Open Collections

UBC Theses and Dissertations

UBC Theses Logo

UBC Theses and Dissertations

Pollutant formation in a gaseous-fuelled, direct injection engine McTaggart-Cowan, Gordon Patrick 2006

Your browser doesn't seem to have a PDF viewer, please download the PDF to view this item.

Item Metadata

Download

Media
831-ubc_2006-200469.pdf [ 18.2MB ]
Metadata
JSON: 831-1.0080746.json
JSON-LD: 831-1.0080746-ld.json
RDF/XML (Pretty): 831-1.0080746-rdf.xml
RDF/JSON: 831-1.0080746-rdf.json
Turtle: 831-1.0080746-turtle.txt
N-Triples: 831-1.0080746-rdf-ntriples.txt
Original Record: 831-1.0080746-source.json
Full Text
831-1.0080746-fulltext.txt
Citation
831-1.0080746.ris

Full Text

POLLUTANT FORMATION IN A GASEOUS-FUELLED, DIRECT INJECTION ENGINE by GORDON PATRICK M c T A G G A R T - C O W A N B.Eng., University of Victoria 1999 M . A . S c , University of British Columbia 2001 A THESIS SUBMITTED IN PARTIAL FULFILLMENT OF THE REQUIREMENTS FOR THE DEGREE OF DOCTOR OF PHILOSOPHY in THE F A C U L T Y OF G R A D U A T E STUDIES (Mechanical Engineering) THE UNIVERSITY OF BRITISH C O L U M B I A October 2006 © Gordon Patrick McTaggart-Cowan, 2006 Abstract Heavy-duty natural gas engines offer air pollution and energy diversity benefits. However, current homogeneous-charge lean-burn engines suffer from impaired efficiency and high unburned fuel emissions. Direct injection offers the potential of diesel-like efficiencies, but requires further research. To improve understanding of the combustion process and pollutant formation mechanisms in a pilot-ignited, direct injection of natural gas engine with intake charge dilution, the effects of enhanced gaseous jet kinetic energy, gaseous fuel composition (including ethane, propane, hydrogen, and nitrogen), and filtering the recirculated gases were studied. An experimental investigation was carried out on a single-cylinder heavy-duty engine. Fuel consumption, in-cylinder performance and gaseous and particulate emissions (total mass, size distributions, and black carbon content) were measured. The results indicated that increasing the jet kinetic energy significantly reduced particulate matter (PM) emissions due to improved fuel-air mixing, especially at high load. The addition of hydrogen to the fuel reduced emissions of carbon monoxide (CO), unburned fuel (HC) and P M . The largest effects were observed at high load conditions. The addition of ethane and propane to the fuel resulted in increases in P M and CO emissions at all operating conditions tested; no effect on the combustion progression was detected. The addition of nitrogen to the fuel significantly reduced emissions of CO, P M , and HC due to enhancement of the late-cycle combustion event from increased in-cylinder turbulence. Removing P M from the recirculated gases revealed that these particles had no significant effect on the combustion event or on P M emissions. In conclusion, mixing and kinetic enhancement both reduced the gaseous fuel ignition delay. The overall combustion event was, at high load, mixing limited; the combustion rate was unaffected by fuel reactivity but was increased with turbulence enhancement. Emissions formation was found to be a result of multiple influences whose relative importance varied with operating condition. Increased mixing and lower fuel carbon content reduced P M emissions. Reductions in emissions through the addition of hydrogen and nitrogen to the fuel may offer a potential technique to offset increases in emissions due to variations in ethane and propane levels in natural gas. Table of Contents Abstract ii Table of Contents iii List of Tables vii List of Figures viii Nomenclature xiii Acknowledgements . xv Co-Authorship Statement xvi Chapter I - Introduction 1 1.1 Air Pollution 1 1.1.1 Particulate Matter '.. 2 1.1.2 Gaseous Pollutants 3 1.2 Diesel Emissions Regulations 4 1.2.1 Exhaust Aftertreatment 5 1.3 Natural Gas Fuelling 6 1.4 Objectives and Scope 7 1.5 Thesis Structure 8 1.6 Tables and Figures 9 Chapter 2 - Background Information 10 2.1 Gaseous Fuelling of Heavy-Duty Engines 10 2.1.1 Premixed technologies 11 2.1.2 Direct Injection Technologies 16 2.2 Non-Premixed Gaseous Combustion 18 2.2.1 Combustion Structure 19 2.2.2 Oxides of Nitrogen 21 2.2.3 Carbon Monoxide 22 2.2.4 Hydrocarbons 22 2.2.5 Particulate Matter 23 2.2.6 PM Population Dynamics 28 2.2.7 PM Formation in Pilot Ignited Natural Gas 28 2.3 Pilot Ignited, Direct Injected Natural Gas 29 2.3.1 Exhaust Gas Recirculation 29 2.3.2 Injection Process 31 2.3.3 Diesel Pilot Influence 33 2.4 Summary / Literature Gap 34 2.5 Tables and Figures 35 -iii-Chapter 3 - Apparatus and Procedures 37 3.1 Research Engine 37 3.1.1 Air Exchange System 38 3.1.2 Fuelling System 39 3.1.3 Instrumentation and Data Acquisition 40 3.2 Emissions Measurements 42 3.2.1 Gaseous Emissions 42 3.2.2 Particulate Measurement 43 3.3 Experimental Parameters 49 3.3.1 In-Cylinder Conditions 50 3.3.2 Fuel'Oxidizer Ratio 52 3.3.3 Data Presentation Parameters 53 3.4 Repeatability and Uncertainty Analysis 54 3.5 General Methodology 55 3.6 Tables and Figures 56 Chapter 4 - Injection Pressure 61 4.1 Introduction 61 4.2 Injection Pressure Influences 61 4.3 Experimental Methodology 64 4.3.1 Experimental Conditions 65 4.4 Results 66 4.4.1 Effect of Load 66 4.4.2 Effect of Speed. 67 4.4.3 Effects of Operating Condition 68 4.4.4 Injection Pressure Details 72 4.4.5 In-Cylinder Performance 73 4.4.6 Effects on Particle Size Distributions 75 4.5 Discussion 76 4.6 Conclusions 78 4.7 Tables and Figures 79 Chapter 5 - Hydrogen/Methane Blends 87 5.1 Introduction 87 5.1.1 Hydrogen/Methane Blend Combustion 88 5.1.2 Non-F'remixedHydrogen/Methane Flames . . . 8 9 5.2 Experimental Methodology 90 5.2.1 Experimental Conditions 90 5.2.2 Fuel Blends 91 5.2.3 Replications and Randomization 91 -iv-5.3 Results 92 5.3.1 Emissions 92 5.3.2 Particulate Matter 94 5.3.3 Combustion Analysis 95 5.3.4 Greenhouse Gas Emissions 98 5.4 Discussion 99 5.4.1 Combustion Implications 99 5.4.2 Emissions and Applications 100 5.5 Conclusions 102 5.6 Tables and Figures 104 Chapter 6 - Fuel Composition 112 6.1 Introduction 112 6.2 Previous W o r k 112 6.3 Experimental Information 114 6.3.1 Operating Conditions 115 6.3.2 Statistical Process 116 6.4 Results 116 6.4.1 Combustion Effects 116 6.4.2 Gaseous Emissions 118 6.4.3 Particulate Matter 120 6.5 Discussion 122 6.6 Conclusions 123 6.7 Tables and Figures 124 Chapter 7 - Fuel Dilution with Nitrogen 135 7.1 Introduction 135 7.2 Previous W o r k 135 7.3 Experimental Information 137 7.3.1 Operating Condition 137 7.4 Effects of Fuel Dilution 138 7.4.1 Combustion Effects 138 7.4.2 Gaseous Emissions 140 7.4.3 Particulate Emissions 142 7.5 Discussion 143 7.5.1 Fuel Systems Issues 144 7.5.2 Fuel Composition Parameters 144 7.6 Conclusions 148 7.7 Tables and Figures 150 -v -Chapter 8 - The Effects of Reingested Particles 159 8.1 Introduction 159 8.2 Experimental Methodology 160 8.2.1 Experimental Conditions 161 8.3 Results 161 8.3.1 Filtration Effectiveness 162 8.3.2 In-Cylinder Performance 162 8.3.3 Emissions and Performance 163 8.3.4 Particulate Matter 164 8.4 Discussion 165 8.5 Conclusions 166 8.6 Tables and Figures 167 Chapter 9 - Significant Findings and Recommendations 172 9.1 Significant New Findings 172 9.2 Systems Implications 181 9.3 Study Limitations 183 9.4 Future Work 184 References 186 Appendices . 203 A . l Effects of Dilution Conditions on PM Measurements 203 A.2 Aethalometer Correction Procedure 214 A.3 Instrumentation List 221 A.4 Experimental Uncertainty Calculations 224 A.5 Natural Gas Composition Variability 226 A.6 In-Cylinder Pressure Traces 229 A .7 T E M Images of Particle Samples from Selected Tests 240 -vi-List of Tables Table 1.1 Heavy-duty engine emissions standards 9 Table 2.1 Selected results from in use natural-gas fuelled heavy-duty engine studies... 35 Table 3.1 Single cylinder engine specifications 56 Table 3.2 Single cylinder engine control and output parameters ; 56 Table 3.3 Calculated uncertainty and repeatability analysis for key parameters 57 Table 4.1 Engine test conditions for injection pressure study 79 Table 4.2 Injection pressure A N O V A results as a function of load 79 Table 4.3 Injection pressure A N O V A results as a function of speed 80 Table 4.4 Summary of results from operating condition tests 80 Table 5.1 Engine operating mode for hydrogen/methane blend testing 104 Table 5.2 Gas composition for hydrogen/methane blend testing 104 Table 6.1 Fuel composition for heavy hydrocarbon study 124 Table 6.2 Engine operating conditions for heavy hydrocarbon study 125 Table 7.1 Fuel composition for nitrogen dilution study 150 Table 7.2 Engine operating conditions for nitrogen dilution study 150 Table 7.3 Representative parameters for all fuel composition results 151 Table 8.1 Base engine operating condition and test modes for filtered EGR work 167 Table 8.2 Means of various parameters from filtered EGR tests 167 Table A l .1 Experimental conditions for dilution ratio study 209 Table A2.1 /-values for all data points for Aethalometer validation 218 Table A3.1 List of instrumentation 221 Table A5.1 Fuel composition from reported engine research 227 -vii-List of Figures Figure 1.1 US EPA Heavy-Duty Engine Emission Standards, 1988-2010 9 Figure 2.1 Effect of intake oxygen dilution on peak combustion temperature and N O x emissions 35 Figure 2.2 Effect of delay between natural gas and diesel injection on P M - N O x and HC-NO x trade-offs 36 Figure 2.3 Combustion timing and EGR fraction effects on P M - N O x and fuel consumption-NOx trade-offs 36 Figure 3.1 Air exchange system layout 57 Figure 3.2 Westport HPDI™ injector schematic 57 Figure 3.3 Westport HPDI™ injector injection process 58 Figure 3.4 P M sampling and dilution system schematic 58 Figure 3.5 Correlation between T E O M and gravimetric filter results 58 Figure 3.6 Assumed effective density of SMPS particles as a function of mobility diameter, for three different fractal dimensions 59 Figure 3.7 Mass concentration comparison between T E O M measurements and calculated ultrafine particle mass based on SMPS measurements 59 Figure 3.8 Typical net and-integrated heat-release rate plots, showing the start-of-injection and combustion timings for the pilot and gaseous fuels 60 Figure 4.1 Effect of injection pressure and operating condition on GISFC 80 Figure 4.2 Effect of injection pressure and operating condition on N O x emissions 81 Figure 4.3 Effect of injection pressure and operating condition on CO emissions 81 Figure 4.4 Effect of injection pressure and operating condition on HC emissions 81 Figure 4.5 Effect of injection pressure and operating condition on P M emissions 82 Figure 4.6 Emissions and GISFC variations with injection pressure, at near peak-torque 82 Figure 4.7 Effect of injection pressure and operating condition on gaseous fuel ignition delay 83 Figure 4.8 Effect of injection pressure and operating condition on combustion duration 83 Figure 4.9 In-cylinder pressure and net heat-release rate, at low-load, low-speed 84 Figure 4.10 In-cylinder pressure and net heat-release rate, at high-load, mid-speed 84 Figure 4.11 Particle size distributions at low-load, low-speed 85 Figure 4.12 Particle size distributions at high-load, low-speed 85 -viii-Figure 4.13 Particle size distributions at high-load, high-speed 86 Figure 5.1 Emissions for natural gas fuelling, 10% and 23% hydrogen 105 Figure 5.2 Emissions for natural gas, pure methane, 15% and 35% hydrogen 105 Figure 5.3 BC concentration for natural gas fuelling, 10% and 23% hydrogen 106 Figure 5.4 Particle size distribution for natural gas fuelling, 10% and 23% hydrogen.. 106 Figure 5.5 BC concentration for natural gas, pure methane, 15% and 35% hydrogen .. 107 Figure 5.6 Particle size distribution for natural gas, pure methane, 15% and 35% hydrogen 107 Figure 5.7 Combustion performance comparison for natural gas, 10% and 23% hydrogen 108 Figure 5.8 Comparison of pilot and gaseous fuel ignition delay times for natural gas, 10% and 23% hydrogen 108 Figure 5.9 Pressure trace and estimated heat release rates, comparisons for natural gas, 10% and 23% hydrogen 109 Figure 5.10 Combustion performance comparison for natural gas, methane, 15% and 35% hydrogen 109 Figure 5.11 Net heat release rates for natural gas, methane, 15% and 35% hydrogen 110 Figure 5.12 Pilot and gaseous fuel ignition delay comparison for natural gas, methane, 15% and 35% hydrogen 110 Figure 5.13 CO2 and net GHG emissions for natural gas, 10% and 23% hydrogen 110 Figure 5.14 CO2 and net GHG emissions for natural gas, methane, 15% and 35% hydrogen I l l Figure 6.1 Pilot and gas ignition delay times for ethane and propane additives 125 Figure 6.2 Peak heat-release rate for ethane and propane additives 126 Figure 6.3 Burn duration and end-of-combustion timing for ethane and propane additives 126 Figure 6.4 In-cylinder pressure trace and net heat-release rate for the mid-timing condition for ethane and propane additives 127 Figure 6.5 Combustion stability for ethane and propane additives 127 Figure 6.6 GISFC for ethane and propane additives 128 Figure 6.7 N O x emissions for ethane and propane additives 128 Figure 6.8 Hydrocarbon emissions for ethane and propane additives 128 Figure 6.9 CO emissions for ethane and propane additives 129 Figure 6.10 P M emissions for ethane and propane additives 129 Figure 6.11 Particle size distributions for ethane addition 130 -ix-Figure 6.12 Particle size distributions for propane addition 131 Figure 6.13 BC concentration for ethane and propane additives 131 Figure 6.14 BC fraction for ethane and propane additives 132 Figure 6.15 Volatile mass rate for ethane and propane additives 132 Figure 6.16 Effect of H:C ratio on total change in P M emissions 133 Figure 6.17 Effect of H:C ratio on change in black carbon mass emissions 134 Figure 7.1 Commanded injection timings for nitrogen diluted fuelling 151 Figure 7.2 Pilot and gaseous fuel ignition delay times for nitrogen diluted fuelling 151 Figure 7.3 Peak heat-release rate and combustion duration for nitrogen diluted fuelling 152 Figure 7.4 In-cylinder pressure and net heat-release rate for early and late combustion timings with nitrogen diluted fuelling 152 Figure 7.5 COV of GIMEP and peak cylinder pressure, as well as GISFC, for nitrogen diluted fuelling 153 Figure 7.6 Emissions (CO, N O x , HC, PM) for nitrogen diluted fuelling 153 Figure 7.7 Percentage of total heat released after the end of injection, for natural gas and nitrogen diluted fuelling 154 Figure 7.8 Particle number size distributions, for nitrogen diluted fuelling 154 Figure 7.9 Particle mobility volume, for nitrogen diluted fuelling 155 Figure 7.10 BC mass and fraction, for nitrogen diluted fuelling 155 Figure 7.11 Estimated compression work and storage volume for various fuel blends... 155 Figure 7.12 Combustion duration and peak heat-release rate, for all gas blends as functions of Wobbe Index 156 Figure 7.13 Peak heat-release rate relative to natural gas for all timings and all fuel blends 156 Figure 7.14 Influence of changes in gas ignition delay time (GID) on peak heat-release rate and hydrocarbon emissions from various fuel blends 156 Figure 7.15 Influence of combustion variability on HC and CO emissions for various fuel blends 157 Figure 7.16 Comparison of HC and CO emissions for all fuel blends as a function of £ 157 Figure 7.17 Comparison of P M emissions (total mass, black-carbon mass, ultrafme particle volume) for all fuel blends as a function of £ 157 Figure 7.18 Comparison of P M emissions (total mass, black carbon mass, mobility volume) for methane and heavy hydrocarbon addition as a function of 158 Figure 8.1 Engine air exchange system with EGR filter 167 Figure 8.2 P M sampling system modified for intake sampling 168 -x-Figure 8.3 Intake P M size distributions, modes A and D 168 Figure 8.4 Intake BC concentrations and % reduction in black carbon between filtered and unfiltered EGR at all EGR test modes 168 Figure 8.5 Pressure trace and heat-release rate at mode D for filtered and unfiltered cases 169 Figure 8.6 Ignition delay and combustion progression for filtered and unfiltered conditions 169 Figure 8.7 Operating condition variations as a function of test mode for filtered and unfiltered conditions 170 Figure 8.8 Power-specific emissions as a function of test mode for filtered and unfiltered conditions 170 Figure 8.9 Particle size distributions in intake and exhaust streams at modes A and D 170 Figure 8.10 BC concentration, BC fraction and mobility volume at all test modes 171 Figure 9.1 Summary of fuel composition effects on emissions 173 Figure 9.2 Summary of nitrogen and hydrogen effects on P M 175 Figure 9.3 Summary of minimum P M emissions levels 176 Figure 9.4 Summary of effects of various parameters on gaseous ignition delay 177 Figure 9.5 Summary of effects of various parameters on combustion duration 178 Figure 9.6 Summary of effects of various parameters on GISFC 180 Figure A l . l The effect of primary dilution ratio on particulate mass emissions at low-speed 209 Figure A1.2 The effect of primary dilution ratio on particulate size distributions at low-speed 210 Figure A1.3 The effect of primary dilution ratio on particulate mass emissions at mid-speed 210 Figure A 1.4 The effect of primary dilution ratio on black carbon concentration at mid-speed 211 Figure A l .5 The effect of primary dilution ratio on particulate size distributions at mid-speed 211 Figure A l .6 The effect of secondary dilution ratio on black carbon concentration 212 Figure A l .7 The effect of secondary dilution ratio on particulate size distributions 212 Figure A2.1 Effect of attenuation coefficient (craln) on reported black carbon concentrations 218 Figure A2.2 Absorption curve for various representative test conditions 219 Figure A6.1 Pressure traces and heat-release rates from injection pressure study at low-speed, low-load 229 -xi-Figure A6.2 Pressure traces and heat-release rates from injection pressure study at low-speed, mid-load 230 Figure A6.3 Pressure traces and heat-release rates from injection pressure study at low-speed, high-load 230 Figure A6.4 Pressure traces and heat-release rates from injection pressure study at mid-speed, mid-load 231 Figure A6.5 Pressure traces and heat-release rates from injection pressure study at mid-speed, high-load 231 Figure A6.6 Pressure traces and heat-release rates from injection pressure study at high-speed, high-load . 232 Figure A6.7 Pressure traces and heat-release rates from detail injection pressure study.. 233 Figure A6.8 Pressure traces and heat-release rates for hydrogen addition testing at low-speed 234 Figure A6.9 Pressure traces and heat-release rates for hydrogen addition testing at mid-speed and high-load 235 Figure A6.10 Pressure traces and heat-release rates for ethane addition testing 236 Figure A6.11 Pressure traces and heat-release rates for propane addition testing 237 Figure A6.12 Pressure traces and heat-release rates for nitrogen addition testing 238 Figure A6.13 Pressure traces and heat-release rates with and without EGR filter.... 239 Figure A7.1 T E M images of particles sampled from exhaust stream on natural gas fuel and high ethane additive 240 Figure A7.2 T E M images of particles sampled from exhaust stream with high propane additive and high nitrogen additive 241 Figure A7.3 T E M images of particles sampled from exhaust stream with and without filter in EGR line 241 Figure A7.4 T E M images of particles sampled from intake stream with and without filter in EGR line 242 -xii-Nomenclature 50%IHR Mid-point of integrated heat-release rate °CA Crank angle degree % H 2 Volume percent hydrogen in fuel % N 2 Volume percent nitrogen in fuel A N O V A Analysis of variance A T D C After top-dead-center BC Black carbon BD Burn duration COV Coefficient of variation (standard deviation / mean) D M A Differential mobility analyzer D p Particle diameter DF Degrees-of-Freedom EGR Exhaust gas recirculation EOC End of combustion EPA United States environmental protection agency G H G Greenhouse gas GID Gas ignition delay time GikWhr Gross indicated kilowatt hour GIMEP Gross indicated mean effective pressure GISFC Gross indicated specific fuel consumption GSOC Gas start-of-combustion GSOI Gas start-of-injection H:C Molar hydrogen-to-carbon ratio H A C A Hydrogen abstraction, acetylene addition HC Hyrdrocarbons (total) HCCI Homogeneous charge compression ignition H H V Higher heating value HPDI™ High-pressure direct injection (trademarked by Westport Innovations Inc.) HRR Heat-release rate IHR Integrated heat-release rate L H V Lower heating value L N G Liquefied natural gas M W Molecular weight NDIR Non-dispersive infra-red N G Natural gas nmHC Non-methane hydrocarbons N O x Oxides of nitrogen (include NO, NO2) P A H Polycyclic aromatic hydrocarbons PID Pilot ignition delay time PIDING Pilot-ignited, direct-injection natural gas Pinj — Gaseous fuel injection pressure P M Particulate matter -xii i-P M 2 . 5 Particulate matter less than 2.5 urn in diameter Pmax — Maximum in-cylinder pressure PSOC Pilot start-of-combustion PSOI Pilot start-of-injection RIT Relative injection timing (delay between pilot and gas injections) R P M Revolutions per minute SMPS Scanning mobility particle sizer SOC Start of combustion Tadiabatic ~~ Adiabatic flame temperature TDC Top-dead-center T E O M Tapered element oscillating microbalance V O C Volatile organic carbon compounds Yjnt02 = Intake oxygen mass fraction Equivalence ratio (on oxygen basis) 5 New index parameter (HHV/H:C ratio) -xiv-Acknowledgements The work described in this thesis could not have been completed without the support of the exemplary colleagues with whom I was fortunate to work. First, I would like to thank my supervisors, Dr. Kendal Bushe and Dr. Steve Rogak, for their unending support, patience, and advice during my time at UBC. The development of both my technical proficiency and my abilities as an independent researcher are a direct result of their guidance and tutelage. I also want to extend special recognition to Dr. Martin Davy, for his support and direction, both during my research program and in my future career direction. And a special thank-you to Dr. Phillip Hi l l , whose unwavering encouragement helped me consistently achieve more than I would have thought possible. Without the research staff at both UBC and Westport Innovations, I would not have had an experimental facility to use for my research. At UBC, the ceaseless efforts of Heather Jones and Bob Parry ensured that the research engine was available and operational for me when I needed it. Support from Westport was always quick and comprehensive; however, I want to especially acknowledge the assistance of Brian Buik, Mark Dunn, Dale Goudie, Guowei L i , Sandeep Munshi, and Patric Ouellette. I also want to acknowledge all the excellent graduate students with whom I have been honoured to work; but in particular, thanks to Jin Bei, Jian Huang, Wu Ning, Conor Reynolds, and Maggie Wang. They always provided me with an ear to listen, and never failed to encourage me when I encountered a pitfall along the way. Finally I want to thank my family, and especially my parents, for their ongoing support of my educational pursuits. Lastly, and most importantly, to my wife Helen, for your unconditional love and support over the past two years. I hope that I have been able to provide you with even half of the support that you have always given me. This work is dedicated to you; without you, I would never have been able to complete this project. -xv-Co-Authorship Statement The work presented in this thesis was conceived, conducted, and disseminated by the doctoral candidate. The co-authors of the manuscripts that comprise part of this thesis made contributions only as is commensurate with a thesis committee or as experts in a specific area as it pertains to the work. The co-authors provided direction and support. The co-authors reviewed each manuscript prior to submission for publication and offered critical evaluations; however, the candidate was responsible for the writing and the final content of these manuscripts. -xvi-Chapter 1 Introduction For over a century, heavy-duty internal combustion engines have been a key component to trade and economic activity throughout the world. These engines, fuelled with liquid diesel and running on cycles approaching the thermodynamic 'diesel' cycle, have facilitated cost-effective transportation of people and goods by road, sea, and rail. They have also powered much of the construction of our modern infrastructure and have provided reliable stationary power generation in remote and emergency situations. The principal attractions of diesel engines are their high efficiency, advantageous torque characteristics, and enviable reliability. However, they emit harmful pollutants and use an increasingly scarce natural resource; as such, there is a pressing need to develop improved combustion systems which simultaneously reduce emissions and improve efficiency while reducing the burden on liquid fossil fuels through the use of alternative fuels. Traditional diesel engines offer many advantages in heavy-duty applications, but also suffer from relatively high levels of regulated and unregulated emissions. Diesel engines are reliable and robust, provide high torque at low speeds, and are as much as 25% more efficient than equivalent gasoline-fuelled engines [1]. As a result of their high efficiency, greenhouse gas emissions are low compared with other in-use transportation motive power sources. However, emissions of harmful species, including 'criteria' pollutants such as fine particulate matter (PM) and oxides of nitrogen (NO x), as well as air toxics such as benzene, are significantly higher. In Canada in 2000, diesel engines emitted 43% of all anthropogenic N O x emissions [2]. Although the contribution of diesel engines to regulated P M mass emissions is relatively small, they pose a significant health concern because of their composition, small size, and high concentrations in urban areas. 1.1 A i r P o l l u t i o n Air pollution has long been recognized as a public nuisance. In the early 20 t h century, increasing uncontrolled emissions led to acute air pollution events, culminating in London's 'killer fogs' of the early 1950's [3]. As a direct result of these high-mortality incidences, clean-air act legislation was enacted in many countries. In general, these actions were successful in preventing acute events. However, epidemiological studies have subsequently shown that increased mortality results from exposures to levels of pollutants far below those first legislated in the clean-air acts [4]. Air pollutants of concern include solid and liquid aerosols, classic gaseous pollutants (such as N O x , sulphur oxides, carbon monoxide and ozone), and carcinogenic 'air toxics' (approximately 200 separate species which are thought to be toxic). In Ontario a recent report estimated that almost 6000 excess deaths per year could be attributed to air pollution [5]. 1.1.1 Particulate Matter Particulate matter is the term used to refer to solid and liquid aerosol particles suspended in the atmosphere. P M is composed of a wide range of natural and anthropogenic species. It can be emitted directly from a source or it can be formed by reactions between gases in the atmosphere. The sizes of particles vary, with typical definitions based on an aerodynamic or effective diameter. Widely-accepted categories include ultrafine (sub 100 nm) and fine (sub 2.5 jam, PM2.5) particles. Coarse mode particles (diameter > 2.5 (am) tend to settle out of the atmosphere in hours or days, while ultrafine particles tend to agglomerate into larger particles (diameters between 0.1 and 1.0 um) in minutes to hours after emission. These agglomerated particles have the longest residence time, as they do not tend to agglomerate further but also have low settling velocities [6]. Combustion processes, such as in diesel engines, tend to generate smaller particles (<1.0um) either directly in the flame or by post-exhaust condensation of volatile species. Larger particles (>1.0um) are typically formed from mechanical processes, including road dust, soil tilling, and natural sources. In Canada in 2000, 90% of PM10 (PM < lOum in diameter) and 60% of P M 2 5 emissions on a mass basis originated from such mechanical (non-combustion) sources [2]. Numerous epidemiological studies have shown that P M significantly increases premature mortality through cardiovascular or pulmonary endpoints [3,4,7,8,9]; other health effects include impaired lung development [10]. PM's adverse health effects are attributed to deposition on the alveoli and other sensitive lung tissues, causing irritation and immune system reactions [11]. Peak deposition of particles in the alveolar region of the lungs occurs for particles with diameters between 0.01 and 0.1 urn; particles larger than 2.5 um are unlikely to penetrate into the lungs [12,13]. For an equal mass, smaller particles (with larger surface areas) have a greater inflammatory effect and cause greater epithelial damage -2-[12,14,15]. Smaller particles are also more likely to pass into the bloodstream, thereby directly affecting the cardiovascular system [16]. There is some evidence that observed discrepancies in mass-based P M health effects studies may be at least partially explained by variations in total particle surface area [17]. The composition of the particles can also have a significant impact on health outcomes. Diesel P M includes a significant quantity of adsorbed volatile organics and other potentially toxic or carcinogenic species [18]. Occupational and animal toxicological studies have shown that exposures to diesel P M can have very significant health impacts [19]. Whether these effects are a result of either the chemical composition of the particles or the high concentrations of ultrafine particles in diesel exhaust is unclear. The California Air Resources Board (CARB) has defined diesel P M as an 'air toxic', and estimates that in 2003 diesel P M was responsible for approximately 75% of the cancer cases attributed to all air toxics [20]*. 1.1.2 Gaseous Pollutants Of the gaseous air pollutants, the most significant relating to diesel engines is N O x . The direct health impacts of exposures to moderate levels of N O x (specifically nitrogen dioxide, NO2) include eye and respiratory tract irritation [6]. Indirect environmental and health impacts result from NO x ' s role as an important precursor for the secondary formation of nitrate aerosols and in the formation of nitric acid, which leads to acid rain [6]. However, the primary environmental impacts of N O x are generally considered to be the formation of ground level ozone (O3) and photochemical smog. The formation of ground-level ozone is dependant on reactions between N O x and volatile organics (VOCs). VOCs are emitted from combustion processes (primarily light-duty vehicles) but also from natural and other industrial (non-combustion) sources. The local geographic and meteorological conditions and the relative concentrations of N O x and VOCs control the rate and quantity of ozone generation; sometimes the process is N O x limited and other times it is V O C limited [21,22]. As a result, reductions in N O x may or may not have a substantial impact on ambient O3 levels, and in fact reductions in N O x may somewhat increase ambient O3 under certain conditions [23]. Despite this, regulatory agencies around the world are aggressively pursuing C A R B recently added environmental tobacco smoke to the air toxic list; due to its prevalence, it poses a greater cumulative health risk than diesel P M , and hence in more recent analyses the relative contribution of diesel P M is less substantial. -3-the reduction of N O x emissions. While other ambient air pollutant emissions are also important - for example carbon monoxide (CO), sulfates, and lead - current diesel engines are relatively insignificant emitters of these species. Another class of gaseous pollutants is the group of gases, commonly referred to as greenhouse gases (GHG's), which have radiative absorption properties in the infrared range. GHG's affect the global climate by absorbing some of the earth's emitted infrared radiation and re-radiating it back towards earth. Their presence influences the earth's radiative energy balance and thus, as their concentrations increase to unprecedented levels, they are altering the earth's climate [6]. Carbon dioxide (CO2) and methane (CH4) are the two most significant GHG's whose ambient atmospheric concentrations have been significantly increased by human influences [6,24]. While the specific long-term effects of changes in G H G levels on the global atmospheric-ocean system are uncertain, comparisons with past climatic records indicate that greater climatic instability and higher global average temperatures correlate with higher G H G levels [6]. While the high efficiency of diesel engines, relative to most other motive power sources, minimizes GHG emissions [1], emission levels are still a concern. 1.2 Diesel Emissions Regulations Heavy-duty engine emission-control legislation is designed to minimize the adverse health and environmental impacts of air pollutants from these engines. In most jurisdictions, including the United States and Canada (Canadian emissions regulations are typically aligned with those for the US), the reduction of emissions from on-road vehicles and engines is achieved by establishing emission limits for those substances [25,26]. For an engine model to "pass" its emissions test, it must meet the legislated levels for each pollutant over a standardized test-cycle. Engine emissions standards have become progressively more stringent over time. The United States Environmental Protection Agency (EPA) emissions standards for N O x and P M for heavy-duty vehicles are shown in Figure 1.1 [25]. Dramatic reductions in permitted emission levels have resulted in ever cleaner new vehicles; however, due to the relatively long lifetime of heavy-duty vehicles, many vehicles still on the road are emitting at much higher levels. Of particular note is the order-of-magnitude reduction in P M between the current (2004) and the 2007-10* standards. A sample of current or upcoming * For 2007 half of a manufacturers new heavy-duty vehicles must meet the 2010 N O x levels, while the rest may emit at the 2004 standard level. Or all vehicles can operate at half of the 2004 standard level until 2010 [25]. -4-emissions standards for various jurisdictions is shown in Table 1.1 [27]. Care must be taken when comparing emissions regulations between jurisdictions because the test-cycles used are not always equivalent (for example, they have different ratios of transient to steady-state operation). Furthermore, data published by engine researchers or developers are frequently based on non-standard test-cycles. 1.2.1 Exhaust After treatment To meet the upcoming emissions standards, substantial reductions in emission levels from heavy-duty vehicles will be required. One of the techniques which will be critical to meeting these standards is exhaust gas aftertreatment, where harmful pollutants are removed from the exhaust stream downstream of the engine's exhaust port. In light-duty automotive applications, three-way catalytic converters have been shown to effectively reduce N O x , CO and hydrocarbon (HC) emissions by more than 95% over a wide range of engine operating conditions [28]. These units depend on a very low concentration of oxygen in the exhaust stream to function properly. If there are significant quantities of oxygen in the exhaust (such as from typical diesel engines), the catalytic converter will function as an oxidative catalyst, oxidizing HC and CO emissions, but not reducing N O x . Advanced strategies to reduce N O x emissions are under development for diesel engines, including lean-NOx traps and urea-based selective catalytic reduction [29]. A problem in applying conventional oxidative catalytic converters to natural gas fuelled engines is that CH4 is not significantly oxidized at normal exhaust temperatures [30,31]. However, special palladium-based catalysts are under development which are capable of achieving CH4 oxidation rates on the order of 90% in the temperature range typically found in heavy duty engine exhaust [32,33]. To reduce P M emissions, manufacturers are pursuing techniques that use filters in the exhaust stream to capture the particles. The filters need to be regenerated (heated up so that the collected carbon will be oxidized to CO2); this is normally achieved by increasing the engine exhaust temperature or by combusting a small amount of fuel in the post-exhaust stream. Continuously regenerating traps, such as those under development by Johnson-Matthey Inc., offer the potential to oxidize the collected P M at a much lower temperature [34]. While effective, these aftertreatment devices impose a restriction on the exhaust stream, increasing the amount of work that the engine must do to expel the burned gases. However, optimizing the combustion and then using aftertreatment to remove the harmful pollutants -5-could improve overall system efficiency while achieving the emissions standards. Notwithstanding the above possibilities, aftertreatment system complexity and capital cost are significant issues that need to be addressed., 1.3 Natural Gas Fuelling One alternative for fuelling heavy-duty engines, which offers both air pollution and energy diversity benefits, is the use of natural gas. Natural gas, like diesel, is a fossil-based hydrocarbon. Although estimates of the worldwide reserves of conventionally recoverable natural gas vary widely, there appear to be slightly greater energy reserves in readily extractable natural gas than there are in oil [35]. Distribution of the natural gas is a concern, with liquefied natural gas (LNG) being regarded as one alternative; for example, L N G imports into the USA have more than doubled over the past five years [36]. Other alternatives, such as gas-to-liquid (Fischer-Tropsch) conversion to create synthetic liquid fuels, require costly physical plants which are highly energy-intensive, even compared to L N G installations. Despite these concerns, by providing an alternative energy source, natural gas offers an opportunity for a more flexible and diverse energy system. Furthermore, so long as the growth in demand for liquid fuels continues, there is significant potential for economic incentives for natural gas use in transportation. The main driving force for the consideration of natural gas as an alternative fuel in transportation has been reduced air pollutant emissions due to its relatively clean combustion. Natural gas burns at a lower temperature than most hydrocarbons, resulting in lower N O x emissions. It also has the lowest carbon/energy ratio of all stable hydrocarbon fuels, resulting in low CO2 emissions. In addition, its principal component, C H 4 , does not have carbon-carbon molecular bonds; the result is a substantially lower probability of benzene (CgF^) ring formation. This reduces the formation of carcinogenic polycyclic aromatic hydrocarbons (PAH) and solid carbon particles (soot), lowering P M mass emissions [37]. A detailed discussion of the advantages and drawbacks of natural-gas fuelled heavy-duty engines is presented in Chapter 2. One technology that is under development to use natural gas in an otherwise unmodified diesel engine is Westport Innovations Inc.'s high-pressure direct injection (HPDI™) system. This system involves an injection of natural gas directly into the combustion chamber late in the compression stroke, resulting in a non-premixed combustion -6-process similar to that of a traditional diesel-fuelled engine. A small quantity of diesel, injected prior to the natural gas, is used to initiate the combustion. The system has been the focus of previous research, which has shown that aggressive in-cylinder emissions control techniques can reduce emissions sufficiently to meet either the P M or N O x standards, but not both [38]. Exhaust-gas recirculation (EGR) has been shown to be the most effective in-cylinder technique for reducing N O x . However, substantial increases in CO and reductions in efficiency have been demonstrated at the EGR levels required to meet the 2007-10 N O x standards [39]. Calibration of such a pilot-ignited, direct-injected natural gas (PIDING) engine over an entire test cycle has shown that diesel engine-like performance and efficiency can be achieved while P M and N O x emissions are significantly reduced. However, further work is required to reduce emissions to the very low levels required by the 2007/2010 standards [40]. N O x and/or P M aftertreatment may eventually be required even for this relatively clean combustion system. However, to optimize the engine system while minimizing the requirements for complex exhaust treatment equipment, it is necessary to understand the principal sources of the harmful emissions. 1.4 Objectives and Scope The overall objective of this work was to improve understanding of the combustion and pollutant formation processes in a PIDING engine. Specific interest was placed on the degradation of the combustion process, indicated by increased cyclical variability and slower energy release rates, with intake charge dilution (EGR). The research project was experimentally-based using a single-cylinder research engine to investigate various aspects of the combustion process under moderate and high EGR conditions. Specific project objectives included to: 1) Understand the influence of physical (mixing) and kinetic (chemical) influences on combustion rate, and the corresponding influence on emissions. Variations in injection pressure and modification of the chemical composition of the gaseous fuel were used to investigate this effect. 2) Investigate the role of ignition delay and combustion stability on pollutant emissions, through the addition of hydrogen to the gaseous fuel. 3) Explore the role of heavier hydrocarbons on ignition delay and combustion stability, and the corresponding influence on the pollutant formation process. 4) Identify the sensitivity of the combustion process to the overall carbon-hydrogen ratio, and identify whether specific species had a significant influence on combustion or emissions. 5) Examine the influence of particles recirculated with EGR on combustion instability and PM emission levels. By meeting these objectives, a better understanding of the pollutant formation in this engine system was to be achieved, specifically under high intake charge dilution conditions. From this, it may be possible to identify improved operating modes to optimize the combustion process, with the goal of maximizing efficiency while minimizing emissions. 1.5 Thesis Structure Three papers have been either published or submitted for publication based on the results presented in Chapters 4, 5 and 8. The work presented in Chapters 6 and 7 are currently in preparation for submission. To avoid redundancy and to improve the readability of this work, the papers have been modified such that common elements, including background information on the combustion system and the experimental facility have been presented as separate chapters (2 and 3). The thesis content is laid out as follows. The current chapter (Chapter 1) provides a general background and the motivation for the research. Chapter 2 provides a detailed review of the current state of knowledge regarding natural gas fuelled heavy-duty engines and of PIDING fuelling applied to heavy-duty engines. The mechanisms of pollutant formation under non-premixed natural gas combustion are also presented. In Chapter 3, a description of the experimental apparatus and procedures are supplied. Then, Chapter 4 presents the effects of increasing the in-cylinder turbulence through higher injection pressures on the combustion process. In Chapter 5, the effect of enhancing local diffusion and flame stability in a non-premixed gaseous jet through the addition of hydrogen to the fuel is examined. Chapter 6 discusses the effect of various fuel compositions on the overall combustion process, focusing on the effects of heavier hydrocarbons on ignition delay time and particle formation. The effect of diluting the fuel with nitrogen on combustion and pollutant formation is explored in Chapter 7. Chapter 8 discusses the effect of recirculated particulate on the combustion process, and specifically identifies its role in the formation of new particles. Finally, a summary of the main conclusions from the thesis and suggestions for future work are provided in Chapter 9. The tables and figures for each study are provided at the end of the corresponding chapter; references are numbered sequentially from the beginning of the thesis and are located after Chapter 9. Appendices provide further information on the experimental apparatus, procedures, and more details regarding the experimental results. 1.6 Tables and Figures Table 1.1: Heavy-Duty Engine Emissions Standards, in g/kWhr Country USA [25] Canada [26] Europe* [27] Japan [27] India [27] China [27] Year into effect 2007-2010 2007-2010 2008 2005 2010 2002 Standard Euro V Euro III Euro II CO 20.8 20.8 4.0 1.5 2.22 2.1 4.0 nmHC 0.19 0.19 0.55 0.46+ 0.17 0.661" 1.1* CH 4 1.1* NOx 0.27 0.27 2.0 2.0 2.0 5.0 7.0 PM 0.013 0.013 0.03 0.02 0.027 0.1 0.15 * - all engines are to meet both stationary (left) and transient (right) test cycle limits t - total hydrocarbons (including CH4 and nmHC) * - natural gas fuelled heavy-duty engines only Note: Levels are not directly comparable, as they are based on different test cycles Figure 1.1: US E P A Heavy-Duty Engine Emission Standards, 1988-2010 [25] -9-Chapter 2 Background Information Heavy-duty engines fuelled with natural gas provide an opportunity to achieve substantial air pollution benefits compared to conventional fuelling. Furthermore, price differences between natural gas and diesel provide a significant economic incentive for alternative fuelling in some markets. Today, substantial numbers of heavy-duty vehicles, notably in urban transit bus fleets, are fuelled with natural gas. To evaluate the benefits of natural gas fuelling, a number of studies have compared the emissions and performance of natural gas engines to traditional diesel-fuelled engines. Currently, the next generation of more efficient, less polluting natural gas vehicles is under development. One promising technology, which is the focus of this thesis, is pilot-ignited, direct-injected natural gas (PIDING) fuelling. This chapter is intended to provide an overview of the current status of in-service gaseous-fuelled heavy-duty engines, to identify some of the promising technologies for future generations of gaseous fuelled engines, and to review the current state-of-knowledge regarding gaseous combustion under conditions typically found in heavy-duty engines. 2.1 Gaseous Fuelling of HD Engines A wide range of engine technologies is either in use or under development for gaseous fuelling of heavy-duty engines. One fundamental distinction in these technologies is the method of delivering the fuel to the engine. In externally premixed charge engines (as for most automotive gasoline engines), the fuel is mixed with the air before it is ingested into the cylinder. In direct-injection engines (as in diesel engines), the fuel is injected directly into the combustion chamber just before the combustion is initiated. A second important distinction is the relative amounts of fuel and oxidizer which are present in the combustion chamber (the stoichiometry). A stoichiometric mixture is when there is just enough oxygen in the combustion chamber to oxidize all the fuel to final products (CO2 & H2O). This is the approach typically used in gasoline engines. In the case of lean fuel-air mixtures, there is more oxygen than is needed to burn all the fuel. Virtually all direct-injection engines operate at an overall lean mixture ratio, even at full load. -10-A key factor in the pursuit of an optimal fuelling technology is the inherent differences in the combustion characteristics of diesel versus natural gas. An important property of natural gas is that it is more resistant to auto-ignition ('engine knock') than most liquid hydrocarbons. This allows premixed natural gas engines to operate at compression ratios as high as 14:1, resulting in higher efficiencies [41]. Conversely, this resistance to auto-ignition results in one of the principal barriers to using natural gas in a compression ignition engine, where the cylinder temperature and pressure are insufficient for the fuel to auto-ignite. Typical ignition delay times for diesel fuel are repeatable and are on the order of 1 ms [42]. Experimental studies have shown that natural gas ignition delay times, under conditions similar to those found in a diesel engine, exceed 5 ms and are highly variable [43]. To overcome this ignition uncertainty in natural gas fuelled engines, a separate reliable ignition source is required. 2.1.1 Premixed technologies Adding the fuel to the air before it is ingested into the cylinder has a number of benefits, including lower fuel pressures, which reduces parasitic compressor work. The intake and compression strokes also allow time for the fuel to mix with the air, resulting in a nearly homogeneous mixture. However, the need to avoid premature autoignition of the fuel (which can damage the engine) limits the maximum compression ratio to less than that of a diesel engine [41], reducing the engine's theoretical maximum efficiency. A further barrier is the need for a higher-energy spark system to provide reliable ignition of lean natural gas mixtures at high compression ratios [44]. These high-energy ignition systems require a higher voltage input and have increased maintenance requirements. Lean-Burn with Spark Ignition The most common natural gas fuelled heavy-duty engines use lean premixed mixtures with a spark ignition system. With lean operation, the throttle (typically used to maintain stoichiometric mixture fractions) is used less, reducing the work required to ingest the fresh charge. However, as the mixture becomes leaner, the rate at which the flame propagates through the combustion chamber is reduced, lengthening the total duration of the combustion event. For very lean mixtures, the flame propagation rate is so slow that the combustion will be terminated by the expansion of the combustion gases in the power stroke before all the fuel-air mixture is consumed. While techniques such as artificially creating a rich mixture in -11-the vicinity of the spark plug improve the strength of the flame kernel and thereby reduce the overall combustion duration, they do not significantly affect the flame propagation rate [45,46]. Manufacturers currently (spring 2006) offering heavy-duty spark-ignition lean-burn natural gas fuelled engines include John Deere, M A N , and Cummins-Westport. Detroit Diesel's Series 50G and 60G natural gas engines (200-300 kW) appear to be no longer in production. John Deere's 8.1L 6081 series (187-209 kW) is certified at the California Air Resources Board optional low N O x level, with emissions of 1.6 g/kWhr combined N O x and non-methane hydrocarbons (nmHC) using an oxidation catalyst for CO and nmHC control [47]. M A N produces a 12L, 180-228 kW engine rated to EURO III standards [48]. Cummins-Westport offers a series of natural gas engines, with the B-gas (5.9L displacement, 145-172 kW), C-gas (8.3L, 186-205 kW), and L-gas (8.9L, 239 kW) engines all using oxidation catalysts. The Cummins-Westport engines are also certified at the optional low-N O x levels, with NO x+nmHC emissions of 2.4 g/kWhr (B&C-gas) and 1.88g/kWhr (L-gas) and P M emissions certification at US EPA 2010 levels of 0.013 g/kWhr [49]. A number of studies have compared emissions and performance of in-use natural gas fuelled engines with equivalent conventional medium and heavy-duty diesel-fuelled engines. A summary of some of these studies involving both urban transit buses and trucks is shown in Table 2.1. In most of the studies, multiple 'identical' vehicles were tested to reduce the influence of vehicle-to-vehicle variability. The studies used a range of vehicle test cycles, with some of the studies carrying out tests on two or more cycles (for these studies, the range of results is indicated in the table). A consistent reduction in P M and N O x emissions was observed with the natural gas fuelled engines. However, a significant loss in efficiency with natural gas fuelling was observed, although in most cases the net CO2 emissions were somewhat reduced due to the lower carbon content of the natural gas. The loss in efficiency was primarily attributed to lower compression ratios (reduced from their diesel-engine levels to avoid engine knock) and high unburned fuel emissions at light load conditions. Emissions of nmHC varied substantially between the tests; Chandler et al. [50] showed a significant nmHC reduction with natural gas fuelling, while the rest of the studies indicated increases that ranged from a few percent to over an order of magnitude. Those studies with the largest increases in nmHC -12-(such as Ayala et al. [51]) typically used older model natural-gas engines without oxidative aftertreatment. The effectiveness of oxidation catalysts used with lean-burn natural gas engines was shown by Frailey et al. [52], where one of three otherwise identical test vehicles was not equipped with an oxidation catalyst. CO, P M , and nmHC emissions were factors of 20, 8, and 4 higher, respectively, for the vehicle without the catalytic converter. McCormick et al. [53] report that conventional oxidation catalysts were found to effectively reduce emissions levels of all non-methane hydrocarbon emissions. Aldehyde emissions were measured, and were found to have mass ratios between 7% and 30% of the nmHC emissions, with lower ratios for more modern engines equipped with oxidation catalysts. Compared to the reference diesels, the nmHC emissions from the natural gas engines were found to have higher masses but were, generally, less toxic and had a lower smog-forming potential [53]. Kado et al. [54] found that P A H emissions (semi-volatile and particle-associated) were generally lower for natural gas engines than for diesels. Seagrave et al. [55] indicated that the mutagenicity of P M from newer-model natural gas engines equipped with oxidation catalysts was in general substantially lower than from equivalent diesel engines and was similar to low-emission gasoline engines. Holmen and Ayala [56] reported that using a natural gas engine with an oxidation catalyst resulted in particle number concentrations 10-100 times less than from diesel engines. Only Chandler et al. [50] and Kamel et al. [57] reported total HC (including methane) values, at levels of 10.9-13.4 g/km and 3.7-6.2 g/km, respectively. Both studies indicated that greater than 95% of the HC's were methane. Although methane emissions are currently unregulated in North America, they are of concern due to methane's role as a greenhouse gas (21 times as absorptive of infra-red radiation as CO2) [6]. By combining the CH4 and CO2 emissions, the net infrared absorption potential of the exhaust was increased by 3-7% (Kamel et al.) and by 25-40% (Chandler et al). These studies suggest that older-model natural gas engines generate a net greenhouse gas increase when compared to diesel engines, especially in transport truck applications. A large contribution to this is that the efficiency of the natural gas engines is 20-30% lower than the equivalent diesel, due to throttled operation at idle and light load. The effect is substantially less significant for more modern natural gas engines with better low-load combustion control. A recent study released by the National Renewable -13-Energy Lab (NREL) reports that, in a comparison between modern natural gas (2001 Cummins-Westport and 2004 John Deere) and diesel (2003 Detroit Diesel Corporation) powered buses, net infrared absorption potential is 20-25% lower for the natural gas fuelled vehicles [58]. The main reason for the difference was that, on the urban transit bus test cycle used in the N R E L study, the natural gas engines were only approximately 4% less efficient than the diesels. A further example of the effect of vehicle useage on efficiency is evident when comparing the Chandler et al. [50] and McCormick et al. [53] studies. McCormick et al. report natural-gas fuelled vehicle efficiencies approaching those of diesel engines over certain bus test cycles, while Chandler et al. report efficiency more than 30% worse on an urban heavy truck test cycle using very similar natural-gas fuelled engines. Stoichiometric with Spark Ignition The primary benefit of using premixed stoichiometric combustion is that an automotive style three-way catalytic converter can be used, allowing substantial reductions in CO, N O x , and HC emissions. The major disadvantage of such an approach from a commercial perspective is the need for an intake throttle to carefully control the mass of air inducted into the engine at all load conditions, resulting in a significant reduction in efficiency. M A N produces a stoichiometric version of its 12L engine that meets European enhanced environmentally friendly vehicle (EEV) regulations [48]. Of the studies outlined in Table 2.1 only Chiu et al. [59] studied the effects of stoichiometric operation with a three-way catalyst. N O x (0.2 g/kWhr), CO and nmHC were below 2010 levels although P M (0.02 g/kWhr) was above the standard. Efficiency penalties of 10-25%) were observed compared to the lean-burn engines. Nellen and Boulouchos [60] demonstrated the use of EGR in place of a throttle on a small (300kW) stationary engine. Low emissions with diesel-like efficiencies were achieved at steady state; however, the aptitude of this method to cope with transient operation was not investigated and unburned methane emissions were not reported. Lean Burn with Pilot Ignition Another ignition option that has been suggested for homogenous charge engines is the use of a diesel pilot injection. The pilot fuel will auto-ignite after injection, leading to near-simultaneous ignition of the natural gas charge at multiple points centrally located in the combustion chamber. This results in a more rapid combustion of the premixed charge, leading to lower unburned hydrocarbon emissions and improved efficiency at part load [61]. -14-Manufacturers including Wartsila [62], Fairbanks-Morse [63] and M A N [64] have developed homogenous charge pilot ignition systems for larger stationary and marine applications (>1MW). For heavy-duty on road automotive applications, Park et al. [61] have reported full-load efficiencies higher than an equivalent diesel engine, at the expense of increased N O x emissions. Singh et al. [65] and Zuo and Yang [66] showed that even with pilot injection, lean air-fuel mixtures at light load resulted in increased HC emissions and reduced efficiency due to reduced flame propagation rates. Shenghua et al. [67] suggested that pure diesel operation (no natural gas) was required at low loads to avoid over-lean charge mixtures. A l l the studies indicated reductions in N O x and P M could be achieved, at the expense of efficiency and CO and HC emissions [67,68]. A n application of lean premixed natural gas with diesel pilot ignition in a commercial setting is the Dual-Fuel™ system, under development by Clean Air Power Inc. This system has been applied to a Caterpillar C-12 engine, whose performance emissions were evaluated over the European Stationary Cycle 13 mode steady-state test cycle. N O x emissions of 0.72 g/kWhr and P M levels of 0.0054 g/kWhr were achieved in conjunction with a particulate filter [69]. Brake-specific energy consumption roughly equivalent to 249 g of diesel per kWhr (based on a diesel heating value of 43.2 MJ/kg [28]), was reported, approximately 7% greater than the baseline diesel [69]. To improve low-load performance, skip-firing (where natural gas was supplied to only four of the six cylinders) and a compressor bypass system on the turbocharger were used. Despite these efforts, high hydrocarbon emissions were still observed at part-load conditions, due to slow flame propagation in the premixed lean mixture. Lean Burn with Compression Ignition A different ignition concept is that of homogenous charge compression ignition (HCCI) where the premixed charge auto-ignites in the high temperatures and pressures found at the end of the compression stroke. Combustion stability and control have proven to be substantial barriers, while the need to avoid excessive in-cylinder pressures limit its use to low and moderate load conditions [70]. Various methods have been suggested to overcome the variable auto-ignition characteristics of natural gas, including blending with a second fuel (for example, naptha) before the fuel is ingested [71] and the careful control of recirculated -15-exhaust gases [72]. Significant issues of reliability and control under transient operating conditions remain to be resolved. 2.1.2 Direct Injection Technologies Natural gas may also be directly injected into the combustion chamber late in the compression stroke, similar to the fuel injection in diesel engines. This results in a primarily non-premixed combustion of a turbulent gaseous jet. Due to the poor auto-ignition qualities of natural gas, a separate ignition source is required to achieve reliable and repeatable combustion initiation. Direct Injection with Spark Ignition Using a spark plug as the ignition source has been suggested for direct-injected natural gas. For successful ignition, there needs to be the right proportion of fuel and air in the vicinity of the spark plug at the time of the spark discharge. Preliminary research carried out in quiescent combustion chambers has shown that the spark can repeatably ignite a non-premixed natural gas jet [73]. For real combustion chambers, where charge motion is significant, reliable ignition at different engine operating conditions is challenging to achieve. Changes in charge motion with operating condition and engine age affect the mixing status of the jet in the vicinity of the spark plug, resulting in degradation of the combustion event. Similar difficulties have been encountered in gasoline fuelled, light-duty direct-injection spark ignition engines, where careful optimization of the combustion system and spark timing control has been required to maintain stable operation [74]. Further drawbacks of this system include the fact that the flame is initiated close to the combustion chamber wall, and hence has to propagate from a single remote point through the rest of the combustible mixture. Failure to achieve ignition also results in misfire and high emissions of unburned fuel. Direct Injection with Hot Surface Ignition Using a hot surface igniter in a direct-injection engine overcomes some of the difficulties encountered with timing the spark discharge. A continuous transfer of heat is supplied to the charge in the vicinity of the injected fuel. Aesoy & Valland [75] demonstrated that efficiencies higher than those of an equivalent diesel engine could be achieved with direct injection of natural gas and a hot surface igniter. Surface temperatures of 1200-1400 K -16-were required, and careful control of the injection process was necessary to achieve stable combustion. Caterpillar [76] demonstrated the application of hot surface ignition of direct-injected natural gas for a heavy-duty engine. Reductions in engine-out N O x were achieved, at the expense of efficiency. Difficulties in ensuring repeatable ignition over a range of operating conditions were also reported [77]. Isuzu and Westport Innovations Inc. [78] have developed a direct-injection of natural gas hot-surface ignition system for a 4.5 L (100 kW) medium-duty diesel engine. Reductions in emissions were reported compared to the standard diesel engine, with no effect on efficiency. This system included an oxidation catalyst (modified to oxidize CH4) and a separate selective catalytic reduction system for N O x control. Low emissions of N O x , nmHC, and CO2 were observed; CO2 emissions of 631 g/kWhr were particularly noticeable, given that conventional lean burn natural gas engines emit on the order of 800 g/kWhr CO2 [78]. While hot surface ignition is promising, the longevity and reliability of the hot-surface materials, which are held at high temperatures for extended periods, are not yet known. Direction Injection with Pilot Ignition Using a diesel pilot spray as the ignition source for a non-premixed natural gas jet has also been proposed. Initial development of systems which directly inject both diesel and natural gas fuels have been carried out by Wartsila Engines, for 6-12 M W marine and stationary engines, and by Westport Innovations Inc., for heavy-duty transportation and smaller stationary power generation applications. A single concentric-needle injector is used to inject both fuels. The natural gas fuel is typically injected shortly before the diesel pilot injection (Wartsila) or after it (Westport). This system results in improved efficiency and higher power densities, as well as less sensitivity to natural gas composition, compared to other natural gas technologies [46]. Westport Innovations Inc.'s High Pressure Direct Injection (HPDI™) system has demonstrated low N O x emissions levels while retaining diesel engine efficiency through the use of recirculated exhaust gases and careful operating condition optimization. N O x emissions levels of 0.8 g/kWhr and P M of 0.04 g/kWhr were achieved in transient tests using an oxidation catalyst [40]. Due to poor transient response of the recirculated exhaust gas system, N O x levels were higher than had been found from an ESC-13 mode steady-state cycle test (0.48 g/kWhr). Further modifications of the injection control processes during -17-transient conditions may permit improved control over emissions levels. The reliability of the combustion system has been demonstrated in a series of long-duration commercial tests [79,80]. In March 2006, the engine was certified by the California Air Resources Board to meet the 2007-2009 heavy-duty engine phase in requirements for NO x+nmHC, at 1.6 g/kWhr; P M emissions, at 0.027 g/bkWhr still exceed the 2007 requirements [81]. Premixed and Direct Injection with Compression Ignition Based on Westport's HPDI™ combustion system, research has also been pursued on a combined system that includes a premixed lean natural gas charge [82]. A carefully controlled injection of diesel pilot early in the compression stroke results in HCCI-like combustion of the premixed fuel-air charge (where the fuel is a mixture of natural gas and evaporated diesel). To increase the engine's power without excessive cylinder pressures, an injection of natural gas shortly after the peak of the premixed combustion event results in non-premixed combustion of this gaseous jet. This later combustion does not significantly increase the peak pressure, but it does significantly increase the net power output. In a stationary power generation system (1 .6-2 MW), N O x levels of 0.67 g/kWhr were achieved at diesel-like efficiencies and power densities. Emissions were generally low, with CO2 levels of 480-520 g/kWhr, C H 4 of 3.75-4.15 g/kWhr, nmHC of 0.54 g/kWhr, and CO of 1.5-2.7 g/kWhr; P M emissions were not reported [82]. Similar developmental results were demonstrated on a 19L 6-cylinder engine. 2.2 Non-Premixed Gaseous Combustion The various technologies which are either in use or under development for fuelling heavy-duty engines with natural gas all involve significant tradeoffs. Premixed-charge engines face reduced efficiency and high unburned fuel emissions at part load. Late-cycle direct-injection technologies, while more efficient, are harder to ignite. A fixed ignition source (spark or hot-surface) is faced with the challenge of reliably igniting the injected fuel at the desired time. A pilot injection provides better control over the timing and location of the ignition event. This system is, arguably, the most promising of the developmental natural-gas technologies. Unfortunately, the combustion system is based on the non-premixed combustion of a turbulent high-pressure gaseous jet, which has not been as extensively studied as other more typical combustion systems. As such, understanding of the combustion -18-process, and in particular of the pollutant formation mechanisms, is not as well developed as that of either liquid diesel fuelled systems or of premixed combustion systems. 2.2.1 Combustion Structure The general structure of combustion in a PIDING system can be divided conceptually into three separate (although overlapping) regimes. The first regime involves the pre-combustion phase, where natural gas is injected into the (relatively) cold oxidizer. The natural gas jet behaves generally as described by Turner [83], with a travelling vortex ball preceding a quasi-steady state jet. Both Hi l l and Ouellette [84] and Rubas et al. [77] have applied this general model to natural gas injection into a diesel engine. These studies have indicated that air entrainment is restricted to the quasi-steady jet with little entrainment into the leading ball. A more recent study by Cossali et al. suggested that treating the leading portion as a toroidal structure rather than a ball was more appropriate [85]. Further, significant entrainment of air into the head of the jet was identified during the early stages of the injection process. These processes are not limited to natural gas; similar results have also been identified for transient hydrogen jets [86]. The momentum of the injected fluid provides the principal impetus for the jet propagation. Momentum transfer to the head from the jet increases the size of the head, and the corresponding diameter of the jet, as it extends into the combustion chamber. Prior to ignition, oxidizer will be mixing into the jet, with a general distribution from nearly pure fuel at the core of the jet to a steadily weaker mixture at the jet perimeter. Turbulent mixing will result in spatial and temporal non-uniformity in the mixture fraction around the jet. The total mass in the jet (m{x)), including both fuel and entrained oxidizer, at a given downstream distance x is given by [84]: m(x)=m0K-> (2-1) a where m is the mass rate of injected fuel, Ks is a constant (0.32) and d is the diameter of the nozzle. The location of the external ignition source strongly influences the ignition of the natural gas jet. Typically, the ignition will occur along the sides of the quasi-steady jet rather than in the region of the vortex head [87,88,89]. The flame then propagates from the ignition source through the combustible mixture on the outer periphery of the gas jet. Once the -19-premixed mixture has been consumed, the flame stabilizes into a non-premixed turbulent gaseous jet flame, which may be considered to be the second quasi-steady stage of the combustion event. Fuel diffuses towards the flame front from the core of the jet while oxidizer diffuses towards the core from the surroundings. The reaction zone w i l l be a surface (fluctuating with turbulent perturbations) around the jet, where the fuel-air mixture is at a near-stoichiometric ratio. Due to the high relative velocity between the gas and the oxidizer at the nozzle exit, and the correspondingly high local strain rates, the flame is lifted from the nozzle. A short distance downstream, the 'lifted' jet burns, in the form of a triple flame. A rich premixed phase (oxidizing air entrained in the lifted section of the jet) burns towards the core of the jet, while a lean premixed flame burns towards the oxidizer. This combustion event 'anchors' the non-premixed flame [90]. Only a small amount of oxidizer mixes into the fuel jet before the triple-flame. A s a result the rich combustion zone is limited in extent. Therefore, after ignition the jet contains virtually pure fuel, with only small quantities of oxidizer or combustion by-products. The third combustion phase occurs at the end of injection. A s the injector needle closes, the rate at which fuel is added, and its corresponding momentum transfer, to the jet diminishes rapidly. Effectively, the separated jet now acts as a 'puf f jet [84]. M i x i n g of the tail end of the jet with oxidizer results in combustion spreading around the fuel cloud, which continues to mix and burn as its momentum carries it away from the nozzle. The combustion process w i l l end when either there is insufficient fuel to sustain the reactions, insufficient oxidizer (which should not occur for diesel-type engines) or when the bulk expansion of the combustion gases (due to the motion of the piston in the power stroke) lowers the temperature of the reactants sufficiently that the reactions are no longer self-sustaining. The relative importance of these three phases depends on the relative durations of the ignition delay and the injection processes. If ignition does not occur until near the end of the injection, the combustion w i l l never develop into the quasi-steady jet, and w i l l instead only burn as a partially premixed cloud. Therefore, processes which either increase the ignition delay or reduce the injection duration tend to reduce the importance o f the quasi-steady mixing-controlled component of the combustion event. -20-2.2.2 Oxides of Nitrogen The mechanisms by which N 0 X are formed in internal-combustion engines are well understood [28,37]. The bulk of the N O x is formed as nitrogen oxide (NO) in the combustion chamber, some of which will react to NO2 during the exhaust process. Some NO2 may also be formed directly in the combustion chamber. The primary mechanism for NO formation is the thermal (Zeldovich) mechanism [37]. This mechanism is highly dependent on temperature due to the high activation energy of its rate-limiting step. The thermal mechanism is also slow, such that not only high temperatures, but a long (on engine time-scales) residence time at those temperatures is required to reach equilibrium. Due to the turbulent mixing between burned gases and cool unburned charge which typically occurs in diesel engines, the thermal mechanism does not normally reach equilibrium conditions. However, it is still the dominant formation mechanism under most conditions. Other NO formation mechanisms include the prompt and nitrous oxide routes [37]. The prompt mechanism results in the immediate formation of NO within the flame zone, unlike the thermal mechanism which, due to its low initial rate, does not contribute significantly to flame-front NO. This prompt mechanism involves the reaction of the C H radical with an N2 molecule to form a series of intermediate species which may eventually form NO. The controlling factor of this reaction is the C H radical, which is highly reactive and is typically found only within the flame region. The nitrous oxide (N2O) route involves the reaction of N2 with an oxygen radical and a third body to form N2O (N2 + O + M -> N2O + M). The N2O will then react with another oxygen atom to form two NO molecules. This reaction is limited by the oxygen radical concentration. Typically, it is only significant between 1000 and 2000 K , where there is a non-negligible quantity of O but where the thermal mechanism rate is very slow. A fourth source of NO that is discussed in the literature [37] is the fuel-bound NO route. This mechanism is most significant for fuels where significant quantities of atomic nitrogen are chemically bound in the fuel, such as coal or ammonia. Although natural gas may contain significant quantities of nitrogen, it is typically as molecular nitrogen (N2) and hence participates in the NO forming reactions similarly to the N2 in the oxidizer. Not all the NO produced in the combustion will be emitted, as some will decompose later in the process. One proposed mechanism for this is a reaction with the HCCO radical, -21-which reacts with NO to form H C N and CO2 [91]. As the HCCO radical is present in significant quantities within the flame zone, in a non-premixed flame, where some of the burned gases (which contain NO) may pass through the reaction zone again, some of that NO will be removed. Another route proposed for NO decomposition in diesel engines is the reverse of the thermal and prompt mechanisms [92,93]. Independent of which mechanism is dominant, these results indicate that a small but significant quantity of the N O contained in the oxidizer when using recirculated exhaust gases will be decomposed in the combustion reaction. How much of the species formed in the decomposition then recombine to form NO is unclear. 2.2.3 Carbon Monoxide Carbon monoxide is one of the main intermediate species in the oxidation of hydrocarbon fuels to CO2. In both stoichiometric and rich (overall) combustion conditions, CO results from a lack of oxygen in the reaction chamber. However, in a system which is overall lean, there should always be sufficient oxygen present to oxidize the CO. From such systems, CO is emitted when the fuel-air mixing is insufficient to provide enough oxygen to the reaction zone to oxidize the CO to CO2. The oxidation of CO can be stopped if the temperature in the reaction zone falls too far. As a result, events which quench the combustion will tend to result in increases in CO emissions. Typically, CO is not a significant concern for direct-injection engines in stable operation [28]. However, in engines operating under more extreme operating conditions (such as in excess of 50% EGR), high CO emissions do start to become a concern [128]. 2.2.4 Hydrocarbons The bulk of the HC emitted from natural gas engines is in the form of unburned methane [50,57]. One potential source of hydrocarbons is unburned fuel retained in the injector sac and nozzle holes, which will gradually enter the combustion chamber as the pressure drops during the expansion stroke. Another potential source is overleaning of the methane which was injected early in the cycle [28]. Some of this may have mixed beyond the lean limit of combustion before ignition occurs. Depending on the amount of excess air and the charge motion, some of this ultra-lean premixed mixture will participate in the combustion event from the oxidizer side. However, some may also be dispersed to the walls or crevices of the combustion chamber, and hence avoid being consumed. Bulk quenching at -22-the end of the combustion event is also a potential source of unreacted fuel. This occurs in the case of retarded combustion, where the pressure and temperature in the combustion chamber fall as a result of the motion of the piston in the expansion stroke. As the combustion chamber cools, the energy released during the combustion event will be insufficient to maintain the reactions, resulting in unreacted fuel and partial combustion by-products being emitted. This process can be minimized by enhanced late-combustion burn-up, potentially through higher turbulence intensities. Local extinction events caused by excessively high turbulent strain rates may also result in higher hydrocarbon emissions. In the case of local extinction events, 'holes' are created in the flame front through which unreacted fuel and partially reacted species (including CO and PM) may enter the oxidizer [37]. Some of these species will be reacted later in the combustion event; however, some may escape and be emitted. Flame extinction due to high heat losses as the flame approaches a wall (or a cold piston) is also a potential source of unburned hydrocarbons, as is penetration of the fuel jet into the relatively narrow and cold 'squish' region. Studies have suggested that, in the case of liquid fuels, these are significant sources of unburned hydrocarbon emissions due to impaired evaporation of the fuel [94]: while this effect could be expected to be less significant for a gaseous fuel, it is still potentially important. The relative importance of these different hydrocarbon emission sources remains to be clarified. While methane is the dominant HC emission from a HD natural-gas fuelled engine, emissions of other species are also of potential concern. Aldehydes, and specifically formaldehyde (CH2O), are a significant health concern. They are also key intermediates in the methane oxidation pathway [37]. Other higher-order hydrocarbons may also be formed, as a significant proportion of C H 4 will pass through ethane (C2H6) during the oxidation process. This can lead to the formation of C2H2 (acetylene) which is a precursor for the formation of polycyclic aromatic hydrocarbons (PAH's) as well as for carbonaceous particles. 2.2.5 Particulate Matter Particulate Matter emitted from direct-injection compression ignition engines is composed of a wide range of species, including volatile organics, solid carbon, metallic oxides or ash, and sulphates [18]. The relative fractions of these species vary as a function of, among other parameters, engine age and technology, fuel formulation, and engine operating -23-condition. The volatile organics are typically derived from unburned liquid fuel and lubricating oil, which has evaporated in the high-temperature combustion chamber and then recondenses during the expansion stroke or in the exhaust stream. These species are typically oxygenated or nitrated PAHs. Ash is formed from metallic compounds, primarily either additives or wear particles in the lubricating oil. Sulphates are formed from the sulphur in the diesel fuel, a small fraction of which is oxidized to SO3, some of which then reacts with water vapour to form H2SO4. This sulphuric acid then reacts in the exhaust and dilution processes to form sulphate particles [95]. It has been estimated that on the order of 4% of the sulphur in the fuel is converted to H2SO4 [96]. Because of the low sulphur content of natural gas (on the order of 1 ppm [97]), sulphate formation from the fuel is not a significant concern for a direct-injection natural gas combustion process [98]. Other sources, such as diesel pilot fuel and lubricating oil, will be the primary contributors to sulphate emissions. However, it should be noted that sulphur contribution from the lubricating oil for a heavy-duty engine has been estimated at approximately the equivalent of 20-50 ppm sulphur in the fuel [99]. This level was estimated to be below the level required to allow sulphate nucleation in the exhaust stream. Solid carbon (soot) is historically the greatest contributor to high-load P M emissions [18]; it is also the most widely studied component of P M . The formation mechanisms for soot in diesel engines are thought to be sufficiently well understood to allow the development of semi-empirical predictive models [100,101]. Understanding of the soot formation processes in high pressure, turbulent, non-premixed methane/air flames is not as well developed. However, the processes have many similarities; as a result, a short description of the current state of knowledge of the soot formation process in diesel engines is relevant. Soot Formation in Diesel Engines Diesel soot formation is a result of high-temperature fuel rich oxidation processes. As the liquid spray is injected, it entrains high-temperature air. As the diesel droplets evaporate, the hydrocarbon vapours mix with the surrounding air to form a fuel-rich mixture. This fuel-rich mixture will then undergo an exothermic reaction where the available oxygen is consumed [102]. However, as there is insufficient oxygen for the reactions to proceed to completion, long chain hydrocarbons such as PAHs and polyynes are present in this high-temperature environment. Through the widely accepted H A C A (hydrogen abstraction -24-acetylene addition) mechanism, hydrogen atoms react with the aromatic species, removing a hydrogen atom from the aromatic. This is followed by the addition of an acetylene molecule to the aromatic in place of the removed hydrogen atom [100]. Repeated steps of this process result in the growth of the aromatic P A H molecule. Particle inception occurs as P A H molecules collide and bind at the same time as they continue to grow via the H A C A mechanism. The exact point at which the agglomerated P A H molecules form an actual particle is a subject of some debate [6]. Another mechanism which has been proposed for pyrolitic (oxygen-deficient) formation of soot is the polyyne model (polyynes are hydrocarbon species of the form C2„H 2, where n is any positive integer greater than 1) [103]. This model suggests that polyynes grow relatively quickly through the addition of C2H radicals, releasing a hydrogen radical from the polyyne structure in the process. Particulate inception occurs at some stage during the ongoing growth of the polyyne molecule by polymerization of supersaturated polyyne vapours. While these models differ in detail, in general they agree that formation of soot particles occurs in fuel-rich environments where reactive hydrocarbon species, including acetylene and other radicals, are in abundance. Once the particles pass into a more oxidative environment, oxidation by O2 and the OH radical occurs [100]. Oxidation of particles emitted from diesel engines has been shown to be substantially different from that of pure graphite particles, possibly due to the presence of metallic ash and other non-hydrocarbon species in the diesel particles [104,105]. PM Formation in Methane Combustion The principal difference between diesel and methane soot formation is that for methane, recombination reactions are required to form C2 and higher hydrocarbons. For soot formation from methane, the initial stage involves reacting CH4 to CH3 through H, O, or OH radical reactions [37]. Methyl (CH3) radical concentration has been identified as a key factor in promoting the formation of aromatic rings [106]. If two CH3 molecules react, then either an ethyl (C2H5) or a C2H6 molecule may be formed. Hydrogen abstraction from this molecule leads to C2H2 [37]. While most of the C2H2 will be oxidized by an oxygen radical to form CO and CH2, some of the C2H2 may react to form benzene (C^), the base ring structure for the formation of PAH's . The relative amount of CH4 that is converted into C2 containing -25-molecules varies with temperature, pressure, and mixture stoichiometry, but can be as high as 80% [37]. Although soot formation in methane flames under diesel-engine conditions has not been extensively investigated, a number of studies have looked at soot formation from methane at a more fundamental level. McCain and Roberts [107] and Thomson et al. [108] have studied the role of system pressure on a laminar methane diffusion flame. Both studies showed soot varying with pressure raised to an exponent between one and two (P y, with l<y<2), for pressures from atmospheric to 2 MPa. At pressures between 2 and 4 MPa, the pressure sensitivity of the soot formation process was significantly reduced (exponential factor y approximately 0.1) [108], suggesting that at higher pressures, variations in pressure will have less impact on soot formation. For a turbulent non-premixed methane flame, Brooks and Moss demonstrated that increasing the pressure from 0.1 to 0.3 MPa increased soot volume fraction by almost an order of magnitude [109]. The effect of turbulence was investigated by Bohm et al. [110], who found that increasing turbulent strain reduced the formation of soot due to reductions in the formation of P A H precursors (C2H2, C3H3) as well as enhancing soot oxidation. The applicability of these results to turbulent non-premixed combustion in an engine environment is unclear; however it is apparent that the fluid dynamics of the reacting system will have a strong influence on particulate formation. The effect of fuel-oxidizer ratio was studied using rich premixed mixtures by Skjoth-Rasmussen et al. [ I l l ] , who found that at low pressures a soot inception limit was approached at a oxygen-to-carbon ratio of ~0.5 (that is, when there were more than ~2 oxygen atoms per carbon atom, soot formation would not occur). They also identified that soot and soot precursor (Ctfits, C2H2) formation were maximized at temperatures around 1500K. At low temperatures, the collected particulate was primarily agglomerated PAH's, while at higher temperatures the P M was carbonaceous soot. The effects of mixture stoichiometry and temperature were also studied by Gruenberger et al., using an atmospheric pressure non-premixed burner apparatus [112]. They found that both incomplete combustion and thermal decomposition could be significant contributors to carbonaceous soot emissions from natural gas combustion. Thermal decomposition was found to lead to solid carbon formation in highly fuel-rich regions, but much of this carbon was oxidized in the flame front. Solid carbon was also formed in the reaction zone due to incomplete oxidation -26-processes. Both processes were found to be enhanced at higher temperature, although the oxidation of soot formed from thermal decomposition was also hastened. The thermal decomposition was found to have a temperature limit (dependent on stoichiometry) below which soot production did not occur. Although these results are not directly applicable to high-pressure non-premixed diffusion flames, they do indicate that there is a range of combustion temperatures and mixture stoichiometrics above and below which soot formation is limited. Oxidation of PAHs and incipient soot particles was also studied for methane based flames. At low pressures, Smooke et al. [113] identified that OH oxidation was the principal mechanism, due to OH concentrations more than 10 times greater than the equilibrium concentration. Skjoth-Rasmussen also found that adding H2O to the fuel increased oxidation by increasing the concentration of OH radicals [111]. Zhu and Gore [114] identified that at high pressures (4 MPa), the OH oxidation rate was reduced to such an extent that it was similar to the O2 oxidation rate. In a non-premixed combustion event, O H radicals are typically found mainly within the reaction zone, but O2 molecules are abundant in the unburned gases. As a result, the increasing importance of the O2 mechanism at high pressures may have a significant influence on soot oxidation in a natural gas fuelled engine. The primary particles whose formation (and oxidation) are discussed above are typically formed on the fuel-rich side of a non-premixed flame. These primary particles are typically spherical and of similar diameter, on the order of 5-50 nm [18]. When the particle concentrations are sufficiently high, a significant number of collisions will occur which will result in agglomeration and the formation of particle chains. As the particles (both primary particles and agglomerate chains) pass through the flame front, they will be exposed to high-temperature oxidative environments. Oxidation will reduce both the number and size of the particles by surface reactions, as discussed previously. If the oxygen concentration, temperature, and residence time are high enough, all the particles will be fully oxidized; however, in most diffusion combustion events, there is not sufficient time at high temperature for this to occur. Those particles that survive the oxidation process will pass out of the reaction zone, where they will continue to interact with other particles and form larger agglomerate chains. After the combustion, the full range of aerosol population dynamics results in significant particulate growth and agglomeration, resulting in engine-out P M which - 2 7 -is greatly different, in shape and in composition, from the P M initially formed in the combustion process. 2.2.6 PM Population Dynamics P M changes significantly in the post-combustion expansion and exhaust process. Volatile hydrocarbons and sulphates will continue to condense onto the existing particles as the exhaust cools, increasing the P M size and coating the soot with a layer of material with significantly different properties from the particle's core. Existing particles will continue to agglomerate, resulting in a reduction in total particle number but an increase in size. New particles may be created by nucleation of sulphur or volatile organic species, i f the concentrations of these species are sufficiently high and there is not sufficient surface area for condensation to bring the concentration into equilibrium. These processes will continue through the expansion and exhaust strokes and during passage through the exhaust system [115]. To simulate the dilution process that would typically be undergone by actual exhaust, as well as to provide a more consistent and repeatable measure of particulate matter, samples are normally diluted before being measured. This dilution process significantly influences the particles being measured, by reducing the gas stream temperature (enhancing condensation of volatiles) and simultaneously diluting the exhaust by a factor of 10-20 (reducing partial pressure of volatile species). The effects of the dilution process on sampled particulate matter have been studied extensively [95,99,116,117]; variations in dilution air temperature, relative humidity, and residence time have been identified as the key sources of sampling variability. By holding these parameters constant, inconsistency in the measured particulate attributable to the sampling process can be minimized. 2.2.7 PM Formation in Pilot-Ignited Natural Gas Natural gas is not pure methane; depending on source, it may contain significant quantities of ethane (C2H6) and propane (C3H8), as well as trace amounts of higher hydrocarbons [118]. Given that these species have higher sooting tendencies than does methane, they may contribute disproportionately to the formation of soot in the non-premixed combustion of natural gas. Although they contain carbon-carbon bonds, it has been suggested that their primary role in the soot formation process is through enhanced decomposition to methyl (CH3) radicals [119]. Reaction pathway analysis has also indicated -28-that the oxidation of both ethane and propane pass through acetylene, generating a larger pool of soot precursor species [37,120,121]. Thus, in PIDING combustion, the natural gas flame may be a significant particulate source. Other sources similar to those in conventional diesel engines include the diesel pilot and the lubricating oil. A l l three sources can be expected to contribute to the overall P M emissions. 2.3 Pilot Ignited, Direct Injected Natural Gas A significant amount of research has been carried out on the PIDING combustion process. Ongoing development of PIDING fuelling systems has improved performance and combustion stability while reducing emissions. The initial research focused on optimizing the injection process; more recent studies have investigated the effects of charge dilution and the role of the diesel pilot on pollutant formation. These studies have been carried out at both the research and the commercial development levels. One of the key technologies which has been identified to allow the direct injection of natural gas system to minimize emissions is charge dilution through the use of exhaust gas recirculation. 2.3.1 Exhaust Gas Recirculation The use of recirculated exhaust gases (EGR) has been shown to substantially reduce N O x emissions from a pilot-ignited, direct-injection of natural gas engine [39,122]. Similar results have been reported for heavy-duty diesel fuelled engines [123,124]; however, these studies have indicated that P M emissions are substantially increased with high levels of EGR, especially at higher load conditions. However, the natural-gas system's inherently lower P M emissions allow higher levels of EGR over the entire operating range while maintaining acceptable P M exhaust concentrations. In direct-injection compression-ignition engines, EGR works by reducing the intake charge oxygen concentration. This increased dilution results in the presence of more inert species which must be heated by the chemical energy released in the combustion, reducing the overall temperature in the flame zone. This reduces the formation of NO, as its dominant mechanism is highly temperature-dependant, as discussed in Section 2.2.2. Results from testing at high intake dilution levels with natural gas fuelling suggest that the exponential dependence of NO formation on combustion temperature is consistent to the lowest temperatures at which combustion stability can be maintained [39]. The exponential -29-dependence of N O x and linear relationship of combustion temperature with intake dilution are shown in Figure 2.1 [122]. An overall activation energy for N O x formation in this combustion system was estimated [39,122] based on the peak in-cylinder temperature and was found to be within 5-10% of previous tests on diesel engines [125]. This suggests that, as would be expected, the mechanisms for N O x reduction with EGR are independent of the fuel being consumed in the direct-injection engine. EGR, and the corresponding reduction in combustion temperature, also increase combustion instability. A fundamental lower combustion temperature limit, below which the combustion instability was excessive and the combustion event was unsustainable, was proposed for diesel engines [126]. This hypothesis has been contradicted in more recent work, where very high EGR levels were used to achieve combustion temperatures below those where either soot or N O x formation was significant [127]. However, substantial increases in combustion instability, leading to higher emissions of unburned fuel and degraded fuel efficiency, were observed [128]. With lower oxygen concentrations in the oxidizer, local strain extinction of the flame is more likely to occur at a given temperature [129]. Thus, both the lower oxygen content and lower combustion temperatures due to EGR increase the probability of local strain extinction, substantially increasing combustion instability and emissions of hydrocarbons and CO. Bulk quenching, which occurs late in the combustion process as the cylinder temperature and pressure fall during the expansion stroke, is also a significant concern with higher levels of EGR. The longer combustion durations due to charge dilution result in the combustion continuing later in the expansion stroke. Bulk quenching results in premature termination of the late-stage oxidation processes which would normally consume byproducts of incomplete combustion (including hydrocarbons, CO, and soot precursors) [28]. Although natural gas burns under a wider range of fuel-air conditions than diesel, increases in combustion instability with EGR are still observed. These instabilities are particularly notable at later injection timings, where larger increases in CO, HC and P M , and greater reductions in efficiency are observed at lower EGR fractions [38,122,130]. Increases in the delay between the start of injection and the observed start-of-combustion of the natural gas (gas ignition delay) have been roughly correlated with increases in unburned fuel emissions [131]. This suggests that over-leaning of early-injected fuel may be a contributing -30-factor. Similar indications of the aetiology of the observed increases in P M and CO emissions have not been established. 2.3.2 Injection Process For a PIDING engine, the natural gas injection parameters have been shown to influence emissions and efficiency substantially. Injection timing, duration, and pressure are among the parameters that have been investigated. The number of injector holes, their angle relative to the firedeck, and the angle between the pilot and main gas nozzles are also significant, although fewer published results are available. One parameter of particular interest in a dual-fuel direct-injection application is the interlace angle between the diesel sprays and gas jets. Early computational studies suggested that this interlace angle could influence both the timing and location of the ignition of the natural gas jet, with direct influences on the overall heat-release rate and on N O x emissions [87]. In the Westport dual-fuel injector, the diesel pilot nozzles are located in the tip of the gas needle, which rotates during operation. As a result, the interlace angle between the gas jets and the pilot sprays varies over time. With equal numbers of pilot holes and gas jets, it was found that when the jet and spray were aligned, combustion stability was substantially degraded. Unequal numbers of pilot and gas holes were required to ensure that some of the diesel sprays were unaligned with the gas jets [88]. Unfortunately, this also increases the probability of some sprays and jets being in fact aligned, resulting in variations in the ignition times for the different jets. The effect was attributed to the natural gas jet overtaking the diesel spray before ignition occurred, resulting in a cool, fuel-rich mixture which was more difficult to ignite and from which flame propagation was impaired. The results also suggested that having the ignition source closer to the sides of the gas jet, rather than the tip, improved the combustion stability. Interactions between the gas and diesel injections were also identified when the delay between the diesel and gas injections was reduced. One hypothesized effect was that the cool natural gas reduced the temperature in the combustion chamber prior to ignition of the pilot fuel. This increased both the diesel and gas ignition delay times [132]. While efforts to develop more advanced models continue, modelling of the combustion process under EGR conditions is still unsatisfactory [133]. -31-The relative injection delay time (RIT, the delay time between the diesel pilot and main natural gas injections*) was studied in a series of tests. Initial studies under non-EGR conditions were inconclusive as the effects of the relative delay were confounded with variations in the overall combustion timing [88]. With the combustion timing fixed, relatively small effects of the RIT were observed. At very short RIT's, when the gas and diesel injections were occurring simultaneously, increased combustion instability was observed. Similarly, at long relative delays, the ignition of the natural gas was impaired by dissipation of the hot gases from the pilot flame [134]. With EGR, the RIT was found to have a more substantial influence on the combustion event. P M emissions were shown to be substantially reduced by shortening the delay so that the diesel and natural gas injections overlapped, as shown in Figure 2.2 [130,135]. The combustion intensity was increased and the duration decreased compared to a more conventional RIT (typically on the order of 1 ms). When the gas and pilot injections overlapped, the gas injection was completed before the diesel pilot ignited. The combustion of the natural gas was mainly premixed (although neither homogeneous nor necessarily stoichiometric), resulting in lower particulate formation. The higher intensity combustion resulted in a significant increase in N O x emissions. Unburned fuel emissions were also substantially increased, the result primarily of increased overleaning of the natural gas prior to ignition. The results were found to be sensitive to engine speed and load and to the phasing of the combustion event [130]. A parameter which has been widely recognized as a technique to control P M and N O x emissions from late-cycle direct injection engines is varying the absolute injection (and hence combustion) timing [28,136]. Later combustion timings result in lower N O x emissions due to lower combustion temperature at the expense of generally higher P M and reduced efficiency. Similar effects were observed with and without EGR for direct injection of natural gas engines [88,134]. With EGR, increases in combustion variability at later timings were also detected. Of particular interest is the comparison of the trade-offs between N O x , P M , and fuel consumption as a function of both EGR and combustion timing, as shown in Figure 2.3 [38]. To achieve equivalent N O x and P M emissions, either higher EGR fractions (-40%) with early combustion timing or moderate EGR (-20%) with late timings could be used. * HPDI™ injection timing processes are discussed in more detail in section 3.1.2 -32-Efficiencies were almost 10% higher at the high EGR, advanced timing condition. The fact that higher EGR is preferable raises system complexity issues, as larger EGR fractions require larger cooling systems and have a greater impact on the engine's air exchange system. Significant interactions between RIT, EGR level, and injection timing indicated that all three parameters need to be optimized over the entire engine operating map [130]. The injection pressure of both the diesel and the natural gas may also have a significant impact on the combustion process. Initial studies suggested that higher injection pressures reduced fuel consumption but increased N O x and unburned hydrocarbon emissions [88]. The N O x and efficiency effects were due to earlier combustion (the injection timing being held constant), as the ignition delay was shorter at high injection pressures. At a single operating condition, fixing the combustion timing and increasing the injection pressure reduced P M emissions with little impact on combustion efficiency or N O x emissions [135]. However, the effects appeared to be sensitive to operating conditions, particularly the relative pressure between the injector and the cylinder gases. The testing was unable to determine the fundamental effects of injection pressure [137]. Further study is required to clarify the effects of this parameter. 2.3.3 Diesel Pilot Influence The role of the diesel pilot flame has been investigated in both premixed and non-premixed combustion systems. For premixed charge conditions, the timing of the pilot injection is used to control the combustion timing. Earlier pilot injections advance the combustion, improving efficiency and reducing P M emissions but increasing N O x emissions [138]. The use of a diesel pilot (or a pre-injection) has also been studied as a technique to reduce emissions and combustion noise from diesel engines [139,140]. Although experimental results have varied, in general the pilot injection has been found to reduce the main fuel ignition delay time by increasing the temperature in the combustion chamber as well as providing an increase in radical concentration. The shorter ignition delay time of the main fuel leads to higher N O x emissions due to higher combustion temperatures. Excessively early pilot timing results in overleaning of the pilot while late injection results in interactions with the main fuel and impairs the penetration of the main fuel spray [141]. In a PIDING engine, the combustion timing depends primarily on the timing of the natural gas injection. The natural gas ignition delay, however, is dependant on the timing and -33-relative location of the diesel pilot flame [87]. This ignition delay time influences the amount of fuel-air mixing that occurs before the jet is ignited. The strength of the pilot flame, along with its location (as discussed in section 2.3.2), will influence both the timing and location of the natural gas ignition. The diesel pilot has also been shown to have a significant impact on P M emissions. A source apportionment study revealed that the diesel pilot contributes between 5 and 40% of the black-carbon component of the P M , with higher load or EGR conditions resulting in lower diesel pilot contributions [142]. Of specific interest was that increasing the pilot quantity from ~3 to 6% of the total energy substantially increased both the total P M and the relative contribution from the natural gas. This outcome may have been due to a reduction in the non-premixed natural gas combustion phase. A shorter ignition delay time may also have resulted in a significantly richer early combustion phase. Higher concentrations of particulate precursors, formed in the diesel combustion event, may also have contributed to the observed increase in natural-gas sourced P M . 2.4 Summary / L i terature Gap Testing on both diesel and natural-gas fuelled direct-injection engines has indicated that EGR can be used to substantially reduce combustion temperature, leading to lower N O x emissions. However, increased combustion variability, in conjunction with slower heat release, results in higher emissions of partial combustion by-products as well as reductions in efficiency. A survey of published results suggests that there are a number of competing influences which lead to higher emissions and lower efficiency with EGR in diesel engines, including local extinction and bulk quenching. Similar effects may be occurring for natural gas combustion; however, fundamental differences between the combustion of a gaseous jet and a diesel spray mean that different processes may be dominant. Further research is required to attempt to clarify these effects and to develop a better understanding of the principal pollutant formation mechanisms in direct-injection natural gas combustion systems. There are certain areas of interest in natural gas direct-injection combustion where significant gaps in the literature exist. The role of injection pressure has been shown, over a limited number of test conditions, to provide substantial emissions benefits; whether this is consistent over a wider range of controlled conditions needs to be determined. As well, the influence of fuel composition has been identified as a potential contributor to emissions; -34-however little research has been published studying the effects of heavier hydrocarbons, hydrogen, or fuel diluents in non-premixed combustion applications. Finally, the increases in P M emissions with EGR have been well described; however the influences of these recirculated particles on the combustion process, and their corresponding effects on emissions, have not been assessed. The work described in the following chapters will attempt to address these issues, and thereby develop an improved overall understanding of combustion and pollutant formation in a non-premixed gaseous-fuelled engine. 2.5 Tables and Figures Table 2.1: Selected results from in use natural-gas fuelled HD engine studies Study McCormick ef al. Chandler ef al. Frailey ef al. Kamel ef al. Ayala ef al. Ch u ef al. Publication year 1999 1999 2000 2002 2002 2004 2004 Manufecturer Cummins Cummins Cummins CWI DDC Mack Mack Engine model (year) B5.9G (1997) L10-300 (1997) B5.9G (1997) BGas+ (2001) 50G (2000) E7G E7G NG stoichiometry Lean Lean Lean Lean Lean Stoich Lean NG aftertreatment oxidation none oxidation oxidation none 3-way cat. oxidation Engine Power (kW, NG) 300 145 280 242 242 Displacement (L, NG) 5.9 9.8 5.9 8.3 12 12 Compression Ratio (NG) 10.5 10.5 10 11.5 11.5 efficiency (% of diesel) 85-98% 70% 8 3 % 80% 75-90% 90-95% CO (% of diesel emissions) 5-10% 265% 16% 10-15% 400% 4.8'g/kWhr. nmHC (% of diesel emissions) 1000-1500% 4 1 % 220% 100-120% 1000-1500% 0.0'gftWhr NOx (% of diesel emissions) 65% 2 1 % 4 0 % 55-75% 30-50% 6% 32 P M (% of diesel emissions) 5 % 5 % 3 % <10% 15-33% 0.02'g/Awhr C 0 2 (% of diesel emissions) 80-90% 102-113% lillPBBillillii 9 3 % 80-95% percentages calculated as NG/diesel*100, in comparison with closest equivalent diesel engine used in study, except "-absolute emissions values (g/kWhr) o 111 o ro X O z --S< o z Q5 X o 0.01 2800 2300 1800 1300 O ro i cu C L E <D I-<D E ro 0.23 0.21 0.19 0.17 0.15 0.13 Intake 0 2 mass fraction (g 02/g intake) Figure 2.1: Effect of intake oxygen dilution on peak combustion temperature and N O x emissions over a range of engine test conditions. N O x emissions are represented as emissions relative to the non-EGR emissions at the same engine test condition. -35-0.10 | o . 0 8 § 0 . 0 6 O l J 0 . 0 4 0.02 0.00 Figure 2.2: ^o—RIT = -0 .4 -| - -• RIT = 0.2 X RIT =1 .8 2 4 6 8 NOx (g/GikWhr) 10 10 O 6 O 4 x 2 0 -0— RIT = -0.4 • Rrr = o.2 -X-— RiT = 1.8 \ \ \ \ 4 6 8 N0 X (g/GikWhr) 10 Effect of delay between natural gas and diesel injection on P M - N O x and HC-N O x trade-offs as EGR level is varied (lower N O x at lower EGR levels). 0.08 _ 0.06 CO 0 . 0 4 0.02 0.00 Figure 2.3: -$-40% EGR -•-20% EGR -X-No EGR Retarding Timing -X-3 6 9 NOx (g/GikWhr) 12 CD c o CL E w c o O a5 LL. 190 185 •« 180 175 170 •40% EGR •20% EGR -X-No EGR Retarding Timing 3 6 9 NOx (g/GikWhr) 12 Combustion timing and EGR fraction effects on P M - N O x and fuel consumption-NOx trade-offs over a range of EGR conditions. Mid-point of combustion timing varied between 2.5° (early) and 17.5° (late) after top-dead-centre. -36-Chapter 3 Apparatus and Procedures 3.1 Research Engine The single-cylinder research engine facility was developed to provide the widest achievable range of test conditions on a heavy-duty gaseous-fuelled engine. The core of the facility is a heavy-duty Cummins ISX series engine (serial #4026530). This inline six-cylinder direct injection diesel engine series, typically used in transportation and stationary power generation applications, covers a power range between 300 and 480 kW (400 and 650 hp). The base multi-cylinder engine was modified so that only a single cylinder (number six, nearest the flywheel) fired. The other five cylinders were deactivated by sealing the valves, replacing the injectors with dummies, and drilling holes through the pistons to reduce compressive work. To maintain the engine's balance, the mass removed in drilling through the pistons was replaced by mass added to the wrist pin. To retain the best achievable speed stability under single-cylinder conditions, the largest available flywheel (inertia ~6.8 kgm ) was provided. Previous work has shown that the single-cylinder engine performance closely resembles that of an equivalent multi-cylinder engine [143]. Details regarding the single-cylinder engine's specifications are shown in Table 3.1. Two compression ratios are shown as two separate combustion chambers were used during different stages of the testing described in this work. The first set of tests (described in Chapter 4) involved the use of the higher compression-ratio piston, which was the standard Cummins ISX piston for the .1998 model year. The later tests were carried out using a newer Cummins piston, developed to meet more recent emissions standards through the use of EGR. The lower compression ratio was required to avoid exceeding the cylinder pressure limits with the higher charge masses associated with EGR. Only the clearance volume was affected; the engine bore, stroke, displacement, etc. were unchanged between the two pistons. To overcome the single-cylinder engine's high internal friction, an external 30kW (40hp) motor was used. This motor was controlled by a torque-control vector-type drive, with the motor providing torque greater than the friction being absorbed by the engine. A General Dynamics 150 kW (200 hp) eddy-current dynamometer coupled to a Digilog controller was used to absorb extra torque. The dynamometer and vector drive were mounted -37-in series, with the vector drive providing constant torque (greater than required to drive the engine) and the dynamometer absorbing varying torque levels to maintain the engine speed at the desired setting. This system proved capable of controlling the engines speed to within ± 1 % of the desired speed. Engine cooling was provided by a closed-loop recirculating system filled with a 50/50 blend of glycol and distilled water. The engine's internal thermostats were used to maintain the coolant temperature at the outlet from the block at 80±2°C. A liquid-to-liquid heat exchanger was used to reject excess heat from the coolant to the city water mains. The coolant regulated the oil temperature; the oil temperature varied between 90 and 110°C depending on the engine operating condition, but was repeatable on a day-to-day basis (±2°C). Energy rejection from the dynamometer was also provided by a separate mains water-cooling system. A sufficient supply of water was provided to ensure that the discharge temperature of either water system did not exceed 40°C at any operating condition. Intake and exhaust fans were used to provide ventilation and temperature control inside the test cell. The exhaust fan, ducted to the outside, was sized larger than the intake fan to ensure that the air pressure inside the test cell was lower than that in the control area. 3.1.1 Air Exchange System A highly flexible air exchange system was developed to provide the widest possible range of intake air and EGR conditions. Combustion air was supplied to the engine from an industrial rotary screw-type air compressor (Atlas Copco GA45FF). Water vapour was removed from the compressed air by a refrigerated air dryer (dew point of -40°C) and any oil carried over from the compressor was filtered with a pair of high-efficiency filters. Sampling of the intake air showed that aerosol concentrations were below those of ambient air. Boost pressures from atmospheric to 350 kPa (absolute) pressure were achieved through a two-stage regulation system. The air flow-rate was measured with a custom-built subsonic venturi. A 50 L surge tank installed upstream of the engine reduced the amplitude of acoustic waves in the intake system by approximately 95%. The intake air was then supplied to the stock internal intake manifold. The intake header, port, and valve seat geometry were unmodified from the base engine. Details of the intake air system are shown in Figure 3.1. The stock engine's external exhaust manifold was replaced by a custom-designed system in order to more closely simulate the multi-cylinder engine. An electronically -38-controlled butterfly valve was used to exert the back-pressure which would normally have been supplied by a turbocharger. An insulated 50L high-temperature surge tank was used to damp out pulsations in the exhaust stream. Both exhaust emission samples (gaseous and aerosol) were drawn from the exhaust downstream of the surge tank. While this resulted in some ageing of the exhaust (average residence time in the surge tank at 1200 R P M was on the order of 10 s), it was required to avoid sample biasing caused by inhomogeneity in the exhaust stream interacting with acoustic pulsations in the exhaust system between the engine and the exhaust surge tank. Upstream of the back-pressure valve, a fraction of the exhaust was directed through the EGR loop. This loop was composed of a second remotely controlled butterfly valve (to control pressure-drop and hence EGR flow rate) and a cooler. The cooler controls were set to hold the EGR temperature at 50±10°C throughout the testing. The EGR was mixed into the intake air upstream of the intake surge tank, to ensure complete mixing of the charge. A small sample (1 litre/minute) of the intake flow was drawn from downstream of the intake surge tank to determine the EGR fraction and to measure the intake aerosol loading. 3.1.2 Fuelling System The fuel supply system provided diesel and natural gas to the engine at pressures up to 31 MPa. A hydraulic pump driven by an 11.2 kW (15 hp) electric motor was used to compress the pilot diesel fuel. The natural gas was compressed by a multi-stage electrically driven compression system which maintained the supply gas pressure between 32.4 and 34 MPa (4700 and 4900 psi). A high-pressure accumulator was used to reduce the compressor cycling time and to minimize the fluctuations in the supply pressure. A manual regulator was used to set the diesel pressure, while a dome-loaded regulator held the natural gas at a pressure -200 kPa below the diesel pressure, to prevent leakage of the natural gas into the diesel. The fuel was supplied to the engine's internal fuelling rails, which delivered the fuel to the injector. The diesel fuel was also used to control the operation of the injector, and excess flow was returned to the low-pressure reservoir. The pilot fuel for all tests was road-grade low-sulphur (<500 ppm) diesel which met Canadian General Standards Board specification CAN/CGSB-3.520. The main fuel was commercially supplied natural gas (Terasen Inc.); while the composition varied somewhat (analyses are provided in the individual chapters), sulphur was at all times < 7 ppm (by volume). -39-The injector used in this work was a Westport J36 prototype injector. The injector, very similar to that shown in Figure 3.2, injected the diesel and the gas through separate injection holes. Concentric needles, actuated by separate solenoids, were used to allow independent control of the diesel and natural gas timings. Details of the injector are provided in Table 3.1; more information on the injector design and development are provided in the references [144,145]. The injection process used in this work involved an initial injection of diesel (at approximately 5 mg/injection) followed after a 1 ms delay by the gaseous-fuel injection. The timing process and related command parameters are shown in Figure 3.3. Briefly, the quantity of fuel injected was controlled by the duration of the injection (pulse width) for both the liquid pilot and gaseous main fuels. The start-of-injection of the pilot was controlled relative to top-dead-centre (TDC) of the operating piston while the gaseous injection was fixed relative to the commanded end of injection timing for the pilot. 3.1.3 Instrumentation and Data Acquisition The research engine facility was fully instrumented, including temperature, pressure and flow sensors as well as in-cylinder condition monitors. A brief description is given below, further details on the instruments used, including ranges, accuracy, and serial numbers, are provided in Appendix 3: • Airflow was measured using a custom-built subsonic venturi (estimated uncertainty ± 1%) located in the combustion air supply line between the two regulator stages. The differential pressure between the upstream and throat was measured with a differential pressure transducer (Omega PX2300-2DI). • The mass flow rate of gaseous fuel was measured using a Micromotion corriolis mass flow meter (sensor CMF010P, transmitter RFT9739). As the sensor measured mass flow directly, it was insensitive to the composition of the fuel. • Diesel pilot flow was measured using a custom-built gravimetric system, where the change in mass of fuel in the diesel reservoir over time was monitored. Regression analyses of the diesel mass data collected over sample durations in excess of 5 minutes were found to provide measurements within 5% of long-duration means. • Temperatures in the intake manifold, exhaust and EGR system, as well as standard engine and auxiliary system temperatures, were measured using standard accuracy Omega k-type thermocouples. -40-• Intake and exhaust stream pressures were measured using Setra type 209 diaphragm transducers. • The intake manifold pressure was measured at high frequency (every half crank angle -!/2°CA) with a PCB (model PCB1501) piezo-resistive transducer. • The in-cylinder pressure was also measured every '/2°CA with an AVLQC33C flush-mounted water-cooled transducer. The signal from the sensor was amplified and converted into a measurable voltage by a Kistler model 503 charge amplifier. • The crank location was measured using a BEIXH25D shaft encoder. This unit output both an index pulse and a square wave to provide !/2°CA resolution. The encoder was indexed to the engine's TDC. The pulse offset from true TDC was measured and used as a correction factor in the data acquisition system. The offset used in the work presented here was 1.7° before TDC. The data was collected through a custom-built data acquisition system. The core of the unit was a National Instruments SCXI1001 bus connected to a PCI-MIO-16E-1 card in a PC computer. The PC computer was running Labview 6.0, in which all the data acquisition routines were written. Two separate routines were used; one collected 45 consecutive cycles (at '/2°CA resolution) of the crank angle, intake manifold pressure and in-cylinder pressure, while the other collected all the other data (flows, temperatures, pressures) at a frequency of 1 Hz. The index pulse was imported directly into the timing circuit and was used to trigger the high-frequency data collection. Three separate data acquisition modules were used for low-frequency sampling in the SCXI1001 bus. One was dedicated for thermocouples (with a thermistor to provide the reference junction temperature, SCXI1303/1102), while the other two measured voltages. One of the voltage boards was unfiltered (SCXI1100) and the other-was filtered at 200 Hz (SCXI1102). The measurements were conducted through the unfiltered module unless excessive noise was observed on the sensor. Data collection periods varied for different test conditions, with a minimum duration of 5 minutes for low-frequency data. For most test points, two or more high-frequency data sets were collected (sampled at the start and end of the low-frequency sampling) to ensure that the engine's in-cylinder conditions had not changed appreciably over the sampling duration. -41-3.2 Emissions Measurements Emissions measurements were carried out with a fractional raw system for gaseous emissions and a fractional diluted system for particulate. As was mentioned in section 3.1.1, separate samples for the gaseous and particulate emissions analysis were drawn from the exhaust system downstream of the exhaust surge tank. For the gaseous emissions sample, an insulated l m long diameter tube was used to connect the exhaust sample point to the emissions handling system. A heated sample line (set to 190°C) was used to transfer the sample to the gaseous emissions bench. For the particulate sampling system, a separate lm long V" diameter tube was used to transfer the sample to the dilution point. The sample handling and dilution process for the particulate sample will be discussed in section 3.2.2. Further details of the sample handling systems are shown in Figures 3.1 and 3.4. 3.2.1 Gaseous Emissions The gaseous emissions system was a raw exhaust system custom assembled in 1992 by Galvanic Analytical Inc. Sample handling included a pair of heated filters (coarse and fine), a heated pump, and a chiller to remove water vapour (dew point <1°C). Upstream of the chiller, a sample was drawn through the heated flame-ionization detector to measure hydrocarbon concentration (Ratfisch RS232). After the water separator, the remaining dry sample was fed to the specific gas analyzers. Non-dispersive infrared (NDIR) analyzers were used for the exhaust CO2 (Beckmann 880), CO (Siemens Ultramat 2IP) and C H 4 (Siemens Ultramat 22P). A chemiluminescent analyzer (API 200AH) equipped with a Mini-HICON high temperature NO2—»NO converter was used to measure total N O x . A paramagnetic sensor (Siemens Oxymat 5E) was used to measure oxygen concentration. A dedicated sampling and analysis system was used to measure the CO2 concentration in the intake to determine the intake charge dilution level. The sample was drawn from the intake manifold downstream of the intake surge tank, ensuring that the charge was homogeneous. The sampling system included a pressure regulator, to reduce the manifold pressure to ambient, as well as a particle filter and a liquid-water separator. A Permapure nafion-membrane gas-to-gas dehumidifier removed water vapour from the sample line. The manufacturer's stated performance for the unit was -90% removal for the flow rates used. A condensation test was conducted and no condensate was detected in the stream after -42-the drier at temperatures as low as 2°C. A California Analytical (Model 100) NDIR analyzer was used to measure the CO2 concentration in the intake stream. 3.2.2 Particulate Measurement A separate sampling and measurement system was used for the particulate matter measurements. To coincide with standard vehicle P M sampling procedures [146], as well as to ensure a stable and repeatable P M sample with low enough concentrations to be measured accurately, dilution of the exhaust sample was required. A micro-dilution tunnel was used to dilute a small fraction (1-2 LPM) of the exhaust flow at dilution ratios of approximately 15:1. Due to the wide range of exhaust flow-rates at different engine operating conditions and the fixed flow requirements of the emissions measuring system, it was not possible to ensure isokinetic sampling for all conditions. However, as the aerosol particles of interest were significantly less than 1 pm in diameter, the effect of velocity biasing on the sample was negligible [147]. The P M sampling and dilution system is shown in Figure 3.4. To reduce the pressure in the P M sample system, an insulated V" sample line (~1 m long) connected the exhaust sample point to the dilution system. A bypass regulator was used to control the pressure at this point, with a fraction of the sample being vented through the regulator to atmospheric pressure. The remaining sample passed through a pressure-reducing orifice, after which it was mixed with dry, bottled nitrogen. As the mixing occurred immediately downstream of the orifice plate, a high degree of local turbulence and hence more rapid mixing was achieved. The diluted sample then passed through a 3 m long (3/8" diameter) mixing section. The dilution ratio was controlled with the bypass regulator by varying the pressure upstream of the orifice. By maintaining a constant dilution ratio and flow rate of diluent, the sample's residence time in the sampling system was held constant. The orifice and mixing section were heated to maintain the diluted gas temperature at 55°C. The dilution ratio, on a molar (or standard-volume) basis, was calculated based on the CO2 concentrations in the exhaust, diluent, and diluted sample streams, by: DR= ^ L - t ^ L ^ , — ^ ( 3 1 ) V--^2 Idiluted sample ~\^^2ldiluenl Where [C02]exh was the measured CO2 concentration in the exhaust stream from the raw emissions sample, [CO2]diluent was the dry nitrogen CO2 concentration (measured at less than 10 ppm) and [CO2]diluted sample was the measured CO2 concentration in the diluted sample. A -43-California Analytical NDIR analyzer (model 100) measured the diluted sample CO2 concentration. The CO2 concentrations, which were measured on a 'dry' basis, were corrected for the presence of water vapour by assuming a concentration of water vapour in the raw exhaust stream based on the hydrogen : carbon ratio of the fuel. Dilution ratios between 14-18:1 (by volume) were used, as at these levels the dilution ratio was found to have little or no influence on the measured P M levels. An analysis of the effects of dilution ratio on P M mass and size distribution measurements is provided in Appendix 1. The residence time of the sample was approximately 4 seconds in the mixing section, which allowed sufficient time for complete mixing as well as evolution of the aerosol population to steady-state conditions. After the mixing section, samples were drawn off the main stream for the various particulate measurement units. The temperature of the diluted sample in the diluted sample section was held at 50°C. Excess sample was exhausted to atmosphere. Total PMmass The total mass of particulate measured at 50°C was composed of solid and liquid particles which had either been emitted from the engine as an aerosol or had condensed during the exhaust or dilution processes. A tapered element oscillating microbalance (TEOM - Rupprecht and Patashnick Model 1105) was used to provide on-line total mass measurements. The T E O M collected all aerosol mass (irrespective of composition) on an oscillating filter, with the frequency of oscillation being a function of the collected mass. Previous work has shown the T E O M to be sensitive to sample humidity and filter face temperature [148,149]. By maintaining high filter face temperatures (50±2°C), a constant dilution ratio, and sufficient stabilization time, the effects of water content on the T E O M were minimized. To validate the T E O M results, gravimetric filter samples were collected for approximately 1/3 of the test points; the filters were then weighed to determine the mass of particulate collected over the sampling period. EPA standard protocols were followed for the P M sampling, with a pair of 47 mm Pallflex Emfab™ (borosilicate microfibers reinforced with woven glass cloth and bonded with PTFE) filters sampled in series. The filters and housing were pre-heated prior to being installed in the sample system to minimize condensation on the filter face. The filters were pre- and post-weighed following EPA -44-recommended practices [150] with a Sartorius CP2 microbalance, after being left in a temperature and humidity controlled room (20±2°C, 35±5%RH) for between 24 and 72 hours. The mass collected on the second filter was subtracted from the mass collected on the first filter to remove the effects of vapour absorption on the filter substrate [151]. The mass concentration in the exhaust stream was then calculated from the mass collected on the filter, the sample flow rate, the dilution ratio, and the sample duration. A strong correlation between the T E O M and the gravimetric filters was achieved, as shown in Figure 3.5. The T E O M results were consistently approximately 15% below the gravimetric filters; this is a similar offset to results reported by other researchers [148,152]. A Chi-squared goodness-of-fit test showed a high degree of correlation for particle concentrations less than 2 mg/hr (p<0.00\). However, for the few test points where the particle concentrations were greater than 2 mg/hr a substantially different response was observed, with the T E O M reading approximately 40% below the gravimetric filters. While the reasons for this result are unclear, the response of the T E O M was very similar to gravimetric filter measurements over a wide range of test conditions. This demonstrates that for virtually all the test conditions, the T E O M provided a valid representation of the actual P M mass in the exhaust stream. However, as can be seen in Figure 3.5, significantly more scatter in the correlation was observed at the lowest P M loadings. As a result, P M levels measured by the T E O M at the lowest loadings may not be as accurate as those at intermediate P M concentrations. Ultrafine particle Concentration To provide some insight into the particulate loadings at very low emission levels, as well as to develop an improved understanding of the particulate structure, number/size distributions were also measured with a TSI Model 3936 Scanning Mobility Particle Sizer (SMPS). The SMPS is composed of two principal units, a short-column differential mobility analyzer (DMA, TSI Model 3085) sized for ultrafine particles (5-155 nm mobility diameter) and a condensation particle counter (CPC, TSI Model 3022A). The fundamental operating principles of the D M A have been described by Knutson and Whitby [153], while Wang and Flagan [154] have developed a comprehensive description of the SMPS. A series of recent review studies have demonstrated the applicability of the SMPS system to P M emissions measurements from both diesel and gasoline-fuelled vehicles [155,156]. Some concern has -45-been expressed regarding the SMPS's ability to measure particles greater than 1000 nm in mobility diameter; however, the use of an impactor with a 50% cut diameter (diameter at which 50% of particles are removed from the sample stream) substantially less than 1000 nm resolved this issue [157]. The functioning of the SMPS has been described in detail in other work [154]. In brief, the sample first passed through an impactor to remove large particles. This was followed by a K r 8 5 charge neutralizer, where the particles achieved a standard charge distribution (dependant on particle size, as predicted by Fuchs charging theory [158]). The neutralized particles then entered a sample column, through which a laminar flow of clean air was maintained and across which an electric field was generated. The rate at which a particle crossed the streamlines was a function of the charge on the particle, its size and shape (related to atmospheric drag forces), and the strength of the electric field. Only those particles with the right 'mobility' crossed the stream in the proper time to reach the sample port. By varying the strength of the electric field, different particle sizes were sampled through the same port. The monodisperse aerosol being drawn from the sample port was then directed through the CPC. Inside the CPC, the sample passed through a super-saturated alcohol environment that caused all the particles to grow by condensation until they reached an optically-detectable size. By inverting the measured concentrations over time and correcting for multiple charges on the individual particles (based on Fuchs charging theory), the number of particles in 50 size ranges between 5 and 150 nm were determined. To avoid saturating the CPC, dilution ratios in excess of 100:1 were required. A filtered recirculation system was used to provide secondary dilution at a ratio of 5-14:1, resulting in overall dilution ratios between 200 and 300:1 (total diluted flow : raw exhaust). The effects of secondary dilution on particle size distributions are shown in Appendix 1. Validation of the results from the SMPS was difficult, as there was no 'gold standard' with which it could be reliably calibrated. The mobility diameter measured by the SMPS was a function of both the size and the shape of the particle, and was not the same as the diameter that a spherical particle of the same mass would occupy. Hence, the mobility size distribution cannot simply be integrated to provide the particle volume or mass. Previous work for diesel ultra-fine particles has identified effective densities in the range of 0.57 to 1.2 g/cm3 with a strong dependence on particle size, with higher effective densities for smaller diameter -46-particles [159,160] . The f racta l - l ike structure o f the agglomerate has also been s h o w n to have a s igni f icant inf luence o n the effective density o f the particulate for larger part ic les ; at smal ler s izes, the projected area has the pr imary inf luence o n m o b i l i t y [161]. A n estimate o f the effect ive density for a g iven size range can be made based on the fractal d i m e n s i o n for part icles o f a g iven m o b i l i t y diameter [162]. U s i n g the results f r o m the prev ious w o r k , a re lat ionship between effective density and diameter can be developed [159]: P , W « ^ > (3-2) where pe(<f>) is the effective density at the g iven diameter, fa is the m o b i l i t y diameter , and d/m is the fractal d imens ion . Fractal d imensions f r o m 1.8 to 3 have been suggested for diesel exhaust part ic les [159,163,164] . Fractal d imensions approaching 3 i m p l y that the agglomerates are essential ly spher ical and that the density is constant w i t h s i ze ; for the current w o r k , T E M images ( A p p e n d i x 7) indicate that the larger agglomerates are more c h a i n - l i k e . H o w e v e r , for smal ler part icles (made up o f fewer pr imary part ic les, whose diameter is o n the order o f 2 0 - 4 0 nm) , the fractal d imens ion w i l l approach 3 . F o r the purpose o f compar ing the S M P S results to measured P M mass emiss ions , three different fractal d imensions were assessed: 1.8, w h i c h was the lowest value ident i f ied i n the literature for d iese l - l i ke combust ion aerosol ; 2 .5 , w h i c h prov ided a reasonable intermediate va lue ; and 3 .0 , w h i c h results i n a constant density. The corresponding effective density for the part ic les, calculated us ing equation 3.2 and assuming a pr imary part icle density o f 1.2 g/cm and a m a x i m u m diameter o f 40 n m , is s h o w n i n F igure 3.6. U s i n g this re lat ionship and the number o f part icles i n a g iven size range recorded by the S M P S , it was poss ib le to estimate a total mass o f part icles o f a g iven mob i l i t y diameter range: (3.3) v 2 y where N(fa) is the number o f particles counted at that mob i l i t y diameter. B y s u m m i n g this parameter over the fu l l range o f part icle sizes measured, an estimate for the total mass o f part ic les smal ler than 150 n m i n mob i l i t y diameter was developed. Integrating the S M P S measurements and compar ing them to the T E O M data provides an ind icat ion o f the capabi l i ty o f the S M P S to resolve trends s imi la r to those observed by the T E O M . The results o f the integration for a l l the data points for w h i c h both T E O M and S M P S data were co l lected for three different fractal d imensions are s h o w n i n F igure 3 .7 . A strong - 4 7 -correlation between the T E O M and integrated SMPS results was observed, independent of fractal dimension. However, the fractal dimension had a strong influence on the strength of the association between the total integrated mass and the T E O M readings. This was anticipated, given that the two instruments are measuring different properties, with the total mass expected to be greater than the ultrafine particle mass observed by the SMPS. The fact that the constant density case tended to over predict the ultrafine particle mass, while the other two tended to under predict it, suggests that the appropriate fractal dimension lies somewhat below 3, but probably above 2. However, further work is required to determine this value more precisely. Another notable feature of the SMPS integration results in Figure 3.7 is that there is a zero offset of approximately 2 mg/m rawexhaust for the TEOM. This indicates that the T E O M is collecting mass, either from condensed lubricating oil or other sources, which are not in the ultrafine particle range even at the lowest P M levels. Although a total mass can be calculated from the SMPS results, this value is highly uncertain due to the uncertainties in the assumed equivalent density. As a result, in the remainder of this thesis, integration of the SMPS results will be used to provide a mobility volume. This value provides a representation of the total volume (which correlates with mass) of ultrafine particulate observed; however it cannot be used for direct mass-basis comparison. The mobility volume will provide a reasonable indication of the trends in total ultrafine particle concentration, allowing further insight to be gained from these measurements. In general, comparison of the SMPS and T E O M results indicate that, although integration of the SMPS results may not give a mass measurement which can be compared directly with the T E O M results, the mobility volume can be used to investigate P M emissions at levels below the observation threshold of the T E O M . Black-Carbon Fraction Another parameter of interest in measuring the particulate emissions from engines is the amount of the P M which is composed of soot (or black carbon) compared to the amount originating from condensed volatiles. In the current work, an Aethaelometer (Magee scientific, AE21) was used to measure the black-carbon content of the P M . First developed in 1984 [165] the Aethalometer is based on the fact that the black carbon (BC) component of P M deposited on a filter reduces the intensity of visible light based on the Beer-Lambert law: I = Ioe"^, (3.4) -48-where / is the intensity of the attenuated light, I0 the intensity of the incident light, b a b s the absorption coefficient and x the distance that the light has passed through the sample [166]. In the Aethalometer, the particles are deposited on a prefired pure quartz filter (Pallflex Q250F). A light is then shone at these collected particles, and the percent attenuation (normally given by ln[.W0]) can be measured. By measuring the attenuation of the light intensity, the quantity of BC deposited on the filter can be determined [167]. The Aethalomter used in the work reported here was equipped with two light sources, one at 880 nm (in the visible range) and the other at 370 nm (in the near-ultraviolet range). The two wavelengths are typically used to identify black carbon (visible light) and the volatile concentration (ultraviolet). Unfortunately, as the absorption of light in the ultraviolet regime by volatile organics depends strongly on the specific nature of the volatile species collected, a quantitative measure of the volatile concentration cannot be determined. As a result, only the black-carbon concentration is reported in the current work. One characteristic of the Aethalometer is that over a finite sampling time, the light absorption of the deposited particulate will change because of the increased path length and scattering by collected particles. As a result, a series of post-analysis corrections are required to generate meaningful results from this unit [167]. Further information regarding the functioning of the Aethalometer, the application of the required correction factors, and the main uncertainties in the analysis procedure are provided in Appendix 2. 3.3 Experimental Parameters The conditions under which the single-cylinder engine is operated have been shown to have a substantial impact on the engine's performance and emissions [38,122]. The parameters used to describe the operating condition of the single-cylinder engine included measures of power, overall mixture stoichiometry, combustion timing, and intake dilution. To indicate power, the brake torque (torque at the engine's output) could be used, calculated as: T =T -T , (3.5) brake dyno veclordriw where Tdyno was the torque absorbed by the dynamometer and TvectordnVe was the torque being supplied by the external motor. The engine's high internal friction led to brake torques which were often negative, and were always substantially lower than for an equivalent multi-- 4 9 -cylinder engine at equivalent operating conditions [168]. As a result, the brake power, calculated from the torque by [28]: ^brake ^TUNT.^f)raj(e 3 where N is the engine speed (in revolutions per second), was found to be a poor indicator of the engine's load condition. As a result, a more independent measure of load was required. 3.3.1 In-Cylinder Conditions To provide a more reliable representation of engine performance, the in-cylinder pressure trace may be integrated to provide the net work done on the piston over the engine cycle. The engine's friction is thereby excluded, providing a more repeatable and representative measure of the single-cylinder engine's performance. By integrating the pressure trace over the compression and power strokes only, the engine's power and torque were defined on a gross-indicated basis. The gross indicated work per cycle (WGI) was calculated as suggested by Heywood [28]: per cycle jpdV, (3.7) comp, expansion where p was the in-cylinder pressure at a given crank angle and V was the instantaneous cylinder volume at the same crank position. By including the intake and exhaust strokes, the net indicated work could also have been calculated. However, the work in these strokes was primarily a function of the conditions in the intake and exhaust manifolds. Increased measurement inaccuracies due to the small difference in pressure over these strokes led to higher uncertainties in the calculations for the pumping work than for the gross-indicated work. As a result, the base unit used in this thesis to define the engine's power was the gross indicated work, which was used to calculate the gross-indicated power (Pgross) from: p _ ^Gl per cycle- ^ ? (3.8) 2 where the 2 was needed to convert N from revolutions/second to cycles/second for a 4 stroke-cycle engine. Normalizing by the cylinder displacement volume (VJ) and engine speed gave the gross indicated mean effective pressure (GIMEP): P -2 GIMEP = — , (3.9) N-Vd y } -50 -which was used as the base representation of engine power as it is independent of speed and cylinder volume, thereby allowing for easier comparison with other heavy-duty engines. The in-cylinder pressure was also used to estimate the heat-release rate (Q n et), as given by: *Q™=^p*L+J-v&, (3.io) dO y-\ dO y-\ dO where 6 was the crank angle and y was the specific heat ratio (Cp/cv - assumed constant). The net heat release rate represented the rate of energy release from the combustion processes less wall heat transfer and crevice flow losses. By integrating the heat-release rate the total heat released during the combustion process (IHR) was estimated with: IHR i net v de de (3-11) By integrating up to a certain crank angle (CA), and normalizing by the total heat release, the fraction of the energy released up to that point was estimated. Typical points of interest in this work included the 5%, 10%, 50% and 90% points (i.e. the crank angle at which the IHR reaches the stated fraction of its total). Of these, the midpoint (50% IHR) was used to define the combustion timing, as it provided the most reliable representation of when the bulk of the combustion occurred, and was independent of ignition delay variability which could have influenced other combustion timing measures. Evaluation of the heat-release rate was also used to estimate the timing of the ignition process. A sample of the net and integrated heat-release rate of a typical pilot-ignited gaseous fuel combustion event is shown in Figure 3.8. The timing of the diesel ignition (pilot start-of-combustion, PSOC) has been identified as the first significant increase in the heat release rate as indicated in Figure 3.8 [131]. The timing of the gas start-of-combustion (GSOC) was typically less distinct, with a gradual increase in heat release rate being observed over a couple of crank-angle degrees. To define a reproducible 'start-of-combustion' timing for the gaseous fuel, the slope of the IHR between 30 and 70% of the total heat release was calculated. This slope was then extrapolated back to the point where it reached the 5% IHR level to represent the gas start-of-combustion. The SOC timings for each test condition were calculated, but they were also validated visually to ensure that the definitions were reliable. The commanded start-of-injection timings for both the gaseous (GSOI) and pilot (PSOI) -51-fuels are also shown in Figure 3.8. To calculate the ignition delay times, the difference between the commanded injection and observed combustion timings were used. The timing parameters of greatest interest include the gas ignition delay time (GID), defined as GSOI -GSOC, the pilot ignition delay time (PID), defined as PSOI-PSOC, and the relative combustion delay, PSOC-GSOC. The in-cylinder pressure data collection process was the same for all the test conditions. Pressure as a function of crank angle data was collected for 45 consecutive cycles. The relative pressures recorded from the piezo-electric in-cylinder sensor were referenced to the intake manifold pressure at bottom-dead-center, just before the start of the compression stroke and approximately 15° before the intake valve closed. Cyclical variability in the pressure trace, including variations in peak cylinder pressure and the GIMEP, were found to be useful measures of combustion variability when reported in the form of the coefficient of variation (COV - standard deviation / mean). In most cases, the cylinder pressure was averaged before the heat-release rates were calculated. 3.3.2 Fuel/Oxidizer Ratio The indicated power provided a reasonable representation of the power being developed in the combustion event. However, it did not provide any insight into the overall mixture stoichiometry. For a direct-injection engine, the relative proportions of fuel and oxidizer are typically represented by the equivalence ratio ((()), which is defined in Heywood However, when using EGR, the exhaust contains oxygen that is not included in this equation, was found to provide a better representation of the overall fuel : oxidizer ratio in the combustion chamber. Normally, the dilution of the intake charge by recirculated exhaust is represented by the EGR mass fraction: [28] as: (3.12) definition. As such, an oxygen equivalence ratio, where m{ oxygen replaces ma(> in the preceding -52-where msGR is the mass of recirculated exhaust (per cycle, or as a mass flow rate) and m f r e s h a i r is the amount of fresh air in the ingested charge. For an engine running overall lean (as for almost all late-cycle direct-injection engines) this is not a good representation of the charge dilution, as the recirculated exhaust contains a varying quantity of oxygen. Thus, the dilution effect for a given EGR fraction varies with the engine's equivalence ratio. A better measure of the dilution of the charge was found [131] to be the intake oxygen mass fraction, Yjnto2 (volume or mole fraction would be equally effective): m 0 2 _ M EGR ' Y 0 2 e x h + m freshair'Y02 freshair Y i n t 0 2 - - , (3-14) M EGR + m freshair M EGR + m freshair where Yo2exh represented the mass fraction of oxygen in the exhaust, which varied with load condition, and Yo2freshatr represented the mass fraction of oxygen in the fresh air (constant at 0.2295). 3 . 3 . 3 D a t a P r e s e n t a t i o n P a r a m e t e r s By defining the engine speed, indicated power, § (oxygen-based), Yjnto2, and timing (50%IHR) the operating condition of the single-cylinder engine was fully constrained. Examples of the use of these parameters to provide a complete definition of the engine operating condition have been provided elsewhere [131]. Other parameters unique to the pilot-ignited direct-injection of natural gas system used in this work included the diesel pilot fraction and the gross indicates specific fuel consumption (GISFC). The diesel pilot fraction is defined as: „, ., mpilot 'QLHV,diesel ,~. , C N %pilot = : (3.15) mpilot 'QLHV,diesel + M N G 'QLHV,NG where QLHV represented the lower heating value of the fuel on a mass basis. The GISFC provides a measure of the fuel conversion efficiency, and is calculated on a diesel energy-equivalent basis from: QLHV.NG m , + m v r pilot (VO v-v QJQp(2 — LHV,diesel , (3.16) -53-where Pgross was the gross-indicated power. Most of the emissions data were also normalised by this latter parameter, resulting in emissions units of mass per gross-indicated power. A summary of the control and output parameters is given in Table 3.2. 3.4 Repeatability and uncertainty analysis Establishing the degree of certainty in the experimental results is mandatory for assessing whether observed variations in output are actually significant. The uncertainty of the measured outcomes based on the uncertainties in the measuring instruments can be determined using a traditional calculated uncertainty analysis [147]. This procedure has been executed for most of the presented data and the results are shown in Table 3.3 (details of the uncertainty analysis procedure are given in Appendix 4). The percentage errors were found not to vary substantially with operating condition, as the published instrument accuracies were typically given in percentage values. The values in this table are actually only a representative uncertainty, as they represent simply the potential error from the various instrument readings based on the manufacturer's published values. This excludes potential contributions to overall non-repeatability of the results due either to small variations in engine operating condition or to random effects. A repeatability analysis at a single non-EGR operating condition was carried out, the results of which are included in Table 3.3 [168]. However, this form of analysis neglects the fact that the repeatability of the operating conditions, and even the variability of the engine, differs substantially between different operating conditions. To conduct similar repeatability tests at all the engine operating conditions would be prohibitively time consuming. The development of statistical techniques has allowed the pooling of experimental errors from multiple operating conditions. From this pooled error, it is possible to determine whether differences between readings are statistically significant [169]. In this work, experimental design/statistical analysis techniques have been applied as often as possible to evaluate the significance of differences within the data. In most cases, sufficient repeated data was available to provide an estimate of the uncertainty through the use of 95% confidence intervals or other statistical measures. In those cases, the uncertainty ranges in the data are represented by these confidence intervals. Otherwise, uncertainty ranges are based on the calculated measurement uncertainty shown in Table 3.3. For all the results, the sources of the uncertainty estimates are identified. -54-3.5 General Methodology The spec i f ic test condit ions used var ied for the different sections o f this work . In Chapter 4 (the effects o f in ject ion pressure), a w ide range o f engine operating condi t ions was used i n order to investigate the effects o f in ject ion pressure over the entire engine map . Th i s also serves as a demonstrat ion o f the effects o f the var ious engine operat ing condi t ions on performance and emiss ions . F o r the hydrogen/methane b lend w o r k (Chapter 5), a selected l o w - s p e e d , low-charge mass (but moderate equivalence ratio) mode was selected. Th i s was chosen to m i n i m i z e the required gaseous fuel mass f l o w rates. A l i m i t e d set o f tests was also conducted at a more representative m i d - s p e e d , h igh - load condi t ion . F o r the fuel compos i t i on w o r k (Chapters 6 and 7) a condi t ion very s imi lar to the m i d - s p e e d , h i g h - l o a d point f r o m Chapter 5 was selected to prov ide a reasonable ind icat ion o f engine condi t ions near the peak-torque operat ing cond i t ion . F ina l l y , the f i l tered E G R testing (Chapter 8) was carr ied out at a m i d - l o a d cond i t ion , where the P M loading i n the exhaust stream was h i g h enough to prov ide reasonable resolut ion o f the inf luences o f r e m o v i n g the P M . Th is cond i t ion also reduced the mass f low- rate o f E G R , result ing i n less mass deposit ion on the filter and hence longer test-t imes before the filter needed to be changed. F o r a l l the test condit ions var ious combust ion t imings were used to prov ide a range o f combust ion condit ions by w h i c h to evaluate the inf luences o f the var ious effects under invest igat ion. F o r a lmost a l l the test sets, data points were repl icated and statistical analyses were carr ied out. The nature o f the analyses used var ied somewhat, depending o n the test set. M o r e repl icat ions were used for the filtered P M work , to prov ide the best poss ib le p rec is ion for the test. Fewer repl icat ions were used for the in ject ion pressure work , due to the large number o f test condi t ions ; however , through the use o f A N O V A , it was possib le to analyze both the m a i n effects and interactions between operating condit ions and in ject ion pressure. W h i l e the unavoidable non- randomness o f the sampl ing procedure for the fuel c o m p o s i t i o n w o r k prec luded the use o f statistical techniques for data analysis , suff ic ient rep l icat ion o f the data points p rov ided a reasonable estimate o f the uncertainty i n the exper imental results. Further details regarding the test points and statistical analysis techniques used, as w e l l as the mot i va t ion for select ing these condit ions, are p rov ided i n the ind i v idua l results chapters. - 5 5 -3.6 Tables and Figures Table 3 . 1 : S ing le cy inder engine specif icat ions Engine Single cylinder 4-stroke Fuelling Direct injection, diesel pilot, natural gas main fuel Displacement 2.5 L Compression Ratio 17:1; 19:1 Bore/Stroke/Connecting Rod Length 137/169/261 mm Lubricating O i l Esso 3 X D Extra 15W-40; sulphates <1% by weight Injector Westport Innovations Inc. dual-fuel concentric needle prototype Injection control Separate diesel and C N G solenoids Injector holes 7 pilot, 9 gas Injection angle 18° below firedeck Tab le 3 .2 : S ing le cy l inder engine contro l and output parameters. Control Parameter Units Key Measurements Pilot fraction % Diesel mass; Time; Gaseous fiiel flow GIMEP bar In-cylinder pressure; Crank angle; Engine parameters 50%IHR °CA In-cylinder pressure; Crank angle; Engine parameters V i n t 0 2 Air flow; Exhaust C O 2 ; Intake C 0 2 <|>02 Air flow; Exhaust C 0 2 ; Intake C O 2 ; Exhaust 0 2 ; Gaseous fuel flow Output Parameter Units Key Measurements GISFC 8(diesel equivalent/ GikWhr Gaseous fuel flow; Gross indicated power Emissions g/GikWhr Emissions measurement; Gross indicated power; Air flow rate Ignition Delay °CA In-cylinder pressure; Crank angle C O V GIMEP % GIMEP - 5 6 -Table 3.3: Calculated uncertainty and repeatability analysis for key parameters Calculated Repeatability Parameter Units Uncertainty Uncertainty [168] Principal Sources of Uncertainty GIMEP bar 2% 5% In-cylinder pressure; Crank angle Yjnt02 5% N/A Air flow; Exhaust C 0 2 ; Intake C 0 2 <(>02 5% N/A Air flow; Exhaust C02; Intake C02 GISFC g/GkWhr 2% 2% Gross power; Gaseous fuel flow CO g/GikWhr 5% 10% CO analyser; Air flow C 0 2 kg/GikWhr 5% N/A C 0 2 analyser; Air flow N O x g/GikWhr 4% 5% N O x analyser; Air flow HC g/GikWhr 4% 10% HC analyser; Air flow P M mg/GikWhr 8% N / A TEOM; Air flow To Emissions To Exhaust B e n c h l ; Pressure Relief Wastegate Exhaust* Insulated Sample Line Back-pressure Valve From Air Compressor Shut-off Valve EGR Cooler <;> From Ambient ~~*—, Venturi Insulated!! Sample JJ I 1 Line Jj Exhaust Surge Tank EGR Valve T o p$j S y s t e m Intake Surge Tank Filters \ Pressure Z1 Regulators Figure 3.1: Air Exchange System Layout O Engine Figure 3.2: Westport HPDI™ Injector Schematic - 5 7 -Compression Stroke Top-Dead-Center — • Expansion Stroke Pilot Inject 1 Duration on Relative Injection Timing + • Gas Injection Duration Crank Angle Pilot Start- Pilot End-of- Gas Start-of-lnjection Gas End-of-lnjection of-lnjection Injection Figure 3.3: Westport HPDI™ injector commanded injection timings. Raw Back-Pressure Regulator Insulated Sample Line Heated ^ Mixing Orifice Vent to Atmosphere . Gravimetric TEM Sampler n | t e r s I X Vent to Atmosphere T JL Rotameter * Heated Mixing Section Exhaust TEOM Bottled Nitrogen Secondary Dilution System Pump, Filters, & Flow Meters Aethalometer <H Figure 3.4: P M sampling and dilution system schematic. Gravimetric PM (mg/hr) Gravimetric PM (mg/hr) Figure 3.5: Correlation between T E O M and gravimetric filter results. The plot on the left includes P M > 2 mg/hr (measured by the gravimetric filters), while the plot on the right is only P M < 2 mg/hr. -58-10 w c 0 Q CD > o CD t t UJ 0.1 constant density .fractal dimension = 2.5 fractal dimension = 1.8 10 100 Mobility Diameter (nm) 1000 Figure 3.6: Assumed effective density of SMPS particles as a function of mobility diameter, for three different fractal dimensions (d/m = 3 for constant density). Assumed size of primary particles is 40 nm. Based on work from Virtanen et al. [159]. 0 5 10 15 0 5 10 15 TEOM mass (mg/m3raw exhaust) TEOM mass (mg/m 3 r a w e xhaust) Figure 3.7: Mass concentration comparison between T E O M measurements and calculated ultrafine particle mass based on SMPS measurements. Left; Ultrafine particle mass calculated based on effective particle density calculated from three fractal dimensions (3, 2.5, 1.8). Right: Ultrafine particle mass for fractal dimension of 2.5 only. -59-Figure 3.8: Typical net and integrated heat-release rate plots, showing the commanded start-of-injection and observed start-of-combustion timings for the pilot and gaseous fuels. -60-Chapter 4 Injection Pressure 4.1 Introduction The combustion process in a PIDING engine depends on both the reaction kinetics and on the fluid mechanics of the injection and mixing processes. The fluid mechanics of the gaseous jet prior to the start of combustion are highly dependent on the properties and structure of the injected fuel. The characteristics of the medium into which the fuel is being injected, including composition (presence of recirculated exhaust gases), density (temperature, pressure), and motion are also critical. These factors can have a substantial influence on the initiation of the combustion event, its propagation, and its eventual termination. Key parameters include the differential pressure across the injector nozzle, the kinetic energy of the jet, the quantity of air entrained into the jet, the distance that the jet penetrates into the combustion chamber, and the turbulent strain rate in the air where the fuel and oxidizer are mixing. This chapter aims to elucidate the effects of injection pressure on performance and emissions over a wide range of in-cylinder conditions. From these results, further insight into the combustion and pollutant formation processes in the PIDING engine may be derived. 4.2 Injection Pressure Influences Increasing the injection pressure in diesel engines generates significant reductions in P M emissions. These reductions are attributed primarily to improved, atomization and enhanced air entrainment, resulting in a leaner mixture in the core of the spray, and hence lower carbonaceous particle (soot) formation [170,171,172]. Conversely, in-cylinder studies suggest that soot formation may actually increase, due to the more rapid evaporation of the fuel; however, the improved mixing also increases soot oxidation rates, resulting in a net reduction in the engine-out soot [173,174]. A modified version of this chapter was presented at the A S M E Internal Combustion Engine Division Fall Technical Conference, 2005. The paper is also in press for publication in the A S M E Journal of Engineering for Gas Turbines and Power. McTaggart-Cowan, G.P., H.L. Jones, S.N. Rogak, W.K. Bushe, P.G. Hil l and S.R. Munshi. The Effects of High-Pressure Injection on a Compression-Ignition, Direct Injection of Natural Gas Engine. A S M E Technical Paper ICEF2005-1213. 2005. A l l tables and figures reprinted with permission from A S M E , © A S M E 2005,2006. -61-One drawback to the use of higher injection pressures in diesel engines is higher N O x emissions. Increases as great as factors of two or three are reported in the literature, although this magnitude varies substantially with engine type and operating condition [171,174]. In-cylinder gas sampling indicates that the concentration of the H C N radical is substantially increased; this radical is an important precursor in the prompt-NO mechanism [175]. However, in general the increase in N O x is primarily attributed to higher combustion temperatures induced by the more rapid mixing and evaporation of the fuel spray, leading to a larger initial pre-mixed combustion event. On the other hand, some researchers suggest that at high EGR fractions, N O x is relatively insensitive to injection pressure [176]. One confounding effect in most of the reported studies is that the injection timing was fixed: because of higher injection rates and shorter ignition delays, the combustion event occurs earlier in the cycle. That these reductions in ignition delay are not as significant with EGR may explain why the injection pressure has less influence on N O x under high EGR conditions. In general, more advanced combustion substantially increases N O x emissions [136], a result of higher in-cylinder temperatures which lead directly to more NO formation via the Zeldovich thermal NO mechanism (see section 2.2.2). To avoid this influence, in investigating the effect of injection pressure it is important to fix the combustion, rather than injection, timing. As mentioned above, higher injection pressures tend to reduce the ignition delay time of diesel fuel. The ignition delay of a non-premixed combustion event can be separated into a physical delay (the time required for the spray droplets to mix with the air, evaporate, and reach a combustible mixture) and a chemical delay (the time required for the break-up of long-chain hydrocarbons and the formation of radicals in the pre-combustion reactions) [171,176]. For a liquid fuel, increasing the injection pressure tends to reduce the physical delay through improved atomization and mixing. However, it does not significantly affect the chemical delay, which is more dependent on the in-cylinder conditions, including temperature, pressure and oxygen concentration [42,174]. Experimental results suggest that, for an auto-igniting natural gas jet, higher injection pressures tend to reduce ignition delay' times primarily due to a reduction in the physical delay [89]. In a PIDING engine, the gaseous jet is not auto-igniting, but rather is ignited by the burning diesel; as a result, the -62-relative locations of the gaseous jet and pilot spray in the combustion chamber have a strong influence on the delay between the injection and the ignition of the gaseous fuel. The reduced ignition delay time has an effect on the rest of the combustion process, as do the increased injection rate and the more rapid mixing induced by higher shear stresses between the jet and surrounding charge. In general, higher injection pressures tend to increase the intensity of the premixed combustion event. This is somewhat counter-intuitive, given that the ignition delay time has been reduced; however, the unaffected chemical delay combined with the higher injection rate and enhanced mixing result in more fuel having mixed to a combustible level before ignition occurs [171]. The effects of injection pressure depend strongly on the engine operating condition. In particular, increased engine speeds result in enhanced squish-related charge motion and mixing, thereby reducing the absolute physical ignition delay (in ms). Increases in the cylinder pressure and temperature also act to reduce the chemical delay [171,174]. Higher load conditions will typically lead to longer combustion durations, which result in more P M formation for both diesel and PIDING engines [38,170]. With the same injector, higher injection pressures will typically reduce the overall injection duration. Particulate size distributions are also substantially influenced by injection pressure. For heavy-duty diesel engines, early timings combined with high injection pressures reduce total particle mass but increase the number of nucleation mode particles [172]. Under certain conditions, some researchers suggest that there is a very strong shift from primarily accumulation-mode (particle mobility diameter, D p , > 50 nm) to mainly nucleation-mode (D p < 50 nm) particles [172]. This effect is most pronounced at light load; at higher load conditions, increases in nucleation mode particles are not observed while fewer accumulation mode particles are emitted [172]. Higher engine speeds, which correspond to higher turbulence and potentially more enhanced particulate agglomeration, also reduce the magnitude of this shift towards smaller particles [177]. These effects have been attributed to increased concentrations of volatiles species at high injection pressures. Furthermore, the particulate surface area on to which these volatiles condense is reduced substantially (surface area scales with the square of particle diameter). As a result, higher volatile concentrations are present in the exhaust stream; under these conditions the volatiles may be self-nucleating, -63-generating the substantial nucleation-mode particle concentrations observed experimentally [178]. For a PIDING engine, increasing injection pressure will most likely reduce the physical component of the pilot ignition delay (where auto-ignition occurs). Due to fundamental differences between a high-pressure evaporating liquid spray and an underexpanded gaseous jet, the influence of injection pressure in the gaseous combustion event may differ substantially from the effects observed in diesel engines. Rubas et al. [77] and Hi l l and Ouellette [84] have studied the fundamental effects of injection pressure on gaseous jets under engine-like conditions. In both studies, the gas jet is considered as a two-part system, with a traveling vortex ball being supplied mass and momentum from a quasi-steady jet, as discussed in section 2.2.1. Increasing the injection pressure increases the momentum transfer and the gas density relative to the cylinder charge. This increases the jet penetration distance; however, the increase is only proportional to the lA power of the change in density (i.e. doubling the density of the jet will only increase the penetration by 20%) [84]. Because kinetic energy increases with the square of velocity, increasing the injection velocity with a higher pressure differential between the fuel jet and the combustion chamber leads to a higher total kinetic energy transfer to the charge. This results in increased in-cylinder turbulence even after the end of the injection process. Varying the injection pressure, in conjunction with other combustion modifications, has been used as one of a number of techniques to optimize the performance of PIDING engines over a full test cycle [40,137]. These previous studies made no attempt to systematically identify the influence of injection pressure independent of other parameters. Preliminary studies suggest that, at one operating condition, increased injection pressures reduce P M with little effect on N O x emissions or other engine performance measures [131,135]. The focus of the current work is to further investigate the effects of increased injection pressure over the full range of engine operating conditions. 4.3 E x p e r i m e n t a l M e t h o d o l o g y The effects of injection pressure on a PIDING engine were investigated using the single-cylinder research engine described in Chapter 3. The combustion chamber used in this work was from the baseline 1998 model ISX engine, with a 19:1 compression ratio. The fuelling system was configured to provide rail pressures of up to 31 MPa for both the diesel -64-and the natural gas. F o r the particulate emiss ions measurements, T E O M and grav imetr ic f i l ter measurements o f the total mass as w e l l as part iculate s ize distr ibut ions were recorded. 4.3.1 Experimental Conditions One o f the p r inc ipa l objectives o f this study was to evaluate the inf luences o f in ject ion pressure over a w ide range o f engine operating condit ions. A s a result, testing was conducted over the fu l l range o f engine operating condit ions accessible to the s ing le -cy l inder research engine. The operating condit ions, based on the parameters def ined i n sect ion 3 . 3 , are ident i f ied in Table 4 . 1 . Other relevant parameters that were held constant were combust ion t i m i n g ( 5 0 % I H R at 1 0 ° A T D C ) , the separation t ime between pi lot and gas inject ions (1.0 ms) , and the p i lo t in ject ion mass (5 mg/injection). The constant p i lo t f l o w resulted i n a p i lo t f ract ion that var ied (on an energy basis) f r o m 1 1 % at l o w load to 3 % at h i g h load . T o investigate the effects o f in ject ion pressure over the entire engine operating map , a complete ly randomized exper imental design was developed. The design was based on a repl icated factor ia l analysis (wi th two independent samples per test condi t ion) ; this resulted i n eight test points per speed/load operating condi t ion and a total test matr i x o f 48 independent test points. A further set o f tests was carried out at the peak-torque cond i t ion (point #5 i n Tab le 4.1) . F o r these tests, the in ject ion pressure was var ied over a w ider range (four va lues , f r o m 18 to 30 M P a ) ; this was intended to prov ide in fo rmat ion o n the l inear i ty o f the in ject ion pressure inf luences. For this second set o f tests, two charge d i lu t ion condi t ions ( n o n - E G R , Yjnto2 = 0 . 2 3 ; and moderate E G R , Yjnto2 = 0.19) were compared . F o r the E G R case, the combust ion t i m i n g was also retarded ( 5 0 % I H R delayed f r o m 10° to 1 7 . 5 ° A T D C ) to evaluate the effect o f in ject ion pressure o n enhancing the late-stage combust ion process. A n analysis o f variance ( A N O V A ) was used to evaluate the statistical s ign i f icance o f the measured results. T o avo id high-order interactions between engine speed and load , two separate A N O V A s were conducted; one for the three condit ions where load was constant (3, 5 and 6) and one where speed was constant (1 , 2 and 3). These two analyses independently ident i f ied the effect o f speed (with in ject ion pressure and E G R level) and load (also w i t h in ject ion pressure and E G R level) . The general representative m o d e l is g i ven by: 7 „ =p + OCl+Yint02 . + Pinjk + OCYint02v + OC • Pinjik + ^ ^ Pinj • YmW2Jk + OC • Pinj -Ymt02uk + em, - 6 5 -In this equation, OC represents the effect of operating condition - either load or speed (whichever was varying). Yyki is the measured response (emissions, fuel consumption, etc.), /u represents a general effect for the whole experiment, OC, Yint02, and Pinj the main effects (combinations represent the interaction terms), and s ^ i represents the error 'nested' within the condition (that is, the variation between the repeated values at each test condition). In carrying out the analysis, an a-level of 0.05 (1 in 20 chance of identifying a significant effect when one was not present) was used. Further details on the model development and analysis procedure can be found in statistical design text-books (e.g., Hicks and Turner [169]). 4.4 Results The influences of the various engine operating condition parameters on PIDING engine performance and emissions interact in a statistically significant and highly non-linear fashion [130]. Therefore, any analysis that includes multiple operating condition parameters will routinely exhibit third-order and higher interactions, making interpretation of the results difficult. In this work, the primary interest in the effects of operating condition is how they relate to the effects of injection pressure. As a result, it is more informative to analyze the effects of injection pressure as a function of speed and load independently. 4.4.1 Effect of Load Investigating the interactions between injection pressure and load involved three operating conditions at constant speed (800 R P M ; Table 4.1, conditions 1, 2, 3). Equation 4.1 provides the statistical model used for this analysis. The A N O V A results are shown in Table 4.2. Significant associations (P-value < 0.05) are shown in the shaded cells. Degrees of freedom (DF) represent the number of unconstrained data points used in calculating the influence of the parameter. The A N O V A results indicate that for most emissions, load interacts significantly with both the injection pressure and the intake oxygen mass fraction. (Recall that an interaction term indicates that the response of the dependent variable to one independent variable varies for different levels of a second independent variable; i.e., if dependent variable C responds differently to a change in parameter A at different levels of parameter B, then there is an interaction effect between A and B on C.) As the results include significant second-order interaction terms, the lower order terms (main effects) may not be accurately represented by -66-the A N O V A [169]. In these cases, it is necessary to evaluate the results graphically to identify whether consistent trends are present. The results indicate that at fixed speed, the GISFC is independent of injection pressure or intake dilution (EGR); only load has a significant effect. The effects of load and intake dilution on combustion duration interact, due to the longer injection durations at higher loads (more fuel to inject at a given injection pressure), while intake dilution tends to reduce the combustion rate. The fact that injection pressure is a significant factor suggests that it directly affects the combustion duration; however, this result is not statistically reliable given the higher-order interaction between load and intake dilution. The analysis of CO and N O x emissions also displays significant second-order interactions between injection pressure and load and intake dilution and load. The fact that no significant interaction between injection pressure and intake dilution is present indicates that the response to injection pressure is consistent with and without intake dilution. Interestingly, while the HC emissions (>95% unburned CH4) show an interaction between load and intake dilution, there are no interactions with injection pressure. This indicates that the effect of injection pressure is independent bf both load and intake dilution. The fact that injection pressure does not statistically influence HC emissions (at either interaction or main-effect level) suggests that injection pressure does not have a consistent influence on unburned fuel. Not surprisingly, given their complicated formation mechanisms, the particulate matter mass emissions show a third-order interaction between load, injection pressure, and intake oxygen mass fraction. Further interpretation of the P M results must be conducted through graphical means. 4.4.2 Effect of Speed Investigating the effect of injection pressure over a range of engine speeds (RPM), involved three test points with constant engine load & equivalence ratio (conditions 3,5, and 6 from Table 4.1). The results of this A N O V A are shown in Table 4.3, with an a-level of 0.05. As with the load results (Table 4.2), higher-order interactions are present for most of the measured parameters. As in the load effect analysis (section 4.4.1), the GISFC is independent of injection pressure; it depends only on speed and intake dilution independently. -67-Engine speed, intake dilution, and injection pressure all interact for the C O and H C emissions. A s a result, evaluation of the lower-order interactions and main effects must be deferred to a later section. N O x emissions vary with speed and intake dilution, but are independent of injection pressure. Unlike the load-effects analysis, P M emissions depend only on the main effects of speed, intake dilution, and injection pressure; no significant interactions occur. While the fact that these factors are significant is not surprising, it is interesting that no higher-order interactions are significant. This indicates that the effects o f injection pressure on P M are consistent, at high load, over the full range of speeds and intake dilution conditions. In general, the A N O V A results indicate that operating condition has a strong effect on the measured results, while the injection pressure tends to interact with speed and load but not as commonly with intake dilution ( E G R level). These second-order interactions limit the statistical evaluation of the significance of the main effects; graphical interpretation is required to provide further insight into the influences of injection pressure. 4.4.3 Effects of Operating Condition A s shown in the preceding statistical analyses, the emissions and performance of the PIDING engine depend strongly on operating condition. Most of the dependent variables show substantial interactions between operating condition, intake dilution, and injection pressure. To provide further insight, the results are presented graphically in Figures 4.1-4.5, with each dependent parameter plotted, individually. Each plot includes all six operating conditions (speed/load), while the individual bars on the plots represent the low and high injection pressure cases, with and without intake dilution (EGR) . The values shown are averages of the two values collected for each test point. Error bars represent the calculated uncertainty (Table 3.3). The results are also presented in tabular form in Table 4.4. This table also includes the pressure ratio between the injection pressure and the peak cylinder pressure (Pinj/Pmax) for each condition. This pressure differential indicates that for the low-load cases, the pressure ratio is as much as five times greater than the cylinder pressure; however, at the high-speed, high-load, low injection pressure condition, the pressure in the fuel rail is only 30% higher than the peak cylinder pressure. Furthermore, as the pressure drop through the injector is estimated at 15-25% of the rail pressure, the actual differential pressure across the -68-nozzle outlet may be quite small. As will be discussed further, this has a substantial influence on the effects of injection pressure at different operating conditions. Injection pressure has an insignificant effect on GISFC, as shown in Figure 4.1; this agrees with the A N O V A results in Tables 4.2 and 4.3. Of the other parameters, load has the largest influence, with optimum efficiency (lowest GISFC) at the mid-load cases. Higher speeds (at equivalent load) also tend to improve efficiency. The fact that GISFC reaches a minimum at part load for a given speed agrees with basic compression-ignition understanding, due to the trade-off between quasi-constant friction losses and increasing heat transfer [28]. The high HC emissions seen at the low load case also tend to reduce efficiency. No statistically consistent influence of injection pressure on N O x emissions is observed (Figure 4.2, Tables 4.1 and 4.2); the emissions are, however, sensitive to operating condition. Intake dilution reduces N O x emissions due to the reduction in combustion temperature, as has been demonstrated in previous work [122,130,135]. Also consistent with previous results, higher speeds reduce power-specific N O x emissions [130]. This may be attributed to the higher turbulent intensities at higher speeds, which reduce the time the burned gases are at high temperatures before they mix with cool charge. This results in less time for the relatively slow thermal NO mechanism (section 2.2.2) to approach equilibrium, thereby reducing NO formation. Higher injection pressures tend to increase N O x without EGR (except at 1600 RPM), as suggested previously [135]; however, the magnitude of the effect varies substantially with operating condition. The fact that N O x emissions depend primarily on the peak combustion temperature has been demonstrated previously [122]. The peak temperature relates to the intensity of the initial combustion event, and is directly related to the rate of pressure rise in the combustion event. Higher injection pressures increase the maximum rate of pressure rise at all conditions except the high-speed (1600 RPM) case. This suggests that at 1600 R P M the ignition and early-combustion phases are kinetically (rather than mixing) limited; as a result, a higher injection pressure does not significantly increase the combustion rate, resulting in no significant change in the combustion temperature for the early combustion stages. This effect may explain the reduction in N O x emissions with increased injection pressures without EGR at 1600 R P M , a finding contrary to the results at the other non-EGR conditions. -69-With E G R , the N 0 X results show no consistent dependence on injection pressure. Only at mid-load (0.45<j), at both 800 and 1200 R P M ) are N O x emissions significantly increased at higher injection pressures. The particularly large increase at 800 R P M is due to test condition repeatability, where both low-injection pressure cases had, by chance, intake oxygen mass fractions of approximately 0.185, compared to 0.191 for the high injection pressure cases. The extreme sensitivity of N O x to intake dilution has previously been demonstrated [131]. This was the only condition where the difference in intake oxygen concentration is greater than 0.003; as a result, this effect should not influence the accuracy of the results at any of the other conditions. It does, however, mean that the results at this condition (mid-load, low-speed with intake dilution) are somewhat less certain than would be desirable; it also demonstrates the importance of ensuring operating condition consistency. The statistical results indicate that the CO emissions depend on injection pressure, intake dilution, and operating condition; however, the statistical significance of these parameters individually cannot be evaluated due to high-order interaction terms. A plot of the CO emissions, Figure 4.3, demonstrates these complex interactions. In general, the CO emissions are greatest at the highest loads. Intake dilution doubles the CO emissions at almost all the operating conditions, which agrees with previous results [122]. At low speed (and mid-speed mid-load) higher injection pressures tend to reduce CO emissions. However, at high load at intermediate speed, the CO emissions are relatively unchanged by higher injection pressure, while at high speed the CO emissions are significantly increased. This non-linear behaviour of CO with injection pressure may be due to the competing influences of enhanced mixing and increased turbulent shear stresses as the injection pressure increases, as will be discussed further in section 4.4.4. Similar to the CO emissions response, the effects of injection pressure on HC emissions, Figure 4.4, vary with operating condition. At all the conditions, the HC emissions are primarily unburned methane. Hydrocarbon emissions are, on a power-specific basis, highest at low-load. This may be at least partially a result of unburned fuel from fixed-volume sources such as the nozzle sac; the amount of fuel which is retained during the combustion event, and is then released late in the combustion cycle, will be essentially constant on a per-cycle basis, leading to higher power-specific emissions at low power levels. Another factor, which may be contributing to the high HC emissions at low load, is over--70-leaning of the injected fuel prior to ignition of the gas jet by the pilot. One possible influence of higher injection pressures is more over-leaning of the fuel, resulting in higher HC emissions. Increases in HC emissions, potentially attributable to this effect, occur at many operating conditions, including most of the higher speed and EGR test conditions. However, at a number of conditions increasing the injection pressure reduces HC emissions; this may be due to changes in gaseous fuel ignition delay time, which will be discussed in section 4.4.5, as shorter ignition delay times will tend to result in less fuel having overmixed before ignition occurs. Other factors, including increased turbulent strain leading to more local extinction events may also be contributing to the HC emissions. The variability in the HC emissions with operating condition demonstrates the complex interactions between these various potential mechanisms. At all the operating conditions the effects of injection pressure on HC emissions are secondary to the influences of operating condition and intake dilution. Injection pressure has a strong and relatively consistent influence on P M emissions, as shown in Figure 4.5. The P M emissions are also affected by intake dilution and operating condition, as demonstrated in Table 4.2. At low- and mid-load at low speed, P M emissions are relatively insensitive to intake dilution or injection pressure, although there is an apparent slight reduction in P M with injection pressure at low load. At high load, the P M emissions at the low injection pressure case are substantially higher; increasing the injection pressure results in reductions in P M emissions by a factor of almost three both with and without intake dilution. Similar influences are seen at 1200 and 1600 R P M . In general, at low-load P M emissions are relatively low, and increases in intake dilution do not substantially increase them. This is consistent with previous results that have shown that most of the P M emissions at low load conditions are volatiles [142], probably formed from evaporated (but uncombusted) lubricating oil and liquid fuel which condense in the exhaust stream. While increasing the pressure of the diesel pilot will likely influence the volatile concentrations attributable to the pilot spray, much of the pilot contribution to the volatiles may be originating from diesel fuel leaking from the injector sac volume late in the expansion stroke, after the combustion has terminated. These species would be emitted unreacted, as the cylinder temperature would be too low to fully oxidize them; this process would not be substantially influenced by injection pressure. At higher load conditions, and especially with -71-higher intake dilution, the natural gas contribution to the P M emissions is substantially greater than the volatile fraction [142]. The observation that injection pressure significantly reduces solid carbon P M (soot) formation is consistent with results from diesel engine research [170-176]. In diesel engines, the most commonly proposed mechanism is increased air entrainment into the fuel spray, which leans out the soot-inducing fuel-rich reactions occurring in the core of the spray as it evaporates, as proposed in the phenomenological model of Flynn, Dec et al. [102,179]. While a diesel spray combustion model does not completely describe the combustion of pilot-ignited natural gas jets, higher gas velocities are expected to increase air entrainment just downstream of the injector. Based on shock-tube experiments it is known that the natural gas flame is lifted from the nozzle, and parameters that increase air entrainment (i.e. reduction in nozzle hole size) decrease soot formation [89]. The pilot combustion has a strong influence on the gas combustion, but at high load, most of the soot is derived from the natural gas [142]. Thus, soot emissions from the natural gas engine may be being reduced by mechanisms similar to those found in a conventional diesel engine. A n alternative explanation of soot emission reduction is that with higher injection pressures, less fuel is burned late in the cycle, where the falling temperatures will lead to slower soot oxidation. In-cylinder heat release rates suggest that this may not be the case (see below); this is an area of ongoing research involving in-cylinder turbulence enhancement through combustion chamber optimization. 4.4.4 Injection Pressure Effect Details In the preceding section, a range of operating conditions was investigated at only two injection pressures (21 and 30 MPa). As such, any non-linear effects of injection pressure would not be identified. A more detailed study was carried out at the near-peak-torque operating condition (test condition #5 in Table 4.1). A wider range of injection pressures (18-30 MPa) were tested, with and without intake dilution. For the diluted case, both the standard combustion timing (combustion mid-point at 10°CA after top-dead-center) and a late-timing case (combustion mid-point at 17.5°CA) were studied. In conducting the tests, variations in the operating condition were minimized by restricting the randomization of the tests. Due to this restriction, the results were not analyzed statistically. The operating condition was set (including intake oxygen mass fraction), then -72-the injection pressure was varied (in random order) over the testing range. While this procedure allows greater precision in studying the individual effects of injection pressure, the accuracy of comparisons between the operating conditions is reduced. In general, the trends in emissions and performance for the three conditions, shown in Figure 4.6, are comparable to those in the earlier tests. For all three cases, a generally linear reduction in P M emissions with increased injection pressure is observed; in all cases, the reduction between the lowest and highest pressures is approximately 50%. The effects on N O x and HC are relatively small, with N O x being slightly increased and HC slightly reduced with increased injection pressure. There are no consistent trends in GISFC. The CO emissions, however, show a substantially non-linear response to injection pressure. At all three conditions, a maximum is reached between 22 and 26 MPa injection pressure. The strong influence of operating condition on CO emissions (Figure 4.3) may be explained by this response, as at different operating conditions the location of the peak CO emissions levels varies substantially. At 1600 R P M , the peak may be shifted towards higher injection pressures, resulting in the observed increase between 21 and 30 MPa. This is most likely due to changes in in-cylinder turbulence intensity, as well as the increased overall combustion duration (in °CA) for the higher speed condition. Comparisons between operating conditions suggest that dilution tends to somewhat increase CO, HC, and P M emissions, while substantially reducing N O x and slightly reducing GISFC. Delaying the combustion for the diluted condition results in a further increase in CO and HC emissions, as well as an increase in GISFC; however, N O x and P M emissions are substantially reduced. At the lowest injection pressure, the late timing results in significantly higher P M , due possibly to the low pressure-ratio between the cylinder and the injected fuel. At this operating condition, it would appear that the general trends in emissions and performance due to changes in the injection pressure are relatively independent of intake dilution and combustion timing. However, the specific levels of the emissions vary substantially. 4.4.5 In-Cylinder Performance The in-cylinder pressure measurement provides an indication of the influences of injection pressure and operating condition on the combustion event. The gaseous fuel ignition delay, shown in Figure 4.7, is in general reduced with higher injection pressures, -73-most likely due to enhanced mixing of the injected fuel with the charge air. This suggests that the mixing is limiting the gaseous fuel ignition process over a wide range of operating conditions. The fact that the effects of injection pressure on ignition delay are relatively small (on the order of 5-15%); suggests that other mechanisms (for example chemical kinetics) are also posing a limitation on the ignition process. Variations in the mechanical injector operation, due to changes in fuel rail and cylinder pressure, could also be affecting the observed results. However, increasing the rail pressure from 24 to 28 MPa has been shown to shorten the mechanical injector delay by less than 0.13°CA, indicating that the mechanical effects are relatively insignificant [180]. The combustion duration, Figure 4.8, (measured as the crank angle range between the 5% and 90% integrated net heat release points) is reduced with injection pressure. The fact that the rate at which the fuel is introduced into the combustion chamber has a significant impact on the combustion duration indicates that the overall duration of the combustion process is at least partially controlled by the mixing rate. As such, any method which increases the injection rate will tend to reduce the combustion duration. For most cases, intake dilution reduces the combustion duration, primarily due to the longer ignition delay (Figure 4.7), which results in more of the fuel being at a combustible mixture prior to ignition. As the combustion duration is measured from start-of-combustion (approximately) to end of combustion, a later start of combustion leads to a shorter combustion duration, even if the overall process (start-of-injection to end-of-combustion) is longer. While intake dilution has been shown to increase combustion duration, this occurs only at dilution levels greater than those tested in this work [39]. Further information about the combustion event can be discerned from the in-cylinder pressure trace and its derivative, the heat-release rate. These parameters are shown for the low-load, low-speed condition (point #1, Table 4.1) in Figure 4.9. The peak HRR is higher without dilution (left plot), while the injection pressure has only a very minor influence on the combustion rate or combustion intensity. The most significant difference is that the peak heat release rate is increased with the higher injection pressures in both cases. This is a result of the more rapid injection process, which is approximately 25% faster for the higher injection pressure case. This leads to more fuel being available to burn in the initial combustion event, resulting in a shorter combustion duration and a higher peak HRR. -74-However, the fact that the combustion duration is significantly influenced by the higher injection pressure indicates that, even at this low-load, short injection duration condition, the rate of mixing between the fuel and oxidizer is still controlling the overall combustion rate. The pressure trace and heat-release rate for the mid-speed, high-load case is shown in Figure 4.10 (Appendix 6 shows the remaining pressure traces and HRR's). Increasing the injection pressure substantially increases the combustion rate, with and without intake dilution. The main-fuel combustion starts significantly (~2°CA) earlier for the lower injection pressure, but the late-phase combustion process (after 15°CA) is essentially the same. At this condition, and at the other higher-load operating conditions, the combustion duration is substantially reduced by the higher injection rate of fuel into the combustion chamber and the correspondingly more rapid mixing process. 4.4.6 Effect on particle size distributions The preceding sections included total mass measurements of particulate matter emissions. However, the accuracy of these measurements for low particle concentrations is uncertain (see section 3.2.2). Furthermore, smaller particles (especially those in the ultrafine range with a diameter <100 nm), pose significant health and environmental concerns but are poorly represented by total mass measurements [178]. As a result, an analysis of the size distribution of the emitted particles is pertinent. The size distributions presented here cover the range of aerodynamic diameters between 5 and 150 nm. On the distribution plots, the low and high injection pressures are plotted together, with the undiluted and diluted distributions shown in adjacent plots. The distributions have been adjusted for dilution ratio, such that they represent the number concentration per unit volume of raw exhaust. The particle size distributions at the low-speed, low-load condition are shown in Figure 4.11. Both with and without intake dilution, the injection pressure has very little influence on the number distribution. In both cases, there is a discernable 'nucleation-mode' peak, although with intake dilution, a significant 'accumulation mode' peak is also present. For diesel engines, the nucleation mode particles are thought to be primarily condensed liquid droplets, comprised of sulphuric acid, water, and volatile hydrocarbons that have nucleated in the exhaust stream [99]. Diesel-engine researchers report that the mode of the nucleation peak occurs at particle diameters of approximately 10 nm [181], which is similar -75-to the mode of the peak shown in Figure 4.11. This provides further evidence that the bulk of the particulate emissions at this condition are attributable to the emissions of volatile species. At the same speed but high load (point #3 in Table 4.1) the particle size distribution (Figure 4.12) has been significantly altered from the low-load case. The particle distribution is now uni-modal with no indication of a separate nucleation-mode peak. This indicates that engine-out particles are making a greater contribution to the total emissions, as these particles will tend to be larger due to a longer residence time as well as higher concentrations in the pre-dilution exhaust system resulting in greater particulate agglomeration. At the same time, any volatiles in the exhaust will tend to condense onto these particles rather than forming new nucleation mode particles. Higher injection pressures reduce both the number and the mode of the ultrafine particle distribution, both with and without EGR. The number concentration of ultrafine particles (less than ~ 25 nm in mobility diameter) is as much as an order of magnitude lower at higher injection pressures. The fact that the particles are both smaller and fewer in numbers suggest that the higher injection pressure is leaving less time for particulate inception to occur. This would be consistent with higher fuel-air mixing rates, resulting in a shorter residence time for the particle precursors to form particles in the high-temperature fuel-rich region of the combustion zone. Size distributions for the high-speed, high-load case (point #6 in Table 4.1) are shown in Figure 4.13. As at the previous condition, intake dilution tends to shift the particle distribution towards more and larger particles. However, at this speed the effect of injection pressure is significantly reduced, which coincides with the smaller (relative) influence on total particle mass shown in Figure 4.5. The reduced effect of injection pressure on the size distributions at higher speeds agrees with results from diesel-fuelled engine research [177]. This provides further evidence that the injection pressure is having less of an effect on improving the turbulent mixing process at higher speeds. For this operating condition, it is apparent that many of the particles are greater than 150 nm in diameter, and as a result are not measurable with the particle sizing instrumentation. 4.5 Discussion The results show that the injection pressure can significantly influence emissions, without impacting combustion efficiency. The effect varies with engine speed and load due to changes in the in-cylinder conditions. At low speed, the in-cylinder turbulence will be at -76-its lowest levels. As such, higher injection pressures will tend to enhance the turbulence, aiding the rapid completion of the combustion event. While this appears to be significant at high load (where the reduction in P M is very substantial), it has little effect at lower loads. The most likely reason for this is that the 21 MPa injection pressure is already substantially higher than the in-cylinder pressure, providing sufficient turbulence for the combustion to complete rapidly. Hence, turbulent mixing is not the limiting factor for the combustion event at low-load conditions, with and without EGR, at constant oxygen equivalence ratio. From the particulate size distributions, it appears that at low-load conditions nucleation of volatile hydrocarbons and sulphates is a significant contributor to the P M loading. The fact that injection pressure does not influence the size distribution suggests that it also has little influence on volatile emissions. At higher loads, the effect of injection pressure is more significant, as the in-cylinder pressure is higher. The relative change with injection pressure of the differential pressure between the combustion chamber and the fuel is also substantially greater. The higher injection pressure tends to increase the density gradient between the gaseous jet and the combustion chamber while also increasing the kinetic energy of the jet. This has a substantial influence on the combustion process, primarily through enhanced turbulence in the combustion chamber. The higher gas-jet velocity is also more likely to entrain air, which will tend to reduce the equivalence ratio in the core of the jet. As soot precursor formation is most likely occurring in the core region, the lower equivalence ratio at higher injection pressures generates fewer soot precursors, resulting in less soot formation. Soot oxidation in the late stages of combustion may also be enhanced by the higher turbulent intensity, resulting in the observed reduction in P M emissions. More rapid mixing also reduces the amount of time for formation of particle precursors in the high temperature fuel-rich region of the combustion zone. As the high equivalence ratio cases typically have a higher soot fraction (volatiles are less significant), the reduction in soot emissions leads directly to substantially lower P M mass emissions and number densities. This effect is even more significant with intake dilution, where the in-cylinder pressure is higher and the combustion duration is longer. Increased injection rates with higher injection pressures also make a significant contribution to this effect. Offsetting these influences at the higher speeds are the increases in the in-cylinder turbulence due to the more rapid piston motion and the longer -77 -in ject ion durat ion (relative to the p iston m o t i o n ; the absolute t ime for in ject ion is essential ly unchanged). 4 .6 Conclusions 1. A t h igh loads, higher in ject ion pressures substantial ly reduce both mass and number o f P M emiss ions . A t l o w load , post-exhaust nucleat ion mode part ic les dominate the P M load ing ; the mass and number o f these part icles are independent o f in ject ion pressure. The effects o f in ject ion pressure o n P M become less s igni f icant as in - cy l inder turbulence increases w i t h engine speed. 2. The combust ion process at a l l operating condit ions is restricted by the rate at w h i c h the fuel and ox id i ze r are m i x i n g . Increasing the in ject ion pressure increases both the mass f lux o f fue l into the combust ion chamber and the in -cy l inder turbulence, result ing i n enhanced m i x i n g , reducing the combust ion durat ion and increasing the peak combust ion intensit ies. 3 . The in ject ion pressure has a s igni f icant impact on emiss ions but l itt le inf luence o n fuel consumpt ion . The nature o f the inf luence depends strongly o n operat ing cond i t ion parameters i n c l u d i n g engine speed, load/equivalence rat io, and intake oxygen mass f ract ion. S igni f icant h igh-order interactions between operating cond i t ion parameters indicate that most o f the emiss ions are a result o f mul t ip le compet ing inf luences, the relative importance o f w h i c h varies substantial ly w i t h operating cond i t ion . 4. The effects o f in ject ion pressure on C O are h igh ly non- l inear , d isp lay ing a m a x i m u m at intermediate in ject ion pressures. The response is also sensit ive to operating cond i t ion ; this is poss ib ly due to changes i n the relative contr ibutions o f cy l inder and gaseous-jet induced turbulence to the combust ion process. The effects o f in ject ion pressure on H C emiss ions are inconsistent and ins igni f icant compared to the effects o f operating cond i t ion and intake d i lu t ion . 5. A t higher in - cy l inder pressures (high load) , the di f ferent ial pressure between the injected fuel and the cy l inder is lower ; hence increasing the in ject ion pressure has a greater inf luence o n the combust ion process. W i t h h igh di f ferent ia l pressures, the f l o w through the nozz le is choked , result ing i n less sensit iv i ty to in - cy l inder condi t ions . 6. W i t h o u t E G R , N O x emiss ions are s l ight ly increased w i t h higher in ject ion pressures due to more rap id and more intense combust ion . W i t h E G R , the effects o f in ject ion -78 -pressure on N O x are not significant; differences due to variability in the intake dilution level dominate the observed variations in the N O x emissions. 7. Relatively small variations in the operating condition setting have a significant influence on the emissions results, even when the test points are replicated. As a result to ensure that the experimental precision is adequate, it may sometimes be necessary to sacrifice absolute accuracy by restricting test point randomization. 4.7 Tables and Figures Table 4.1: Engine test conditions ID Speed GIMEP Y i n t 0 2 Pinj Pilot (RPM) (bar) (MPa) % 1 800 3 0.25 0.23, 0.19 21,30 11 2 800 8.5 0.45 0.23, 0.19 21,30 6 3 800 13.5 0.6 0.23, 0.19 21,30 3 4 1200 8.5 0.45 0.23, 0.19 21,30 6 5 1200 13.5 0.6 0.23, 0.19 21,30 3 6 1600 13.5 0.6 0.23,0.19 21,30 3 Table 4.2: A N O V A results as a function of load at 800 R P M P-value Term DF GISFC CO N O x HC P M Comb. Dur. Load 2 (• 0 0.01 0 0 0 Yint0 2 1 0.16 0 0 0 0 0.05 P i n i 1 0.86 0 0 0.06 0 0 Load * Yint0 2 2 0.06 II (i 0 0 0 Load * P i n i 2 0.17 0.03 0.13 (1 0.14 Yint0 2 * P i n i 1 0.24 0.09 0.07 0.3 0.06 0.06 Load * Yint0 2 * P i n i 2 0.35 0.33 0.75 0.93 0.02 0.49 Experimental Error 23 -79-Table 4.3: A N O V A resu ts as a function of speed at 13.5 bar G I M E P P-value Term DF GISFC C O N O x HC P M Comb. Dur. Speed 2 0 0 0 0 0.03 0 Yint0 2 1 0.04 0 0 0 0 0.00 Pinj 1 0.77 0.004 0.85 0.38 0.04 0 Speed * Yint0 2 2 0.14 0.027 0 0 0.23 0.97 Speed * P i n i 2 0.17 0 0.25 0.01 0.08 0.00 Yint0 2 * P i n i 1 0.07 0.169 0.84 0.4 0.49 0.20 Speed * Yint0 2 * P i n i 2 0.69 0.019 0.34 0 0.25 0.07 Experimental Error 23 Table 4.4: Summary of results from operating condition tests Operating Condition 800RPM 800 R P M 800 R P M 1200 R P M 1200 R P M 1600 RPM 2.5 bar 8.5 bar 13.5 bar 8.5 bar 13.5 bar 13.5 bar Pinj 21 30 21 30 21 30 21 30 21 30 21 30 Parameter Y m t02 GISFC 0.23 210 210 188 189 203 196 183 184 193 192 189 189 (g/GikWhr) 0.19 215 224 189 187 199 199 181 184 191 191 180 184 NO, 0.23 7.2 8 6.2 7.9 5.8 6.1 4.4 4.8 4.3 4.4 4.3 3.6 (g/GikWhr) 0.19 1.1 1.1 0.94 1.9 1.2 1.3 0.83 1 0.96 0.93 0.62 0.6 C O 0.23 3.9 3.1 0.96 0.97 11.6 3.1 2.7 1.6 10.3 9.3 6 8.5 (g/GikWhr) 0.19 6.7 6.2 4.2 2.6 19.6 8.8 5.3 4.1 17.6 17.7 7.8 15.5 H C 0.23 3.4 2.5 0.43 0.55 0.3 0.4 0.46 0.55 0.28 0.32 0.38 0.4 (g/GikWhr) 0.19 8.3 6.9 2.34 1.82 0.82 0.67 1.58 2.01 1.2 0.98 1.14 1.5 PM*100 0.23 2 1.7 1.1 1.1 3.4 0.84 1.1 1 5.4 2 4.6 3.8 (g/GikWhr) 0.19 2.5 2.1 1 1 10.8 3.5 2.5 1.2 19.2 9.3 12.9 9.3 Comb. Dur. 0.23 5 4.7 15.5 11 24.5 19.8 24.3 22.3 31 27.3 38 31.8 (°CA) 0.19 8 7 11.5 10 22.3 18 22 18 28.5 25.3 38.5 27.3 Pinj/Pmax 0.23 3.3 4.8 2.1 2.7 1.5 2.2 2.1 3 1.5 2.2 1.5 2.1 0.19 2.8 3.9 1.8 2.4 1.4 1.9 1.9 2.7 1.4 2 1.3 1.9 100 uou iviraj B21 MPal y _ n l Q • 30 M P a J Y | n , 0 2 ~ a 1 9 800 RPM, 800 RPM, 800 RPM, 1200 RPM, 1200 RPM, 1600 RPM, 0.25EQR 0.45 EQR 0.6 EQR 0.45 EQR 0.6 EQR 0.6 EQR Figure 4.1: Effect of injection pressure and operating condition on GISFC -80-800 RPM, 800 RPM, 800 RPM, 1200 RPM, 1200 RPM, 1600 RPM, 0.25EQR 0.45 EQR 0.6 EQR 0.45 EQR 0.6 EQR 0.6 EQR Figure 4.2: Effect of injection pressure and operating condition on N O x emissions 2 1 M P a ~ L Y - r m ? 30 MPa j i n l 0 2 800 RPM, 800 RPM, 800 RPM, 1200 RPM, 1200 RPM, 1600 RPM, 0.25EQR 0.45 EQR 0.6 EQR 0.45 EQR 0.6 EQR 0.6 EQR Figure 4.3: Effect of injection pressure and operating condition on CO emissions 3 o X 10 9 8 7 6 5 4 3 2 1 0 • 21 MPa l y - n rw 1 • 3 0 M P a J Y i n , O 2 _ a 2 3 2 • 2 1 M P a l Y _ n i q m 3 0 M P a J Y i n , O 2 - a 1 9 lx ft p I h rf A m . dl 800 RPM, 800 RPM, 800 RPM, 1200 RPM, 1200 RPM, 1600 RPM, 0.25EQR 0.45 EQR 0.6 EQR 0.45 EQR 0.6 EQR 0.6 EQR Figure 4.4: Effect of injection pressure and operating condition on HC emissions -81-0.25 0.2 ikWhr; 0.15 rj 0.1 CL 0.05 0 • 21 M P a ~ l v • 30 MPa _Tinl02 • 21 MPa 1 y _ n i q • 30 MPa X Y " < ° 2 - ° - 1 9 800 RPM, 800 RPM, 800 RPM, 1200 RPM, 1200 RPM, 1600 RPM, 0.25EQR 0.45 EQR 0.6 EQR 0.45 EQR 0.6 EQR 0.6 EQR Figure 4.5: Effect of injection pressure and operating condition on P M emissions 25 20 I 15 b •3 10 o o 5 Y i n l 0 2 = 0.23 ; 50% IHR = 10°ATDC 250 • C O o GISFC 230 c i 210 | 190 O LL W 170 u 25 20 I« o s 10 o o 5 Yi„t02 = 0.19; 50% IHR = 10°ATDC 150 ^•co o GISFC 10 20 30 Pinj.cton (MPa) 40 20 30 , (MPa) 40 5 C 4 5 3 6 z 1 A NOx x HC • P M 0.5 0.4 C 0.3 | 0.2 3 s CL 0.1 0.0 5 c 4 a 3 ^> £ 2 o z 1 A NOx u a • x H C • PM X X X • A A 250 230 5 210 I 3 190 O L L CO 170 ° 150 0.5 0.4 0.3 | 0.2 3 s E 0.1 0.0 25 20 1 15 3 10 o CJ 5 0 5 o 4 S 3 3 § S 2 <5 z 1 0 Yi„to2 = 0.19; 50% IHR = 17.5°ATDC • • • • 0 o • • > C O o GISFC 250 230 210 3 a 190 O CO 170 5 150 10 0 30 'inieeton (MPa) 40 n A NOx x H C x • • PM X • A A * c 0.5 0.4 0.3 § 0.2 a s 5. 0.1 0 0 10 20 30 Pinj.cn (MPa) 10 20 Pin 30 40 10 20 30 Pinj.Con (MPa) 40 h'i j.cton (MPa) Figure 4 6- Emissions and GISFC variations with injection pressure, at near peak-torque. Operating conditions (L-R): Y i n , 0 2 = 0.23, 50%IHR = 10°ATDC; Y i n t 0 2 = 0.19, 50% IHR = 10°ATDC; Y i n t 0 2 = 0.19, 50%IHR = 17.5°ATDC. For all: 1200 R P M , 4> = 0.6, 13.5 bar GIMEP. -82-16 800 RPM, 800 RPM, 800 RPM, 1200 RPM, 1200 RPM, 1600 RPM, 0.25EQR 0.45 EQR 0.6 EQR 0.45 EQR 0.6 EQR 0.6 EQR Figure 4.7: Effect of injection pressure and operating condition on gas ignition delay (GID) 800 RPM, 800 RPM, 800 RPM, 1200 RPM, 1200 RPM, 1600 RPM, 0.25EQR 0.45 EQR 0.6 EQR 0.45 EQR 0.6 EQR 0.6 EQR Figure 4.8: Effect of injection pressure and operating condition on combustion duration -83--84 -1.E+09 I1.E+08 E gl.E+07 Q •a 21.E+06 T3 1.E+05 Y i n t 0 2= 0.232 7* • 21 MPa o 30 MPa 1 10 100 Mobility Diameter (nm) 1000 1.E+09 S1.E+08 E -&M.E+07 Q . Q _c Z1.E+06 T3 1.E+05 ' int02 ' 0.19 . 21 MPa o 30 MPa 10 100 Mobility Diameter (nm) 1000 Figure 4.11: Particle size distributions at low-load, low-speed (800 R P M , 3.5 bar GIMEP) 1.E+09 1.E+09 1 .E+05 1 1.E+05 1000 10 100 1000 1 10 100 Mobility Diameter (nm) Mobility Diameter (nm) Figure 4.12: Particle size distributions at high-load, low-speed (800 R P M , 13.5 bar GIMEP) -85-1.E+09 11.E+08 I CO E •S21.E+07 Q _c |l.E+06 1.E+05 Y i n t 0 2 = 0.232 # ! / «• . 21 MPa o 30 MPa O " «. i i 1.E+09 1 10 100 Mobility Diameter (nm) 1000 1.E+05 10 100 1000 Mobility Diameter (nm) Figure 4.13: Particle size distributions at high-load, high-speed (1600 R P M , 13.5 bar GIMEP) - 8 6 -Chapter 5 Direct-Injected Hydrogen-Methane Mixtures 5.1 Introduction The use of gaseous fuels in internal combustion engines has long been seen as a possible method for reducing emissions while maintaining engine performance and efficiency. Most research has focused on the use of natural gas, due to its wide availability and relatively low cost (compared to other gaseous fuels); in fact, an estimated 5 million natural gas fuelled vehicles are currently in service world-wide [182]. More recently, hydrogen has received both popular and scientific attention as a potential long-term replacement for liquid hydrocarbon fuels in transportation applications. Both natural gas and hydrogen have benefits and drawbacks as mobile vehicle fuels. In an internal combustion engine, natural gas provides excellent anti-knock properties, but suffers from low flame propagation rates and high auto-ignition temperatures. Hydrogen's low ignition energy results in a stronger tendency to knock compared to natural gas, limiting the compression ratio (and hence maximum theoretical efficiency) for homogeneous-charge hydrogen engines. When added to the air upstream of the intake port, hydrogen's low volumetric energy density also reduces the energy content of a given volume of inducted charge [183]. However, hydrogen does have a higher flame speed than natural gas, and it is easier to ignite. This suggests that a combination of these two fuels could be a superior vehicle fuel than either individually. While hydrogen production and onboard storage are issues that have yet to be overcome, a relatively small amount of hydrogen, potentially derived from renewable sources and blended with compressed natural gas, could provide substantial benefits with little modification to an engine system developed for natural gas fuelling. A significant amount of research has been conducted investigating hydrogen/methane blend combustion in spark-ignition engines; however, few studies were identified which investigate the non-premixed combustion of hydrogen/methane blends. The purpose of this * Part of the work presented in this chapter was previously presented at the SAE World Congress in Detroit, MI, April 2 2006. McTaggart-Cowan, G . P , H.L. Jones, S.N. Rogak, W.K. Bushe, P.G. Hi l l and S.R. Munshi. Direct-Injected Hydrogen-Methane Mixtures in a Heavy-Duty Compression Ignition Engine. SAE Technical paper 2006-01-0653. 2006. Extracts reprinted with permission from SAE. -87-work is to understand the influences of hydrogen addition to the fuel in a PIDING engine. This study aims to identify the benefits and drawbacks of hydrogen/methane blend fuelling. Furthermore, in the previous chapter, higher injection pressures reduced the overall combustion rate through enhanced mixing; the current chapter investigates whether enhancing the reaction kinetics (through hydrogen addition) has a similar influence. 5. J. 1 Hydrogen/Methane Blend Combustion The concept of using hydrogen as an additive to improve the combustion rate in spark-ignition engines was first suggested for conventional gasoline fuelling [184,185]. Several more recent studies have investigated the effects of blending natural gas and hydrogen for use in homogenous charge, spark-ignition engines [186—192] . These results have shown varying positive and negative results. The most important influence of hydrogen addition is under lean premixed conditions, where the lean limit is substantially extended [186,187,190]. This has been attributed to an enhanced combustion rate and shorter ignition delay [186,193]. For a given air-fuel ratio (including both stoichiometric and lean operation), N O x emissions are higher with hydrogen addition, due to the higher flame temperature, while CO and HC emissions are reduced [187,190]. These effects become more significant as the lean limit is approached. However, because of hydrogen's ability to extend the lean limit, lower N O x emissions are achieved by running at leaner air-fuel ratios with hydrogen addition [188,190]. Flame stability in the presence of EGR is also improved at all air-fuel ratios [189,191]. The effects of hydrogen addition on efficiency appear to depend on operating condition, with some studies indicating improved efficiency [186,187,194], and others reporting reduced efficiency [191,192]. The fraction of hydrogen in the fuel (typically reported on a per-volume basis) varies between the different studies. Values of 15-20% generally achieve substantial improvements without impairing knock resistance [188,189,194]. Above 3 0 % , substantial reductions in the charge energy density, coupled with a higher potential for knock, pose substantial handicaps while generating little further benefits in emissions or stability [191,192]. The lower energy density of the gaseous charge can be overcome through turbocharging: however, this further increases the chance of knock at high hydrogen concentrations [188,190]. More fundamental premixed studies indicate that the preferential diffusion of hydrogen in a turbulent combustion event results in a higher flame propagation rate, even -88 -when the laminar flame speed is constant [195]. The flame's greater resistance to stretch results in fewer local extinction events, reducing CO and HC emissions [196]. The presence of H2 in the lean premixed flame increases the concentration of H, OH, and O radicals [196]. Increased OH concentrations may contribute to the more rapid oxidation of the methane; using 20% hydrogen may increase peak OH radical concentrations by as much as 20% [191,197]. 5.1.2 Non-Premixed Hydrogen/Methane Flames Non-premixed combustion of hydrogen/methane blends has not been as extensively studied as the premixed case. In a low-pressure, low-temperature co-flow burner experiment, non-premixed flame stability is enhanced by the higher flame speeds and improved mixing associated with hydrogen addition to either the fuel or the oxidizer [198]. Differences in fuel-stream density with hydrogen addition are secondary [198,199]. The higher diffusivity of the hydrogen also increases flame thickness under partially-premixed conditions [200]. In industrial gas turbines and boilers, hydrogen addition enhances prompt NO formation (due to high H and OH radical concentrations) while flame stability is improved [201]. The auto-ignition of direct-injected hydrogen, on a time-scale similar to diesel, requires a combustion chamber temperature in excess of 1100K [202]. Interestingly, reducing the charge oxygen concentration (by, for example, EGR) has little effect on ignition delay or combustion rate. Shock-tube studies of the auto-ignition of methane-hydrogen mixtures show that while hydrogen somewhat reduces the ignition delay, it is insufficient to achieve ignition under diesel-engine conditions within the time scales needed for compression-ignition engine operation [203]. In non-premixed counterflow methane/heated air jet experiments, the concentration of hydrogen in the methane influences the ignition mechanism. At concentrations below 30% by volume, the presence of H radicals enhances the methane ignition, while above this value hydrogen ignition dominates the process, with delays independent of the relative H2/CH4 concentration [204]. In internal combustion engines, the use of hydrogen mixed in the oxidizer improves the combustion rate of a spark-ignited direct-injected natural gas jet. Emissions of CO and HC are reduced, but the higher flame temperature results in increased N O x emissions [205]. Similar results are reported for hydrogen addition to diesel-fuelled engines, where substantial reductions in particulate matter (PM) emissions are also observed [206]. These studies -89-indicate substantial potential benefits of direct-injection, non-premixed fuelling with hydrogen/natural-gas fuel blends. 5.2 Experimental Methodology The experimental work presented in this section was carried out on the single-cylinder heavy-duty research engine (SCRE) whose general configuration was discussed in sections 3.1 and 3.2. For this work, the 17:1 compression ratio piston was installed in the engine. A n alternate fuel supply system was also developed, which fed gas from storage bottles directly into the high-pressure compressor, thus .providing fuels of varying composition to the engine at high pressures. As the gaseous fuel mass-flow sensor worked on the coriolis force principle, it was insensitive to the composition of the fuel. Finally, to maximize the accuracy of the fuel blends, as well as to alleviate safety concerns, the hydrogen/methane blend fuel was supplied in industrial style compressed gas cylinders. As a result of the high costs of fuelling the engine in this manner, the quantity of fuel available was limited; this restricted the number and scope of the tests conducted. 5.2.1 Experimental Conditions The operating parameters used for controlling the S C R E are discussed in Section 3.3. For the current work, two separate test conditions were used to evaluate the influence of hydrogen/methane blend fuelling over a range of engine operating modes. The first condition was selected to minimize the fuel flow rate (hence increasing the amount of testing that could be carried out using the limited quantity of fuel available). The second set of tests was carried out at a test condition more representative of actual engine operation. For the first round of tests, the experimental condition was selected to minimize fuel consumption while still generating relatively high emissions levels, so that the effectiveness of hydrogen at enhancing poor natural gas combustion could be determined. The condition selected (Table 5.1) used a high E G R fraction, at low speed and low load but moderate oxygen-equivalence ratio (<j)). To establish influences over a range of combustion conditions while minimizing the required changes to the operating mode, a range of combustion timings was used. By varying combustion timing, highly stable conditions (early timings) and very unstable conditions (late timings) could be tested at the same baseline condition (EGR fraction, load, speed). By fixing the operating condition, then varying the timing, it was -90-possible to minimize variations due to non-repeatability of the operating condition setpoint. Replication of timing sets was used to establish repeatability. Most of the testing was carried out at 16MPa, to ensure that the commanded injection opening durations were repeatable (in excess of 0.9 ms). The second test mode used in this work was more representative of a mid-speed, high-load (-75% of rated power) condition at a relatively high EGR rate. Due to the high gaseous fuel flow rate (in excess of 5 kg/hr) and limits on the available quantity of fuel, only one set of combustion timings (0-15°ATDC for the 50% IHR) for each fuel blend could be conducted. This test condition was similar to that used in Chapters 6 and 7, and to test condition 5 from Chapter 4 (Table 4.1). 5.2.2 Fuel Blends The fuel used in these tests included natural gas (from the building supply), pure methane, and 10, 15, 23, and 35% hydrogen (by volume) in methane. While it would have been interesting to investigate higher H2 concentrations, the upper limit was selected out of concern regarding the safety of the fuel supply system. Of the gas blends, the methane was supplied at a certified grade 2.0 (>99% purity [207]), while the 10%, 23%, and 35% H 2 (balance methane) were supplied at certified standard accuracy (analyzed accuracy ± 2%, impurities <0.1%). The 15%) H 2 bottles contained some heavier hydrocarbons, as they were partially blended with natural gas as well as pure methane. Gas chromatograph analyses of the natural gas, pure methane, and 15% H 2 blends were performed by Westport Innovations Inc. For the remaining bottles the industrial supplier, Praxair Inc., performed the composition analysis. The analytical results for the gas blends are provided in Table 5.2. Heating values, molecular weights, and H/C ratio are based on the reported compositions. The natural gas, and the 10% and 23% hydrogen in methane blends, were tested in the first set of tests (mode 1). The second set of tests (mode 2) used natural gas, methane, and 15% and 35% hydrogen in methane. 5.2.3 Replications and Randomization For the first set of tests, fuel flow rates were sufficiently low that the test points were replicated. Three separate tests were conducted for the natural gas while the 23% hydrogen in methane tests were replicated twice. At the 10% case, only a single data set was collected. The tests were conducted in a paired manner, with each test point being tested on a hydrogen -91-blend and then on natural gas. The 10% and 23% tests were carried out on different days. A l l timings were tested at one operating mode before the operating condition was changed. Due to this restriction on randomization, the results cannot be analyzed in a statistically valid manner. For the 23% hydrogen and natural gas fuel blends, the uncertainties presented in the experimental results are based on the 95% confidence intervals. As only a single data set was collected for the 10% hydrogen case, the average of the percentage errors for the natural gas and 23% hydrogen cases were used to represent the uncertainty. For the second round of tests, the high fuel flow rates limited the testing to only one set of timings for each fuel blend. As a result, the uncertainties presented are those derived from the replicated natural gas tests carried out in Chapter 6 at a similar operating condition. While not statistically valid, this procedure provides an indication of the uncertainty in the measured results including both experimental variability and instrumentation errors; as such, it is superior to the estimated uncertainty calculated in Section 3.4. 5.3 Results There were significant differences between the fuel blends and the operating condition used in the two sets of tests. The presentation of the results is divided into first and second-round results, relating to these two separate test sets. The general emissions results are presented first, followed by the detailed particulate-matter analysis and the effects on the combustion process as indicated by the in-cylinder conditions. Finally, the influence of hydrogen addition on net greenhouse gas emissions is presented. 5.3.1 Emissions At the low-load operating condition, the effect of hydrogen addition is most significant on the CO and HC emissions at 23% H2, as shown in Figure 5.1. The 10% H2 level has a relatively minor influence on emissions, with no detectable effect on N O x or P M . The CO and HC emissions appear to be slightly reduced, although the magnitudes of the differences are quite small (on the order of 5-10%, less than the experimental uncertainty). The use of 23% H2 in the methane fuel has a much more substantial impact on the emissions than does 10% H2. Further, these effects are largest at later combustion timings; for the earliest timings, adding hydrogen to the fuel has little effect on emissions. Specifically, CO and HC emissions are reduced, while the N O x emissions are slightly -92-increased. The only observed significant influence is at the latest timings, where P M is substantially reduced with 23%H2. The presence of hydrogen in the combustion zone may affect pollutant emissions due to an increase in the concentration of the OH radical. This highly reactive molecule provides more rapid oxidation of unburned fuel and partial-combustion species such as CO and HC. Hydrogen also effectively reduces local flame extinctions induced by high turbulent strain-rates, events that may generate substantial pollutant emissions. These influences are more substantial at later combustion timings, where the combustion temperature is lower and instability is greater, leading to higher emissions of partial combustion by-products. The fact that N O x emissions are slightly increased at the 23% H2 case is most likely due to the small increase in flame temperature. An increase in the prompt-NO mechanism may also result from the higher OH concentrations [202]. Irrespective of the source, the increase in N O x emissions is relatively small. High-Load Emissions The effects of hydrogen addition to the fuel are more significant at high-load operating conditions, as investigated in the second round of testing. The pure methane results are presented to provide a comparison between the natural gas used as a reference in the first round of testing and the pure methane used as the balance gas in the majority of the hydrogen/natural gas blend fuels. The emissions results (CO, HC, P M , and NO x ) are presented in Figure 5.2. The results indicate that using pure methane in place of natural gas does not substantially influence the N O x or HC (unburned fuel) emissions, but has a more substantial effect on P M and CO. As would generally be expected, the methane fuel has lower P M emissions than does the natural gas, due primarily to the smaller quantities of ethane and propane in the fuel (Table 5.2). Pure methane does, however, increase the CO emissions, although the cause of this effect is not clear. A more comprehensive investigation of the effects of these additives on emissions and performance at a very similar operating condition is provided in Chapter 6. In general, the differences between methane and natural gas are relatively small, suggesting that the composition of the balance gas should not have a significant effect on the influences of hydrogen addition. Both the 15 and 35% H2 cases (Figure 5.2) show a significant reduction in CO compared to natural gas. This may be due either to a kinetic effect (increasing OH radical concentration) or a physical/thermal effect (higher combustion temperature, better mixing). -93-As well, some reduction in CO would be expected due to the reduction in the carbon content of the fuel (Table 5.2). The fact that the addition of hydrogen results in higher N O x emissions at the 35% H2 case may also be due to either thermal effects or kinetic effects. Kinetically, increases in OH radical concentration tend to enhance the thermal NO mechanism [37]. The 80K increase in the adiabatic flame temperature due to the hydrogen addition (Table 5.2) also increases the rate of the thermal NO formation mechanism. The higher flame temperature explains the increase observed at 35% H 2 (and at 23%). The relatively small difference in temperature at the lower concentration results in changes in the NO formation rate that cannot be differentiated from background variability. Hydrogen addition to the fuel substantially reduces HC and P M emissions. The presence of higher concentrations of H and OH radicals are the most likely cause for reduced HC emissions, as they tend to accelerate the oxidation of methane. Similarly, the higher concentration of these radicals substantially increases the oxidation rate of solid carbon particles and P M precursors. Under these conditions, either reductions in P M formation or increases in oxidation dominate the P M emission rate. There is no evidence of a reduction in hydrogen abstraction resulting in increases in P M formation through the H A C A mechanism (Section 2.2.5). However, this effect may be occurring but being offset by increases in oxidative mechanisms. 5.3.2 Particulate Matter Further insight into the effects of hydrogen addition to the fuel is provided by the particle size distributions collected by the mobility particle sizer and the black-carbon concentration measured by the Aethaelometer. For the low-load case (first round of tests), the black carbon content is reduced for both the 10% and the 23% H 2 cases, as shown in Figure 5.3. The black-carbon fraction of the total measured P M is also reduced with 23% H 2 . In both cases, the earliest timing cases, where the P M levels are • in general lowest, show no significant differences. The particle size distributions (Figure 5.4) show similar results, with significant reductions in particle sizes numbers at the later timings with the 23% H 2 fuel but no significant changes at early timings or with 10% H 2 . These results are consistent with the mass measurements discussed in section 5.3.1, and suggest that the hydrogen addition is having a greater effect at reducing P M at those conditions where the P M loading is highest. They also suggest that the reduction in P M mass is mainly attributable to a reduction in -94-black-carbon mass, whereas the volatile contribution to the total P M is relatively unchanged. This indicates that the volatile P M was originating primarily from either the liquid fuel or the lubricating oil, as was discussed in Chapter 4. More details on the P M emissions under high load are shown in Figures 5.5 and 5.6. The black carbon content of the P M (Figure 5.5, both absolute emission level and the relative fraction of the total P M as measured by the TEOM) is not significantly different between the natural gas and methane cases. The 15% H2, however, significantly reduces the black carbon level, while for the 35% H 2 case emissions levels approach the detection limits of the Aethalometer. This corresponds to a reduction in the black-carbon fraction of the P M from as much as 45% to 5% for the same operating condition. This indicates that the P M measured by the T E O M is primarily volatiles which are either condensing in the exhaust stream or on the face of the T E O M filter. Further evidence of the presence of significant volatile concentrations in the exhaust is demonstrated by the particle size distributions, Figure 5.6. These measurements indicate that at 35% H2, there is a significant 'nucleation mode' of particles; these particles are typically attributed to post-dilution nucleation of volatile species [99]. This 'nucleation mode' is not present for conditions where P M levels are higher, most likely due to the presence of sufficient solid P M (black carbon) surface area for volatile condensation to occur. In general, volatile condensation onto existing particles is thermodynamically favoured over direct nucleation of new particles [6]; however, at the 35% H2 case, the lack of existing surface area leads to significant volatile nucleation rates. 5.3.3 Combustion Analysis The sensitivity of the pollutant emissions to the addition of hydrogen to the fuel indicates that the hydrogen is significantly influencing the combustion event. Further insight into the combustion process can be gained from the in-cylinder pressure trace and its derivative, the net heat-release rate, as discussed in section 3.3.1. For the low-load condition, the burn duration (10-90% of integrated heat release), fuel consumption (GISFC), peak heat-release rate, and coefficient of variation (COV) of the GIMEP are shown in Figure 5.7. The GISFC (calculated based on an energy-equivalent mass of diesel) shows no significant influence of either timing or fuel composition. On the other hand, the burn duration is substantially reduced for the hydrogen-fuelling cases at late timing, especially with 23% H2; however, the burn duration for the earlier timings is unaffected. This suggests that different -95-mechanisms may be restricting the combustion rate at early and late timings, with a chemical kinetic limit at late timings, compared to a mixing-limited condition for early timings. The peak heat-release rate (corresponding roughly to the maximum rate of chemical energy being released from the fuel) averages approximately 20% higher at 23% H2 fuelling than for the natural gas. The difference at the 10% H2 case is much less significant, although there is a slight increase in peak H R R at most timings. The use of hydrogen/methane blend fuel also substantially reduces the combustion variability ( C O V of the GIMEP). For the 10% H2 fuel, the variability is significantly reduced only at the latest combustion timings. The 23% H2 reduces variability at all timings, although the effect is most substantial at the latest timing. This reduction in combustion variability, most likely due to increased flame stability with the addition of hydrogen, contributes directly to the observed reductions in C O and H C emissions. Switching from natural gas to hydrogen/methane blended fuel has no significant effect on the pilot ignition delay, as shown in Figure 5.8. Similarly, the gas ignition delay at 10% H2 fuelling is no different from natural gas; however for the 23% H2 case, the gas ignition delay is reduced by, on average, 20%. Both pilot and gaseous fuel delays are defined as the time between the commanded start of injection and the observed start of combustion. As such, they include any physical delay within the injector, as well as both mixing and chemical delay times for the injected fuel. The commanded start-of-injection is a recorded value while the start-of-combustion timing is determined by examination of the heat-release rate. Further details on the timing definition process are provided in section 3.3.1. The shorter gas ignition delay time with H2 addition is consistent with previous premixed and non-premixed auto-ignition of methane tests, which show that hydrogen addition substantially reduces ignition delay times [203,204]. While the process in this situation relates to a non-premixed jet being ignited by a pilot flame, and hence is not directly comparable to either of the cited works, it is not surprising that the addition of H2 reduces the ignition delay. However, the previous work suggests that even at 10% H2, a noticeable reduction in ignition delay occurs; the current results indicate that a more substantial quantity of hydrogen is required before it has a significant effect. The shorter gas ignition delay has a number of effects on the combustion process, including reducing the time available for mixing prior to ignition. In general, the lower -96-density of the blended fuel reduces the turbulent mixing rate and the penetration distance of the gaseous jet. It is unlikely that the higher molecular diffusivity of the hydrogen offsets this effect, as the mixing process is turbulence dominated. The shorter ignition delay leads to less methane having over-mixed during the ignition delay period, contributing to the reduced HC emissions. The addition of hydrogen also results in wider flammability limits of the mixture during the initial combustion-initiation phase, leading to more fuel being available to burn during the immediate post-ignition partially premixed burn; this increases the peak heat-release rate. The influence of hydrogen addition on the observed pressure trace and the heat-release rate is relatively small, as shown in Figure 5.9 for the 5 and 15°ATDC cases. The 10% H 2 case shows no significant difference from natural gas, while the 23% H 2 case has a more significant effect; the shorter ignition delay is one of the main contributing factors. The 23% H 2 fuelling also substantially increases the peak heat-release rate, as suggested in Figure 5.8. This effect is relatively consistent, although the increase in HRR is more substantial at 15°ATDC than at the earlier timings. These results indicate that the greatest influence of hydrogen addition is on the early stages of the combustion event, with a shorter ignition delay and a more intense early combustion phase. High-Load Combustion Performance The effect of hydrogen addition to the fuel on combustion performance at the high-load condition differs significantly from the influences observed at the low-load case. The effects on the peak heat-release rate, combustion duration, and two measures of combustion variability (COV of the peak pressure and of the GIMEP) are shown in Figure 5.10. The heat release-rate plots at combustion timings of 5 and 15°ATDC are shown in Figure 5.11. Contrary to the low-load case, the maximum heat-release rate is substantially reduced by the addition of hydrogen to the fuel. The net heat-release rates shown in Figure 5.11 indicate that the methane and low-hydrogen cases both have a significant impact on the combustion process at earlier timings, but that at the latest timings only the highest hydrogen concentration significantly influences the combustion event. This suggests that the combustion event at the early timings may be sensitive to changes in chemical kinetics, but that at the later timings the main factor affecting the combustion rate is the mixing process. -97-At this operating condition, both 15% and 35% H 2 addition significantly reduce the gas ignition delay (Figure 5.12), while the methane and natural gas delays are indistinguishable. The shorter gas ignition delay for the hydrogen/methane cases result in less gas having pre-mixed to a combustible mixture when the gas ignites, resulting in less energy release in the early premixed burn phase of the combustion event and hence lower peak heat-release rates. The fact that the wider flammability range of the hydrogen/methane fuel does not offset this reduction in the ignition delay time, as it did in the low-load case, suggests that there is an injection-rate limit to the initial combustion at high load, while the limiting factor at low loads is kinetic-limited. As the overall burn duration (Figure 5.10) is not influenced by hydrogen addition to the fuel, it is evident that the total combustion duration is primarily dependent on the rate of injection and mixing of the fuel, and that kinetic effects related to the easier flammability of the hydrogen are not significantly influencing the overall burn duration. The observation that the early-stage combustion stability (measured by the COV of the peak cylinder pressure) is substantially improved suggests that hydrogen is not only reducing the ignition delay but also that it is providing a more repeatable ignition event. That the overall combustion stability (indicated by COV GIMEP, Figure 5.10) is not significantly influenced provides further evidence that the primary role of the hydrogen addition on the overall combustion event is to enhance the ignition process. 5.3.4 Greenhouse Gas Emissions The preceding emissions and performance results demonstrate that hydrogen addition to natural gas fuel can substantially reduce emissions of harmful combustion by-products while maintaining combustion efficiency. A further advantage of hydrogen is that it contains no carbon, resulting in reduced engine-out C 0 2 emissions. The effects of hydrogen on the C 0 2 emissions for the two sets of tests carried out are shown in Figures 5.13 and 5.14. That the addition of hydrogen to the fuel also reduces the unburned fuel (primarily methane) emissions provides a further GHG benefit, as methane is a key contributor to global IR absorption. Combining methane and C 0 2 emissions provides a net infrared-absorption value which can be expressed as an equivalent mass of C 0 2 . At the low-load case, the relative contribution of the unburned fuel is substantial. The 23% H 2 reduces C 0 2 emissions by ~8%; at early timings, the reduction in net GHG emissions is equivalent to this reduction in C 0 2 , due to the low level of unburned fuel emissions (see Figure 5.1). However, at the late timings -98-where the unburned fuel emissions are significantly greater, the net G H G emissions are reduced by almost 25%. At the high-load case, the majority of the G H G emissions are attributed to CO2, with the 35% hydrogen blend reducing CO2 emissions by -12% (equivalent to the fraction of carbon replaced, on an energy basis, in the fuel); including the reduction in unburned fuel, the net G H G emissions are reduced by -15%. Although the use of H2 can substantially reduce the GHG emissions (in excess of the simple reduction in CO2 due to carbon replacement), this reduction applies only for engine-out emissions; the net G H G impact depends on how the H2 is generated, a topic which is beyond the scope of this thesis. 5.4 Discussion The results demonstrate that the addition of varying quantities of hydrogen to natural gas has a significant impact on the overall combustion event. These results can be used as guidance in developing an optimum procedure for blending natural gas and hydrogen to achieve substantial reductions in harmful pollutant emissions while maintaining or improving system efficiency. 5.4.1 Combustion Implications The implications of the hydrogen addition on the combustion process are discussed in detail in earlier sections. In general, the results indicate that the ignition process is dependant at least in part on chemical kinetics, with hydrogen addition substantially reducing ignition delay. It also improves combustion stability, especially early in the combustion cycle and at those conditions with normally greater instabilities. The improved stability most likely contributes to the observed reduction in combustion by-products. The effect of hydrogen addition to the gaseous fuel on the particulate matter emissions provides significant insight into the relative importance of different P M sources and formation processes. Higher hydrogen concentrations (23%) and 35%) reduce the emission of solid carbon by a substantially greater fraction than would be expected simply due to the reduction in fuel carbon content (-8% and -11% on an energy basis, respectively). It is not possible to identify whether the initial solid carbon formation process is being impaired or i f the later-stage oxidation is being enhanced. The simultaneous appearance of a significant concentration of nucleation-mode particles indicates that volatile nucleation may -99-be significant in the absence of sufficient particulate surface area upon which condensation may occur. The significantly greater influence of 35% H2, compared to 23%, indicates that there is benefit to going to higher hydrogen concentrations; however, given that the black-carbon content is negligible at 35% H2, it is unlikely that concentrations greater than this value will further reduce P M levels. This suggests that there is a lower limit on the P M mass emissions, which may be attributed to volatile contributions from the lubricating oil or the diesel pilot fuel. The addition of hydrogen to the fuel at two different operating conditions has generally similar effects on emissions, although the effects on the net heat-release rate are significantly different. This indicates that the overall combustion progression, as described by the net heat-release rate, is not the primary factor influencing most of the pollutant formation processes. The fact that the influence of hydrogen addition on the combustion process varies with operating condition is similar to the observed variations with operating condition discussed in Chapter 4. Specifically, at low-load conditions, the addition of hydrogen substantially increases the heat release rate, most likely due to enhanced chemical kinetics. Conversely, at the higher load case (where the differential pressure between the fuel rail and the combustion chamber is lower) the addition of hydrogen reduces the intensity of the combustion event. This is most likely due to the slower injection rate and reduced mixing caused by the lower density of the injected fuel. Furthermore, it suggests that at higher loads, the overall combustion rate is mixing-limited, while at lighter loads there is a significant kinetic limit to the combustion rate. However, the fact that the emissions are uniformly reduced, especially at the higher load case, suggests that enhancement of the kinetics and reduction of combustion variability have very significant impacts on the emissions of combustion by-products. 5.4.2 Emissions and Applications The application of hydrogen-blended natural gas to a direct-injection engine suggests a number of intriguing possibilities. Hydrogen-blended natural gas offers the potential to achieve substantial net G H G reductions. Assuming an equal energy requirement, an equivalent amount of hydrogen could be used to fuel either a single pure hydrogen powered vehicle or 15 vehicles using 20% (by volume; -7% on an energy basis) hydrogen in natural gas. While the reduction in CO2 emissions would be equivalent, the reduction in CH4 -100-emissions from the 15 vehicles would be substantially greater than the reduction that could be achieved by removing a single vehicle; this could result in as much as a 50% greater reduction in net G H G emissions for an equivalent amount of hydrogen. Although hydrogen-fuelled vehicles (either internal-combustion or fuel-cell powered) may offer some potential improvements in efficiency, benefits in well-to-wheel efficiency will most likely not be of the same magnitude as the net reduction in GHG emissions indicated here. The addition of varying levels of hydrogen to natural gas over a range of operating conditions substantially reduces emissions and improves combustion performance. This provides a number of potential opportunities for application to a heavy-duty mobile power plant. Through the addition of hydrogen, the engine's operating condition range can be extended before excessive levels of harmful pollutants and combustion instability are reached. This could be combined with higher levels of EGR and more advanced timings to achieve lower N O x levels and higher combustion efficiency while maintaining low P M , CO, and HC emissions. Further work is obviously required to evaluate the true potential for this concept over an entire engine duty cycle. The benefits discussed above are somewhat offset by the need to provide significant amounts of hydrogen. Furthermore, practical on-board storage would most likely be limited to compressed gases, and a greater work input would be required to compress the blend to the injection pressure, due to the lower density of the gaseous mixture (discussed in section 7.5.1). A n alternative for applying the blended H 2 and natural gas combustion system that could overcome some of these drawbacks would be to provide varying control over the hydrogen content of the fuel. This could be achieved by providing separate gaseous reservoirs (which increases system complexity), and then varying the rate of blending. As such, larger quantities of H 2 could be provided to the engine at times when combustion instability is higher and emissions of P M , HC, and CO greater, for example during start-up or transient conditions. Whether the potential offered by such a system could compensate for the increased cost and complexity requires further study. Compared to pure hydrogen fuelling, hydrogen blended with natural gas has a number of practical advantages. One is that on-board fuelling systems (either port- or direct-injection) can operate with little or no modification on the blended fuel; modifications, in terms of materials, design, and safety requirements, for pure hydrogen will substantially -101-enhance the near-term cost and complexity of these vehicles. Furthermore, natural gas supply and distribution networks currently exist, while fuelling stations could then blend the natural gas with hydrogen (delivered in liquefied form by tanker-truck, for example) at the pump. These results identify some of the potential benefits to using hydrogen-blended natural gas as a near-term transportation fuel alternative; the implications will be discussed further in Chapters 6 and 7. 5.5 Conclusions 1. Both the engine's operating condition and hydrogen concentration have a substantial impact on the influence exerted by hydrogen addition to the natural gas. At low load, the addition of 10% H2 to methane has little impact on performance or emissions, while higher concentrations (23%) have significant impacts. At high load, both low (15% H 2 ) and high (35%) concentrations significantly impact both the combustion progression and emissions. 2. In general, particulate matter levels are reduced by H 2 concentrations greater than 10%, with larger reductions at those conditions where the P M levels are highest. Higher hydrogen concentrations also significantly reduce CO and HC emissions. Hydrogen addition at high load generates significantly greater effects on emissions than at low load. 3. At very high (35%) H 2 concentrations, black-carbon (soot) concentrations are negligible, but significant quantities of nucleation-mode particles are formed. These particles are composed primarily of volatiles that self-nucleate due to insufficient surface area being available for condensation. This suggests that there is a baseline P M level composed of volatiles, attributable to lubricating oil, diesel pilot, etc. which are not significantly influenced by the gaseous combustion process. 4. At low loads, the addition of hydrogen to the gaseous fuel improves the combustion stability, increases the peak heat-release rate, and reduces the combustion duration. This indicates that chemical kinetics can enhance the combustion at low loads. 5. At high loads, the peak combustion intensity is reduced because of the lower energy density of the fuel; the overall combustion duration, however, is not affected. The fact that enhanced chemical kinetics do not reduce the combustion duration suggests that -102-mixing limitations are the primary restriction on the combustion process at higher load conditions. 6. Hydrogen addition to a non-premixed gaseous jet combustion system has no significant effect on fuel conversion efficiency. Net greenhouse gas emissions are reduced, through both the lower carbon content of the fuel and the reduction in unburned methane emissions. 7. Higher H 2 concentrations generate higher N O x emissions. This is most noticeable at the highest H 2 levels (35%), although a small (non-significant) increase is also detectable at 23% H 2 . These increases may be attributed to increases in the combustion temperatures of the hydrogen/methane blended fuels. 8. The addition of hydrogen to the fuel reduces the gaseous fuel ignition delay time, by approximately 15% for 15% H 2 , 20% for the 23% H 2 , and 30% for the 35% H 2 cases. This indicates that chemical kinetics are playing a substantial role in the ignition process, as the effects of the hydrogen addition on the mixing process would be to reduce the jet density leading to lower penetration and less mixing. 9. The principal difference between the natural gas tested (composition in Table 5.2) and methane fuels at high-load conditions is a reduction in P M and an increase in CO for the pure methane fuel. Ignition delays are slightly longer with pure methane, resulting in higher peak heat-release rates. Early-stage combustion variability is also slightly greater. The effects of ethane and propane in the natural gas are investigated in greater detail in Chapter 6. -103-5.6 Tables and Figures Table 5 . 1 : Eng ine operating mode for hydrogen/methane b lend testing Parameter 1 s t R o u n d 2 n d R o u n d Speed ( R P M ) 800 1200 G I M E P (bars) 6 13.5 ( >02 0.5 0.6 E G R (mass %) 40 30 Ymt02 0.175 0.19 F u e l Pressure ( M P a ) 16,21 21 5 0 % I H R ( ° A T D C ) 0, 5, 10, 15 0, 5, 10, 15 P i l o t (mg/inj) 5 5 Tab le 5 .2 : G a s compos i t i on as analyzed by gas chromatography C o m p o n e n t Natura l Gas C H 4 1 0 % H 2 * 1 5 % H 2 2 3 % H 2 * 3 5 % H 2 * H e 0 0 0 0 0 0 0 2 0 0 0 0 0 0 H 2 0 0 10.0 14.7 23.3 35.1 C O 0 0 0 0 0 0 n -Butane 0.09 0 0 0.06 0 0 i -Butane 0.10 0 0 0.03 0 0 i -Pentane 0.03 0 0 0 0 0 n-Pentane 0.02 0 0 0 0 0 H e x a n e 0.005 0.001 0 0 0 0 Heptane 0 0 0 0 0 0 Octane 0 0 0 0 0 0 N i t r o g e n 1.05 0 0 0.34 0 0 Methane 96.12 99 .999 90.0 84.0 76.7 64 .9 C a r b o n D i o x i d e 0.43 0 0 0.05 0 0 Ethane 1.70 0 0 0.75 0 0 Propane 0.45 0 0 0.19 0 0 M W (kg/kmol) 16.76 16.04 14.6 14.14 12.74 11.09 L H V (kJ/kg) 48404 50007 50959 51151 52561 54433 H H V (kJ/kg) 53661 55489 56685 56837 58664 60977 Tadiabatic (K-) 2485 2470 2490 2500 2520 2 5 5 0 H : C ratio 3.94 4.00 4.22 4.32 4.61 5.08 - c o m p o s i t i o n ana! ysis by Praxai r (a l l other species <0.1%) - 1 0 4 -2.5 2.0 S 1.5 x 1 0 O Z 0.5 0.0 0.05 0.04 I 0 0 3 b S 0.02 5 Q. 0.01 0.00 i • •• 10% H 2 - - X - - 23% H2 7 6 | 5 ^ 4 3 3 x 2 1 0 —A— NG ••••• 10% H 2 - X - - 23% H2 - X ' 10 50%IHR (°CA) 15 20 10 15 50%IHR (°CA) 20 - & - N G ••••• 10% H 2 - X - 23% H 2 AT 6 - —A— NG 5 - . . . . . . 10% H2 4 - - X — 23% H 2 ro 3 O o O 2 1 0 10 50%IHR (°CA) 15 20 10 15 50%IHR (°CA) 20 Figure 5 . 1 : E m i s s i o n s compar ison between natural gas fue l l ing , 1 0 % and 2 3 % H2. 800 R P M , 6 bar G I M E P , 0.5 (j), 0.175 Yinto2, Injection Pressure = 16 M P a . (reprinted with permission from SAE 2006-01-0653. © SAE International 2006) o 3 X o 5.0 4.0 3.0 2.0 1.0 0.0 0.04 -00.03 (3 0.02 S a-0.01 0.00 . . . A . . . NG • 15% H2 1. - x - 3 5 % H2 - ^ - C H 4 1.4 1.2 I 1.0 ^ 0 8 O S 0.6 I I 0.4 0.2 0.0 . . . A - NG - • - 15% H 2 - x - 3 5 % H 2 - - ^ - C h U A i"—*—* £ * X . j 5 10 50%IHR (°CA) 15 20 5 10 50%IHR (°CA) 15 20 . . . A . . NG - « - 1 5 % H 2 - x - 3 5 % H 2 - ^ - C H 4 A 1~ A - . NG - - - 1 5 % H 2 • x - 3 5 % H 2 - » - C H 4 -5 5 10 50%IHR (°CA) 15 20 5 10 50%IHR (°CA) Figure 5 .2 : E m i s s i o n s compar ison between natural gas, pure methane, 1 5 % and 3 5 % H2. 1200 R P M , 13.5 bar G I M E P , 0.6 <|>, 0.19 Y i n to2, In ject ion Pressure = 21 M P a . - 1 0 5 -8000 6000 E .3 4000 O 2000 —•— NG 10% H 2 - - X - - 2 3 % H 2 Figure 5 .3 : 0.9 0.8 •2 0.7 o 2 0.6 u_ g 0.5 JD S 0.4 o ^ 0.3 o 3 0-2 0.1 0 —•— NG _ •••£]••• 10% H 2 - - X - - 2 3 % H 2 H • /f 10 50%IHR (°CA) 15 20 15 20 0 5 10 50%IHR (°CA) B l a c k - c a r b o n P M content compar ison between natural gas f u e l l i n g , 1 0 % and 2 3 % H 2 . 800 R P M , 6 bar G I M E P , 0.5 d), 0.175 Y i n t 0 2 , Inject ion Pressure = 16' M P a . 1.E+08 I 1.E+07 m "fe O 5 Q c .E+06 1.E+05 1.E+08 10 100 Mobility Diameter (nm) 1000 1.E+05 10 100 Mobility Diameter (nm) 1000 Figure 5.4: Part ic le size distr ibut ion compar ison between natural gas f u e l l i n g , 10%) and 2 3 % H 2 . 800 R P M , 6 bar G I M E P , 0.5 <j), 0.175 Y i n t 0 2 , In ject ion Pressure = 16 M P a . - 1 0 6 -5 10 50%IHR (°CA) , • • - A • • NG C H 4 -« -15%H2 - X - 3 5 % H 2 j , — — f r H \ \ 1 -5 5 10 50%IHR (°CA) 15 20 Figure 5 .5 : B l a c k - c a r b o n P M content compar ison between natural gas, pure methane, 1 5 % and 3 5 % H 2 . 1200 R P M , 13.5 bar G I M E P , 0.6 (>, 0.19 Y i n t o 2 , In ject ion Pressure = 21 M P a . 1.E+08 | 1.E+07 g 1.E+06 Q . Q I 1.E+05 "D 1.E+04 1.E+08 1 1.E+07 -0°ATI rr™:r.rrrri-• NG ° CH 4 - 15%H2 x 35% H 2 I * i i .» • r-» i * f \ . o i l l ; ^ Jj •- - - -x _T y *«. Ml t i M l ; ; ; 1 | I ; *x ; x ; 10 100 Mobility Diameter (nm) 1000 "E o 1.E+06 1.E+05 1.E+04 ;-10°ATDC ., i "A" • NG ° CH 4 - 15% H 2 x 35% H 2 i i o: X. . | * ' • X ! 1 1 10 100 Mobility Diameter (nm) 1000 1.E+08 1 1.E+07 | § 1.E+06 Q . D c. I 1.E+05 1.E+04 1.E+08 1 1.E+07 5°ATDC • NG o CH 4 - 15% H 2 x 35% H 2 ! ] j i n| ! i ! i i in : T j u ; x< #, x i k ....... x • • , • i i i h i l l 1 . 10 100 Mobility Diameter (nm) 1.E+06 1.E+05 1.E+04 10 100 Mobility Diameter (nm) 15°ATDC \ . t i f: = 3 ; . NG 0 C H 4 - 15% H 2 x 35% H 2 i 1 • EEH;~H::±Hr!:ll"I 'A 0 * i * ' V * ' * X 1000 Figure 5.6: Part ic le size distr ibut ion compar ison between natural gas, pure methane, 1 5 % and 3 5 % H 2 . 1200 R P M , 13.5 bar G I M E P , 0.6 <|>, 0.19 Y i n t o 2 , Inject ion Pressure = 21 M P a . - 1 0 7 -250 I 200 O | 150 t J 100 o u- SO C3 0 250 < 200 ^ 150 g 100 S 50 Q-rr * J — A — NG 10% H 2 23% H2 10 50%IHR (°CA) 15 5 10 50%IHR (°CA) 15 20 -&-NG 10% H 2 -X- 23% H2 20 _ 15 < o ° . 12 o s 9 6 3 0 i r=C_ ie. . - -. . . - ' X •• •'••'r""' " x 1 1 5 * - A - N G 10% H 2 23% H2 ] g 3 a, I 2 > o 10 50%IHR (°CA) 15 5 10 50%IHR (°CA) 15 20 — A — NG • 10% H 2 - - * - 2 3 % H 2 20 Figure 5.7: Combustion performance comparison for natural gas, 10% and 23% H 2 . 800 R P M , 6 bar GIMEP, 0.5 <j), 0.175 Y i n to2, Injection Pressure = 16 MPa. (reprinted with permission from SAE 2006-01-0653. © SAE International 2006) Figure 5.8: Comparison of pilot and gaseous fuel ignition delay times for natural gas, 10% and 23% H 2 . 800 RPM, 6 bar GIMEP, 0.5 <j), 0.175 Y i n t 0 2 , Injection Pressure = 16 MPa. (reprinted with permission from SAE 2006-01-0653. © SAE International 2006) -108--40 -20 0 20 40 60 -20 -10 0 10 20 30 Crank Angle ( °CA) Crank Angle ( °CA) Crank Angle ( °CA) Crank Angle ( °CA) Figure 5.9: Pressure trace and estimated heat release rates, compar isons between natural gas, 1 0 % and 2 3 % Ff 2 . 800 R P M , 6 bar G I M E P , 0.5 <|>, 0.175 Y i n t o 2 , Inject ion Pressure = 16 M P a . (reprinted with permission from SAE 2006-01-0653. © SAE International 2006) 4.0 3.5 3.0 X CO 2.5 E s- 2.0 CL 1.5 > o 1.0 o 0.5 0.0 5 10 50%IHR (°CA) . A . - NG - » - 1 5 % H 2 - x - 3 5 % H 2 - ^ - C H 4 A s / A • .' t X 5 10 50%IHR (°CA) 15 20 30 ~ 25 < ° " 20 s 1 5 2 10 Q m 5 . . . A . . . N G - ^ 1 5 % H 2 - x - 3 5 % H 2 - ^ - C H 4 5 10 50%IHR (°CA) 15 20 . . . A . . . N G 15% H 2 x - 3 5 % H 2 ^ - C H 4 5 10 50%IHR (°CA) 20 Figure 5 .10: C o m b u s t i o n performance compar ison for natural gas, methane, 1 5 % and 3 5 % H 2 . 1200 R P M , 13.5 bar G I M E P , 0.6 <>, 0.19 Y i n t o 2 , In ject ion Pressure = 21 M P a . - 1 0 9 -250 50%IHR @ 5°ATDC 15% H2_ y Methane Natural Gas 250 < 200 p 150 E & 100 or a: x 50 50%IHR@ 15°ATDC xfr- Natural Gas 35% H2 * / 15% H 2 >k -10 0 10 Crank Angle (°CA) 20 -10 30 0 10 20 Crank Angle (°CA) Figure 5.11: Net heat release rates for natural gas, methane, 15% and 35% H2 at 50%IHR of 5 and 15°ATDC 1200 RPM, 13.5 bar GIMEP, 0.6 <|>, 0.19 Y i n t o2, Injection Pressure = 21 MPa. 5 10 15 0 5 10 50%IHR (°CA) 50%IHR (°CA) Figure 5.12: Pilot and gaseous fuel ignition delay comparison for natural gas, methane, 15% and 35% H 2 . 1200 RPM, 13.5 bar GIMEP, 0.6 <|>, 0.19 Y i n t o2, Injection Pressure = 21 MPa. 0 7 0.6 1,0 5 8 0.4 0 3 -h NG 10% H 2 2 3 % H2 - I 10 50%IHR ( °CA) 15 20 •o 0.7 S 3 0.6 o> 2 1 0.5 I § 0.4 O I U 0.3 —A— NG 10% H 2 - ^ - 2 3 % H 2 Tr ...A 10 50%IHR ( ° C A ) 15 20 Figure 5.13: C 0 2 and net GHG emissions for natural gas, 10% and 23% H 2 . 800 R P M , 6 bar GIMEP, 0.5 <)>, 0.175 Y i n t o2, Injection Pressure = 16 MPa. -110-0.6 0.3 A NG --©-- C H 4 - » - 1 5 % H 2 -x -35% H2 5 10 50%IHR (°CA) 15 20 0.6 C3 O O 0.3 -A - - NG - ^ - C H 4 15% H 2 x - 35%H 2 5 10 50%IHR (°CA) 15 20 Figure 5.14: CO2 and net GHG emissions for natural gas, methane, 15% and 35% H 2 . 1200 R P M , 13.5 bar GIMEP, 0.6 <j), 0.19 Y i n t o2, Injection Pressure = 21 MPa. - I l l -Chapter 6 Fuel Composition 6.1 Introduction Like many fuels, natural gas is a mixture of various hydrocarbon molecules. Commercial grade natural gas compositions vary from 70-95% methane (CH4), with the balance composed of heavier hydrocarbons (primarily ethane, C2H6, and propane, C3H8) as well as diluents such as N2 and CO2 (see Appendix 5). This chapter focuses on the role of the heavy hydrocarbons (ethane and propane); the effects of diluting the fuel with nitrogen will be discussed in Chapter 7. By studying the effects of varying the quantity of C2H6 and C3H8 in the fuel, it should be possible to identify the sensitivity of the combustion system to these species. Specifically, knowing whether the heavy hydrocarbons promote the gaseous ignition process will help to identify whether that process is limited by chemical kinetics or by mixing. Furthermore, by identifying the sensitivity of the particulate matter formation processes to the concentrations of these heavy hydrocarbons, their contribution to P M emissions can be identified. These objectives are important not just from an academic standpoint but also operationally, given that natural gas compositions vary significantly by season, geographical area, and supplier. As a result, understanding the effects of varying fuel compositions on both engine performance and emissions is vital. As well, as more effort is focused on non-conventional gases (such as coal-bed methane and biologically derived synthetic gases), improved understanding of the sensitivity of the combustion process to variations in fuel composition is required. Understanding these effects will also be critical for engine developers and regulators to ensure that engines (and fuels) can meet current and future emissions standards. 6.2 Previous W o r k The effect of fuel composition on the combustion process and on the emissions from natural gas fuelled engines has been addressed in both fundamental and applied studies. As the majority of in-service internal combustion engines are premixed charge spark ignition types (see Chapter 2), these have received the most research attention. However, some -112-fundamental research has also been carried out on the non-premixed autoignition of natural gas under conditions similar to those found in a heavy-duty diesel engine. For premixed charge combustion, the greatest influences of ethane and propane on the combustion of natural gas are in the, ignition and early combustion phases. The ignitability of the mixture is enhanced, due primarily to pre-combustion increases in the concentration of H, OH, and HO2 radicals from decomposition of the heavier hydrocarbons [60,208]. The formation of hydrocarbon radicals, including C2 species such as the ethyl radical (C2Hs) and acetylene (C2H2), is also enhanced, although formation of these hydrocarbons also occurs in reactions involving only pure methane, through recombination of the methyl (CH3) radical [209]. The importance, of the H, O, and HO2 radicals appears to apply mainly at higher temperatures; at low temperatures, significant formation of the methylperoxy (CH3O2H) radical by either ethane or propane decomposition appears to be the primary mechanism for enhanced ignition [210]. At these low temperatures, ethane and propane have very similar ignition promoting effects, as the two fuels have equivalent CH2O2H formation rates. At higher temperatures (>1200K) the formation of H radicals from propane and of OH from ethane appear to be the dominant ignition enhancing mechanisms [210]. The fundamental and applied studies available from the literature generally agree that significant enhancement of the ignition process is achieved with ethane or propane addition due to enhanced radical formation. The improved ignitability of the fuel-air mixture has a direct impact on engine system development. For example, the spark energy required to ignite the mixture is reduced. However, the propensity for auto-ignition of the fuel/air mixture (engine knock) increases. Thus, more advanced spark timings are required to avoid knock in the presence of significant quantities of ethane or propane in spark-ignited engines [60,211]. Conversely, the auto-ignition process is used as the desired ignition source in homogeneous charge compression ignition (HCCI) engines; while ethane and propane addition can improve the ignitability of the fuel-air mixture, varying quantities of these species pose a significant challenge for HCCI engine control [212]. While the effects of ethane and propane addition to natural gas are greatest on the ignition process, flame propagation and pollutant formation are also affected. By promoting radical formation in the combustion process, the heavier hydrocarbons enhance combustion -113-stability and extend the lean combustion limit [213,214]. The flame's resistance to turbulent stretch (which can lead to local extinction events) is also enhanced, as is the flame propagation rate [215]. The effects of hydrocarbon addition on emissions are not as repeatable between studies as the ignition and combustion results. In some studies, N O x emissions are slightly increased as a result of higher flame temperatures [216]; however this effect is not consistently observed [217]. No significant effects on CO emissions are reported in most of the studies; however, some researchers report reductions in HC emissions with increased ethane and propane concentrations [218]. As well as variations in reactivity of the fuel-air mixture, changes in the energy density of the fuel also affect the overall engine performance. Even when the composition varies, little influence on performance is observed at constant Wobbe indices (ratio of fuel energy content to specific gravity) [213,219]. Higher Wobbe values tend to result in reduced hydrocarbon emissions and higher engine power levels [220]. The development of closed-loop engine control [219] and fuel quality sensors [221] has allowed spark-ignition engines to operate over a wide range of gaseous fuel compositions with minimal impacts on engine performance. Typically, impacts on emissions for in-use vehicles are minimized through the use of efficient exhaust catalyst systems. The principal influence of natural gas composition on non-premixed combustion is also on the ignition process. At high temperatures (>1400 K) fuel additives have little effect as this process is mixing limited [222]. At lower temperatures, the addition of either ethane or propane is found to reduce ignition delay times by as much as 0.7 ms. However, there is a limit to the effectiveness of improved kinetics, especially at higher temperatures; beyond a certain point mixing limitations dominate the ignition processes. The shorter ignition delay time has also been identified as a potential source for increased N 0 X emissions with ethane addition to the fuel [89]; however, substantially more work is required to understand the effects of the heavier hydrocarbons on the ignition and pollutant formation mechanisms of a natural gas engine using a non-premixed direct-injection combustion system. 6.3 Experimental Information UBC' s single-cylinder research engine was used to investigate the effects of heavier hydrocarbons on a direct-injection of natural gas engine. The facility was in most respects unmodified from that described in Chapter 3. The re-entrant bowl-in-piston geometry, with a -114-17:1 compression ratio, was used for all the testing. Modifications were also made to the fuel supply system to permit the addition of ethane and propane additives to the standard building-supplied natural gas. To prepare the fuel blends, four large high-pressure natural gas storage tanks (internal volume 115L/tank) were emptied of gas. .Measured amounts of the specific additives from compressed gas supply bottles (all industrially-supplied, 99.9% purity grade 2.0 or better) were put into the storage tanks. The tanks were then refilled with commercial natural gas and allowed to sit for at least 2 days prior to use, to ensure that the fuel mixture was homogenous. Composition control during mixing was achieved through the use of partial pressures, while gas chromatograph analysis was carried out on the blended gas (sampled just upstream of the engine part way through the testing) to verify the actual fuel composition. The fuel compositions reported in this work refer exclusively to the gas chromatograph measurements. Four blends were prepared and compared to standard commercial natural gas. These were low and high C2H6 (targets were 5% and 10% by volume) and low and high C3H8 (targets 2% and 4% by volume). These concentrations were selected to minimize the chance of condensation of the hydrocarbons after the blend was compressed to full system pressure (up to 30 MPa). Although the gas composition varied, the fuel flow rate measurements were not affected as the flow meter (MicroMotion coriolis force) measured mass flow directly. The compositions of the various blends, as analysed at Westport Innovations Inc., are shown in Table 6.1. In some cases, and in particular for the C2H6 case, the actual compositions were significantly different from the target compositions. As the maximum concentrations were still well below the saturation partial pressures, and as the original target concentrations were arbitrarily selected, the final conclusions are not affected by this discrepancy in composition. 6.3.1 Operating Condition The engine operating condition selected for this testing was at high-load but moderate speed, approximating the peak-torque condition for the ISX series engine. This point was very similar to the second-round engine condition investigated in Chapter 5 (Table 5.1). The constants and variables for this operating condition are outlined in Table 6.2. The only variable was combustion timing, which was controlled (based on the midpoint of the heat release, 50%IHR, as discussed in section 3.3.1) from 0° to 15° after top-dead-center (ATDC). A l l other parameters, including the oxygen equivalence ratio ((j)), speed, and intake oxygen -115-mass f ract ion (Y j n t 02) were f i xed . Th i s process m i n i m i z e d the var iab i l i ty i n the operating cond i t ion w h i l e m a x i m i z i n g the testing t ime avai lable w i t h the l im i ted quantity o f fuel avai lable . B y vary ing the combust ion t i m i n g , a range o f combust ion condit ions was achieved without s igni f icant ly affect ing the air -exchange parameters. 6.3.2 Experimental Process F o r a l l the fuel b lends, at least three repetitions o f each test cond i t ion were carr ied out. F u l l randomizat ion o f the testing was not possib le , as only one gas b lend c o u l d be prepared at a t ime. The test sequence was as f o l l o w s : First , a f u l l set o f natural gas tests was per formed. T h e n , to m i n i m i z e the number o f t imes the tanks needed to be empt ied , the tests w i t h the h igh concentrat ion gas b lend o f each species was conducted, f o l l o w e d by the tests w i t h the l o w concentrat ion gas mixture . T o complete the sequence, the f u l l set o f natural gas tests was repeated. C o m p a r i n g the natural gas results f r o m before and after the test sets showed no s igni f icant var iat ion i n operating cond i t ion setting, emiss ions or performance. The sequence o f the test points was fu l l y randomized w i t h i n each fuel compos i t ion . A s a result o f the lack o f randomizat ion or repl icat ions between b locks , statistically v a l i d tests between the fuel compos i t ions were not possib le , as the inf luence o f fuel compos i t i on c o u l d not be differentiated f r o m day - to -day var iabi l i ty . H o w e v e r , the fact that no s igni f icant differences i n performance or emiss ions were detected between the in i t ia l and f ina l natural gas tests indicates that no systematic errors were present w h i c h w o u l d have biased the results. 6.4 Results The role o f the heavier hydrocarbons i n the overal l combust ion event is o f part icular interest because o f their vary ing concentrations i n natural gas. Prev ious results have found that the ign i t ion process, combust ion rate, and emissions format ion are a l l affected by fuel compos i t ion . These preceding studies relate p r imar i l y to p remixed combust ion ; the effects on direct injected natural gas under condit ions s imi la r to those o f a modern diesel engine have not been prev ious ly investigated i n detai l . 6.4.1 Combustion Effects One o f the pr inc ipa l effects o f the presence o f ethane and propane i n the natural gas is o n the ign i t ion process, as s h o w n i n F igure 6.1. The p i lot ign i t ion delay t ime (see sect ion 3.3.1 for the ign i t ion t i m i n g def init ions) is unaffected by the addi t ion o f ethane or propane. - 1 1 6 -Th is is expected, g i ven that the p i lot fuel compos i t ion is no different and that p i lo t ign i t ion occurs before the start o f gas in ject ion. The gas ign i t ion delay t ime, however , is reduced by as m u c h as 1 0 % w i t h either propane or ethane added to the gaseous fue l . The addi t ion o f s igni f icant quantit ies o f either hydrocarbon s igni f icant ly increases the ign i tabi l i ty o f the natural gas by increasing the concentrat ion o f speci f ic radicals [210]. A s the diesel p i lot ignites the gaseous jet, the gaseous star t -of -combust ion is not truly an auto - ign i t ion process. H o w e v e r , the presence o f more reactive radicals in the gaseous jet w i l l increase the ign i tabi l i ty o f the jet. Ignit ion can thus occur at lower temperatures and over a w ider range o f fue l -a i r s to ichiometr ics than w i t h pure methane. In compar ison , there is no consistent dif ference i n ign i t ion delay between the pure methane and natural gas cases (see section 5.3.3) . H o w e v e r , this is most l i ke l y a result o f an inabi l i ty to resolve the effects o f the re lat ively s m a l l concentrations o f ethane and propane i n the natural gas. The fact that both l o w and h igh ethane cases show the same reduct ion i n the gaseous ign i t ion delay suggests that the upper l i m i t to the ign i t ion -p romot ing effects o f ethane may be be ing approached. Th is c o n c l u s i o n agrees w i t h previous work , w h i c h suggests that once the c h e m i c a l (k inet ic -related) component o f the ign i t ion delay is m i n i m i z e d , a s igni f icant delay remains due to the t ime required for the fuel and air to m i x to a combust ib le stoichiometry [222]. F o r both ethane and the h igh propane cases, the gaseous ign i t ion delay is s imi lar . Th i s suggests that the potential benefits o f i m p r o v i n g kinet ics have been achieved, and that the pr imary factor affect ing the ign i t ion delay at this condi t ion is phys ica l relat ing to the m i x i n g o f the fuel and air. A s igni f icant effect o f the shorter gas ign i t ion delay t ime is a reduct ion i n the p r e m i x e d component o f the combust ion event (Figure 6.2). The reduct ion i n peak heat-release rate indicates that the amount o f energy released dur ing the early part ia l ly p r e m i x e d burn phase is dependent o n the amount o f fue l injected and m i x e d to a combust ib le mix ture pr ior to ign i t ion . W i t h the shorter ign i t ion delay, less combust ib le fue l -a i r mix ture is avai lable w h e n ign i t ion occurs and hence the peak energy release is reduced. The overal l burn durat ion is also increased by a corresponding amount (approximately 5 % , F igure 6.3), result ing f r o m the earl ier ign i t ion combined w i t h constant end -o f - combust ion t i m i n g . (The e n d - o f - c o m b u s t i o n t i m i n g is designated i n F igure 6.3 by the crank angle at w h i c h 9 0 % o f the fuel energy has been released, 9 0 % I H R . ) These relat ively smal l effects, i n c l u d i n g the - 1 1 7 -reduct ion i n peak heat-release rate and earl ier s tar t -o f -combust ion t i m i n g , are also shown i n the pressure trace and net heat release rates shown (for a single combust ion t iming) i n F igure 6.4. The pressure traces at the other t imings indicate s imi la r trends; these m a y be found i n A p p e n d i x 6. These observations indicate that the p r inc ipa l factors in f luenc ing the combust ion event are the in ject ion and m i x i n g rates; the inf luence o f i m p r o v e d k inet ics is l im i ted to enhancing the ign i t ion process. The importance o f combust ion stabi l i ty on performance and emiss ions is demonstrated i n Chapter 5, w i t h hydrogen s igni f icant ly enhancing combust ion stabi l i ty . W i t h the hydrocarbon addit ives, the ign i t ion process is enhanced, but overa l l combust ion stabi l i ty , represented by the C O V o f the G I M E P , is not s igni f icant ly affected (Figure 6.5). A t the operating cond i t ion tested, the combust ion stabi l i ty is general ly very good ( < 1 % C 0 V ) for a l l the fuel condi t ions . O n l y one cond i t ion , w i t h l o w propane addi t ion , has a C O V consistently higher than the other cases. The reasons for this effect are unclear, but this m a y have a s igni f icant impact on emiss ions for this fue l b lend. The potential for ethane or propane addi t ion to improve stabi l i ty at condit ions w i t h h igh combust ion var iab i l i ty was not investigated. Prev ious results [60] have suggested that fue l -convers ion ef f ic iency is in f luenced by the presence o f heavier hydrocarbon addit ives. In this work , the ef f ic iency is represented by the spec i f ic fue l consumpt ion ( G I S F C ) , shown i n F igure 6.6. F o r a l l cases, the fuel consumpt ion is calculated o n an energy-equivalent mass o f diesel basis (see sect ion 3.3.3) . The addi t ion o f propane s l ight ly reduces the fuel consumpt ion , w h i l e the ethane s l ight ly increases it. H o w e v e r , these effects are relat ively s m a l l , and i n general the convers ion o f c h e m i c a l energy i n the fuel to p iston w o r k is independent o f fuel c o m p o s i t i o n over the range o f compos i t ions and the engine condit ions tested. Th is is not surpr is ing, g i ven the relat ively smal l in f luence o f the addit ives o n the overal l combust ion progression ( indicated by the i n -cy l inder pressure and heat-release rate, F igure 6.4). 6.4.2 Gaseous Emissions A d d i n g propane or ethane to the gaseous fuel has a s igni f icant inf luence on some o f the measured emiss ions . These effects are, i n most cases, more s igni f icant than w o u l d have been expected f r o m the relat ively smal l changes i n the observed combust ion progression. Interestingly, The hydrocarbon addit ives d i d not s igni f icant ly inf luence the N O x emiss ions - 1 1 8 -( shown i n F igure 6.7). Th is cou ld be expected g iven that the adiabatic f lame temperature is not s ign i f icant ly affected (Table 6.1) and the peak combust ion intensity (F igure 6.2) is on ly s l ight ly reduced. The H C emiss ions are, i n general , s l ight ly reduced w i t h the addi t ion o f ethane or propane to the gaseous fuel (Figure 6.8). In a l l cases, the hydrocarbon emiss ions are indist inguishable f r o m the exhaust methane measurements, suggesting that at least 9 5 % o f the exhaust H C are unburned methane. The observed reduct ion in unburned fuel emiss ions m a y have been a result o f reduced over - leaning due to the shorter gaseous ign i t ion delay p rov ided by the ethane and propane addit ives. The fact that the l o w propane case shows no s igni f icant improvement in H C emissions may be related to the higher combust ion var iabi l i ty observed at this condi t ion (Figure 6.5). For most o f the condit ions tested, the largest reductions i n H C emiss ions occur at the latest t i m i n g , w h i c h is the on ly cond i t ion w i t h s igni f icant reductions i n the combust ion var iabi l i ty compared to the natural gas fue l . S i m i l a r to the H C emiss ions , the effects o f fuel compos i t ion o n the emiss ions o f C O (Figure 6.9) are sensit ive to combust ion t i m i n g and fuel compos i t ion . The general trend w i t h t i m i n g - reaching a m a x i m u m at intermediate t imings , w i t h lower emiss ions at both early and late t imings - is contrary to prev ious results at lower loads, as shown both i n Chapter 5 and i n some prev ious w o r k [135], as w e l l as d i f fer ing f r o m convent ional d iesel -engine emiss ion format ion w i s d o m [28]. However , some other previous results suggest that, at mid - speed and h i g h - l o a d condi t ions , C O emissions may be higher at intermediate t imings [130,134] . In the prev ious w o r k , this effect was attributed to interactions between the burn ing gas jet and the p iston, w i t h both earl ier and later t imings result ing i n less interact ion, due to changes i n charge m o t i o n , p is ton locat ion , and combust ion rate. In general , the addi t ion o f ethane or propane to the fuel increases C O emiss ions . A t the latest combust ion t i m i n g , the C O emiss ions f r o m the natural gas fue l l i ng case are substantial ly higher than most o f the fuel addit ive cases; this may be a result o f the higher combust ion var iab i l i ty encountered at this point (Figure 6.5). The combust ion var iab i l i t y also appears to inf luence the l o w propane addi t ion case's C O emiss ions levels , w h i c h are s igni f icant ly higher than the emissions at the high-propane fue l l ing cond i t ion . These results p rov ide further evidence that combust ion instabi l i ty is one factor that can contribute s ign i f icant ly to C O emiss ions. However , whether the increased var iab i l i t y is a result o f - 1 1 9 -compet ing mechanisms i n the propane ox idat ion process or s i m p l y due to day - to -day performance or measurement var iabi l i ty cannot be conc lus ive ly determined. In general , C O emiss ions are substantial ly increased through the addi t ion o f heavier hydrocarbons to the natural gas. A s the combust ion stabi l i ty is not s igni f icant ly in f luenced, the results suggest that the inf luence o f the heavy hydrocarbons on C O emissions is through the chemica l k inet ics either increasing C O format ion or reducing its ox idat ion to CO2. A s the late stages o f the combust ion event and the overal l oxygen concentrations i n the combust ion chamber are general ly unaffected by the fuel compos i t ion , it is l i ke l y that the addi t ion o f the heavy hydrocarbons is enhancing the chemica l format ion o f C O . Th i s agrees w i t h basic understanding o f C O format ion i n methane f lames, w h i c h tends to be enhanced by the presence o f ethane or propane i n the fuel [37,217]. 6.4.3 Particulate Matter The P M total mass emiss ions , F igure 6 .10, demonstrate trends general ly s imi la r to those o f the C O emiss ions . The addit ion o f ethane or propane tends to substantial ly increase P M emiss ions , w h i l e higher P M levels are also observed at the intermediate combust ion t imings . Increases in P M w i t h the hydrocarbon addit ives are greater than a factor o f two at certain condi t ions , demonstrat ing the very strong inf luence o f ethane and propane on the P M emiss ions . The fact that the l o w propane addit ive case was indist inguishable f r o m the h igh propane case m a y be a result o f effects s imi la r to those that induced the higher C O emiss ions at this cond i t ion , as discussed i n the preceding section. The P M was most s igni f icant ly increased w i t h the h igh ethane concentrat ion, w h i c h also has the lowest H : C ratio o f any o f the fuels (Table 6.2). The increases i n emiss ions are generally consistent for the intermediate and late t i m i n g s ; on ly at the earliest t imings are substantial ly smal ler increases observed. Th i s suggests that at the earliest combust ion t imings , the greater residence t ime i n the h i g h -temperature post - combust ion gases m a y result i n more P M being o x i d i z e d . A t the intermediate t imings , impairment o f late-cycle ox idat ion through interact ion between the burn ing fue l -a i r mixture and the p iston may result in premature quenching o f the combust ion . S ince the unburned fuel emissions are not increased (Figure 6.8) it w o u l d appear that it is the very- late stage combust ion w h i c h is be ing impai red . A p lausib le exp lanat ion is that the major i ty o f the methane w h i c h was go ing to burn has already been consumed ; on ly part ial combust ion products ( C O , P M ) are not be ing fu l l y ox id i zed . - 1 2 0 -Further insight into the P M formation and oxidation processes with the addition of ethane and propane to the natural gas can be identified from the particle size distributions (Figures 6.11 and 6.12). These demonstrate general agreement with the measured mass distributions, with no discernable difference in number concentrations between the low and high propane cases, although both are greater than the natural gas case. For ethane addition, substantial increases are observed between the natural gas, low ethane, and high ethane cases. In general, a shift towards both a greater number of particles and larger diameters is observed, consistent with accepted theories of particulate growth during the dilution process [6]. If a larger number of particles are formed initially, there will be a greater likelihood of these particles combining into agglomerates and thereby forming longer chains. This process will tend to reduce the individual number of particles measured; therefore, the observation of a larger total number of particles, along with a larger mobility diameter, suggests that the number of particles formed in the initial combustion event is substantially increased. This leads to more agglomeration resulting in the observed increase in particle size. These results are confirmed by T E M images (Appendix 7), which do not indicate a difference in primary particle size for different fuel compositions. None of the conditions tested show a significant nucleation-mode peak, indicating that the particle surface areas are consistently large enough to ensure that adsorption of volatiles occurs preferentially to the post-exhaust nucleation of new volatile particles. The relative fraction of volatile material in the particle mass, measured using an Aethalometer, is shown in Figure 6.13. The black carbon content of the exhaust stream increases substantially at all timings with the addition of either ethane or propane to the fuel. The black carbon fraction (BC mass / P M total mass), shown in Figure 6.14, also increases at the earlier injection timings with the fuel additives. However, at the later timings, this ratio is independent of the fuel composition. This indicates that the mass of both BC and of volatiles in the P M increases with the addition of the heavier hydrocarbons to the methane fuel. The estimated mass rate of volatile emissions (PM total mass less BC mass) is shown in Figure 6.15; the large error bars are a result of the combined uncertainty from the total mass and BC measurements. One possible explanation for the larger volatile concentration is that the gaseous fuel is contributing significantly to volatile emissions. As discussed in section 2.2.5, the formation pathways for carbonaceous P M involve the formation of many polycyclic -121-hydrocarbons; not all o f these species w i l l form P M during the combustion event, and i f they are not oxidized, they w i l l be emitted as volatiles which could condense post-exhaust. Another potential explanation is that the increased surface area of post-combustion particulate adsorbed substantially more of the volatile materials, which would otherwise be emitted in the vapour phase. 6.5 D i s c u s s i o n Ethane or propane addition to the natural gas results in a significant increase in P M . A s shown in the previous chapter (Chapter 5) the use of pure methane, compared to commercial natural gas, has only a small effect on P M emissions (Figure 5.2). A s suggested by reaction kinetics analyses, a significant fraction of methane may combine (via the methyl radical, CH3) to form either ethane (C2H6) or ethyl (C2H5), most of which w i l l pass through acetylene (C2H2) as they dissociate [223]. B y increasing the ethane concentration, the initial response may be a reduction in the amount of ethyl recombination forming ethane. A s a result, small quantities of ethane in the fuel may not significantly increase the formation of acetylene. However, as the ethane quantity increases further, acetylene formation is enhanced as the fuel-side concentration of ethane becomes larger than the amount which would have been formed via ethyl recombination. This may explain both the relatively small differences in P M between pure methane and natural gas and the much larger differences between natural gas and higher ethane and propane additives. The addition of either propane or ethane to the natural gas fuel results in reductions in the global hydrogen to carbon (H:C) ratio in the fuel (Table 6.1). Whether the global H : C ratio or kinetics relating to the specific hydrocarbon molecule has a greater influence on P M emissions may be deduced from the plots of total P M mass and black carbon mass as a function of H : C ratio, shown in Figures 6.16 and 6.17 for two timings (50% IHR at 5 and 15°ATDC). For both timings, the particulate mass increases dramatically at H : C ratios less than 3.9, but decreases relatively slowly at ratios greater than 4 (pure methane and the hydrogen results from Chapter 5). Similar results are observed for the black carbon mass, providing further indication that the gaseous fuel has a significant impact on carbonaceous particulate formation. The right-hand plot in both figures (6.16 & 6.17) shows the H : C ratio effects on P M and B C for ratios between 3.8 and 4.0 (pure methane, natural gas, ethane and propane addition cases). These indicate that the propane (for an equal H : C ratio) generates -122-significantly more total and black carbon P M than the ethane; this suggests that the H:C ratio is not the only factor relating to P M formation'. The results indicate that relatively small quantities of heavy hydrocarbons are not substantially influencing P M emissions. However, as concentrations increase above the baseline level, significant increases in P M emissions are observed. One technique that could mitigate these increases would be to change the engine's operating mode (either through the use of more advanced timing or lower EGR levels); however, this would tend to increase N O x emissions. Another option would be to consider adding a small amount of another additive to the fuel to offset the observed increases in P M and CO. Hydrogen is one obvious alternative, as was shown in Chapter 5. Some natural gas fuels do contain significant quantities of diluents, including nitrogen; i f the level of these diluents could be used to offset the effects of propane or ethane addition, a low-cost technique for achieving low emissions over a wide range of fuel compositions could be achieved. Chapter 7 investigates the effect of fuel dilution with nitrogen on emissions and performance, as well as providing further discussion of the effects of fuel composition. 6.6 Conclusions 1. CO and P M emissions are generally increased with the addition of ethane and propane to the fuel. The substantial increases in P M are attributed to the formation of acetylene as a reaction intermediary of the decomposition of ethane and propane [223], leading to increased polycyclic aromatic hydrocarbon formation and higher levels of P M . 2. The global H:C ratio of the gaseous fuel generally predicts the observed trends in both P M and black carbon. However, consistent differences in P M levels between the ethane and propane cases at equivalent H:C ratio suggests that kinetic effects relating to the specific molecules play a significant role in the formation of black carbon and total P M . 3. The addition of either ethane or propane to the natural gas increases the ignitability of the gaseous mixture. However, increasing the ethane from moderate to high levels did not reduce ignition delay any further; this indicates that kinetic limitations have been minimized and that the ignition process has become dependent on the time for the gaseous jet to mix to a combustible stoichiometry. 4. The shorter gaseous-fuel ignition delay results in a smaller amount of fuel having premixed to a combustible level, resulting in a reduced peak heat-release rate. This -123-results in a longer duration for the gaseous fuel combustion, suggesting that the overall duration of the combustion event is limited by the rate at which the fuel and air mix. 5. The addition of heavier hydrocarbons did not significantly affect N O x emissions. HC emissions were slightly reduced, possibly due to small improvements in combustion stability resulting from the ethane and propane in the fuel. 6. The effects of ethane and propane in the fuel on P M emissions have significant implications for engine development. Aftertreatment may be required to ensure that ultra-low P M emissions levels are retained over the range of fuel compositions that may be experienced. An alternative would be to provide more careful control over the composition of natural gas fuels supplied to vehicles. 6 .7 T a b l e s a n d F i g u r e s Table 6.1: Fuel composition as measured by gas chromatography Natural Low High Low High Gas Ethane Ethane Propane Propane n-Butane 0.05 0.05 0.04 0.08 0.08 i-Butane 0.07 0.07 0.06 0.09 0.09 i-Pentane 0.02 0.02 0.02 0.02 0.03 n-Pentane 0.02 0.02 0.01 0.02 0.02 Hexane 0.01 0.01 0.00 0.01 0.01 Heptane 0.00 0.00 0.00 0.00 0.00 Octane 0.00 0.00 0.00 0.00 0.00 Nitrogen 1.47 1.09 1.19 2.02 2.82 Methane 96.6 91.43 89.39 93.5 91.3 Carbon Dioxide 0.44 0.40 0.34 0.41 0.38 Ethane 1.41 6.60 8.62 1.59 1.50 Propane 0.32 0.32 0.28 2.23 3.73 M W (kg/kmol) 16.70 17.37 17.63 17.34 17.85 L H V (kJ/kg) 48080 48279 48220 47540 46883 H H V (kJ/kg) 53313 53462 53372 52658 51891 Tadiabatic (K) 2485 2530 2550 2495 2510 H:C ratio 3.94 3.88 3.86 3.91 3.88 -124-Table 6.2 Operating condition constants and variables Parameter Value Speed (RPM) 1200 G I M E P (bars) 13.5 <k>2 0.6 E G R (mass %) 30 Yi„t02 0.19 Fuel Pressure (MPa) 21 50%IHR ( ° A T D C ) 0, 5, 10, 15 Pilot (mg/inj) 5 5 10 15 50%IHR (°CA) 5 10 50%IHR (°CA) - NG low ethane - - X - - high ethane 14 12 O o g O 10 - N G low propane - - x - - h i g h propane T > f ' ^  » •5 i-- ] 5 10 50%IHR (°CA) 15 20 -5 15 20 0 5 10 50%IHR (°CA) Figure 6.1: Pilot (PID) and gas ignition delay (GID) times (relative to commanded start-of-injection) for ethane (left) and propane addition (right) additives, compared to natural gas fuelling. -125-100 250 < O220 E 190 or 160 or 3 130 0) a. 100 •NG • Q - low propane - - x - high propane 5 10 50%IHR (°CA) 15 20 15 20 -5 0 5 10 50%IHR (°CA) Figure 6.2: Peak heat-release rate for ethane (left) and propane (right) additives, compared to natural gas fuelling. 5 10 50%IHR (°CA) 5 10 50%IHR (°CA) 5 10 50%IHR (°CA) 15 20 5 10 50%IHR (°CA) Figure 6.3: Burn duration and end-of-combustion (EOC) timing (90%IHR) for ethane (left) and propane (right) additives, compared to natural gas fuelling. -126-200 •c- 150 CD § 100 CO CO °- 50 0 250 5 0 % I H R @ 10°ATDC 5~ 200 p ^ 150 S 100 or or x 50 0 -20 natural gas - low ethane -high ethane -30 -15 0 15 30 Crank Angle (°CA) 45 x natural gas — low ethane •high ethane 0 20 Crank Angle (°CA) 250 natural gas - low propane - high propane -15 0 15 30 Crank Angle (°CA) 45 0 20 Crank Angle (°CA) Figure 6.4: In -cy l inder pressure trace and net heat-release rate for the m i d - t i m i n g cond i t ion ( 5 0 % I H R at 1 0 ° A T D C ) for ethane (left) and propane (right) addit ives, compared to natural gas fue l l ing . NG low ethane --x--high ethane - N G • D - - low propane --x--high propane 5 10 50%IHR (oCA) 5 10 50%IHR (oCA) — . — N G --a- low ethane --x-- high ethane / .1 r 5 10 50%IHR (°CA) 15 20 5 10 50%IHR (°CA) 15 20 Figure 6 .5 : C o m b u s t i o n stabil ity ( C O V o f G I M E P and C O V o f peak cy l inder pressure) for ethane (left) and propane (right) addit ives, compared to natural gas fue l l ing . - 1 2 7 -?00 t3 80 > s 3 60 O LL CO O 40 - N G •••o-- low propane high propane 5 10 15 50%IHR (°CA) -5 15 20 0 5 10 50%IHR (°CA) Figure 6 .6 : G I S F C w i t h ethane (left) and propane (right) addit ives, compared to natural gas fue l l ing . 5 10 15 20 -5 0 5 10 15 5 0 % I H R ( ° C A ) 5 0 % I H R ( ° C A ) Figure 6 .7 : N O x emiss ions as a funct ion o f t i m i n g , for ethane (left) and propane (right) addit ives, compared to natural gas fue l l ing . - N G o low ethane - -X-- high ethane - N G - n - l o w propane --x--high propane 0 5 10 15 20 - 5 0 5 10 15 50%IHR (°CA) 50%IHR (°CA) Figure 6 .8 : Hydrocarbon emissions as a funct ion o f t i m i n g , for ethane (left) and propane (right) addit ives, compared to natural gas fue l l ing . - 1 2 8 -5 _ 4 i s b u> 2 o o - N G low ethane - -x--high ethane > s -1 • 1 10 15 20 5 10 50%IHR (°CA) 50%IHR (°CA) Figure 6 .9 : C O emiss ions as a funct ion o f t i m i n g , for ethane (left) and propane (right) addit ives, compared to natural gas fue l l ing . - 5 - N G • ••D-- low ethane - - X - - h i g h ethane \ C ,\ / / / / _..[ I \ / / . ' \ T 10 15 20 0.06 0.05 I 0.04 b 0.03 3 5 0.02 o_ 0.01 0.00 - N G •••a- low propane - - X - - high propane -5 5 10 50%IHR (°CA) 15 20 50%IHR (°CA) Figure 6 .10: P M emiss ions as a funct ion o f t i m i n g , for ethane (left) and propane (right) addit ives, compared to natural gas fue l l ing . - 1 2 9 -o 1.E+08 1.E+07 1.E+06 5* CL a c 1 1.E+05 •a 1.E+04 1.E+08 3 1.E+07 0°ATDC JT VM • natural gas • low ethane x high ethane / . _ L . / _ • „ x " x *& 10 100 Mobility Diameter (nm) 1.E+06 1.E+05 1.E+04 10°ATDC & * M \ • natural gas • low ethane x high ethane * $ 4, .* < .^«#° •rf Sc • 1.E+08 I 1.E+07 *j 1.E+06 CL Q I 1.E+05 1.E+04 1.E+08 | 1.E+07 •"E g 1.E+06 a. Q c I 1.E+05 1.E+04 5 A T D C • natural gas o low ethane x high ethane 10 100 Mobility Diameter (nm) 1000 15°ATDC / \\ • natural gas ° low ethane x high ethans X — , . 10 100 Mobility Diameter (nm) 1000 10 100 Mobility Diameter (nm) 1000 Figure 6 .11: Part ic le size distr ibutions for natural gas and two levels o f ethane addi t ion at four combust ion t imings (c lockwise f r o m top left: 0, 5, 10, 1 5 ° A T D C ) -130-1.E+08 i 1.E+07 | 1.E+06 2 1.E+05 TJ 1.E+04 1.E+08 | 1.E+07 1 g 1.E+06 Q. a c 1 1.E+05 TJ 1.E+04 0 A T D C • natural gas ° low propane x high propane 1.E+08 I 1.E+07 10 100 Mobility Diameter (nm) 1000 10 A T D C • natural gas ° low propane x high propane 2 1.E+05 TJ 1.E+04 1.E+08 I 1.E+07 5 4 ^ 1.E+06 Q. Q c I 1.E+05 TJ 1.E+04 5 A T D C • natural gas »low propane x high propane 10 100 Mobility Diameter (nm) 1000 15°ATDC x * • natural gas • low propane x high propane X 10 100 Mobility Diameter (nm) 1000 10 100 Mobility Diameter (nm) 1000 Figure 6.12: Particle size distributions for natural gas and two levels of propane addition at four combustion timings (clockwise from top left: 0, 5, 10, 15°ATDC) 1 "jE 6000 a> c | 3000 to O •g s 0 — • — N G low ethane high ethane / / / / \ f- —) { \ i \ \ \ \ \ 5 10 50%IHR (°CA) 15 20 5 10 50%IHR (°CA) Figure 6.13: Aethalometer results, BC concentration as a function of timing for ethane (left) and propane (right) additives, compared to natural gas fuelling. -131-1 0 0 % N G --a- lowethane - - X - - high ethane 1 0 0 % NG • lowpropane - - X - - high propane 5 10 50%IHR (°CA) 15 20 15 20 5 10 50%IHR (°CA) Figure 6 .14: B l a c k - c a r b o n fract ion ( B C / P M total mass) as a funct ion o f t i m i n g for ethane (left) and propane (right) addit ives, compared to natural gas fue l l ing . _ 6000 | 5000 „ i 4000 E "5) 3000 3, jjj 2000 J2 1000 o > 0 -5 - N G •• a•• low ethane high ethane > P T \ \ \ y.k "' 1 ^ 10 15 20 _ 6000 | 5000 „I 4000 E o> 3000 3, jjj 2000 ™ 1000 o > -5 - N G • -o •• low propane - - x - h i g h propane T 5 10 50%IHR (°CA) 15 20 50%IHR (°CA) Figure 6 .15 : V o l a t i l e mass rate ( P M total mass - B C mass) as a funct ion o f t i m i n g for ethane (left) and propane (right) addit ives, compared to natural gas fue l l ing . - 1 3 2 --sr 3 9> m 2 5 0 % I H R @ 5 ° A T D C —•— hydrogen/methane • natural gas - - o - - ethane —x— propane 3.8 4.2 4.6 5 Hydrogen:Carbon molecular ratio 5.4 50%IHR@ 15°ATDC - • - hydrogen/methane • natural gas -o-- ethane -x— propane ^ - — B natural gas methane •••o-- ethane propane 3.85 3.9 3.95 4 Hydrogen:Carbon molecular ratio 4.05 3.8 5.4 3.85 3.9 3.95 4 Hydrogen:Carbon molecular ratio 4.05 4.2 4.6 5 Hydrogen:Carbon molecular ratio Figure 6 .16 : E f fect o f g lobal hydrogen to carbon ratio on total change i n P M emiss ions ( P M measured / P M measured under natural gas fuel l ing) at 5 and 15°CA A T D C combust ion t i m i n g . The r ight -hand plots exc lude the hydrogen results. - 1 3 3 -5 0 % I H R @ 5 ° A T D C — • - hydrogen/methane m natural gas - - o - - ethane —x— propane i < 1 1 — "— i~• 3.8 4.2 4.6 5 Hydrogen:Carbon molecular ratio 5.4 hydrogen/methane natural gas o - - ethane propane 4.2 4.6 5 Hydrogen:Carbon molecular ratio CD n <= c 7, £ 5 0 % I H R @ 5 ° A T D C T m natural gas —•— methane •••<>•• ethane —x— propane 3.85 3.9 3.95 Hydrogen:Carbon molecular ratio £ 3 CQ ro 50 %IHR g 15°ATC C 0 natural gas • methane - - o - - ethane x propane < >--... , 3.85 3.9 3.95 HydrogervCarbon molecular ratio Figure 6.17: Effect of global hydrogen to carbon ratio on change in black carbon mass emissions (measured black carbon / black carbon with natural gas fuelling) at 5 and 15°CA A T D C combustion timing. The right-hand plots exclude the hydrogen results. -134-Chapter 7 Fuel Dilution with Nitrogen 7.1 Introduction Var ia t ions i n the compos i t ion o f natural gas due to suppl ier , seasonal or geographical differences pose a s igni f icant chal lenge for engine developers. A s d iscussed i n Chapter 6, re lat ively l o w concentrations o f heavy hydrocarbons can have a s igni f icant impact o n the c o m b u s t i o n process and emiss ions f r o m a p i lo t - ign i ted , d i rect - in ject ion o f natural gas ( P I D I N G ) engine. D i luents i n the fue l , such as N 2 and C 0 2 , m a y also have a s igni f icant impact o n the combust ion process, a result o f both reduced energy density as w e l l as direct effects on the pol lutant format ion k inet ics . B y studying the effects o f vary ing the quantity o f N 2 i n the fue l , it should be possib le to ident i fy the sensit iv i ty o f the c o m b u s t i o n system to fue l d i lu t ion . The p r inc ipa l inf luences o f this d i lu t ion process are expected to be on the turbulent m i x i n g o f the gaseous jet, as w e l l as on the turbulent k inet ic energy o f the post -in ject ion combust ion chamber gases. 7.2 Previous Work The addi t ion o f di luents, i n the f o r m o f E G R , to the ox id i ze r substantial ly reduces diesel and natural gas d i rect - in ject ion engine N O x emiss ions by lower ing the combust ion temperature [122,123] . In a n o n - p r e m i x e d combust ion event, the react ion rate and temperature depend o n the concentrat ion o f the fue l , the ox id i zer , and any inert species. Whether these inert species are or ig inat ing f r o m the fuel side or the ox id i ze r side is not an important factor. W h i l e the addi t ion o f exhaust gases to the fuel cou ld raise certain safety and re l iab i l i ty issues, some di luents, i n part icular molecu lar nitrogen and carbon d iox ide , are present i n foss i l natural gas at up to 4 % by v o l u m e (see A p p e n d i x 5). U n c o n v e n t i o n a l gases, i n c l u d i n g coal -based synthetic gas and landf i l l b iogas, also contain substantial quantit ies o f these .species [224]. W h i l e their presence w o u l d pose numerous systems issues (larger quantity fue l storage devices, lower power density for the fue l , h igher compress ion w o r k requirements, etc.) they may also have a direct inf luence on the combust ion process. The effect o f d i lu t ing a gaseous fuel w i t h nitrogen has been invest igated i n var ious contexts. C rookes et al. [225] found that for a p remixed charge engine, the results are - 1 3 5 -essential ly ident ical to increasing the E G R fract ion, w h i c h w o u l d be expected as the effect is essential ly the same as d isp lac ing oxygen f r o m the charge. B y adding up to 1 4 % N 2 in the fue l , N e l l e n and B o u l o u c h o s [60] found that w h i l e k n o c k resistance is i m p r o v e d ef f ic iency is impai red at constant fue l -a i r stoichiometry . These results are equivalent to those o f tests w i t h 'synthetic E G R ' where nitrogen is added to the intake air [123]. T h e on ly case where nit rogen addi t ion to the fuel in a p remixed combust ion system cou ld s ign i f icant ly impact performance w o u l d be in cases w i t h s ignif icant non-homogenei ty i n the fue l -a i r mix ture . The use o f nitrogen as a di luent in fundamental non -p remixed combust ion studies is re lat ively c o m m o n , p r imar i l y as a technique to reduce fuel concentrations. The effects o f this ni t rogen d i lu t ion have been quant i f ied for a non -p remixed opposed f l o w d i f f u s i o n f lame [226]. T y p i c a l l y , no s igni f icant effects are observed unt i l the fuel stream contains > 8 0 % N 2 (by v o l u m e ) ; above this leve l the temperature required,for ign i t ion to occur increases. These effects m a y be attributed to the increased heat capacity o f the fue l , and are general ly s imi la r to the inf luences o f increased energy d iss ipat ion through higher turbulent strain rates. N 2 concentrat ions b e l o w 8 0 % have no s igni f icant effects. In a c o - f l o w laminar f l a m e , G u l d e r et al. report that the soot vo lume fract ion i n a f lame is reduced proport ional ly w i t h the reduct ion i n methane concentrat ion [227]. The authors attribute this direct ly to fue l d i l u t i o n ; no s igni f icant effect o f the nitrogen o n soot fo rmat ion or ox idat ion k inet ics is proposed. These results indicate that the pr inc ipa l inf luence o f nitrogen addi t ion manifests i tse l f by reduc ing the energy density o f the fue l . There is no evidence o f direct effects on the react ion k inet ics , even at very h igh N 2 concentrations. A s a result, it is reasonable to expect that nitrogen addi t ion to a turbulent non -p remixed gaseous jet w i l l not have a s igni f icant direct impact on the c h e m i c a l k inet ics . One o f the pr inc ipa l effects o f nitrogen addi t ion to the gaseous jet is reduc ing the energy density o f the injected gas, resul t ing i n a longer in ject ion durat ion to p rov ide the same amount o f avai lable c h e m i c a l energy. Increasing the total mass injected also s ign i f icant ly increases the total k inet ic energy transfer to the combust ion chamber gases. H o w e v e r , the author was unable to ident i fy any studies reported in the literature that attempt to evaluate the inf luence o f ni t rogen content in a turbulent non -p remixed jet combust ion system. - 1 3 6 -7.3 Experimental Information The heavy -duty single cy l inder research engine descr ibed in Chapter 3 was used to investigate the effects o f nitrogen d i lu t ion on a gaseous- fuel led direct in ject ion engine. F o r the w o r k descr ibed here, the engine was operated us ing the more modern (17:1 compress ion ratio) p is ton . The di luted gaseous fuel was p remixed and suppl ied through the standard fuel supply system. G a s b lend preparation was carr ied out in large stationary storage vessels, w i t h concentrations in i t ia l l y set us ing part ia l pressures. The fuel b lend compos i t ions tested were 2 0 % and 4 0 % by v o l u m e nitrogen w i t h the remainder be ing c o m m e r c i a l natural gas. The actual fue l compos i t ions , as measured by gas chromatographic analysis o f samples taken f r o m just upstream o f the engine, are shown i n Table 7 .1 . These levels were selected to attempt to ensure that s ignif icant variat ions i n combust ion and emiss ions c o u l d be discerned. T h e 40%) leve l was the upper l i m i t to w h i c h the fuel tanks c o u l d be f i l l e d w h i l e st i l l p r o v i d i n g suff ic ient fue l to conduct the desired number o f test condi t ions. 7.3.1 Operating Condition Th is testing was carr ied out at a h i g h - l o a d , moderate speed operat ing cond i t ion , equivalent to the test point used i n Chapter 6, and s imi la r to the h i g h - l o a d point i n Chapter 5 and the h i g h - l o a d m i d - s p e e d point in Chapter 4. The test cond i t ion is out l ined in Table 7 .2 ; the on ly var iable was combust ion t i m i n g , w h i c h was contro l led (based o n the midpo in t o f the heat release, 50%>IHR, as discussed i n Chapter 3) f r o m 0° to 15° after top-dead-center ( A T D C ) . A l l other parameters, i nc lud ing power ( G I M E P ) , oxygen-based equivalence ratio (()>), speed, and intake oxygen mass f ract ion (Yi n t02) were f i xed to m i n i m i z e operating cond i t ion var iabi l i ty . A s the power and in ject ion pressure were h e l d constant w h i l e the energy density o f the fuel was reduced, the in ject ion durat ion had to be increased to prov ide the same amount o f chemica l energy. A s the t i m i n g was set to prov ide a constant m i d - p o i n t o f combust ion , for the longer in ject ion process w i th fuel d i lu t ion the start o f in ject ion had to be advanced, leading to an earlier s tar t -of -combust ion. T h e addi t ion o f nitrogen to the fuel results in increases i n the concentrat ion o f ni t rogen i n the exhaust and o f the total mass o f exhaust. Th i s results i n lower concentrations o f the other species ( inc lud ing oxygen) : for the h igh nitrogen concentrat ion tests, this was reduced by - 3 % (relative to pure natural gas). Therefore, to achieve the same Yj n t o2, the E G R - 1 3 7 -mass flow-rate was slightly reduced (by -2%), leading to reductions in the inducted mass of charge and EGR fraction (both <1% relative to the natural gas case). As these changes were substantially smaller than the variability in the experimental conditions, they did not have a significant impact on the operating condition set-point or on the measured results. To improve the experimental accuracy, all the data points were replicated three times. Full randomization of the test plan was not possible, as only one gas blend could be prepared at a time. However, the sequence of the test points was fully randomized within each fuel composition. In view of the restrictions on randomization necessitated in the testing, the experimental design represents a blocked design, with the various fuel compositions representing the blocks and the other controlled variable being the combustion timing. As the only replications were within the blocks (rather than between blocks), it is not possible to identify any interaction term between the fuel composition variable and the combustion timing. Furthermore, as a result of this lack of randomization statistically valid tests between the different fuel blends are not possible, as the influence of fuel composition cannot be isolated from day-to-day random fluctuations in experimental conditions. Although these day-to-day fluctuations are small, they nonetheless restrict the statistical analyses. 7.4 Effects of Fuel Dilution Previous studies relating to fuel dilution with nitrogen indicate that the principal influence is on physical mixing processes. In non-premixed gaseous jet combustion, these effects are expected to include increased in-cylinder turbulence (induced by the higher mass of gaseous mixture injected) as well as changes in the density gradient and mixing processes between the fuel jet and the charge. Changes in the fuel composition shift the stoichiometric combustion point towards a richer mixture (on a total mass of fuel including nitrogen / mass of charge basis). The lower rate of energy transfer into the combustion chamber also imposes greater restrictions on the combustion rate. These processes are expected to have a significant impact on the combustion process and pollutant emissions. 7.4.1 Combustion Effects The addition of nitrogen to the fuel while holding the engine power constant results in a longer injection duration, lower energy-based injection rate, and a higher transfer of kinetic energy into the combustion chamber. To maintain the combustion timing (50%IHR) -138-constant, the in ject ion t i m i n g must be advanced as nitrogen is added to the fuel (F igure 7.1). Contrary to ox id i ze r d i lu t ion , w h i c h tends to increase ign i t ion delay [42,130] , fuel d i lu t ion s igni f icant ly reduces the gaseous ign i t ion delay t ime (Figure 7.2). The p i lot delay (also F igure 7.2) increases at the earliest combust ion t imings w i t h 4 0 % N 2 ; this is a result o f the earl ier in ject ion , where the p i lot enters a lower temperature and pressure envi ronment . The fact that the 2 0 % N 2 had a shorter gaseous fuel ign i t ion delay than the 4 0 % case is unexpected. These differences may be a result o f compet ing inf luences between increased turbulent m i x i n g o f the ear ly - in jected fue l , higher shear rates between the fuel and the ox id i ze r , and lower energy content o f the p remixed fuel c loud . A s w e l l , the greater penetration o f the gas jet (due to its higher density) may result i n the ignitable mix ture be ing i n the v i c in i t y o f the p i lot f lame sooner, result ing in faster ign i t ion . These results indicate that, at this cond i t ion , the in ject ion process, and the subsequent m i x i n g o f the gaseous fue l , is exert ing a substantial inf luence o n the ign i t ion o f the gaseous fue l . Th i s agrees w i t h the observations f r o m Chapters 5 and 6 that m i x i n g is a key constraint on the ign i t ion process. T h e shorter ign i t ion delay and reduced energy content o f the d i lu ted natural gas result i n less c h e m i c a l energy be ing contained in the p remixed fuel c loud pr ior to ign i t ion , result ing i n a correspondingly lower peak heat-release rate (Figure 7.3). The energy release rate was also reduced, due to the lower (energy-basis) in ject ion rate. Despi te this , the overa l l combust ion durat ion, as represented by the 1 0 - 9 0 % I H R (also shown i n F igure 7.3) is not increased by ni t rogen addi t ion, and at some condit ions is actual ly s l ight ly reduced. T h i s is especia l ly not iceable i n the later combust ion t i m i n g cases. Th i s result m a y be attributed direct ly to the increased turbulent k inet ic energy imparted f r o m the gaseous jet, resul t ing i n i m p r o v e d la te -combust ion phase m i x i n g and more rapid comple t ion o f the combust ion event i n the post - in ject ion per iod . Further evidence o f this more complete late-stage burnup is p rov ided f r o m the i n -cy l inder pressure and heat-release traces for early and late combust ion t imings (F igure 7.4). These plots c lear ly show the reduced peak heat-release rate, lower premixed-phase combust ion , and longer mix ing - con t ro l l ed combust ion w i t h the 4 0 % N 2 d i lu t ion . Pressure traces and heat-release rates for the other t imings are prov ided i n A p p e n d i x 6. D i l u t i n g the fuel w i t h nitrogen also improves the combust ion stabi l i ty , as s h o w n i n F igure 7.5. A t the h igh nitrogen d i lu t ion level retarding the combust ion enhances the - 1 3 9 -combust ion stabi l i ty ; this is contrary to most prev ious results where retarding the combust ion tended to degrade stabil ity. Th is improvement i n stabil ity is observed for both the overa l l combust ion stabi l i ty ( C O V o f the G I M E P ) and the variat ions i n the early combust ion event ( C O V o f the peak cy l inder pressure). Improved late-cycle m i x i n g due to the higher turbulence levels i n the cy l inder may exp la in these observed improvements i n combust ion stabil i ty . The higher turbulence results in more complete ox idat ion o f the combust ion by -products and more complete ut i l i zat ion o f the hydrocarbon fuels. The i m p r o v e d combust ion stabi l i ty m a y also contribute to the reduced fuel consumpt ion w i t h ni t rogen d i lu t ion , also s h o w n i n F igure 7.5. These results demonstrate the substantial combust ion benefits o f reduc ing the amount o f p remixed fuel pr ior to ign i t ion and o f increasing the late -cyc le i n -cy l inder turbulence. 7.4.2 Gaseous Emissions D i l u t i o n o f the fuel w i t h nitrogen s igni f icant ly inf luences the emiss ions as w e l l as the combust ion event. A s indicated i n F igure 7 .6 , the emiss ions o f a l l the measured gaseous emiss ions species ( H C , C O , and N O x ) are s igni f icant ly reduced. The hydrocarbons, w h i c h are more than 9 5 % CH4, are consistently reduced by approx imate ly 0 .35 g / G i k W h r ( - 3 0 0 ug/cycle) at a l l t imings for the h igh N2 case. The reduct ion for the l o w ni t rogen case is approx imate ly h a l f that o f the h i g h nitrogen case ( - 0 . 1 6 g/GikWhr ) . These emiss ions o f unburned fuel are on ly approximately 0 . 2 % o f the total mass o f CH4 in jected per cyc le ( - 1 3 0 mg/cycle). H o w e v e r , the fact that the magnitude o f the reduct ion is constant at a l l t im ings , and that the reduct ion is approx imately 4 0 % at the lowest emiss ions leve l (earliest t iming) , suggests that there is a constant vo lume o f fuel that is not be ing consumed. S ince the magnitude o f the reduct ion is constant, it appears that there is more than one source o f unburned fuel emiss ions ; at the earlier t imings , a vo lume-based effect is the p r inc ipa l source. One potent ial contr ibutor to these emissions w o u l d be the injector, where gas i n the nozz le holes and sac v o l u m e gradual ly vents into the combust ion chamber as the cy l inder pressure fa l ls after the end o f the in ject ion process. Fue l that enters the combust ion chamber i n this manner is u n l i k e l y either to m i x to a combust ib le mixture or to ignite, thereby p r o v i d i n g a m i n i m u m v o l u m e o f unburned fuel w h i c h w i l l be emitted w i t h the exhaust. A t the later t imings , the fact that the baseline emiss ions are substantial ly h igher suggests that other mechan isms , such as increased bu lk quenching or loca l turbulent strain ext inct ion , m a y also - 1 4 0 -be contr ibut ing to unburned fuel emiss ions. G i v e n that these mechanisms do not appear to be reduced by the addi t ion o f nitrogen to the fue l , it w o u l d appear that contr ibutory mechanisms such as combust ion var iabi l i ty (wh ich was reduced) or late-cycle combust ion rate (wh ich was increased) are not the pr inc ipa l sources. The lower intensity, longer durat ion combust ion as w e l l as the lower adiabatic f lame temperature have s igni f icant impacts o n the N O x emiss ions (F igure 7.6). The N O x emiss ions are no different between the natural gas and the l o w d i lu t ion leve l ; a s igni f icant reduct ion is not iceable , however , between the l o w and h igh d i lu t ion cases. The change i n adiabatic f lame temperature (Table 7.1) is on ly 3 0 K for the l o w nitrogen d i lu t ion case. D u e to the thermal m e c h a n i s m ' s h i g h temperature dependence, this w o u l d be expected to reduce N O x emiss ions by approx imate ly 1 5 % based on the f lame temperature correlations deve loped i n prev ious w o r k [39,122] . F o r the h igh d i lu t ion case, the reduct ion i n N O x is expected to be o n the order o f 4 0 % ; on ly at the earliest t i m i n g is the observed reduct ion this s ignif icant . A t later t imings , reduct ions in N O x emiss ions are only - 2 0 % . The fact that the reductions i n N O x emiss ions are not as s igni f icant as the changes i n the adiabatic f lame temperature w o u l d predict suggests that the nitrogen i n the fuel may be in f luenc ing the p r inc ipa l N O format ion mechanisms (section 2.2.2) through other effects. F o r the dominant thermal m e c h a n i s m , the format ion o f N radicals is the l i m i t i n g step i n the react ion; hence, a higher N 2 concentrat ion, especial ly on the fue l - r i ch side o f the combust ion zone, c o u l d increase N O format ion . H o w e v e r , as the thermal m e c h a n i s m typ ica l l y occurs i n the pos t - combust ion gases, where N 2 f r o m the ox id i zer is abundant, the addi t ion o f N 2 to the fuel increases the N 2 concentrat ion by less than 1 0 % . A s w e l l , the residence t ime o f the burned gases at h igh temperature is reduced by the more rapid turbulent m i x i n g in the c o m b u s t i o n chamber result ing f r o m the higher jet k inet ic energy. It is also poss ib le that the contr ibut ions o f other N O x fo rmat ion mechanisms are affected by ni t rogen add i t ion ; for example , the prompt mechan ism m a y be b e c o m i n g more s igni f icant as the combust ion temperature is reduced. A s for the other gaseous emiss ions , the C O emissions are substantial ly reduced w i t h ni t rogen addi t ion (Figure 7.6). The fact that the C O emissions are independent o f combust ion t i m i n g w i t h d i luted fuel (as opposed to the increase at intermediate t imings for the undi luted natural gas) suggests that the C O increase at m i d - t i m i n g s may not be due to jet interaction - 1 4 1 -w i t h the p is ton . The higher density and longer in ject ion duration o f the d i luted gaseous jet substantial ly increases the jet penetration distance, w h i c h w o u l d be expected to exacerbate any piston- jet interact ion effects. It is l i ke l y the higher turbulent k inet ic energy s igni f icant ly increases the late -combust ion burn -up o f the C O , result ing i n a s igni f icant reduct ion i n emiss ions o f C O as w e l l as other combust ion by -products . 7.4.3 Particulate Emissions The enhanced ox idat ion o f the combust ion by -products by increased turbulence late i n the c o m b u s t i o n stroke may also have a substantial inf luence on the P M mass emiss ions (F igure 7.6). P M emiss ions are substantial ly reduced at the intermediate t imings w i t h fuel d i lu t ion . The higher turbulence intensity imparted by the higher density fuel m a y result i n s igni f icant increases i n the late-stage combust ion event and burn -up , result ing i n a substantial reduct ion i n P M emiss ions at most t imings . That the ta i l -end combust ion is more complete is suggested i n F igure 7.4; furthermore, F igure 7.7 shows that the heat released i n the late combust ion stages (post injection) is substantial ly higher for the nit rogen cases than for the pure natural gas. Interestingly, at the earliest t i m i n g P M emiss ions are increased w i t h ni t rogen d i l u t i o n ; the reasons for this effect are unclear. Th is suggests that the nit rogen addi t ion is a lso in f luenc ing other P M format ion and ox idat ion mechan isms . These mechan isms m a y inc lude compl icated chemica l and phys ica l effects re lat ing to the increased concentrat ion o f nitrogen i n the fue l . For example , the presence o f a relat ively inert species (N2) i n the fuel c o u l d enhance any third body reactions; changes i n the loca l gradients i n the react ion zone w o u l d strongly affect both energy and species d ist r ibut ion i n the f lame zone; or even direct part ic ipat ion o f dissociated nitrogen molecules i n the reactions may be important. Fundamenta l studies conducted under condit ions directly equivalent to those in a direct -in ject ion engine are required to further elucidate these effects. In general , however , these results indicate the P M mass emiss ions are fo rmed f r o m mul t ip le compl icated mechanisms w h i c h respond differently to changes i n fuel compos i t ion . The total mass emiss ions (Figure 7.6) indicate that the P M mass at the latest t i m i n g cond i t ion is essential ly the same for a l l three fuel composi t ions . H o w e v e r , this value is s imi la r to the m i n i m u m quanti f iable P M mass ident i f ied i n earl ier w o r k (Chapter 5). The ul t raf ine number distr ibut ions and integrated m o b i l i t y v o l u m e , s h o w n i n F igures 7.8 and 7.9, suggest that both the number and the m o b i l i t y v o l u m e o f the part icles have been reduced by - 1 4 2 -as m u c h as 8 0 % at the latest t i m i n g through fuel d i lu t ion . A t the earl ier t imings , the integrated number distr ibut ions showed trends equivalent to those o f the measured masses. Th i s di f ference between the P M mass measurements and the m o b i l i t y v o l u m e at the lowest mass concentrations is most l i ke l y due to volat i le condensat ion, as d iscussed i n Chapters 5 & 6. The size distr ibut ions also indicate that a nuc leat ion -mode peak occurs at the later t imings w i t h both l o w and h igh N2 levels . These nucleat ion mode peaks are typ ica l l y attributed to vo lat i le nucleat ion processes in the absence o f suff ic ient so l id P M surface area for condensat ion to occur (as discussed in section 5.3.2). Interestingly, at the latest t i m i n g , w i t h h igh nit rogen d i lu t ion , the nucleat ion mode peak is not apparent. The lack o f so l id part icles i n the pos t - combust ion gases may result i n less volat i le nucleat ion, result ing i n higher vapour -phase concentrations. These concentrations may be suff ic ient to induce substantial ly earl ier se l f -nuc leat ion o f the volat i le species (for example upstream o f the surge tank). W i t h such earl ier nuc leat ion , more part icle growth by accumulat ion and condensat ion m a y occur . A s the total quantity o f volat i les should not have been substantial ly affected by the nit rogen d i lu t ion o f the f u e l , the earl ier nucleat ion and subsequent growth m a y result in the lower number o f larger part icles observed in the size distr ibutions at the latest combust ion t i m i n g . The mass o f b lack carbon i n the exhaust stream is substantial ly reduced w i t h nitrogen d i lu t ion , especia l ly at the latest combust ion t imings , as shown in F igure 7.10. A t the h igh nit rogen d i lu t ion case, the amount o f b lack carbon i n the sample is not detectable. Th i s suggests that the observed ultrafine P M (Figure 7.8) is p r imar i l y nucleated vo lat i le species. Hence , either the format ion o f carbonaceous P M is reduced, or ox idat ion is enhanced. It is l i k e l y that both effects are relevant; h igher turbulence results i n enhanced pos t - combust ion m i x i n g , poss ib ly enhancing ox idat ion . S imul taneous ly , the lower temperatures and reduced carbon - to -oxygen ratio i n the gas jet upstream o f the combust ion zone m a y result i n reduced carbonaceous particulate format ion . 7.5 Discussion D i l u t i n g the fuel w i t h an inert species, nitrogen, is the on ly technique invest igated i n this thesis w h i c h s imultaneously reduces al l measured emissions wh i le retaining the P I D I N G engine's e f f ic iency and performance. A l t h o u g h the use o f h igh quantit ies o f ni t rogen di luent m a y pose technica l and economic chal lenges, part icular ly for veh icu lar appl icat ions , these results do prov ide an ind icat ion that nitrogen d i lu t ion cou ld be used to offset increases in - 1 4 3 -emiss ions result ing f r o m higher concentrations o f heavier hydrocarbons i n different geographic locat ions. 7.5.I Fuel Systems Issues One o f the m a i n drawbacks o f d i lu t ing the gaseous fuel w i t h ni t rogen is the need to compress an increased quantity o f fue l . A s s u m i n g adiabatic compress ion w i t h constant spec i f ic heats and ideal gas behaviour , the w o r k required to compress an equal quantity o f fuel on an energy basis can be roughly estimated. The results o f this ca lcu la t ion , s h o w n i n F igure 7 . 1 1 , indicate that more than twice the p u m p i n g w o r k is required to compress the fuel w i t h 4 0 % nitrogen. The w o r k required to compress 2 0 % nitrogen is on ly 3 0 % greater than the w o r k for pure methane. Th i s work is , however , s t i l l a relat ively smal l f ract ion o f the total energy content o f the fuel . F o r compar ison purposes, the w o r k (per unit energy stored) required to compress 1 5 % and 3 5 % (by vo lume) blends o f hydrogen are also s h o w n ; substantial ly less compress ion w o r k is required for the hydrogen addit ive than for an equivalent amount (by vo lume) o f nitrogen. An oth e r s igni f icant drawback to nitrogen (or hydrogen) b lend ing i n the fuel is the reduct ion i n vo lumet r i c energy density. The effect o f 4 0 % nit rogen and 3 5 % hydrogen b lend ing o n the gas v o l u m e required to store an equivalent amount o f c h e m i c a l energy (again assuming ideal gas behaviour) is also shown i n F igure 7 .11 . A t a pressure o f 35 M P a (roughly 5000 psi ) , the vo lume required for the hydrogen is 5 0 % greater than for pure methane and the v o l u m e required for nitrogen is almost twice the methane v o l u m e . Whether the improvements i n engine performance and emiss ions warrant such substantial increases i n paras i t i c load and fuel storage v o l u m e depends on the appl icat ion. C ryogen ic storage might be an opt ion , since it provides improved energy density and reduced parasit ic compressor w o r k for convent ional natural gas. Potent ial var iat ions i n fuel c o m p o s i t i o n dur ing extract ion cou ld pose a chal lenge for mainta in ing the desired fuel b lend. Separate cryogenic storage o f the addit ive might be required; this w o u l d impose a substantial increase i n system complex i t y and cost. M o r e research is required to further understand the cryogenic storage o f blends o f natural gas w i t h h igh concentrations o f nitrogen. 7.5.2 Fuel Composition Parameters A number o f different parameters have been prev ious ly suggested to prov ide a general o v e r v i e w o f the effects o f natural gas compos i t ion on engine performance. The most - 1 4 4 -w i d e l y used o f these inc lude the W o b b e Index (discussed i n sect ion 6.2) [221], the molar hydrogen- to -carbon ( H : C ) ratio and the higher heating value ( H H V ) . The calculated values for these parameters for a l l the fuel composi t ions tested i n this thesis are s h o w n i n Tab le 7 .3 . In part icular , prev ious w o r k suggests that for a constant W o b b e Index, combust ion performance and emiss ions are not s igni f icant ly affected by fuel compos i t i on [213]. The W o b b e Index, w h i c h has the same units as the mass -spec i f i c heating value ( typical ly MJ/kg) is def ined as: f specific gravity where the spec i f ic gravity is the density o f the fuel b lend relative to air at standard temperature and pressure. A compar ison o f the combust ion durat ion and the peak heat-release rate as funct ions o f the W o b b e Index are shown in F igure 7.12. The compar ison inc ludes the nit rogen blends tested i n the current chapter, as w e l l as the hydrogen b lend results (Chapter 5) and the ethane and propane addi t ion tests (Chapter 6). A s the results indicate, the W o b b e Index does not prov ide a reasonable representation o f the combust ion parameters for the composi t ions tested. Th is is most l i ke l y due to the fact that it was establ ished for premixed-charge combust ion , where the speci f ic gravity o f the fuel has a more s igni f icant effect than in a non -p remixed combust ion event. A n invest igat ion o f the effect o f fuel compos i t ion on the peak heat-release rate for P I D I N G combust ion is in format ive , as this parameter is representative o f the amount o f energy released dur ing the in i t ia l p r e m i x e d combust ion phase. F igure 7.13 demonstrates that the fuel c o m p o s i t i o n has a s ignif icant inf luence on the peak heat-release rate (relative to the peak heat-release at the same t i m i n g cond i t ion us ing natural gas). O n l y the pure methane case indicates a higher peak heat-release rate than the natural gas, w h i l e the most s igni f icant reductions are observed for the h igh nitrogen d i lu t ion cases. A l l four o f the fuel blends that inc lude heavier hydrocarbons have v i r tual ly ident ical results. The sensit iv i ty o f peak H R R to fue l c o m p o s i t i o n comes f r o m two sources: the ign i t ion delay t ime and the energy density o f the fue l . E i ther shorter ign i t ion delays or lower energy-density fuels result i n less c h e m i c a l energy be ing avai lable to be released in the in i t ia l par t ia l l y -p remixed combust ion event. The inf luence o f gas ign i t ion delay (G ID) on the peak H R R for a l l the fuel blends is s h o w n in F igure 7.14. The outl ier points are for the h igh nitrogen case, where the c h e m i c a l energy - 1 4 5 -avai lable for the p remixed combust ion is reduced by d i lu t ion o f the fue l . The effects o f the shorter gas ign i t ion delay t ime also contribute to the observed reduct ion i n peak H R R at this condi t ion . The fact that w i t h a shorter G I D , less fuel is avai lable to burn after ign i t ion occurs results i n less o f the fuel hav ing m i x e d beyond a combust ib le stoichiometry , thereby reducing H C emiss ions . The s igni f icance o f this inf luence on H C emiss ions is s h o w n i n F igure 7.14, a l though it is evident f r o m the var iabi l i ty i n the results that factors other than the G I D are also contr ibut ing to H C emiss ions. N o r do these parameters prov ide any further insight into the changes i n emiss ions o f other species. Another attribute w h i c h has a direct impact on emiss ions is the combust ion var iab i l i t y , represented by the C O V o f the G I M E P . F igure 7.15 demonstrates that processes w h i c h reduce the combust ion var iabi l i ty w i l l also reduce emiss ions o f both C O and H C . A s w i t h the G I D , the var iabi l i ty i n the data indicates that other mechanisms are also important. F r o m the results observed in the current chapter, as w e l l as those f r o m Chapters 5 and 6, fue l compos i t i on clear ly has a s ignif icant impact on emiss ions o f a l l combust ion by -products. T h i s result dif fers s igni f icant ly f r o m premixed charge combust ion f ind ings , w h i c h general ly suggest that fuel compos i t ion has relat ively smal l inf luences o n emiss ions [217,228] . In the current w o r k , the least substantial effects are, i n general , found for the N O x emiss ions . Th i s m a y be attributed to relat ively s m a l l changes i n combust ion temperature for most o f the fuel blends. The only two fuel blends that show signi f icant differences i n N O x emiss ions (high hydrogen and h igh nitrogen) are the two where the changes i n f lame temperature are most s ignif icant. The higher f lame temperatures w i t h hydrogen fue l l i ng increase N O x emiss ions (section 5.3.1) w h i l e the lower temperatures associated w i t h nitrogen d i lu t ion reduce N O x (section 7.4.2). Var iat ions i n other aspects o f the N O format ion mechanisms are also important; however , the results indicate that the p r inc ipa l effect is the combust ion temperature, as suggested i n prev ious w o r k [39]. The effects o f fuel compos i t ion on emiss ions o f C O , H C , and P M are m u c h more s igni f icant . It is expected that many phys ica l and chemica l effects are contr ibut ing to these var iat ions. H o w e v e r , some o f the p r inc ipa l inf luences m a y be encompassed i n the vo lumetr i c energy content o f the fuel (as represented by the H H V o n a molar basis) and the carbon content o f the fuel (represented by the H : C ratio). T o prov ide further insight into the roles o f - 1 4 6 -these parameters o n the pol lutant fo rmat ion mechanisms, a new index ( ^ ) may be def ined as: cf = ^ L (7.2) H:C w i t h units o f M J / k m o l . The effects o f this index o n C O and H C emiss ions (relative to the basel ine natural gas fue l l ing condit ion) over the f u l l range o f fuel compos i t ions and t imings tested are s h o w n in F igure 7.16. A general increase in C O and H C emiss ions w i t h increasing £ (i.e. higher energy content and/or lower H : C ratio) is observed. The C O emiss ions are sensit ive to combust ion t i m i n g , ind icat ing that other parameters are also h a v i n g a s igni f icant impact . A potential cause for this is that variat ions in the in -cy l inder pressure and temperature substantial ly inf luence C O emiss ions ; a s imi lar inf luence is apparent i n the in ject ion pressure studies discussed i n Chapter 4. Hydrocarbon emiss ions ( w h i c h i n a l l cases were > 9 5 % C H 4 ) are less sensitive to the combust ion t im ing . Th i s also agrees w i t h the results from the in ject ion pressure study (Chapter 4), w h i c h indicate that H C emiss ions levels are less sensit ive to variat ions i n cy l inder pressure and temperature than are C O emiss ions . In general , the H C emiss ions s h o w substantial correlat ion w i t h £ for the hydrogen and nitrogen addi t ion cases; however , for the heavier hydrocarbon cases, the effects are less clear. The greatest inf luence o f H : C ratio and the heating value (as represented by the parameter is observed o n the particulate matter emiss ions , s h o w n i n F igure 7.17. In general , P M emiss ions are substantial ly greater for higher values o f The total mass o f P M , as measured by the T E O M , f o l l o w s this general t rend; however , the substantial contr ibut ion o f vo lat i le species to the P M emiss ions at the h igh ni t rogen- and hydrogen-addi t ion condi t ions results i n s ignif icant P M mass var iabi l i ty at these fue l l ing condi t ions . Th i s is not surpr is ing , as the £ term relates most direct ly to carbonaceous particulate fo rmat ion , as represented by the b lack carbon measurements. The mob i l i t y v o l u m e values , calculated o n the basis o f the integration o f the S M P S results (see section 3.2.2) reveal a s imi la r l y strong dependence on the £ ratio. Th is provides further evidence that the b lack carbon and the ultraf ine part ic les are c lose ly related to each other and are p r imar i l y attributable to so l id carbon (soot), poss ib ly covered w i t h condensed volat i les . In general , the £, ratio provides a useful ind icat ion o f the changes i n P M emiss ions , and especia l ly o f carbonaceous soot part icles, w h i c h may be expected to occur w i t h - 1 4 7 -variat ions in fue l compos i t ion . D u e to the m u c h wider var iat ion o f \% w i t h the hydrogen and nitrogen d i lu t ion studies, the relat ionship between the 1% term and P M emiss ions for the heavier hydrocarbon cases are less clear f rom Figure 7.17. T o prov ide further insight , the same plots are reproduced i n F igure 7.18 for only the heavier hydrocarbon and pure methane c o m p o s i t i o n tests presented i n Chapters 5 and 6. The results indicate that the combust ion t i m i n g has a more s ignif icant effect on the P M emiss ions as a funct ion o f i\ for the heavier hydrocarbons than for the nitrogen and hydrogen test results. H o w e v e r , it is apparent that, yet again , i\ prov ides an ind icat ion o f the relative potential increase i n b lack carbon emiss ions w i t h the addi t ion o f the heavier hydrocarbons, w h i c h s imultaneously increase the H H V o f the fuel and reduce the H : C ratio. Th i s prov ides further conf i rmat ion that the H : C ratio and the H H V are t w o o f the most important parameters relat ing to changes i n P M emiss ions due to differences i n fue l compos i t ion . B y ensuring that the i\ parameter for a g iven fuel is re lat ively consistent, it should be possible to m i n i m i z e the effects o f fuel compos i t i on var iat ions o n emiss ions f r o m a d i rect - in ject ion o f natural gas engine. A further c o m p l i c a t i n g factor to the ca lcu lat ion o f the i\ value pertains to the addi t ion o f carbon conta in ing species such as C O or CO2 to the f u e l ; for these species, it is l i ke l y that the carbon w i l l need to be exc luded f r o m the H : C ratio. H o w e v e r , further research is required to val idate this supposi t ion . 7.6 Conclusions 1. The late -cyc le combust ion process is substantial ly enhanced by nitrogen addi t ion . T h i s may be attributed to higher in -cy l inder turbulence, due to the larger mass o f fuel injected and corresponding increase in gaseous jet k inet ic energy. T h i s more rap id , more complete combust ion results i n lower emissions o f unburned fuel as w e l l as contr ibut ing to reductions in the combust ion var iabi l i ty . 2. Part iculate matter total mass, b lack -carbon content, and ultrafine part ic le emiss ions are substantial ly reduced by nitrogen d i lu t ion . For the h igh -n i t rogen d i lu t ion cond i t ion , b l a c k carbon emiss ions are not detectable at the later combust ion t imings . The measured mass o f P M under h igh fuel d i lu t ion condit ions m a y be attributed to vo lat i les condens ing i n the sample system. 3. D i l u t i o n o f the fuel results i n s ignif icant reductions in emiss ions o f C O and H C (>95% CH4). Reduct ions i n H C are consistent at a l l operating condi t ions and are direct ly - 1 4 8 -proport ional to the reductions i n the concentrat ion o f methane in the fue l . The reduct ion i n C O emiss ions is p r imar i l y attributed to the more rapid late -cyc le combust ion process; this factor may also contribute to the reduced H C emiss ions . 4. N O x emiss ions are not affected w i t h intermediate levels o f nitrogen addi t ion to the fuel but are s igni f icant ly reduced at higher (40%) concentrations. Th is can be attributed to the substantial reduct ion i n combust ion temperature due to the h igh concentrat ions o f di luent i n the react ion zone. 5. The fuel convers ion ef f ic iency is s l ight ly improved through the addi t ion o f ni t rogen to the fue l . Th i s may be a result o f reduced unburned fuel and combust ion by -product emiss ions . 6. D i l u t i n g the fuel w i t h nitrogen reduces the peak heat-release rate, i gn i t ion delay, and cyc le - to - cyc le var iabi l i ty o f the gaseous combust ion event. Despi te the lower heat-release rate and longer in ject ion durat ion, the overa l l combust ion durat ion is s l ight ly reduced. 7. The parameter \% provides a useful measure o f the relative changes i n emiss ions w h i c h may be expected f r o m signif icant changes i n fuel compos i t ion . H o w e v e r , it is somewhat sensit ive to combust ion t i m i n g ; as a result, any speci f ic parameters der ived f r o m this value w i l l be appl icable on ly for a g iven operating cond i t ion . - 1 4 9 -7.7 Tables and Figures Table 7 . 1 : Gaseous fuel composi t ions as measured by gas chromatograph analysis Natura l L o w H i g h Species Gas N i t rogen N i t r o g e n n -Butane 0.05 0.04 0.03 i -Butane 0.07 0.05 0.05 i -Pentane 0.02 0.02 0.01 n-Pentane 0.02 0.01 0.01 Hexane 0.01 0.00 0.00 Heptane 0.00 0.00 0.00 Octane 0.00 0.00 0.00 N 2 1.47 18.89 38.88 Methane 96.6 78.2 59.4 C 0 2 0.44 0.28 0.24 C2H6 1.41 2.31 1.17 C 3 H 8 0.32 0.25 0.22 M W (kg/kmol) 16.70 18.83 21.04 L H V (kJ/kg) 48080 35489 23792 H H V (kJ/kg) 53313 39338 26378 Tadiabatic (K-) 2485 2440 2400 H/C ratio 3.94 3.93 3.94 Tab le 7 .2 : E n g i n e operating constants and variables Parameter V a l u e Speed ( R P M ) 1200 G I M E P (bars) 13.5 0.6 Yint02 0.19 F u e l Pressure ( M P a ) 21 5 0 % I H R (°ATDC) 0, 5, 10, 15 P i l o t (mg/inj) 5 - 1 5 0 -Table 7 .3 : Representative parameters for a l l fuel compos i t ion results F u e l H i g h e r Heat ing Dens i ty H : C Rat io W o b b e V a l u e ( H H V ) ( @ S T P ) Index ( H H V / H : C ) MJ/kmol kg/m3 MJ/kg MJ/kmol Natura l G a s 888.6 0.697 3.96 70.1 226 .4 Pure Methane 888.0 0.677 4.00 74 .6 222 .0 L o w Ethane 926.7 0.726 3.86 68.9 241.7 H i g h Ethane 938.7 0.737 3.83 68.1 2 4 6 . 9 L o w Propane 910.1 0.723 3.88 68 .0 236 .2 H i g h Propane 922.8 0.745 3.83 66.0 242.7 L o w N i t r o g e n 726.6 0.784 3.96 48 .7 189.5 H i g h N i t r o g e n 547.5 0.883 3.96 31.2 141.4 L o w H y d r o g e n 803.7 0.598 4.32 79.6 187.8 H i g h H y d r o g e n 676.0 0.469 5.08 96.4 133.0 - 5 0 5 10 15 20 -5 0 5 10 15 20 5 0 % I H R ( ° C A ) 5 0 % I H R ( ° C A ) Figure 7 . 1 : C o m m a n d e d Injection T i m i n g s (pi lot start -of - in ject ion t i m i n g o n left, gas start-o f - in jec t ion on right). - 5 0 5 10 15 20 -5 0 5 10 15 20 50%IHR (°CA) 5 0 % l H R (°CA) Figure 7 .2 : Igni t ion delay t imes (relative to commanded start -of - in ject ion) for p i lo t (left) and gaseous fuel (right) w i t h natural gas and nitrogen d i luted fuels. - 1 5 1 -250 O 2 0 0 H150 or 100 or « 50 <u CL 0 - N G 2 0 % N 2 — X - - 4 0 % N 2 Figure 7 .3 : 5 10 50%IHR (°CA) - 5 15 20 50%IHR (°CA) Peak heat-release rate and combust ion durat ion ( 1 0 - 9 0 % I H R ) for natural gas and nitrogen di luted cases. top) and late ( 5 0 % I H R at 1 5 ° A T D C , bottom) combust ion t imings . - 1 5 2 -1.5 1.2 LU u a V 0.6 > O " 0.3 0.0 • NG • n - 2 0 % N 2 - - X - 4 0 % N 2 NG • • D • • 2 0 % N 2 - -x - 4 0 % N 2 5 10 15 50%IHR (°CA) 20 -5 0 5 10 15 50%IHR (°CA) 0 5 10 15 20 50%IHR (°CA) Figure 7 .5 : C O V o f G I M E P and peak cy l inder pressure, as w e l l as G I S F C for natural and the nitrogen di luted cases. 5 10 50%IHR (°CA) 3.0 2.5 I 2.0 ej 1.5 O 1 0 o 0.5 0.0 • N G ••-D-- 2 0 % N 2 - - X - 4 0 % N 2 % N - 5 5 10 15 50%IHR (°CA) 20 -5 5 10 15 50%IHR (°CA) Figure 7 .6 : E m i s s i o n s ( C O , N O x , H C , P M ) for natural gas and nitrogen d i lu ted cases. - 1 5 3 -- 5 0 5 10 15 20 50%IHR (°CA) gure 7.7: Percentage o f total heat released after the end o f in ject ion, for natural gas and nitrogen di luted fue l l ing . 1.E+08 1.E+08 1000 Mobility Diameter (nm) Mobility Diameter (nm) 1.E+08 | 1.E+07 I 1.E+06 | 1.E+05 1.E+04 10°ATDC NG 2 0 % N 2 4 0 % N2 1.E+08 1.E+07 I 1.E+06 a. Q 1.E+04 15°ATDC / \ f \ * & "\* \ ^ NG ^ 2 0 % N 2 '+ 4 0 % N 2 10 100 1000 10 100 Mobility Diameter (nm) 1000 Mobility Diameter (nm) gure 7.8: Part ic le number size distr ibutions as measured by the S M P S for pure natural gas and nitrogen di luted fue l l ing . - 1 5 4 -— • — NG - ° - 2 0 % N2 - - X - - 4 0 % N 2 / T V , k ^=^! ^ 15 20 -5 0 5 10 50%IHR (°CA) Figure 7.9: Part ic le m o b i l i t y vo lume (integrated size distr ibution) as a funct ion o f combust ion t i m i n g , for pure natural gas and nitrogen d i luted fue l l ing . _ 100% 5 10 50%IHR (°CA) c o 80% T3 m 60% LL c o . a 40% ro O 20% o ro CO 0% —•—NG •••0-- 20% N 2 -x--40%N2 •p.-- ) ' - ^ 5 10 50%IHR (°CA) 15 20 Figure 7 .10 : B l a c k carbon mass measured by the Aethalometer (left) and b l a c k - c a r b o n f ract ion ( B C mass / total P M mass, right) as a funct ion o f combust ion t i m i n g , for pure natural gas and nitrogen di luted fue l l ing . Pressure (MPa) Pressure (MPa) Figure 7 . 1 1 : Es t imated compress ion w o r k (as a % o f the chemica l energy i n the fuel) and the storage v o l u m e for var ious fuel blends. Ca lcu lated assuming an adiabatic ideal gas compress ion process w i t h constant speci f ic heat ratio. - 1 5 5 -1.3 0 1.2 1 1.1 a> .o « 1 0.9 0.8 • 0 ° A T D C • 5 ° A T D C A 10°ATDC x 15°ATDC 50 100 W o b b e Index (MJ/kg) 150 O 1 2 Z 1 T J § 0.8 | 0.6 | 0.4 ro E 0 0 I X u * : * . 0 ° A T D C • 5° A T D C A 1 0 ° A T D C x 15° A T D C 50 100 W o b b e Index (MJ/kg) 150 Figure 7 .12: C o m b u s t i o n duration ( 1 0 - 9 0 % I H R , left) and peak heat-release rate (right), for a l l gas blends (hydrogen, hydrocarbon, and nitrogen addit ion) relative to natural gas as funct ions o f W o b b e Index. 1.2 1 0 ° A T D C • 5° A T D C • 1 0 ° A T D C • 15° A T D C J> J> J& J> ^ & v ^ ' * ^ v ^ Figure 7 .13 : Peak heat-release rate relative to natural gas for a l l t imings and a l l fuel blends. CD 1.2 1 T 3 £ 0.8 0.6 g 0.4 I x 0.2 . 0 ° A T D C • 5° A T D C A 10° A T D C x 15° A T D C \ High nitrogen dilution case 0.6 1.2 1 0.8 0.6 0.4 0.2 0 P A j - ^ ^ V ^ A / 0 * • 0 ° A T D C • 5 ° A T D C A 1 0 ° A T D C x 15° A T D C 1 / A ' x 1.2 0.6 0.8 1 G ID . fue l blend / N G 1.2 0.8 1 G ID . fue l b l e n d / N G Figure 7.14: Influence o f changes in gas ign i t ion delay t ime (G ID ) on peak heat-release rate and hydrocarbon emissions f r o m var ious fuel blends. L i n e s represent best fit through the data. - 1 5 6 -1.2 o 1 z T3 0.8 c <u 0.6 0.4 o 0.2 0 n j.K ^ — x j > ^ • x A < a • • O " A T D C • 5 ° A T D C A 1 0 ° A T D C x 1 5 ° A T D C 1 • X o Z 1.5 •a c - 1 O 0.5 o • u . x ° • 0 U A T D C • 5° A T D C A 10° A T D C x 15° A T D C ° X 0.5 1 C O V G IMEP, fuel blend / NG 1.5 1.5 0 0.5 1 C O V G IMEP, fuel blend / NG Figure 7 .15: Influence o f combust ion var iabi l i ty ( C O V G I M E P ) o n H C and C O emiss ions for var ious fuel blends. L ines represent best fit through the data. 1.5 "55 d o 0.5 • 0 ° A T D C • 5° A T D C • A 10° A T D C x 15° A T D C * • A • V » A xg A 1.2 z 1 | 0.8 1 0.6 | 0.4 2 0.2 • 0 ° A T D C • 5 ° A T D C A 10° A T D C x 15° A T D C - / * %S X • • A X 100 2 0 0 t; (MJ/kmol) 3 0 0 3 0 0 0 100 2 0 0 t] (MJ/kmol) Figure 7 .16: C o m p a r i s o n o f H C and C O emissions for a l l fuel blends relat ive to natural gas as a funct ion o f i\ ( H H V / H : C ) . L i n e represents best fit through the data. z •o c cu -§ 2 in m co E ° - 0 • 0 ° A T D C • 5 ° A T D C A 10° A T D C x 15° A T D C A z x> 3 c O CO . 0 ° A T D C o 5 ° A T D C " A 10° A T D C x 15° A T D C • • A • . K » — 1 0 0 2 0 0 c; (MJ/kmol) 4 3 0 0 100 2 0 0 t\ (MJ/kmol) 3 0 0 CD Z - 3 •o c a) £ 2 . 0 ° A T D C " o 5 ° A T D C A 1 0 ° A T D C - x 15° A T D C • •a * /i* • J 3 0 0 0 100 2 0 0 \ (MJ/kmol) Figure 7 .17: C o m p a r i s o n o f P M emiss ions (total mass, b lack -ca rbon mass, ul traf ine part icle vo lume) for a l l fuel blends relative to natural gas emiss ions as a funct ion o f i\ ( H H V / H : C ) . L i n e s are best fit (exponential) curves through the data. - 1 5 7 -C3 •o 3 aJ 2 3 «*— in ro 1 E . 0 ° A T D C o 5 ° A T D C A 1 0 ° A T D C x 1 5 ° A T D C * *ft S • ft B 210 a> 2 o m • 0 ° A T D C • 5 ° A T D C A 1 0 ° A T D C x 1 5 ° A T D C xx 220 230 i, (MJ/kmol) 240 250 210 220 230 i, (MJ/kmol) 240 250 O Z 3 T3 C 0) f 2 <u •5 1 • 0 ° A T D C • 5 ° A T D C A 1 0 ° A T D C x 15°ATDC • • 0 • A * x X X 8 M 210 220 230 240 250 5 (MJ/kmol) Figure 7 .18: C o m p a r i s o n o f P M emissions (total mass, b lack carbon mass, m o b i l i t y vo lume) for methane and heavy hydrocarbon addi t ion , relative to natural gas emiss ions as a funct ion o f ^ ( H H V / H : C ratio) - 1 5 8 -Chapter 8 The Effects of Reingested Particles 8.1 Introduction In an effort to meet emiss ions standards for ox ides o f ni t rogen ( N O x ) , exhaust gas rec i rcu lat ion ( E G R ) systems are w i d e l y used in n e w on - road heavy-duty diesel engines. E G R w o r k s by reduc ing the concentrat ion o f oxygen i n the intake charge, thereby l o w e r i n g the combust ion temperature and hence i m p a i r i n g nitrogen ox ide format ion [123]. H o w e v e r , h igh levels o f E G R increase emiss ions o f particulate matter ( P M ) and impai r fue l e f f i c iency [229]. A l t h o u g h very h igh E G R fractions (>50%) reduce P M and N O x emiss ions , this strategy is l im i ted to l o w - l o a d condit ions and suffers f r o m impai red fuel economy and h igh carbon m o n o x i d e ( C O ) and unburned hydrocarbon ( H C ) emissions [230]. A s noted i n prev ious chapters, the use o f p i lo t - ign i ted , direct - in jected natural gas ( P I D I N G ) is another technique to reduce N O x and P M emiss ions f r o m heavy-duty engines [40]. W h i l e very l o w N O x leve ls can be ach ieved by c o m b i n i n g h igh E G R fractions w i t h direct - in jected natural gas, emiss ions o f C O , H C and P M are unacceptably h igh and ef f ic iency is impai red [38]. F o r both diesel and natural gas d i rect - in ject ion engines, the recirculated exhaust gases conta in substantial quantities o f P M . Subsequent part icle nuc leat ion , as w e l l as growth by condensat ion and agglomerat ion, occurs w h e n the E G R is m i x e d into the c lean intake air at d i lu t ion ratios o n the order o f 3 -7 :1 ( E G R fractions o f 3 0 % - 1 5 % by mass) . S u c h low - ra t io d i lu t ion processes strongly promote particulate nucleat ion and growth [99]. The effects o f these recirculated particles on intake system fou l ing and engine wear have been studied extensively . The concentrat ion o f P M i n unf i l tered reci rculated exhaust gas precludes int roduc ing the E G R into the intake o f the compressor system due to concerns over excessive f o u l i n g [229,231] . Depos i t i on i n the intake m a n i f o l d and o n the intake valves is also o f concern . Increases i n carbon contaminat ion o f the lubr icat ing o i l , and corresponding increases i n engine wear , are also associated w i t h E G R use [232,233] . Th i s has been attributed m a i n l y to higher overal l in -cy l inder soot concentrations i n the post -* A modified version of this chapter will be presentated at the SAE 2006 Fall Powertrain and Fluid Systems Conference. McTaggart-Cowan, G.P., S.N. Rogak, S.R. Munshi, P.G. Hi l l , and W.K. Bushe. The Effects of Reingested Particles on Emissions from a Heavy-Duty Direct Injection of Natural Gas Engine. SAE Technical Paper 2006-01-3433. 2006. Reprinted with permission from SAE. - 1 5 9 -combust ion per iod [234] ; however , increases i n sulphur ic and ni t r ic ac id concentrations in the ingested charge due to E G R may also be important sources o f wear under some condi t ions [235,236] . These effects o n engine wear and deposi t ion are t radi t ional ly attributed to the P M recirculated w i t h E G R . F r o m a combust ion standpoint, studies us ing either filtered or synthetic E G R have general ly s h o w n increases i n P M emiss ions w i th increasing charge d i l u t i o n ; however , no compar isons between filtered and unfi l tered E G R condit ions have been reported [123,237] . Inert part icles i n an atmospheric-pressure non -p remixed f lame inf luence the ign i t ion process both by c o o l i n g the surrounding gases and ( i f the part icle number density is h i g h enough) by m o d i f y i n g the ve loc i ty field [238]. The author has been unable to ident i fy any pub l i shed w o r k to date that spec i f ica l ly addresses the quest ion o f the effect o f the recirculated carbonaceous part icles on the combust ion event. The current w o r k , therefore, focuses spec i f i ca l l y o n ident i fy ing whether the recirculated part icles contr ibute s ign i f icant ly to the observed increase i n P M emiss ions w i t h E G R . 8.2 Experimental Methodology T h e testing was carr ied out on the heavy-duty single cy l inder research engine descr ibed i n Chapter 3 , us ing the 17:1 compress ion ratio p iston. The two major adaptations for the current set o f tests were the instal lat ion o f a filter i n the E G R system (as shown i n F igure 8.1) and the addi t ion o f an intake air sample l ine to the P M s a m p l i n g system (F igure 8.2). The f i l ter selected for the E G R system was a Head l ine Fi l ters grade 5 0 C , constructed o f P o l y V i n y l i d i n e D i F l u o r i d e ( P V D F ) f luorocarbon resin [239]. A s the f i l ter 's rated temperature was on ly 150°C, the filter was instal led downstream o f the E G R cooler (Figure 8.1) to prevent thermal over loading. The filter was rated to remove i n excess o f 9 9 . 9 9 % o f so l id and l i q u i d part icles larger than 100 n m . Performance for sizes smal ler than 100 n m was not stated. A s the engine exhaust contains many particles w i t h m o b i l i t y diameters less than 100 n m , the part ic le concentrat ion i n the intake was measured to determine whether the f i l ter was adequately r e m o v i n g the smal l particles f r o m t h e , E G R . A s the P M load ing i n the intake stream w i t h the f i l ter was very l o w , the T E O M was not able to accurately measure the intake P M concentration. A s a result, on ly the S M P S and the Aetha lometer were used to measure intake concentrations. The intake air sample - 1 6 0 -bypassed the pr imary d i lu t ion system and was drawn direct ly into the secondary d i lu t ion system (Figure 8.2). The d i lu t ion ratio prov ided by the secondary d i lu t ion system was f i xed at 6:1 by the constant f l o w rates through the two instruments and the m a x i m u m f l o w through the rec i rcu lat ion system. A f u l l - f l o w ba l l va lve was used to select between the intake and exhaust sample streams. F o r the intake system, a secondary vent l ine into the test ce l l ensured that the sample entering the d i lu t ion system was at ambient pressure. The effectiveness o f the E G R f i l ter at the var ious exper imental condit ions is presented i n sect ion 8 .3 .1 . 8.2.1 Experimental Conditions A s discussed i n section 3 . 3 , the operation o f the s ing le -cy l inder research engine can be fu l l y def ined f r o m the equivalence ratio ((j)), speed, indicated power , the intake oxygen mass f ract ion (Yj n to2), and the combust ion t i m i n g ( 5 0 % I H R ) . F o r the tests descr ibed here, the engine operat ing cond i t ion and the speci f ic test modes used are s h o w n i n Table 8 .1 . E a c h test mode was repeated to establish a statistically v a l i d sample for compar i son between the f i l tered and unf i l tered E G R condit ions (the number o f repl icat ions for each point are also s h o w n i n Tab le 8.1). The test modes were chosen to generate a range o f exhaust particulate concentrations w h i l e m i n i m i z i n g the var iabi l i ty in engine operation. The n o n - E G R mode , A , p rov ided a reference cond i t ion , as the presence o f the f i l ter should not in f luence the results w i t h no E G R . A n intermediate E G R case ( B , Yj n to2 = 0.21), and a pair o f h i g h E G R cases ( C and D , Yj n t 02 = 0.19) prov ided a range o f recirculated P M loadings and intake d i lu t ion levels . A t the h i g h E G R cond i t ion , the two t imings (early t i m i n g , C , and late t i m i n g , D ) p rov ided l o w and h igh P M loadings in the recirculated exhaust at equivalent intake d i lu t ion levels . C o m p l e t e randomizat ion o f the tests was impract ica l , so the tests were conducted i n this order: 2 sets o f non- f i l tered tests; 7 sets o f f i l tered tests; 5 sets o f non- f i l tered tests. The order o f test ing o f the modes w i t h i n the test sets was fu l l y randomized . F o r each point , the engine operat ing cond i t ion was held stable for at least 5 minutes pr ior to co l lec t ing data. D a t a co l lec t ion durat ion was a m i n i m u m o f 10 minutes. 8.3 Results T o assess the inf luence o f the E G R f i l ter on engine operat ion, the in - cy l inder pressure trace, gaseous emiss ions , and detai led particulate matter emiss ions are evaluated. The i n -cy l inder pressure and gaseous emissions w o u l d be affected on ly i f the recirculated particulate - 1 6 1 -has a major effect on the overal l combust ion process. The particulate matter fo rmat ion process, o n the other hand , c o u l d be inf luenced by the recirculated part ic les without s ign i f icant ly in f luenc ing the overal l combust ion . H o w e v e r , before the impacts o f the reci rculated part icles o n the combust ion and particulate format ion processes are assessed, the effectiveness o f the E G R fi lter needs to be discussed. 8.3.1 Filtration Effectiveness T o assess the effectiveness o f the E G R f i l ter at reducing part iculate l o a d i n g i n the ingested charge, samples f r o m the intake stream were analyzed w i t h the S M P S and the Aethalometer . F igure 8.3 indicates the P M size distr ibutions in the intake for modes A and D i n a l o g - l o g format. The S M P S shows a 9 8 % reduct ion in the peak part ic le number between the f i l tered and unf i l tered condit ions. S i m i l a r l y , the Aethalometer indicates a reduct ion o f over 9 6 % i n the intake b lack carbon concentrat ion at a l l test modes (Figure 8.4) w h e n the f i l ter is added to the E G R l ine. Images o f T E M grids sampled f r o m the intake (at the same f l o w rate, d i lu t ion ratio, and sample t ime) show a very substantial reduct ion i n part ic le concentrat ion w i t h the f i l ter (see A p p e n d i x 7). The b lack carbon mass and size distr ibut ions i n the intake for case A (Figure 8.3) are indist inguishable f r o m background no ise , as w o u l d be expected for this n o n - E G R case. These results indicate that the filter is ef fect ively r e m o v i n g the P M f r o m the intake stream. Furthermore, since no increase in nuc leat ion mode part ic les is observed, it can be conc luded that no direct nucleat ion o f vo lat i les species is occur r ing i n the intake downstream o f the filter. Therefore, compar i son o f the filtered and unf i l tered E G R cases should ident i fy any substantial inf luences o f the recirculated P M on the combust ion and pol lutant format ion processes. 8.3.2 In-Cylinder Performance I f the recirculated particles do s igni f icant ly inf luence the combust ion event, di f ferences i n the in - cy l inder pressure trace between the filtered and unf i l tered cases w o u l d be expected. N o s igni f icant differences i n the pressure trace or heat release rate are observed at mode D , as s h o w n i n F igure 8 .5 . S i m i l a r results apply for the other operat ing modes (shown i n A p p e n d i x 6). A compar ison o f combust ion progress parameters, i n c l u d i n g the ign i t ion delay t imes (for both the diesel p i lo t and the natural gas), the peak heat-release rate, and the burn durat ion, are s h o w n i n F igure 8.6. The p i lo t ign i t ion delay t ime is not s ign i f icant ly different at any o f the observed condit ions. A n increase in the natural gas - 1 6 2 -i gn i t ion delay t ime (commanded start o f natural gas in ject ion to observed start o f gas combust ion) is observed at on ly one mode , C . Th is can be attributed to the higher E G R fract ion at that mode w i t h the f i l ter as shown i n F igure 8.7. The longer ign i t ion delay is consistent w i t h prev ious results w h i c h suggest that increases in E G R can s igni f icant ly increase gaseous ign i t ion delay t imes [131]. The peak heat-release rate is also s l ight ly h igher at this mode , due to a higher intensity p remixed burn phase result ing f r o m the longer ign i t ion delay per iod for the natural gas. The burn durat ion (as def ined by the 1 0 - 9 0 % integrated heat release rate), however , does not show any s ignif icant differences between the f i l tered and unf i l tered cases. F r o m these results, it is apparent that on ly at mode C are there s igni f icant di f ferences, and that even at that cond i t ion , the recirculated particulate is not s igni f icant ly in f luenc ing the overa l l combust ion event. However , m i n o r effects in f luenc ing the gaseous or part iculate emiss ions are not necessari ly discernable f r o m the pressure trace. 8.3.3 Emissions and Performance A s indicated by the in -cy l inder pressure results, the engine's operating cond i t ion remains re lat ively constant between the f i l tered and unfi l tered cases; the most s igni f icant dif ference is the higher E G R level at mode C . Th is is a result o f the greater var iab i l i ty in E G R leve l w h i c h results f r o m the vary ing pressure drop through the f i l ter. The higher E G R leve l has a s igni f icant inf luence o n the combust ion and emiss ions fo rmat ion processes at mode C . A l l the other modes show good operating condi t ion repeatabil i ty , i n c l u d i n g the E G R fract ion and G I M E P (shown i n F igure 8.7) as w e l l as the overa l l equivalence rat io, combust ion t i m i n g , and p i lot f ract ion (not shown) . N o statistically s igni f icant var iat ions are detectable at any mode i n any o f these parameters. A s the engine operating condi t ions (aside f r o m mode C ) remain consistent between the f i l tered and non- f i l tered E G R tests, any differences i n measured emissions are attributable to the recirculated P M . H o w e v e r , because the testing was not complete ly randomized , some error due to day - to -day var iab i l i ty may also have in f luenced the readings. A compar ison o f the results at mode A quantif ies this effect, s ince the filter's presence or absence should have no inf luence w h e n there is no f l o w through the E G R system. C o m p a r i s o n s o f the average emiss ions .o f N O x , C O , H C , and P M between the f i l tered and non - f i l te red cases are s h o w n i n F igure 8 .8 . The error bars o n the plot represent the 9 5 % conf idence intervals about the mean o f the col lected data. The data, i n c l u d i n g the - 1 6 3 -uncertainties, are shown i n Table 8.2. For the gaseous emiss ions , the var iat ions i n the emiss ions are relat ively s m a l l . The on ly statistically s ignif icant var iat ion (at a 9 5 % conf idence l e v e l , determined through compar ison o f means us ing Student's T-test) is at mode C , where N O x emiss ions are s igni f icant ly lower w i t h the f i l ter than without , w h i l e P M levels are s ign i f icant ly higher (Figure 8.8); this is direct ly attributable to the differences i n E G R leve l at this cond i t ion that were discussed earlier. A t a l l other modes , the reci rculated P M does not s igni f icant ly inf luence emissions levels. However , the uncertainties on the P M total mass measurements are suff ic ient ly large that subtle effects are not discernable. 5.3.4 Particulate Matter T h e var iab i l i t y i n the P M mass measurements is due p r imar i l y to the re lat ively l o w P M levels at most o f the condit ions. A day - to -day bias is also apparent, w i t h the days where filtered testing was carr ied out hav ing , on average, higher measured P M mass levels . Th i s is observable even under condi t ions where the presence o f the E G R f i l ter shou ld not in f luence the results (mode A , 0.23 Yjn to2)- Further insight into the P M emiss ions may be developed f r o m the average part ic le size distr ibutions for the A and D cases. In F igure 8.9 under n o n -E G R condi t ions (case A ) , the number o f part icles for the unf i l tered cases can be seen to be s l ight ly increased. Th i s is generally consistent w i t h the observed increase i n mass at this cond i t ion , and is indicat ive o f the var iabi l i ty in both the mass and number measurements. A t h igher E G R levels (w i th late t i m i n g , case D ) , the total number o f part ic les is reduced. The plots i n F igure 8.9 also show the intake part icle number concentrat ions; however , on ly the unf i l tered case D sample shows a discernable number o f particles. T h e size distr ibut ions are the average values for a l l the data co l lected at the g iven operating cond i t ion . Further insight may be gained f r o m integrating the size distr ibut ions to estimate a m o b i l i t y v o l u m e o f particles i n the S M P S measurement range (ultra- f ine part ic les, m o b i l i t y diameter o f 5 -150nm) . The mob i l i t y v o l u m e is not a measure o f the true v o l u m e occup ied by the part ic les, due to their non-spher ica l shape; however , it is representative o f the total quantity o f u l t ra - f ine particulate.. B y m u l t i p l y i n g the m o b i l i t y v o l u m e by an effect ive density, an estimate o f the mass can be generated [159,160,161] . H o w e v e r , f r o m the current w o r k there is insuff ic ient in format ion to determine the effective density for these part ic les, as d iscussed in sect ion 3 .2 .2 . Despite this, as the effective density depends p r i m a r i l y on size, changes i n the mob i l i t y vo lume w i l l be generally representative o f changes in ultrafine - 1 6 4 -part ic le mass. The m o b i l i t y vo lume results are shown i n F igure 8.10. F o r mode D , a reduct ion i n u l t ra - f ine part icle v o l u m e o f approx imately 1 5 % is suggested w i t h the E G R f i l ter . Converse ly , the other modes a l l show an increase i n part ic le v o l u m e , general ly consistent w i t h the increase in total mass indicated in F igure 8.8. The effect o f part icles larger than 150 n m i n diameter, as w e l l as adsorption o f volat i les and water vapour on the T E O M fi l ter , may exp la in the substantial ly lower vo lumes at l o w E G R levels (relative to the h igh E G R modes) measured by the S M P S i n compar ison to the T E O M readings. Spec i f i ca l l y , evaporated lubr icat ing o i l recondensing on the T E O M filter m a y be generating a basel ine P M mass measurement that is not detectable by the S M P S . D a y - t o - d a y var iab i l i ty o f the T E O M at l o w emiss ions levels may also be contr ibut ing to the discrepancies between the T E O M and S M P S measurements. The b l a c k - c a r b o n concentrations i n the exhaust stream, as measured by the Aethalometer , are shown i n F igure 8.10. The mass fract ion o f b lack -ca rbon (Aethalometer result / total mass as measured by T E O M ) is also shown. These results indicate that the use o f the f i l ter substantial ly reduces the total mass o f b lack carbon i n the exhaust at the highest P M load ing cond i t ion (mode D ) . A t mode C , the increases in b lack carbon content and ul t ra -fine part ic le mass are on ly approx imately o n e - h a l f those o f the increase i n mass measured by the T E O M . Th i s results i n a reduct ion in the fract ion o f the total P M that is attributable either to the u l t ra - f ine part icles or to b lack carbon. The fact that these results are not observed i n the T E O M data is most l i k e l y due to a corresponding increase i n vo lat i le fractions i n the exhaust, poss ib ly resul t ing f r o m reduced particle surface area avai lable for condensat ion. 8.4 Discussion The general emiss ions and performance results are consistent in suggesting that the part ic les i n the intake stream are not hav ing a s igni f icant inf luence o n the overa l l combust ion event. H o w e v e r , there is ind icat ion o f a reduct ion i n b lack -carbon part icle mass (soot) in the exhaust stream due to the f i l ter ing o f the recirculated particles. If a l l the recirculated part icles were pass ing unreacted through the combust ion chamber, a reduct ion i n exhaust P M mass o f - 3 0 % w o u l d be expected f r o m the reduct ion o f P M i n the ingested charge. S ince such a reduct ion is not observable i n the col lected data, it appears that m u c h o f the P M is be ing o x i d i z e d dur ing the combust ion process. A s the engine was operating overa l l lean (<j>02 o f 0.6) , approx imate ly 4 0 % o f the charge d id not participate direct ly i n the combust ion event. I f - 1 6 5 -the P M i n this unreacted charge was not o x i d i z e d , a 1 0 - 1 5 % reduct ion i n exhaust P M mass w o u l d be expected; this f igure is i n fact roughly equivalent to the reductions suggested by the S M P S and the Aethalometer . These results indicate that the P M w h i c h is contained in the unreacted charge passes through the cy l inder and is emitted, w h i l e the P M i n the por t ion o f the charge w h i c h participates i n the combust ion is consumed. Th i s is not surpr is ing , g i ven that the env i ronment i n the v i c in i t y o f the combust ion event w o u l d be at higher temperatures and w o u l d contain a s ignif icant poo l o f radicals to enhance the ox idat ion rate o f the reci rculated part ic les. Th i s reasoning suggests that the bu lk o f the P M emitted dur ing E G R operat ion is newly fo rmed i n the combust ion event. A t mode C , the increase in P M due to the higher E G R leve l for the f i l tered tests masks these effects. A t mode B , the antic ipated inf luence o f the recirculated particles w o u l d be only ha l f that at mode D , g iven the lower E G R f l o w . A 5 % reduct ion w o u l d be expected at mode B ; but this is not discernable in the data due, i n a l l l i k e l i h o o d , to the var iab i l i ty i n the P M levels . These results have a number o f s ignif icant impl icat ions for an operat ional system where E G R is used i n conjunct ion w i t h an exhaust particulate remova l dev ice . There is no apparent benefit , i n terms o f P M format ion , to locat ing the E G R loop downst ream o f the exhaust P M trap. Nonetheless , there are other benefits o f c leaning the exhaust gases before it is rec i rculated, i n c l u d i n g reducing particulate bu i ldup i n the E G R and intake systems as w e l l as i m p r o v i n g engine longevi ty and reducing o i l degradation. 8.5 Conclusions 1. A s l o n g as the E G R rate is equivalent between the filtered and unf i l tered cases, no statist ical ly ident i f iable differences i n gaseous emiss ions , total part ic le mass, or combust ion performance occur. 2 . In general , the fo rmat ion o f P M under E G R operation is independent o f the quantity o f the recirculated particulate. Th i s suggests that the bu lk o f the P M emitted is freshly fo rmed. 3 . The b lack -ca rbon masses and ultrafine particle size distr ibut ions suggest that reci rculated P M far f r o m the react ion zone is emitted f r o m the engine unreacted. The P M that is i n the charge that participates in the combust ion event is fu l l y o x i d i z e d and does not inf luence the format ion o f new particles. - 1 6 6 -4. W i t h o u t a f i l ter i n the E G R l ine , s igni f icant quantities o f particulate are reingested into the engine w h e n us ing E G R . Instal l ing a f i l ter can reduce the total P M load ing i n the intake stream by more than an order o f magnitude. 8.6 Figures and Tables Table 8 . 1 : Base engine operating cond i t ion and test modes Speed 1200 R P M Load (G IMEP) 13.5 bar 0 2 E Q R 0.6 Injection Pressure 21 M P a Diesel Quantity 5 mg/inj Mode Y i n t 0 2 E G R % 50%IHR ("ATDC) Sampl Unfiltered s size Filtered A 0.23 0 10 6 6 B 0.21 15 10 8 6 C 0.19 30 0 7 6 D 0.19 30 10 7 7 (reprinted with permission from SAE 2006-01-3411. © SAE International 2006) Table 8 .2 : M e a n (Standard Dev iat ion) o f var ious parameters M o d e Filter G I M E P bar <(>02 E G R % G I S F C g/GikWhr N O x g/GikWhr C O g/GikWhr H C g/GflcWhr P M g/GikWhr A un filtered 13.42(0 . 1 17) 0.599(0.009) UO) 185.4(0.44) 5.30(0.10) 0.69(0.07) O-i(t(0 01) 0.007(0.002) lil'.ered 13.42(0.08) 0 . 6 ( 1 7 ( 0 . 0 1 0 ) 0(0) 186 1(0 91) 5.02(0.08) 0.81(0.08) 0 30(0 1)1) 0 010(0 0 0 3 1 1) unfiltered 13.43(0.08) 0.601(0.007) 14.0(1.5) 184.6(0.74) 2.22(0.44) 1.43(0.39) 0.50(0.07) 0 .014(0.006) filtered 13.43(0.07) 0.596(0.008) 13.8(3.2) 184.8(0.71) 2.29(0.43) 1.55(0.36) 0.47(0.06) 0 .015(0.004) r untilleivd I3.90(0.O7) 0.594(0.020) 24.3(0.8) 177.5(0.44) 3.45(0.30) 1.41(0.19) 0.69(0 05) 0.009(0.002) filtered 13.89(0.06) 0.593(0.013) 2o.4t2.Ui 177.3(0.50) 2.95(0.40) 1.67(0.24) 0 .71(0 .04) 0 .0 )4(0 .003) D unfiltered 13.52(0.04) 0.588(0.008) 26.4(0.7) 183.1(0.88) 0.96(0.08) 2.92(0.22) 0.81(0.07) 0 .027(0.003) filtered 13.43(0.08) 0.582(0.009) 26.4(2.4) 183.5(2.03) 1.00(0.18) 3.03(0.52) 0.81(0.09) 0.027(0.006) (reprinted with permission from SAE 2006-01-3411. © SAE International 2006) To Exhaust To Emissions Bench/h Exhaust < Pressure Relief Wastegate Back-pressure Valve Insulated Sample Line [ From Air Compressor Shut-off Valve EGR Cooler <'f From Ambient — * Insulated] Sample Line SJ Exhaust Surge Tank EGR Valve ToPtvi System Engine Filters \ Pressure Z1 Regulators Figure 8 . 1 : Eng ine air exchange system w i t h E G R fi lter. - 1 6 7 -T Back-Pressure Regulator Insulated Sample Line Vent to Atmosphere Heated f i x i n g Orifice TEM Sampler t Gravimetric Filters Vent to Atmosphere T t Rotameter Raw Exhaust Heated Mixing Section TEOM Bottled Nitrogen From Intel Manifo Aethalometer Secondary Dilution System Pump, Filters, & Flow Meters Figure 8 .2 : P M sampl ing system m o d i f i e d for intake sampl ing . 1.E+07 1.E+03 Figure 8 .3 : 500 S 400 3 c "I 300 cn - 2 0 0 o3 Q 100 m 0 10 100 Mobil ity D iameter (nm) - 'D', intake, unfiltered • 'D' intake, filtered x 'A' intake, unfiltered • 'A' intake, filtered 1000 Intake P M size distr ibut ions, modes A and D. (reprinted with permission from SAE 2006-01-3411. © SAE International 2006) • Unfiltered • Filtered j /IrK • 11  1 0 0 % 9 8 % 9 6 % 9 4 % 9 2 % 9 0 % llli IB ll|ll|p ^^^B life?! ^^Si «ittiiiit Figure 8 .4 : Aethalometer measurements o f intake concentrations and % reduct ion i n b lack carbon between f i l tered and unf i l tered E G R at a l l E G R test modes . E r ro r bars are based o n the poo led uncertainty i n the measured exhaust values (they are not s h o w n for the r ight -hand plot as they are greater than the scale o f the graph). (reprinted with permission from SAE 2006-01-3411. © SAE International 2006) - 1 6 8 --60 -40 -20 0 20 40 60 -20 -10 0 10 20 30 40 Crank Angle (°CA) Crank Angle (°CA) Figure 8 .5 : Pressure trace and heat-release rate at mode D (0.19 Y i n t 0 2 , 5 0 % I H R at 1 0 ° A T D C ) for ind iv idua l f i l tered and non- f i l tered test points , (reprinted with permission from SAE 2006-01-3411. © SAE International 2006) ^ 2 0 o cu ° 10 o a. 0 il Unfiltered • Filtered i - E - i 11 < o >. ra cu Q 16 14 12 10 8 6 4 2 0 H Unfiltered • Filtered 250 < " 200 "E 2 150 or 100 ro 50 cu °- 0 ! Unfiltered • Filtered 30 CA) 25 20 -c o 15 -Q 10 -E CO 5 ) Unfiltered • Filtered ill Figure 8 .6 : Igni t ion delay and combust ion progression for f i l tered and unf i l tered condi t ions . E r ro r bars indicate 9 5 % conf idence intervals, (reprinted with permission from SAE 2006-01-3411. © SAE International 2006) - 1 6 9 -30 ~ 25 $ 2 0 I 15 g 1 0 m 5 0 • Unfiltered • Filtered 15 ~ 14 El Unfiltered • Filtered LU 13 12 11 • 10 i A B C D A B C D Figure 8 .7 : Operat ing cond i t ion variat ions as a funct ion o f test mode for f i l tered and unf i l tered condit ions. Er ror bars indicate 9 5 % conf idence intervals, (reprinted with permission from SAE 2006-01-3411. © SAE International 2006) m Unfiltered • Filtered 1 0.8 1 0.6 O S 0.4 O x 0.2 0 B Unfiltered • Filtered 3 Unfiltered • Filtered r t r f I — •uu X 0.04 f 0.03 b S 0.02 5 o LU J| 0.01 a. B Unfiltered o Filtered Figure 8 .8 : P o w e r - s p e c i f i c emiss ions as a funct ion o f test mode for f i l tered and unf i l tered condit ions. Er ror bars indicate 9 5 % conf idence intervals, (reprinted with permission from SAE 2006-01-3411. © SAE International 2006) 5.E+06 4.E+06 1 " E 3.E+06 o Jt CL Q 2.E+06 _C z TJ 1.E+06 0.E+00 Test Condition "A" unfiltered exhaust 5 ~*£Ti>° filtered exhaust P • unfiltered intake x filtered intake 5.E+07 to 4.E+07 1 3.E+07 o Q- 2.E+07 c JJ Z TJ 1.E+07 0.E+00 10 100 1000 Mobility Diameter (nm) 10 unfiltered exhaust filtered exhaust - unfiltered intake x filtered intake 100 1000 Mobility Diameter (nm) Figure 8 .9 : Part ic le size distr ibutions i n intake and exhaust streams at modes A and D. (reprinted with permission from SAE 2006-01-3411. © SAE International 2006) - 1 7 0 -O) to to to E O CO 5000 4000 3000 2000 1000 0 • Unfiltered • Filtered .9 S T> <S 0.8 I Q O . 2 0.0 n Unfiltered • Filtered Q . CD § l - e > E I S o E 7 -, 6 5 4 3 2 1 0 • Unfiltered • Filtered "T IT • F ^ T ^ T gure 8 .10: B l a c k - c a r b o n concentrat ion, f ract ion o f total particle mass and m o b i l i t y v o l u m e ( P M vol . ) at a l l test modes. Er ror bars indicate 9 5 % conf idence intervals. (reprinted with permission from SAE 2006-01-3411. © SAE International 2006) - 1 7 1 -Chapter 9 Significant Findings and Recommendations The object ive o f this research was to improve understanding o f the gaseous fuel combust ion process o f a p i lo t - ign i ted , direct in ject ion, gaseous- fuel led engine w i t h intake-charge d i lu t ion . Th is objective has been addressed through studies where the gaseous jet energy, the fuel compos i t ion , and the charge compos i t ion were var ied . T h i s thesis has e x a m i n e d factors relat ing to the phys ica l in ject ion and m i x i n g processes as w e l l as fuel chemistry and charge compos i t ion . Because o f interactions between these factors, it is d i f f i cu l t to isolate f r o m any one o f the studies the pr inc ipa l factors affect ing the combust ion process. H o w e v e r , by consider ing a l l the studies presented here i n aggregate, it is poss ib le to develop more robust conclus ions . The conc lus ions for the ind iv idua l studies are set out i n the corresponding chapters and w i l l not be reproduced in detail here. The present chapter offers a general ove rv iew o f the salient f ind ings o f the research and a ims to prov ide an interpretation o f these results i n the context o f i m p r o v e d understanding o f the combust ion process and the key factors leading to pol lutant generation. S u m m a r y plots w h i c h combine results presented i n prev ious chapters are presented where they augment the d iscuss ion . The impl icat ions o f these f ind ings to engine systems development are discussed. F i n a l l y , recommendat ions for future research to develop further understanding o f the combust ion and po l lu tant - fo rmat ion processes, and to improve the combust ion system, are presented. 9.1 Significant New Findings Gaseous Fuel Composition The c o m p o s i t i o n o f the gaseous fuel has a substantial impact o n the combust ion process and emiss ions . H igher ethane and propane concentrations increase P M and C O emiss ions even though they enhance the igni tabi l i ty o f the fue l . The addi t ion o f hydrogen reduces P M , H C and C O , most s igni f icant ly at condit ions where the combust ion is most unstable and where the emiss ion levels o f these species are highest. D i l u t i n g the gaseous fuel w i t h n i t rogen substantial ly reduces emiss ions o f P M , H C and C O , as w e l l as N O x . These effects, s u m m a r i z e d i n F igure 9.1 (note that the magnitude o f the var iat ions i n P M are m u c h larger than for the other species), are i n general substantial ly greater than prev ious ly reported - 1 7 2 -150% O z ° 125% cu 100% 75% methane 35% H 2 4% prop. 9% eth. 40% N 2 125% 0 100% CU 1 75% O" 50% O 25% methane 35% H 2 4% prop. 9% eth. 40% N 2 125% O z 100% Q I 75% CD O 50% I 25% 1 1 1 1 — 1 300% O z 3 200% cu 100% 0% methane 35% H 2 4% prop. 9% eth. 40% N 2 methane 35% H 2 4% prop. 9% eth. 40% N 2 Figure 9 . 1 : S u m m a r y o f effects o f fuel compos i t ion on emiss ions for the h igh levels o f a l l fuel compos i t ions at 1200 R P M , 13.5 bar G I M E P , 5 0 % I H R = 1 0 ° A T D C (note • scale change for P M plot) . for homogeneous-charge spark - ign i t ion systems; this is due to fundamental di f ferences in the P I D I N G engine 's non -p remixed combust ion system. F u e l compos i t ion has s ignif icant effects on emiss ions f r o m a P I D I N G engine: as a result, var iab i l i ty i n the compos i t i on o f commerc ia l -g rade natural gas (not to ment ion the l i k e l i h o o d o f even larger variat ions w i t h the use o f non -convent ional fuel sources), w i l l have a s igni f icant impact on the system's abi l i ty to meet emiss ions standards. It is evident that an easy w a y to evaluate the inf luences o f different fue l blends o n the c o m b u s t i o n process is required. P rev ious ly proposed values, such as the W o b b e Index and the H : C rat io , are insuff ic ient to describe the effects o f fuel compos i t ion on a d i rect - in ject ion o f natural gas engine. A n e w index is proposed, w h i c h includes both the vo lumetr i c energy density (molar heat ing value) o f the fuel and the H : C ratio. Th i s index is able to differentiate the effects o f fue l compos i t i on o n total P M and b lack -ca rbon emiss ions over a range o f operating condi t ions (Figures 7.17 & 7.18). Therefore, it may be o f use i n determining what fuel b lends w i l l be able to meet future emissions standards, or to ensure that a g i ven engine cal ibrat ion is capable o f meet ing emiss ions standards over a range o f gaseous fuel compos i t ions . - 1 7 3 -Recirculated Particulate R e m o v i n g the recirculated particulate f r o m the ingested charge has a negl ig ib le inf luence o n the measured mass o f P M i n the exhaust, ind icat ing that the b u l k o f the particulate emitted w h e n operating w i t h E G R is freshly formed i n each cyc le . Prev ious studies o f the effects o f the P M recirculated w i t h E G R have focused on engine longevi ty and wear ; the current study was the first to investigate the effects on the combust ion process. A l t h o u g h the recirculated particulate had no measurable inf luence on the gaseous emiss ions or the total particulate matter emissions (Figure 8.8), both ultraf ine part ic le v o l u m e and b l a c k - c a r b o n mass are s l ight ly reduced (Figure 8.10). The results indicate that the b u l k o f the reci rculated P M is be ing o x i d i z e d i n the high-temperature, oxygen - r i ch zone o n the ox id i ze r side o f the n o n - p r e m i x e d combust ion event, and the emitted particulate matter is freshly f o r m e d i n the combust ion chamber dur ing each cyc le . H o w e v e r , g i ven that f i l te r ing the reci rculated exhaust gases does reduce the ultrafine particle v o l u m e and b lack carbon mass, it appears that some part icles - most l i ke l y those located in the extremities o f the combust ion chamber - are not be ing complete ly o x i d i z e d . The fact that this relat ively s m a l l d i f ference i n emiss ions is not observed i n the total P M mass results may be attributed to uncertainties i n the mass measurements or the signif icant contr ibut ion o f condensed volat i le substances. Injection Pressure and Operating Conditions Increasing the in ject ion pressure i n a d i rect - in ject ion o f natural gas engine consistently reduces P M emiss ions at m e d i u m and h igh load condit ions w i t h and wi thout charge d i lu t ion ( E G R ) . The effects o f in ject ion pressure on a l l other emiss ions interacted strongly w i t h engine speed, l o a d , and E G R leve l . P rev ious diesel -engine research has shown that increasing the in ject ion pressure o f d i rect - in ject ion engines can reduce P M emiss ions ; however , substantial differences between the combust ion o f a gaseous jet and a l i q u i d spray mean that the results' are not direct ly comparable to a P I D I N G engine. Pre l iminary studies at a single operating cond i t ion also demonstrated that higher in ject ion pressures may reduce P M f r o m a P I D I N G engine. The current research, o n the other hand, revealed that wh i le higher in ject ion pressures reduce P M at h igh load condi t ions , at lower loads, where the differential pressure between the fuel ra i l and the combust ion chamber is higher, increasing the in ject ion pressure has less o f an effect o n P M (Figure 4.5) . S i m i l a r l y , the effects are less s ignif icant at h igh speed, as the relative - 1 7 4 -contr ibut ion o f the gaseous jet to the overal l turbulent mot ion in the combust ion chamber is reduced. The other emissions (Figures 4 .2 - 4.4) showed even stronger dependence on operating cond i t ion , w i t h relat ively less sensit iv i ty to in ject ion pressure. T h i s demonstrates that interactions between the gaseous jet and the in -cy l inder condi t ions ( inc lud ing temperature, pressure, and charge mot ion) have a substantial impact o n the combust ion event, and o n the result ing emiss ion levels. Particulate Matter Emissions The quantity and compos i t ion o f P M emitted f r o m a P I D I N G engine are s ign i f icant ly affected by combust ion parameters inc lud ing in -cy l inder turbulence levels and fuel compos i t i on . A d d i n g large amounts o f nitrogen or hydrogen to the fuel reduces carbonaceous P M to undetectable levels ; however , s ignif icant total mass emiss ions are st i l l observed. These are attributed p r i m a r i l y to vo lat i le species. Increasing the in -cy l inder turbulence (wi th higher in ject ion pressures, F igure 4 . 5 , or by d i lu t ing the fuel w i t h ni t rogen, F igure 7.6) substantial ly reduces P M e m i s s i o n levels . Increasing radica l concentrations through the addit ion o f hydrogen to the gaseous fuel can achieve s imi la r results (Figures 5.1 & 5.2). A d d i n g ethane or propane to the fue l increases P M emiss ions substantial ly (Figure 6.10), demonstrat ing the importance o f the c h e m i c a l structure o f the fuel on P M format ion . In general, b lack carbon and ultraf ine part ic le emiss ions are reduced more than the total P M (as demonstrated for the hydrogen and nitrogen addi t ion cases i n F igure 9.2). A t condit ions where ultraf ine and b lack carbon 20% N 2 40% N 2 15% H 2 35% H 2 Figure 9 .2 : S u m m a r y o f average P M total mass, ultrafine part icle v o l u m e and b lack carbon mass e m i s s i o n reductions for nitrogen and hydrogen fue l l i ng at 1200 R P M , 13.5 bar G I M E P , 5 0 % I H R = 1 0 ° A T D C - 1 7 5 -0.02 0.01 0 Pinj, low Pinj, mid 20% N2 40% N2 23% H2 35% H2 Filter speed speed (non mid load mid load EGR) Figure 9 . 3 : M i n i m u m P M mass levels f r o m a l l tests ( inc lud ing in ject ion pressure, n i t rogen, hydrogen at l o w and h igh loads, and f i l tered E G R tests for the n o n - E G R condi t ion) . particulate emiss ions approach zero, a baseline level o f P M mass emiss ions is st i l l observed; the lowest P M total mass is relat ively consistent for var ious test condi t ions , as s h o w n i n F igure 9 .3 . A large fract ion o f these total mass emiss ions is composed o f n o n - b l a c k carbon (al l the points s h o w n had B C concentrations less than 1 5 % o f the total mass) . These n o n - B C emiss ions are p r imar i l y condensed volat i les , most l i ke l y or ig inat ing f r o m the lubr icat ing o i l or p i lo t fue l . These emiss ions , whose levels are not substantial ly affected by mechan isms w h i c h reduce P M format ion f rom the gaseous combust ion event, are a serious concern i n that they approach the u p c o m i n g regulatory l imi ts for heavy-duty engine emiss ions (Table 1.1). Gaseous Fuel Ignition Delay E n h a n c i n g the react iv i ty o f the gaseous fuel results i n s igni f icant reduct ions i n the ign i t ion delay t ime. H o w e v e r , l imitat ions to the reductions achievable through k inet ic enhancements indicate that the fue l -a i r m i x i n g process is also a cr i t ica l component i n this n o n - p r e m i x e d p i lo t - f lame- ign i ted combust ion event. T h e add i t ion o f hydrogen, ethane, or propane to natural gas fue l has been s h o w n i n the literature to increase the igni tabi l i ty o f natural gas. The current study investigated the inf luence o f such k inet ic effects on the ign i t ion delay o f the p i lo t - ign i ted gaseous fue l . Tests were conducted w i t h moderate intake d i lu t ion ( E G R ~ 3 0 % ) . K i n e t i c enhancement w i t h hydrogen was tested at both mid -speed h igh - load and low-speed l o w - l o a d ; ethane or propane were on ly tested at the former condi t ion . For the condit ions tested, a l l three fuels indicated that k inet ic enhancement s igni f icant ly reduces the ign i t ion delay o f the gaseous jet (F igure - 1 7 6 -0 0 </> ro 125% 100% 75% 50% low speed, low load, with E G R Gaseous Ignition Delay mid speed, high load, with E G R o - CM O X o"- CM CO X CM CL 0 CM X ro 0 m E co 0 ^ ro CD £ (D 0 c CO vP 15 ™ 0 c ro CL O 0 c ro "* p CM o •<fr Q. F igure 9 .4 : Changes (relative to natural gas) i n gaseous fuel ign i t ion delay (SOI to S O C ) , at l o w load and h igh load for fuel compos i t ion and in ject ion pressure (Pinj ) tests, a l l at 5 0 % I H R - 1 0 ° A T D C , ~ 3 0 % E G R 9.4). H o w e v e r , the ethane addi t ion tests suggest a possible l imi ta t ion on the potential for k inet ic enhancement o f the ign i t ion process, w i t h no further reduct ion i n ign i t ion delay between the intermediate and h igh ethane concentrations. Th i s suggests that, as kinet ics are i m p r o v e d , f u e l - o x i d i z e r m i x i n g becomes the c ruc ia l factor l i m i t i n g igni tabi l i ty . Reduct ions i n ign i t ion delay are somewhat greater w i t h h igh levels o f hydrogen addi t ion ( 3 5 % by v o l u m e ) ; at these h igh concentrations, the f l u i d mechanics o f the gas in ject ion and m i x i n g are s igni f icant ly affected by the jet compos i t ion . B o t h nitrogen d i lu t ion o f the fuel and higher in ject ion pressures demonstrate that enhanced fuel -a i r m i x i n g can s igni f icant ly reduce the ign i t ion delay. W h i l e higher in ject ion pressures increase the turbulent m i x i n g i n the early stages o f the in ject ion process, the nitrogen d i lu t ion may relate more to changing the spatial and temporal locat ion o f ignitable fue l -a i r mixture i n the gaseous jet relative to the diesel p i lo t f lame. Gaseous Fuel Combustion Rate Increasing the rate o f fuel -a i r m i x i n g reduces the gaseous fuel combust ion durat ion at a l l operat ing condit ions. Enhanc ing the reactivity o f the fuel also increases the combust ion rate at l o w l o a d ; however , it has no signif icant effect at h igh load. Th i s indicates that the h igh load c o m b u s t i o n is p r imar i l y l im i ted by fue l -a i r m i x i n g . A t l o w loads, both m i x i n g and c h e m i c a l k inet ics are important, because o f the relat ively greater importance o f the ign i t ion and p r e m i x e d combust ion event at these condit ions. - 1 7 7 -g 120% CD CO CO CD > ro 100% 80% 60% low speed, low load Combustion Duration mid speed, high load CM CN C CL CD c CD CD E CN X LO CO CN o CD co c5* O ^ D . CL Figure 9 .5 : Changes (relative to natural gas) i n combust ion durat ion ( 1 0 % to 9 0 % I H R ) , at l o w load and h igh load for fue l compos i t ion and in ject ion pressure (Pinj ) tests, a l l at 5 0 % I H R = 1 0 ° A T D C , - 3 0 % E G R H i g h e r in ject ion pressures increase the fuel in ject ion rate, the turbulent shear stress, and the k inet ic energy transfer to the charge. These factors a l l contribute to higher turbulent m i x i n g intensities and s ignif icant reductions i n the combust ion durat ion (Figure 4.8) . Th i s has more o f an impact at higher loads, where the in -cy l inder pressures are higher , suggesting that m i x i n g l imi tat ions become more s igni f icant as the di f ferent ial pressure between the injected fue l and the cy l inder charge is reduced (Table 4.4). B y adding hydrogen, ethane, or propane to the fue l , the overa l l reactivity o f the gaseous jet is enhanced; at l o w load , the hydrogen addi t ion results i n a s ignif icant reduct ion i n combust ion durat ion, but at h igh load , these addit ives do not s igni f icant ly affect the overal l combust ion durat ion (Figure 9.5). O n l y higher in ject ion pressures or the addit ion o f nitrogen to the fuel reduces the combust ion durat ion at h igh load ; this is a result o f increased k inet ic energy transfer f r o m the gaseous jet to the combust ion chamber and hence enhanced late-cycle m i x i n g . Th i s f i n d i n g is part icular ly interesting as the in ject ion durat ion for the nitrogen d i lu t ion was tw ice that o f convent ional natural gas, because o f the lower energy density o f the fue l . These results show that at h igh load w i t h moderate levels o f intake d i lu t ion , m i x i n g l imi tat ions dominate the combust ion process, w h i l e at l o w loads both chemica l k inet ics and m i x i n g can substantial ly enhance the combust ion rate. Processes w h i c h improve m i x i n g , either for the gaseous jet or i n the post - in ject ion combust ion chamber , reduce emiss ions o f unburned hydrocarbons and particulate matter. O n the other hand , carbon m o n o x i d e product ion depends on operating cond i t ion , suggesting that at some condi t ions , mechanisms other than m i x i n g (possibly increased loca l ext inct ion events due to h i g h turbulent shear stresses) are contr ibut ing to the fo rmat ion o f this pol lutant. - 1 7 8 -Improved m i x i n g does not s igni f icant ly affect the fuel convers ion ef f ic iency , but under certain condi t ions N O x emiss ions are increased. M o r e rapid m i x i n g may also reduce the amount o f t ime at w h i c h the fue l - r i ch zones are at h igh enough temperatures to f o r m soot; the result is a reduct ion i n P M emiss ions. Combustion Stability Increased combust ion stabi l i ty correlates direct ly w i t h reduced emiss ions o f C O and H C (Figure 7.15). The addi t ion o f hydrogen or nitrogen to the fuel improves the combust ion stabi l i ty and correspondingly reduces emiss ions. Ethane or propane addi t ion has l i tt le impact on combust ion stabi l i ty at most condi t ions, and hence has l ittle impact on H C emiss ions ; however it does cause the C O emiss ions to increase. Th is increase indicates that other mechan isms besides combust ion stabil ity are contr ibut ing to these emiss ions . W h i l e processes w h i c h improve combust ion stabi l i ty also reduce emiss ions o f C O and H C , it is unclear whether these are caused direct ly by improvements in combust ion stabi l i ty . Poss ib le explanat ions inc lude fewer loca l turbulent shear-related ext inct ion events, more stable ign i t ion processes, or more complete late-cycle ox idat ion o f the combust ion by -products . Fuel Conversion Efficiency F u e l convers ion ef f ic iency is relat ively unaffected by modi f i cat ions to the fue l c o m p o s i t i o n or enhancement o f the turbulent k inet ic energy through gaseous jet modi f i ca t ions . W h e n corrected for the energy density o f the var ious fuel b lends, ni t rogen or propane addi t ion s l ight ly improved fuel convers ion ef f ic iency , w h i l e ethane s l ight ly reduced it ; neither hydrogen addi t ion nor pure methane fue l l ing had any effect o n e f f i c iency (F igure 9.6). H o w e v e r , these effects were substantial ly smal ler than the effects o f operating cond i t ion , i n c l u d i n g in ject ion t i m i n g (also shown i n F igure 9.6) , engine speed, and load (Figure 4.1) . Th i s reinforces the general ly recognized dominance o f operat ing cond i t ion on fuel convers ion e f f i c iency ; it is interesting that intake d i lu t ion ( E G R ) has no s igni f icant effect on ef f ic iency at the E G R levels investigated. It should be noted that very h igh d i lu t ion ratios (>40% E G R ) were not tested; it is l ike ly that, at condit ions where combust ion instabi l i ty is very h i g h , processes that reduce the combust ion var iabi l i ty w i l l enhance the combust ion ef f ic iency . - 1 7 9 -o 110% 0 jg 100% ro .a 0 •M 90% m 0 80% 0 CM 0 c I c ro -~P CO eth LO ° : C L ^ O E CO C L 0 ^ ra 0 CM o b_ • T 3 CT> IS «1 Figure 9 .6 : Changes (relative to natural gas) i n fuel ef f ic iency ( G I S F C ) , at 1200 R P M for fue l compos i t ion , in ject ion pressure (Pinj ) , and f i l tered E G R tests, as w e l l as for retarded t i m i n g condit ions. A l l (except retarded) at 5 0 % I H R = 1 0 ° A T D C , - 3 0 % E G R Engine Operating Conditions The strong interactions between the engine's operating cond i t ion , the fuel compo s i t i o n , and the in ject ion pressure indicate that mul t ip le mechanisms are contr ibut ing to the format ion o f any g i ven pol lutant. These f indings reinforce results f r o m studies o f other combust ion systems w h i c h have shown that at different operating condi t ions , different pol lutant fo rmat ion mechanisms are dominant . F o r P I D I N G combust ion , the combust ion process responds dif ferently to hydrogen addit ion under l o w - l o a d and h i g h - l o a d condi t ions , w i t h increases in combust ion intensity at l o w load , but reduced combust ion intensity at h igh load. Furthermore, the effects o f in ject ion pressure o n C O emiss ions were contradictory between l o w and h i g h engine speeds at the same intake d i lu t ion fract ion and engine load. E v e n var iat ions i n combust ion t i m i n g at constant operating condi t ion have a s igni f icant effect; for example , w h i l e ni t rogen d i lu t ion reduced P M emissions at most t imings , at the earliest t i m i n g , the P M emiss ions increased s igni f icant ly . These f indings demonstrate the importance o f conduct ing tests over the widest possible range o f combust ion condi t ions . - 1 8 0 -9.2 Engine System Implications The conc lus ions out l ined in the prev ious section have s igni f icant impl icat ions for o p t i m i z i n g engines us ing d i rect - in ject ion gaseous- fuel led combust ion systems. W h i l e not a l l the results are direct ly appl icable to commerc ia l l y compet i t ive appl icat ions, they do prov ide considerable guidance for engine development. The results indicate that increasing in ject ion pressure tends to result in s igni f icant emiss ions reduct ions w i t h no fuel convers ion ef f ic iency penalty (exc lud ing compress ion w o r k ) , especia l ly at h igh load condit ions. The impact o f increased compress ion w o r k can be m i n i m i z e d w i t h a l iquef ied natural gas fuel storage system us ing a cryogenic p u m p to reach the desired pressures; however , this adds substantial system complex i t ies . Increased in ject ion pressures are most benef ic ia l at condit ions w i t h higher in - cy l inder pressures, w h i c h are typ ica l ly the condi t ions where engine loads are highest. A s a result, for a compressed natural gas system, increasing in ject ion pressures w i t h load w o u l d m i n i m i z e the parasit ic compressor w o r k as a f ract ion o f total engine power over the entire engine operating range. The in ject ion pressure c o u l d be contro l led us ing the engine's on -board computer. The exper imental results further indicate that techniques w h i c h enhance the late stage combust ion process through enhanced in -cy l inder turbulence also help to m i n i m i z e emiss ions . W h i l e the addi t ion o f large quantities o f nitrogen to the fuel as a di luent can satisfy this requirement, this increases both the compressor w o r k and the storage v o l u m e , not to ment ion increased infrastructure system complex i ty . One potential technique for i m p r o v i n g the late -cyc le combust ion w o u l d be to enhance the in -cy l inder charge m o t i o n , either through modi f i cat ions to the combust ion chamber shape or through var iat ions i n the in ject ion process, i n c l u d i n g mul t ip le inject ions or variat ions i n the in ject ion rate. The use o f hydrogen as an addit ive to natural gas is a more l i k e l y alternative than ni t rogen addi t ion . E v e n i n this case, however , there are s ignif icant potent ial barriers i n this case, i n c l u d i n g reduced energy densit ies, higher compress ion w o r k requirements, and more c o m p l e x infrastructure requirements. However , this technique can achieve s igni f icant local air pol lutant and G H G emiss ion reductions. A potential opportunity w o u l d be to develop mul t ip le cal ibrat ions o f the engine system on a variety o f hydrogen - natural gas b lend compos i t ions . U s i n g this technique, engine e f f i c iency c o u l d be o p t i m i z e d through the use o f more advanced t imings and higher levels o f E G R (to contro l N O x ) w h e n hydrogen levels are -181-high enough to ensure that P M emiss ions remain b e l o w legislated levels . Ano ther poss ib i l i t y w o u l d be to offer a reduced aftertreatment system (potential ly e l iminat ing the P M trap, for example) i f the engine is only to be run on hydrogen-blended natural gas. These are just two examples o f techniques that cou ld be used to m a x i m i z e the effectiveness o f hydrogen-b lended gaseous- fuel led combust ion systems V a r y i n g natural gas composi t ions pose a s ignif icant chal lenge to engine developers, as the engine must be cal ibrated to meet emiss ions standards w h i l e st i l l operat ing ef f ic ient ly over the f u l l range o f fue l composi t ions . The results presented here indicate that w h i l e heavy hydrocarbon content does not s igni f icant ly affect the combust ion progress ion, it does substantial ly inf luence emiss ions. A s a result, development o f the engine system, i n c l u d i n g in ject ion processes and aftertreatment, w i l l need to ensure that emiss ions are not excessive w i t h the worst fuel b lends; this w i l l result in higher system costs and reduced ef f ic iency , even w h e n n o r m a l l y h igh purity fuels are used. A d j u s t i n g the in ject ion t i m i n g based on onboard sensing o f fue l compos i t i on is one alternative to overcome this d rawback ; however , the substantial increases i n P M observed w i t h the levels o f ethane and propane tested i n this w o r k suggest that substantial ly greater P M aftertreatment requirements are unavoidable i f 'd i r t ier ' fue l b lends are to be accommodated . One opt ion w o u l d be for regulators to prov ide a fue l qual i ty standard for natural gas, s imi la r to the standards in place for gasol ine and diesel . Ano the r so lut ion w o u l d be to prov ide a fuel qual ity requirement for the engine, potent ial ly i n the f o r m o f a t, va lue, w h i c h w o u l d a l l o w increased levels o f heavier hydrocarbons to be offset by H2 or N 2 addi t ion to the fue l . The f indings o f this research indicate that there is no s igni f icant benefit for the combust ion process i n f i l ter ing the exhaust before it is recirculated i n an E G R system. If a particulate filter is required to control P M emiss ions , the increased back-pressure generated by this system c o u l d be used to help generate the pressure dif ferential needed to force some o f the exhaust gases back into the intake stream. Furthermore, by reducing the f l o w rate o f exhaust through the exhaust particulate filter, the size o f the f i l ter and its cor responding cost cou ld be reduced. H o w e v e r , offsett ing these benefits w o u l d be the potential for s igni f icant f o u l i n g o f the intake system through the particulate carr ied w i t h the reci rculated exhaust gases. These are trade-offs w h i c h must be considered i n the development o f the engine -182-system; the results f r o m this w o r k may prov ide some further insight w h i c h can be appl ied w h e n m a k i n g des ign decis ions. 9.3 S t u d y L i m i t a t i o n s The w o r k presented i n this study prov ides insight into the combust ion process o f a d i rect - in ject ion , natural gas fuel led engine. In general , fuel ava i lab i l i t y and system restr ict ions resulted i n the testing be ing carried out over a l imi ted range o f operating condi t ions . The appl icabi l i ty o f these results to extreme operating condi t ions , such as those encountered under very h igh charge d i lu t ion condit ions or wh i le undergoing transient engine operations is uncertain. The experiments were conducted on a mul t i - cy l inder engine m o d i f i e d for single cy l inder operat ion. Th i s m a x i m i z e d the s imi lar i ty o f the single cy l inder engine, as the stock combust ion chamber and intake and exhaust f l o w geometries were retained. The operating condi t ions were also selected to be as close as possible to those o f an equivalent m u l t i -cy l inder engine. Despi te these efforts, potential differences between the s ing le -cy l inder and an equivalent mut l i - c y l inder engine inc lude differences i n the magnitude and phas ing o f the pressure pulsat ions i n the intake and exhaust system. These m a y cause differences i n the charge m o t i o n and residual gas fract ion between the s ingle - and m u l t i - c y l i n d e r engines. Pressure f luctuat ions i n the fuel rai ls o f the mul t i - c y l i nder engine due to the f i r i n g o f ne ighbour ing cy l inders may also introduce s ignif icant inf luences w h i c h w o u l d not have been observed i n the s ing le -cy l inder engine. These factors are o f s igni f icant concern w h e n cons ider ing absolute performance and emiss ions values; however , there is no reason to expect that the general effects ident i f ied i n this w o r k should be different between the s ingle and m u l t i - c y l i n d e r engines. The results presented here, and the conclus ions based o n them, are d rawn f r o m an understanding o f the compos i t i on o f the intake and exhaust streams and a basic understanding o f the combust ion process as prov ided by the in - cy l inder pressure. A s no opt ical access into the combust ion chamber was avai lable , it was not possib le to ver i fy the conc lus ions based on v isua l observations o f the combust ion process. Potent ia l ly , further insight into the overa l l combust ion process, as w e l l as details regarding pol lutant fo rmat ion , c o u l d be ach ieved through a study o f s imi la r combust ion processes i n an opt ica l l y -access ib le engine. - 1 8 3 -9.4 Future Work The results f r o m the studies discussed here prov ide insight into some o f the aspects o f a direct in ject ion natural gas combust ion system. They also prov ide guidance for what aspects o f future research might be pursued to achieve further reductions i n emiss ions and improvements i n ef f ic iency . The current w o r k suggests that i m p r o v i n g the late-cycle combust ion process may achieve s igni f icant reductions i n emissions o f particulate matter, hydrocarbons and carbon m o n o x i d e . O n e o f the drawbacks o f the current combust ion system is that the p is ton b o w l shape is o p t i m i z e d for diesel combust ion . G i v e n the differences i n the c o m b u s t i o n processes between l i q u i d diesel sprays and gaseous jets, it is very l i k e l y that the current geometry is subopt imal for the combust ion o f natural gas. A s a result, substantial performance improvements m a y be achieved by m o d i f y i n g the combust ion chamber geometry. H o w e v e r , addressing this issue w i l l require careful p lanning , as the best technique w o u l d use an accurate combust ion m o d e l val idated against h igh qual i ty exper imental results. Current ly , there is no combust ion m o d e l that is adequately va l idated; nor do the exper imental techniques prov ide suff ic ient in format ion to re l iab ly val idate the details o f combust ion models . A l m o s t certainly , quant i f ied opt ical measurements o f the charge m o t i o n w i t h i n the combust ion chamber w i l l be required to prov ide m o d e l va l idat ion . In the meant ime, some ind icat ion o f poss ib le combust ion chamber improvements may be achieved through the exper imental invest igat ion o f engine performance and emissions us ing var ious combust ion chamber geometries. These results can be compared to s imulat ion results to prov ide potential techniques to opt imize the overa l l combust ion process. The use o f mul t ip le injections or in ject ion rate-shaping may offer benefits w h i c h are complementary to combust ion chamber opt imizat ion . B y avo id ing excessive gaseous-fuel p r e m i x i n g , emiss ions o f unburned hydrocarbons can be reduced. Further, by p r o v i d i n g a secondary in ject ion o f fue l , it may be possible to enhance fue l -a i r m i x i n g and thereby augment the late -cyc le combust ion process. Th i s cou ld prov ide s igni f icant reduct ions i n C O , H C , and P M emiss ions , al though there cou ld be a fue l ef f ic iency penalty. The use o f rate-shaping, where the in ject ion rate is var ied dur ing the in ject ion process by vary ing the injector needle l i f t , may also prov ide substantial benefits. A s an in i t ia l step, starting w i t h a relat ively l o w in ject ion rate pr ior to ign i t ion w o u l d reduce early o v e r m i x i n g o f the fue l and charge. -184-Increasing the injection rate late in the cycle could achieve substantial late-cycle combustion benefits. These techniques have the advantage that they could be optimized individually at every operating condition. Other parameters, such as the physical features of the injector (hole size, injection angle, or hole number, for example) are fixed and hence cannot be optimized for different operating conditions. However, optimization of the injector geometry, in conjunction with injection process and combustion chamber geometry improvements, is required to provide the best possible combustion system. The interrelatedness of a multitude of variables underlines the complexity of the optimization challenge. The fundamental requirement to assess all these variables concurrently reinforces the need to develop an efficient numerical model which can provide reasonable predictions of performance and pollutant formation to supplement experimental tests. The influence of fuel composition on the combustion process and corresponding emissions are highlighted in the current work. The effects of some o f the principal constituents in natural gas have been studied. However, further research over a wider range of operating conditions could provide information on the effect of small variations in natural gas composition. A s well , there is a requirement for further research into the effects on engine operation of non-traditional gaseous additives. For example, coal synthesis gas (composed of large quantities of C O and H2), landfill gas, or oxygenated biogases might be added to natural gas. The implications for global climate change and energy security concerns would be, in all likelihood, highly beneficial. However, negative impacts on combustion performance or overall emissions could put an unacceptably high price on any benefits. Substantially more research is required, at both fundamental and applied levels, to determine what role these fuel compositions could play as the engine evolves to meet the inseparable criteria of performance, efficiency, economic optimization, and environmental responsibility. -185-References [I] Pischinger, S. The future o f vehicle propulsion - combustion engines and alternatives. Topics in Catalysis. 30/31. 2004. Pp. 5-16. [2] Environment Canada. 2000 Criteria A i r Contaminant Emissions Summary. Available from: www.ec.gc.ca/pdb/cac/ape_tables/canada2000_e.html. Accessed 28-11-2005. [3] Seaton, A . , W . MacNee, K . Donaldson, and D . Godden. Particulate air pollution and acute health effects. Lancet, 1995, 345 pp 176-178. [4] Dockery, D . W . , C A . Popel l l , X . X u et al. A n association between air pollution and mortality in six U S cities. New England Journal of Medicine, 329, pp 1753-1759. 1993. [5] Ontario Medica l Association, The Illness Costs o f A i r Pollution. 2005-2026 Health & Economic Damage Estimates. 2005. Available from: www.oma.org/Health/smog/report/ICAP2005_Report.pdf. Accessed 27/11/2005. [6] Seinfeld, J .H . and S .N. Pandis. Atmospheric Chemistry and Physics of A i r Pollution: From A i r Pollution to Global Climate Change. Wiley-interscience, N e w York , 1998. [7] Pope, C A . I l l , R .T. Burnett, M . J . Thun et al. Lung Cancer, Cardiopulmonary Mortality, and Long-Term Exposure to Fine Particulate A i r Pollution. Journal of the American Medical Association, 287. 2002. Pp. 1132-1143. [8] Burnett, R .T . , J.Brook, T. Dann, et al. Association between Particulate and Gas-Phase Components o f Urban A i r Pollution and Dai ly Mortality in Eight Canadian Cities. Inhalation Toxicology. 12(supp. 4). 2000. Pp. 15-39 [9] Jerrett, M . , R . T . Burnett, R . M a , et al. Spatial Analysis o f A i r Pollution and Mortal i ty in Los Angeles. Epidemiology. 16(6). 2005. Pp. 727-736. [10] Gauderman, W.J . , E . A v o l , F . Gi l l i land, et al. The Effect of A i r Pollution on Lung Development from 10 to 18 Years o f Age. N e w England Journal of Medicine. 351(11). 2004. Pp. 1057-1067. [II] Hahn, F.F. , E . Barr, M . G . Menache and J .C. Seagrave. Particle Size and Composition Related to Adverse Health Effects in Aged, Sensitive Rats. Health Effects Institute Research Report 129. 2005. Available from: http://www.healtheffects.org/Pubs/Hahn.pdf. Accessed 7/12/2005. [12] Oberdorster, G . Toxicology o f ultrafine particles: in vivo studies. Philosophical Transactions o f the Royal Society of London, A . 358. 2000. Pp. 2719-2740. [13] Biswas, P. and C . Y . W u . Nanoparticles and the Environment. Journal o f the A i r and Waste Management Association. 55. 2005. Pp. 708-746. [14] Renwick. L . C . , D . Brown, A . Clouter and K . Donaldson. Increased inflammation and altered macrophage chemotactic responses caused by two ultrafine particle types. Occupational and Environmental Medicine. 61. 2002. Pp. 442-447. -186-[15] Donaldson, K., X .Y . L i and W. MacNee. Ultrafine (nanometer) particle-mediated lung injury. Journal of Aerosol Science, 29. 1998. Pp. 553-560. [16] Nemmar, A., P.H.M. Hoet, B. Vanquickenborne et al. Passage of inhaled paticles into the blood circulation in humans. Circulation 105(4), 2002. Pp. 411-414. [17] Delfino, R.J., C. Sioutas and S. Malik. Potential Role of Ultrafine Particles in Associations between Airborne Particle Mass and Cardiovascular Health. Environmental Health Perspectives, 113(8). 2005. Pp. 934-946. [18] Kittelson, D.B. Engines and Nanoparticles: A Review. Journal of Aerosol Science, 29(5/6). pp. 575-588, 1998. [19] Kagawa, J. Health effects of diesel exhaust emissions - a mixture of air pollutants of worldwide concern. Toxicology. 181-182. 2002. Pp. 349-353. [20] California Air Resources Board. A R B Almanac 2003—Chapter 5: Toxic Air Contaminant Emissions, Air Quality, and Health Risk. Available from: www.arb.ca.gov/aqd/almanac/almanac03/pdf/chap503.pdf. Accessed 30/11/2005. [21] Marr, L . C . and R.A. Harley. Spectral analysis of weekday-weekend differences in ambient ozone, nitrogen oxide, and non-methane hydrocarbon time series in California. Atmospheric Environment. 36. 2002. Pp. 2327-2335. [22] Fujita, E . M . , W.R. Stockwell, D.E. Campbell, R.E. Keislar and D.R. Lawson. Evolution of the Magnitude and Spatial Extent of the Weekend Ozone Effect in California's South Coast Air Basin, 1981-2000. Journal of the Air and Waste Management Association. 53(7). 2003. Pp. 802-815. [23] Lawson, D.R. The Weekend Ozone Effect - The Weekly Ambient Emissions Control Experiment. EM(AWMA). July 2003. Pp. 17-26. [24] Intergovernmental Panel on Climate Change (IPCC). Third Assessment Report, 2001. Available from: http://www.ipcc.ch/pub/online.htm. Accessed 07/04/06. [25] Code of Federal Regulations (CFR) 40 Parts 69, 80, and 86: Control of Air Pollution from New Motor Vehicles: Heavy Duty Engine and Vehicle Standards and Highway Diesel Fuel Sulfur Control Requirements; Final Rule. US Federal Register. 66(12). Pp. 5002-5050. 2001 [26] On Road Vehicle and Engine Emissions Regulations. Canada Gazette Part II. 137(1). 2003. Pp. 21-27. [27] Diesel Emissions Standards as published on www.dieselnet.org . Available from: www.dieselnet.org/standards. Accessed 23/01/2006. [28] Heywood, J.B. Internal Combustion Engine Fundamental. McGraw-Hill, New York. 1988. [29] Johnson, T.V. Diesel Emission Control Technology - 2003 in Review. SAE Technical Paper 2004-01-0070. 2004. [30] Poulopoulos, S.G., D.P. Samara and C.J. Philippopoulos. Regulated and speciated hydrocarbon emissions from a catalyst equipped internal combustion engine. Atmospheric Environment (35). 2001. Pp. 4443-4450. -187-[31] Ayala, A., M.E . Gebel, R.A. Okamoto et al. Oxidation Catalyst Effect on C N G Transit Bus Emissions. SAE Technical Paper 2003-01-1900. 2003. [32] Andersson, B., N . Cruise, M . Lunden et al. Methane and Nitric Oxide Conversion Over a Catalyst Dedicated for Natural Gas Vehicles. SAE Technical Paper 2000-01-2928. 2000. [33] Hirasawa, Y. , Y. Tanaka, Y. Banno and M . Nagata. Development of Methane Oxidation Catalyst and Its Mechanism. SAE Technical Paper 2005-01-1098. 2005. [34] Chatterjee, S., C. McDonald, R. Conway, et al. Emission Reductions and Operational Experiences with Heavy Duty Diesel Fleet Vehicles Retrofitted with Continuously Regenerated Diesel Particulate Filters in Southern California. SAE Technical Paper 2001-01-0512 [35] Anon. Global Activity Blossoms as World Demand Climbs. World Oil. 226(9). 2005. [36] US Energy Information Agency (EIA). US Natural Gas Imports. Available from: http://tonto.eia.doe.gov/dnav/ng/ng_move_impc_sl_a.htm. Accessed 13/07/2006. [37] Warnatz, J., U. Maas, and R.W. Dibble. Combustion. Physical and Chemical Fundamentals, Modeling and Simulation, Experiments, Pollutant Formation. 2n d Ed. Springer-Verlag, Berlin. 1999. [38] McTaggart-Cowan, G.P., S.N. Rogak, P.G. Hill, W.K. Bushe and S.R. Munshi. Effects of Operating Condition on Particulate Matter and Nitrogen Oxides Emissions from a Heavy-Duty Direct Injection Natural Gas Engine using Cooled Exhaust Gas Recirculation. International Journal of Engine Research. 5(6). 2004. Pp. 499-511. [39] Hill, P.G. and G.P. McTaggart-Cowan. Nitrogen Oxide Production in a Diesel Engine Fueled by Natural Gas. SAE Technical Paper 2005-01-1727. SAE Transactions, Journal of Engines. 114(3). 2005. [40] Goudie, D., M.Dunn, S.R. Munshi, E . Lyford-Pike, J. Wright, V . Duggal and M . Frailey. Development of a Compression Ignition Heavy Duty Pilot-Ignited Natural Gas Fuelled Engine for Low NOx Emissions. SAE Technical Paper 2005-01-2954. 2005. [41] Attar, A .A . and G. A. Karim. Knock Rating of Gaseous Fuels. A S M E Journal of Engineering for Gas Turbines and Power. 125, April 2003. Pp. 500-504. [42] Kabori,S., T. Kamimoto and A.A. Aradi. A Study of Ignition Delay of Diesel Fuel Sprays. International Journal of Engine Research. 1(1). 2000. Pp. 29-39. [43] Huang, J., P.G. Hill, W.K. Bushe and S.R. Munshi. Shock-Tube Study of Methane Ignition Under Engine-Relevant Conditions: Experiments and Modelling. Combustion and Flame. 136(1-2). 2004. Pp. 25-42. [44] Dale, J.D., M.D. Checkel and P.R. Smy. Application of High Energy Ignition Systems to Engines. Progress in Energy and Combustion Science. 23(5-6). 1997. Pp. 379-398. [45] Reynolds, C.C.O.B. and R.L. Evans. Improving emissions and performance characteristics of lean burn natural gas engines through partial stratification. International Journal of Engine Research. 5(1), 2004. Pp. 105-114. - 1 8 8 -[46] Chen, S.K. and N.J. Beck. Gas Engine Combustion Principles and Applications. SAE Technical Paper 2001-01-2489. 2001. [47] John Deere Power Systems, Specification Sheet for 6081H Natural Gas Engine. Available from: http://jdpsesc.deere.com/ESComp/servlet/com.deere.enginedb.servlet.model2.Engine SearchResuitsServlet. Accessed 19-01-2006. [48] M A N Engines and Components. Available from: http://www.man-engines.com/en/Productsand Solutions/Vehicle_engines/Gas_engines.jsp. Accessed 20-01-2006. [49] Cummins-Westport Inc. Specification sheets for B-gas, C-gas, and L-gas engines. Available from: http://www.cumminswestport.com/products/index.php. Accessed 19-01-2006. [50] Chandler, K . L , P. Norton and N. Clark. Interim Results from Alternative Fuel Truck Evaluation Project. SAE Technical Paper 1999-01-1505. 1999. [51] Ayala, A., N.Y Kado, R. A. Okamoto et al. Diesel and CNG Heavy-Duty Transit Bus Emissions Over Multiple Driving Schedules : Regulated Pollutants and Project Overview. SAE Technical Paper 2002-01-1722. [52] Frailey, M . , P. Norton, N.N. Clark and D. W. Lyons. An Evaluation of Natural Gas versus Diesel in Medium Duty Buses. SAE Technical Paper 2000-01-2822. [53] McCormick, R.L., M.S. Graboski, T. Alleman et a/.In-Use Emissions from Natural Gas Fueled Heavy-Duty Vehicles. SAE Technical paper 1999-01-1507. [54] Kado, N.Y, R. A. Okamoto, P.A. Kuzmicky et al. Emissions of Toxic Pollutants from Compressed Natural Gas and Low Sulfur Diesel-Fueled Heavy-Duty Transit Buses Tested over Multiple Driving Cycles. Environmental Science and Technology. 39. 2005. Pp. 7638-7649. [55] Seagrave, J . C , A. Gigiliotti, J.D. McDonald, et al. Composition, Toxicity, and Mutagenicity of Particulate and Semivolatile Emissions from Heavy-Duty Compressed Natural Gas-Powered Vehicles. Toxicological Sciences. 87(1). 2005. Pp. 232-241. [56] Holmen, B.A. and A. Ayala. Ultrafine PM Emissions from Natural Gas, Oxidation-Catalyst Diesel and Particle-Trap Diesel Heavy-Duty Transit Buses. Environmental Science and Technology. 36. 2002. Pp. 5041-5050. [57] Kamel, M . , E . Lyford-Pike, M . Frailey et al. An Emissions and Performance Comparison of the Natural Gas Cummins Westport Inc. C-Gas Plus Versus Diesel in Heavy-Duty Trucks. S A E Technical Paper 2002-01-2737. [58] Melendez, M . , J. Taylor, J. Zuboy, W.S. Wayne and D. Smith. Emission Testing of Washington Metropolitan Area Transit Authority (WMATA) Natural Gas and Diesel Transit Buses. N R E L Report NREL/TP-540-36355. December 2005. Available from: http://www.nrel.gov/docs/fy06osti/36355.pdf. Accessed 26-04-2006 [59] Chiu, J.P., J. Wegrzyn and K . E . Murphy. Low Emissions Class 8 Heavy-Duty On-Highway Natural Gas and Gasoline Engine. SAE Technical Paper 2004-01-2982. -189-[60] Nellen, C. and K. Boulouchos. Natural Gas Engines for Cogeneration: Highest Efficiency and Near-Zero Emissions through Turbocharging, EGR and 3-Way Catalytic Converter. SAE Technical Paper 2000-01-2825. [61] Park, T., R.J. Atkinson, N.N. Clark et al. Operation of a Compression Ignition Engine with a HEUI Injection System on Natural Gas with Diesel Pilot Injection. SAE Technical Paper 1999-01-3522. 1999. [62] Wartsila. Wartsila 32DF Technology Review. Available at: http://vvww.wartsila.com/Wartsila/docs/en/ship_power/media_publications/brochures/product /engines/w32df_tr.pdf. Accessed 20-01-2006. [63] Fairbanks-Morse. Available from: http://www.fairbanksmorse.com/engines/index.htm. Accessed 20-01-2006. [64] M A N engines. Available from: http://www.manbw.com/products. Accessed 20-01-2006. [65] Singh, S., S.R. Krishnan, K.K. Srinivasan, K . C . Midkiff and S.R. Bell. Effect of pilot injection timing, pilot quantity and intake charge conditions on performance and emissions for an advanced low-pilot-ignited natural gas engine. International Journal of Engine Research, 5(4). 2004. Pp. 329-348. [66] Zuo, C. and M . Yang. Operating Characteristics and Description of a Dual Fuel Engine for Diesel-Natural Gas Heavy-Duty Operation. SAE Technical Paper 1999-01-3523. 1999. [67] Shenghua, L. , W. Ziyan and R. Jiang. Development of compressed natural gas/diesel dual-fuel turbocharged compression ignition engine. Proceedings of the Institution of Mechanical Engineers, Part D - Journal of Automobile Engineering. 217. 2003. Pp. 839-845. [68] Pappagianakis, R.G., and D.T. Hountalas. Combustion and exhaust emissions characteristics of a dual fuel compression ignition engine operated with pilot Diesel fuel and natural gas. Energy Conversion and Management. 45. 2004. Pp. 2971-2987. [69] Wong, H.C. Next Generation Natural Gas Vehicle Program Phase I: Clean Air Power 0.5 g/hp-h NOx Engine Report - Final Report. N R E L Report, NREL/SR-540-33957. July, 2003. [70] Hiltner, J., S. Fivelan, R. Agama and M . Willi. System Efficiency Issues for Natural Gas Fueled HCCI Engines in Heavy-Duty Stationary Applications. SAE Technical Paper 2002-01-0417.2002. [71] Stanglmaier, R.H., T.W. Ryan and J.S. Souder. HCCI Operation of a Dual-Fuel Natural Gas Engine for Improved Fuel Efficiency and Ultra-Low NOx Emissions at Low to Moderate Engine Loads. SAE Technical Paper 2001-01-1897, 2001. [72] Tominaga, R., S. Morimoto, Y. Kawabata et al. Effects of Heterogeneous EGR on the Natural Gas Fueled HCCI Engine Using Experiments, CFD and Detailed Kinetics. SAE Technical Paper 2004-01 -0945. 2004. [73] Huang, Z., S. Shiga, T. Ueda et al. Combustion Characteristics of Natural-Gas Direct Injection Combustion under Various Fuel Injection Timings. Proceedings of the Institution of Mechanical Engineers, Part D. Journal of Automobile Engineering. 217. 2003. Pp. 393-401. -190-[74] F. Zhao, M.-C. Lai and D.L. Harrington, Automotive spark-ignited direct-injection gasoline engines. Progress in Energy and Combustion Sciences. 25. 1999. Pp. 437-562. [75] Aesoy, V . and H. Valland. Hot Surface Assisted Compression Ignition of Natural Gas in a Direct Injection Diesel Engine. SAE Technical Paper 960767, 1996. [76] Caterpillar Inc. Development of a Direct-Injected Natural Gas Engine for Heavy Duty Vehicles. Phase 1 Report. N R E L Report NREL/SR-540-27500. 2000. [77] Rubas, P.J., M.A. Paul, G.C. Martin, R. Coverdill, R. Lucht, J. Peters, K. DelVecchio. Methane Jet Penetration in a Direct-Injection Natural Gas Engine. SAE Technical Paper 980143, 1998. [78] Okada, M . , H. Sugii, T. Wakao et al. Development of CNG Direct Injection Diesel-Cycle Engine. Proceeding of N G V 2004, 2004. [79] Chandler, K. and K. Proc. Norcal Prototype L N G Truck Fleet: Final Results. U.S. Department of Energy Advanced Technology Vehicle Evaluation Report. July, 2004. Available from: http://www.eere.energy.gov/afdc/pdfs/36707.pdf Accessed 26-04-2006. [80] Garner, G., K. Komuniecki, K. Wadley and B. Zehr. Clean Air Corridor Mid-Trial Report. Westport Innovations Inc. March 10, 2006. Available from: http://www.westport.com/programs/norcal.php. Accessed 26-04-2005. [81] California Air Resources Board. Executive Order A-343-0003. Available from: http://www.arb.ca.gov/msprog/onroad/cert/mdehdehdv/2006 /westport_hhdd_a3430003_l4d9_ld2-0d02_cngplusdiesel.pdf. Accessed 26-04-2006. [82] Williams, G., K. Tanin, R. Ancimer et al. High Performance HCCI Natural Gas Technology for Distributed Generation. Proceedings of the 2005 Spring Technical Conference of the A S M E Electric Power Conference. Paper # ICES05-55859. 2005. [83] Turner, J.S. The 'Starting Plume' in Neutral Surroundings. Journal of Fluid Mechanics, 1962. 13(3). Pp. 356-371. [84] Hill, P.G. and P. Ouellette. Transient Turbulent Gaseous Fuel Jets for Diesel Engines. A S M E Journal of Fluids Engineering. 121. 1999. Pp. 93-101. [85] Cossali, G.E., A. Coghe and L. Araneo. Near-Field Entrainment in an Impulsively Started Turbulent Gas Jet. Journal of the American Institute of Aeronautics and Astronautics. 39(6). 2001. Pp. 113-1121. [86] Petersen, B.R. and J.B. Ghandi. Transient High-Pressure Hydrogen Jet Measurements. SAE Technical Paper 2006-01-0652. 2006. [87] L i , G. , P. Ouellette, S. Dumitrescu, and P.G. Hill. Optimization Study of Pilot-Ignited Natural Gas Direct-Injection in Diesel Engines. SAE Technical Paper 1999-01-3556. SAE Transactions, Journal of Fuels and Lubricants. 108(4). 1999. [88] Dumitrescu, S., P.G. Hill, G. Li and P. Ouellette. Effects of Injection Changes on Efficiency and Emissions of a Diesel Engine Fueled by Direct Injection of Natural Gas. SAE Technical Paper 2000-01-1805.2000. -191-[89] Sullivan, G.D., J. Huang, T.X. Wang, W. K. Bushe and S.N. Rogak. Emissions Variability in Gaseous Fuel Direct Injection Compression Ignition Combustion. S A E Technical Paper 2005-01-0917, 2005. [90] Muniz, L . and M.G. Mungal. Instantaneous Flame-Stabilization Velocities in Lifted-Jet Diffusion Flames. Combustion and Flame. 111(1-2). 1997. Pp. 16-30. [91] Nishioka, M . , Y. Kondoh and T. Takeno. Behaviour of Key Reactions on NO Formation in Methane-Air Flames. 26th International Symposium on Combustion. The Combustion Institute. 1996. Pp. 2139-2145. [92] Easley, W.L. and A . M . Mellor. NO Decomposition in Diesel Engines. SAE Technical Paper 1999-01-3546. 1999. [93] Ahmad, T. and S.L. Plee. Application of Flame Temperature Correlations to Emissions from a Direct-Injection Diesel Engine. SAE Technical Paper 831374. 1983. [94] Tsurushima, T., L . Zhang and Y. Ishii. A Study of Unburnt Hydrocarbon Emission in small DI diesel engines. SAE Technical Paper 1999-01-0512. 1999. [95] Shi, J.P. and R.M. Harrison. Investigation of Ultrafine Particle Formation during Diesel Exhaust Dilution. Environmental Science and Technology. 33(21). 1999. Pp. 3730-3736. [96] Baumgard, K. and J.H. Johnson. The effect of fuel and engine design on diesel exhaust particle size distributions. SAE Technical Paper 960131. 1996. [97] Lev-On, M . , C. LeTavec, J.Uihlein, T.L. Alleman, D.R. Lawson, K. Vertin, G.J. Thompson, M . Gautam, S. Wayne, B. Zielinska, J. Sagebiel. Chemical Speciation of Exhaust Emissions from Trucks and Buses Fueled on Ultra-Low Sulfur Diesel and CNG. SAE Technical Paper 2002-01-0432. 2002. [98] Lloyd, A .C. and T.A. Cackette. Diesel Engines: Environmental Impact and Control. Journal of the Air and Waste Management Association 51(6). 2001. Pp. 809-847. [99] Khalek, I. A., D.B. Kittelson and F. Brear. Nanoparticle Growth During Dilution and Cooling of Diesel Exhaust: Experimental Investigation and Theoretical Assessment. SAE Technical Paper 2000-01-0515.2000. [100] Frenklach, M . Reaction Mechanism of Soot Formation in Flames. Physical Chemistry and Chemical Physics. 4. 2002. Pp. 2028-2037. [101] Tao, F., V.I. Golovitchev, J. Chomiak. A Phenomological Model for the Prediction of Soot Formation in Diesel Spray Combustion. Combustion and Flame. 136. 2004. Pp. 270-282. [102] Flynn, P.F., R.P. Durrett, G.L. Hunter, A.O.Z. Loye, O.C. Akinyemi, J.E. Dec and C.K. Westbrook. Diesel Combustion: An Integrated View Combining Laser Diagnostics, Chemical Kinetics, and Empirical Validation. SAE Technical Paper 1999-01-0509. 1999. [103] Krestinin, A .V. Detailed Modeling of Soot Formation in Hydrocarbon Pyrolysis. Combustion and Flame. 121. 2000. pp. 513-524. [104] Higgins, K.J. , H. Jung, D.B. Kittelson, J.T. Roberts and M.R. Zachariah. Kinetics of Diesel Nanoparticle Oxidation. Environmental Science and Technology. 37. 2003. Pp. 1949-1954. - 1 9 2 -[105] Song, H. , N . Ladommatos and H. Zhao. Diesel Soot Oxidation under Controlled Conditions. SAE Technical Paper 2001-01-3673.2001. [106 ] Roesler, J.F., S. Martinot, C S . McEnally, L.D. Pfefferle, J.L. Delfau and C. Vovelle. Investigating the Role of Methane on the Growth of Aromatic Hydrocarbons and Soot in Fundamental Combustion Processes. Combustion and Flame. 134. 2003. Pp. 249-260. [107] McCrain, L . L . and W:L. Roberts. Measurements of the soot volume field in laminar diffusion flames at elevated pressure. Combustion and Flame. 140. 2005. Pp. 60-69. [108] Thomson, K.A. , O.L. Gulder, E J . Weckman, R.A. Fraser, G.J. Smallwood and D.R. Snelling. Soot Concentration and Temperature Measurements in Co-Annular, Non-Premixed C H i / A i r Laminar Flames at Pressures up to 4 MPa. Combustion and Flame. 140. 2005. Pp. 222-232. [109] Brookes, S.J. and J.B. Moss. Measurements of Soot Production and Thermal Radiation From Confined Turbulent Jet Diffusion Flames of Methane. Combustion and Flame. 116. 1999. Pp. 49-61. [110] Bohm, H. , K. Kohse-Hoinghaus, F. Lacas, C. Rolon, N. Darabiha and S. Candel. On PAH Formation in Strained Counterflow Diffusion Flames. Combustion and Flame. 124. 2001. Pp. 127-136. [ I l l ] Skjoth-Rasmussen, M.S., P. Glarborg, M . Ostberg, J.T. Johannessen, H. Libbjerg, A.D. Jensen and T.S. Christense. Formation of Polycyclic Aromatic Hydrocarbons and Soot in Fuel-Rich Oxidation of Methane in a Laminar Flow Reactor. Combustion and Flame. 136. 2004. Pp. 91-128. [112] Gruenberger, T . M . , M . Moghiman, P.J. Bowen and N. Syred. Dynamics of Soot Formation by Turbulent Combustion and Thermal Decomposition of Natural Gas. Combustion Science and Technology. 174. 2002. Pp. 67-86. [113] Smooke, M.D., C S . McEnally, L.D. Pfefferle, R.J. hall and M.B. Colket. Computational and Experimental Study of Soot Formation in Coflow, Laminar Diffusion Flame. Combustion and Flame. 117. 1999. Pp. 117-139. [114] Zhu, X . L . and J.P. Gore. Radiation Effects on Combustion and Pollutant Emissions of High-Pressure Opposed Flow Methane/Air Diffusion Flames. Combustion and Flame. 141. 2005. Pp. 118-130. [115] Ning, Z., C S . Cheung, S.X. Liu. Experimental Investigation of the Effect of Exhaust Gas Cooling on Diesel Particulate. Journal of Aerosol Science. 35. 2004. Pp. 333-345. [116] Luders, H. , M . Kruger, P. Stommel, B. Luers. The Role of Sampling Conditions in Particle Size Distribution Measurements. SAE Technical Paper 981374. 1998. [117] Mathis, U., J. Ristimaki, M . Mohr, J. Keskinen, L. Ntziachristos, Z. Samaras and P. Mikkanen. Sampling Conditions for the Measurement of Nucleation Mode Particles in the Exhaust of Diesel Engines. Aerosol Science and Technology. 38. 2004. Pp. 1149-1160. [118] Ingersoll, J.G. Natural Gas Vehicles. Fairmont Press, Upper Saddle River, New Jersey. 1996. p. 44. -193-[119] Yoon, S.S., S.M. Lee and S.H. Chung. Effect of Mixing Methane, Ethane, Propane, and Propene on PAH and Soot Formation in Ethylene Base Counterflow Diffusion Flames. Proceedings of the Combustion Institute. 30. 2005. Pp. 1417-1424. [120] Kuo, K. Principles of Combustion. 2nd Ed. Wiley, New York. 2005. [121] Huang, J. and W.K. Bushe. Experimental and Kinetic Study of Autoignition in Methane/Ethane/Air and Methane/Propane/Air mixtures under Engine-Relevant Conditions. Combustion and Flame. 144. 2006. Pp. 74-88 [122] McTaggart-Cowan, G.P., W.K. Bushe, P.G. Hill and S.R. Munshi. NOx Reduction from a Heavy-Duty Diesel Engine with Direct Injection of Natural gas and Cooled Exhaust Gas Recirculation. International Journal of Engine Research, 5(2). 2004. Pp. 175-191. [123] Ladommatos, N. , S. Abdelhalim, H. Zhao. The Effects of Exhaust Gas Recirculation on Diesel Combustion and Emissions. International Journal of Engine Research. 1(1). 2000. Pp. 107-126. [124] Kreso, A . M . , J.H Johnson, L.D. Gratz, S.T. Bagley and D.G. Leddy. A Study of the Effects of Exhaust Gas Recirculation on Heavy-Duty Diesel Engine Emissions. SAE Technical Paper 981422. 1998. [125] Plee, S.L., T. Ahmad, J.P. Myers and G.M. Faeth. Diesel N O x Emissions - A Simple Correlation Technique for Intake Air Effect. 19th International Sypmosium on Combustion. The Combustion Institute, Pittsburgh, Pa. 1982. Pp. 1495-1502. [126] Flynn, P.F., G.L. Hunter, R.P. Durrett, L .A. Farrell and W.C. Akinyemi. Minimum engine flame temperature impacts on diesel and spark-ignition engine NOx production. SAE Technical Paper 2000-01-1177. 2000. [127] Idicheria, C A . and L . M Pickett. Soot Formation in Diesel Combustion under High-EGR Conditions. SAE Technical Paper 2005-01-3834. 2005. [128] Alriksson, M . , T. Rente and I. Denbratt. Low Soot, Low NOx in a Heavy Duty Diesel Engine Using High Levels of EGR. SAE Technical paper 2005-01-3836. 2005. [129] Mastorakos, E. , A . M . Taylor and J.H. Whitelaw. Extinction of Turbulent Counterflow Flames with Reactants Diluted by Hot Products. Combustion and Flame. 102. 1995. Pp. 101-114. [130] McTaggart-Cowan, G.P., W.K. Bushe, S.N. Rogak, P.G. Hill and S.R. Munshi. PM and N O x Reduction by Injection Parameter Alterations in a Direct Injected, Pilot Ignited, Heavy Duty Natural Gas Engine with EGR at Various Operating Conditions. SAE Technical Paper 2005-01-1733. SAE Transactions, Journal of Engines. 114(3). 2005. [131] McTaggart-Cowan, G.P., W.K. Bushe, S.N. Rogak, P.G. Hill and S.R. Munshi. The Effects of Varying EGR Test Conditions on a Direct Injection of Natural Gas Heavy-Duty Engine with High EGR Levels. SAE Technical Paper 2004-01-2955. SAE Transactions, Journal of Engines. 113(3). 2004. Pp. 1500-1509. [132] Ouellette, P., P. Mtui and P.G. Hill. Numerical Simulations of Directly Injected Natural Gas and Pilot Diesel Fuel in a Two-Stroke Compression Ignition Engine. SAE Technical Paper 981400.1998. - 1 9 4 -[133] L i , G., T. Lennox, D. Goudie and M . Dunn. Modeling HPDI Natural Gas Heavy Duty Engine Combustion. Proceedings of the A S M E Internal Combustion Engine Division Fall Technical Conference. A S M E Technical Paper ICEF2005-1307. 2005. [134] Larson, C. Master's Thesis. University of British Columbia. 2002. [135] McTaggart-Cowan, G.P., W.K. Bushe, S.N. Rogak, P.G. Hill and S.R. Munshi. Injection Parameter Effects on a Direct Injected, Pilot Ignited, Heavy Duty Natural Gas Engine with EGR. SAE Technical Paper 2003-01-3089. SAE Transactions, Journal of Fuels and Lubricants. 112(4). 2003. Pp. 2103-2109. [136] Yu, R.C. and S.M. Shahed. Effects of Injection Timing and Exhaust Gas Recirculation on Emissions from a D.I. Diesel Engine. SAE 811234. 1981. [137] Harrington, J., S. Munshi, C. Nedelcu, P. Ouellette, J. Thomson and S. Whitfield. Direct Injection of Natural Gas in a Heavy-Duty Diesel Engine. SAE Technical Paper 2002-01-1630.2002. [138] Shen, J., J. Qin and M . Yao. Turbocharged CNG/Diesel Dual Fuel Engines with Intercooler: Combustion, Emissions, and Performance. SAE Technical Paper 2003-01-3082. 2003. [139] Zhang, L . A Study of Pilot Injection in a DI Diesel Engine. SAE Technical Paper 1999-01-3493. 1999. [140] Badami, M . , F. Millo and D.D. D'Amato. Experimental Investigation on Soot and N O x Formation in a DI Common-Rail Diesel Engine with Pilot Injection. SAE Technical Paper 2001-01-0657.2001. [141] Trueba, A., B. Barbeau, O. Pajot and K. Mokaddem. Pilot injection timing effect on the main injection development and combustion in a DI diesel engine. SAE Technical Paper 2002-01-0501.2002. [142] Jones, H.L. , G.P. McTaggart-Cowan, S.N. Rogak, W.K. Bushe, S.R. Munshi and B.A. Buchholz. Source Apportionment of Particulate Matter from a Diesel Pilot-Ignited Natural Gas Fuelled Heavy Duty DI Engine. SAE Technical paper 2005-01-2149. SAE Transactions, Journal of Fuels and Lubricants. 114(4). 2005. [143] McTaggart-Cowan, G.P., W.K. Bushe, P.G. Hill and S.R. Munshi. A Supercharged Single-Cylinder Heavy-Duty Engine for High Pressure Direct Injection of Natural G