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Performance, emissions and combustion characteristics of natural gas fueling of diesel engines Douville, Brad 1994

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PERFORMANCE, EMISSIONS AND COMBUSTION CHARACTERISTICS OFNATURAL GAS FUELING OF DIESEL ENGINESbyBRAD DOUVILLEB.Sc. in Mechanical Engineering, University of Alberta, 1992A THESIS SUBMITI’ED IN PARTIAL FULFILLMENT OFTHE REQUIREMENTS FOR THE DEGREE OFMASTER OF APPLIED SCIENCEINTHE FACULTY OF GRADUATE STUDIESMechanical Engineering DepartmentWe accept this thesis as conformingto the required standardTHE UNIVERSITY OF BRITISH COLUMBIAApril 1994© Brad Douville 1994In presenting this thesis in partial fulfilment of the requirements for an advanceddegree at the University of British Columbia, I agree that the Library shall make itfreely available for reference and study. I further agree that permission for extensivecopying of this thesis for scholarly purposes may be granted by the head of mydepartment or by his or her representatives. It is understood that copying orpublication of this thesis for financial gain shall not be allowed without my writtenpermission.(Signature)____________________Department ofThe University of British ColumbiaVancouver, CanadaDate çDE.6 (2/88)AB ST RACTThe performance, emissions and combustion characteristics of natural gas fueling of dieselengines have been investigated. The natural gas fueling system employs electronically-controlledlate-cycle direct injection of high pressure natural gas with a small amount of diesel fuel (dieselpilot). Since the temperature in the combustion chamber at the end of compression is below theautoignition temperature of natural gas, the diesel pilot is required for ignition. Diesel engineperformance and emissions have been measured using both natural-gas and conventional dieselfueling, over a wide range of operating conditions. The combustion process in diesel engines hasbeen modeled based on measured cylinder pressure to learn about the formation rates of oxides ofnitrogen and fuel burning rates. This combustion model has been developed to deal specificallywith the non-uniformities and pollutant formation associated with stratified-charge combustion indiesel engines. The test results demonstrate that the thermal efficiencies for both natural gas andconventional diesel fueling at low and medium engine loads, are almost identical. The thermalefficiencies at high loads for natural gas fueling are greater than for conventional diesel fueling.Upon optimization of the natural gas and diesel pilot injection, lower pollutant exhaust emissionswill be produced over the entire engine load range, while obtaining higher peak engine loadcapabilities than with conventional diesel fueling.IiTABLE OF CONTENTSABSTRACT iiLIST OF TABLES vuLIST OF FIGURES viiiACKNOWLEDGMENT xi1 INTRODUCTION1.1 Vehicle Exhaust Emissions and Alternative Fuels 11.2 Natural Gas Fueling of Diesel Engines 41.3 Intensifier-Injector for Natural Gas Fueling of Diesel Engines 91.4 Objectives of This Research 102 ENGINE TESTING APPARATUS2.1 Introduction 122.2 Engine and Engine Control 132.3 Dynamometer and Dynamometer Control 142.4 Fuel Injection System 152.5 Engine Instrumentation 172.5.1 Output Torque 182.5.2 Engine Speed 182.5.3 Diesel Fuel Flow 182.5.4 Natural Gas and Air Flows 202.5.5 Cylinder Pressure 21hi2.5.6 Mean Engine Temperatures and Pressures 232.5.7 Exhaust Gas Analyzers 242.6 Data Acquisition System 263 DATA ACQUISITION AND CALCULATED PERFORMANCE PARAMETERS3.1 Thermal Efficiency 293.2 Mean Effective Pressure 313.3 Cyclic Variation 323.4 Wet-Basis and Brake Specific Emissions 333.5 Unburned Mass Fraction of Fuel 364 PREVIOUS COMBUSTION ANALYSIS MODELS4.1 The Stratified Charge Combustion Process 374.2 Thermodynamic Combustion Analysis 384.3 Estimating Instantaneous Heat-Transfer to Cylinder Walls 404.4 Simple-HeaLing Combustion Analysis 424.5 Chemical Reaction Combustion Analysis 454.6 Limitations of Pervious Combustion Analysis Models 495 PROPOSED THERMODYNAMIC COMBUSTION ANALYSIS MODEL5.1 Formulation of Combustion Model 535.2 Calculation of Mass Fraction of Fuel Burned 585.3 Combustion Stoichiometry 615.4 Trapped Air and Residuals 635.4.1 Scavenging Process 645.4.2 Mass of Trapped Air 655.4.3 Mass of Trapped Residuals 665.4.4 Residual Gas Composition 68iv5.4.5 Residual Gas Temperature 685.5 Thermodynamic Properties of the Unburned Gas Zone 705.6 Thermodynamic Properties of the Diesel Fuel and Natural Gas Zone 725.7 Thermodynamic Properties of the Burned Gas Zones 765.8 Thermodynamic Properties of the Mixed Gas Zone 795.9 Calculation of NOx Formation 815.10 Calculation Procedure 846 PERFORMANCE, EMISSIONS AND COMBUSTION CHARACTERISTICS6.1 Discussion of Combustion Analysis Results 876.1.1 Effect of Computation Method 886.1.2 Effect of Unburned Fuel in the Combustion Chamber 926.1.3 Computed Temperatures 926.2 Variables that Effect Diesel Engine Performance, Emissions, and Combustion 936.3 Effect of Fuel Injection Timing 956.4 Effect of Engine Load and Fuel Injection Rate 1067 CONCLUSIONS AND RECOMMENDATIONS7.1 Conclusions 1207.2 Recommendations 1228 REFERENCES 123VAPPENDICESAppendix I - Hydraulic Wheatstone Bridge Diesel Fuel Measuring Device 126Appendix II - Pressure Transducer Mounting ftr DDC 1-71 128Appendix III - Pressure Transducer Mounhing for DDC 6V-92TA 129Appendix IV - Software for Processing Cylinder Pressure Data 130Appendix V - Brake Power Correction (SAE Standard J1349 JUN85) 132Appendix VI - Calculation of Specific Humidity 134Appendix VII - Calculation of Engine Cylinder Volume 136Appendix VIII - Combustion Analysis Program XMF.BAS (QuickBASIC) 139viLIST OF TABLESTable 1-1 Urban Bus Heavy-Duty Engine Emission Standards 1Table 2-1 General Engine Specifications 13Table 2-2 Measured Engine Parameters 27Table 4-1 Szekely and Alkidas Combustion Analysis Model 48Table 5-1 Rate Constants for the Extended Zeldovich Mechanism 82Table 6-1 Local Equivalence Ratio as a Function of Injection Timingfor Diesel Fueling 87Table 6-2 Normalizing Constants for Figures 6-7 through 6-10 96Table 6-3 Normalizing Constants for Figures 6-11 and 6-12 99Table 6-4 Measured and Estimated NO (DDC 1-7 1) 119viiLIST OF FIGURESFigure 1-1 Effect of Fuel Injection Timing (Normalized) 2Figure 1-2 Methods of Using Natural Gas in Diesel Engines 5Figure 2-1 Schematic of Engine Test Set-up and Instrumentation 12Figure 2-2 Schematic of Engine Fuel Supply System 16Figure 2-3 Natural Gas Electronic Unit Injector 17Figure 2-4 Diesel Mass Flow Rate Measuring Device Schematic(DDC 6V-92TA) 19Figure 2-5 Effects of Charge Amplifier Time Constant on CylinderPressure Data 22Figure 2-6 Emissions Analysis System Schematic (heated instruments) 25Figure 2-7 Emissions Analysis System Schematic (cool instruments) 26Figure 3-1 Typical P-V diagram for a Two-Stroke Diesel Engine 32Figure 4-1 Schematic of Combustion Chamber Thermodynamic System 39Figure 5-1 Co11LpLual Diagram of Proposed Combustion Model 53Figure 5-2 Engine Mass Flow Schematic 65Figure 5-3 Scavenging Correlation 66Figure 5-4 Initial NO Formation Rate as a Function of Temperature 84Figure 5-5 Combustion Analysis Calculation Procedure 86Figure 6-1 Effect of Computation Method on Mass Fraction of Burned Fuel 89Figure 6-2 Effect of Computation Method on Heat Transfer Estimates 90Figure 6-3 Effect of Computation Method on Initial Burned Gas Temperature 90Figure 6-4 Effect of Computation Method on Estimated Exhaust NO 91Figure 6-5 Computed Temperatures 92VII’Figure 6-6 Engine Performance Map (DDC 6V-92TA) 94Figure 6-7 Effect of Fuel Injection Timing at 3 bar bmep, 1800 rpmDDC 6V-92TA- Diesel Fueling (normalized) 97Figure 6-8 Effect of Fuel Injection Timing at 3 bar bmep, 1200 rpmDDC 6V-92TA - Diesel Fueling (normalized) 97Figure 6-9 Effect of Fuel Injection Timing at 3 bar bmep, 1250 rpmDDC 1-71 - Diesel Fueling (normalized) 98Figure 6-10 Effect of Fuel Injection Timing at 1 bar bmep, 1250 rpmDDC 1-71 - Diesel Fueling (normalized) 98Figure 6-11 Effect of Fuel Injection Timing at 1 bar bmep, 1250 rpmDDC 1-71 - Natural Gas Fueling (normalized) 100Figure 6-12 Effect of Fuel Injection Timing at 3 bar hmep, 1250 rpmDDC 1-71 - Natural Gas Fueling (normalized) 100Figure 6-13 Effect of Fuel Injection Timing on Mass Fraction of BurnedFuel (DDC 1-71 - Diesel Fueling at 3 bar bmep) 102Figure 6-14 Effect of Fuel Injection Timing on Mass Fraction of BurnedFuel (DDC 1-71 - Natural Gas Fueling at 3 bar bmep) 102Figure 6-15 Effect of Fuel Injection Timing on Initial Burned GasTemperature (DDC 1-71 - Diesel Fueling at 3 bar bmep) 103Figure 6-16 Effect of Fuel Injection Timing on Initial Burned GasTemperature (DDC 1-71 - Natural Gas Fueling at 3 bar hmep) 103Figure 6-17 Effect of Fuel Injection Timing on Estimated Exhaust NO(DDC 1-71 - Diesel Fueling at 3 bar bmep) 104Figure 6-18 Measured and Estimated NO Emissions 105Figure 6-19 DDC 1-71 Thermal Efficiency - 0.016’ dia. Gas Holes 107Figure 6-20 DDC 1-71 Thermal Efficiency - 0.020” dia. Gas Holes 107Figure 6-21 DDC 1-71 Carbon Monoxide - 0.016” dia. Gas Holes 108Figure 6-22 DDC 1-71 Carbon Monoxide - 0.020” dia. Gas Holes 108ixFigure 6-23 DDC 1-71 Oxides of Nitrogen - 0.016” dia. Gas Holes 109Figure 6-24 DDC 1-71 Oxides of Nitrogen - 0.020” dia. Gas Holes 109Figure 6-25 DDC 1-71 Methane - 0.016” dia. Gas Holes 110Figure 6-26 DDC 1-71 Methane - 0.020” dia. Gas Holes 110Figure 6-27 DDC 1-71 Non-Methane Hydrocarbons0.016” dia. Gas Holes 111Figure 6-28 DDC 1-71 Non-Methane Hydrocarbons0.020” dia. Gas Holes 111Figure 6-29 DDC 1-71 Diesel-to-Gas Energy Ratio0.016” dia. Gas Holes 115Figure 6-30 DDC 1-71 Diesel-to-Gas Energy Ratio0)20” dia. Gas Holes 115Figure 6-31 Effect of Load on Mass Fraction of Burned FuelDiesel Fueling of the DDC 1-71 116Figure 6-32 Mass Fraction of Burned Fuel for Both Natural Gas andDiesel Fueling of the DDC 1-71 (3 bar bmep) 117Figure 6-33 Effect of Fuel and Load on Initial Burned Gas Temperature 118Figure A-i Pressure Transducer Mounting in Cylinder of DDC 1-71 128Figure A-2 Pressure Transducer Mounting in Cylinder of DDC 6V-92 129Figure A-3 Engine Cylinder Geometry 138xACKNOWLEDGMENTSSpecial thanks are in order to those who helped make this thesis a reality. I wish toexpress my deepest gratitude to my supervisor, Dr. P. G. Hill, for the opportunity to work withhim and more importantly to learn from him. His enthusiasm for research and discovery isinspirational.I would like thank Bruce Hodgins, research and project engineer, for driving the overallproject, designing and developing the natural gas injector, and his technical guidance and supportin the engine test cell set-up and instrumentation. His assistance in collecting and interpreting themeasured data is also appreciated.Thanks are due to Hardi Gunawan and Yinchu Tao for doing much of the preliminarywork to give this thesis a starting point. Thanks are also due to my fellow graduate students inthe combustion lab; Alain (Jack) Touchette, Patric Ouellette, Christoph Aichinger, Dehong Zang,and Alexander Chepakovich for being good friends.Thanks are also in order for NSERC for financial support during this work.xi1. INTRODUCTION1.1 Vehicle Exhaust Emissions and Alternative FuelsVehicle exhaust emissions have been identified as a substantial source of air pollution inurban centres world wide. Significant contributors to exhaust emissions in particular are diesel-powered urban buses. Diesel engines are the engine of choice in buses because of their greaterdurability and higher thermal efficiencies compared with gasoline engines. However, diesel enginemanufacturers are facing increasingly stringent emissions regulations which will be difficult tomeet using existing technology.Table 1-1 Urban Bus Heavy-Duty Engine Emission StandardsUnited States EPA Clean Air Act Amendments (1991)(g/bhp-hr measured during EPA heavy-duty engine test)Model year li1990 6.0 1.3 15.5 0.601991 5.0 1.3 15.5 0.251993 5.0 1.3 15.5 0.101994 5.0 1.3 15.5 0.051998 4.0 1.3 15.5 0.05The Environmental Protection Agency (EPA) in the United States was given the mandateto establish the Clean Air Act, which places a high priority on the reduction of air pollution andsmog levels in urban America. As part of the Clean Air Act, regulations are placed on vehicleexhaust emissions. The most recently proposed EPA emissions regulations for urban buses aregiven in Table 1-1. While no reduction in HC or CO emissions are required between the years1990 and 1998, 33% reduction of NON, and 92% reduction in PM are required. While dieselengine manufacturers are presently meeting regulations, considerable improvements will berequired to meet the proposed 1998 emissions regulations.Diesel exhaust generally has the same concentrations of nitrogen oxides (NOr) as gasolineengines and slightly lower concentrations of hydrocarbons (HC), C02, and CO. Diesel enginesare however a major source of particulate matter (PM), which is the emission that is facing thetoughest regulation. Particulate matter consists primarily of combustion generated carbonaceousmaterial (soot) on which some organic compounds absorb. 159 161 163 165 167 169 171 173 175 177BOl [°ABDC]Figure 1-1 : Effects of Fuel Injection Timing (Normalized)While the oxidation rate of PM generally increases with cylinder temperature, theformation rate of NO does as well. At a given engine speed and load, changes in the fuelinjection timing have a significant effect on exhaust emission levels of both NO and PM withlittle effect on thermal efficiency. Figure 1-1 illustrates the effects of fuel injection timing onemissions and thermal efficiency of a typical diesel engine. It demonstrates the traditional tradeoff between the reduction of either PM or NO in conventional diesel engines. Changinginjection timing to decrease PM emissions will increase NO emissions and vise versa.Strategies that can be employed to reduce diesel engine pollutant emissions include:2• Improvement of diesel engine design and electronic control• Use of particulate traps or other exhaust after-treatments• Use of alternative fuelsBecause of the trade-off between particulate matter and NO emissions tinder conventional dieseloperation, the amount by which pollutant emissions can he reduced by improvements of dieselengine design or electronic control is limited. Particulate traps’ could be used to reduce PMemissions which would allow the engine designer to employ NO reduction strategies. Heywood[112 points out that particulate traps are difficult to implement with heavy-duty diesel engines dueto high particulate loading and relatively low exhaust temperatures. The reliability and cost ofparticulate traps are also concerns. Other exhaust after-treatment equipment such as catalyticconverters and steady-flow thermal reactors are not effective with heavy-duty diesel engines.Catalysts in catalytic converters are fouled by particulate matter and exhaust temperatures are nothigh enough for thermal reactors to be effective [1].Diesel engine emissions can be improved by using a cleaner burning fuel. Preliminarytesting of alternative fuels in engines has shown promise in reducing emissions sufficiently to meetthe new regulations without concessions on thermal efficiency [2, 3, 4, 5, 6, and 7]. Natural gasis one alternative fuel in particular that seems to be attractive since it is relatively inexpensive andreadily available. Since the carbon content of natural gas is less than that of diesel fuel, CO2 andPM emissions can be lower which allows the flexibility of using NO reduction techniques. Thiseffectively decreases the prominence of the traditional trade-off between PM and NO emissions.One drawback of using natural gas as an alternative fuel for vehicles is its low volumetricenergy content. This requires large storage tanks to obtain reasonable vehicle range. Anotherdrawback is the high autoignition temperature of natural gas compared with diesel fuel. In orderfor natural gas to ignite in a conventional diesel engine, either a supplementary ignition source, ora higher compression ratio is required. With no supplementary ignition source, a compression1A temperature-tolerant filter or trap that removes particulate material from exhaust gases; the filter is “cleanedoff” by oxidizing the accumulated particulates. [1]2Numbers in square brackets designate references listed at the end of this thesis.3ratio on the order of 60 would be required for autoignition of natural gas. This is about threetimes conventional diesel engine compression ratios.Compared to other pollutant reduction strategies, alternative fuels have the added benefitof offsetting the increasing dependency and demand for conventional fuels. Development ofviable alternative fueling schemes for vehicles will have a significant impact on extending theworld’s usable energy resources. Hence considerable motivation exists for investigating the use ofalternative fuels in vehicle power plants.1.2 Natural Gas Fueling of Diesel EnginesThree principal methods for natural gas fueling of diesel engines are:• Manifold Injection• Timed Port Injection• Direct InjectionFigure 1-2 illustrates each method schematically. In the first method, natural gas isinjected into the engine inlet manifold where it mixes with the inducted air to form a fully premixed fuel-air mixture before the initiation of combustion. The mixture is compressed and thenignited near top dead center by either a spark plug or a small pilot amount of diesel fuel directlyinjected into the cylinder. While this method is relatively easy to implement, it has sonic seriousshortcomings when applied to vehicle engines.Manifold injection is not suitable for two-stroke engines as a considerable amount ofnatural gas would find its way through the engine cylinder, into the exhaust, during thescavenging process. Furthermore, part-load efficiency is much lower than with its dieselcounterpart since throttling of the intake air is required to maintain the pre-mixed air and fuelmixture ratio to within combustible limits. Knock can be encountered as a consequence ofsubjecting pre-mixed air and fuel mixtures to high compression ratios. Manifold injection ofnatural gas is best applied to engines that operate only at full load and constant speed conditions,such as engines for driving generators, compressors, or pumps.4NATURAL GASAIRAIR(A) NATURAL FUMIGATION(B) TIMED PORT INJECTIONAIRDIESEL FUEL(C) DIRECT INJECTIONFigure 1-2 : Methods of Using Natural Gas in Diesel Engines(courtesy of H. Gunawan (SI)PREMIXEDDIESEL FUELGASMIXERGAS INJECTION DIESEL FUELHIGH—PRESSURENATURAL GAS5The second method of natural gas fueling (timed port injection) is an improvement on thefirst but still has some of the same drawbacks. In this design, natural gas is injected near theintake port of each cylinder at timed intervals. With proper injection timing and little mixingduring the compression stroke, charge stratification can occur to some extent which reduces theproblems of fuel loss during scavenging and knock. Engines utilizing this method of natural gasfueling still require throttling of intake air and either a spark or diesel pilot to initiate combustion.The final method of natural gas fueling of diesel engines uses direct injection of high-pressure natural gas into the engine cylinder near the end of the compression stroke. Since thetemperature in the combustion chamber of a conventional diesel engine at the end of thecompression stroke is below the autoignition point of natural gas, a supplementary ignition sourceis required. If a small amount of diesel is also directly injected into combustion chamber near theend of the compression stroke, then the autoignition of the diesel will initiate combustion of thenatural gas. By directly injecting the natural gas, no throttling of the intake air is required formixture control as the charge is fully stratified. This suggests that part and full-load efficienciescan be comparable to those of a conventional diesel engine.In 1989 Beck et al. [3] reviewed the different concepts of natural gas fueling ofcompression ignition engines. They concluded that the high-pressure direct injection methoddescribed above has the following advantages:• Uses the basic diesel cycle with compression ignition.• Avoids engine knock.• Experiences no throttling losses.• Requires no mixture ratio control.• Obtains diesel cycle efficiency.• Produces negligible unburned fuel emissions.Miyake et al. [4] in 1983 presented the results of research carried Out at the MitsuiEngineering and Ship Building Company in Japan studying high-pressure gas injection in engines.They investigated two ignition methods in a large bore (420 mm) single cylinder 4-stroke engine.The first involved direct injection of compressed natural gas (CNG) and pilot diesel separately6with diesel injection preceding CNG. The second method involved direct injection of CNG anddiesel pilot simultaneously as a mixture.Better performance and lower amounts of diesel needed for stable operation wereachieved with the first method. With CNG injected at about 250 bar, thermal efficiencycomparable to that obtained with diesel-only operation was obtained using 5% (of the totalcalorific value at full load) of diesel pilot at 85% and 100% of full load. No results werepresented at low loads or at varying injection timings. A heat release analysis was performed,however no details of the analysis were given. The only emissions measurements presented wereCO and smoke.Similar work was presented by Einang, Koren, Kvamsdal, Hansen, and Sarsten [5] also in1983. Their project involved high-pressure, direct injection natural gas fueling with diesel pilot ofa single cylinder, two-stroke, loop-scavenged, 300 mm bore engine which produces 196 kW at375 rpm. Since their main objective was to have the capability of running with up to 100% diesel,the diesel injector and its location were unaltered as gas was injected through a separate injectordisplaced off-center in the cylinder head. In preliminary tests without gas injection optimization,it was demonstrated that with 73% (by energy) natural gas, injected at pressures below 200 bar,thermal efficiencies were slightly better, smoke readings were slightly higher, and NO emissionswere 24% lower than 100% diesel operation.Later Einang, Engja, and Vestergren [6] were able to achieve reliable and stable ignitiondown to 2% diesel pilot in a medium speed, four-stroke diesel engine. The same thermalefficiency as the 100% diesel operation was obtained with 5% pilot. A single injector was used toinject both natural gas and the diesel pilot. The pilot fuel nozzle and spray pattern were identicalto that of the standard diesel engine with the gas jets directed so as to avoid collisions with thediesel jets. In this work, details are presented for preliminary tests of their prototype injector inone cylinder of a multi-cylinder engine. They did not present results from full range performanceor emissions tests, and a detailed combustion analysis was not undertaken.In 1987 Wakenell, O’Neal and Baker [7] presented the results of testing of high-pressure,late cycle, direct injection of natural gas in a medium speed diesel engine. They injected natural7gas at 5000 psig (345 bar) into a two-cylinder, two-stroke, blower scavenged locomotive researchengine (216 mm bore and 835 rpm maximum speed). Separate injectors were used for natural gasand diesel. With the engine running on natural gas, rated speed and power were achieved withslightly lower thermal efficiency than with standard diesel operation.The effects of diesel pilot quantity were investigated. Full power was achieved with aratio of 99% natural gas and 1% diesel fuel, however audible knock was detected. It was foundthat engine knock was unacceptable unless the pilot ratio was increased to approximately 10% ofthe total fuel input which gave a thermal efficiency 21% less than the diesel baseline efficiency of28.7% at full load. Part load was unobtainable at these natural gas injection pressures because ofunstable engine operation. While this work included a brief mention of fuel injection timing andheat release analysis, results from a thorough investigation of either were not presented. Thiswork did however include some emissions measurements of smoke, NO, CO and hydrocarbons atdifferent engine operating conditions.While much work has been done by other researchers on natural gas fueling of dieselengines, there is presently no commercially available direct injection natural gas fueling system forurban bus diesel engines that meet proposed EPA emissions regulations. An important factor inapplying natural gas fueling technology to urban buses, as compared with other applications, isthe requirement of good performance over the entire operating range of the engine. Performance,emissions and combustion characteristics of a viable direct-injection natural-gas fueling system forurban-bus diesel engines has not been comprehensively investigated over the entire engineoperating range.Combustion analysis based on measured cylinder pressure time histories are importantengine diagnostic tools which also provide insight into in-cylinder processes, such as fuel burningrates, fuel injection phenomena, heat transfer and pollutant formation. An accurate combustionmodel is of primary importance in gaining a thorough understanding of the advantages andlimitations of natural gas fueling of diesel engines. As will be shown, previous combustionanalysis techniques have shortcomings when applied to stratified charge engines (i.e. diesel8engines). The need exists for a combustion analysis technique which can give a more accurateindication of the combustion phenomena occurring inside of a diesel engine.1.3 Intensifier-Injector for Natural Gas Fueling of Diesel EnginesAn intensifier-injector system for natural gas fueling of diesel engines is presently in thedevelopment stages at UBC. The intensifier-injector concept employs electronically-controlled,late-cycle, direct injection of high pressure natural gas with diesel pilot. It is proposed that naturalgas be compressed from storage tank pressure up to injection pressure by an engine-drivenintensifier compressor. Both natural gas and diesel fuel are directly injected into the combustionchamber at high pressure through a single injector of roughly the same outer dimensions as thestandard diesel injector it is intended to replace.Fuel injection is initiated toward the end of the compression stroke with diesel injectionpreceding CNG. Since the temperature in the combustion chamber at the beginning of injection isabove the diesel fuel’s ignition point, diesel and air that have mixed to within combustible limitswill spontaneously ignite after a delay period of a few crank angle degrees. Since the ignitionpoint of the natural gas can not be achieved by compression alone, combustion of the diesel fuel isrequired to raise the temperature such that the natural gas will ignite.The goal of the intensifier-injector project is to develop a retro-fit system for natural gasfueling of urban buses capable of meeting the stringent EPA emissions standards whilemaintaining high efficiency. In obtaining this goal, it is important to have a minimum of changesto engine components and systems such that this retro-filting is commercially viable. The work inthis thesis is part of the ongoing effort at UBC to achieve this goal.Preliminary work at UBC concentrated on the evaluation of an electronically controlledpoppet type injector from which high-pressure natural gas along with a small pilot amount ofdiesel fuel are injected [8, 9, 10, 11]. Just before leaving the injector tip through the poppetvalve, the diesel pilot is gas-blast atomized by the flow of high-pressure natural gas. Hence thenatural gas and diesel fuel are injected simultaneously as a mixture. Along with being tested in a9naturally aspirated, two-stroke, single-cylinder diesel engine, this prototype injector has beenanalyzed using flow visualization techniques.The main findings of this preliminary work with the poppet type injector are that:• NO and CO2 emissions3 are lower than with conventional diesel operation over the entireoperating range of the engine, while CO and hydrocarbon (both methane and non-methane)emissions are greater.• at high loads, thermal efficiencies exceeded those for conventional diesel operation; howeverlower part load thermal efficiencies were found4.• a minimum of 25% liquid diesel fuel (by energy) was required for stable engine operation overthe entire load range.Suggestions for improvements based on these findings are to:• redesign the injector nozzle geometry to obtain better penetration and distribution of fuel inthe combustion chamber.• separate injection of diesel and natural gas, with diesel injection preceding that of natural gas,to improve ignition.• improve diesel spray and atomization to reduce ignition delay periods and CO andhydrocarbon emissions.Implementation of these suggestions forms the basis for the objectives of the present research.1.4 Objectives of This ResearchThis research is concerned with the investigation of diesel engine perfonnance, emissionsand combustion using natural gas fueling. The objectives of this work are as follows:1) To measure diesel engine performance and emissions, using both natural-gas and conventionaldiesel fueling, over a wide range of operating conditions.2) To model the stratified-charge combustion process in diesel engines, based on measuredcylinder pressure.3WhiIe CO2 is not a regulated emission, it is thought to be a contributor to global warming.4High cycle-to-cycle variations were found at low loads.103) To learn about the formation rates of oxides of nitrogen in diesel engines to gain a betterunderstanding of how to reduce NO emissions.4) To learn about fuel burning rates to gain insight into the differences between combustion ofdiesel fuel and natural gas in diesel engines.112. TESTING APPARATUS2.1 IntroductionTwo different engines have been used to test the natural gas fueling concept. Each enginewas instrumented in order to measure key parameters such that a full analysis of engineperformance and combustion could be carried out. Measured engine parameters include outputtorque, engine speed, fuel flow, air flow, intake air pressure and temperature, exhaust pressureand temperature, cylinder pressure, crank angle position, and exhaust emission levels. Measuredexhaust gas constituents include C02, CO, N0, 02, CH4, and THC (total hydrocarbons).Figure 2-1 : Schematic of Engine Test Set-up and Instrumentation (DDC 6V-92TA)A computer controlled data acquisition system is used to tie the whole system together.Engine sensor signals are sent to an IBM 286 PC after signal modification (i.e. AID conversion,I1Amb:ntAirempI-Engine andDynamometer I I AirboxL Control Panel T&PAi I DseLV Flow \SignalConditioning Gas Flowand [cLTerminal Board — — —ngneSpd_Dynamometer TorquelAirFlowI I-Exhaust GasV Exhaust Gas I & PCylinder BDC &Pressure I Crank AngleOptical Incoder_IISAACITC = turbochargerB = blowerAC = after cooler12filtering, amplification, etc.). Software, developed in-house at UBC , is used to coordinate theacquisition of performance, emissions and high speed cylinder pressure data. Figure 2-1 is aschematic of the engine test set-up and instrumentation.2.2 Engine and Engine ControlThe engines used for testing are a single-cylinder, naturally aspirated Detroit Diesel 1-71and a six-cylinder, turbo-charged and after-cooled Detroit Diesel 6V-92. These engines werechosen because six-cylinder versions of the 71 and 92 Series Detroit Diesel engines power morethan 90% of the urban buses in North America. Both engines operate on a two-stroke cycle anduse blower-forced uniflow scavenging. Table 2-1 lists the general specifications of these engines.Table 2-1 : General Engine SpecificationsDDC 1-71 DDC 6V-92TABasic Engine 2 cycle 2-cycle-VeeNumber of Cylinders 1 6Control DDEC I DDEC IIBore and Stroke 4.25 x 5.0 in (108 x 127 mm) 4.84 x 5.0 in (123 x 127 mm)Displacement 70.93 cu in. (1.162 litres) 552 cu in (9.05 litres)Compression Ratio 16.0 to 1 17.0 to 1Gross Rated Power Output 15 BHP (11.2 kW) @ 1200 RPM 300 BHP (224 kW) @ 2100 RPMGross Rated BMEP 4.8 bar @ 1200 RPM 9.2 bar @ 1200 RPMRated Peak Torque 56lbft (76Nm) @ 1200 RPM 975lb1t (1322Nm) @ 1200 RPMInlet Port Closure 60 0 ABDC 550 ABDCThe DDC 1-71 had originally been equipped with a mechanically controlled unit injector.Its injection system was upgraded to a Detroit Diesel Electronic Control (DDEC) system similarto the one used by the DDC 6V-92TA. The DDC 1-71 uses the first generation DDEC I system,13while the DDC 6V-92TA uses the improved DDEC II system. The electronic fuel injectionsystem consists of electronic unit injectors, sensors and an electronic control module (ECM). TheECM contains a microprocessor which control fuel injection timing and quantity by sendingactuation signals to the injector solenoid. The ECM monitors the current in the injector solenoidsto sense when the injector valve closes and uses this information as feedback to compute timingfor subsequent injection events.By interfacing with the ECM, the beginning and duration of fuel injection fuel can befreely selected manually. This creates the opportunity to fully explore engine performance,emissions, and combustion characteristics under a variety of fueling strategies. The DDC 6V-92TA uses two separate ECMs; one to control the diesel injectors, the other to control thenatural-gas injector independently.2.3 Dynamometer and Dynainorneter ControlA dynamometer is used to dissipate mechanical energy developed by the engine and tomeasure engine output torque. The DDC 6V-92TA is coupled to an eddy current or inductordynamometer, while the 1-71 is coupled to a water-brake dynamometer. In an eddy currentdynamometer, the output shaft of the engine is connected to a rotor which rotates in a magneticfield provided by a DC current flowing through coils in a stator. Voltages induced in the rotorcauses local currents to flow in short circuit paths or eddies. The energy of these eddy currents isdissipated in the form of heat which is carried away by the flow of water through a jacket on thestator.In a water-brake dynamometer, mechanical energy dissipation is created by a volume ofwater within the dynamometer acting upon straight vanes on the stator and rotor. [121 Thevolume of water within the dynamometer is determined by the water flow rate passing through theunit. This flow rate is controlled by both a water inlet control valve and onfices in the outletwater line. At the same time, the power developed by the engine is absorbed within thedynamometer, converted into heat, and carried away by the outlet water.The eddy current dynamometer has a rated absorption capacity of 200 hp, which is only14two thirds of the maximum rated power output of the six-cylinder engine. While testing however,instead of the dynamometer being power limited, it was found to be torque limited (independentof engine speed) at about 630 lbft. Looking at 1200 and 2100 RPM, for example, only 65% and84% of maximum power could be achieved respectively. The water-brake dynamometer has arated absorption capacity of 550 hp, which is more than sufficient for full load operation of thesingle cylinder engine.In an eddy current dynamometer, torque is controlled by manipulating the DC field currentsince field current is proportional to torque. In the eddy current dynamometer controller used,both rpm and torque are measured and fed back to a PID controller. The proportional gain,integral time and derivative time can be tuned via pots on the control module to achieve optimumtransient control. The manufacturer specifies that the precision is +1- 1 RPM for speed controland +1- 0.1% (0.02% typ) for load control. Accept at high speed low load operation, these claimswere found to be true. PID control is also integrated into the water-brake dynamometercontroller, but is by-passed because the response times associated with changes in water flow rateare slow and lead to unstable dynamometer operation. Hence the water flow rate (whichregulates torque) is controlled manually using an electro-pneumatically operated valve.2.4 Fuel Injection SystemBoth natural gas and a small amount of diesel (pilot) are directly injected into the enginecylinder at high pressure through a single injector. The natural gas and diesel fuel supply systemsare independent of one another, each with its own mass flowmeter. Figure 2-2 is a schematic ofthe natural gas and diesel fuel supply systems. The diesel fuel, from a storage tank, flows into aflow measuring device. The diesel fuel returning from the engine (not shown on the figure) alsoflows into this measuring device such that the flow meter measures the net diesel fuelconsumption of the engine. Natural gas, from the main supply at a pressure of 14 kPa, iscompressed to 19 MPa by a commercially available compressor and stored in pressure bottles.Natural gas drawn from these bottles flows first through a regulator and then through the flowmeasuring device before reaching the injector.15O1PRESSORFigure 2-2 : Schematic of Engine Fuel Supply System(courtesy of H. Gunawan [8])In conventional mechanical fuel injeclors, injection timing and injection rate aremechanically controlled by ports and helices machined in the hushing and plunger assembly. In aneiectromcally controlled fuel injector, liexibility and control are enhanced because timing andduration of the fuel injection is performed electronically. Figure 2-3 shows the schematic of theelectronically controlled natural-gas unit injector. A cam is used to depress the plunger forpressurization of the liquid diesel fuel. A solenoid-operated spool valve performs the fuelinjection and metering functions.When the solenoid coil is energized, the spool valve closes the diesel fuel flow passage.Closure of the spool valve initiates pressurization of the liquid diesel fuel. Once the pressure hasreached a preset value, the diesel needle lifts to begin injection of diesel fuel. Pressure continuesto build until the opening pressure of the natural-gas needle has been reached to begin injection oftot34PRESSEDNATURAL GASSUT-DFFDUAL rti..UNITIN.ECTORnatural gas. Opening of the spool valve causes pressure decay and end of injection. The durationof spool-valve closure, referred to as the pulse width (PW), determines the quantity of natural gasinjected. The amount of diesel fuel injected is constant with changing pulse width. The enginecontroller sends actuation signals to the solenoids according to fuel timing and quantityrequirements.Figure 2-3 : Natural-Gas Electronic Unit Injector(courtesy of K.B. Hodgins)2.5 Engine InstrumentationThe principal measured parameters are output torque, engine speed, fuel and air flows,cylinder pressure, intake and exhaust temperatures and pressures, and emissions levels. Thefollowing sub-sections outline the measuring techniques used in each case.DDECSOLENOIDACCESSORY SHAFTDRIVEN ACTUATORDIESELSUPPLY/RET URNVALVE CHECK VALVEPILOT NEEDLEGAS NEEDLECHECKVALVECNQDleset PI(otCNG StorcAge(20—200 bar)172.5.1 Output TorqueThe torque developed by the engine is measured with the use of the dynamometer and astrain gage load cell. The dynamometer is trunnion mounted such that the stator housing is freeto rotate. A load cell, attached a to support, is used to restrict the rotation of the stator. As therotor rotates within the stator, a torque is developed which as an applied force is measured by theload cell. The load cell can be calibrated by placing weights at the end of an extended momentarm bolted to the dynamometer casing. Maximun error in this instrument is ±0.1%.2.5.2 Engine SpeedThe engine speed measurement is acquired with a magnetic induction probe shaft encoder.The magnetic induction probe is placed over a 60-tooth gear rotating with the crank shaft. Thissensor provides a signal frequency proportional to engine speed (1RPM = 1 Hz). A hand-helddigital tachometer was used to calibrate this sensor. With a piece of reflective tape attached tothe engine output shaft and the hand-held tachometer pointed toward it, the unit sends out acontinuous beam of light and counts the light pulses reflected by the tape. Maximun error in thisinstrument is ±0.1%.2.5.3 Diesel Fuel FlowThe diesel mass flow rate measuring device used on the DDC 6V-92TA operates on thehydraulic “Wheatstone Bridge” principle. The system consists of three primary pieces ofequipment: a fuel boost pump, a recirculating tank and the mass flow meter itself. Figure 2-4 is aschematic of this system. As illustrated, the diesel fuel, drawn from the fuel tank, passes througha filter and is then pressurized by the boost pump to a specified level set with a relief valve.Excess diesel fuel not used downstream is returned to the pump inlet through this relief valve.This provides a clean smooth supply of fuel to the mass flow measuring device.Diesel fuel pumped from the storage tank flows through the mass flow measuring deviceinto the recirculating tank. From the recirculating tank, the diesel fuel passes through a heatexchanger which maintains its temperature at 40°C ± 3°C to conform with SAE standard J199518JUN90 [16] and then flows to the engine. Since diesel fuel returning from the engine also flowinto the recirculating tank, the mass flow measuring device measures the net fuel consumption ofthe engine. A vertical tube with small cross-sectional area, referred to as the stand pipe, sits atopthe recirculating tank. By regulating the pressure of the supply fuel entering the recirculationtank, the mass flow meter measures the make up diesel flow required to maintain a constant levelin the standpipe. Although sudden changes in the engine’s diesel fuel demands will cause the fuellevel in the stand pipe to deviate, this set-up is designed to minimize the time for the level toreturn to its preset position.Figure 2-4: Diesel Mass Flow Rate Measuring Device Schematic (DDC 6V-92TA)The mass flow rate of the make up diesel fuel is measured directly using an arrangement offour orifices and a constant volume pump in a “Wheatstone Bridge network [13]. When thesupply flow (Q) is greater than the recirculating flow (q) through the constant volume pump, themass flow rate can be directly related to the pressure difference P., — P3 if the products of the flowRelief ValveRegulatorCP4 RegulatorEngineEngineI Shell-in-TubeHeatExchangerTankW.19coefficient and area of orifices ‘a’ and ‘d” are equal. When Q is less than q, the mass flow rate isdirectly related to the pressure difference P1 — P4 assuming the products of the flow coefficientand area of orifices “b’ and ‘c” are equal. The “Wheatstone Bridge” network allows the massflow rate of diesel fuel to he determined by measuring only differential pressures without havingto uniquely determine the fuel density. The derivation of the proportionality between mass flowrate and differential pressure is given in Appendix I.While the layout of the diesel mass flow rate measuring device used on the DDC 1-71 issimilar to one used on the DDC 6V-92TA, it operates on a different principle than that describedabove. Here the weight of the recirculation tank is measured directly to give an estimate of themass flow rate. Fuel from the recirculation tank flows to the engine while fuel returning from theengine flows into the recirculation tank. Hence the diesel fuel consumption rate of the engine canbe determined by measuring the change in weight of the recirculation tank. When therecirculation tank is nearly empty, it is re-filled from the diesel storage tank. A solenoid operatedvalve is used to control the re-filling process, during which diesel flow measurements can not bemade. For accurate diesel fuel flow measurements, it is important to bleed each system to removeany trapped air. The maximum diesel flow measurement error is ±2%.2.5.4 Natural Gas and Air flowsThe compressed natural gas mass flow is measured using an instrument which operates onthe Coriolis acceleration principle. The natural gas flows through a U-shaped tube which vibratesat a frequency directly proportional to the product of the density and velocity of the gas. Thefrequency of the signal produced is converted to a 4-20 mA signal and sent to the data acquisitionsystem via a remote flow transmitter. The manufacturer’s calibration, which was used for thismeter, demonstrated an average measurement error of ±0.4%.The air volume flow rate is measured using a turbine meter located immediatelydownstream of the air filter in the air intake. The maximum error, quoted by the manufacturer, is±0.5% of full scale (DDC ÔV-92TA). The mass flow rate of the air is calculated from the airdensity based on ambient temperature and pressure measurements.202.5.5 Cylinder PressureEngine cylinder pressure is a key variable required in fundamental engine performance andcombustion analysis. Engine cylinder pressures are used to determine the rate of combustion,indicated work, cyclic variation and engine friction. However to achieve adequate results in anyanalysis based on pressure data, the importance of its accuracy is paramount. Therefore sometime has been spent on determining the most satisfactory way of collecting quality cylinderpressure data.To measure cylinder pressure a piezoelectric pressure transducer and a charge amplifierare required. Piezoelectric pressure transducers generate an electrical charge proportional to thepressure they are sensing. A charge amplifier (charge amp) is required to produce an outputvoltage (generally -1OV to +1OV) proportional to this charge. When dealing with electricalcharges, high insulation resistance in the transducer and all connections and cables between thetransducer and charge amplifier are required. If this resistance is lowered in anyway, chargeleakage will occur which will be seen as signal drift. Hence, care must be taken to clean allconnections thoroughly using a cleaning spray which leaves no residue. Bending or stretching ofcables can also lead to resistance degradation.Proper setup of the charge amplifier is important for obtaining accurate pressure data.The piezoelectric transducer has a high output impedance (on the order of a thousand megaohms)[14] and a low output signal. Thus it is critical that the charge amplifier has a high inputimpedance to match that of the transducer. If the impedances are mismatched, loading of thetransducer will occur. The loading effect is when the measured signal is distorted by themeasurement equipment. The longer the time constant of the charge amplifier, the higher theinput impedance. However, with time constants that are too long, signal drift can be encountered[15].To investigate the effect that charge amp time constant has on measured cylinder pressure,pressure data at short, medium and long time constants were collected from a cylinder in the DDC6V-92TA when running at 1200 rpm and a medium-high load condition (6 bar bmep). The datacollected at the medium and long time constants settings are almost identical throughout the entire21cycle. On the other hand, while the data collected at the short time constant setting matches theother two data sets near top-dead-centre, the pressures during compression are significantlyhigher and during expansion are significantly lower. The result is a narrower compression andexpansion loop on the log-log plot of pressure versus volume with an exaggerated scavengingloop.These effects can be seen in Figure 2-5. Figure 2-5 is a cylinder pressure versus volumelog-log plot of the superposition of an average of 20 consecutive cycles of pressure data collectedat both short and medium time constant setting on the charge amp at the same engine operatingcondition. A time constant of medium was found to give the most accurate results.Figure 2-5 : Effects of Charge Amplifier Time Constant on Cylinder Pressure DataThe type of transducer and its mounting are equally important in obtaining accuratepressure data. The hostile thermal environment of the diesel engine combustion chamber caninduce thermal strains on the piezoelectric material to give erroneous pressure readings. Hence,the type of transducer and the manner in which it is mounted should be chosen in such a way thatthe thermal effects are minimized while still measuring true cylinder pressures. Both watercooling and recess mounting can be used to reduce thermal effects. Recessing the transduceraway from the fire deck can avoid the flame coming in direct contact with the transducer since the3. 2.1 2.3 2.5 2.7Iog(VcyI)I I2.9 3.1 3.322flame will quench on the passage walls. With a passage mounting there can be the danger ofacoustic delay and/or passage resonance. As an example of acoustic delay, assuming a cylindercomposition of air at a temperature of 1000K, the speed of sound is approximately 620 rn/s. Thismeans that for a transducer recessed 6mm, the pressure wave reaching the transducer’s sensingsurface will be delayed by 0.7 °CA at 1200 rpm.To investigate transducer thermal effects, a number of different types and mountings oftransducers were tested. Thermal effects on pressure data are similar to the effects of charge amptime constants shown above in Figure 2-5. As thermal effects become more apparent, thecompression and expansion ioop becomes narrower and the scavenging loop becomes moreexaggerated on the log-log plot of cylinder pressure versus volume. An immediate indication oferroneous pressure data is seen when the resulting calculation of frictional work is unrealisticallylow or negative.Pressure transducers tested included a water cooled AVL 8QP 500c and several non-water cooled PCB transducers with model numbers 112A05, 112A10, and 114M297. The PCB112A05 is a high temperature model. The PCB 112A10 is the same as the 1 12A05, however itincludes an extended slotted head which is intended to further reduce thermal effects byquenching combustion flames before they contact the transducer’s sensing surface. The 114M297is a larger high temperature model. Several different recess lengths were experimented with whenmounting each of the transducers. The most satisfactory pressure data was collected using thePCB 112A10. The pressure transducer mounting details for both the DDC 1-71 and the DDC6V-92TA are given in Appendices II and Ill respectively2.5.6 Mean Engine Temperatures and PressuresMean temperatures and pressures of fluids flowing through the engine are required tooperate, monitor and analyze the engine. Thermocouples are used to measure temperatures forthe following...• ambient air• air in the airhox23• exhaust productsi) directly outside each exhaust port in the exhaust manifoldii) downstream of the turbocharger (DDC 6V-92TA)• supply diesel fuel• engine cooling water both entering and leaving• engine oil temperature• water leaving the dynamometerPressures were measured for the following...• ambient air• air in the airbox• exhaust products upstream of the turbocharger (DDC 6V-92TA)2.5.7 Exhaust Gas AnalyzersThe concentration of exhaust gas constituents in the engine tail-pipe are measured on avolumetric basis. The measured constituents include NOR, C02, CO. THC (total hydrocarbons),Cl4, 02, and smoke. The emissions instrumentation is shared between three engines used by theAlternate Fuels Research Group at UBC. The system has been given full treatment in otherpublications such as the M.A.Sc. thesis written by Y. Tao [9]. For completeness however, a briefdescription of the system will be given here.The emissions analysis system consists of six distinct analyzers. Each of the exhaust gasconstituents are measured using separate analyzers except for NO and H4 which are combinedinto a single analyzer. The Non-dispersive Infrared (NDIR) absorption detection method is usedto measure the concentrations of CO, CU2, NON, and CH4. The Flame Ionization Detection(FID) method is used to measure the concentration of total hydrocarbons. The paramagneticproperties of oxygen are exploited to measure its concentration. Exhaust smoke levels aremeasured independently of the rest of the system using a Bosch “Spot” Smokemeter.So that hydrocarbons and nitrogen dioxides do not condense out of the sample gas beforereaching the measuring instruments, the exhaust gas analyzers are fed via a heated sampling24system which includes heated sampling lines, filters, pump, and valves. The FID instrument andN07 to NO converter are also heated for the same reason. The NDIR instruments, on the otherhand, require a water free exhaust sample. Therefore a chiller is located upstream of theseinstruments to condense all of the water Out of the sample.Figure 2-6 : Emissions Analysis System Schematic (heated instruments)(courtesy of Y. Tao f9j)#1*2*3To conform with the SAE recommended practice J 177 APR82 [16], the sampling probe islocated in the exhaust line at a distance of 1- to 3m from the exhaust manifold outlet flange or theoutlet of the turbocharger. From the sampling probe exhaust gases flow through the heatedsampling lines to the heated instruments enclosure (Figure 2-6). After passing through particulatefilters, the flow is separated into two paths. One path leads to the FID to measure hydrocarbons,the other leads to the heated pump. After passing through the pump, the flow is again separatedinto two paths. One line goes through the heated NO7 to NO converter and then through theSAMPLE INI IICABINET 1I —zciEXHAUST f25chiller to the CH4JNO instrument in the cool instruments enclosure. The other line flows throughthe chiller and is then separated into three paths in the cool instruments enclosure, each leading tothe remaining instruments (C07, CO, and 02). A schematic of the cool emissions instruments isgiven in Figure 2-7.#1#2NOx SAMPLE’N2Figure 2-7 : Emissions Analysis System Schematic (cool instruments)(courtesy of Y. Tao [9])2.6 Data Acquisition SystemOutput signals from all of the engine sensors (except for cylinder pressure) are sent to aninterconnect board. The signals are then conditioned and converted from differential signals tosingle-ended signals with a common ground before being sent to a 12 bit AiD board (PCL 818)installed in an IBM PC. The PCL-818 AID board can accept up to 16 single-ended signals. Table2-2 lists the 16 measured engine parameters along with their corresponding AID board channelnumbers for both the DDC 1-71 and DDC 6V-Q2TA data acquisition systems.26CABINET 2The cylinder pressure output signal from the charge amp is sent to a self-contained highspeed data acquisition unit (ISSAC) along with bottom dead centre (BDC) and crank anglesignals from a crank shaft mounted optical encoder. The BDC signal is used to trigger dataacquisition while the crank angle signal is used as an external clock to determine the intervals atwhich the cylinder pressure signal is to be sampled. Up to 100 consecutive cycles of pressuredata can be taken and stored by the ISSAC. This data, in binary format, can be downloaded tothe PC through a general purpose interface board.Table 2-2: Measured Engine ParametersCHANNEL DDC 1-71 DDC 6V-92TA0 CNG Mass Flow CO emission1 Beginning of Injection CO2 emission2 Ambient Temperature NO emission3 Pulse Width THC emission4 Torque CNG Pressure5 RPM RPM6 Intake Air Pressure Torque7 Diesel Mass Flow CNG Mass Flow8 Intake Air Flow Diesel Mass Flow9 CNG Pressure Airbox Temperature10 CH4 emission Ambient TemperatureU 02 emission Airbox Pressure12 CO2 emission Exhaust Pressure13 CO emission Intake Air Flow14 THC emission CH4 emission15 NO emission 02 emission27Software, written in-house at UBC, is used to coordinate the data acquisition, convert thesignals to selected engineering units, perform calculations and display selected parameters. Eachof the 16 channels on the AID board is sampled 1D() times and then averaged. Averaged data isused to calculate engine performance and emission parameters such as thermal efficiency, poweroutput, and wet basis emissions. The averaged raw data and calculated data can be saved to diskwhen commanded to do so. The software is also used to interface with the ISSAC to acquirecylinder pressure data. A description of the software used for processing cylinder pressure isgiven in Appendix IV.283. CALCULATED ENGINE PERFORMANCE AND EMISSIONS PARAMETERSThe parameters which are useful in comparing the performance and emissionscharacteristics of different engines, or the same engine at different operating conditions arethermal efficiency, mean effective pressure, coefficient of variation, wet basis and brake specificemissions, and unburned fuel fractions.3.1 Thermal EfficiencyThermal efficiency (lth) is defined as the ratio of the engine output work to a measure ofthe chemical energy of the consumed fuel. Thermal efficiency can be written asPhlthjjilj.IV (3-1)where th and LHV are the mass flow rate and lower heating value of the fuel, respectively, andb is the engine brake power. When more than one fuel is involved in the combustion, theexpression for thermal efficiency can be written more generally as1th = (3-2)The engine brake power is calculated from measurements made at the engine output shaft.It is given byPb=NWb (3-3)29where N is the engine speed and Wb is the engine output work. For a two-stroke engine theoutput work per crank shaft revolution is given byWb=2JtTb (3-4)where Tb is the measured engine torque. In order to have a common basis for comparison undervarying atmospheric conditions, the brake power is corrected to standard conditions (i.e. 100 kPainlet air pressure, 99 kPa dry inlet air pressure, 25°C inlet air temperature). The correctionprocedure is in accordance with SAE Standard J1349 JUN85 [16] and is given in Appendix V.The calculated engine brake power and thermal efficiency have an associated error due tothe error in the measured parameters. The error in each calculated parameters is found by takingthe differential of the expression representing that parameter. The relative error can then be foundby dividing by the original expression. The relative error in brake power is given byb Tb NFrom chapter 2, the relative error in engine torque and speed measurements were both ±0.1. Thisgives a relative error in brake power of ±0.2%. Similarly, the relative error in thermal efficiency isgiven byTI P diLHVi+Ihg•LHVgwhich can be rewritten as30am m am gss= 8P — rh g thgsLHV1 + LHVrngasAlso from Chapter 2, the relative error in natural gas and diesel flow measurements were ±0.4%and ±2%, respectively. Using lower heating values of 49680 kJ/kg and 43200 kJ/kg for naturalgas and diesel, respectively, and using a typical mass ratio of diesel fuel to natural gas of 30%,then the maximum relative error in thermal efficiency is ±0.9%.3.2 Mean Effective PressureA parameter that scales out the effect of engine size on engine output work is the meaneffective pressure (mep). It is defined as followsmep=-- (3-5)where Vd is the engine displacement volume. As the name implies, the units of mep are pressureunits. The brake mean effective pressure (bmep) is the output work, as defined in equation (3-4),divided by the engine displacement volume. The bmep can therefore be written asbmep= — b (3-6)VdThe indicated mean effective pressure (imep) is the sum of the bmep and the total frictionmean effective pressure (tfmep). The tfmep includes the work required to pump mixture into andout of the cylinder, to overcome resistance to relative motion of all moving parts, and to drive31engine accessories such as the blower, water pump, oil pump, alternator, etc.. The imep is given• Wdimep =by(3-7)where Wfld is the indicated work and VdC,,l is the displacement volume of a single cylinder. Theimep is calculated directly from measured cylinder pressures since the indicated work, defined byWmd=fPdV (3-8)is the area contained within the pressure-volume curve as shown in Figure 3-1.700060005000Cylinder 4000P res sure(kPa) 30002000100000 0.5 1 1.5 2Cylinder Volume (litres)Figure 3-1 : Typical P-V diagram for a Two-Stroke Diesel Engine3.3 Cyclic VariationObservation of successive cycles of engine cylinder pressure on an oscilloscope indicatescycle-to-cycle variability in the pressure rise due to combustion. Significant losses in power andefficiency can occur with high levels of cyclic variation. From a diagnostic point of view, it isdesirable to quantify this combustion variability. One measure of cyclic variability is derived frompressure data and is called the coefficient of variation in imep (COVjmp)• It is defined as thestandard deviation in imep divided by the mean imep as follows32COVmep = 0lmep xlOO (3-9)imepHeywood [1] points out that vehicle driveability problems usually result when COV exceedsabout 10 percent.3.4 Wet-Basis and Brake-Specific EmissionsA volumetric analysis of the engine tail-pipe exhaust emissions gives a great deal of insightinto the combustion phenomena and overall engine performance. It is informative to look at boththe concentration of exhaust gas constituents measured in parts per million (ppm) or percent byvolume and the brake specific emissions levels. Brake specific emissions are simply the mass flowrate of each exhaust gas constituent divided by the engine brake power.As described in section 2.5.7, all but the total hydrocarbon (THC) analyzer require thesample to be free from water vapour. Since water vapor constitutes a significant portion of theengine tail-pipe exhaust, these “dry-basis” concentration measurements will overestimate the trueconstituent concentrations. Therefore a conversion to “wet-basis” is required to account for thewater vapor present in the engine exhaust. The wet-basis values are the ones used in the brakespecific emissions calculations.Heywood [1] derives the following relationship between wet and dry mole fractions usingthe notation y. and y’ to denote respectively the wet and dry mole fraction of species i for a fuelcomposition CnHm•y1=(lYH2o) i (3-10)where* *m y + YcoYH2or * / * i * * (311)[1+y0/K1. Yco2)+(m/2Yco + Yco2)j33Since two fuels are present (natural gas and diesel), the fuel composition CmHn represents anaveraged composition based on the fuel flow rates of both fuels. The constant K1 relates the wetbasis concentrations of C02, CO, H20, and 2 based on available exhaust gas composition datain the following way:K1= YcoYH2o (3-12)Yco2 YH2K1 is an empirical constant determined from published exhaust gas composition data. It isrequired since the concentration of H2 is not measured directly. Values of 3.8 and 3.5 arecommonly used for K1. The difference between these values has little effect on the computedmagnitude of the mole fraction of water. The relationship given in equation (3-10) is used withthe equations (3-11) and (3-12) to convert all of the mole fractions of the measured exhaust gasconstituents from a dry to a wet basis with the exception of THC, as it is the only one measuredon a wet basis. In addition to the measured constituents (C02, CO, NO, 02, CH4, and THC),the use of the above equations enables the wet mole fractions of N2, 112 and H20 to becalculated as well.An additional correction must be made to the mole fraction of N0 since the humidity ofthe inlet air has an effect on the amount of NO chemically formed in combustion. The correction,from the SAE Recommended Practice J177 APR82 [16], is applied to the wet basis NO molefraction as follows:YNOrr=(3-13)whereK2 =1+ 7A(H — 10.714)÷ 1.8B(T — 29.444)A = 0.044(F/A)_0.0038B = _0.116(F/A)÷ 0.005334T is the intake air temperature (in °C), F/A is the fuel-air mass ratio (dry basis), and H is thespecific humidity in grams of water per kilogram of dry air. The calculation of specific humidity isgiven in Appendix VI. At 25°C, 20% relative humidity, and equivalence ratios of 0.2 and 0.8, K2equals 1.12 and 1.07, respectively. At the same temperature and respective equivalence ratios,but at 80% relative humidity, K2 equals 0.85 and 0.96 respectively.To calculate the brake specific emissions, the molecular weight of the exhaust must first bedetermined. The molecular weight of the exhaust is given by:MWe . MW1) (3-14)where the species included in this summation are C02, CO, N2, NO, 02, H20, H2, and THC.Since NO consists mostly of NO, the mole fraction and molecular weight of NO are used in thecalculation of the exhaust molecular weight. However, the SAE recommendation is to expressNO as NO2 for the brake specific calculations. Also, THC are analyzed on a carbon atom basis.Assuming the THC have the same hydrogen to carbon ratio as that of the average of the two fuels(i.e. CnHm), then the molecular weight of THC per carbon atom is expressed asMW =12.01+(-.1.01) (3-15)The brake specific emission of a component (bs1) is equal to the mass flow rate of thatemission (m1) divided by the engine brake power (b)bs=-’- (3-16)The mass flow rate of a component is given by35mMW. (3-17)1VlW/where the exhaust mass flow rate (me) is given bymCXh mair + rnG + rnDSL (3-18)where thr and rnDSL are the measured flow rates of air, compressed natural gas and dieselrespectively.3.5 Unburned Mass Fraction of FuelThe unburned mass fraction of fuel is required for the mass fraction of fuel burnedcalculation, which is described in chapter 5. The unburned mass fraction of natural gas is given byXUG = (3-19)mcG + mDSLand the unburned mass fraction of diesel fuel is given bym -thx = THC CH4 (3-OUDSLm(G +364. PREVIOUS COMBUSTION ANALYSIS MODELSThis chapter begins with a description of the combustion phenomena occurring in thecylinder of a diesel engine. A description of basic combustion analysis techniques is then given,followed by an examination of previously developed combustion analysis models intended toidentify their strong points and shortcomings.4.1 The Stratified Charge Combustion ProcessThe stratified charge combustion phenomenon occurring in the cylinder of a diesel engineinvolves complex interactions of physical and chemical processes. A qualitative description of thecomplexities associated with the stratified charge combustion process is given by Edwards,Siebers, and Hoskins as an introduction to their study of the autoignition process of a diesel sprayvia high speed visualization [17]. The stratified charge combustion process can briefly bedescribed as follows.The process begins when fuel is injected with high velocity into the combustion chambershortly before the piston reaches top-dead-center (TDC). The high momentum fuel mixes andbegins to chemically react with the relatively quiescent, but high temperature (9OOK) mixture ofcompressed air and residual combustion products. This mixing process is dependent on both timeand location; hence a competition exists between parcels of fuel and air to see which, if any, willautoignite. Autoignition can be described as the point where the preferred chemical pathways ofthe fuel/air mixture are no longer endothermic or leading to non-participative intermediatespecies, but lead to the release of chemical energy. Sufficient chemical energy must be released toremain exothermic during subsequent mass and energy exchanges resulting from the mixingprocess.37Parcels of fuel and air adjacent to parcels which are autoignition sites exchange mass andenergy, and may become favorable reaction sites themselves. Other sites which are chemicallyless mature, but are suddenly exposed to a flux of heat or species, may be consumed by a wave ofchemical reaction or flame. The rate of combustion for the remainder of the process is dependentupon mixing. Volumetric expansion of the burning parcels generates convective motions which,along with residual motions from the fuel injection and compression-generated swirl, continue tomix reactants and products.In short, the compression-ignition combustion process involves mixing with heat and masstransfer between parcels of varying compositions of species while complex chemical kinetics leadto the release of chemical energy and the production of pollutants. ‘While the combustion processis exceedingly complex on a microscopic scale, the application of simplifying yet realisticassumptions and standard thermodynamics enable the progress of combustion to be investigatedquantitatively using measured engine cylinder pressure dataA thermodynamic combustion model based on measured cylinder pressure data canpotentially be a valuable diagnostic tool used to investigate the performance and emissioncharacteristics of an engine under various operating conditions. If the model can provide insightinto the processes occurring inside the combustion chamber (such as heat transfer, fuel injection,and pollutant formation), then steps can be taken to improve performance and reduce pollutantemissions of an engine.4.2 Thermodynamic Combustion AnalysisA thermodynamic combustion analysis based on measured cylinder pressure typicallybegins with an energy balance on the engine cylinder contents. A schematic of the thermodynamicsystem as applied to the combustion chamber is illustrated in Figure 4-1. The first law ofthermodynamics for this open system for a small crank change interval can be written asoQ—oW=dE+2dm1h (4-1)38dmhmI ControlVolumeboundaryFigure 4-1: Schematic of Combustion Chamber Thermodynamic SystemIgnoring potential and kinetic energy changes in the cylinder gas, the change in total energy dE isequal to the change in total internal energy dU. The flow of mass and energy into and out of thesystem is accounted for in the summation of the differential mass (dm) multiplied by itscorresponding enthalpy (h). The flow of enthalpy into the cylinder is associated with the injectedfuel, while the flow out of the cylinder is due to the leakage past the piston rings or through thevalves. The work transfer from the system, oW, is given byOW = PdVwhere P is the average pressure over a small time increment.To estimate the instantaneous heat-transfer, OQ, to the cylinder walls, an estimate of aheat-transfer coefficient is required. A discussion of estimates of heat transfer to cylinder walls isgiven in the following section. Implementation details of the combustion analysis model hingeupon the thermodynamic description of the cylinder contents and the complexity of sub-models ofcertain in-cylinder processes (such as fuel injection, heat-transfer, pollutant formation, etc.).While the implementation details vary with the application, there are two broad categories ofmethods. The first method treats combustion as simple heating. The second method accounts forchemical reaction. Discussions of these two methods are given in sections 4.4 and 4.5.394.3 Estimating Instantaneous Heat-Transfer to Cylinder WallsAnnand [18] in 1963 reviewed existing formulae for internal combustion engineinstantaneous heat-transfer rates and concluded that the dimensionless experimental convectiveheat-transfer coefficient (i.e. the Nusselt number) could be represented byNu=a(Re)°7 (4-2)with0.35 a0.8Here the characteristic length used in both the Nusselt and Reynolds numbers is the cylinder bore,and the characteristic velocity used in the Reynolds number is the mean piston speed. Thekinematic viscosity is evaluated at the mass-averaged gas temperature.Woschni [19] in 1967 introduced a new method for determining the heat-transfercoefficient for internal combustion engines based on an overall energy balance of a four-stroke,open-chamber diesel engine. He suggested that the characteristic gas velocity used in theReynolds number should account for the convection resulting from both piston motion andcombustion. Woschni determined that the Nusselt number could be expressed asNu = 0.035(Re)°8 (4-3)the exponent of 0.8 being the same as for turbulent fluid flow in pipes.Woschni also used the cylinder bore as the characteristic length in both the Nusselt andReynolds numbers, but expressed the characteristic velocity asw=Clcm+C2cb 4-4)Here C1 and C2 are constants, cm is the mean piston speed and Cb is a gas velocity which can beattributed to the burning process. This gas velocity is estimated by40VTcbr rwhere V, P, and T are the volume, pressure and temperature, respectively. The subscript r refersto a convenient reference point such as inlet closure or beginning of combustion, and (P - P0) isthe pressure difference between fired and non-fired cycles. The constants C1 and C2 depend oncylinder geometry and are selected such that approximately 85% of the total gas velocity resultsfrom piston motion, with the remaining 15% resulting from combustion intensity. Due tocomplexities associated with temperature gradients and fluid motion in the combustion chamber,this heat transfer estimate is expected to have a relatively high degree of uncertainty.Radiation heat transfer inside the combustion chamber should also be considered.Radiation in diesel engines consists of luminous (soot) and non-luminous (gaseous) radiation [20].Soot particles formed during combustion assume almost the same temperature as the gases in theflame. If the soot concentration is high enough, this cloud of particles radiates like a solid body.Soot emissivity is dependent on the soot concentration and the thickness of the soot cloud [19].The non-luminous radiation from the combustion gases is primarily due to emission contributionsfrom the tri-atomic molecules of carbon dioxide, and water vapor.As reported by Szekely and Alkidas [20], the contribution of radiation to the total heattransfer in conventional diesel engines is significant. Their analysis shows that in the case of anopen-chamber diesel engine, the ratio of radiation to total heat transfer could be 24% at full load.They also reported findings from other researchers; Ebersole et al. found that the radiationcontribution in a two-cycle open-chamber diesel to increase from 5 to 40% as load increased;Oguri and Inaba found that the maximum contribution of radiation to be about 30%; and Dent andSulaiman indicated that radiation was under 20% of the total. Szekely and Alkidas also reportedthe non-luminous radiation to be negligible compared with the soot radiation which wasconfirmed by other researchers.Since radiation heat transfer is only significant with high soot concentrations, no errorshould be incurred by ignoring radiation heat transfer calculations when no smoke is present in the41engine exhaust. Smoke levels at part loads in diesel engines are low. Smoke levels in natural gasfueled diesel engines are also low.4.4 Simple-Heating Combustion AnalysisRassweiler and Withrow developed a simple-heating model in the 1930s [21]. It is easyto implement and requires only cylinder pressure measurements and corresponding cylindervolume details as inputs. Their analysis is based on observations from combustion ofhomogeneous mixtures in constant-volume bomb experiments.In constant-volume bomb experiments with combustion of homogeneous mixtures, there isno work done and no flow of enthalpy into or out of the bomb during the combustion event.Hence the first law of thermodynamics (equation 4-i) applied to a constant volume bomb reducestoQ=AU (4-5)The release of chemical energy during combustion is thought of as a simple heat addition process.Assuming that the heat transfer to the walls of the bomb during combustion are negligible, thenthe heat transfer, Q, is a fictitious heat addition.Using the ideal gas law, the change in internal energy, AU, can be written asAU= mCAT=.y -i(4-6)y -1where V is the volume of the combustion bomb, ‘ is the ratio of specific heats, and AP is thechange in pressure in the bomb. Combining equations (4-5) and (4-6) gives42(4-7)‘ -1Assuming that the equivalent heat addition, Q, is proportional to the mass of fuel burned, that theratio of specific heats is constant, and that there is no change in the molecular weight of themixture during combustion, then it follows from equation (4-7) that for a constant volume bomb,the mass fraction of burned fuel at some point in time can be written asX=’ (4-8)In this expression, P is the measured pressure, P1 is the initial pressure, and P is the maximumpressure achieved during combustion. For a given amount of heat addition, 0, it also followsfrom equation (4-7) that pressure rise due to combustion is inversely proportional to volume.(4-9)In the combustion chamber of an engine however, the pressure change is a result of thereciprocating piston motion as well as combustion. Assuming these two effects on pressure canbe considered independent then the following expression can be written.combustion = APmeasured -A1piston (4-10)Assuming that the compression and expansion processes are polytropic, then the pressure changedue to piston motion AP0 between states 1 and 2 can be expressed by:n=- P1 (4-11)43where n is the polytropic exponent obtained from the pressure data. Since APmeed = P2 - P1. thenequations (4-10) and (4-11) can be combined to give:n= P2-(4-12)If states 1 and 2 are considered respectively to be the beginning and end of successivesmall crank angle intervals, then the change in pressure due to combustion can be calculated atevery crank angle between the beginning and end of combustion. However, since each crankangle corresponds to a different volume, to apply the analogy of equation (4-9) for the constantpressure bomb these pressures need to be referenced to a common volume (for exampleHence the mass fraction of burned fuel in the engine cylinder at some crank angle will be the sumof the pressure changes due to combustion (referenced to a common volume) divided by the totalsum from beginning to end of combustion of these pressure changes as follows:8 /(AP1) 0o \\VTDCXmf = BOC / (4-13).V0°BOCVTDCThe subscripts BOC and EOC refer to the beginning and end of combustion, respectively.While this method is widely used, it suffers from the following drawbacks:1. Although heat-transfer effects are accounted for to the extent that the polytropic exponent ndiffers from the isentropic exponent, the polytropic exponent is difficult to estimate because itis not constant during combustion.2. The calculation of the mass fraction of burned fuel depends strongly on a measure of theEOC. Since the EOC is difficult to obtain accurately from pressure data, the calculation ofmass fraction of burned fuel at a given crank angle can be quite uncertain.443. The ratio of specific heats does not remain constant during combustion, which is inconsistentwith equation (4-13).4. The energy changes of the cylinder charge represented by the pressure-time history of theengine are not explicitly determined; hence spatially non-uniform properties (i.e. burned gastemperatures) and pollutant-formation rates can not be computed.Gatowski et al. [22] in 1984 developed, tested, and applied a one-zone combustion rateanalysis procedure with the aim to maintain simplicity while including “all phenomena ofsignificance”. A one-zone model describes all of the cylinder contents in terms of averageproperties. The cylinder contents are treated as uniform and homogeneous, with no distinctionmade between burned and unburned gases. The heat transfer term in the first law energy balance(equation 4-1) is treated as the difference between the simple heat addition from combustion andthe heat transfer from the cylinder gases to the combustion chamber walls. Using a heat transfermodel and a representation of the sensible internal energy change of the cylinder contents, theheat release rate can be calculated. Spatially non-uniform properties (i.e. burned gastemperatures) can not be computed since only a single zone is used.The ratio of specific heats (y) is the primary thermodynamic property representing thecylinder contents. The variance of ‘ with temperature was approximated by a linear fit which isevaluated at the bulk cylinder temperature. The correlation used for calculating cylinder heattransfer is based on the form proposed by Woschni. Details of the fuel injection event were notincluded in this analysis. For simplicity, the fuel mass was considered to be vaporized andpremixed with the combustion chamber air at the beginning of the analysis. Finally, crevice effectswere accounted for by using a simple modeling assumption. Crevice effects are the flow of gasinto and Out of crevices in the combustion chamber which constitute a couple of percent of theclearance volume.4.5 Chemical Reaction Combustion AnalysisInstead of treating combustion as a simple heat addition process, it would be moreaccurate to describe the release of chemical energy resulting from a change in composition from45unburned reactants to burned products. This can be represented thermodynamically byconsidering the differences in energy between the unburned and burned gases. The heat transferterm in the first law energy balance (equation 4-1) then strictly accounts for heat transfer from thecylinder gases to the combustion chamber walls.Krieger and Borman [23] presented a one-zone thermodynamic combustion rate model in1966. Their treatment assumes that the entire cylinder contents are made up of a homogeneousmixture of air and combustion products which are in thermodynamic equilibrium at each instant.The products of combustion are assumed to be the result of the oxidation of CH2 in air. Datafrom the JANAF tables of thermodynamic properties are used to obtain mathematical expressionsfor the internal energy and gas constant as functions of temperature, pressure, and equivalenceratio. Dissociation of the combustion products is ignored. The release of chemical energy isrepresented by changes in the internal energy and composition of the cylinder contents withpressure and temperature.Phenomena such as temperature gradients, pressure waves, nonequilibrium compositions,fuel vaporization, mixing and so on are ignored in this model. The instantaneous heat-transferfrom the gas is computed from the Annand correlation. Five metal-surface areas representing thehead, piston, sleeve, and valves are each assigned a different constant temperature. These surfacetemperatures must be either estimated from experimental data, or computed by the use of a cycleanalysis program.Burning is assumed to take place incrementally. Fuel is assumed to be introduced into thecombustion chamber at the same rate that it is consumed. Thus unreacted fuel in either liquid orvapor phase is never present. The equations of energy and mass continuity together with theequations of state and the mathematical expressions for internal energy and the gas constant areintegrated to obtain the mass of fuel burned during each crankangle increment.In 1985 Bedran and Beretta [24] discuss the elements of a generalized multi-zonethermodynamic analysis method that can be applied to any level of modeling (zero-dimensional,quasi-dimensional, or multi-dimensional) and any type of internal combustion engine(homogeneous-charge, direct injection, indirect injection, etc.). The model uses a mass and46energy balance to determine the instantaneous mass fraction of burned gas mixture in differentialequation form. The only assumptions made are as follows:1. The mass of gas which is instantaneously burning is negligible.2. The pressure is uniform throughout the control volume.3. The burned and unburned gases behave as Gibbs-Dalton mixtures of ideal semi-perfect gases(i.e. with temperature dependent specific heats).4. The liquid fuel in the combustion chamber behaves as an incompressible fluid so that specificvolume and internal energy depend on temperature only.Before any useful information can be attained from this analysis, it must be complementedby a set of modeling assumptions to account for such effects as fuel distribution after injection andvaporization, wall heat transfer, mass exchange with crevice volumes, mixing between unburnedand burned gases, non-uniform distributions in temperature and composition and so on. While anapplication of the analysis was not developed and tested, Bedran and Beretta described how onemight set up the problem of estimating, from cylinder pressure measurements, the mass fraction ofburned gases as a function of time in a diesel combustion chamber with no residual burnedfraction.In this application the control volume coincides with the walls of the combustion chamberand a number of zones are defined. There are two burned gas zones; one corresponding tocombustion at the cylinder core temperature and the other at the crevice temperature. There areN+2 unburned zones which include a liquid fuel zone, unburned fresh charge at crevicetemperature, and N gaseous zones with increasing fuel concentrations. Assuming enoughindependent models (i.e. vaporization, heat transfer to the walls, etc.) are available, a step by stepsolution scheme starting from an instant in time at which all the variables are known can be usedto calculate the mass burned fraction. Each stepwise estimate of the mass fraction burned must beiteratively refined. Szekely and Alkidas [20] in 1986 presented a two-stage two-zonecombustion analysis model for diesel engines. The two stages refer to two distinct modes ofburning. In the first stage, combustion occurs at a stoichiometric equivalence ratio. In the secondstage, combustion occurs at an equivalence ratio below the cylinder-averaged equivalence ratio47(i.e. fuel lean). These distinct burning modes are arbitrarily chosen such that the calculatedexhaust NO corresponds to the measured value.Table 4-1 Szekely and Alkidas Combustion Analysis Model(burning stages and zones)unburned zone ( = 1)Stage I unburned zone ( < p)burned zone (( = 1)unburned zone (q: < py)Stage II burned zone ( = 1)burned zone (( < 4avu)Each burning stage has a corresponding burned and unburned zone, for a maximum offour zones. However, only three of the four zones are actually active at any one time. The zonewhich begins to react first is completely consumed prior to reaction of the second zone, and nomixing occurs between unburned zones or between the burned and unburned zones. Table 4-1lists the corresponding zones for each burning stage. Each zone contains a homogeneous mixtureof air, fuel and residual combustion products in thermodynamic equilibrium. Temperaturegradients within the zones are ignored as is flow leakage through the valves and past the pistonrings. Uniform pressure within the cylinder is assumed and the ideal gas law is assumed to bevalid for each zone.The equivalence ratio of the second stage is empirically derived by comparing calculatedNO emission levels to those directly measured in the exhaust. Szekely and Alkidas deduced that67% to 75% of the fuel is burned in the second stage at an equivalence ratio of about 0.1 belowthe cylinder averaged value. The NO formation is described by the extended Zeldovichmechanism (described in chapter 5). Convective and radiative heat transfer are also accounted foras is the injection of liquid fuel. The convective heat transfer is described by Woschm’s empiricalequation.48When comparing the results of this model to a single stage model, Szekely and Alkidasfound that in the initial stages of combustion, the calculated rate of heat release was higher thanthe corresponding rate calculated by the single stage model.’ They also found the computedcombustion temperatures for the stoichiometric burning stage to he consistently higher (7 to 22%on an absolute scale) than published flame temperature measurements. Model calculations alsosuggest that NO-decomposition reactions are insignificant which is contrary to some experimentalfindings.Gunawan [81 in 1992 presented results from a two-zone combustion rate model to analyzediesel engine combustion when fueled with direct-injection natural-gas with diesel pilot. The twozones he considered were an unburned and burned zone. In this model the key assumptions are asfollows:1. The cylinder pressure is uniform.2. Both the unburned and burned zones behave as ideal gases.3. The combustion occurs at the cylinder averaged equivalence ratio.4. The combustion products are in thermodynamic equilibrium.5. All of the fuel is present in the cylinder at the beginning of injection.6. The mass burned fraction of diesel fuel and natural gas are the same at any instant duringthe combustion process (i.e. proportional burning of natural gas and diesel).7. The vaporization of the liquid diesel fuel occurs instantaneously with allowance madefor heat of vaporization.Heat transfer to the cylinder walls is ignored in this model which requires the calculated massburned fraction curve to be normalized so that the final calculated mass burned fraction agreeswith the measured exhaust composition.4.6 Limitations of Previous Combustion Analysis ModelsThe combustion analysis techniques presented in this chapter have been classified into twobroad categories: a) simple-heating models and b) models that account for chemical reaction.tDetails of the single stage model were not given.49While the latter are more tedious to implement than the former, they represent the thermodynamicproperties of the cylinder contents more accurately.A key feature of a combustion rate model is how the contents of the cylinder aredescribed. One-zone models describe all of the cylinder contents in terms of average properties.The cylinder contents are treated as uniform and homogeneous and there is no distinction madebetween burned and unburned gases. The advantages of a one-zone model are that the effects ofheat-transfer and gas flow phenomena are simple to describe and the combustion can beconsidered either as a separate heat addition process or a chemical reaction process. Averagecylinder temperatures, however, can not be used for modeling pollutant-formation processes (i.e.NO1)due to the high sensitivity of these processes to gas temperature.To permit more accurate treatment of thermodynamic properties of the cylinder gases,different zones can be considered. This approach is referred to as multi-zone modeling. Anynumber of zones can be specified. Each zone has uniform and homogeneous properties and iseffectively treated as a separate thermodynamic system. While the implementation of a multi-zonemodel is more difficult than its one-zone counterpart, realistic gas temperatures can be computed.The number and complexity of modeling assumptions used in a combustion model willhave an effect on its accuracy. Modeling assumptions describing radiation and convection heattransfer, fuel injection, crevice flows, pollutant formation, etc. can be incorporated. The accuracyof the results from these modeling assumptions depend not only on their own level of complexityand accuracy, but also on the accuracy of the representation of the thermodynamic properties ofthe cylinder gases.The accuracy with which the model represents stratified-charged, compression ignitioncombustion versus homogeneous-charge, spark-ignition combustion will also affect the results.Most of the models presented in this chapter are based on models describing homogeneouscharge combustion. While homogeneous-charge combustion can be accurately described usingonly an unburned and burned zone, the high degree of non-uniformity in composition andproperties in stratified-charge combustion require a greater number of specified zones for proper50treatment. The basic models needed to simulate the stratified-charge diesel combustion processand the homogeneous-charge spark-ignition combustion process are quite different.While the model developed by Szekely and Alkidas distinguishes homogeneous-chargefrom stratified-charge combustion, it does not deal with temperature gradients in the burnedgases, or mixing between burned and unburned gas zones. Both of these factors have a significantinfluence on the accuracy of the predicted burned gas temperature. Since diesel engines operatewith 50% to 400% excess air with turbulent flow inside the combustion chamber, it seemsreasonable that the burned gases will not remain independent of unburned gases for the entireduration of combustion. Some mixing between burned and unburned gases must take place.A problem immediately comes to mind when specifying a single zone to describe all of theburned gases. As the combustion process involves mixing, non-uniformities in the properties ofthe burned gases formed at different times and locations in the cylinder would have to beexpected. Temperature gradients, for example, exist in the burned gases because products ofcombustion formed early in the combustion process are compressed to a higher temperature ascombustion proceeds and cylinder pressure increases, while the last of the reactants to burn arecompressed prior to combustion with no compression afterward.An approximation to more effectively deal with the non-uniformities in the burned gases isto use many distinct burned gas zones, each with unique and independent properties. Instead ofthe size of a single burned gas zone growing as combustion proceeds, a new burned gas zone canbe created during each calculation step (generally one crank angle). The size of each zonetherefore depends on the instantaneous combustion rate. Once formed, the properties of eachburned gas zone are updated with changing cylinder pressure for subsequent calculation steps.The question which now arises is: should each of these burned gas zones be independentof each other and the other zones in the cylinder throughout the remainder of the combustionprocess’? The answer is no. Since the combustion in a diesel engine involves mixing of fuel withexcess air, one would expect that burned gases formed at the flame front either diffuse away orare swept away by turbulence and then forced to mix with either unburned gases or other burnedgases.51There is an important need for a combustion model that deals with the non-uniformities inthe cylinder of a stratified-charge internal-combustion engine. A more accurate treatment of thestratified-charge combustion process will give more accurate pollutant-formation, burned gastemperature, and fuel burning rate predictions. A model which addresses non-uniformities in thecombustion chamber has been developed and tested as part of this thesis. A detailed descriptionof this model is given in the following chapter.525. PROPOSED THERMODYNAMIC COMBUSTION ANALYSIS5.1 Formulation of Combustion ModelThe combustion analysis model developed here uses a mutli-zone description of thecylinder contents. As the intended application of this model is to investigate fuel burning rates,non-uniformities of cylinder gas properties, heat transfer, nitrogen oxide formation, and fuelinjection for both diesel and natural gas fueling regimes, appropriate modeling assumptions areused. A conceptual diagram of the combustion analysis model is given in Figure 5-1. Thefigure illustrates the different zones and their interaction with one anotherFigure 5-1 : Conceptual Diagram of Proposed Combustion Model53By definition zones have homogeneous and uniform thermodynamic properties, thus thekey to the model formulation is how the zones are defined. The diesel fuel zone consists of theinjected liquid diesel fuel. The natural gas zone consists of the injected compressed natural gas.The unburned zone consists of a mixture of fresh intake air and residual combustion productsfrom the previous cycle. The burned gas zones consist of the products of combustion betweenthe fuel and the unburned zones. The ‘mixed” zone consists of a mixture of burned andunburned gases.A new burned gas zone is formed during every calculation step. It remains independentwith its properties changing with cylinder pressure for only a specified “mixing delay” period.At the end of the mixing delay period the burned gas zone mixes out with a proportional amountof unburned zone in the mixed zone. This formulation is somewhat empirical in that the mixingdelay period and the proportion of unburned gas that mix with the burned gas zone must beselected.Application of conservation of mass and energy to the engine cylinder serves as thestarting point in the combustion analysis. If m,, mg, m, m, and mm are the instantaneousmasses of the diesel, natural gas, unburned gas, burned gas and “mixed” gas zones respectively,then the total mass in the cylinder at a given time ism0 = md +mg +m -f-mb +rnm (5-1)Since the burned gases are represented by a number of zones, the total mass of the burned gases(mb) is written as mbI. If there is no leakage past the piston rings, then the only change in thetotal cylinder mass once the intake and exhaust ports are closed is due to injection of fuel. Thetotal mass in the cylinder prior to fuel injection is the mass of air and residuals trapped at intakeand exhaust port closure (i.e. entirely unburned gas zone).Using the same subscripts as above, the total cylinder volume at any given time in thecombustion process is54or=V + Vg + + + VmV1= mdvd + mgvg + mv + mmvm + (mbvb)1 (5-2)where v is specific volume and k is the number of burned gas zones present at any time. Thetotal cylinder volume can be determined as a function of crank angle from cylinder geometry (ascalculated in Appendix VII). The specific volumes of each zone are functions of temperatureand pressure. In a similar manner, the total energy of the cylinder contents can be written asorEcyI=Ed+Eg+Eu+Ebi+Em= mu + mgug + muuu + mmum + (mbub) (5-3)where u is the specific internal energy.Applying equations (5-1), (5-2) and (5-3) at every crank angle forms the basis of thecombustion analysis. To proceed at this point, the following assumptions are made:• Pressure is uniform throughout the cylinder.• The combustion chamber gases exhibit ideal gas behavior.• All of the diesel fuel burns before any of the natural gas (referred to as the sequential burningassumption).• The composition of the unburned zone does not change during combustion.• Natural gas and diesel fuel are assumed to be fuels composed of only one chemical species;pure methane (cH4)represents natural gas and CH18 represents diesel fuel.• Burned gas zones result from stoichiometric combustion.55• The composition of the residuals consists only of oxygen, nitrogen, carbon dioxide and watervapour.• The composition of air is 21% oxygen and 79% nitrogen.• NO is the predominant oxide of nitrogen inside the combustion chamber and is formed in thepost flame gases by the extended Zeldovich mechanism.The computation begins at the beginning of injection (BOI). At BOI, equations (5-1), (5-2) and (5-3) are greatly simplified since only one zone, the unburned gas zone, is present. Asthe cycle proceeds, the changes in properties of each zone can be calculated independently.Then by using a combination of equations (5-1), (5-2) and (5-3), the properties of the mostrecently formed burned gas zone can be found in one of two ways.The first method for determining the properties of the most recently formed burned gaszone requires an estimate of the heat transfer to the cylinder walls. Based on this heat transferestimate, the change in total energy from one crank angle to the next can be calculated from afirst law energy balance as described in the previous chapter (section 4.2). Woshni’s convectiveheat transfer correlation is used to calculate the total heat transfer to the cylinder walls.Radiation and convection are not treated separately, because the total heat transfer prediction isuncertain.Rearranging equations (5-2) and (5-3) with m from (5-1) givesV1—mg(vg vU)md(vd vU)+vUmb 2(mbvb) fflm(Vm —vs)mtot m0 m0— mg(ug — un)— md(ud — u)+ uUmbI — (mbub)I — — u) k mki=1mtot m0 m056The superscript k denotes the most recently formed burned gas zone. For convenience, lettingthe left hand sides of the above equations equal the variables v and u, respectively, furtherrearrangement givesye—vu =-L(v_v) (5-4)m0(5-5)from whichmbvCvU UU (5-6)VbVU UbUUSince the specific volumes and internal energies are functions of temperature and pressure, aniterative approach must be used to calculate the mass of the most recently formed burned gaszone, mt, from equation (5-6) as follows. LettingA = v0 — v = — vuc—uu u—uuthenv—v = A(u— u)or= Au — Au +IfB =—Aug + vthenv=Au-i-B (5-7)57A and B can be directly determined at the end of each time step since they are functionsonly of the unburned gas properties and the variables v and u which are determinedindependently. With A, B and the pressure known, a guess for the burned gas temperature ismade. From tabulated burned gas properties, Vb and u, can be found. Each temperature guessis refined until equation (5-7) is satisfied; this signifies convergence. With the burned gasproperties and the total mass of cylinder contents known, equation (5-6) is used to calculate themass of the most recently formed burned gas zone.The second method to determine the properties of the most recently formed burned gaszone requires an estimate of the burned gas temperature. The temperature of a small burned gaszone burned in a short time interval can be approximated by the adiabatic constant-pressureflame temperature. Once the burned gas temperature is known then the other burned zoneproperties can be found from tabulated values since the pressure is also known. By combiningequations (5-1) and (5-2), the mass of the most recently formed burned gas zone can becalculated as follows:Vc—mg(vg—vu)—md(vd—vu)+vumbj—(mbvb)j—mm(vm—vu)m=In this case, since equation (5-3) is not required for determining the mass or properties of themost recently formed burned gas zone, it can be used to calculate the total energy of the cylindercontents. A first law energy balance can then be used to determine the total heat transfer to thecylinder walls.5.2 Calculation of Mass Fraction of Fuel BurnedOnce the mass of the most recently formed burned gas zone is known, the cumulativemass fraction of fuel burned (Xmf) can be determined. First a formal definition of X isrequired. Let m1 and m be defined respectively as the total mass of diesel and natural gas58injected into the cylinder for one complete cycle. Further, let mlbm, mgm and mthm bedefined respectively as the cumulative mass of diesel, natural gas and air burned at some point intime during the combustion process. Then X is defined as follows,—+ mgmmf+ mBy expressing the cumulative mass of burned gas m in terms of its constituents, an expressionfor X can be found as follows:m= mlbm + mgbm + fllj= (m1 + mg(Xmf +mi+mg)f ‘bm m1 += kmdSl + m X,,, +gmIbffl + m8b m1 + mgRearranging and solving for X givesXmf= (5-8)mdslbm +where AF and AF are the respective air/fuel mass ratios for diesel and natural gascombustion as determined from stoichiornetry. Since the mass of diesel and natural gas burnedcan not be found explicitly, to arrive at a solution for equation (5-8) a key assumption about thecombustion process must be made. It is assumed that all the diesel fuel burns before any naturalcum59gas burns. This assumption, referred to as the sequential burning assumption, is valid for tworeasons. First, only a small amount of diesel fuel is injected and it is injected prior to the naturalgas. Second, the autoignition point of natural gas is not achieved in the combustion chamber bycompression alone at typical diesel engine compression ratios. Since the autoignition point ofdiesel fuel is reached under these conditions, it would be the first the bum.For the initial diesel-only burning phase of the combustion processm10 Xmf <m1 + mHere mm = 0 thus equation (5-8) reduces tommX=f b (59)j mI + mg A’ + AF)For the subsequent natural gas-only burning phasem1Xmtlm1 + mFor X,11 in this range, mlbffl = m1 since all the diesel has combusted. The mass of natural gasburned can be expressed bymm =Xmf(mi+mg)—ml (5-10)Substituting equation (5-10) into equation (5-8) gives60cumxm,mf=mlAl+[Xmf(rnthJ +mg)_mdsI]1+mgl+ /m +Xmfkml +m)—mthIxm,mf— m1 + mg + _-(miMi +[Xmf(mthl + m)_ mi]A1)xm,,— m + m + — A1)+ (m1 + mg)A1Xfflf(mJSI + m)(1+ = m —m1(M — Ai)m”1— mI(AFdSI — AFS)Xmf= (5-li)(m +m)(1÷M)To proceed with the results of equations (5-9) and (5-11) it is necessary to specifymethods for determining the properties of each zone as well as the mass of air and residualstrapped in the cylinder. Before the properties can be determined however, the composition ofthe burned gases must be established by the combustion stoichiometry. Hence the followingsections give detailed descriptions of the combustion stoichiometry, calculation of trapped airand residual mass, and the determination of the zone properties.5.3 Combustion StoichioanetryNear top dead center (TDC) before injection of fuel, the combustion chamber is filledwith a mixture of fresh intake air and residual combustion products from the previous cycle.This is identified as the unburned zone and is assumed to have an unchanging compositionthroughout the combustion process. At the beginning of injection (BOl), jets of fuel issue intothe cylinder from a series of orifices in the injector tip. With propagation of the jet, entrainmentand mixing with the air and residuals occur. The jet has the structure of a fuel-rich core61surrounded by a progressively leaner mixture of fuel and air. Experiments where theautoignition sites were recorded showed that autoignition occurred in a concentration bandbetween the equivalence ratios of 1 and 1.5. Subsequent flame development was seen to occuralong contours close to stoichiometric [1]. Consequently, it is reasonable to assume that burnedgas zones result from stoichiometric combustion.It is important to note that the fuel reacts with a mixture of air and residual combustionproducts. The composition of the residuals is assumed to consist only of oxygen, nitrogen,carbon dioxide and water vapour since the concentrations of other species such as carbonmonoxide and oxides of nitrogen are comparatively small. At the flame front, as chemicalenergy is released from oxidization of the fuel; nitrogen, carbon dioxide and water vapor diffusethrough the flame, acting as diluents to lower the flame temperature. If natural gas and dieselfuel can be assumed to be fuels composed of only one chemical species, then one simplecombustion reaction can be written. Using ‘y” to denote the mole fraction of the constituents ofthe air-residual mixture, for complete combustion the reaction can be written asCHfl+A.(y02.02+yNNHO.11+ COC0)(1+Ayc)CO+(+AyH2o).110+A.yN . N2Balancing oxygen atoms for stoichiometric combustion gives1+A =y02Thus the combustion reaction can be written as62YH2Oy02 y02 - y02(i+(i+).i).co2+(+(i+). °)H2o+(1+). N2Since this reaction assumes a single species fuel, the single species representing naturalgas and diesel must be identified. While the composition of natural gas varies with geographiclocation, it is not uncommon for it to consists of upwards of 95% methane (CH4). Thus toassume that natural gas is pure methane is reasonable. Diesel fuel, on the other hand, can consistof up to 100 hydrocarbons and 100 to 200 trace species, and can show significant variations inproperties with time, geographic location, and intended service [25j. Thus representing dieselby a single chemical species may lead to some inaccuracies. The best estimate of a singlechemical species for diesel is obtained by looking at a typical chemical analysis that reveals itshydrogen-to-carbon atom ratio. The hydrogen-to-carbon ratio typical of the diesel fuel use forthese experiments was found to be around 1.8; thus CH18 is used to represent diesel fuel.Furthermore, to use the above reaction to determine the composition of the burned gases,the composition of the unburned mixture with which the fuel reacts must be known. This mustbe determined iteratively along with the moles of residuals trapped in the cylinder. Thisiteration procedure is described in the following section.5.4 Trapped Air and ResidualsTo determine the mass of air and residual combustion products trapped in the cylinder,the cylinder gas exchange process during scavenging and the in-cylinder mixing after inlet andexhaust port closure must be investigated. The temperature and composition of the trappedgases must also be determined since they are important for heat transfer calculations anddetermination of gas properties. A look at the fluid mechanics and thermodynamics of the63cylinder contents will give insight to the physics of this problem, and with a few assumptions acalculation procedure can be developed.5.4J Scavenging ProcessUniflow scavenging is the configuration used in both the DDC 1-71 and 6V-92TA. Inthis configuration inlet ports, evenly spaced around the circumference of the lower part of thecylinder, are used to direct the incoming air such that swirling flow is created in the cylinder.Exhaust valves located in the cylinder head allow the exhaust flow to exit. During scavenging,air delivered from the blower or compressor of the turbocharger is used to displace thecombustion products from the cylinder. It is the question of how well the air displaces thecombustion products that must be answered. Short-circuiting of fresh air and dead volumes ofcombustion products were found to occur in flow visualization experiments in liquid analogs ofcylinder flow [1]. Using these observations, the following simplified model was used todescribe the scavenging process.With inlet and exhaust ports open, incoming fresh air forces all the combustion productsout of the cylinder except for those found in the dead volumes. The flow into the exhaustmanifold therefore consists of the discharged combustion products and the excess air that eithershort circuits through the cylinder or follows the combustion products Out. Adiabatic mixing ofthe air and combustion products is assumed to occur downstream of the exhaust port at constantpressure such that the exhaust manifold mounted thermocouple measures the mixed outtemperature. Upon closure of the inlet and exhaust ports, it is assumed that the fresh air andresidual combustion products contained in the cylinder mix adiabatically at constant volume.Figure 5-2 schematically describes the mass flows of the entire process. Here m is thedelivered air (mass flow as measured in the engine intake), m is the mass of air trapped in thecylinder, m0 is the initial mass of the unburned gas zone, f is the fraction of unburned gaswhich participates in the combustion, mre, is the mass of residuals in the cylinder, rn is themass of diesel injected, m is the mass of natural gas injected, rnbumed is the final total mass ofthe burned gas zones, and is the mass of the exhaust leaving the engine.64m11 mmrn m burnedma1. m atrap m00 (1-f) m0 ÷ m m atrap m1 ex11m4maff - m atrapFigure 5-2: Engine Mass Flow Schematic5.4.2 Mass of Trapped AirThe mass of air trapped in the cylinder after intake and exhaust ports are closed iscalculated using the following correlation [26] from Detroit Diesel based on their measuredscavenging results. The values calculated with this correlation match values typical of uniflowscavenged engines published by Heywood [1].matmp = m1 [0.9— (i.o —0.34 e1h11] (5-12)for0.5<R8<1.8where= p11 VR m1/scav— /The density of the air in the air box (p) is calculated using the ideal gas law with air boxtemperature and pressure. V is the displacement volume of one cylinder. All masses haveunits kg/cycle/cylinder.The relationship between the mass of air trapped and the mass of delivered air providedby equation (5-12) is illustrated in Figure 5-3 using the ideal mass (m1) as defined above to650.90.850.80.750.70.650.60.550.50.450.4equation (5-12)0.5 0.7 0.9 1.1 1.3 1,5 1.7rn/rnair idealnon-dimensionalize both variables. Once the mass of air trapped in the cylinder is known, aniterative approach is required to determine the mass, composition and temperature of theresiduals.matrapm.idealFigure 5-3: Scavenging Correlation5.4.3 Mass of Trapped ResidualsTrapped air and residuals are assumed to mix adiabatically at constant volume at intakeand exhaust port closure (ipc). Since ipc represents the beginning of compression, the cylinderpressure (P) and trapped air temperature (Tar) at ipc can be approximated as the measuredairbox pressure and temperature. The cylinder volume at ipc (V) can be calculated fromcylinder geometry knowing when ipc occurs. Hence, the first law of thermodynamics for thisinstantaneous mixing process on a molar basis can be written asUmix = n Uatrap +flres Ures (5-13)but u = C T for an ideal gas; thus (5-13) becomes66(iap + ires)• vmix = t1atlap Cajr Tatp + n Cvres (5-14)where n and n are the number of moles of fresh air and residual combustion products,respectively, trapped in the cylinder. The ideal gas law written for the mixture is(5-15)Solving (5-14) and (5-15) for T and equating givesPV = n v +n Cvres T (5-16)R Cwhere=+ (5-17)n+nCombining equations (5-16) and (5-17) and rearranging produces•j n +[nap (vres . T + CVr )— n— I p.\\+nap CvaiHflatiap Ta — )=O (5-18)Note that equation (5-18) is a quadratic in the moles of residuals. Since the moles,temperature and specific heat of the trapped air are known, equation (5-18) can be used tocalculate the moles of trapped residuals once the specific heat and temperature of the residualsare determined. Due to the dependence of the specific heat of the residuals on composition andtemperature, equation (5-18) must be solved iteratively. The following sections describe thecalculations used to estimate the residual gas composition and temperature in this iteration.675.4.4 Residual Gas CompositionIn a diesel engine the amount of air trapped in the cylinder exceeds the amount requiredto react with all of the injected fuel. In other words, not all of the air trapped in the cylinderparticipates in the combustion. The residuals are therefore composed of a mixture of theproducts of stoichiometric combustion and the unburned gas zone that did not participate in thecombustion from the previous cycle. Since the air trapped in the cylinder and the amounts ofdiesel and natural gas injected in one cycle are known, the moles of each of the constituents ofthe residuals can be found. From the combustion stoichiometry worked out in section 5.3, forevery mole of fuel burned (1+n/4) moles of 02 are consumed and one mole of CO2 and n/2moles of 1120 are produced.1 Assuming the composition of air is 21% oxygen and 79%nitrogen, the total moles of each species in the cylinder after complete combustion (ignoringdissociation) are given byn0 =0.21n +y0 . —L45n1—2nN2 =0.79n +YNflS=+ +H2O = Y112O, +0.9n1÷2nwhere y02, YN2r’ 02r’ and YH2O, are the mole fractions of the residual constituents iterativelycalculated. The iteration begins by setting the composition of the residuals equal to that of pureair (i.e. y0 = 0.21, YN2 = 0.79, y = 0, and YH2O = 0) and then proceeds using the aboveexpressions to calculate new mole fraction estimates until convergence is achieved. Generallyfive or six iterations are required for convergence.5.4.5 Residual Gas TemperatureThe residual gas temperature can be calculated by considering the flow out of thecylinder. The flow into the exhaust manifold during scavenging is made up of the discharged1n=1.8 for diesel and 4 for natural gas68combustion products and the excess air that either short circuits through the cylinder or followsthe combustion products out. It is assumed that the scavenging air and the products ofcombustion mix adiabatically downstream of the exhaust port at constant pressure in the exhaustmanifold. The thermocouple mounted in the exhaust manifold measures the mixed temperature,while the scavenging air is assumed to be at airhox temperature.From Figure 5-2 during scavenging a mass balance in a control volume in the exhaustmanifold just downstream of the exhaust valve isme)th = (m — m) ÷ (m + m1 + m)ormexh = mr + m + m(1 (5-19)Considering the same control volume in the exhaust manifold, an energy balance of the constantpressure mixing process would be(map + mg + mdSI) +(majr — map) hair = mexh (5-20)but h = C, T for an ideal gas; thus (5-20) becomes(map + m + m1) C + (ma — map) Cpair T = mexh C0 xh (5- 21)whereCpexh=[(ma + m + mdSL) C ÷(m — map) CJ (5- 22)exhCombining (5-21) and (5-22) gives69= xh+(X-(5-23)The specific heat of the residuals is a function of the temperature and composition of theresiduals. Thus equation (5-23) is used to improve the estimate of the temperature of theresiduals using the best estimate of the specific heat for that iteration.5.5 Thermodynamic Properties of the Unburned Gas ZoneIn the previous section the mass of trapped air and the mass and composition of residualswere calculated. Using this information the total moles of unburned zone prior to combustioncan be written asn0 = O.2lna + YO2, aresN2 _°.79nauap +YN2rflIeSn02rn0= YH2Or areswhere Yo2, YN2’‘2r’and YH2Or are the mole fractions of the residual constituents. While thetotal number of moles of unburned gas zone decrease as combustion proceeds, the unburned gascomposition is assumed to remains the same. The assumption of an unchanging compositionalong with the assumption that the unburned gases exhibit ideal gas behavior are used todetermine’ the zone properties.If y1 represents the mole fractions of the unburned gas constituents, then the molecularweight and gas constant are calculated as followsM =y1M70R=—UMThe temperature of the unburned gas changes as a result of compression or expansion and heattransfer to the walls only in the absence of mixing with other zones. The first law written for theunburned gas zone as a control mass isôq =dh —vdPusing v,, = RT11,dh, = CPdTU and= Ry leads toP 1-1óq=y dT dPRET,, y—1T PordT—1dP KÔQW (5-24)T ‘ P mCTwhere K is some fraction of the total heat transfer to the cylinder walls (oQ) and ‘ is the ratioof specific heats for the unburned gas. Since unburned gases are consumed during combustion,K should vary over the combustion duration period. It would therefore seem reasonable to set Kas the ratio of the mass of the unburned zone to the total cylinder mass. Making this substitutionin equation (5-24) yieldsdi’,, ( —ldP÷ ôQ1 (5-25)T,, y PWith a known temperature at the beginning of the time step, equation (5-25) can be usedto calculate the temperature at the end of the time step. To begin the computation, the71temperature at the beginning of the first time step must be known. By beginning thecomputation at the beginning of injection (BOl), only unburned gas is in the cylinder. Thismeans that the total mass of the cylinder contents is equal to the trapped air and residuals ascalculated in the previous section. Since the volume and pressure are known at this point, theideal gas law can be used to calculate the BOl temperature.In order to calculate a value for y, an estimate of the specific heat is required. Thespecific heat at constant volume can be determined using the following expression.y[(T)]. -k=M(5-26)As indicated in section 5.1, an estimate of the internal energy of unburned zone is required. Theenthalpy of the unburned gas can be determined usingI c(T)dTjh= i-i 29.15K (5- 27)U N’Ithen the internal energy is found usinguuUwhere M is the molecular weight of the unburned gas and the overbar denotes a molar basis.5.6 Thermodynamic Properties of the Diesel and Natural Gas ZonesThe injection of fuel can be handled in a couple of different ways. The simplestapproach would be to assume that at any time there is no unburned fuel in the combustion72chamber. The fuel would be introduced at the same rate that it is consumed. This approach isbased on the assumption that the presence of unburned fuel has little effect on the mass andenergy balance in the cylinder since the amount of unburned fuel in the cylinder at any time issmall compared to other cylinder constituents.A more involved approach to deal with injected fuel would be to assume a rate of fuelinjection independent of the fuel burning rate. The properties and mass of unburned fuel in thecylinder would then have to be accounted for. Assuming a constant rate of fuel injection overthe injection period, the changes in unburned diesel and natural gas properties can be determinedin the following way.Diesel fuel is injected into the combustion chamber as a liquid. The temperature of thediesel fuel entering the combustion chamber is roughly the same as the working temperature ofthe injector. In a liquid state, the diesel fuel volume is negligible compared to that of thecylinder gas and its internal energy and enthalpy are approximately equal. The temperature risein the liquid diesel due to compression in the cylinder is negligible. Hence by assuming that thediesel does not evaporate or mix with other zones until the moment it combusts, the temperatureof the diesel zone does not change.Natural gas injected into the combustion chamber is assumed to exhibit ideal gasbehavior. The temperature of the natural gas entering the combustion chamber is also roughlythe same as the working temperature of the injector. Ignoring heat transfer to the fuel from thehot cylinder gases, the temperature of the injected natural gas will increase due to compression.As stated earlier, each zone identified in the model has a uniform temperature. Therefore inorder for the natural gas zone to have a uniform temperature, it is necessary to assume that thereis a continuous mixing between the natural gas being injected and the accumulated unburnednatural gas already in the cylinder from prior injection.By assuming that the natural gas undergoes a two-stage process during each timeinterval, the temperature of the natural gas zone can be calculated. First, there is adiabatic,constant pressure mixing between the incremental mass of natural gas injected and theaccumulated mass of unburned natural gas in the cylinder from prior injection. This change in73the temperature is referred to as dTg1. Second, there is a change in temperature dTg2 fromisentropic compression (or expansion) of the gas. The total change in temperature of the naturalgas is then the sum of dTg1 and dTg2.For the first stage the first law of thermodynamics as a control mass with is written asÔQ—ÔW=dE (528)The heat transfer (öQ) is assumed to be negligible. The work (oW) is equal to PdVg. Thechange in energy of the natural gas (dE) in the cylinder is equal to the change in internal energyif kinetic and potential energy changes are ignored. The internal energy at the initial state isgiven byU1 = mgug + dmgu0where mg and u are the initial mass and internal energy per unit mass of the natural gas zone,and dmg is the mass of injected natural gas for the time step with its corresponding internalenergy per unit mass u0. The internal energy at the final state is given by= (dmg + mg)ug2where ug2 is the internal energy per unit mass of the natural gas zone after mixing. Combiningthe expressions for internal energy and work into equation (5-28) gives_PdVg = (dmg + mg)ug2— mgug1 — dmgu0 (5-29)Expanding equation (5-29) further leads to_P{(dmg + mg)vg2— mgvg1 — dmgvo] = (dmg + mg)ug2— mgug1 — dmgu074but h = u + Pv thuso = (dmg + mg)hg2— mghg1 — drngho0= mgdhg+dmg(hg2_ o)0 = mCd11+ drngCp (Tg2—c) (5-30)where C, is the constant-pressure specific heat of the natural gas evaluated at the initialtemperature, and Cpavg is the average constant-pressure specific heat of the natural gas and isgiven byc +cc— PTg2 PinPavg 2Finally solving equation (5-30) for dTg1 givesdmC (T-T)dT1=g (5-31)For the isentropic compression of the second stage of the process, the following can bewritten:Tds=dh—vdP=00 = CpdTg2 — vdPvdP RTdT =—= 2p (5-32)g2c c75Now by adding equations (5-31) and (5-32) the total change in temperature of the natural gaszone dTg is accounted for. By applying this expression for temperature over every time step, thespecific volume and internal energy for the zone at the end of each time step can be calculated.The specific volume is calculated using the ideal gas law as follows:RTg2g p2To calculate the internal energy the enthalpy is determined first usinghgh+ fc(T)dT2. 15Kthen the internal energy is found usinghgu =——Pv&gwhere Mg is the molecular weight of methane and the overbar denotes a molar basis.5.7 Thermodynamic Properties of the Burned Gas ZonesThe burned gases form as a result of the combustion of the fuel with the unburned gaszone. Mixture which burns early in the combustion process is subsequently compressed ascombustion proceeds and cylinder pressure rises. If no mixing occurs, then this compression canbe thought of as isentropic. Further, dissociation, which occurs in the engine due to the hightemperature of the combustion products, can not be ignored. Its effects must be accounted forsuch that the composition and properties of the burned gas are representative.76The properties and composition of the burned gases can be found using a chemicalequilibrium solver. The products of combustion are assumed to be in equilibrium, except foroxides of nitrogen. The formation of oxides of nitrogen is discussed in a following section.STANJAN [27] is an equilibrium solver that uses the method of element potentials to find theminimum Gibb’s function for a chemical system subject to atom population constraints. Eachspecies is treated as an ideal gases using thermodynamic properties from the JANAF tables. Tofind the burned gas properties, STANJAN requires as inputs the atom populations of carbon,hydrogen, oxygen, and nitrogen (CHON) in the reactants.The CHON atom ratios relative to carbon can be found by looking at the combustionreaction from section 5.2 which is rewritten here.CH ÷(1+).(o2 N2 H2O+CO)y02 y02_*(i+(i+ CO2+(+(i+).) 1120 +(i+ )--. N2y02The CHON ratios relative to carbon are therefore( nyu2on+12+— IH \ 2Jy077( \( Y2o Yco21÷— 2+—+2-———0 4,y0y,J’ flYN21+ hN 4jy023=where Y02, Y2 y, and y110 are the mole fractions of the constiLuents of the unburned gaszone. Since the mole fraction ratios y /y0 and in the unburned gas are small, theabove CHON ratios are closely approximated byH—=nC0/nC \ 4,1Y02The ratios of hydrogen and oxygen to carbon are now simple expressions of n (the ratio ofhydrogen to carbon for the fuel). By noting that YN, /y0 in the unburned gas is nearly the sameas that for air (i.e. 3.76) and nearly constant, then the ratio of nitrogen to carbon is also a simpleexpression in terms of n. Thus two tables of burned gas properties (one for diesel and one fornatural gas) can be created using the above CHON ratios as inputs to STANJAN. The matrix of78burned gas properties for both diesel and natural gas was generated for temperatures andpressures ranging from 1300 to 3100 K and 1 to 95 atmospheres respectively.5.8 Thermodynamic Properties of the Mixed Gas ZoneAs mentioned in the opening section of this chapter, the mixed zone is a mixture ofburned and unburned gases. Burned gases only remain at the high temperatures of the flamefront for a specified mixing delay period until they are forced to mix with a portion of unburnedgas and previously mixed out burned gases. By assuming a two-stage process, the properties ofthe mixed zone can be determined. The first stage of the process is an instantaneous mixing atconstant-pressure between the burned gas zone which has reached the end of the mixing delayperiod with a proportional amount of unburned zone and the contents of the previously formedmixed zone. The second stage of the process deals with the change in properties resulting fromisentropic compression (or expansion).The first law of thermodynamics applied to the mixed zone as a control mass for theconstant-pressure instantaneous mixing process isÔQ—ÔW=dE (5-33)where the work is equal to= PdV = P[(mm + dmb + dmu)vm2— mmv,i,i — dmbvb — dmv] (5- 34)and the change in energy is equal todE= (mm + dmb + dm)u2— mmumi —dmbub —dmu (5-35)The subscripts b and u refer to the burned and unburned zones respectively and dm is the massof burned or unburned zone that mixes. The subscripts I and 2 refer to the respective states79before and after mixing, and m is the mass of the mixed zone before mixing. Assuming heattransfer during the instantaneous mixing is negligible, then using equations (5-34) and (5-35) inequation (5-33) gives—P[(mm + dmb + dmu)vm2— mmvmi — dmbvb — dmv] =(m + dmb + dmu)um2— mmumj —dmbub —dmu (5-36)using u + Pv = h and Pv = RT and solving (5-36) for urn2 yieldsm h ÷dm h +dm h— m ml b b u —m2mm ÷dmb ÷dmIf the temperature at state 2 Tm2 is approximated by the mass-averaged value, then equation (5-37) gives the internal energy of the mixed zone after mixing.For the isentropic compression of the second stage of the process, the first law ofthermodynamics as a control mass isóq—ôw=du (5-38)where the work isôW=Pavg(Vm2Vmi) (5-39)Here the subscripts 1 and 2 refer respectively to the states before and after compression. Theaverage pressure avg is an average of the pressures at these states. If as a first approximation theheat transfer is neglected, then the change in internal energy of the mixed zone due to expansionis given by substituting equation (5-39) into equation (5-38) as follows:du=—Pavg(vm2-vmi) (5-40)80The specific volumes can be estimated using the ideal gas law. To use the ideal gas lawhowever, the temperature change due to expansion is required. The change in temperature of themixed zone can be calculated assuming isentropic expansion as follows:dTmIfl3ldP (5-41)y Pwhere ‘ is the ratio of specific heats for the mixed zone.5.9 Calculation of NOx FormationOxides of nitrogen (NOr) are composed of nitric oxide (NO) and nitrogen dioxide(NO2). Chemical equilibrium considerations indicate that for burned gases at typical flametemperatures, N02/NO ratios should be negligibly small. Mole fractions of NO are typicallymore than 1000 times that of N02.[27] Thus by assuming that inside the combustion chamberof an engine nitric oxide is the predominant oxide of nitrogen produced, an estimate of theformation rate of NO is obtained by determining the formation rate of NO.The mechanism of NO formation in the combustion of near-stoichiometric fuel-airmixtures is widely accepted to be described by the following reactions which are referred to asthe extended Zeldovich mechanism. [1]O÷N2—*NO+N (1)N+0—*NO÷O (2)N÷OH—NO+H (3)The forward and reverse rate constants (k7 and k respectively) of these reactions have beenmeasured experimentally. Heywood [1] has made a critical review of this published data andrecommends using the values given in Table 5-1.81Table 5-1 : Rate Constants for the Extended Zeldovich MechanismReaction Rate Constant(cm3 / mol s)(1) 0+ N2 —*NO+N 7.6x10’exp[-38000/TJ(-1) N + NO — N2 + 0 1.6 x 1013(2) N+ 02 —*NO+O 6.4xlO9Texp[-3150!T](-2) 0 + NO —* 02 + N 1.5 x Texp[-19500/T1(3) N+OH—*NO+H 4.1x1013(-3) H + NO —* OH + N 2.0 x iO’ exp[—23650/T]Using the law of mass action, the rate of formation of NO can be written as:—k[NO][N]—k [N0][0]—k[N0][H] (5-42)where []denotes species concentration in mole / cm3. Similarly, the rate of formation of N canbe written as:— k1 [NO] [N] + k [NO] [0] + k [NO] [H] (5- 43)Since [N] is much less than the concentrations of other species of interest ( 10 mole fraction)[1], the steady-state approximation, d[N] = 0, is used to eliminate [N].The NO formation rate then becomes82[NOjd[NO]=2k[0][N1+K[O,][N2] (544)k[0.,] + k[0H]where K = (ç/k1-)(k;/ ) = 20.267exp(_21650/T)If it is assumed that all of the NO forms in the postflame gases, then the concentrationsof 0, 02, OH, H, and N2 can be approximated by their equilibrium values at the pressure andtemperature of the burned gas zone. By using [ le to denote equilibrium concentration andsubstituting the reaction rates from Table 5-1 into equation (5-44), then the NO formation ratebecomes (in units mol / cm3 . s)- [NOrd[N0]120.267exp(-21650/T)[0 L[N ]dt =t52x1O’exp(-38OOO/T)[0]e[N 1+ [NO]2 24 x 1OTexp(_315O/T)[O]+ 2i625[0H]To eliminate [Ole from the above expression, use can be made of the equilibrium oxygen atomconcentration given by:K(0)[O2]I ]e= (T)”2where K(0) is the equilibrium constant for the reaction-402 = 0 and is given byK(0) = 3.6 x exp(—31090/T) atm112The final form of the expression used to calculate the NO formation rate becomes831—[NO]2d[NOJ 6x1016dt — T112exp(69o9o/T)[O12[NL1+20.267exp(-21650/T)[O,]e[N,]e[NO]4x 1O4Texp(—315O/T)[O + 2.5625[OHLwhich is also in units mol / cm3 s.The strong temperature dependence of NO formation rate can be demonstrated byconsidering the initial value of d[NO]/dt when [NO] = 0. Figure 5-4 illustrates the initialformation rate of NO at 35 atm.Figure 5-4 : Initial NO Formation Rate as a Function of Temperature5.10 Calculation ProcedureThe input data required in this computation are cylinder pressure versus crank angle,engine speed, air manifold temperature and pressure, exhaust temperature, air and fuel mass flowrates, and the unburned fuel fractions. After the mass of air and residuals trapped in the cylinderare determined, the combustion analysis begins at the beginning of injection (BOl). At BOl onlyunburned gas is present in the cylinder, thus the total cylinder energy can be determined at the100001000d[NOJ 100cit[nI/cm3 SI 100.12200 2300 2400 2500 2600 2700 2800T enperciture [KI84beginning of the first time step as it is equal to the internal energy of the unburned gas at BOlconditions. The calculation of BOl conditions are based on the measured air manifoldtemperature and pressure and the exhaust temperature.As the process proceeds, the changes in properties of each zone can he calculatedindependently. As mentioned in the opening section of this chapter, the properties and mass ofthe most recently formed burned gas zone can be found in one of two ways. Either the heattransfer to the cylinder walls must be estimated or the burned gas temperature must be estimatedby the adiabatic flame temperature. When the second method is used, the actual heat transfer tothe cylinder walls can be calculated directly.Once combustion begins, a new burned gas zone is created every calculation step(generally one crank angle). As the combustion process continues, the properties of thepreviously formed burned gas zones that have not yet mixed out must be updated with thechanging cylinder pressure. The changes in these properties are assumed to be the result ofisentropic compression (or expansion). Therefore at every calculation step, the tables of burnedgas properties are interpolated at the cylinder pressure and constant value of entropy of eachpreviously formed burned gas zone.Figure 5-5 illustrates the entire combustion analysis calculation procedure in the form ofa flow chart. The source code (in QuickBASIC) of the analysis has also been included(Appendix VIII).85()READINPUTDATACALCULATE MASS OFAIR TRAPPED IN CYLINDERITERATE TO FIND MASS, COMPOSITION ANDTEMPERATURE OF TRAPPED RESIDUALSSPECIFY PROPERTIES OF EACH ZONE AT BOlSTEP ONE CRANK ANGLE INtERVALDErERMINE FNFHALPY OF INJECTEDUPDATE PROPERTIES OF EACH ZONEDUE TO COMPRESSLONIIIXPANSION1CALCULATE WORIC DONE DURING CRANK ANGLE INTERVALSPECIFY WHICH FUEL IS BURNING AND CALCIJLATE THE(X)RRESPNDING STOICHIOMETRIC AIR-FUEL RATIOEITHERCALCULATE HEAT TRANSFERTO CYUNDER WAIlSCYLINDER ENERGY BALANCEITERATE TO FIND BURNED-GAS-ZONE PROPERTIESORITERATE TO FIND ADIABATICFLAME TEMPERATUREICYLINDER ENERGYII CALCULATE HEAT TRANSFER_________________TO CYUNDER WAIlSCALCULATE MASS OF NEWLY FORMED BURNED GAS ZONECALCULATE MASS PRACI’ION OF BURNED FUELCOMPUTE AMOUNT OF NO FORMED DURING CRANK ANGLE INTERVAL1COMPUTE BULKTEMPERATURE IN COMBUSTION CHAMBER71PIUNTOUTPUTDATACALCULATE PROPERTIES OF MIXED ZONE AFtER MIXINGiiEPARE VARIABLES fOR NEXT CRANKANGLE INTERVALANALAN(YI’HERRANKANGIEINTERVAL? YESNOFigure 5-5 Combustion Analysis Calculation Procedure866 PERFORMANCE, EMISSIONS AND COMBUSTION CHARACTERISTICS6.1 Discussion of Combustion Analysis ResultsThe analysis described in Chapter 5 and the computer code given in Appendix VIII areconsistent with the nominal requirement that in the diffusive burning zone the equivalence ratio isunity. During testing of the combustion analysis, it was found that the calculated mass fraction ofburned fuel would not reach the value deduced from an exhaust composition analysis. In theresults which follow, the local equivalence ratio was adjusted to differ somewhat from unity tocause the calculated mass fraction of burned fuel to reach the measured final value.Table 6-1 : Local Average Equivalence Ratio as a Function of Injection Timing for Diesel Fueling.Local Equivalence Local EquivalenceInjection Timing Ratio Ratio(°ABDC) (1 bar brnep) (3 bar bmep)173 0.97 1.07171 1.03 1.09169 1.03 1.10167 1.05 1.11165 1.05 1.15163 1.10 1.17The departures from an equivalence ratio of one are shown in Table 6-1 for different fuelinjection timings with diesel fueling of the DDC 1-71. These results indicate that the averagelocal equivalence ratios are generally slightly rich of stoichiometric. As fuel injection timing isadvanced and load is increased, the combustion becomes more fuel rich. The overall cylinder87averaged equivalence ratios at 1 bar and 3 bar bmep are 0.17 and 0.28, respectively. Thecalculation of burned gas temperature is affected by changes in the local equivalence ratio, hut thishas not been taken into account in the computation. Thus the results which follow reveal thegeneral effects of wall heat transfer, mixing delay, and method of estimating the burned gastemperatures, but they do not account for the full thermodynamic effect of equivalence ratio.6.1.1 Effect of Computation MethodAs described in Chapter 5, two different methods can be used to calculate the propertiesof the burned gases. The output from the combustion model is dependent on which method isused. The first method for determining the temperature of the most recently formed burned gaszone requires an estimate of the heat transfer to the cylinder walls. Based on this heat transferestimate, the change in total energy from one crank angle to the next can be calculated from a firstlaw energy balance. Woshni’s convective heat transfer correlation is used to calculate the totalheat transfer to the cylinder walls. This method is referred to as the“Qwl” method.The second method for determining the temperature of the most recently formed burnedgas zone is to use the constant-pressure adiabatic flame temperature. In this method, while thecombustion process is assumed to take place adiabatically, the heat transfer to the cylinder walls istaken into account. This method is referred to as the “Tad” method. While the follow resultsshow in general the effects of the different methods for computing the burned gas temperatures,they do not allow for the full thermodynamic consequences of the changing equivalence ratio.The equivalence ratio used in the computation in generating the following results was heldconstant at 1.05.Figures 6-1 through 6-4 illustrate the differences in output from the combustion modelusing these two different methods. For the Qwl method, results have been generated using threedifferent constants (Cl) in Woshni’s heat transfer correlation. Values for Cl of 0, 0.02, and 0.04have been used. When Cl = 0, there is no heat transfer to the cylinder walls. As Cl increases, sodoes the heat transfer.8$The effect of computation method on the calculation of mass fraction of burned fuel isillustrated in Figure 6-1. The results from the two methods agree when Cl = 0. For the 0wlmethod, as the heat transfer to the cylinder walls is increased, the calculated mass fraction ofburned fuel increases at a given crank angle in the later stages of combustion.Li10.90.80.7x 0.6mf0. 6-1 Effect of Computation Method on Mass Fraction of Burned FuelThe calculated values of heat transfer to the cylinder walls and the initial burned gastemperature are shown in Figures 6-2 and 6-3, respectively. For the Tad method, uncertainty isapparent in the calculation of heat transfer. For the Qwl method, uncertainty is apparent in thecalculation of the initial burned gas temperature. This indicates that the inherent uncertainty canbe associated with the energy equation in each case. With this incremental burning model, themass of the newly formed burned gas zone is small compared to the masses of the other zones.Hence the uncertainty in the results can be attributed to small differences between large numbers(i.e. subtractive cancellation results in a large loss of significant digits). In each computationmethod, the uncertainty shows up in the dependent variable.89Tad+ Qwl (Ci=0)Qwl (Ci=0.02)Qwl (Ci=0.04)190 210crank angle (°ABDC)2300.0090.0080.0070.0060.0050.0040.0030.0020.00 10-0.001-0.002-0.003-0.004Qi(kJ)170 190 210 230crank angle (°ABDC)Figure 6-2 : Effect of Computation Method on Heat Transfer Estimates (kJ)2700 2100190017001500Tad+ Qwl (C1=0)° Qwl (C1=0.02)QwI (Cl =0.04)I I180 200 220 240crank angle (°ABDC)Figure 6-3 Effect of Computation Method on Initial Burned Gas Temperature (K)90As with the calculation of mass fraction of burned fuel, the best agreement in the heattransfer and burned gas temperature calculation between the two methods is found when Cl is setto zero in the Qi method. This same result, however, does not occur in the calculation of NO.Figure 6-4 illustrates the estimates of exhaust NO. As shown in chapter 5, NO formation ishighly sensitive to burned gas temperature. As a result of the fluctuations in the calculated burnedgas temperature with the Qwl method, different characteristics in the NO formation are seen.The calculated NO formation rate for the Qi method is higher in the initial stages ofcombustion as a result of the fluctuating burned gas temperatures. Due to the high sensitivity ofNO to burned gas temperature, the computation method based on the constant-pressureadiabatic flame temperature has been used for analysis in the remainder of this thesis since itprovides better burned gas temperature estimates.NO(ppm)400350300250200150100500 170 190 210 230crank angle (°ABDC)Figure 6-4 : Effect of Computation Method on Estimated Exhaust NOx (ppm)916.1.2 Effect of Unburned Fuel in the Combustion ChamberIn Chapter 5, it was stated that two methods could be used to describe the injected fuel.The simplest method assumes that at any time there is no unburned fuel in the combustionchamber, because fuel is introduced into the combustion chamber at the same rate that it isconsumed. In this case, there is no fuel zone present. The more involved method assumes a rateof fuel injection independent of the fuel burning rate. The properties and mass of unburned fuel inthe cylinder are then accounted for in the fuel zone. Computations for both diesel and natural gasfueling showed no differences in combustion rate, burned gas temperature or NO formation withor without fuel zones. While these results show in general the effects of unburned fuel in thecombustion chamber, they do not allow for the full thennodynamic consequences of the changingequivalence ratio.6.1.3 Computed TemperaturesTypical computed temperatures for each of the zones specified in the combustionanalysis are illustrated in Figure 6-5. This particular analysis is based on pressure data collectedfrom the DDC 1-71 operating on diesel fuel at 1250 rpm and 3 bar bmep. The initial burned gastemperature is denoted as Tb,, the temperature of the “mixed” zone is Tm the cylinder massaveraged bulk temperature is TbUIk, the temperature of the unburned zone is T, and thetemperature of the natural gas zone is Tg• The initial burned gas temperature is the temperatureat which each new burned gas zone forms.Since at BOl, the cylinder contains only unburned gases, the bulk temperature is the sameas the unburned gas zone temperature. Upon initiation of combustion, the bulk temperaturebegins to depart from the unburned gas zone temperature and approach the mixed zonetemperature as combustion proceeds. This indicates that the unburned zone is being depleted andthe mixed zone is growing. It also demonstrates that the mass of burned gases, at high flametemperatures, is considerably less than the mass of either the unburned zone or the mixed zone.9228002400 Tbj200()T T(K)16001200800T400 I I170 190 210 230crank angle (°ABDC)Figure 6-5 : Computed Temperatures (K)6.2 Variables that Effect Diesel Engine Performance, Emissions, and CombustionFor a given diesel engine design, the variables that affect engine performance, emissionsand combustion are:• fuel injection timing• fuel injection rate• engine load• engine speedCombustion begins shortly after an ignition delay period from the beginning of fuelinjection. The ignition delay period depends oh the temperature of the air into which the fuel isinjected. Hence fuel injection timing essentially controls the crank angle at which combustionbegins. Fuel injection rate refers to the rate at which fuel enters the combustion chamber. Itdepends on fuel injection pressure and fuel-injector nozzle area. Higher fuel injection ratesrequire a shorter injection duration for a given mass of injected fuel.As the effects of engine load and speed are engine-specific, a common way to present theperformance characteristics of an engine over its full load and speed range is to plot contours of93thermal efficiency on a plot of brake mean effective pressure versus engine speed. These plots arereferred to as performance maps and are used to illustrate the performance characteristics of aparticular engine. In generating a performance map, the fuel injection rate will generally he heldconstant, hut the fuel injection timing will vary. The injection timing is selected at different engineloads and speeds to provide the best possible thermal efficiency with emission levels low enoughto satisfy constraints imposed by emission regulations.The performance map for the DDC 6V-92TA using diesel fueling with factory selectedfuel injection timing values and a set fuel injection rate has been generated based on engine testcell measurements. This map, shown in Figure 6-6, does not show the entire engine load rangesince the dynamometer, to which this engine is coupled, is only capable of absorbing a brnep ofjust over 6 bar. The rated peak bmep of this engine is 9.2 bar at 1200 rpm. Ultimately, a similarperformance map of this engine when fueled with directly injected natural gas would be the besttool for fully evaluating this newly developed fueling concept.BMEP(bar) 6-6 : Engine Performance Map (DDC 6V-92TA)1000 1250 1500 1750 2000Engine Speed (rpm)94Presented in this chapter are results from tests performed using the newly developedprototype natural gas injector described in Chapter 2. These results illustrate the effects of fuelinjection timing, fuel injection rate, engine load and engine speed with the prototype injectoroperating in both the naturally-aspirated single-cylinder diesel engine and in one cylinder of theturbo-charged and after-cooled six-cylinder diesel engine. The performance and emissionscharacteristics of the natural gas fueling system are compared with those of conventional dieselfueling. Results of combustion analysis of both diesel and natural gas fueling using thecombustion analysis model developed in the previous chapter are presented.6.3 Effect of Fuel Injection TimingThe effect of fuel injection timing on performance and emissions near optimum are shownin Figures 6-7 and 6-8 for the DDC 6V-92TA and Figures 6-9 and 6-10 for the DDC 1-7 1. Theresults presented for the DDC 6V-92TA are at the same load (3 bar bmep), but at different speeds(1200 and 1800 rpm). The results presented for the DDC 1-7 1, on the other hand, are at thesame speed (1250 rpm), but at different loads (1 and 3 bar bmep). The behavior of thennalefficiency (11th), Bosch Smoke, carbon monoxide (CO), oxides of nitrogen (NOr), and totalhydrocarbons (THC) with fuel injection timing are illustrated. At the DDC 1-71 low load case,and the DDC 6V-92 low speed case, smoke was not present in the exhaust.The presented measurements have been normalized by dividing by the maximum value ofeach parameter. These normalizing constants are given in Table 6-2. The much improvedemissions with the DDC 6V-92 compared with the DDC 1-71 can be attributed to better fuelinjection characteristics. The figures illustrate that while thermal efficiency is quite insensitive tochanges in injection timing, NO, smoke and carbon monoxide (CO) emissions are stronglyeffected. Retarding injection timing decreases NO1 emissions at the expense of increasing COand smoke emissions. This makes it impossible to achieve optimum thermal efficiency with lowlevels of all emissions with conventional diesel fueling at any engine load and speed.95Table 6-2 : Normalizing Constants for Figures 6-7 through 6-1()Engine DDC 6V-92TA DDC 6V-92TA DDC 1-71 DDC 1-71rpm 1800 1200 1250 1250brnep (bar) 3 3 3 11hmax (%) 29.6 33.7 25.5 16.3NOxm (ppm) 484 677 794 436Bosch 0.6 - 2.7-SmokemaxCOm 92 102 459 538(ppm)THCm 10 11 322 413(ppmc)9610. 6.-7 : Effect of Fuel Injection Timing at 3 bar bmep, 1800 rpmDDC 6V-92TA - Diesel Fueling (normalized) [°ABDC]Figure 6-8 Effect of Fuel Injection Timing at 3 bar bmep, 1200 rpmDDC 6V-92TA - Diesel Fueling (normalized)161 163 165 167 169 171 173 175 177 179BOl [°ABDC]161 163 165 167 169 171 173 175 177 1799710. 6-9 : Effect of Fuel Injection Timing at 3 bar bmep, 1250 rpmDDC 1-71 - Diesel Fueling (normalized) 6-10 Effect of Fuel Injection Timing at 1 bar bmep, 1250 rpmDDC 1-71 - Diesel Fueling (normalized)159 161 163 165 167 169 171 173 175 177BOl [°ABDC1163 165 167 169 171 173 175 177BOl [°ABDCI98The effect of fuel injection timing on performance and emissions near optimum for naturalgas fueling of the DDC 1-71 are shown in Figures 6-11 and 6-12. The results presented are at thesame speed (1250 rpm), but at different loads (1 and 3 bar bmep). The normalizing constants forthese figures are given in Table 6-3.Table 6-3 : Normalizing Constants for Figures 6-11 and 6-12Engine DDC 1-71 DDC 1-71rpm 1250 1250bmep (bar) 3 1lthmax (%) 25.1 14.8NOxm (ppm) 790 395CO 246 608(ppm)THCm 932 1584(ppmc)CH4max 118 920(ppm)As with diesel fueling, these results demonstrate thaI thermal efficiency is quite insensitiveto changes in fuel injection timing, while NO is strongly effected. With natural gas fueling,however, carbon monoxide (CO) emissions are much less sensitive to changes in fuel injectiontiming, and no particulate matter emissions are produced at any fuel injection timing at theseloads. At the low load case (1 bar bmep) shown in Figure 6-1 1, the tradeoff between NO andCO is much less prominent than with diesel fueling. At the medium-high load case (3 bar bmep)shown in Figure 6-12, there are no tradeoffs between any of the emissions. Therefore, in this casethe fuel injection timing can be selected such that all emission levels are low without seriouslysacrificing thermal efficiency.9910. 6-1110. of Fuel Injection Timing at 1 bar bmep, 1250 rpmDDC 1-71 (normalized)0.020” dia. Gas holes, 140 bar Gas pressureBOl [°ABDCJFigure 6-12 Effect of Fuel Injection Timing at 3 bar bmep, 1250 rpmDDC 1-71 (normalized)0.020” dia. Gas holes, 140 bar Gas pressure159 161 163 165 167 169 171 173BOl [°ABDC]159 161 163 165 167 169 171 173100The effect of fuel injection timing on combustion characteristics near optimum are shownin Figures 6-13 through 6-18. Results using six different fuel injection timings at two crank angledegree increments are presented. The equivalence ratios listed in Table 6-1 have been used in thisanalysis. While the following results show in general the effects of fuel injection liming oncombustion characteristics, they do not allow for the full thermodynamic consequences of thechanging equivalence ratio. The effects of fuel injection timing on mass fraction of burned fuel forboth diesel and natural gas fueling are illustrated in Figures 6-13 and 6-14, respectively. Asshown in both cases, the characteristic shape of each curve does not change with injection timing.The ignition delay time is also unaffected. Hence combustion rate is unaffected by injectiontiming near optimum for diesel and natural gas fueling.10110. angle (°ABDC)Figure 6-13 : Effect of Fuel Injection Timing on Mass Fraction of Burned Fuel(DDC 1-71 Diesel Fueling at 3 bar bmep)X mf10. 6-14: Effect of Fuel Injection Timing on Mass Fraction of Burned Fuel(DDC 1-71 Natural Gas Fueling at 3 bar bmep)180 200 220 240160 180 200 220 240crank angle (°ABDC) angle (°ABDC)Figure 6-15 : Effect of Fuel Injection Timing on Initial Burned Gas Temperature (K)(DDC 1-71 Diesel Fueling at 3 bar bmep)T.hi(K)2600250024002600Figure 6-16 : Effect of Fuel Injection Timing on Initial Burned Gas Temperature (K)(DDC 1-71 Natural Gas Fueling at 3 bar bmep)163°ABDC°ABDC160 180 200 220 240160 180 200 220 240crank angle (°ABDC)103The effects of fuel injection timing on the initial burned gas temperature for both dieseland natural gas fueling are shown in Figures 6-15 and 6-16, respectively. The effect of advancinginjection timing is seen clearly in Figure 6-15 to increase the initial burned gas temperature duringthe early stages of combustion. This same effect can not be seen as clearly in Figure 6-16 as aresult of the diesel pilot combustion phase. In the natural gas fueling case, it is assumed that all ofthe diesel pilot bums prior to the combustion of the natural gas. Due to the lower combustiontemperature associated with natural gas, the burned gas temperatures shown in Figure 6-16 arelower than in Figure 6-15.An increase in burned gas temperature, with more advanced injection timings, has asignificant effect on the amount of NO that forms. For the diesel fueling case, these effects areseen in Figure 6-17, which illustrates the effect of injection timing on NO formation. It can alsobe seen that NO formation occurs early in the combustion process, when burned gastemperatures are highest.800700600500NO(ppm) 4003002001000 160 180 200 220 240crank angle (°ABDC)Figure 6-17 : Effect of Fuel Injection Timing on Estimated Exhaust NOx (ppm)(DDC 1-71 Diesel Fueling at 3 bar bmep)104The estimates of exhaust NO depend on the number of burned gas zones present at anytime during combustion. The more burned gas zones present, the longer the burned gases remainat high temperatures before mixing out. Hence more NO is formed. The presence of two, threeand four burned gas zones correspond to mixing delay periods1 of 0.27 ms, 0.40 ms, and 0.52 msrespectively. The measured and estimated NO values at different injection timings are plotted inFigure 6-18 for the diesel fueling case. Estimates of exhaust NO are shown using either two,three, or four burned gas zones present at any time during combustion. At less advanced injectiontimings, the measured NO values are bracketed by NOestimates using two and three burnedgas zones. At the highly advanced injection timings, the measured values fall in-between theestimates using three and four burned gas zones. Assuming that the burned gas temperaturepredictions are accurate for each injection timing case, this suggests that at highly advancedinjection timings, mixing delay periods are longer.750700650600NO 550(ppm) 500450400350300250 I I I I163 165 167 169 171 173Injection Timing (crank angle °ABDC)Figure 6-18: Measured and Estimated NO Emissions (ppm exhaust)1 At 1250 rpm and using a one crank angle degree calculation step.105° Measured NO+ Estimated NO(2 burned gas zones)° Estimated NO(3 burned gas zones)Estimated NO(4 burned gas zones)6.4 Effects of Engine Load and Fuel Injection RateThe effect of engine load on the performance and emissions of the DDC 1-71, usingvarious natural gas fueling rates and operating at a constant speed of 1250 rpm, are illustrated inFigures 6-19 through 6-28. The emissions are given on a wet basis, which accounts for watervapour in the exhaust. The natural gas fueling results are compared with diesel fueling results.For diesel fueling, the fuel injection rate is held constant. Natural gas injection pressures of 100,120 and 140 bar are used along with two different natural gas fuel-injector nozzle areas. Injectortips with gas hole diameters of either 0.016” or 0.020” are used. The injection timing is selectedfor both diesel and natural gas fueling such that the beginning of combustion occurs at TDC.1062BMEP (bar)Figure 6-19 : DDC 1-71 Thermal Efficiency (%) - 0.016” dia. Gas holes2622181410620 1 2 3 4BMEP (bar)Figure 6-20 : DDC 1-71 Thermal Efficiency (%) - 0.020” dia. Gas holes302622th 18(%)14106th(%)Baseline0 100 bar+ 120 bar140 bar1072000Co(ppm)Co(ppm)15001000500200015001000500BMEP (bar)Figure 6-21 DDC 1-71 Carbon Monoxide [wet] (ppm)0.016” dia. Gas holesBMEP (bar)Figure 6-22 : DDC 1-71 Carbon Monoxide [wet] (ppm)0.020” dia. Gas holes0 1 2 3 40 1 2 3 4108500450400350300250200150BMEP (bar)Figure 6-23 : DDC 1-71 Oxides of Nitrogen [wet] (ppm)0.016” dia. Gas holes5004504003503002502000 1 2 3BMEP (bar)Figure 6-24 : DDC 1-71 Oxides of Nitrogen [wet] (ppm)0.020” dia. Gas holes100 bar+ 120 bar140 bar0 1I I I I I——NO(ppm)NO(ppm)2 3 4150 4109D 100 bar+ 120 bar140 barDiesel BaselineI I ICH4(ppm)CH4(ppm)I I I I I800700600500400300200100080070060050040030020010000 1 2 3 4BMEP (bar)Figure 6-25 : DDC 1-71 Methane [wet] (ppm)0.016” dia. Gas holes100 bar+ 120 bar° 140 bar0 1Diesel BaselineI I3 4BMEP (bar)Figure 6-26: DDC 1-71 Methane [wet] (ppm)0.020” dia. Gas holes1109008007006005004003002001000Figure 6-27 DDC 1-71 Non-Methane Hydrocarbons [wet] (ppmc)0.016” dia. Gas holes9008007006005004003002001000100 bar120 bar140 barNMHC(ppm)NMHC(ppm)I I I I I I0 1 2 3 4BMEP (bar)100 bar+ 120 bar° 140 barDiesel Baseline0 1 2 3 4BMEP (bar)Figure 6-28 : DDC 1-71 Non-Methane Hydrocarbons [wet] (ppmc)0.020” dia. Gas holes111For a given gas pressure engine operation becomes unstable below a certain load, whichmay be designated the low load limit. For example, at a gas pressure of 140 bar, engine operationbelow 1 bar bmep experiences misfiring or high cycle-to-cycle variations. The high load limit at agiven gas pressure is set by an upper limit on fuel injection duration. An injection duration ofabout 15 crank angle degrees is used as the upper limit to avoid over-pressurization of the dieselpilot which can damage the injector. Larger gas hole diameters (0.004h1 larger) and higher gasinjection pressures allow greater load capability for a given injection duration, because ofincreased gas mass flow.Figures 6-19 and 6-20 illustrate the effects of gas injection pressure and gas-fuel-injectornozzle area on the thermal efficiency. At low and medium loads, the thermal efficiencies for bothnatural gas and conventional diesel fueling are almost identical. At high loads, the thermalefficiencies for natural gas fueling are greater than those for conventional diesel fueling.Furthermore, while the diesel thermal efficiency begins to decrease with increasing load beforemaximum load is reached, natural gas thermal efficiencies have not yet begun to decrease.Carbon monoxide emissions, illustrated in Figures 6-21 and 6-22, dramatically increase athigh load for the diesel fueling case. A similar increase in CO emissions does not occur for thenatural gas fueling case. Large amounts of CO in the exhaust generally correspond to richcombustion and high levels of smoke, since too much fuel has been injected for proper airutilization. These high smoke levels at high loads usually determine the maximum engine load.The fact that the natural gas fueling regime is not smoke limited, and because the thermalefficiencies do not drop off at high loads, a higher load capability than with conventional dieselfueling is implied.Figures 6-23 and 6-24 illustrate the effects of gas injection pressure and fuel-injectornozzle area on emissions of oxides of nitrogen. These figures show that at injection pressures of120 and 140 bar, emissions of NO are greater than for diesel fueling. However, as shown in theprevious section, the injection timing can be retarded to produce lower NO emissions withoutpenalties in thermal efficiency or CO and smoke emissions.112Figures 6-25 and 6-26 illustrate the effects of gas injection pressure and fuel -injectornozzle area on emissions of methane. As natural gas is about 95% methane, an increase inmethane emissions over diesel fueling is expected. There can be a number of reasons why themethane and non-methane hydrocarbon emissions from natural gas fueling are higher than fordiesel fueling. Some of the fuel-air mixture may become too lean or too rich to support apropagating flame. The trends depicted in Figure 6-25 and 6-26 show that methane emissionsgenerally increase with increasing gas injection rate at a given load. Assuming that mixingbetween natural gas and air increases with increasing gas injection rate, then this suggests that asthe injection rate increases, the gas-air mixture becomes increasingly overlean due to overmixing.Methane emissions are particularly high at low loads. Ignition delay periods are longer atlow loads due to lower gas temperatures. Hence more time is available at low loads for theinjected gas to mix with the air to become too lean to support a propagating flame. It also seemsreasonable that overlean mixtures would tend to occur when the overall cylinder averagedequivalence ratio is lower, which is the case at low loads.Figures 6-27 and 6-28 illustrate the effects of gas injection pressure and fuel-injectornozzle area on emissions of non-methane hydrocarbons (NMHC). While only about 5% ofnatural gas consists of NMHC, some NMHC compounds are formed as intermediate species inthe oxidation reaction of natural gas. Figures 6-27 and 6-28 indicate that NMHC emissionsincrease with increasing load. Assuming that the majority of NMHC emissions can be attributedto unreacted diesel pilot, since the amount of diesel fuel injected remains more or less constantwith increasing load, the increase in NMHC emissions may be the result of a competition for airbetween the diesel fuel and the natural gas. The implication is that the natural gas injectioninterferes with the diesel pilot injection such that sufficient air for complete combustion of thediesel fuel is not available.Figures 6-27 and 6-28 also indicate that for natural gas fueling NMHC emissions aregenerally higher than for diesel at all loads. CO emissions, shown in Figures 6-21 and 6-22, arealso higher for natural gas fueling (at low and medium loads). CO emissions are associated withrich combustion, but are also formed as intermediate species during oxidation. The higher CO113emissions could be the result of poor atomization of the diesel pilot resulting in an over-richdiesel-air mixture. This is assuming that the natural gas combustion is lean and that CO emissionsprimarily result from the combustion of the diesel pilot.Methane, NMHC, and CO emissions and ignition delay periods are greater for the injectortip with 0.020” gas holes. While the diesel pilot holes in the two different injector tips wereintended to be identical, small differences have resulted in slightly different diesel pilot injectioncharacteristics. Greater diesel pilot ignition delay periods allow the natural gas to mix longer andbecome leaner before it begins to combust. This may be another reason why methane emissionsare greater with the injector tip with 0.020” gas holes. It appears that the diesel pilot injectioncharacteristics in the injector tip with the 0.020” gas holes are not as good as with the tip with the0.016” gas holes.The prototype natural-gas injector has been designed to inject a constant amount of dieselfor all injection durations. Thus the diesel-to-natural gas ratio varies with engine load. Figures 6-29 and 6-30 show the diesel-to-natural gas ratios (by energy) for each gas nozzle diameter. Thediesel-to-natural gas energy ratio varies from about 50% at no-load to 20% at full-load, and hassimilar characteristics for all gas injection rates.The effect of engine load on the mass fraction of burned fuel (for diesel fueling of theDDC 1-71) is shown in Figure 6-31. The combustion analysis is performed on an average of 20consecutive engine cycles of pressure data. The figure shows that slightly slower burning occursin the higher load case. Slower burning can be associated with mixing controlled combustionversus premixed combustion. This suggests that at lower loads a greater fraction of the injectedfuel premixes with air during the ignition delay period. The more prernixing that occurs prior toignition, the leaner the fuel-air mixture becomes. Hence, the local equivalence ratio is expected tobe smaller at lower loads to indicate leaner combustion. The local equivalence ratios determinedfrom the combustion analysis were 1.03 at 1 bar bmep and 1.09 at 3 bar bmep.I 14555045403530252015555045403530252015IJ 100 bar+ 120 bar140 bar0 1 2 3 4DieselRatio(%)BMEP (bar)Figure 6-29 : DDC 1-71 Diesel-to-Gas Energy Ratio (%), 0.016” dia. Gas holesDieselRatio(%)0 3 4BMEP (bar)Figure 6-30 : DDC 1-71 Diesel-to-Gas Energy Ratio (%), 0.020” dia. Gas holesu 100 bar+ 120 bar140 bar11510. 6-31 : Effect of Load on Mass Fraction of Burned FuelDiesel Fueling of the DDC 1-71The mass fractions of burned fuel for both natural gas and diesel fueling of the DDC 1-71at 3 bar bmep are shown in Figure 6-32. The shapes of the burning curves are quite similar,which is also the case at 1 bar bmep (not shown). In the natural gas fueling case, a distinctionbetween combustion of the diesel pilot and the natural gas can not be made. It was thought thatbecause these curves were generated using 20 cycles of averaged cylinder pressure data, details ofdistinct burning phases would be masked through averaging. However, single cycles of pressuredata were also analyzed, and again no distinction could be made between the combustion of eachfuel. This suggests that some of the natural gas begins to burn before combustion of the dieselfuel is complete.Low cycle-to-cycle variability indicates that the 20 cycles of averaged pressure data isrepresentative of each cycle. For the diesel fueling cases the calculated coefficient of variation inindicated mean effective pressure (COVimep) was about 0.7%. For the natural gas fueling cases,the calculated COVimeP was about 1.0%. A COVimeP greater than about 10% indicates excessivecrank angle (°ABDC)116cycle-to-cycle variability. The local equivalence ratios for the natural gas burning phasedetermined from the combustion analysis are 0.78 at 1 bar bmep and 0.95 at 3 bar bmep. Thelower equivalence ratio at low load suggests (in accordance with the earlier argument) that thenatural gas combustion is leaner at lower loads. 0.6mf0. 6-33 illustrates the initial burned gas temperatures for both diesel and natural gasfueling at 1 and 3 bar bmep. The burned gas temperatures for natural gas fueling are lower due tothe lower combustion temperatures associated with natural gas. In the early stages ofcombustion, the burned gas temperatures for natural gas fueling are the same order of magnitudeas for diesel fueling because of combustion of the diesel pilot.160 180 200 220 240crank angle (GABDC)Figure 6-32: Mass Fraction of Burned Fuel for both Natural Gas andDiesel Fueling of the DDC 1-71 (3 bar bmep) 180 200 220crank angle (°ABDC)Figure 6-33 : Effect of Fuel and Load on Initial Burned Gas Temperature (K)DDC 1-71While the above results show in general the effects of fuel and load on combustion rateand burned gas temperature, they do not allow for the full thermodynamic consequences of thechanging equivalence ratio. Table 6-4 gives, for each case, the measured and estimated NOvalues along with the number of burned gas zones used in the combustion analysis. Accurateestimates of NO for natural gas fueling require more burned gas zones than for diesel fueling.Assuming that the predicted burned gas temperatures are correct for both diesel and natural gascombustion, this implies that the burned gas mixing delay period is longer for natural gas fueling.ibar dsl+ lbar gas3bar dsl3bar gasI I I I I I I I240118Table 6-4 : Measured and Estimated NO (DDC 1-71)Engine Fueling bmep Measured NO Estimated NO Number of(bar) (ppm) (ppm) Burned GasZonesDiesel 1. 265 265 3Natural Gas 1 240 240 4Diesel 3 345 395 3Natural Gas 3 440 420 51197. CONCLUSIONS AND RECOMMENDATIONS7.1 ConclusionsBased on the investigation of performance, emissions and combustion characteristics ofnatural gas fueling of diesel engines, the following conclusions can be made:1. In a conventional diesel engine, retarding fuel injection timing reduces NO emissions at theexpense of increasing CO and smoke emissions. With natural gas fueling of diesel engines, NOemissions can also be reduced by retarding fuel injection timing, however there is no penalty inCO or smoke emissions.2. The thennal efficiencies for both natural gas and conventional diesel fueling at low andmedium engine loads, are almost identical. The thermal efficiencies at high loads for natural gasfueling are greater than for conventional diesel fueling.3. A sharp increase in carbon monoxide and smoke emissions occurs when approaching peakengine load with conventional diesel fueling but not with natural gas fueling. This along withhigher thermal efficiencies at high loads implies higher peak engine load capability with natural gasfueling.4. The high methane emissions at low loads with natural gas fueling are consistent withoverleaning of the natural gas due to overmixing prior to ignition.1205. Accurate predictions of engine exhaust NO emissions can be made using the presentedcombustion analysis model. The accuracy of NO predictions strongly depend on the accuracy ofthe burned gas temperature results, due to the high sensitivity of NO formation to burned gastemperature. From combustion analysis results, it is evident that NO forms early in thecombustion process when burned gas temperatures are highest. Hence, reduction of peak burnedgas temperatures results in a reduction of NO emissions.6. Accurate estimates of both heat transfer from the combustion gases to the cylinder walls andburned gas temperatures can not be made simultaneously using the presented combustion analysismodel. This is attributed to numerical error associated subtractive cancellation1and summationerrors, and measurement error of engine parameters (such as air and fuel flow rates).7. Before the presented combustion analysis model can be used to give conclusive informationabout the local equivalence ratio at which combustion takes place inside the engine, it needsmodification to account for the effect of lean or rich burning on the computed burned gastemperatures.8. For natural gas fueling, the rate of combustion is greater than for conventional diesel fueling asa result of more predominate pre-mixed combustion versus mixing-controlled combustion. Thecombustion rates for both fueling regimes decrease with increasing load as mixing-controlledcombustion predominates over pre-mixed combustion.1Subtractive cancellation is the subtraction of two nearly equal numbers that cause a large loss in the number ofsignificant digits.1217.2 RecommendationsBased on what was learned in the investigation of performance, emissions and combustioncharacteristics of natural gas fueling of diesel engines, the following recommendations are bemade:1. Improve diesel pilot spray characteristics to reduce hydrocarbon and carbon monoxideemissions.2. Delay the injection of natural gas after the injection of diesel pilot a greater amount at lowerloads to reduce methane emissions at low loads.3. Minimize the amount of diesel pilot to minimize exhaust emissions and maximize thermalefficiencies. Stable engine operation and low cycle-to-cycle variability will indicate the minimumallowable amount of diesel pilot.4. Account for non-stoichiometric equivalence ratios in the calculation of composition andproperties of the burned and unburned gases to obtain greater accuracy from the combustionanalysis.5. Investigate the effects of engine speed on performance, emissions and combustion to obtain amore complete comparison between conventional diesel and natural gas fueling.1228. REFERENCES1. Heywood, J. B., “Internal Combustion Engine Fundamentals”, McGraw-Hill Inc., New York,1988.2. Communication, “Information Update”, Detroit Diesel Corporation, Sept. 19933. Beck, N. J., Johnson, W.P., George, A. F., Peterson, P. W., van der Lee, B., and Klopp, G.,“Electronic Fuel Injection for Dual Fuel Diesel Methane”, SAE Technical paper 891652,Aug., 1989.4. Miyake, M., Biwa, T., Endoh, Y., Shimotsu, M., Murakami, S., and Kornoda, T., “TheDevelopment of High Output, Highly Efficient Gas Burning Diesel Engines”, CJMAC PaperD11.2, Conference Proceedings, Paris, France, June 1983.5. Einang, P., Koren, S., Kvamsdal, R., Hansen, T., and Sarsten, A., “High-Pressure,Digitally Controlled Injection of Gaseous Fuel in a Diesel Engine, With Special Reference toBoil-Off from LNG Tankers”, Proceedings CIMAC Conference, Paris, France, June 1983.6. Einang, P., Engja, H., and Vestergren, R., “Medium Speed 4-Stroke Diesel Engine Using HighPressure Gas Injection Technology”7. Wakenell, J. F., O’Neal, G. B., and Baker, 0. A., “High-Pressure Late Cylce DirectInjection of Natural Gas in a Rail Medium Speed Diesel Engine”, SAE Technical paper872041, Nov., 1987.8. Gunawan, H., “Performance and Comb ustion Characteristics of a Diesel-Pilot GasInjection Engine”, Unpublished M.A.Sc. Thesis, Department of Mechanical Engineering,University of British Columbia, June, 1992.9. Tao, Y., “Performance and Emission Characteristics of a Gas-Diesel Engine”, UnpublishedM.A.Sc. Thesis, Department of Mechanical Engineering, University of British Columbia,August, 1993.12310. Ouellette, P., “High Pressure Injection of Natural Gas for Diesel Engine Fueling”,Unpublished M.A.Sc. Thesis, Department of Mechanical Engineering, University of BritishColumbia, January, 1992.11. Chepakovich, A., “Visualization of Transient Single- and Two-Phase Jets Created by DieselEngine Injectors”, Unpublished M.A.Sc. Thesis, Department of Mechanical Engineering,University of British Columbia, April, 1993.12. Go-Power Corporation Publication “Operation, Installation, Service and Repair of Models Dand DA-316 -516 Dynarnomters”, May, 1984.13. Pierburg Instruments Inc. “Instruction Manual - Model FT1OE Electrical Mass FlowTransmitter”14. deSilva, C. W.,”Control Sensors and Actuators”, Prentice-Hall inc., New Jersey, 1989, p. 3715. Lancaster, D. R., Krieger, and R. B., Lienesch, J. H., “Measurement and Analysis of EnginePressure Data”, SAE Technical paper 750026, 1975.16. “1992 SAE Handbook Volume 3”, Society of Automotive Engineers, 199217. Edwards, C. F., Siehers, D. L., and Hoskins, D. H., “A Study of the AutoignitionProcess of a Diesel Spray via High Speed Visualization”, £4E Technical paper 920108, 1992.18. Annand, W. J. D., “Heat Transfer in the Cylinder of Reciprocating Internal CombustionEngines”, Proc Insin Mech Engrs, Vol.177 No. 36, 1963.19. Woschni, G., “A Universally Applicable Equation for the Instantaneous Heat TransferCoefficient in the Internal Combustion Engine”, SAE Technical paper 670931, 1967.20. Szekely, G. A. and Alkidas, A. C., “A Two-Stage Heat-Release Model for DieselEngines”, SAE Technical paper 861272, 1986.21. Shayler, P. J. and Wiseman, M. W., “Improving the Determination of Mass Fraction Burnt”,SAE Technical paper 900351, 1990.22. Gatowski, J. A., Balles, E. N., Chun, K. M., Nelson, F. E., Ekchian, and I. A., Heywood, J.B., “Heat Release Analysis of Engine Pressure Data”, SAE Technical paper 841359,1984.23. Krieger, R. B. and Borman, 6. L., “The Computation of Apparent Heat Release forInternal Combustion Engines”, ASME paper 66-WA/DGP-4, Nov, 1966.12424. Bedran, E. C. and Beretta, G. P., “General Thermodynamic Analysis for EngineCombustion Modeling”, SAE Technical paper 850205, 1985.25. Ferguson, C. R., “Internal Combustion Engines - Applied Thermosciences”, John Wiley &Sons, Inc., 198626. Personal Communication from Detriot Diesel Corporation.27. Reynolds, Wm. C., “STANJAN chemical equilibrium solver v 3.93 IBM-PC”, Mech. Eng.Dept., Stanford University, 198728. Van Wylen, G. J. and Sonntag, R.E, “Fundamentals of Classical Thermodynamics”, thirdedition, John Wiley & Sons, inc., 1985125APPENDIX IHYDRAULIC WHEATSTONE BRIDGE DIESEL FUEL MEASURING DEVICEWith reference to Figure 2-4, if the recirculating flow, q, is less than the measured flow,Q, then the flow through orifices ‘b” and “d” will be half of 0 + q and the flow through orifices“a” and “c’ will be half of Q — q. The flow versus pressure drop relationship for orifice “d” is0±q=KCdAdP1P3and the same relationship for orifice “a” is°= /P1p22By squaring the above expressions and then subtracting the expression for orifice “a” from orifice“d” and assuming K2CA = K2CA = k1 (a constant) the result isQq=k1pRearranging gives126Qp=-(P2—P3)qSince q is a preset constant flow rate and k1 is also a constant, this expression shows that thediesel mass flow rate (Qp) is directly proportional to the differential pressure (P, — P3).By a similar development, if the recirculating flow, q, is greater than the measured flow,Q, the final expression for the diesel mass flow rate isQp=h(P1—P4)qHence once the flow characteristics of the orifices are determined, the constant of proportionalitycan be calculated. Then measuring the proper differential pressures from the 11WheatstoneBridge’ network will give a measure of the diesel mass flow rate directly without uniquelydetermining the diesel density.127APPENDIX IIPRESSURE TRANSDUCER MOUNTING IN CYLINDER OF DDC 1-71EXHAUST VALVECYLI A ,‘\ _(_\/iDTMEPRESSLR TRANSDUcERCPtB M, 11EAFigure A-I : Pressure Transducer Mounting in Cylinder of DDC 1-71(courtesy of I-i. Gunawan [8])PRESSURE SIGNALID CHARC AMPLIFIERSECTV A—A128(jC/I—--,-‘C-3C—)3—— —ri—CDflC0 z(_)I-9 C ri I-/ri x I C V - E ri V213I V I-n -< F z m ;tJ InU 0-niXF xwrnin-U V zAPPENDIX IVSOFTWARE FOR PROCESSING CYLINDER PRESSURE DATAPressure data acquired using the ISAAC are stored in Basic binary format. The ISAACsamples the pressure transducer at a specified rate (i.e. every crank angle), digitizes the sampledvalue and stores it as an integer. Each engine cycle of pressure data collected is stored in aseparate file with a numerical file extension. Pressure data from the first cycle will be stored in afile with the extension .00 1. Similarly, the twentieth consecutive cycle of pressure data will bestored in a file with the extension .020. This binary data must be converted to ASCII format,scaled, and then averaged for use in the combustion analysis.PDC (Pressure Data Conversion Utility)The program PDC is used to convert binary ISAAC pressure data files to ASCII text.Since the ISAAC pressure data is a digitized value stored as an integer, when the data isconverted to ASCII it must also be scaled. The system gain and the transducer-charge amplifiercalibration factor are used to convert the integer data to relative pressures. The relative pressuresare shifted by a constant value to obtain absolute cylinder pressure (in kPa). There are severaltechniques for determining the magnitude of the shift required. In PDC the assumption made isthat the cylinder pressure at BDC is equal to the mean intake air manifold pressure.In PDC there are two methods in which the conversion is done; 1) Convert each binaryfile to its ASCII equivalent directly, and 2) Extract pressure data from a set of files and mergeinto a single file, in preparation for mass of fuel burned analysis.PDC is also used to attach other pertinent engine data (such engine speed, exhaust temperature,etc.) to the ASCII file in the form of a header. PDC is menu-driven. Menu options are given in130Figures AX-i and AX-2. If the file PDC.INI exist in the directory that you started PDC from,then PDC uses it to establish initial settings for the various menu options. The current session’ssettings are saved in PDC.INI upon exiting the program (changes are not saved if Ctrl-C ispressed to exit). The .INI file is in ASCII format and can be modified using a text editor.XPDATA (Pressure Data Averaging and Indicated Work Calculation Utility)The program XPDATA uses the output data file from PDC directly if the ‘prepare formass burn analysis” option in PDC was selected (no modifications to this file are necessary).XPDATA averages any number of engine cycles of pressure data, calculates the indicated workfor each cycle, then calculates the average indicated work and coefficient of variation in imep.The reason why combustion analysis uses averaged pressure data is because the engine isitself an averaging device which responds to mean values of air and fuel flows by generating amean power output. It is therefore appropriate to use the mean pressure data of many cycles forcombustion analysis as the other quantities (e.g. fuel and air flow, exhaust and inlet temperature,etc.) used in the analysis are mean engine measurements.Different pressure data averaging techniques can be used. The technique used inXPDATA is to calculate the mean pressure at each crank angle.XPDATA requires some user input. It first prompts the user to select the engine type(DDC 6V-92 or DDC i-71), then to enter the input file name (and directory). The output file ofXPDATA is used directly as the input file for the combustion analysis program (XMF). Hencethe user is prompted to enter the unburned fractions of diesel and natural gas, even though theyare not directly used for computation in this program.As the computation proceeds, XPDATA graphically displays each cycle of pressure dataon a log-log plane of cylinder pressure and volume. The indicated work for each cycle is alsodisplayed. After the averaging is complete, the average pressure data and the average indicatedwork and coefficient of variation in imep are displayed. The user is then asked if more data fromfiles are to be averaged.131AP1ENDIX VBRAKE POWER CORRECTION (SAE STANDARD J1349 JUN85)The correction is made against standard inlet air conditions:Inlet Air Pressure (Absolute) : 100 kPaInlet Air Temperature : 25°CDry Inlet Air Pressure (Absolute) : 99 kPaThe correction factor f applied to the observed brake power is a function of theatmosphere factor a and the engine factor m The following empirical relationship is used:= (fa)fmThe atmosphere factor is calculated based on the measured dry inlet air pressure BdO (in kPa) andthe measured inlet air temperature t0 (in °C) as follows:a LBdO]L 298 JFor the DDC 1-71 a=1.0 and for the DDC 6V-92TA a=0.7. The engine factor is a function ofthe fuel flow F (in g/s), the engine displacement D (in litres), the engine speed N (in rpm), the fueldelivery q, and the pressure ratio r of measured inlet manifold pressure to measured inlet airpressure. The value ofmis given as:132= 0.3 for (q/r) <40= (0.036)_ H4 for 40 < q/r < 65f=1.2 for (q/r)>65where q = 60000 —f-DNThe correction factor applied to the observed brake power is within the range of 0.90to 1.10.133APPENDIX VICALCULATION OF SPECIFIC HUMIDITYSpecific humidity H (g of H20 per kg of dry air) of an air-water vapour mixture is definedas the ratio of the mass of water vapour m (kg) to the mass of dry air ma (kg) [28].H=1000-- (1)maThe equation of state for both water vapour and dry air can be written in terms of partialpressures as follows:= mRT (2)PaV = maRaT (3)The gas constant R for water vapour is 461.62 i/kg-K. The gas constant Ra for air is 287.0J/kg-K.Dividing equation (2) by equation (3) gives:mw(RaIPw (4ma wIaSubstituting equation (4) into equation (1) and using Ra/Rw = 0.6219 gives:H=621.9-- (5)134Relative humidity is defined as the ratio of the partial pressure of water vapour P to thesaturated steam pressure P evaluated at the dry bulb temperature.P(6)SThe barometric pressure Pb is the sum of the partial pressures of dry air and water vapour.Pb=Pa+Pw (7)Substituting equations (6) and (7) into (5) gives:H=621.9 PS (8)—To calculate the humidity ratio using equation (8), barometric pressure b (inHg), relativehumidity (%), and dry bulb temperature T (°F) are measured directly. The saturated steampressure (inHg) is expressed in terms of T (°F) by an equation from a curve fit on the SaturatedSteam Tables [28] from 40 to 140 °F.P5 = —0.46192 + 0. (i29439T —0. 00045T + 0. 000004T3135APPENDIX VIICALCULATION OF ENGINE CYLINDER VOLUMEThe cylinder volume calculations for both the DDC 1-71 and the DDC 6V-92 were donein the same manner. An illustration of the engine cylinder geometry is given in Figure A-3.The cylinder volume is given by1BVwhere B is the cylinder bore, V is the clearance volume, and the distance y is as defined inFigure A-3. The clearance volume is given byVdr —cr-iwhere is the cylinder displacement volume and cr is the compression ratio. The length y canbe expressed byy = Li— L2whereSLi = — + r + L3and136L2 = sin(O— 90)+ qr2 _[cosO_ 90)1 += [sinOcos9o — cos0sin90]+ 2_()[cosOcos9() + sinOsin9O]2+ L3= —cosO+ r2— sinO2+ L32 2 )hencey = r4(1+cosO)_(r2 _(sinü)2)137LiL2Figure A-3 : Engine Cylinder Geometrys = strokeB = borec = clearance heightr = connecting rod lengthfYTr0 /138APPENDIX VIIICOMBUSTION ANALYSIS PROGRAM- XMF.BAS (Qu1ckBASIC)1= XMF.BASThis program calculates the mass fraction of fuel burned in thecylinder of a DDC 6V92-TA or 1-71 diesel engine based on pressurevs. crank angle data. This analysis is valid for two fueling scenerios.It can handle fueling with natural gas and diesel pilot as well asfueling with straight diesel. It is assumed that burning occurs atstoichiometric air/fuel ratios, with all of the diesel fuel burningprior to the natural gas.Either the adiabatic flame temperature is used to estimate the burnedgas temperature or it is calculated based on an estimate of heattransfer to the cylinder walls.Written by Brad DouvilleThe arrays “PcyP’ and ‘ca” contain the pressure and crank angle dataDIM ca(360), Pcyl(360), P(120), x(10), y(10)DIM mbz(10), vbz(10), ubz(10), sbz(10), Thz(10, 10), NOxz(10)DECLARE SUB unburned (Tu!, hu!, uu!, Cvu!, visc!, 1%)DECLARE SUB gas (Tg!, Cpg!, ug!, 1%)DECLARE SUB burned (P!, Th!, vu!, vm!, vb!, uu!, urn!, ub!, sb!, xmbi!, dNO!, NOe!, 1%)DEClARE SUB tad (P!, Th!, vu!, vb!, hu!, ub!, sb!, dNO!, NOe!, 1%)DECLARE SUB qwall (Tm!, P!, Po!, vol!, asurf!, dqwl!, 1%)DECLARE FUNCTION Vcyl! (ca!)DECLARE FUNCHON Cvair! (Tair!)DECLARE FUNCTION Cvres! (Tres!)DECLARE SUB table (Thx!, Pbx!, hbx!, sbx!, ubx!, vbx!, dNO!, NO!, NOe!, 1%)DECLARE SUB cubics (np%, x!O, y!, xset!, ycalc!)DECLARE FUNCTION Acyl! (ca!)DIM ShARED bore, stroke, rod, VcIrDIM SHARED mtot, nres, natrap, Ru, dtDIM SHARED yO2r, yN2r, yCO2r, yH2Or residual compositionDIM SHARED y02, yN2, yCO2, yH2O, wtmol a mole fractions and molar mass of unburned mixtureCLSPRINTINPUT” Enter pressure data file name (eg. octO2a): “, pdata$PRINTPRINT” Output to File? (yin)”oputfile$ = INPUT$(1)IF oputfile$= “y” OR oputfile$ = Y” THENPRINT139PRINT” Data output filename?INPUT ‘ (do not include file extension) “,out$PRINTPRINT ‘ Output will be stored as c:\wk\brad\outV’; out$; “.out’END IF‘PRINT‘INPUT” Enter a mass ratio of unburned gases that mix with burned gases: “, mixratiomixratio = IPRINTINPUT” Enter local equivalence ratio for diesel burning phase: “, phidslPRINTINPUT” Enter local equivalence ratio for gas burning phase: “, phigasPRINTDOINPUT” Enter the number of existing burned gas zones at any instant (ito 6): “, brndzones%LOOP UNTIL brndzones% < 7 AND brndzones% > 0DOPRINTPRINT ‘ Select computation method:PRINT‘ 101 Estimate the burned gas temp. using the adiabatic flame temp.”PRINT” [1] Calculate the burned gas temp. based on an estimate of”PRINT “ heat transfer to the combustion chamber walls.”PRINTINPUT” “, methodLOOP UNTIL method = 0 OR method = 1Read the engine fluid properties and operating parameters:rpm is the engine crank shaft rotational speed (rev/mm)Pabox is the engine air box pressure [kPajTabox is the engine air box temperture [KITres is the engine exhaust port temperature [Kimair is the delivered air flow [kgJcycle/cylinderjmdsl is the diesel flow [kg/cycIe/cylindermgas is the natural gas flow [kgJcycle/cylinderlnote: if mdsl=0 then mgas must also equal zerocapri is the crank angle of the first pressure recordcaboi is the crank angle of the beginning of injection of diesel fueldca is the crank angle interval sizecapw is the total pulsewidth of fuel injection in crank angle degreescapwd is the pulsewidth of diesel pilotnpr% is the # of pressure records.nca% is the # of ca intervals to be analyzedncyc% is the # of engine cycles of averaged pressure databore is cylinder bore (m)stroke is the cylinder stroke (m)rod is conn rod length(m)Cr IS compression ratioVclr is clearance volume(m**3/cyl)Vdisp is displacement volume(m**3/cyl)caipc is a crank angle (ABDC) after all ports are closedxudsl is the unburned fraction of diesel in the exhaustxugas is the unburned fraction of natural gas in the exhaustMWexh is the molar mass of the engine exhaustOPEN “c:\wk\brad\ascii\” + pdata$ + “.xpj” FOR INPUT AS #1INPUT #1, rpm, Pabox, Tabox, TexhINPUT #1, mair, mdsl, mgas140INPUT #1, caprl, caboi, dca, capwINPUT #1, npr%, nca%, ncyc%INPUT #1, engine%, xudsl, xugasFOR i% = 1 TO npr%INPUT #1, ca(i%), Pcyl(i%)NEXT i%CLOSE #1dt= dca/rpm/6MWexh = 28.8IF engine% = 1 THENengine$ = “171”PRINT” Pressure data is from the DDC 1-71 engineELSEengine$ = “6V92”PRINT” Pressure data is from the DDC 6V-92TA engineEND IFPRINTPRINT” working...”IF engine$ = “6V92” THENbore = .12294rod = .257175cr=17Vdisp = .0015075caipc = 55ELSEIF engine$ = “171” THENbore = .10795rod = .254cr = 16Vdisp = .001162caipc = 60END IFstroke = .127VcIr = Vdisp / (Cr - 1)‘set up graphics screenCLSSCREEN 12LOCATE 1,60PRINT pdata$LOCATE 2,60PRINT “mixratio= “; mixratioLOCATE 3,60PRINT “brndzones= “; brndzones%LOCATE 4,60PRINT “phidsi= “; phidsiLOCATE 5,60PRINT “phigas= “; phigasLOCATE 6,60PRINT USING “Xmfmax = #.##“; 1 - xudsl- xugasLOCATE 4,5PRINT “cylinder”LOCATE 5,5PRINT “pressure”LOCATE 6,5141PRINT ‘ (kPa)”LOCATE 2, 16PRINT “7000”LOCATE 3, 16PRINT “4000LOCATE 5,16PRINT “2000LOCATE 8, 17PRINT “500”LOCATE 12,17PRINT “100”LOCATE 15,3PRINT “ mass’LOCATE 16,3PRINT” burned”LOCATE 17,3PRINT “fraction”LOCATE 22,16PRINT “0”LOCATE 18,14PRINT “0.5”LOCATE 14, 14PRINT “1.0”VIEW PRINT 25 TO 30VIEW (100, 0)-(500, 180)WINDOW (-10, 4.5)-(-6, 9.5)LINE (LOG(Vcyl(ca(1))), LOG(i00))-(LOG(Vcyl(ca(npr% / 2))), LOG(i00)), 14LINE (LOG(Vcyl(ca(1))), LOG(500))-(LOG(Vcyl(ca(npr% I 2))), LOG(5(X))), 14LINE (LOG(VcyI(ca(1))), LOG(2000))-(LOG(Vcyl(ca(npr% / 2))), LOG(2000)), 14LINE (LOG(Vcyl(ca(1))), LOG(4000))-(LOG(Vcyl(ca(npr% / 2))), LOG(4000)), 14LINE (LOG(Vcyl(ca(1))), LOG(7000))-(LOG(Vcyl(ca(npr% / 2))), LOG(7000)), 14LINE (LOG(Vcyl(ca(npr% / 2))), LOG(100))-(LOG(VcyI(ca(npr% / 2))), LOG(7000)), 14LINE (LOG(Vcyl(ca(1))), LOG(100))-(LOG(VcyI(ca(1))), LOG(7000)), 14PSET (LOG(Vcyl(ca(1))), LOG(Pcyl(1))), 11FORi%=2TOnpr%LINE -(LOG(Vcyl(ca(i%))), LOG(PcyI(i%))), 11NEXTi%‘Estimate the residual gas mole fraction, fresnres/(nres+natrap)‘and the mass of air trapped in the cylinder, matrap(Kg)‘A correlation of scavenging data from DDC, the ideal gas law and‘an energy balance at ipc are used to determine Fres.‘Ra=[KJ/Kg-K] Rbar4KJ/KmoI-K1Ra — .287rhoabox Pabox I (Ra * Tabox)mideal = rhoabox * (Vdisp)rscav = mair / midealmatrap = mideal * (.9 - (1 - .34 * rscav) * ((2.71828) “(1.11 * rscav)))natrap = matrap / 28.97flair = mair / 28.97ndsl = mdsl / 13.8ngas = mgas /16.04‘specify initial residual composition to begin iterationyO2r = .21yN2r = .79yCO2r=0142yH2Or = 0Tres = Texhiter% = 0DOiter% = iter% + 1yO2old = yO2r‘the equation for nres is a quadratic so use the quadratic formulaa = Cvres(Tres) * Tresb = natrap * (Cvres(Tres) * Tres + Cvair(Tabox) * Tabox) - Pabox * VcyI(caipc) * Cvres(Tres) / 8.3144c = natrap * Cvair(Tabox) * (natrap * Tabox - Pabox * Vcyl(caipc) / 8.3 144)nresl =(b+SQR(b*b4*a*c))/(2*a)nres2=(bSQR(b*b4* a* c))/(2 *a)IF (nresl > 0) AND (nresl <natrap) THENnres = nreslELSEIF (nres2 > 0) AND (nres2 < natrap) THENnres = nres2ELSEPRU\IT nres is out of rangenres = 0END IFcalculate the composition of the residualsn02 = .21 * natrap + yO2r * nres - 1.45 * ndsl- 2 * ngasnN2 = .79 * natrap + yN2r * nresnCO2 = yCO2r * nres + ndsl + ngasnH2O = yH2Or * nres + .9 * ndsl + 2 * ngastotmols = n02 + nN2 + nCO2 + nFI2OyO2r = n02 / totmolsyN2r = nN2 I totmolsyCO2r = nCO2 / totmolsyll2Or nH2O / totmolsmwres = yO2r * 31.999 + yN2r * 28.013 + yCO2r * 44.01 ÷ yH2Or * 18.015‘calculate residual gas temperatureCpair = (Cvair(Tabox) + 8.3144) / 28.97 ‘U/kg KCpres = (Cvres(Tabox) + 8.3144) / mwres ‘kJ!kg KTres = Texh + (Texh - Tabox) * (mair - matrap) * Cpair / (matrap + mgas + mdsl) / CpresLOOP UNTIL ABS(yO2old- yO2r) < .0001Fres = nres / (nres + natrap)mres = mwres * nresrdeliv = mair / (mres + matrap)degp = matrap / (mres + matrap)PRINT USING “## iterations were required to converge on the residual composition”; iter%PRINT USING “rdeliv = #.## degp = #.## fres = #.## mwres = ##.#“; rdeliv; degp; Fres; mwresPRINT USING “matrap = mres = #.## A; matrap; mresPRINT USING ‘Tres = ### K Texh = ### K Tabox = ### K”; Tres; Texh; TaboxPRINT USING “The mass of fuel burned should reach a maximum of #.##“; 1 - xudsl - xugasINPUT “Press any key to continue... “, cont$OPEN “c:\wk\brad\chneq\dsl.dat” FOR INPUT AS #1OPEN “c:\wk\brad\chneq\cng.dat” FOR INPUT AS #2CALL unburned(ql, q2, q3, q4, q5, 1)CALL gas(ql, q2, q3, 1)IF method = I THENCALL burned(ql, q2, q3, q4, q5, q6, q7, q8, q9, qlO, ql 1, q12, 1)143CALL qwall(ql, q2, q3, q4, q5, q6, 1)ELSECALL tad(ql, q2, q3, q4, q5, q6, q7, q8, q9, 1)END IFCLOSE #1CLOSE #2IF oputfile$ = “y” OR oputfile$ = “Y” THENOPEN “c:\wk\brad\out\” + out$ + “.out” FOR OUTPUT AS #1PRINT #1, CHR$(34); “Output from XMFBAS “+ DATE$ +“ file: “+ pdata$ + “.out”; CHR$(34)PRINT #1, CHR$(34); “Engine is DDC” + engine$; CHR$(34)PRINT #1,PRINT #1, CHR$(34);PRINT #1, USING “####rpm Tres = ### K Texh = ### K Tabox = ### K”; rpm; Tres; Texh; Tabox;PRINT #1, CHR$(34)PRINT #1, CHR$(34);PRINT #1, USING “Number of burned gas zones = # mixratio = #.##“; brndzones%; mixratio;PRINT #1, CHR$(34)PRINT #1, CHR$(34);PRINT #1, USING “Gas velocity/mean piston speed proportionality factor = #.#“; Cl;PRINT #1, Cl-i R$(34)PRINT #1, CHR$(34);PRINT #1, USING “rdeliv = #.## degp = #.## fres = #.## mwres = ##.#“; rdeliv; degp; Fres; mwres;PRINT #1, CHR$(34)PRINT #1, CHR$(34);PRINT #1, USING “diesel combustion is complete when xmf =#.### “; (mdsl/ (mdsl + mgas) - xudsl);PRINT #1, CHR$(34)PRINT #1,PRINT #1, CHR$(34);” CAØ P [kPa] Tu [K] Tbi [K] Tm [K] Tg [K] Tbulk[K1 NOx [ppm] xmfxmu xmm xmd xmg dqwl”; CHR$(34)END IFRedefine pressure and crank angle arrays such that theyspan from (BOI÷1)-->359 deg. instead of from PR1-->359 deg.for mass burned fraction analysis.kboi% = (caboi - capri) \ dca + 1Pboi = Pcyl(kboi%)ca(1) = caboi + dcaPcyl(1) = Pcyl(1 + kboi%)FORi%=2TOnpr%-kboi%ca(i%) = ca(i% 1) + dcaPcyl(i%) = Pcyl(i% + kboi%)NEXTi%‘Draw mass fraction of burned fuel graphVIEW (100, 190)-(500, 370)PRINTPRINTPRINTPRINTPRINTWINDOW (caboi - 10, ..2)-(caboi + nca% + 10, 1.1)LINE (ca(1), 0)-(ca(nca%), 0), 4LINE (ca(l), .25)-(ca(nca%), .25), 4LINE (ca(1), .5)-(ca(nca%), .5), 4144LINE (ca(1), .75)-(ca(nca%), .75), 4LINE (ca(1), 1)-(ca(nca%), 1), 4LINE (ca(1), 0)-(ca(1), 1), 4LINE (ca(nca%), 0)-(ca(nca%), 1), 4‘initialize variablescapwd = mdsl / (mdsL + mgas) * capw ‘approximate dsl pulse widthxmf = 0 ‘mass fraction of burned fuelmb = 0 ‘cummulative mass of burned gasmbi = 0 ‘mass of the ith burned gas zonembilast = 0 ‘mass of the burned gas zone to be mixed in the mixed” zonembiml = 0 ‘mass of the i-i burned gas zonembim2 = 0 ‘mass of the i-2 burned gas zonembim3 = 0 ‘mass of the i-3 burned gas zonembim4 = 0 ‘mass of the i-4 burned gas zonembim5 = 0 ‘mass of the i-S burned gas zonembim6 = 0 ‘mass of the 1-6 burned gas zonemm = 0 ‘mass of the mixed zoneNOimi =0NOim2 =0NOim3 = 0NOim4 =0NOimS =0NOim6 =0Thimi = 0Thim2 =0Thim3 = 0Tbim4 = 0ThimS = 0Tbimó = 0sbiml =0sbim2 =0sbim3 =0sbim4 = 0sbim5 =0sbim6 = 0NOx =0‘AFdsl is the stoichiometric air-fuel ratio for diesel‘AFgas is the stoichiometric air-fuel ratio for methaneAFdsl = 1.45 * wtmol / yO2 / 13.8 / phidsiAFgas = 2 * wtmol / yO2 / 16 / phigasfuel$ = “diesel”fuel% = 0 ‘latch required for tracking previously formed burned gas zonesfuelswitch% = 0 ‘stores the increment # when fuel switches from dsl to CNGVi = Vcyl(ca(i))asurfi = Acyl(ca(i))mtot = matrap + mres + (dca / capwd) * mdslmu = matrap + mresTu = Pcyl(1) * Vi / Ru! muCALL unburned(Tu, hu, uu, Cvu, visc, 2)etot = uu * muetoti = uu * muThoi = TuVboi = VIPboi = Pcyl(i)Po = Pcyl(l)P(1) = Pcyl(i)145Tfi = 325 ‘injected fuel temperature in KTg = TfiCALL gas(Tg, Cpgi, ug, 2)Rg = .51835 ‘gas constant for methane in units U/kg Kmwgas = 16.04 ‘molar mass of methanemwdsl = 148.6 ‘molar mass of C10.8H 18.7mwmixed = 28 ‘approximated molar mass of mixed zone‘calculate enthalpy of gas and diesel at injection temp‘enthaply equation for methane integrated from Cp from vW&S pg.652‘enthapy equation for diesel from Heywood pg.132hgas = (-74873 -672.87 * Tfi + 111.246 * Tfi A 1.25 - .449496 * Tfi A 1.75 ÷ 6477.6 * SQR(Tfi)- 39442.3)!mwgas ‘[U/kg]hdsl = (-3.81 * Tfi + .51666 * Tfi A 2 - 2.0047E-04 * Tfi A 3 + 3.3816E-08 * Tfi A 4 - 2167300 / Tfi - 209736)!mwdsl ‘[kJ/kgud = hdslThulk = Tu ‘bulk temperarure of all cylinder contentsdelmd = (dca! capwd) * mdsl ‘amount of diesel injected over dcamd = delmdIF capwd <> capw THENdelmg = (dca / (capw - capwd)) * mgas ‘amount of CNG injected over dcaELSEdelmg 0END IFPSET (ca(1), 0), 11FOR i% =2 TO nca%Determine the mass of fuel present and its enthalpyIF dca * i% <= capwd THENmd = (dca * i% / capwd) * mdsl - xmf * (mdsl + mgas)mg = 0mtot = matrap + mres + (dca * i% / capwd) * mdslHd = hdsl * delmdHg = 0ELSEIF dca * i% > capwd AND dca * i% <= capw THENIF xmf < (mdsl / (mdsl + mgas) - xudsl) THENmd mdsl - xmf * (mdsl + mgas)mg = ((dca * i% - capwd) / (capw - capwd)) * mgasELSEmd = xudsl * (mgas + mdsl)mg = ((dca * i% - capwd) / (capw - capwd)) * mgas - (xmf * (mdsl + mgas) - mdsL)END IFmtot = matrap + mres + mdsl + ((dca i% - capwd) / (capw - capwd)) * mgasHd = 0Hg = hgas * delmgELSEIF dca * i% > capw THENIF xmf < (mdsl / (mdsl + mgas) - xudsl) THENmd = mdsl - xmf * (mdsl + mgas)mg = mgasELSEmd = xudsl * (mgas + mdsl)mg = mgas - (xmf * (mdsl + mgas) - mdsl)END IFmtot = matrap + mres ÷ mdsl + mgasMd = 0Hg = 0END IFIF mg < xugas * (mgas + mdsl) THEN mg = xugas * (mgas + mdsl)146IF md <xudsl * (mgas + mdsl) THEN md = xudsl * (rngas + mdsl)IF dca * (i% = 1) > capw AND mg> mgprev THEN mg = mgprevIF dca * (i% - 1) > capw AND md > rndprev THEN rnd = mdprevmgprev = mg ‘used to prevent formation of fuel aftermdprev = md ‘xmf has reached its maximum valuemtot = matrap + mres + xmf * (mdsl ÷ rngas)md=Omg=OHd=0Hg=OV2 = Vcyl(ca(i%))vol=(V2+V1)/2asurf2 = Acyl(ca(i%))asurf = (asurf2 + asurfi) I 2‘smooth measured cylinder pressureIF xmf> 0 THENP(i%) = (Pcyl(i% - 1) + Pcyl(i%) + Pcyl(i% + 1)) I 3ELSEP(i%) = Pcyl(i%)END IFfind unburned zone propertiesCALL unburned(Tu, hu, uu, Cvu, visc, 3)gu = I + Ru / Cvu ‘isentropic exponentIF method = 1 THENTu = Tu * (1 + (gu - 1) / gu * (P(i%) - P(i% - 1)) / P(i% - 1)) + dqwl / (mtot * (Ru + Cvu))ELSETu = Tu * (1 + (gu - 1)! gu * (P(i%) - P(i% - 1)) / P(i% - 1))END IFIF Tu <300 ThEN Tu = 3(X)vu = Ru * Tu / P(i%)CALL unhurned(Tu, hu, uu, Cvu, visc, 2)find mixed zone propertiesIFTm <>0 THENgm 1.3 ‘approximated isentropic exponentvmprev = 8.3144 I mwmixed * Tm / P(i% - 1)Tm = Tm * (1 + (gm - 1)/gm * (P(i%)- P(i% - 1)) / P(i% - 1))IF Tm <300 THEN Tm = 300vm 8.3144 / mwmixed * Tm / P(i%)urn urn - (P(i%) + P(i% - 1)) / 2 * (vm - vrnprev)END IFfind natural gas zone propertiesIF mg > xugas * (mgas + mdsl) AND Tg> 3(X) THENIF delrng I mg < I THENCALL gas(Tg, Cpg, ug, 2)Tg = Tg + (delmg / mg * (Cpgi + Cpg) / 2 * (Tfi - Tg) + Rg * Tg * (P(i%) - P(i% - 1)) / P(i% - 1))/ Cpgvg = 8.3144 * Tg / 16.04 / P(i%)IF Tg> 300 ThEN CALL gas(Tg, Cpg, ug, 3)END IFEND IFupdate previously formed burned zone properties147IF mbiml <>0 THENIF (i% - fuelswitch%) = 1 THEN fuel$ = ‘dieselCALL table(Thiml, P(i%), h, sbiml, ubimi, vbirnl, dNO, NOimi, q8, 5)NOx = NOx ÷ dNO * vbiml * mbiml * MWexh / (mair + mgas + mdsl) * 10 ‘ 9END IFIF mbirn2 <> 0 AND brndzones% >= 2 ThENIF (i% - fuelswitch%) = 2 THEN fuel$ = “diesel”CALL table(Thim2, P(i%), h, sbim2, ubim2, vbim2, dNO, NOim2, q8, 5)NOx = NOx + dNO * vbim2 * mbim2 * MWexh / (mair + mgas + mdsl) * 10 A 9END IFIF nibim3 <>0 AND brndzones% >= 3 THEN-IF (i% - fuelswitch%) = 3 THEN fuel$ = “diesel”CALL table(Thim3, P(i%), h, sbim3, ubim3, vbirn3, dNO, NOim3, q8, 5)NOx = NOx + dNO * vbim3 * mbim3 * MWexh / (mair + mgas + mdsl) * 10 A 9END IFIF mbim4 <> 0 AND brndzones% >= 4 THENIF (i% - fuelswitch%) = 4 THEN fuel$ = “diesel”CALL table(Thim4, P(i%), h, sbim4, ubim4, vbim4, dNO, NOim4, q8, 5)NOx = NOx + dNO * vbim4 * mbim4 * MWexh / (mair + mgas + mdsl) * 10 A 9END IFIF mbim5 <>0 AND brndzones% >= 5 THENIF (i% - fuelswitch%) = 5 THEN fuel$ = “diesel”CALL table(Thim5, P(i%), h, sbim5, ubim5, vbim5, dNO, NOimS, q8, 5)NOx = NOx + dNO * vbim4 * mbim5 * MWexh / (mair + mgas + mdsl) * 10 A 9END IFIF mbitn6 <> 0 AND brndzones% >= 6 THENIF (i% - fuelswitch%) = 6 THEN fuel$ = “diesel”CALL table(Tbim6, P(i%), h, sbimó, ubim6, vbim6, dNO, NOimô, q8, 5)NOx = NOx + dNO * vbim6 * mbim6 * MWexh / (mair + mgas + mdsl) * 10 A 9END IFcalculate workdwrk = (V2 - Vi) * (P(i% - 1) + P(i%)) /2‘identify the combusting fuelIF mgas = 0 OR xmf < (mdsl / (mdsl + mgas) - xudsl) THENfuel$ = “diesel”ELSEfuel$ = “gas”END IFIF method = 1 THEN‘Calculate heat transfer to cylinder wallsIF ca(i%) < 180 THENpolyexp = 1.3 ‘polytropic exponent for motored comp/expansionELSEpolyexp = 1.4END IFPo = Po * (Vi / V2) A polyexp ‘estimated motored engine PIF Po> P(i%) THEN Po = P(i%)CALL qwall(Thulk, P(i%), Po, vol, asurf, dqwl, 2)cylinder energy balanceetot etot - dwrk + dqwl + 1-Id + Hgprepare for burned gas properties iteration148vc = (V2 - mg * (vg- vu) + md * vu- mbiml * (vbiml - vu)- mbim2 * (vbim2- vu) - rnbim3 * (vbim3 -vu)- mbim4 * (vbim4- vu)- mbim5 * (vbim5- vu) - mbim6 * (vbim6 - vu) - mm * (vm - vu)) / mtotuc = (etot - mg * (ug- uu) - md * (ud- uu) - mbiml * (ubimi- uu) - mbim2 * (ubim2 - uu)- mbim3 *(ubim3 - uu) - mbim4 * (ubim4 - uu) - mbim5 * (ubim5 - uu) - mbim6 * (ubirn6- uu) - mm * (urn- uu)) / mtotCALL burned(P(i%), Thi, vu, vc, vbi, uu, uc, ubi, sbi, xmbi, dNO, NOe, 2)mbi = xmbi * mtot ‘mass of the ith burned gas zoneIF mbi <0 AND xmf <= 0 ThEN mbi =0ELSEcalculate the adiabatic flame temperatureCALL tad(P(i%), Thi, vu, vbi, hu, ubi, sbi, dNO, NOe, 2)cylinder energy balanceetot2 md * ud + mg * ug + mu * uu + mm * urn + mbi * ubi + mbiml * ubimi + mbim2 * ubim2 +mbim3 * ubim3 + rnbim4 * ubim4 ÷ mbim5 * ubim5 + mbim6 * ubim6dqwl = etot2 - etoti + dwrk-lid- Hg‘mass of the ith burned gas zonembi = (V2 - rntot * vu- rng * (vg - vu) + md * vu - rnbiml * (vbiml - vu)- mbim2 * (vbim2 - vu) -mbim3 * (vbim3 - vu)- mbim4 * (vbim4- vu)- rnbim5 * (vbim5 - vu)- rnbim6 * (vbim6- vu) -mm * (vm - vu)) / (vbi - vu)END IFmb mb + mbi ‘total cummulative mass of burned gas‘Calculate the mass fraction of fuel burnedIF mgas = 0 OR mb < (indsi - (mdsl + mgas) * xudsl) * (1 + AFdsl) THENxmf = mb I (mdsl + mgas) / (1 + AFdsl)ELSExmf = (mb + (mdsl - (mdsL + mgas) * xudsl) * (Afgas - AFdsl)) / (mdsl + mgas) / (1 + AFgas)EN]) IFNOi = dNONOx = NOx + dNO * vbi * mbi * MWexh /(mair + mgas + mdsl) * 10” 9mu = mtot - md- mg - mbi - mbiml- mbim2 mbim3 - mbim4 - mbim5 - mbim6- mmThulk = (mbi * Thi + mbiml * Thiml + mbim2 * Thim2 + mhim3 * Thim3 + mbim4 * Thim4 + mbim5 *Thim5 + mbim6 * Thim6 + mg * Tg + md * Tfi + mu * Tu + mm * Tm) / rntotPRINT USING “###ø P=####kPa Tu=####K Thi=####K Tg=####K NOx=###ppm xmm=#.##xmf=#.##”; ca(i%); Pcyl(i%); Tu; Tbi; Tg; NOx; mm / mtot; xmfPRINT USING “###ø P=####kPa Thi=####K Thinil=####K Thim2=##K NOx=I##hIppm xmf=#.##”;ca(i%); Pcyl(i%); Thi; Thimi; Thim2; NOx; xmfPRINT USING “xmu=#.### xmm=#.### xmb=#.### xmg=#.### xmd=#.### sum=#.##’; mu / mtot; mm/ mtot; (mbi + mbiml + mbim2 ÷ mbim3 + rnbim4) / mtot; mg / mtot; md / rntot; (mu ÷ mm + mbi + mbiml +mbim2 + mbim3 + mbim4 + mg + md) / mtotIF oputfile$ = “y” OR oputfile$ = “1” THENPRINT #1, USING” ### #### #### #### #### #### #### #### #.## #.###.## #.## ### ##““; ca(i%); Pcyl(i%); Tu; Thi; Tm; Tg; Thulk; NOx; xmf; mu / mtot; mm / mtot; md Imtot; mg / —mtot; dqwlEND IFIF dca * i% C1NT(capw) ThENLINE -(ca(i%), xmt), 14ELSEIF fuelS = “gas” THENLINE -(ca(i%), xmf, 10149ELSELINE -(ca(i%), xmf), 11END IFIF fuel$ = gas” AND fuel% = 0 THEN ‘set latchfuel% = 1fuelswitch% = i%END IFCalculate the properties of the mixed zone after mixingAdiabatic, constant pressure mixing in the mixed zoneAapproximate the mixed zone temperature using a mass averaged valueIF brndzones% = 1 THENmbilast = mbimlubilast = ubimivbilast = vbimlTbilast = ThimiELSEIF brndzones% = 2 THENmbilast = mbim2ubilast = ubini2vbilast = vbim2Thilast = Thim2ELSEIF brndzones% = 3 THENmbilast = mbim3ubilast = ubirn3vbilast = vbim3Thilast = Thim3ELSElF brndzones% =4 THENmbilast = mbim4ubilast ubim4vbilast = vbim4Thilast Thirn4ELSEIF brndzones% = 5 ThENmbilast = mbim5ubilast = ubim5vbilast = vbim5Thilast = Thim5ELS ElF brndzones% =6 THENmbilast = mbimóubilast = ubim6vbilast = vbim6Thilast = Thim6END IFmui = mixratio * mbilast‘mui is the mass of unburned zone that mixes with last burned zoneIF mu <= 0 THEN mui = 0IFmui > mu THEN mui =muIF rnbilast <>0 OR miii <> 0 OR mm <> 0 THENhm = um + P(i%) * vmhb = ubilast + P(i%) * vbilastTm = (Thilast * mbilast + Tu * mui + Tm * mm) / (mbilast + mui + mm)urn = (hb * mbilast + hu * mui + hm * mm) / (mbilast + mui + mm) - 8.3144 / mwmixed * TmEND IFmm = mm + mbilast + muiprepare for next stepVi = V2150etoti = etot2IF brndzones% = 6 THENmbimô = mbim5Thim6 = ThimSsbim6 sbim5NOim6 = NOim5END IFIF brndzones% >= 5 THENmbim5 = mbim4Thim5 = Thim4sbim5 = sbim4NOim5 = NOim4END IFIF brndzones% >= 4 THENmbim4 = mbim3Thim4 = Thim3sbim4 = sbim3NOim4 = NOim3END IFIF brndzones% >= 3 THENmbim3 = mbim2Thim3 = Thim2sbim3 = sbim2NOim3 = NOim2END IFIF brndzones% >= 2 THENmbim2 = mbimlThim2 = Thimisbim2 = sbimlNOirn2 = NOimiEND IFmbiml mbiThirni = ThisbimL = sbiNOimi NOiNEXT i%PRINT “Cl = “; Cl.CLOSE #1ENDFUNCTION Acyl (Ca)Calculates the cylinder surface area for a given degree capi# = 3.14159265359#Apisten = (pi# / 4) * bore “2car = ca * pi#/ 18()z = rod + (stroke / 2) * (1 + COS(car)) - SQR((rod) “2 - ((stroke / 2) * SIN(car)) “2)Acyl = z * pi# * bore + 2.5 * ApistonEND FUNCTIONSUB burned (P, Th, vu, vc, vb, uu, uc, ub, sb, xmbi, dNO, NOe, 1%) STATICSHARED fuel$IF1%= 1 THENCALL table(ql, q2, q3, q4, q5, q6, q7, q8, q9, 1)xmbi = 0END IFIF 1% =2 THENFind the linear relationship between ub and vh at the flame front151a = (vc vu) / (uc - uu)b = vu - a * uu‘Iterate to find Tb, ub, and vb.TIow = 1500Tint = 200iter% = 0found$ ‘falserangeS = “okay”DOiter% = iter% + ITup = Tiow + TintCALL table(Tlow, P, hhl, sbl, ubl, vbl, dNOl, 0, NOel, 3)CALL table(Tup, P. hbu, sbu, ubu, vbu, dNOu, 0, NOeu, 3)yl = vbl- (a * ubi + b)yu=vbu-(a * ubu+h)IFyl=OTHENfound$ = “true’Tb = TlowELSEIFyu=OTHENfound$ = “true”Tb = TupELSEIF yl * yu < 0 THENTint = Tint / 10IFTint <2 THENfound$ = “true”Tb = (Tiow + Tup) / 2END IFELSETow = TupEND IFIF Tiow = 2900 THEN range$ “exceeded”LOOP UNTIL found$ “true” OR range$ = “exceeded”IF vb <> vu AND range$ = “okay” THENCALL table(Th, P, hb, sb, ub, vb, dNO, 0, NOe, 4)xmbi = (vc - vu) / (vb - vu)ELSExrnbi = 0dNO =0END IFEND IFEN]) SUBSUB cubics (np%, xO, yO, xset, ycaic)np% is the number of x,y data pairsDIM d(10), e(10), f(10), g(I0)= np% - 1mm% = np% - 2CALCULATION OF SECOND DERIVATIVES g(i%)g(1) = 0g(np%) = 0FOR i% =2 TO m%d(i%) = x(i%) - x(i% - 1)e(i%) = 2 * (x(i% + 1) - x(i% - 1))f(i%) = x(i% + 1) - x(i%)g(i%) = 6 / f(i%) * (y(i% + 1) - y(i%)) + 6 / d(i%) * (y(i% - 1) - y(i%))NEXT i%FORi%=2TOmm%152fa d(i% + 1) / e(i%)e(i% + 1) = e(i% + 1) - Ia * f(i%)g(i% + 1) = g(i% + 1)- fa * g(i%)NEXT 1%FOR i% =2 TO rn%g(np% + 1- i%) = (g(np% + 1 - i%) - f(np% + 1 - i%) * g(np% + 2 - i%)) / e(np% + 1 - i%)NEXT i%CALCULATION OF INTERPOLATED VALUE ycaic AT x=xsetd(np%) = x(np%) - x(np% - 1)i% = 1DOi% = i% + 1LOOP WHILE (xset >= x(i%)) AND (i% < np%)deim = xset- x(i% - 1)deip = x(i%) - xsetycaic = g(i% - 1)! 6 / d(i%) * deip ‘ 3 + g(i%) / 6/ d(i%) * delrn” 3 + (y(i% - 1)! d(i%) - g(i% - 1) * d(i%) /6) * deip + (y(i%) / d(i%) - g(i%) * d(i%) / 6) * deimEN]) SUBFUNCTION Cvair (Tair)‘Cvair has units kJ/kmol KTdiin = Tair / 100Tdim2 Tdim * TdimTdim3 = Tdim2 * TdimTdim32 = Tdim 1.5Cp02 = 37.432 + .020102 * Tdim32 - 178.57 / Tdirn32 + 236.88 / Tdim2CpN2 = 39.06 - 512.79 / Tdim32 + 1072.7 / Tdim2 - 820.4 / Tdirn3Cvair = .21 * Cp02 + .79 * CpN2- 8.3144END FUNCTIONFUNCTION Cvres (Tres) STATIC‘Cvres has units kJ/kmol KTdim = Tres / 100Tdimsq = SQR(Tdim)Tdim2 = Tdim * TdimTdirn3 =Tdim2*TdimTdiml4 = Tdim .25Tdirn32 = Tdim” 1.5Cp02 = 37.432 + .020102 * Tdim32 - 178.57 / Tdim32 + 236.88 / Tdim2CpN2 = 39.06- 512.79 / Tdim32 + 1072.7 / Tdim2 - 820.4 / Tdirn3CpH2O = 143.05 - 183.54 * TcIirnl4 + 82.75 1 * Tdimsq - 3.6989 * TdimCpCO2 = -3.7357 + 30.529 * Tdimsq - 4.1034 * Tdim + .024198 * Tdirn2Cvres = yO2r * Cp02 + yN2r * CpN2 + yCO2r * CpCO2 + yH2Or * CpH2O - 8.3144END FUNCTIONSUB gas (Tg, Cpg, ug, 1%) STATIC1% = 1---> INITIALIZE SUBROUTINE1% =2---> CALCULATE SPECIFIC HEAT U/kg K1% =3---> CALCULATE INTERNAL ENERGY kJ/kgIF 1% = I THENmwCH4 = 16.04hfCH4 = -74873ELSETdim=Tg/100Tdimsq = SQR(Tdim)Tdiml4 = Tdim” .25153Tdim54 = Tdim 125Tdim74 = Tdim” 1.75Tdim34 = Tdim ‘ .75END IFIF 1% = 2 THENCpg = (-672.87 + 439.74 * Tdiml4 - 24.875 * Tdim34 + 323.88 / Tdimsq) / mwCH4ELSEIF 1% = 3 THENdhCFI4 = -67287 * Tdim + 35179.2 * Tdim54 - 1421.43 * Tdim74 + 64776 * Tdimsq - 39442.3ug = (hfCH4 + dhCH4 - 8.3144 * Tg) / mwCH4END IFEND SUBSUB qwall (Thulk, P. Po, vol, asurf, dqwl, 1%) STATICCalculates the heat transfer from the gas to the cylinder wallSHARED a, dca, rpm, Thoi, Vboi, Pboi, Cl, C2IF1%= 1THENdqwl = 0INPUT “Enter gas velocity/mean piston speed proportionality factor: ‘, ClC2 = Cl / 700 ‘gas velocity/combustion intensity proportionality factorpi# = 3.14159265359#pisvel = rpm * stroke / 30 ‘mean piston speedELSEIF 1% =2 THENgasvel = c * pisvel + C2 * vol * Tboi / Pboi / Vboi * (P - Po)dens = mtot / volCALL unburned(Tbulk, ho, uu, Cvu, visc, 4)Renum = dens * gasvel * bore / viscCALL unburned(Thulk, hu, uu, Cvu, visc, 3)Cpg = Cvu + Ruthcond Cpg * visc / .7Tw = 450 ‘The wall temperature (K) is assumed to be constantdqwl = asurf * thcond / bore * Renum (.8) * (Tw - Thulk)Convert from kW to kJdqwl = dqwl * 60 * dca / (rpm * 360)END IFEND SUBSUB table (Thx, Pbx, hbx, sbx, ubx, vbx, dNO, NO, NOe, 1%) STATIC1% = 1 ---> INITIALIZATION1% =2 ---> FIND BURNED GAS ENTHALPY GIVEN P AND Th1% = 3 ---> FIND BURNED GAS PROPERTIES GIVEN P AND Th(excludes NO calculation for more efficient iterating)1% =4 ---> FIND BURNED GAS PROPERTIES GIVEN P AND Th(includes NO calculation)1% =5 ---> FIND BURNED GAS PROPERTIES GIVEN P AND sb(includes NO calculation)SHARED fuel$DIM Ptab(10), Tbtab(10), amdtab(10, 10), amgtab(i0, 10)DIM ubdtab(10, 10), sbdtab(10, 10), ubgtab(10, 10), sbgtab(10, 10)DIM cN2dtab(10, 10), cO2dtab(10, 10), cOHdtab(10, 10)DIM cN2gtab(10, 10), cO2gtab(10, 10), cOllgtab(10, 10)DIM cNOdtab(10, 10), cNOgtab(10, 10), hbdtab(10, 10), hbgtab(10, 10)DIM ya(10), yu(10), ys(10), ycN2(10), yCO2(10), ycOH(10), ycNO(10)DIM amp(l0), up(10), sp(10), N2p(10), O2p(lO), OHp(10), NOp(I0)IF 1% = I THENPstore = -10(X)Read tables of burned gas properties resulting from154stoichiometric combustion of diesel or natural gaswith air.suffix “g” --> burned gas properties for NG combustionsuffix “d” --> burned gas properties for dsl combustionFORj% = 1 TO 10FOR i% = 1 TO 10INPUT #1, Thtab(i%), P, amdtab(i%, j%), vbd, ubd, hbd, sbd, yOl-ldtab, yNOdtab, yN2dtab,yO2dtabINPUT #2, Thtab(i%), P, amgtab(i%, j%), vbg, ubg, hbg, sbg, yOHgtab, yNOgtab, yN2gtab,yO2gtab‘convert from J/kg to KJ/kg for h and u‘convert from J/kg K to KJ/kg K for subdtab(i%,j%) = ubd * .001hbdtab(i%, j%) = hbd *sbdtab(i%, j%) sbd *ubgtab(i%, j%) = ubg *hbgtab(i%,j%) = hbg * .001sbgtab(i%,j%) = sbg *‘convert equilibrium mole fractions to equilibriumconcentrations in units mole/cin”3cN2dtab(i%, j%) = yN2dtab / vbd / amdtab(i%, j%) / 1000cO2dtab(i%, j%) = yO2dtab / vbd / amdtab(i%, j%) / 1000cOHdtab(i%, j%) = yOHdtab / vhd / amdtab(i%, j%) I 1000cNOdtab(i%, j%) = yNOdtab / vbd / arndtab(i%, j%) I 1000cN2gtab(i%, j%) = yN2gtab / vbg / amgtab(i%, j%) / 100()cO2gtab(i%, j%) = yO2gtab I vbg / amgtab(i%, j%) / 1000cOHgtab(i%, j%) = yOHgtah / vbg / amgtab(i%, j%) / 1000cNOgtab(i%, j%) = yNOgiab / vbg / amgtab(i%, j%) / 1000NEXT i%Ptab(j%) = P * 101.325 ‘pressure converted from atm to kPaNEXTj%ELSE! F Pbx <> Pstore ThENPstore = PbxFOR i% = 1 TO 10IF fuel$ = “diesel” THENFORj% = ITO 10yu(j%) = ubdtab(i%, j%)yh(j%) = hbdtab(i%, j%)ys(j%) sbdtab(i%, j%)ya(j%) = amdtab(i%, j%)ycN2(j%) = cN2dtab(i%, j%)yCO2(j%) = cO2dtab(i%, j%)ycOH(j%) = cOHdtab(i%, j%)ycNO(j%) = cNOdtab(i%,j%)NEXT j %ELSEFORj% = 1 TO 10yu(j%) = ubgtab(i%, j%)yh(,j%) = hbgtab(i%, j%)ys(j%) = sbgtab(i%, j%)ya(j%) = amgtab(i%, j%)ycN2(j%) = cN2gtab(i%, j%)yCO2(j%) = cO2gtab(i%,j%)ycOH(j%) = cOHgtab(i%, j%)ycNO(j%) = cNOgtab(i%, j%)NEXT j %END IF155CALL cubics(1O, PtabO, yuO, Pbx, up(i%))CALL cubics(10, PtabO, yhO, Pbx, hp(i%))CALL cubics(1O, PtabO, ysO, Pbx, sp(i%))CALL cubics(10, PtabO, yaO, Pbx, amp(i%))CALL cubics(10, PtabO, ycN2O, Pbx, N2p(i%))CALL cubics(1O, PtabO, yCO2O, Pbx, 02p(i%))CALL cubics(10, PtabO, ycOHO, Pbx, OHp(i%))CALL cubics(1O, PtabO, ycNOO, Pbx, NOp(i%))NEXT i%END IFIF 1% =2 THENCALL cubics(10, ThtabO, hpO, Tbx, hbx)ELSEIF 1% = 3 THENCALL cubics(10, ThtabO, upO, Thx, ubx)CALL cubics(10, ThtabO, spO, Thx, sbx)CALL cubics(10, ThtabO, ampO, Thx, amx)vbx=8.3144/amx *Thx/PbxELSEIF 1% =4 THENCALL cubics(10, ThtabO, upO, Thx, ubx)CALL cubics(10, TbtabO, spO, Tbx, sbx)CALL cubics(10, ThtabO, ampO, Thx, amx)vbx= 8.3144/amx * Thx/Pbx‘N2e, 02e, OHe, and NOe are equilibrium‘concentrations in mol/cmA3CALL cubics(10, ThtabO, N2p0, Tbx, N2e)CALL cubics(10, ThtabO,°2p0. Thx, 02e)CALL cubics(10, ThtabO, OHpO, Thx, OHe)CALL cubics(10, ThtabO, NOpO, Thx, NOe)‘calculate the change in NO concentration over one time stepdNO = (6E+16) / SQR(Thx) * EXP(-69090 / Thx) * SQR(02e) * N2e * ((1 - NO “ 2 / 20.267 / EXP(21650 / Thx) I 02e / N2e) / (1 + NO / (.0004 * Thx * EXP(-3150 / Thx) * 02e + 2.5625 * ORe))) * dt‘dNO and NOe are in units mol/cm”3ELSEIF 1% =5 THENCALL cubics(10, spa, ThtabO, sbx, Thx)CALL cubics(10, spO, upO, sbx, ubx)CALL cubics(10, spO, ampO, sbx, amx)vbx=8.3144/amx *Thx/PbxIFThx> 1200THEN‘N2e, O2e, OHe, and NOe are equilibriumconcentrations in mol/cm”3CALL cubics(10, spO, N2p0, sbx, N2e)CALL cubics(10, spO, O2p0, sbx, 02e)CALL cubics(10, spO, OHpO, sbx, OHe)‘calculate the change in NO concentration over one time stepdNO = (6E+16) I SQR(Tbx) * EXP(-69090 / Thx) * SQR(02e) * N2e * ((1 - NO “2! 20.267 /EXP(-21650 I Thx) / 02e/ N2e)/ (1 + NO /(.0004 * Thx * EXP(-3 150 / Thx) * 02e + 2.5625 * OHe))) * dtdNO is in units mol/cm”3END IFEND IFEN]) SUBSUB tad (P. Tb, vu, vb, hu, ub, sb, dNO, NOe, 1%) STATICSHARED fuel$, hdsl, hgas, phidsi, phigasIF1%= 1 THENCALL table(ql, q2, q3, q4, q5, q6, q7, q8, q9, 1)END IFIF 1% = 2 THEN156IF fuel$ = ‘diesel’ THENmwfuel = 14hfuel = hdsln=2Hreact = (hfuel * mwfuel ÷ 1w * (1 + n / 4)! yO2 / phidsi * wtmol) / (mwfuel + wtmol * (1 + n / 4)/yO2) ‘[kJ/kgJELSEmwfuel = 16hfuel = hgasn=4Hreact = (hfuel * mwfuel + hu * (1 + n / 4)! yO2 / phigas * wtmol) / (mwfuel + wtmol * (1 + ii / 4)IyO2) ‘[kJ/kgjEND IFTb = 2900Tint = 200iter% = 0found$ = “false”range$ = “okay”DOiter% = iter% + 1CALL table(Th, P, hb, sb, ub, vb, dNO, 0, NOe, 2)IF hb = Hreact THENfound$ = “true”ELSEIF hb < Hreact THENIFTint> 1 THENTh = Th + TintTint = Tint / 10ELSEfound$ = “true”END IFELSETh = Th - TintEND IFIF Th <= 1300 OR Tb> 3100 THEN range$ = “exceeded”LOOP UNTIL found$ = “true” OR range$ = “exceeded”IF range$ = “okay” THENCALL table(Th, P, hb, sb, ub, vb, dNO, 0, NOe, 4)ELSEPRINT “*** RANGE EXCEEDED IN FLAME TEMP CALC “END IFEND IFEND SUBSUB unburned (Tu, hu, uu, Cvu, visc, 1%) STATIC1% = 1 ---> IMTIALIZE SUBROUTINE1% =2---> CALCULATE ENTHALPY AND INTERNAL ENERGYl%=3---> CALCULATE Cv1% =4---> CALCULATE VISCOSITYIF 1% = 1 THEN‘molar massmwO2 = 31.999mwN2 = 28.013mwH2O = 18.015mwCO2 = 44.011enthalpy of formation [kJ/kmol]157hfO2 = 0hfN2 = 0hfH2O = -241827hfCO2 = -393522n02 = yO2r * nres + .21 * natrapnN2 = yN2r * nres + .79 * natrapnCO2 = yCO2r * nresnH2O = yH2Or * nresdenorn = n02 + nN2 + nCO2 + nH2Oy02 = n02 / denomyN2 = nN2 / denomyCO2 nCO2 / denomyH2O = nH2O I denomwtmol=yO2 * mwO2 + yN2 * rnwN2 + yCO2 * mwCO2 + y112() * mwH2ORu = 8.3 144 I wtmolELSETdim=Tu/100Tdimsq = SQR(Tdim)Tdim2 = Tdim * TdimTdim3 = Tdim2 * TdimTdim4 = Tdim3 * TdimTdiml4 = Tdim A .25Tdim54 = Tdim” 1.25Tdim74 = Tdim A 1.75Tdim32 = Tdim” 1.5Tdim52 = Tdim 2.5Tdim34 = Tdim A 75END IFIF 1% = 2 THEN‘Calculate molar enthalpy differences between 298K and Tu,then calculate the internal energy uu jki/kgjdhO2 = 3743.2 * Tdim + .80408 * Tdim52 + 35714 / Tdimsq - 23688 / Tdim - 23911dhN2 = 3906 * Tdim + 102558 / Tdirnsq - 107270 / Tdim + 41020 / Tdim2 - 39677dhH2O = 14305 * Tdim - 14683.2 * Tdim54 + 5516.73 * Tdirn32 - 184.95 * Tdim2 - 11881.33dhCO2 = -373.57 * Tdim + 2035.27 * Tdirn32 - 205.17 * Tdim2 + .8066 * Tdim3 - 7561.65hu=(yO2 * (hfO2 + dhO2) + yN2 * (hfN2 + dhN2) + yH2O * (hfH2O + dhH2O) + yCO2 * (hfCO2 +dhCO2)) / wtmoluu = hu - 8.3 144 * Tu / wtmolELSEIF 1% =3 ThEN‘Calculate specific heat Cvu [U/kg KICp02 = 37.432 + .020102 * Tdim32 - 178.57 / Tdim32 + 236.88 / Tdim2CpN2 = 39.06 - 512.79 / Tdim32 + 1072.7 / Tdim2 - 820.4 / Tdim3CpH2O = 143.05 - 183.54 * Tdiml4 + 82.751 * Tdimsq - 3.6989 * TdimCpCO2 = -3.7357 + 30.529 * Tdimsq - 4.1034 * Tdim + .024198 * Tdim2Cvu=(yO2 * Cp02 + yN2 * CpN2 + yCO2 * CpCO2 + yFl2O * CpH2O - 8.3 144) / wtrnolELSEIF 1% =4 THEN‘Estimate the mean viscosity of gas mixturesTn = Tu A .645visc=(yO2 * mwO2 * 5.09 + yN2 * mwN2 * 4.57 + yH2O * mwH2O * 3.26 + yCO2 * mwCO2 * 3.71)/ wtmol * 10 A (7) * TnEN]) IFEND SUBFUNCTION Vcyl (ca)‘ca is crank angle degrees ABDC. Vcyl(m**3).158pi# = 3.14159265359#car= ca* pi#/180Vcyl = pi# * ((bore / 2) “ 2) * (rod + (stroke / 2) * (1 + COS(car)) - SQR(rod “ 2 - ((stroke / 2) * SIN(car))2)) + VclrEND FUNCTION159


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