ventoried, " Bureau of Mines Report PB-201896 (July 1971).
3-111
-------
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
Model for the Formation of Nitric Oxide in Direct Injection
Diesel Engines and its Application in Parametric Studies, "
SAE Paper 730083 (January 1973).
3-15. G. Blair Martin and E. E. Berkau, "An Investigation of the
Conversion of Various Fuel Nitrogen Compounds to Nitrogen
Oxides in Oil Combusion, " AIChE National Meeting
(30 August 1971).
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,
37 FR 175 (September 8, 1972).
3-18. "Exhaust Emissions from Uncontrolled Vehicles and
Related Equipment using Internal Combustion Engines, "
APTD-1490, Part I, Locomotive Diesel Engines and Marine
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
-------
3-23. R. E. Bosecker and D. F. Webster, "Precombustion
Chamber Diesel Engine Emissions - A Progress Report,"
SAE Paper 710672 (August 1971).
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
-------
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
-------
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
-------
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
-------
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
-------
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
-------
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
-------
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
-------
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
-------
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
-------
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
-------
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
-------
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
-------
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
-------
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
-------
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
-------
_ 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
-------
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
-------
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
-------
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
-------
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
-------
•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
-------
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
-------
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
-------
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
-------
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
-------
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
-------
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
-------
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
-------
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
-------
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
'000
5
§ 6000
o
I 50»
5 4000
OC
g
3000
1
1
1
SINGLE-CrilNDiR ENGINE
CONSTANT POWER
MB! SPARK TIMING
IUII I
100
MO ^
o
z
400 g
|
z
MO g
o
01
?00 5
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
-------
0 700
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£ 500
£ 400
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£ £600
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20
15
10
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I
CYLINDER EXHAUST TEMP.
V "-•
SPARK PLUG GASKET TEMP.
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
-------
I I I
COMP. RATIO 6.7
-------
700
MX)
5 300
g
8 200
100 -
i \ r
SINGll-CYUNDtR [NGINt
CONSTANT POWIR
AIR-FUEL RATIO 16.0
FUEL ' (BY CHROMATOGRAPHYI
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
1600
1400
1200
B m
<
tt
5
g 600
§
too
200
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
-------
700
600
£ 500
400
en
i/i
a.
a:
CO
to
UJ
700
600
500
O
i 400
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ca
8000
I 7500
^ 7000
ui
CO
a.
x
CO
20
15
CYLINDER EXHAUST TEMP.
SPARK PLUG GASKET TEMP.
FIRING PRESSURE
MASS EMISSIONS
NOV
FUEL CONSUMPTION
to
1 5
l/>
5 0
UJ
HCT
£. A *
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1,1,1,1
02468
1
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IGNITION TIMING, 8BTDC
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
-------
S- 700
LU
^ 600
2 500
Q.
£ 400
O)
Q.
9 uj- 700
600
26
a:
Q.
IS
£2
a. uj
< 13
a: u.
JC
u a- 7500
u. I
CO \
2 7000
CD
a.
CQ
Z
20
15
10
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
90
BO
70
60
50
SPEED 1600 RPM
INTAKE DEPRESSION. 12 IN
SPARK TIMING
101 *v
36
HG
DEC. BTC
COOLANT
1M« v
141
*
\
160
COOLANT TEMP..
f
\
179*
199
/
"F
o
*mp
WATER
ETHYLENE GLYCOL '
120
So 140
\0
1M
""s, 180 271
^
200/
?00 240 280 320
AVC. SURFACE TEMP.. °f
360
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
-------
5000
O
Z 3000
7000
MA
CO
\
%
M. AIR PRES.
WP. RATIO
%
s.
^^
V> in. t
6.7
../
^*— .
9.
16 A/F
114 A/F
l» A/F
12 A/F
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
o
a)
5 300
u
O
a:
o
200
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
-------
1
EARLY-— j — "-LATE
i wmm INT. i
j EXH. | |
BC TC BC
~
NO A-
_O^
A"'
0
^*~-°
MBT
14.5:1 A/F
cTn 1200 rpm
5™- SOpsilMEP
1
ill i i i
30 10 -10 -30 -50
INTAKE VALVE OPENING, deg BTDC
_
—
-7C
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\
1200 rpm VI
50 psl IMEP
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
-------
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
-------
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
-------
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
-------
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
-------
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
-------
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
-------
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
-------
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
-------
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
-------
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
-------
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
-------
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
-------
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
-------
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
-------
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
-------
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
-------
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,
"Design Factors That Affect Diesel Emissions," SAE
paper No. 710484.
4-5. R. P. Wilson, Jr., E. B. Muir, and F. A. Pellicciotti,
"Emissions Study of a Single-Cylinder Diesel Engine, "
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
and Smoke Control Technology Evaluations — A Summary, "
ASME paper 69-WA/APC-3 (16-20 November 1969).
4-26. Presentation to the California Air Resources Board Staff,
Cummins Engine Company, 22 August 1973.
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
SAE Progress for Technology, Vehicle Emissions, Part II,
Vol. 12.
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
Chamber Deposits on Exhaust Hydrocarbon Emissions, " SAE
paper No. 486A, SAE National Automobile Week, Detroit
(March 1962); SAE Technical Progress Series, Vol. 6,
Vehicle Emissions, 175 (1964).
4-31. C. R. McGowin, F. S. Schaub, and R. L. Hubbard, "Emission
Control of a Stationary Two-stroke Spark-gas Engine by
Modification of Operating Conditions, " Shell Development
Company, Emeryville, California.
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, "
SAE paper No. 710835 (October 1971).
4-34. G. J. Nebel, and M. W. Jackson, "Some Factors Affecting
the Concentration of Oxides of Nitrogen in Exhaust Gases
from Spark Ignition Engines, " presented at the Symposium
on Air Pollution, American Chemical Society, New York,
New York (September 1957).
4-35. T. A. Huls, P. S. Myers, and O. A. Uyehara, "Spark Igni-
tion Engine Operation and Design for Minimum Exhaust
Emission, " SAEProgress in Technology, 12, Vehicle
Emissions, Part II, p. 71 (1966) SAE Transactions, 75
(1967).
4-36. R. D. Kopa, "Control of Automotive Exhaust Emission by
Modofications of the Carburetion System, " SAE paper
No. 660114, SAE Automotive Engineering Congress,
Detroit (January 1966); SAE Progress in Technology, 12,
Vehicle Emissions, Part II, 212(1966).
4-37. J. A. Robison and W. M. Brehob, "The Influence of Improved
Mixture Quality on Engine Exhaust Emissions and Perform-
ance, " paper presented at Western States Combustion
Institute Meeting, Santa Barbara (October 1965).
4-38. C. F. Taylor, The Internal Combustion Engine in Theory
and Practice, Volumes I and II, MIT Press (Cambridge,
1968).
4-39. W.A. Daniel and J. T. Wentworth, "Exhaust Gas Hydro-
carbons — Genesis and Exodus, " SAE paper No. 486B,
SAE National Automobile Week, SAE Technical Progress
Series, 6, Vehicle Emissions, Detroit, Michigan,
March 1962, pp 192-196.
4-40. J. T. Wentworth, "Effect of Combustion Chamber Surface
Temperature on Exhaust Hydrocarbon Concentration, "
SAE paper No. 710587, SAE Meeting, Montreal, Canada
(June 1971).
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
Emissions and Power," SAE paper No. 710832, St. Louis,
Missouri (October 1971).
4-43. R. W. Aiman, "Engine Speed and Load Effects on Charge
Dilution and Nitric Oxide Emission, " SAE paper No. 720256,
Detroit, Michigan (January 1972).
4-44. D. F. Hagen, and G. W. Holiday, "The Effect of Engine
Operating and Design Variables on Exhaust Emissions, "
SAE paper No. 486C, SAE National Automobile Week,
Detroit, Michigan (March 1962); SAE Technical Progress
Series, 6, Vehicle Emissions, p. 206 (1964).
4-45. W. W. Haskell, and C. E. Legate, "Exhaust Hydrocarbon
Emissions from Gasoline Engines — Surface Phenomena, "
SAE paper No. 720255, Detroit, Michigan (January 1972).
4-46. R. M. Siewert, "How Individual Valve Timing Events Affect
Exhaust Emissions, " SAE paper No. 710609, SAE Meeting,
Montreal, Canada (June 1971).
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
Control of Oxides of Nitrogen (NOX), " SAE paper
No. 720121, SAE Congress, Detroit, Michigan (January 1972).
4-49. G. B. Meacham Kirby, "Variable Cam Timing as an Emission
Control Tool, " SAE paper No. 700673, SAE Meeting, Los
Angeles, California (August 1970).
4-50. C. E. Scheffler, "Combustion Chamber Surface Area, A Key
To Exhaust Hydrocarbons, " SAE paper No. 660111, SAE
Progress in Technology, 12, Vehicle Emissions, Part II,
p. 60 (1966).
4-51. L. Eltinge, F. J. Marsee, and A. J. Warren, "The Potenti-
alities of Further Emissions Reduction by Engine Modifica-
tions, " SAE paper No, 680123, SAE Congress, Detroit,
Michigan (January 1968).
4-159
-------
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
-------
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
-------
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
-------
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
-------
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
-------
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
-------
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
-------
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
-------
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
-------
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
-------
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
-------
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
-------
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
-------
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
-------
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
-------
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
-------
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
-------
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
-------
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
-------
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
-------
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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