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Performance and combustion characteristics of a diesel-pilot gas injection engine Gunawan, Hardi 1992

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PERFORMANCE AND COMBUSTION CHARACTERISTICS OFA DIESEL-PILOT GAS INJECTION ENGINEbyHARDI GUNAWANB.Eng., Trisakti University (Indonesia), 1977A THESIS SUBMITTED 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 COLUMBIAJune 1992© Hardi Gunawan 1992In 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 of Mer,hanic4a Fhpe&Ifil.The University of British ColumbiaVancouver, CanadaDate jaint aci I tqli?„DE-6 (2/88)ABSTRACTHigh pressure injection of natural gas with diesel liquid pilot fuel has beeninvestigated in a single-cylinder two-stroke diesel engine of low compression ratio.The thermal efficiency and emissions of NO N, CO and HC were determined at 1200RPM as a function of load.Variations of the pilot-diesel energy to total heat energy ratio (in the range 15 -25%) strongly affected efficiency and emissions rate. Gas injection pressure was alsoshown to be an important variable.The thermal efficiency of the gas-diesel engine was shown to exceed that of theconventional diesel at full load, and is less at low load in the present configuration.The combustion rate analysis has been used to determine the pressure rise (ignitiontime) delay and combustion duration as well as the characteristic burning pattern.At low load late burning high cyclic variations, and incomplete combustion areassociated with peak compression temperature lower than 900 K.11TABLE OF CONTENTSABSTRACT^ iiLIST OF SYMBOLS^ xiiiLIST OF TABLES xvLIST OF FIGURES^ xviACKNOWLEDGEMENT xix1 INTRODUCTION^ 11.1^Background 11.2^Use of Natural Gas in Diesel Engine^ 41.3^Objective of This Research^ 71.4^Methodology^ 72 LITERATURE REVIEW^ 92.1^The Diesel Engine 92.2^Natural Fumigation Method^ 122.3^Timed Port Injection Method 152.4^Direct Injection Method^ 162.5^Analysis of Combustion Rate 172.6^Summary^ 23111iv3 EXPERIMENTAL APPARATUS^ 253.1^Introduction^ 253.2^Engine and Test Bed^ 283.3^Fuel System^ 313.3.1 Fuel Supply System^ 323.3.2 Fuel Injection System 333.4^Instrumentation^ 353.4.1 Engine Speed 363.4.2 Torque (Load Cell)^ 363.4.3 Gas Fuel Mass-flow Rate 363.4.4 Pilot Diesel Fuel Consumption^ 373.4.5 Emission Instrumentation 373.4.6 Cylinder Pressure^ 393.5^Data Acquisition System 404 EXPERIMENTAL PROCEDURE^ 434.1^Testing Procedure^ 434.2^Engine Performance Calculation Procedure^ 464.2.1 Thermal Efficiency^ 464.2.2 Brake Mean Effective Pressure (BMEP)^ 484.2.3 Brake Specific Emissions^ 48V5 MEASUREMENTS OF ENGINE PERFORMANCE^ 535.1 Conventional Diesel Performance^ 535.2 Natural Gas-Diesel Performance 575.3 Effects of Gas Injection Pressure^ 585.4 Effects of Diesel Ratio^ 635.5 Effects of Pilot-Diesel Cetane Number^ 665.6 Summary^ 686 COMBUSTION RATE ANALYSIS^ 706.1^General^ 706.2^Formulation of the Combustion Model^ 726.3^Mixture Composition^ 766.3.1 Scavenged Air 786.3.2 Residual Gas^ 816.3.3 The Unburned Gas Composition^ 826.4^Thermodynamics Properties of the Unburned Gas^ 836.5^Thermodynamics Properties of the Burned Gas 866.6^Heat Transfer to the Cylinder Wall^ 886.7^Calculation Procedure^ 906.8^Pressure Measurements 926.9^Indicated Work^ 95vi96981016.10 Cyclic Variation6.11 Mass-burned Fractions6.12 Summary7 CONCLUSIONS AND RECOMENDATIONS9.10 Appendix J - Iteration Procedure9.11 Appendix K - Mass-burned Fraction Program Listing103103105107111111113118119120121122123126129131^7.1^Conclusions7.2^Recommendations8 REFERENCES9 APPENDICES9.1^Appendix A - Diesel Engine Emission Standards9.2^Appendix B - Calibration Curves9.3^Appendix C - Properties of Test Fuels9.4^Appendix D - Composition of Natural Gas9.5^Appendix E - Exhaust Sampling Probe in the Engine Exhaust Pipe9.6^Appendix F - Schematic of the Fuels Injector9.7^Appendix G - Power Correction Factor Calculation Method9.8^Appendix H - Determination of the Cylinder Volume9.9^Appendix I - Determination of Fuel-air Equivalence Ratiovii9.12 Appendix L - Verification of the Computation Procedurefor the Constant Volume Case^147LIST OF SYMBOLSa^constantbsco brake specific of carbon monoxidebso^brake specific emissionsbsfic brake specific of unburned hydrocarbonbsNox brake specific of oxides of nitrogenB^Cylinder boreBMEP Brake mean effective pressureBdo^Dry inlet air pressureBo^Inlet air pressurec^clearance heightCH4 MethaneCO Carbon monoxideCO2 Carbon dioxideCp^Molar specific heat at constant pressureC,^Molar specific heat at constant volumeD^Engine displacementDP^Degree of purity of the chargeEtot^Total energy of the systemf^fractionviiiixfa^Atmospheric factorFA^Fuel-air ratioFA.), Stoichiometric fuel-air ratiofe.,^Correction factor to observed brake horse powerfai^Engine factorhe^Convective heat transfer coefficientEnthalpy of formation at standard state conditionH^Enthalpy112^HydrogenH2O WaterHC Unburned hydrocarbonk^Thermal conductivityLHV Lower heating value of fuelsm^massmad^Delivered air mass to the engine per cyclemale^Mass of inducted airMat,^Mass of air trapped in the cylinder per cycleMdsl^Mass of diesel fuel injected into the combustion chamber per cycleme^Mass of emissionsmep Mean effective pressuremexh Mass of exhaust gasme^Mass of fuelmg. Mass of natural gas injected into the combustion chamber per cycleRot^Mass of cylinder charge during combustion processmfr^Trapped massM^Molecular weightn^Polytropic exponentHydrogen-to-carbon atomic ratioMole numberN^Engine speedN2^NitrogenN„^Mole number of trapped chargeNO Nitric oxideNO, Nitrogen dioxideNO Oxides of nitrogenNi,^Nusselt number02^Oxygenp^Distance between the piston pin to the piston top surfaceP^PressurePowerP,^Initial pressurePb^Pressure of cylinder at any crank position during combustionPer^Corrected pressurexiPc3, 1^Cylinder pressureP.^Maximum-measured pressurePM^Particulate MatterPee., Maximum pressurePmeas Measured cylinder pressurePinot^Polytropic compression or expansion pressure obtained without combustionPo^Inlet manifold pressurePs^Pressure of cylinder at beginning of combustionQ„^Heat transfer from the cylinder contents to the wallr^Crank radiusInlet manifold to inlet air pressure ratioDiesel-to-gas mass ratiore^Compression ratioR^Universal gas constantRa^Air gas constantRe^Reynolds numberRPM Revolution per minuteSp^Mean piston velocityt^timetemperature (in degree Celcius)T^TemperatureTb^Engine torque outputxiiTm^Mass-average gas temperatureT„,^Cylinder wall temperatureu^Specific internal energy of the systemU0^Internal energy of the cylinder contents at a reference pointv^Specific volumeV^VolumeVl^Initial volumeVe^Cylinder clearance volumeVd^Displacement volumeVipc^Cylinder volume when piston at inlet port closure positionVmax Maximum volumev,a Pa Specific volume at ambient temperature and pressure,vo^Specific volume at temperature t and pressure PW^Work done on the pistonWe^Work produced per cyclex^Mass-burned Fractionxmax Maximum mass-burned fraction assuming no heat transferY^Mole fractionGreek LettersSpecific heat ratioScavenged-blower thermal efficiencylibth^Brake thermal efficiency11th^Thermal efficiencyCrank angleA^Delivery ratioViscosity of the mixturep^Density of the mixturePt,P^Density of air at temperature t and pressure PFuel-air equivalence ratioSubscriptabox Air boxamb Ambientb^Burned gascf^Correction factori^Species ires^Residual gastr^trappedu^Unburned gasAbbreviationsABDC After Bottom Dead CentreATDC After Top Dead CentreBDC Bottom Dead CentreBOI Beginning of fuels injectionBTDC Before Top Dead CentreCA Crank AngleCHEMI ChemiluminiscenseEPA Environmental Protection AgencyEVC Exhaust Valve CloseEVO Exhaust Valve OpenFID Flame inoization detectorIPC Inlet Port ClosureLFE Laminar Flow ElementLNG Liquified Natural GasNDIR Non-dispersive InfraredPW Pulse widthTDC Top Dead CentrexivLIST OF TABLESTable 3.1^Engine Specifiations^ 28Table 4.1^Testing Ranges^ 44Table 5.1^Effect of Gas Injection Pressure on Unburned Gas Ratio^62Table 6.1^List of Temperatures and Pressures usedin Burned Gas Properties Table^87Table A.1^Urban Bus Heavy-duty Engine Emission Standard^111Table A.2^Heavy-duty Truck Engine Emission Standard 112Table C.1^Properties of Test Fuels^ 118Table D.1^Composition of Natural Gas^ 119XVLIST OF FIGURESFigure 1.1^Status of Conventional Diesel Emissionstowards US EPA Standards 2Figure 1.2^Methods of Using Natural Gas in Diesel Engines^4Figure 2.1^Graphical Method of Estimated Mass-burned Fraction^19Figure 2.2^Determination of Combustion Pressure fromMeasured Pressure 20Figure 3.1^Schematic of Experimental Apparatus and Instrumentation^26Figure 3.2^Test Engine and Dynamometer in a Cell^ 29Figure 3.3^Engine Control Panel and Data Acquisition System^30Figure 3.4^Schematic of Fuel Supply System^ 32Figure 3.5^Schematic of Fuel Injection System 34Figure 3.6^Exhaust Emission Sampling System^ 38Figure 3.7^Pressure Transducer Location in the Cylinder Head^39Figure 3.8^Data Flow in the Data Acquisition System^41xvixviiFigure 5.1^Diesel Baseline Performance Curve^ 54Figure 5.2^Diesel Oxides of Nitrogen Emissions 55Figure 5.3^Diesel Unburned Hydrocarbon Emission^ 56Figure 5.4^Determination of Best BOI Performance 57Figure 5.5^Effect of Gas Injection Pressure on Thermal Efficiency^59Figure 5.6^Effect of Gas Injection Pressure on Oxides of Nitrogen^60Figure 5.7^Effect of Gas Injection Pressure on Unburned Hydrocarbon^61Figure 5.8^Effect of Diesel Ratio on Thermal Efficiency^63Figure 5.9^Effect of Diesel Ratio on Oxides of Nitrogen 64Figure 5.10^Effect of Diesel Ratio on Unburned Hydrocarbon^65Figure 5.11^Effect of Pilot-Diesel Cetane Number on Thermal Efficiency^66Figure 5.12^Effect of Pilot-Diesel Cetane Number on NO„^67Figure 5.13^Effect of Pilot-Diesel Cetane Number on unburned HC^68Figure 6.1^Typical Cylinder Pressure Data^ 71Figure 6.2^Schematic of the Combustion Model 72Figure 6.3^Schematic of the Uniflow-scavenged Configuration^73Figure 6.4^Scavenging Data Typical of Large Two-stroke Diesels^81Figure 6.5^Typical Mass-burned Fraction Results^ 88Figure 6.6^Typical Pressure Crank-Angle Diagram at different Loads^93Figure 6.7^Typical Log P - Log V Diagram^ 94Figure 6.8^Comparison of Indicated Work to Brake Work^96xviiiFigure 6.9^Standard Deviation and Relative S.D. of the Brake Work^97Figure 6.10^Superposition of 10 Successive Cycles ,^98Figure 6.11^Cylinder Pressure, Temperature and Mass-burned FractionDistribution^99Figure 6.12^Combustion Pattern at different Load^ 100Figure 6.13^Combustion Pattern at different Gas Injection Pressure^101Figure B.1^Speed Calibration Curve^ 113Figure B.2^Torque Calibration Curve 114Figure B.3^Diesel Fuel Mass-flow Calibration Curve^115Figure B.4^Cylinder Pressure Transducer Calibration Curve^116Figure B.5^Laminar Flow Element Calibration Curve 117Figure E.1^Exhaust Sampling Probe in the Engine Exhaust Pipe^120Figure F.1^Schematic of the Two-fuel Unit Injector^121Figure H.1^Piston Top Surface Profile^ 123Figure H.2^Geometry of the Cylinder 124ACKNOWLEDGEMENTSI wish to express my sincere gratitude to Dr. P.G. Hill for his invaluable guidance,help and encouragements throughout all phases of the project and the writing of this thesis.Thanks are due to Bruce Hodgins, research enginer and project manager, for hisadvice and assistance in experimental work. Thanks also are due to Dale Nagata for thehelp in the data acquisition system, and all the technical staff of the MechanicalEngineering Department for the help in construction of the experimental apparatus.Special thanks to my fellow graduate students for their helpful advice, suggestions,to Yinchu Tao for his help performing the experiments, and to Patric Ouellette who helpedtaking the pictures for this thesis.A very special thanks to my wife, Lenny, and all the children (Novi, Yani, Andrew,and Lita), who were a constant source of moral support and motivation during my graduatework.Financial support for this work by the Eastern Indonesia University DevelopmentProject - Simon Fraser University is gratefully acknowledged.xix1. INTRODUCTION.A 'diesel-pilot gas injection engine' is defined as a diesel engine running onhigh-pressure gaseous fuel which is directly injected into the combustion chamber togetherwith a small portion of diesel fuel to initiate the combustion. This method of using naturalgas in diesel engines has been used before but not with the same injector for both fuels.Note that a 'diesel-pilot gas injection engine' is not a 'dual-fuel' engine since it can not runon either diesel or gas alone. It may operate with a wide range of gas/diesel fuel ratios.The general purpose of this investigation was to study the performance, and thecombustion and emission characteristics of a single-cylinder diesel-pilot gas injectionengine. The intent was to provide a method for investigating the effect of injector designon engine performance and emissions in an effort to meet the new emissions regulationsfor heavy duty bus engines.1.1. Background.Since its invention about 100 years ago, the diesel engine has become the mostefficient prime mover that has found wide application. It is used for stationary purposes,such as in pumping stations, as well as in land and sea transportation.The use of diesel engines in urban buses and trucks is facing stricter environmental12regulations than in the last decade. A new set of emission standards for urban busheavy-duty engines, and heavy-duty truck engines has been established by theEnvironmental Protection Agency (EPA) of the United States (Appendix A). In 1993, forurban bus engine these standards require a reduction in allowable oxides of nitrogen (NO,)and particulate matter (PM) to a level that is less than 5.0 g/bhp-hr and 0.1 g/bhp-hr,respectively. For trucks, the same emission concentration levels will be applied in 1994.An indication of what this means is shown in Fig. 1.1 which illustrates the typical statusof a conventional diesel engine emissions in terms of its PM and NO. concentrationtowards the legislated standards.0 . 6 m.11■1■11M,NO: - PARTICULATE TRADE-OFF Production0 . 5 - HEAVY DUTY TRANSIENT CYCLE Engines.2 . 4 -o..cca 0 . 3 - •••■ loftst-•••■•.. •••••'^0 . 2 -1994 Legislated Standard•••■• •■■■• ammoa •■••• May...11s.00. 1 - DevelopmentOEM =NM Ogia 411■1 Engines0 .01^2^3^4^5^6^7NOx (g/bhph)Figure 1.1 : Status of Conventional Diesel Emissions relative toU.S EPA Emissions Standard. 03Particulate matter consists of soot on which some organic compound have becomeabsorbed. Most of it results from incomplete combustion of fuel hydrocarbons. The primaryconstituent of NO is nitric oxide (NO) which forms throughout the high-temperatureburned gases behind the flame through chemical reactions involving nitrogen and oxygenatoms and molecules, which do not attain chemical equilibrium. The other NO constituentis nitrogen dioxide (NO2) which is rapidly converted from NO in the flame zone.As can be seen in Fig. 1.1, the new emissions standards are quite challenging to thediesel engine manufacturer. With present technology, it seems that it will be difficult tomeet the EPA 1993/4 particulates and NOx emission standards with conventional dieselengines. Alternative solutions should be developed if diesel engines are to power futureurban buses and trucks. One possible solution is to use natural gas as an alternative fuelfor the diesel engine. Natural gas fuelling can solve the diesel particulate emission problem;it can also offer the flexibility to apply NOx reduction strategies which will not work withconventionally-fuelled diesels. Besides its potential for reducing particulate emissions,natural gas is preferred because of its ready availability, and relatively low cost. However,natural gas has the disadvantage of low volumetric heat content even when compressed asa gas to 20 MPa. The primary drawback to the use of compressed natural gas intransportation is that massive and bulky tanks are required to store enough compressednatural gas for reasonable vehicle range. This fuel storage problem puts a high value onincreasing efficiency with natural gas.41.2. Use of Natural Gas in Diesel Engines.There are three principal methods of using natural gas in diesel engines as shownschematically in Fig. 1.2. In the following a brief explanation of each method will bepresented together with a short description of previous experiences. A more detaileddescriptions of the combustion characteristics and previous work on these topics will begiven in Chapter 2.The first method is known as natural fumigation. As shown in Fig. 1.2a, the naturalgas is injected into the inlet manifold. It mixes naturally with the inducted air, forms a fullypre-mixed fuel-air mixture which then enters the engine. Thus in the compression strokethis mixture will be compressed instead of air alone. This situation does not favor thetwo-stroke engine due to a considerable fuel loss which would occur in the scavengingprocess. The advantage of the natural fumigation method is associated with its simplicity.This method is relatively easy to adapt to an existing diesel engine. However, previousexperience on using this method in four-stroke diesel engines showed that it generally hasserious drawbacks during part load operation. Unless a high proportion of diesel pilot isused emissions are unsatisfactory and thermal efficiency is low due to incompletecombustion. In addition knock may be encountered as a consequence of premixing and highcompression ratio. For these reasons this method is applied mainly to four-stroke stationarydiesel engines which operate at relatively constant load and speed.PREMIXEDFUEL/AIRAIRGASMIXER<A) NATURAL FUMIGATIONGAS INJECTION^I DIESEL FUELAIR(B) TIMED PORT INJECTIONDIESEL FUELNATURAL GASHIGH—PRESSURENATURAL GASDIESEL FUELAIR (C) DIRECT INJECTIONFigure 1-2 : Methods of Using Natural Gas in Diesel Engines.56The second approach is known as timed-port injection. This method is a step aheadof the natural fumigation for this method would allow the used of natural gas in two-strokediesels as well as four-stroke. Figure 1.2b shows schematically how this method works. Inthis second approach, medium-pressure (perhaps 20 atm) natural gas is directly injected intothe inlet manifold close to the inlet valve. With precise timing and little mixing in thecylinder, a stratification of the gas in the cylinder could overcome some of thedisadvantages encountered at part-load operation in the natural fumigation method.However, work reported to date on timed-port injection shows that knock and incompletecombustion at low loads are still a concern.The last method of using natural gas in diesel engines is direct injection ofhigh-pressure gas into the combustion chamber as shown in Fig. 1.2c. High-pressure gasis injected directly into the combustion chamber near the end of the compression stroke.Thus a full stratification of the fuel-air mixture can be obtained with good flammabilityover the entire load range and with only small quantities of pilot diesel fuel needed.Successful operation has been demonstrated in applying this method in a medium speeddiesel railway engine. In this case separate diesel and natural gas injectors were used withas low as 2% diesel pilot injection quantity. Previous work on marine diesel engines hasshown that natural gas directly injected in the cylinder coupled with diesel pilot ignitioncan provide high efficiency and low emissions.It appears that direct injection of high-pressure natural gas with diesel pilot ignitionis one of the most promising ways of meeting the 1993/4 EPA emissions standard. The7disadvantage of this option is the need for pressurization of the injected gas. A researchgroup at UBC is addressing this problem by development of an intensifier/injector concept.1.3. Objectives of This Research.This research includes both experimental and analytical work on a diesel-pilot gasinjection engine, and has the following objectives :(a) To measure the engine performance over a wide range of load. The dimensionlessparameter used to measure the engine performance in this work is thermalefficiency. It is defined as the ratio of the work produced per cycle to the amountof fuel energy supplied per cycle that can be released in the combustion process.(b) To determine the rate of combustion by analysis of cylinder pressure development.(c) To investigate the dependence of emissions on operating conditions.(d) To provide a method for investigation of the effect of injector design on engineperformance and emissions.1.4. Methodology.A two-stroke Detroit Diesel of series-71 engine was used. This type of engine waschosen because of it is the engine type most widely used for transit buses in NorthAmerica.The test engine has been instrumented and converted to an electronically controlledtest engine, and equipped with a computerized data acquisition system. Performance andemissions data acquired are directly processed by the data acquisition software used, and8can be recorded on a floppy disk. Pressure development data can also be taken, viewed,and recorded for further analysis.Cylinder pressure data taken at specified engine operating conditions is thenanalyzed to revealed its combustion rate using a fuel mass-burned fraction program that ismodified for this specific engine from an existing program.2. LITERATURE REVIEW.This chapter reviews some terminology and the available literature on natural gasfueling of diesel engines and on methods of estimating the mass-burned fraction. Sec. 2.1reviews briefly the history of the natural gas fuelling of diesel engines and clarifies someof the terminology. Sec. 2.2, Sec. 2.3, and Sec. 2.4, briefly discuss the performance and thecombustion characteristics of the diesel engines as affected by the three different methodsof natural gas fuelling. Previous work on each method will be reviewed in the samesection. The fifth section reviews methods of estimating the mass-burned fraction. Then,a summary will be presented before moving to the next chapter which will describe theexperimental apparatus.2.1. The Diesel Engine.The diesel engine is the internal combustion engine which was invented byDr. Rudolph Diesel in 1892. His invention appeared only a few years after the firstprototype of the Otto spark-ignition engine ran in 1876. In the spark-ignition engine, airis throttled so that a nearly stoichiometric fuel-air mixture is available to feed the enginewhile the combustion is initiated by an electric spark. In the diesel engine, only unthrottledair is inducted and compressed. The fuel is injected into the hot air near the end of the910compression stroke. Combustion is initiated by self-ignition of the injected fuel. Since onlyair is inducted to the cylinder, the compression ratio in the diesel engine can be muchhigher than that for the spark-ignition engine. For that reason, and because it operateswithout throttling losses, the thermal efficiency of the diesel engine is considerably higherthan that of the spark-ignition engine.Conventionally, diesel engines are fuelled by liquid fuels derived from petroleumand known as diesel oils or diesel fuels. The diesel combustion process is considered tostart after the beginning of fuel injection. As diesel fuel is injected, it atomizes andpenetrates the hot air in the combustion chamber. The fuel vaporizes and mixes with theair. After a delay period, which is termed the ignition delay, spontaneous ignition ofportions of already mixed fuel and air occurs because the air temperature are above the fuelignition point. After ignition occurs, the flame propagates through the region in which fuelvapour and air are sufficiently mixed. Compression of the unburned portion of the charge(due to burning and piston movement) shortens the delay before ignition for the fuel andair which has mixed to within combustible limits and then burned rapidly. It also reducesthe evaporation time of the remaining liquid fuel. Injection continues until the desiredamount of fuel has entered the cylinder. Atomization, vaporization, fuel-air mixing, andcombustion continue until all the fuel passes through each process. Mixing of the airremaining in the cylinder with burning and already burned gases continues throughout thecombustion and expansion processes.Gaseous fuel has been used with supplementary ignition means in diesel engines formore than 60 years. The necessity of other ignition means for a natural-gas-fuelled diesel11engine was experienced by the C.& G. Cooper Company in 1927, as reported by Boyer andCrooks [11 1 . This is because natural gas by itself will not self-ignite in the combustionchamber of diesel engine at normal compression ratios. As a test, a small portion of dieselwas also injected to self-ignite and subsequently ignite the natural gas which was injectedat high pressure (1000 psi or about 7 MPa) at the end of the compression stroke. The smallportion of diesel injected here is termed as pilot diesel. It was a successful attempt, andthis event can be considered as the birth of the 'gas-diesel' engines. For some years thismethod was not popular because it was considered that the high pressure gas injectionequipment needed were costly, and rather difficult to maintain.A diesel engine is defined as a 'dual-fuel' diesel engine if it is designed to operateon either diesel oil or natural gas, or both at the same time. Its mode of operation isdefined as dual-fuel if the two fuels are used at the same time, and straight diesel if onlydiesel oil is used. Using this definition, a gas-diesel engine is then similar to a dual-fuelengine. It is important to note here that a dual-fuel diesel engine requires a different meansof introducing the gaseous fuel into the combustion chamber in order to provide fuelflexibility in using either gaseous or liquid fuel. In contrast, in this experimental work, boththe gaseous fuel and the diesel pilot are injected simultaneously through the same injector.This kind of diesel engine can not operate solely with gaseous fuel so it can not strictly beclassified as a dual-fuel diesel engine. Thus, it is necessary to define another term. A'diesel-pilot gas injection engine' is defined as a diesel engine which is fuelled withNumber in the bracket is the number of the reference as listed in references.12gaseous fuel and uses a small amount of diesel fuel to initiate combustion. A single injectoris used to inject gaseous fuel and diesel-pilot into the combustion chamber.Combustion occurs in the conventional diesel engine combustion within small zoneswhere the fuel-air ratio is suitable for combustion. The use of natural gas as the main fuelin the dual-fuel diesel engine consequently affects the combustion process. The combustioncharacteristics in dual-fuel operation differ from those of diesel operation, and depend onwhich method of using natural gas is applied. As briefly discussed in Sec. 1.2, we candistinguish three different methods of using natural gas in diesel engines. The threefollowing sections discuss the combustion characteristics as well as review previous workon each method.2.2. Natural Fumigation Method.As shown in Fig. 1.2a, in the natural fumigation method the engine ingests a fullypre-mixed air/fuel mixture. Obviously the combustion occurs in a nearly homogeneousfuel-air mixture. This mixture is first compressed by the piston movement which increasesthe mixture temperature and pressure but ideally not high enough to cause auto-ignition.Near the end of this compression, a small amount of diesel fuel is injected into the hothomogeneous gas-air mixture. The injected pilot diesel subsequently goes through theignition delay before it disintegrates into diesel vapour and ignites to initiate flame frontswhich propagate through the gas-air mixture. Thus, the propagation of flame fronts islargely responsible for subsequent combustion of the remaining gas-air mixture. In part-load13operation, there will be a failure in flame propagation due to mixture weakness. In such acase misfiring will occur or some of the fuel will survive the combustion process.Natural fumigation as a concept of using natural gas in diesel engines has beeninvestigated by a number of researchers over the decades. Much work on this concept hasbeen done. A review of problems associated with the application of natural fumigation indual-fuel diesel engines was made by G.A. Karim in 1983 [2]. He recognized that theperformance of a dual-fuel diesel engine of this type at light load was inferior to that ofthe conventional diesel due to poor combustion of lean mixtures. He concluded that usinglarger pilot quantity, preheating the charge, partly restricting the air flow, recirculatingexhaust gas, or using a combination of these measures, could partially improve the lightload performance of this dual-fuel engine.The other problem encountered with natural fumigation engines discussed wasknock, which is observed when very high outputs or very high intake temperatures andpressures are involved in the engine operation. The knock phenomenon encountered hereis of an autoignition nature. The onset of knock can be delayed by employing a lowercompression ratio and slightly later fuel injection. However, the engine will lose some ofits thermal efficiency if the compression ratio is lowered. The two problems discussed inthe G.A. Karim [2] review paper are typical problems encountered by a natural fumigationtype dual-fuel diesel engine due to the compression and the combustion of a fullypre-mixed fuel/air mixture.Work using the natural fumigation method on a Caterpillar 3304 engine [3] [4], in1985 and 1986 respectively, showed that the method has serious drawbacks at part-load.14The test engine was a four-stroke turbo-charged diesel engine which has a 93 kW ratedoutput at 2000 rpm. The engine is mechanically controlled. It was found that the naturalfumigation method was inherently unsuitable for part-load operation because of lowflammability2 , low efficiency, and excessive emissions compared to straight dieseloperation due to poor combustion of the lean gas/air mixture. Throttling of the inlet airwith existing turbocharged engines to improve flammability introduces the danger ofcompressor surge. In any case, throttling means pumping loss which will lower the thermalefficiency of the diesel engine.In other work on a naturally fumigated dual-fuel engine, a microprocessor was usedto control the liquid pilot/ gas ratio as a function of load and speed [5] [6]. The test engineswere a normally aspirated Caterpillar 3208 and a turbocharged John Deere 6466T dieselengine. Both are four-stroke engines. The amount of diesel and natural gas was varied,depending on engine speed and load in order to optimize the dual-fuel engine performanceThe results show that a significantly lower efficiency was encountered, while no emissionsdata were reported.2 Flammability limits are defined [20] as the upper and lower limits of volume percentagecomposition of mixtures of fuel and air, within which flame propagation takes place whenthe mixture is ignited.152.3. Timed Port Injection Method.As shown in Fig. l .2b, with timed port injection relatively low-pressure natural gasis injected into the inlet manifold close to the inlet valve. It might be expected that someof the problems associated with compressing a fully pre-mixed fuel/air mixture which areencountered in the natural fumigation method can be eliminated or reduced. Precise timingmight allow for stratification of the gas in the cylinder. Gas stratification, together withlittle in-cylinder mixing, would improve the mixture flammability under part-loadconditions. If the injected gas mixes rapidly with the air, the engine will compress a nearlypremixed fuel-air mixture and its combustion characteristics would not differ much fromthose of natural fumigation. In this case, it would be expected its combustion is alsostrongly governed by the propagation of flame fronts.Following work on this concept with the Mercedes OM-352 naturally aspirateddiesel engine, it has been reported [7] that above 50 percent speed and load, the engine canbe operated on gas fuel with an unthrottled air supply. This result suggests that timed-portinjection does not fully overcome the disadvantages of natural fumigation. Thus, asdiscussed before, compression ratio reduction, throttling, or increased diesel proportion ora combination of theses measures may be necessary to provide adequate part loadflammability of the gas mixture and prevent knock. The first two of these methods reducethermal efficiency, with consequent penalty in fuel economy, and satisfactory emissionslevels have not yet been reported.162.4. Direct Injection Method.As illustrated in Fig. 1.2c, the high-pressure gas is injected into the combustionchamber near the end of compression. This method of using natural gas in diesel enginesenables a full stratification of the fuel-air mixture with good flammability over the entireload range and with only small quantities of pilot diesel fuel needed.In this method, air alone is compressed by piston movement. The combustion, incontrast to the two previous methods, is governed by the mixing process which is typicalof straight diesel operation. The combustion process in the direct injection method of usingnatural gas is largely due to auto-ignition of diesel fuel; whereas that of the first twomethod operation depends heavily on both the auto-ignition characteristics of pilot dieseland the propagation of flame fronts.Of three previous investigations of this method which will be briefly reviewed inthe following one was done on a locomotive diesel engine, and two on marine dieselengines, using separate injectors for gas and diesel oil.Early work in 1983 on large bore diesel engines [8] [9] has shown thathigh-pressure natural gas directly injected in the cylinder coupled with diesel pilot ignitioncan provide high efficiency and low emissions. Miyake et al [8] confirmed thatperformance with gas fuel almost equal to that of the oil burning diesel engine wasobtained by adopting a combustion system like that of the usual diesel engine. The testengine used in their work was a single-cylinder marine diesel engine which had a bore of320 min and a stroke of 450 mm. This engine had a rated power of 520 kW at 500 rpmrated speed. The natural gas used in the experiment contains 99% methane. Although the17authors obtained their results with a four-stroke diesel engine, they argued that similarresults should also be obtainable with two-stroke diesel engines.Einang, et al [9] showed that the high pressure gas injection system is a viableconcept and suited to dual-fuel operation with gas as the main source of energy. Theirexperiments were done on a single-cylinder, two-stroke marine diesel engine which has 300mm bore, and 450 mm stroke. Rated output of the test engine was 375 kW at 375 rpm. Thegas used is methane which is boil-off gas from LNG and injected into the cylinder froma gas injector. An electro-hydraulic system, which is electronically controlled, actuated thegas valve.Wakenell et al in 1987 reported [10] that a successful operation had been achievedusing a separate diesel and natural gas injectors in a locomotive research engine. This isan Electro-Motive Division (EMD) 567B two-stroke, two-cylinder medium speed dieselengine, which has a bore of 8.5 inches and a stroke of 10.0 inches, with a compressionratio of 16:1. A liquified natural gas (LNG) fuel was vaporized before being injecteddirectly at 6000 psi (or about 40MPa) to the combustion chamber near top dead centre onthe compression stroke. It was found that rated speed and power were obtained with as lowas 2% or 3% diesel pilot injection quantity without reducing compression ratio; althoughwith slightly lower thermal efficiency.23. Analysis of Combustion Rate.Combustion rate or fuel mass burning rate analysis is used to estimate from pressure18data the fuel mass-burned fraction during the combustion. The pressure data are thecylinder pressures (recorded at each crank angle degree) over the compression andexpansion strokes of the engine operating cycle. Several means for estimating mass-burnedfraction from pressure data have been proposed by researchers.Most of the methods are based on the result that for constant-volume combustionthe mass-burned fraction x can be shown to be(2.1)P - P1where P is the measured pressure, P 1 is the initial pressure, and P^is the maximumpressure achieved in the combustion.In the engine the volume changes but the pressure P., which would have beendeveloped if the volume had stayed constant (without combustion) can be estimated fromPte,,. = p V^(2.2)in which n is the polytropic exponent. Thus to a first approximation,^P V" - P,V,"^ (2.3)P. Vmez)" - P1 V1 "where^is the maximum-measured pressure.The last equation was proposed by McCuiston et. al. in 1977 [11]. This method isvery similar to that of Marvin [12] who in 1927 proposed a graphical method equivalentto Eq. (2.3) except that subscript 1 denotes the TDC volume in the above equations.Illustrating Marvin's procedure, Fig. 2.1 shows states s and e at the beginning and the endx -19of the combustion process, respectively, on a logarithmic PV diagram. An intermediatepressure Pb at state b is corrected to the equivalent pressure Pc corresponding to TDCvolume. Pressure correction is made to this volume by drawing a line parallel to thecompression polytrope, i.e., points a, c, and d at TDC volume are the correction pressuresfor s, b, and e respectively.ln(P)exhaustsopensintakecloses TDCr=> In(V)Figure 2.1 : Graphical Method of Estimated Mass-burned Fraction.[28]Thus, according to Marvin, the mass burned fraction at crank position b could beapproximated asPc Pa (2.4)autasuredpressureei-i^ei^soi=> Crank AngleFigure 2.2 : Determination of Combustion Pressure from Measured Pressure [16].Another method which is based on constant-volume combustion and requires onlycylinder pressure data was proposed by Rassweiler and Withrow in 1938 [13]. Theirapproach utilizes pressure corrected to constant-volume for each small crank angle step.Fig. 2.2 , shows how for a step from 0 ;_, to 0 ; the measured pressure Pei differs from thepressure Pco, 1 which would have been obtained with compression only. The equivalentcombustion pressure rise AP, for this step is P oi - Pa.., . Rassweiler and Withrow showedthat for a number of such steps the mass-burned fraction can be calculated from20PcylriPeaPeon.Pai-1xeVeE Pei Pcorde=e,e„ VE Pei - Pcon) V0=0, 1where V1 is the initial volume, and again subscripts s and e denote beginning and end ofcombustion.Shayler et al [14] in 1990 investigated the Rassweiler and Withrow method andaddressed the problem of uncertainties in the value of polytropic exponent. They proposeda method of determining the correct polytropic exponent from the smoothened pressuredata.Amman in 1985 [15] has reviewed these three methods and concluded that they aresimilar in approach, although the procedures are different.Since these three different methods need only measured pressures and correspondingcylinder volumes, they offer a relatively handy analysis. However, these approaches arebased on approximations which could be avoided by using the energy equation, andavoiding the polytropic assumption (to correct for piston motion).Krieger and Borman in 1966 [16] presented in detail a computational method whichattempted accurately to represent the thermodynamic properties of the cylinder contents.It computes the mass of fuel burned during each crankangle increment by solving theequations of energy and mass continuity together with the equations of state, internalenergy. The disadvantage of this method, in addition to the approximation that the burnedand unburned gases are handled as a mixture, is that it requires a correlation for gas-side21(2.5)22heat-transfer coefficient and an estimate of metal temperature for the exposed cylindersurface in order to account the effect of heat transfer.The amount of heat transfer from the cylinder contents to the wall Qw is given byQw = A h, ( T - T. ) (2.6)where he is the convective heat transfer coefficient, T is the temperature of gases in contactwith the cylinder wall, T. is the wall temperature, and A is the cylinder wall area.One problem in estimating the amount of heat transfer from the cylinder contentsto the wall is how to determine the temperature of the gases in contact with the cylinderwall. The methods available assume that all gases in contact with the wall have a meantemperature T. which is the mass-average or bulk mean gas temperature obtained from anequation of state with given pressure [17]. In the following, two of the widely usedcorrelations are briefly reviewed.Annand [18], after reviewing the existing formulae for internal combustion engineinstantaneous heat transfer rate in 1963, concluded that the experimental convective heattransfer data can be represented by a Nusselt-Reynolds Number relation in the followingform.Nu = a (Re)"^ (2.7)where a is a constant having a value in the range between 0.35 and 0.8.In the above correlation the characteristic length is taken to be the cylinder bore B, and theNusselt Number and Reynolds Number are expressed as follows.Nu=kB/kRe = p Sp B /23where he is the convective heat transfer coefficient, k is the thermal conductivity of themixture, p is density of the mixture, S p is the mean piston velocity, and IA is the viscosityof the mixture.In 1967, Woschni [19] proposed an equation for determining the heat transfercoefficient for internal combustion engines. His correlation is identical with thedimensionless equationN. = 0.035 (Rd" (2.8)These two methods of determining the convective heat-transfer coefficient are ofsimilar form , differing primarily in the constant and the exponent of the Reynolds Number.Both models use bulk mean gas temperature as an approximation of the gas temperaturein contact with the wall. The uncertainty associated with this assumption constitutesconsiderable physical uncertainty (perhaps as much as 20%) in estimating the instantaneousconvection heat transfer in the internal combustion engine. After reviewing the availableheat transfer model available in 1987, Borman and Nishiwaki [16] concluded that "asatisfactory predictive heat transfer model for engines has not yet evolved".2.6. Summary.It has been shown experimentally that the part-load performance of dual-fuel withnatural fumigation diesel engines is inferior to that of the conventional diesel due to poorcombustion of lean mixtures. Using microprocessor control of diesel pilot fuel it is possibleto avoid knock and maintain part-load fuel economy, but at the expense of large diesel fuelconsumption at low load. The second method, timed port injection, was expected would24overcome the disadvantages encountered in the natural fumigation method. However, workto date on a coach diesel engine does not indicate that timed port injection fully overcomethe disadvantages of the first method. It is unlikely that adequate part-load efficiency willbe obtained with natural fumigation or timed-port injection method, or that these twomethods will offer the emissions reduction of direct injection of natural gas into thecylinder. Work on large marine diesel engines and railway engine suggests that directinjection method can provide high efficiency and low emissions.Work on the direct injection method indicates that the traditional diesel mode ofcombustion (stratified fuel charge) could be employed with natural gas, so that theefficiency advantage of the diesel engine can be preserved. In previous works differentinjectors were used for natural gas and diesel fuel. However, no literature found onemploying a two-fuel unit-injector which simultaneously injects natural gas and diesel fuelin medium-size bus and truck diesel engines.As a tool, analysis of combustion rate from the calculated mass-burned fractiongives a pattern of how combustion occured. There are several methods for estimating mass-burned fraction which could be classified into two categories, i.e., methods using onlypressure crank-angle data, and methods employing the mass and energy equations andcylinder pressure information.3. EXPERIMENTAL APPARATUS3.1. Introduction.The purpose of this chapter is to describe the test engine, the apparatus, and the dataacquisition system.The test engine was fuelled with compressed natural gas and diesel fuel used as anignition source (diesel pilot). It was coupled with a water-brake dynamometer, and fullyinstrumented.Principal measurements are torque, engine speed, fuel mass-flow, cylinder pressure,and emissions. The first three operating variables were needed to calculate brake thermalefficiency which is used to describe engine performance. Cylinder pressure data were takenin order to determined combustion rate. Ambient conditions were measured to determinethe correction [201 applied to the calculated power. The composition of NO R, CO, HC, CO2,and 02 of the exhaust gas were measured to determine dependence of emissions onoperating condition.2526Figure 3-1 : Schematic of Experimental Apparatus and Instrumentation.The arrangement of the apparatus and instruments is shown schematically in Figure3.1. The test engine used was a single-cylinder, two-stroke, diesel engine. A water-brakedynamometer was connected to the engine to regulate load. The fuel system consists of twosub-systems, i.e., the compressed natural gas sub-system and the diesel pilot sub-system.27Each has a mass-flow measuring instrument. A laminar flow element was installed in theintake line to measure intake air flow. Ambient pressure and temperature sensors werepositioned close to intake air filter. Exhaust gas was analyzed by flowing it through aconsole. This console has a gas sampling system which is capable of measuring fiveexhaust gas compositions simultaneously. A data acquisition system which consists of amultiplex box, a data acquisition card, an ISAAC% and an IBM2 compatible standardpersonal computer (IBM-PC) was provided. The engine test bed together with almost allthe instrumentation is located in the engine test cell. Engine operation and data acquisitionwere controlled from a control room next to the test cell.Data were classified into two categories: performance data and pressure data. Thefirst one is related to performance and emissions; these data are time-averaged steady statedata taken at particular operating conditions. The pressure data are cylinder pressuremeasurements taken at relatively high frequency, e.g., every crank angle degree. Both aretaken at steady engine operating condition, and both are handled by the data acquisitionsystem. An Ethernet system was added to the IBM-PC to allow a transfer of data from thedata acquisition system to a main frame computer. The data flow will be elaborated later.All instruments installed are electronic except the ones used to read the barometricpressure, relative humidity and dry bulb temperature. They were selected to fit the dataacquisition system adopted. Calibrations of some principal instruments are documented inThe trade mark of a high speed data acquisition computer.2 IBM = International Business Machines. A company name.28the Appendix B.The following section gives details on the test engine used and its test bed. Fuelsystem, which is considered one of the most important engine operation system, isdiscussed afterwards. The testing and calculation procedure will be covered in the nextchapter.3.2. Engine and Test Bed.As stated in Sec. 1.4, a Detroit Diesel Series 71 single-cylinder diesel engine wasused in this experimental work. The engine is a two-stroke naturally aspirated, directinjection diesel engine which has specification [36] as shown in the following table.Bore x Stroke^ 4.25 in.(108 mm) x 5.0 in.(127 mm)Displacement 70.93 cu in. (1.162 liters)Rated Output^ 15 HP (11.2 kW) @ 1200 RPMRated BMEP 70 psi (4.8 bar)Compression Ratio^16 : 1Fuel Injection Direct, Unit InjectorScavenging Type^UniflowNo. of Intake Ports 18No. Exhaust Valve^2 (two)Table 3.1 : Engine Specifications.29The engine had been used to power a 10 kW alternator at 1200 RPM withmechanical governor control of a unit-injector. This fuel injection system has beenconverted to electronic control to gain accurate control and variability of the injectiontiming and the amount of fuel injected to the combustion chamber.Figure 3.2 shows the test bed installation in a UBC engine test cell. The test engine(1) was mounted on a movable test bed. It was coupled with a Go Power Model DA316Dwater-brake dynamometer (2). A cooling tower (3) has been installed to replace the radiatorso that the cooling fan could be removed and so that closer control of engine temperaturecould be achieved.Figure 3-2 : The Test Engine and Dynamometer in a Cell.(1) Test Engine, (2) Water-brake Dynamometer,(3) Cooling Tower, (4) Dynamometer-controlled Valve30Figure 3.3 : The Engine Control Panel and Data Acquistion System.(1) Dynamometer Control Module (DCM),(2) Data Display Module (DDM).The engine control panel and data acquisition system are in the control room. Figure3.3 shows a photograph of these equipment. A Dynamometer Control Module (DCM) anda Data Display Module (DDM) from Environmental Dynamics Inc. are used. These twomodules are labelled as (1) and (2) respectively, and are parts of the engine control panelalso shown in Figure 3.3. The load and the speed of the engine are controlled through thevariable resistor on the DCM front panel. The DDM displays engine speed, dynamometerload, fuel rate, and one out of eighteen temperature options. The DCM also has a safety31shutdown feature. If one of the engine operating indicators, e.g., engine cooling watertemperature, oil temperature, and oil pressure, is higher than a preset level, the DCM willshut the engine down by closing the fuel supply valves.Resistance of the brake is controlled by adjusting the water flow rate passingthrough the dynamometer. An electro-pneumatically operated water inlet valve, (4) onFigure 3.2, linked to a load setting dial in the control panel is used. The capacity of thedynamometer is 130 lb.ft (176.2 N.m) at 1500 rpm or 57 lb.ft (77.3 N.m) at 1000 rpm.Engine torque is measured by load cell attached to the dynamometer case. A magneticpickup type speed sensor develops a frequency signal (1 Hz = 1 RPM) which is thenconverted to an analog voltage proportional to speed. Both load (torque) and speed signalsare sent to data acquisition system and are displayed on the engine control panel.3.3. Fuel System.As shown in Figure 3.1, the engine was provided with gaseous as well as liquid fuelfor its operation as a dual-fuel diesel engine. Unlike conventional dual-fuel diesel engines,compressed natural gas is injected directly into the combustion chamber. Natural gas doesnot ignite easily. Therefore, a small portion of diesel fuel is injected from the same injectorto self-ignite and then ignite the natural gas. The following two sub-headings will elaborateon the fuel supply and fuel injection systems.COMPRESSOR DIESELFUELCOMPRESSEDNATURAL GAS MASS-FLOWMETERSHUT-OFF VALVEI SHUT-OFFVALVE PUMPFILTERBLD. GAS LINEPRESSUREREGULATOR.ELECTRONIC.ILCONTROL1323.3.1. Fuel Supply System.Figure 3.4 shows the fuel supply system which consists of two sub systems, i.e.,natural gas, and diesel fuel supply sub systems. The properties of both fuels and a typicalchemical composition of the natural gas used are found in Appendix C, and Appendix Drespectively.DUAL FUELUNITINJECTORVFigure 3.4: Schematic of Fuel Supply System.The diesel fuels used are a commercial Grade-2 Diesel Fuel Oil and a high-cetanenumber diesel fuel. The fuel is pumped from the fuel tank through a filter to a gravimetricmass-flow measuring instrument (AVL) before it is injected to the combustion chamber.33The return line from the injector is then directed to the AVL, so that the net consumptionof diesel fuel can be determined.Compressed natural gas is drawn from commercial pressure bottles. Natural gasfrom the main supply at a pressure of 14 kPA is compressed to 19.0 MPa by acommercially available Residential Refueling Appliance (RRA). One bottle is used at atime by manually setting the appropriate valves.The compressor in the appliance is a four cylinder, 4-stage, air-cooled, oil-free unit.It is driven by a 1.5 HP electric motor and is equipped with a control system consisting ofpressure and temperature sensors, relief valves and electrically operated valves. Thecompressor stops automatically when the pressure reaches the set level. Storage bottles areconnected in parallel in a system which allows one cylinder to be used while the othercylinders are being filled. The supply gas pressure was manually regulated through apressure regulator.A solenoid shut-off valve was installed in each of fuel supply line to stop the fuelflow to the unit injector in case of emergency. They are controlled by the engine'selectronic control system which is an integral part of the fuel injection system.3.3.2. Fuel Injection System.In the mechanically controlled injector of conventional diesel engines, injectiontiming and injection rate are mechanically controlled by ports and helices machined in thebushing and plunger assembly. An electronic unit-injection engine governing system knownas Detroit Diesel Electronic Control (DDEC) was developed by Detroit Diesel [21] to34enhance flexibility and precision. This electronic system was used to control the engine inthe baseline test series.COMPRESSEDDIESEL FUEL^ NATURAL GAS■DEG. OF INJECTION^CRANK*PULSE WIDTH^POSITION*DSL THROTTLE POS.^SIGNALI INPUT SIGNAL COND, UNIT II MICRO-CONTROLLER BOARD II OUTPUT DRIVE CIRCUIT I^I ELECTR. DISTRIBUTOR LNIT I-^ I HYDR. PRESS ^ SELEMED I STEPPERMOTOR IINJECTORFigure 3.5 : Schematic of Fuel Injection System.A microprocessor controlled unit-injector for gas-fueling diesel engines has beendeveloped at UBC Combustion and Alternative Fuel Laboratory which is part of an"Intensifier-Injector Technology" [22]. Figure 3.5 shows a schematic of this gas injectorin the fuel injection system. The injector injected compressed natural gas and diesel pilotinto the combustion chamber near top dead center. It utilizes a cam-driven plunger toprovide high pressure diesel fuel to actuate the poppet valve which permits flow of the gasand entrained diesel fuel to enter combustion chamber near top dead center. The diesel pilot35fuel is gas-blast atomized by the natural gas flow past the diesel ports. A modification ofthe DDEC is adapted in developing an electronic control system for this two fuel unit-injector.The beginning of injection (BOI), and the duration of injection ( or pulse width,PW ) settings are fed to the electronic control system. They were manually set throughseparate variable resistors mounted on the engine control panel. These input signals,together with crank position signal, were used to actuate a solenoid which will close avalve and consequently develops a hydraulic pressure (Fig. 3.5). The developed hydraulicpressure will open the poppet valve. High-pressure gas and gas-blast atomized diesel fuelenters the combustion chamber with high velocity in a form of conical sheet.A needle valve, which is coupled with a stepper motor, is used to throttle the flowto meter the amount of diesel fuel injected. Needle position was electronically controlledby a variable resistor. The amount of pilot diesel can be manually controlled through avariable resistor mounted on the engine control panel in this way.3.4. Instrumentation.Four principal instruments used to measure engine performance, i.e., engine speed,torque, gas mass-flow, and diesel mass-flow, are described together with emissionsinstrumentation. Cylinder pressure instrumentation used is also discussed. Calibrationprocedure of most these principal measurements is presented. Other instruments wereinstalled to support these measurements, or to monitor the engine working condition.363.4.1.Engine Speed.A speed sensor attached to the shaft of the dynamometer is used to measure enginespeed. This magnetic speed sensor provides a signal frequency proportional to enginespeed. The frequency signal is then converted into an analog signal which is displayed onthe engine control panel and sent to a data acquisition system.A hand digital tachometer (Shimpo, model DT-205) was used in the calibration.This hand tachometer sends out a continuous light beam and counts the pulses reflectedfrom a piece of reflective tape attached to the engine shaft. Calibration showed that thestandard error of the measurement is 0.1% (Appendix B.1).3.4.2. Torque (Load Cell).A strain-gauge load cell mounted on the dynamometer casing is used to measurethe load applied to the engine. The strain-gauge signals are processed and displayed on theengine control display. This low level signal is also amplified in a circuit and then sent tothe data acquisition system.The load cell was calibrated by loading and unloading the load cell with standardweights placed on an arm bolted to dynamometer casing. A standard deviation of 0.1% wasfound in the calibration (Appendix B.2).3.4.3. Gas Fuel Mass-flow Rate.Compressed natural gas mass-flow is measured with a Micro Motion Model D12mass-flow meter. This instrument works on the Coriolis acceleration principle. The gas37flows through a U-shaped tube which vibrates at a frequency which is directly proportionalto the product of the fluid density and velocity. The frequency of the signal produced wasconverted to a 4-20 mA current signal and sent to the data acquisition system by a remoteflow transmitter. The manufacturer's calibration for this meter was used, which showed thatthe averaged measurement error is ±0.4% .3.4.4. Pilot Diesel Fuel Consumption.An AVL Dynamic Fuel Consumption meter series 730 is used to measure dieselmass-flow rate. This determines flow rate by sensing the weight of fluid filling a vessel.Both the supply line and the return line of pilot diesel are connected to this vessel. Thediesel fuel consumption is determined through the effective change in weight of the vessel.The system permits measurement of the total consumption over a selected measuring periodas well as the instantaneous fuel consumption.A five second measuring period is used for the measurements. The analog signalis sent to the data acquisition system, and its digital value is displayed on the evaluationmodule placed on the engine control panel.Calibration is done by weighing the fuel leaving the vessel in a set time period. Itwas found that the measurement standard error is 0.2% (Appendix B.3).3.4.5. Emissions Intrumentation.Emissions instrumentation used was assembled in 1985 at UBC [23]. It can measurethe concentration of unburnt hydrocarbons, oxides of nitrogen, and carbon monoxide; it is38equipped with a gas sample handling system. Modification in 1990 [24] allowsmeasurement of carbon dioxide and oxygen. However, the oxygen analyzer did not workproperly. The four emission analyzers and gas cylinders needed for analyzer calibration andoperation are mounted in a movable cabinet. Figure 3.6 illustrate the schematic diagram ofthe emissions console. Signals are sent to the data acquisition system.ToExiausr <zi GSTACKFROM ENGINESAW LIN6 PROSEOUTLETFigure 3-6 : Exhaust Emission Sampling System.PRESSURE SIGNALTO CHARGE AMPLIFIERADAPTOR SLEEVECYLINDER HEAD/_L^/^-1-- -'--- /X/ l^I^A^1FUEL/ INJECTOR \-%.. I ...--'i^1^1 ^—...._ ■-123'1 /1 ^N. A/^I^\ X^ /1\ ‘^1^/‘^ /INTAKE VALVEPRESSURE TRANSDUCER(PCB Mod. 112A05)EXHAUST VALVE39An exhaust sampling probe was inserted in the exhaust line at about 1.2 m from theengine manifold. A drawing of the sampling probe located in the exhaust line isdocumented in Appendix E. Calibration of each analyzer was made before taking data withspan gas and zero gas. After taking the data, a calibration check were made to ensure thatthe analyzer works properly during measurements.3.4.6. Cylinder Pressure.An air-cooled PCB piezoelectric pressure transducer with maximum frequency of80 kHz was used. The frequency used when taking one pressure datum at each crank angledegree at engine speed 1200 rpm is 7.2 kHz.SECTION A-AFigure 3-7: Pressure Transducer Location in the Cylinder Head.40Figure 3-7 shows how the pressure transducer is mounted in the cylinder headthrough an adaptor sleeve. The sensing surface of the transducer is installed 3 0 mm abovethe fire deck. The adaptor passage has the same diameter as that of the sensing surface (5.6mm) This mounting technique is applied to minimized the effect of thermal shock [25][26] [27]. The signal was amplified in a Model 5004 Kistler Charge Amplifier andtransmitted to the data acquisition system.The pressure transducer was statically calibrated using a dead-weight tester for apressure range of 0 to 1900 psi. Standard error is 0.6% (Appendix B.4).3.5. Data Acquisition System.The two engine test cells, with one engine in each cell, are equipped with an IBM-PC based data acquisition system [28]. This system is controlled by a program whichallows the user to choose the engine, to configure the system to his needs and to calibrateinput channels. The software does calculations, allows data to be saved to a disk, andconverts cylinder pressure data from binary to ASCII 3 .Hardware includes a multiplex box, a PC which has a PCL-818/Data AcquisitionBoard, an IEEE-488/General Purpose Interface Board, and an ISAAC which is a computerto acquire pressure data. The multiplex box consists of a multiplexer board, DT-709/screwterminal board, "Vector" trigger board, and a digital interface board. It does all the3 ASCII = American Standard Code for Information Interchange, which is a set of 256 codesthat many computers use represent letters, digits, special characters, and other symbols.!MICRO NOTIONDI21H-RFT9712FREQ/VOLT --al 0CONVERTMULTIPLEX BOXTc_ Li41switching in the data acquisition system. Figure 3-8 shows a schematic diagram of dataflow.STEADY STATE DATAGAS MASS FLOWBEGINSIG OF INJECTION (Manual setting)^AMBIENT AIR TEMP.^I Electr.Sensorl^PULSE WIDTH^(Manual setting)^TIME LOAD CELL SPED Sensor ^DCNINTAKE PRESSURE^IDDEC Press.SensorDIESEL MSS FLOW AVLFuel BatAIR FLOW (DELTA P)^LanFlow lien.^Delta Press.GAS PRESSURE^[Heise 620#TURNS OF STEPPER^(Manual setting)EMISSIONS (NOKHC,OE1CIE02)^I DOSSIERS CENSILEHIGH SPEED DATACR. ANGLE & BBC INDEX^ISIFTPOT SP-360I ^CYLINDER PRESSURE^I PCB Piezo-Elect. I-4Charue Aro.HYDRAULIC PRESSURE^I PCB Piezo-Elect. ItPar. SuDilvIBM-PCDISK DRIT;IE(Data Acq. Board) PCL - 818^1^"Cr^IIEEE-488 / Ii• 2• 3^ 4I .5• 67891011-15AUX.Maz_4-.P6LILiLI^►_ - nISAAC^IFigure 3-8 : Data Flow in the Data Acquisition System.Up to 16 performance data, e.g., engine speed, torque, etc., can be directed by themultiplexer board to the DT-709/screw terminal board (Fig. 3.8). The incoming signal wasconditioned by the DT-709 with a gain of one. This board converts the 16 differentialsignals to 16 single-ended signals with a common ground. Analog voltage signals are thenconverted to a 12-bit digital number by the PCL-818 data acquisition board. Each of the16 channels data is sampled 100 times, averaged, and standard deviation calculated.42Averaged data are used to calculate engine performance characteristics, e.g., brake power,brake mean effective pressure, thermal efficiency. Any eight of the steady state data andor calculated values can be displayed simultaneously on the PC monitor screen. Data canbe saved to a specified floppy disk on command.Because it is acquired at a relatively higher frequency, the cylinder pressure dataflows in a different way compare to that of the performance data. Crank angle signals,together with bottom dead centre (BDC) signals, are needed to clock data taking. Anauxiliary multiplexer board directs these signals to a trigger board. When commanded toacquire pressure data, the PCL-818 sends a digital signal to the trigger board whichprepares a trigger for the ISAAC to take cylinder pressure data. Using the next BDC signalafter the power stroke as a trigger and crank angle as an external clock, the ISAAC willthen start taking data. Up to 100 cycles of cylinder pressure data can be taken. Uponcompletion of this process, the ISAAC sends signal through the IEEE-488 General PurposeInterface Board, and data is transferred in binary form, one cycle at a time to the specifiedfile. Converting binary pressure data to ASCII format can be done by selecting the relevantoption on the main menu.An Ethernet system is employed to connect the PC to a main frame which allowstransferring and processing pressure data for further analysis.4. EXPERIMENTAL PROCEDURE4.1. The Testing Procedure.Tests were divided into two groups. The first group was done with the dieselinjector to obtain the diesel performance data. The second group used a prototype two-fuelunit injector. The injector has an approximately 10° poppet seat angle. The schematic ofthe injector is filed in Appendix F. Note that the gaseous and diesel fuels injection havea conical-sheet formed jet when leaving the injector. The conical sheet jet was interruptedby the castellated-shrouded the poppet since flow visualization [29] had shown that thisconfiguration could enhance gas penetration in the cylinder which subsequently could affectcombustion rate. The 50% shrouding' was devised so that there were six jets instead ofa continuous conical-sheet jet.Natural gas injection pressure, diesel ratio', diesel-fuel cetane number, and certainengine operating parameters were varied, while engine speed was maintained at 1200 RPMto determine their effects on thermal efficiency, emissions and combustion characteristics.The variable parameters ranges are given in Table 4.1.I Approximately 50% of the initial jet-flow area blocked by the shroud.2 Diesel ratio is defined as the energy percentage of diesel pilot to total fuel energysupplied to the engine.4344Injection Gas Pressure^50 ... 70 barDiesel Ratio^ 15 ... 25Diesel-fuel Cetane Number3^—45, or 62.2Load (BMEP)^ 0.5 ... 4.5 barBeginning of Injection (B01)^24 ... 40°BTDCTable 4.1 : Testing Ranges.Brake mean effective pressure (BMEP), which is defined in Section 4.2.2, is usedto represent the load applied to the engine. Engine speed is kept constant by manuallyadjusting the fuel injection duration. Injection gas pressure is adjusted with a natural gassupply pressure regulator as shown in Fig. 3.4. Diesel ratio is adjusted by metering theamount of diesel fuel injected. The metering valve position is controlled by a stepped motoras described in Sec. 3.3. The beginning of injection (B01) indicates the crank positionwhere the injector solenoid is energized which is followed by pressurization of the dieselfuel and simultaneous injection of both liquid and gaseous fuels.Prior to the tests of the gas injector, a series of tests were completed with the dieselinjector. This first group of tests was done to acquire the diesel performance data against3 Diesel-fuel properties are listed in Appendix C.45which to compare the data on performance, emissions for natural gas fuelling.The procedure below was followed in obtaining the gas-diesel engine performancedata for the chosen gas injector at the specified engine speed.1. Set the gas pressure at a value, i.e., 50, 60, or 70 bar.2. Set the beginning of injection (BOI) at a value, i.e., 24, 28, 32, 36, or40°BTDC.3. Set the diesel ratio by adjusting the pilot-diesel metering valve.4. Take data for the chosen combination of gas injection pressure, diesel ratio,and BOI at different loads in the range 0.5 bar to 4.5 bar (or the maximumachievable load) at constant speed.5.^Repeat for all the other BOI.The main purpose of engine performance evaluation is to determine the dependenceof thermal efficiency on BMEP; this is termed the performance curve. For a particular gasinjection pressure and a specific diesel ratio, the performance curve for every BOI settingwas plotted. At a given load, the BOI which gave the best thermal efficiency wasdetermined by cross-plotting BMEP vs BOI at that load. Best performance for a given gasinjection pressure and diesel ratio is obtained by plotting 2 variables vs load (BMEP):maximum thermal efficiency, and best BOI. This curve termed "best BOI performancecurve" and used to compare engine performance for different gas injection pressure, diesel46ratio, and cetane number.Pressure data were acquired at intervals of one degree crank angle. They were takenat low, medium, and maximum load in each group of tests. Pressure data were analyzedto show combustion rate by estimating the mass-burned fraction of the fuel at each crankangle interval. Combustion rate analysis is discussed in Chapter 6.4.2. The Engine Performance Calculation Procedure.4.2.1. Thermal Efficiency.Thermal efficiency ri th is defined as the ratio of the work produced per cycle W,to the amount of fuel enthalpy supplied per cycle ( h f = mf LHV ) We (4.1a)th m • LHVwhere mf is the mass of fuel inducted per cycle, and LHV is the lower heating value ofthe fuel. By taking the time derivatives of the numerator and the denominator of the Eq.(4.1a), thermal efficiency can be expressed asPth - •mf • LHV(4.1b)in which P is the power.47Power was obtained from speed and torque measured with the water-brakedynamometer. Brake power Pb (kW) was calculated from engine torque Tb (N.m) and theengine speed N (rev/min or RPM), using :P = 2 It60 Tb x10 -3Then by substituting the above equation into Eq. (4.1b) (and considering there are j typesof fuel burned) the brake thermal efficiency of the engine at a defined engine operatingcondition can be calculated as follows:11 b.th2 It0 Tb x10-3(4.2)E mf; LHVi / 3600where N is the engine speed (RPM), T is the torque (N.m), riz.0 is the mass flow of theth fuel burned (kg/hr) which are j=1 for diesel and j=2 for natural gas, and LHVi is theLower Heating Value or the enthalpy of combustion of the i th fuel.The engine speed, the torque, and the mass flow rate of the diesel-pilot liquid andthe compressed natural gas are the measured quantities which are acquired during theexperiments. The enthalpy of combustion (LHV) of the diesel-pilot is 45220 kJ/kg, thenatural gas is 49098 kJ/kg (Appendix C).Since it is not practical to run the experiments at the typical standard inlet aircondition, a correction factor to the observed power was applied conforming to the SAEStandard J1349 JUN85 [20] to provide a common basis of comparison. The procedure to48compute the correction factor is filed in Appendix G. This correction factor is also appliedto the calculated thermal efficiency or the engine performance and the specific emissions.4.2.2. Brake Mean Effective Pressure (BMEP).The mean effective pressure MEP is defined as the work per cycle We per unitcylinder displacement volume Vd. This parameter is a performance measure independentof engine size and which expresses engine load.MEP =^ (4.3)VSubstituting the work per cycle with the engine speed N (rpm), and brake torque Tb (N.m),the bmep (bar) of the engine at a defined operating conditions can be calculated using:BMEP - 2 1G Tb X10^(4.4)Vdwhere the cylinder displacement volume Vd is in cm3 , and 1 bar is 100 kPa.4.2.3. Brake Specific Emissions.The measured concentrations of gaseous emissions in the dry exhaust gases areoxides of nitrogen ( NO, and NO, grouped together as NO„ ), carbon monoxide (CO),unburned hydrocarbons (HC), and carbon dioxide (CO 2). They are measured in parts permillion except for CO, which is measured in percent by volume.The brake specific emissions is used to indicate the level of emissions since it is a49more comparable indicator. Brake specific emissions b se are defined as the mass flow rateof pollutant ME per unit brake power output Pb .bse^m e (4.5)PbThe following assumptions are made in calculating the brake specific emissions ofNON, and the unburned HC.1. The dry exhaust gases consist of CO 2, CO, 02, N2, NON, and unburned HC.2. The unburned HC has the same composition as that of the compressednatural gas fed into the engine.3. The NO  is considered to be NO.4. The intake air consists of 79.0% N2 and 21% 02 by volume, and isconsidered as an ideal gas.The first assumption was made because the concentration of the other constituentsof the exhaust gas, e.g., 0, OH, H, and N, are very small compared to the concentrationsof 02 or N2 .Since the concentration of the emissions gas are measured volumetrically,conversion of these pollutant volumetric flows into mass flows are needed. Dry exhaust50mass flow rate must be calculated first.Dry exhaust gas mass flow rate thexh is calculated from the intake air mass flowrate in , which is assumed to be an ideal gas and water free, and the total fuel massflow rate E thfue by conserving mass flow.mph = may + E fuel (4.6)Compressed natural gas mass flow rate and diesel-pilot mass flow rate are individuallymeasured. Mass flow rate of intake air is calculated from its standardized measured(volume) flow rate 1)41' since its density at the adopted instrument standard condition Pt,Pis known. Their relationship is=nt,P r(4.7)Standardized intake air flow rate is computed from actual flow rate 1. %Pa byapplying the correction factors for pressure Pef and temperature Ta .= 'ta,Pa Pcf Tcf^ (4.8)A Laminar Flow Element (LFE) is used to determined the actual flow rate of the incomingair (Fig. 3.1). Output of LFE is a pressure difference which is translated into flow rate51using the manufacturer calibration curve (Appendix B.5). The base standard condition ofthe curve is 70°F(21.1°C) and 29.92"Hg(101.03 kPa). The correction factor for pressure is_  Pa cf^101.03,and for temperature [30] is294.1T,„ - [ ^ [ 1 - 0.00231 (^- 21.1 ) ]ta +273where P. , and t o are ambient pressure and temperature respectively. The first factor in thetemperature correction adjusts the volume changes due to temperature. The second factorin the temperature correction account for the dependency of viscosity on temperature whichis a linear fit applied for range of 20°C to 40°C.Intake air mass flow is calculated using Eg. 4.9 obtained by substitutingstandardized intake air flow rate in Eq. (4.5) with actual flow rate as given in Eq. (4.8).2.1^Pa= ptpi,rapa [(1 -0.002317(ta -21.1)] [ta +29473 [  101.03](4.9)The next step is converting the mole fraction of a pollutant Y i , e.g. NOR , CO, orunburned HC in the exhaust gas, to its mass flow rate using the equation,(4.10)bsC0(4.5a)(4.5b)(4.5c)ihNO.PbCOPbhv'sncPb52where M 1 is the molecular weight of the P h pollutant species.Then by inserting the mass flow rate of the pollutant calculated from Eq.(4.10) intothe Eq.(4.5), the brake specific emissions of the pollutants are :Each has the unit g/kW.h5. MEASUREMENTS OF ENGINE PERFORMANCE5.1 Conventional Diesel Performance.Conventional diesel performance (termed diesel baseline) was taken with theelectronically controlled diesel fuel injector installed. Gas-diesel test results were comparedto this baseline. As described in Sec. 4.2, the engine performance is expressed by its brakethermal efficiency at different loads and plotted as a thermal efficiency' versus BMEP(defined in Sec. 4.2); this is termed a performance curve. The corresponding engine exhaustemissions, e.g., oxides of nitrogen and unburned hydrocarbons, are presented as brakespecific emissions (defined in Sec. 4.2.3).Since the diesel-fuel injector was electronically controlled the amount of diesel fuelinjected into the combustion chamber and the injection timing, i.e., the beginning ofinjection timing (B01) and the duration of injection (PW), are automatically updated fromtime to time to keep the engine running smoothly at its best performance for a given loadand speed. The performance curve corresponding to best operating conditions for thisengine is termed the best BOI performance curve or performance curve for short.53The term thermal efficiency used for the rest of this thesis implies the brake thermalefficiency.28A 26 0.90.8 017-0.70.8 00.5ct0.4 50.3cc0.2 -Jw0.12422201816141210860BMEP ( bar ) r>Figure 5.1 : Diesel Baseline Performance CurveThe diesel engine performance curve is presented in Fig. 5.1, while Fig. 5.2 andFigure 5.3 show the corresponding emissions. As shown in Fig. 5.1, the maximumachievable load at 1200 RPM with the diesel injector was about 4.5 bar which is 6% lowerthan the engine specification. This test engine (manufactured in 1939) has a relatively lowcompression ratio (16:1) and low load capability compared with currently typical naturally-aspirated two-stroke diesel engines which have BMEP values in the range 7 to 9 bar [31].It was found that the engine maximum thermal efficiency is 25.7%.54a 4EPA 93 /55Fig. 5.1 also shows the fuel-air equivalence ratio of the fuel-air mixture in thecombustion process. The fuel-air equivalence ratio 4) is the ratio of the actual fuel-air massratio to the stoichiometric mass ratio which is the theoretical proportions of fuel and air.It is shown in Fig. 5.1 that the equivalence ratio of the diesel ranges from 0.23 to 0.69,which means the diesel engine operates with excess air of 335% to 45% from low to highload.60A 5550-C^4540co35302520015105011001000900800700600500400300200100AEca.CLX0BMEP ( bar ) L^ >Figure 5.2 : Diesel Oxides of Nitrogen Emission.120115110105100AEPA '93/ '94485 I8075705 651056Concentration of oxides of nitrogen in the exhaust gas increases with load in therange of 440 to 995 ppm as shown in Fig. 5.2. This is higher than the standard. At lowload (BMEP — 1 bar) it is about three times higher, while at high load (BMEP — 4 bar) itexceeds the standard by approximately 80%.5.55ff^4.5-C^43.5C.)^3F• 2.52▪ 1.510.50BMEP ( bar ) -^,›Figure 5.3 : Diesel Unburned Hydrocarbons Emission.As shown in Figure 5.3, the unburned hydrocarbons concentration in the dieselemissions vary with load in the range from 75 to 115 ppm. From medium (BMEP 1.5Bar) to high load the concentration meets the standard, although at low load it fallsconsiderably above the standard.4^4BMEP ( bar )575.2 Natural Gas Diesel Performance.The best BOI performance curve is used to represent the diesel baselineperformance. The gas-diesel performance at a given gas injection pressure and diesel ratiois also presented in best BOI curve. However, the injection timing in the gas-diesel engineis manually adjusted to run the engine smoothly for a given load and speed. The testingprocedure decribed in Sec. 4.1 was followed.Figure 5.4 : Determination of Best BOI Performance.( Gas injection pressure 60 bar, diesel-ratio 25% DF2, Shrouding 50% )In each test, a set of performance data includes the whole range of BOI and loadlisted in Table 4.1. The individual BOI performance curve was plotted. By cross-plotting58the maximum thermal efficiency and BOI versus BMEP, the best BOI performance curveis obtained. Fig. 5.4 shows the typical individual BOI performance curves and the BOIperformance curve that represents the engine performance at a given gas injection pressureand diesel ratio (defined in Sec. 4.1).5.3 Effect of Gas Injection Pressure.Tests to determine the effect on thermal efficiency of gas injection pressure wereperformed for the whole range of load while keeping the engine speed and the diesel ratioconstant.Fig. 5.5 shows the effect of injection pressure on thermal efficiency for the wholerange of load with diesel ratio maintained at 25%. The tests for three different gaspressures exhibit improvements of the maximum load capability when the gas pressure isincreased. This is not surprising since higher gas pressure, to a certain degree, increases thegaseous fuel mass flow with the same injection duration. With either 50 or 60 bar gasinjection pressure, thermal efficiency appears to be about the same up to medium load(BMEP — 2.5 bar); although with 70 bar injection presssure its thermal efficiency is 2-4%lower than with 50 or 60 bar. However, these results clearly indicate that high thermalefficiency (exceeding the baseline by about 2%) is obtained at high load with high injectionpressure.A 2824-2220-18-us-1412-108-4-20 4BMEP (bar) I.^ >RPM^: 1200Diesel Ratio : 25%Shrouding : 50%5930 ^28-Figure 5.5 : Effect of Injection Pressure on Thermal Efficiency.Gas injection pressure affects the NOR concentration in the exhaust gas (Fig. 5.2).At low load, higher gas injection pressure reduces the NO R. In all cases a substantialreduction in NOR as compared to the baseline is observed up to medium load although athigh load the NO R concentration is slightly higher.5045 -A40-35 -a 30 -x0 25 - DIESEL ••••Z^BASELINE20 -15-50 bar10 -a5-0  ^ 4BMEP ( bar ) =>Figure 5.6 : Effect of Gas Injection Pressure on Oxides of Nitrogen.The effect of gas injection pressure on unburned hydrocarbon (HC) is shown inFigure 5.7. HC emissions are strongly dependant on gas pressure and the best gas pressuredepends on load. At high load the HC for any gas injection pressure exceeds the emissionstandard by at least 100%. However, the emission standard regulates only non-methaneunburned hydrocarbons.60RPM^: 1200Diesel Ratio : 25%Shrouding : 50%Pgas : 70 bar80 bar^\528-24 -22 -20 -18- 18-14-1= 12-Fa 10-a 8 -2^-fs.-0^4 -^2^^0 ^ RPM^: 1200Diesel Ratio : 25%Shrouding ; 50%DIESELBASELINEEPA S3P94A^4^6BMEP ( bar ) ,===>.Figure 5.7 : Effect of Gas Injection Pressure on Unburned Hydrocarbon.It can be seen from Fig. 5.7 and 5.5 that the concentration of the unburnedhydrocarbon in the exhaust gas is related to the thermal efficiency. A a higher HC in theexhaust gas suggests a lower thermal efficiency, while a lower HC is associated with ahigher thermal efficiency.The amount of the unburned fuel gas in the exhaust gas is a good indication of howthe combustion progressed. Assuming all measured unburned hydrocarbon in the exhaustgas is methane (CH4), gaseous fuel which survives the combustion process can be6162estimated. Table 5.1 lists the unburned gas ratio' for three injection pressure and for highand low load cases. The high load value at each injection pressure corresponds to minimumHC emissions and highest thermal efficiency.Pio , bar 50 60 70BMEP, bar 0.4 2.8 0.5 2.6 0.5 4.0Unburned GasRatio0.544 0.019 0.447 0.023 0.374 0.031Table 5.1 : Effect of Gas Injection Pressure on Unburned Gas Ratio.As shown in Table 5.1, the gas injection pressure significantly affects the unburnedgas ratio. The lowest unburned gas ratio is 2.4% which is experienced with 60 bar gasinjection pressure at its optimum load 3 (BMEP = 2.6 bar). At low load operation, highergas injection pressure greatly reduces the unburned gas ratio. At low loads 37-54% of theinjected gaseous fuel survives combustion. This is associated with very late burning ratherthan misfiring, especially with higher gas pressure. Injection gas pressure appears to be oneof the important factors controlling the combustion process. This suggests that the fluidmechanics of gas distribution in the combustion chamber are of great importance.2 The unburned gas ratio is defined as the ratio of the unburned gas in the exhaust to theinjected gaseous fuel.3 Optimum load is defined as the load that has maximum thermal efficiency.31-A 29-27-25 -23 -21-19-17-15-13-11-9-7-5-3-1 0^0:5^1:5^2^2:5^6^3:5^4^46^65.4 Effect of Diesel Ratio.63BMEP ( bar) ==>Figure 5.8 : Effect of Diesel Ratio on Thermal Efficiency.Tests were carried out with different diesel ratios to study the effect of pilotquantity on thermal efficiency. Fig. 5.8 shows the effect of diesel ratio on thermalefficiency. With 25% diesel ratio the thermal efficiency is about the same as with 20%.Lowering the diesel ratio to 15% reduced the thermal efficiency at low load operation byabout 1.5%, but the engine operation is less smooth than with the 20% and 25% dieselratio operation. However, at high load the 15% diesel ratio operation is associated with64thermal efficiency of about 3% higher compare with the other diesel ratios. This suggeststhat the gas-diesel engine needs a minimum quantity of pilot fuel to operate smoothly andefficiently.BMEP ( bar) ^Figure 5.9 : Effect of Diesel Ratio on Oxides of Nitrogen.The specific NO  in the gas-diesel engine emission is significantly affected by thediesel ratio (Fig. 5.9). Lower diesel ratio produces lower specific NO„ in most of loadranges except at very high load (BMEP > 3.7 bar). Fig. 5.9 suggests that increasing theRPM^: 1200Pgas 80 barShrouding 50%20%EPA '93/'94Diesel Ratio ; 25%15%,DIESELNE65amount of pilot-diesel tends to increase the NO„ concentration in the exhaust gas. However,regardless of the diesel ratio, the NO„ in the gas-diesel engine is lower than that of thediesel engine.4BMEP ( bar ) ^Figure 5.10: Effect of Diesel Ratio on Unburned Hydrocarbon.Fig. 5.10 shows the specific unburned hydrocarbon in the exhaust gas of the gas-diesel engine operated with 15, 20, and 25% diesel ratio. The minimum in HC emissionsis associated with 20% diesel ratio. Operating the engine up to BMEP — 3.5 bar with 15%diesel and gas injection pressure of 60 bar gives large amounts of unburned hydrocarbonA2422Zo18161412108a4204BMEP ( bar) ,===>31  29-27-25 -23-21 -191715131197-5-3666compared with that of the 20 and 25%. This indicates particularly poor combustion withthe 15% diesel ratio at low load.5.5 Effect of Pilot-Diesel Cetane Number.Figure 5.11: Effect of Pilot-Diesel Cetane Number on Thermal Efficiency.Tests were carried out using a higher cetane number diesel fuel to examine theeffect of ignition delay of the pilot-diesel. Fig. 5.11 shows the effect of pilot-diesel cetanenumber on thermal efficiency. A significant improvement in the part-load thermalRPM : 1200Pgas : 60 barDiesel Ratio : 20%Shrouding : 50%DIESELBASELINEPILOT-DIESEL : DF2CN62EPA '93/ '940^1^ 567efficiency results from using higher cetane number pilot-fuel (CN62) although at higherload the thermal efficiency is substantially lower. Decreasing the quantity of the highercetane number pilot fuel by lowering the diesel ratio seems to overcome the drawback ofusing higher cetane number pilot fuel at higher load operation.As shown in Fig. 5.12, the NO  in the exhaust gas is affected by changing thepilot-diesel cetane number. The use of higher cetane number pilot-diesel causes the loadrange to shift. The concentration of NO shifted in same pattern as the loads thus the NO.concentration at the same load is higher for a higher cetane number in most cases.5550454035302520151050BMEP ( bar ) ^Figure 5.12: Effect of Pilot-Diesel Cetane Number on Oxides of Nitrogen.201088420PILOT-DIESEL : DF2N\ilkA4DIESELBASELINEEPA 13/14RPM^: 1200Pgas : 80 barDiesel Ratio : 20%Shrouding : 50%68Fig. 5.13 shows the effect of using two different pilot-fuel cetane numbers on theconcentration of the unburned hydrocarbon in exhaust gas (HC). The use of higher-cetane-number diesel-pilot significantly reduces the amount of HC, and provides bettercombustion. This is one of the reason for using a higher cetane number pilot-diesel (CN62).BMEP ( bar ) ^Figure 5.13 : Effect of Pilot-Diesel Cetane Number on Unburned Hydrocarbons.695.6 Summary.The effect of engine performance and emissions of the gas injection pressure, dieselratio, and pilot-fuel cetane number in a gas-diesel engine have been examined. The resultswere compared against the diesel baseline.The gas-diesel has better thermal efficiency at full load while at low load thethermal efficiency is nearly as good as that of the conventional diesel.Relative to the 1994 EPA bus standard, emissions are a serious problem both withconventional diesel and gas-diesel pilot opeation. At full load NO„ is excessive, while atpart load unburned HC (probably mostly unburned fuel) considerably exceeds the standards.High unburned HC associated with low thermal efficiency.Gaseous fuel injection pressure affected the load capability. Increasing gas pressurestends to extend load capability, improve thermal efficiency at full load, and reduce NO upto high load operation. However, higher gas pressure at low load operation producesrelatively higher unburned HC.Results of the tests done with different diesel ratios show that relatively smoothoperation at low load was achieved with high diesel ratio, although at high load relativelyhigher thermal efficiency was obtained with low diesel ratio.Tests carried out using a higher-cetane-number pilot-diesel demonstrated that byreducing the pilot-diesel ignition delay time, the thermal efficiency is better and the engineoperation is smoother. At the same time the concentration of unburned HC is reduced.These results indicate problems with combustion which are further investigated inChapter 6 with the aid of combustion rate analysis.6. COMBUSTION RATE ANALYSISThe analysis of combustion rate in internal combustion engines is a technique usedto obtain the fuel mass-burned fraction or the combustion rate from a measured timehistory of the cylinder pressure.This chapter is divided into twelve sections. The first section provides a briefexplanation of the analysis. The second section describes the combustion model used. Themixture composition of the unburned gases is discussed in the third section. Itsthermodynamics properties are evaluated in section four. Section five explains how thetable for burned-gases properties is formed. The method to account the effect of heattransfer to the cylinder wall is described in section six. The computation procedure isexplained in the seventh section. Measurements of cylinder pressure and results of analysisare presented and discussed in the last three sections before the summary.6.1. General.Engine pressure data is the principal data in the combustion rate analysis. It is thecombustion part of the pressure history that the analysis needed. The combustion processis considered to start at the beginning of injection (BOI) near the end of compressionstroke, although the compression auto-ignition occurs a few crank angle degrees after BOI.7071This delay is termed ignition delay. A typical cylinder pressure distribution is shown inFigure 6.1. The fuel mass-burned fraction at a certain crank position is directly related tothe pressure difference AP between the measured pressure P.. (with combustion) and thepolytropic compression or expansion pressure P rnot obtained without combustion.PeY BOl^TDC522* °CAFigure 6.1: Typical Cylinder Pressure DataIn the thermodynamic analysis a two-zone model has been chosen to simplify thecomplex combustion process in the engine. Governing equations of this analyis are themass conservation equation and energy conservation equation.726.2. Formulation of the Combustion Model.Combustion in the gas-diesel engine is initiated by spontaneous ignition of a portionof the injected diesel fuel when the air temperature and pressure in the combustion chamberare above the ignition point of the diesel fuel. This happens after a delay period of a fewcrank angle degrees from the beginning of fuel injection and near the end of compressionstroke; it could happen at several places in the cylinder. The result is that the localtemperature is increased enough to autoignite and bum the gaseous fuel.FUELSINJECTORTime-stepC.AFigure 6.2: Schematic of the Combustion Model73A two-zone combustion model has been adopted to simplify the combustion processto estimate the combustion process. Figure 6.2 shows a schematic of the thermodynamicsystem in the engine cylinder while combustion is in progress. Consider the presence oftwo zones in the control volume, i.e., those of the burned and the unburned gases.The following assumptions are applied :1. The pressure inside the cylinder is uniform.2. The cylinder constituents (both unburned and burned) behave as ideal gases.3. The thermodynamic state of each zone is considered to be homogeneous anduniform.4. The natural gas is considered to be methane (CH 4), and the diesel fuel canbe represented by CH2 . Vaporization of the injected diesel fuel is assumedto take place very quickly, with allowance made for heat of vaporization.All fuel is assumed to be present in the combustion chamber in vapour onedegree after the beginning of injection.5. The combustion products are in thermodynamic equilibrium.6. The mass-burned fraction of diesel fuel and natural gas are the same at anyinstant during the combustion process. This is called the proportionalburning assumption.Temperatures of these two zones are Tb and T. respectively. Specific energy content in theburned zone is u b which is a function of both temperature Tb and pressure P sincecombustion products are partially dissociated. The unburned zone has a specific energy74content of u u which is a function of temperature. For polytropic compression of theunburned gases, their temperature T. will be a function of cylinder pressure P.At any crank angle position, the system specific volume v and its specific energyu can be evaluated as followsV(0)v -^ (6.1a)m totU0 - W + Q.,,and^ u = tot^ (6.2a)mu. mtotwhere V is the cylinder volume, 0 is the crank angle, m,„, is the total mass of the cylindercontents, E t., is the total energy of the system which is a summation of is the internalenergy of the cylinder contents at some reference point U. , the work done on the pistonW , and the heat transfer to the wall Q., (a negative quantity).Define x as the fuel mass-burned fraction at any instant during the combustionprocess , v the specific volume, and u the specific energy of the system (and letsubscripts u and b denote unburned and burned gas properties). The properties of theunburned and burned gas at a given time can be expressed by the following two governingequations :mass conservation,^v = x. vb(P,Tb) + (1-x)• vu(P)^ (6.3)energy conservation,^u = x- ub(P,Tb) + (1-x)• u.(P) (6.4)Given P, v, and u , the system of the above two equations can be solved iteratively for Tband x.75The calculation is done stepwise in a small time interval corresponding to a 1degree crank angle. The volume of the cylinder at any crank angle position V(0) isdetermined from the geometry of the cylinder (Appendix H). The total mass of thecylinder contents m is the summation of the mass of incoming air trapped in thecylinder, the exhaust gas remaining from the previous cycle (residual gas), and the fuelmasses. The procedure used to calculate the residual gas fraction, and the incoming airtrapped in the cylinder is explained later.The specific volume of the system v is the mean specific volume in the timeinterval which is taken to be the specific volume at the end of the step. It is calculated byusing the Eq. (6.1a).The energy of the system is the total energy up to the end of the chosen timeinterval which is calculated by using Eq. (6.4a), if the total internal energy of the systemEtm is known. The latter is determined from the first law of thermodynamics which isapplied to the system for a small time change At as followsdE = 8Q - 8Wwhere Q is the heat transfer to the system and W is the work done to the piston.Then, integrating the Eq. (6.5) from state 1 to state 2 for a given time step At which iscorresponding to 1 degree crank angle, will gives,DES = .1Q2 - .11312^ (6.5a)By defining the average pressure of the system for the given time step asP1 + P2-2the work done to the piston can be written asP1 +P. 1 W2 - ^2+P2 (V2 -Vi)The method of accounting for the effect of heat transfer is explained in Sect. 6.6. For themoment we note only, that if SQ = 0 then substituting Eq. (6.6) into Eq. (6.5a) gives theenergy of the system at the end of the step asP1 +P E2 =^(^22 ) V2 - ) (6.5b)Using this last relationship in Eq. (6.2a), we can compute the specific energy of thesystem.U = E2^ (6.2b)mwr6.3. Mixture Composition.The combustion process is preceded by compression of the cylinder charge. Thecomposition of the cylinder charge depends on the type of gas exchange process in betweentwo consecutive cycles.76(6.6)EXHAUST VALVEOPEN ; WAIDCCLOSE ; dTPABDC1,/"..*°.".--- r=:;• EXHAUSTINTAKE. PORTS (IE)rosmoNsJFRESH AIR^FROM^AIR=-1)SLOWER 1110X73•ARDCIIDC1Figure 6.3 : Schematic of the Uniflow-scavenged Configuration.The test engine is a two-cycle with a uniflow-scavenged configuration as illustratedschematically in Fig. 6.3. A root-type blower develops a slightly higher pressure in the airbox surrounding the cylinder. Through inlet ports the fresh air enters the cylinder anddisplaces the previous-cycle cylinder contents. A part of the incoming fresh air mixes withthe residual gas and is expelled with it. However, some of the burned gas remains in thecylinder after this process. To determine the cylinder charge composition, it is importantto estimate the amount of this remaining gas or residual gas as well as the portion of thefresh air trapped in the cylinder.77786.3.1. Scavenged Air.To estimate the amount of air trapped in the cylinder at the end of the scavengingprocess the following definitions are used.■ The delivered air mass, m ad , is the mass of air delivered to the engine per cycleas measured at the air-intake line.■ The mass of air trapped, m ai,. , is the portion of the delivered air mass per cycletrapped in the cylinder when the piston is at the inlet port closure (IPC) position.■ The residual mass, m res , is the mass of the combustion products remaining fromthe previous cycle.■ The trapped mass, m fr , is the mass of the cylinder contents at IPC which is thesummation of the mass of the air trapped and the residual mass.= matr + inns■ The delivery ratio, A , is defined as the ratio of the delivered air mass per cycleto the trapped mass.A -MadMa.■ The degree of purity of the charge, DP , is defined as the ratio of the mass of airin trapped cylinder charge to the mass of trapped cylinder charge.DP =tr^M res + Matr79It indicates the degree of mixing of the delivered air with burned gases in thescavenging process, and is shown later to be a function of delivery ratio.DP = f(A)■ The residual fraction, fres , is defined as the mole ratio of the residual gas nr. tothat of the trapped charge in the cylinder at IPC n iz. .Tres^NestrIf it is assumed that the residual gas has the same molecular weight as that of thefresh air, we can writeco res cores fres _ Itt^+ cotr res^t,.or^fres = 1 - DPThere are two limiting ideal models in the scavenging process [31] [32], i.e., perfectdisplacement, and complete mixing. In perfect displacement, the incoming fresh gasestotally displace the burned gases without any mixing. Complete mixing occurs if enteringfresh mixture mixes instantaneously and uniformly with the cylinder contents.Following the method described by Heywood [31], imagine a boundary surface inbetween the fresh charge and the burned gases; then we can writeDP = A^for A s 1^ (6.7a)DP = 1^for A > 1 (6.7b)80For the complete mixing, during a small time interval dt in the scavenging processthe delivered air entering the cylinder is dm ad while the instantaneous air mass in cylinderis mat, and the instantaneous total mass in cylinder is mom. Assuming the incoming massflow rate is equal to the exit mass flow rate while keeping in mind that it is a fully mixedprocess, we can writedmaa. = dmad dmad ( m°")mtrIf mtr is constant during the process, then"tar )nta^= 1 - (^- In(1- )mgrtr^M+ CTaking the intital condition, mad/ma. = ma,,/ma = 0 , C = 0.Then ?flair = 1 - e Ma.and by using the definition of Degree of Purity, DP , and Delivery Ratio, Awe get^DP = 1 - CA^ (6.8)In the actual scavenging process, the scavenging configurations of the the engineaffects the degree of purity. Typical data for large two-stroke diesels is shown in Fig. 6.4.It shows that the uniflow scavenging is the most effective scavenging configuration.1.0t! 0.4E. 0.60.84r •to,•"'s%6>42\, '44^111111,1i^/2111 11111111 11 11ili^\ \ 0.2Unillow scavengingLoop scavengingCross scavenging8100.4 0.6^0.8 1.0^1.2Delivery ratio A1.4^1.6Figure 6.4 : Scavenging Data Typical of Large Two-stroke Diesels [31].In the determination of the amount of air trapped in the cylinder, a fit to the averagevalue of the degree of purity as a function of delivery ratio for the uniflow scavenging areaas shown in Fig. 6.4 is used. This third-degree polynomial curve fit is implemented in thecalculation of the fuel-air equivalence ratio filed in Appendix I.6.3.2. Residual Gas.The composition of the residual gas is the same as that of the burned gas. However,in the determination of the residual gas composition, complete combustion (with thesimplifying approximation that the products consist of CO2 , H 2O , 02 , and N2) is82assumed.If r is the diesel-to-gas mass ratio and 4) is the fuel-air equivalence ratio ascalculated in Appendix 1, the complete combustion equation for the methane and dieselfuel (CH2) is2 + 3 16rCH4 + -1 6^2 14r CH2 +^( 02 + 3.76 N2 )14(1+-16r) CO2 + (2+-16 r) H2O14^14(2+-3 —16r) (1-4)) 02 + (2+-3 —16 r) 3.76 N22 14^•^2 14(6.9)To determine the amount of residual gas, the residual fraction^as defined in sub-section 6.3.1 is used.63.3. The Unburned Gas Composition.The fuels are assumed to be injected within the first time step after the beginningof injection. Then the unburned gas in the beginning of the combustion process consist ofthe trapped air, residual gas, and the fuels.Taking 1 mole methane in a complete combustion as shown in Eq. (6.9), thenumber of moles of the i species Ft; of the unburned gases are as follows :nCH4 = 114= — r164)nCH216^fres r)14^1-4.16^fres r)14^1 —fres022+-322+-3nN2co2 =2(1 +nH20 = (1+16 r14^fres  1-,[ 1 +1 -f16 r14^f( 1^r_  ) (336)1 -f83and the sum of the moles of the unburned gases is6En i^n + n CH 2 + n + nN + n + nH0c H4^2^CO2^2where n i is the number of moles of the unburned gases species.The relative population of the unburned gases species is determined using the followingrelationship.YE -ni6E nia=i6.4. Thermodynamics Properties of the Unburned Gas.In the combustion process, the properties of the unburned gas are determined fromknowledge of its composition as derived in Sect. 6.3. Its composition is constant in the84combustion process because of the assumption of proportional burning in the combustionmodel. The molecular weight M. and the gas constant R. of the unburned gas arecalculated as followsMu (6.10)=^ (6.11)Mywhere n, is the number of moles, M 1 is the molecular weight of the constituent gases,and R is the universal gas constant.The unburned gas temperature T. at the end of a step is calculated from its pressureincrement (or decrement) in a polytropic compression (or expansion) process. Note that theinjection of the fuels is assumed to occur in the first crank angle degree after BOI, so thatthe combustion starts one degree after BOI. The unburned gas temperature at BOI iscomputed from the total mass of cylinder charge using the ideal gas relation since itspressure and volume at BOI are known. The molar specific heat at constant volume C„of the unburned gas is calculated from its molar specific heat at constant pressure C p„which is a function of its temperature only [331.C. = Cpu6E nE [ CdT ) lE_Mwbut,Cpu85Then,C.6E^Cp0(T )i+ 1 R.(6.12) MuThe isentropic exponent of the unburned gas is determined from its constant-volumespecific heat value C,,„ and gas constant R„ .R1 + ."C(6.13)The specific volume of the unburned gas is calculated from the ideal gas equation,VuR. T. (6.14)Pwhere P is the cylinder pressure.The unburned gas enthalpy H. is the summation of its constituent enthalpies H 1each of which is a function of temperature as a result of the ideal gas assumption.Hu6E^[ /7f1= 1f EpoM err ],298.15where Y i is the i th unburned gases species, h f is its enthalpy of formation at standardstate condition (298.15K and 0.1 MPa), Cpo is the constant-pressure specific heat of the86unburned gas species. The heat of vaporization of diesel fuel is accounted for at this point.Using the definition of enthalpy,^u = h -Pvthe internal energy of the unburned gas is. - P v,,^ (6.15)6.5. Thermodynamic Properties of the Burned Gas Mixture.The burned gas internal energy u b and specific volume v b are found from aproperties table by specifying the pressure and temperature. This table of properties isgenerated using STANJAN [34] which is a chemical equilibrium solver using the JANAFthermodynamic property tables. Since the engine consumes two different fuels, i.e., dieselfuel and natural gas, we need two different sets of tables as data. To cover adequately allpossible equivalence ratios of the engine working condition, a set of six tables weregenerated with equivalence ratio of 0.20, 0.40, 0.60, 0.80, 0.90, and 1.00. In generatingthe tables, it was assumed that the diesel fuel can be represented by CH 2 , and the naturalgas by CH 4 . Each table lists the properties for ten different temperatures and ten differentpressures as listed in Table 6.1. These two sets of tables provide the burned-gas propertydata.87TemperatureKPressureatm.1500 11700 21900 32100 52300 102500 202700 402900 603100 803300 100Table 6.1 : List of Temperatures and Pressures used inBurned Gas Properties TableAs stated in Sect. 6.2, the combustion of diesel fuel and natural gas are assumedto be proportional so that at any time the mass-burned fraction of two fuels are the same.If r is the mass ratio of diesel to natural gas, then by replacing ( CH 2 + 16/14 r CH2 ) withCH. as described in Appendix I, the burned gas properties can be determined from a tablefor a given n ( 2 < n < 4 ) at the given fuel-air equivalence ratio. The procedure to formthis table form the two sets of tables ( CH4 and CH 2 ) is as follows.1.^For each fuel, interpolate the tables to obtain a table for the given fuel-airequivalence ratio.0XX882.^Interpolate the above two tables to form a table for a given n (carbon to hydrogenatom ratio).Then, the set-up table serves specifically a particular engine operating condition, i.e. atcertain diesel to gas mass ratio at a given fuel-air equivalence ratio.6.6. Heat Transfer to the Cylinder Wall.The amount of heat transfer to the cylinder wall directly affects the calculation ofthe system total energy as can be seen in Eq. (6.2b) and subsequently the calculation of themass-burned fraction. Methods of direct estimation of the instantaneous spatially averagedheat flux to the cylinder wall have uncertainties as shown in Sect. 2.5.8 min^0 maxe Figure 6.5 : Typical Mass-burned Fraction Results.89To account for the effects of heat transfer, an indirect method has been used.Exhaust composition measuments provide the concentration of total unburned hydrocarbons.If it is assumed that all of these hydrocarbons represent unburned (methane) fuel, then themaximum mass-burned fraction x. can be determined. This will substantially differ fromthe value obtained in the adiabatic calculation because the effect of heat transfer issubstantial. Fig. 6.5 illustrates the difference between the calculated mass-burned fractionand the estimated actual values.The broken line in Fig. 6.5 (for the adiabatic calculation) indicates negative valuesof x before the combustion pressure rise begins; this is a direct result of heat transfer andis corrected for by the following procedure. Let X, be the adiabatic calculation value andx be the estimated true mass-burned fraction. The approximate correction procedure is asfollows.(i) For^eboi^< em ioex = ^)(-x)°min - °kg(ii) For Oimn < 0 < 0.(6.16a)x = xa +^+ (  e - mb, ) (x - xl + xmin)0 - Omin^(6.16b)(iii)^For^0^0,nu(6.16c)90This procedure for estimating the distribution of mass-burned fraction is quiteapproximate. The chief reason for this is that the cylinder heat transfer rate is predictableonly with a wide range of uncertainty. However it should be borne in mind that theimportant features of the calculation are :(i) The delay period before there is appreciable burning.(ii) The characteristic burning duration.(iii) The apparent presence (or absence) of distinctive burning periods for thepilot-diesel fuel and natural gas. The curves in Fig. 6.5 do not indicate sucha distinction.6.7. Calculation Procedure.In calculating the combustion rate, determination of the initial conditions isimportant. The mass of air trapped in the cylinder has to be calculated first. This parameteraffects directly the fuel-air equivalence ratio whose calculation procedure is filed inAppendix I.The combustion rate calculation starts when the fuels are injected into thecombustion chamber. The pressure data is used to determined the work done to the pistonand subsequently the specific energy of the system is determined as explained in Sect. 6.2.A table of properties for the burned gases is formed to use in iteratively solving for themass-burned fraction.The combustion rate calculation procedure is as follows:1.^Determine the trapped air mass, matr , and the fuel-air equivalence ratio, (1).912. Determine the gas constant of the unburned gas,^(Eq. 6.11).3. Set the burned-gas table for the given fuel-air equivalence ratio and hydrogen tocarbon ratio.4. Obtain the total mass in the combustion chamber from the trapped mass, rn tr , andthe fuel masses.5. Set the initial condition of the combustion process, i.e., the unburned gasestemperature and internal energy (T„ and u u ). The unburned gas temperature isdetermined using the total mass, pressure, and the volume information.6. Calculate the isentropic compression (or expansion) coefficient of the unburnedgases, yu , and the unburned gas temperature at the end of the calculation step. Asmentioned in Sect. 6.2, the calculation step is taken to be 1 crank angle degree.7. Evaluate the unburned gas properties at the end of the time step. Specific volume,vu , is calculated from the ideal gas relationship. Its internal energy, u u , isdetermined by knowing its temperature (Eq. 6.15).8. Compute the work done during the time step using the cylinder pressure information(Eq. 6.6).9. Evaluate the specific energy of the system at the end of the step u m from Eq. 6.2b,and its specific volume vm from Eq. (6.1a).10. Given P, T, , vu , um , and vm , iteratively solved the two governing equations forburned gases temperature, Tb , and fuel mass-burned fraction, x, by looking at theburned table. The iteration procedure is documented in Appendix J.11. Prepare for next calculating step by resetting the initial condition.9212.^Repeat step 6 to 11 for the next calculating step.To implement the above computation procedure, an existing mass-burned fractioncomputer program for a four-stroke spark ignition engine was modified. The listing of theis documented in Appendix K together with the burned-gas tables.Appendix L contains the record of a verification of the computation procedure forthe constant volume case.6.8. Pressure Measurements.The engine cylinder-pressure data were acquired using a PCB 112A05 air-cooledpiezoelectric pressure transducer. As described in Sec. 3.4.6 and shown in Fig. 3.7, it wasmounted through an adaptor sleeve which has a single passage. This mounting techniqueavoided the problem of thermal shock encountered when the transducer was flush mountedand thus directly exposed to the combustion gas. As explained in Sec. 3.5, the cylinder-pressure transducer analog signals were digitized and recorded. To interpret the signals,they must be referenced to a known pressure at some point in the engine cycle. The air-boxintake port pressure was chosen to be the reference pressure at BDC.Fig. 6.6 shows a typical pressure crank-angle diagram at low and high load. Eachcurve represents the ensemble-averaged pressure for 100 consecutive cycles at optimumB01. Note that the peak pressure magnitude is higher for the higher load than for the lowerone. The location (in terms of degree CA) seems to be moving towards the TDC as theload increases; this suggests a late burning at low-load operation. Operating the engine at1i0 160 1'0 180 i90 260 210 220 260 240e (*ABDO ) ==>93higher load resulted in higher cylinder pressure since more fuel was burned in thecombustion process.Figure 6.6 : Typical Pressure Crank-Angle Diagrams at different Loads.The cylinder-pressure data can also be plotted against the cylinder volume usinglogarithmic scales to show the polytropic of the compression and expansion strokesexplicitly. Fig. 6.7 shows a typical log PV diagram for one cycle each of high and lowload. As can be seen from Fig. 6.7, the polytropic exponent n comp of the compression strokeis similar for both high and low loads. The value of n c,„„p deduced from the compressionshown in Fig. 6.7 is 1.23 (at both loads). For high load expansion stroke (aftercombustion), the value of the polytropic exponent nem, is 1.29. At low load it was notdetermined because of the late burning situation.4i^AS^4A^4.4^•^as^-7!^44Ln (Vcyl) 1=:.( a )94AVa54A^4i^4.4^4.4^4^-Ti^42^4ALn (Vcyl) 4=>(b)Figure 6.7 : Typical Log P - Log V Diagram.956.9 Indicated Work.Work transfer from the gas to the piston, which is calculated from the pressure data,is defined as indicated work Wind. It is equal to the area enclosed in the linear P-V diagramwhich is obtained by integrating around the curve.Wind = c P dV^ (6.17)Brake work Wb is defined as the measured work per cycle which is calculated fromthe brake mean effective pressure BMEP.Wb = 100 BMEP Vd^ (6.18)where brake work is in kJ, BMEP in bar and displaced volume Vd in m3 .Comparing the indicated work with the brake work is one way to check the validityof the pressure data acquired. Since the indicated work has to overcome the friction inengine parts and to drive engine accessories, the indicated work is greater than the brakework.Fig. 6.8 shows the comparison of the indicated work to the brake work of the testengine operated with 60 bar gas injection pressure and 20% diesel ratio at engine speed1200 rpm. As seen in Fig. 6.8, the slope of the line fitted to the indicated work is verynearly parallel to that of the brake work which is exactly a straight line. This demonstratesthat the friction loss is essentially independent of load.RPM^: 1200P9as : 60 barDose' Ratio : 20%Shrouding : 50%960.60.5OA0.30.20.102^3^4BMEP ( bar ) c==>Figure 6.8 : Comparison of Indicated Work to Brake Work.6.10 Cyclic Variation.Variations of the pressure data, and indicated work, from cycle to cycle are termedcyclic variations. Assessing cyclic variations is one way to evaluate the quality of thecombustion process.Fig. 6.9 shows the standard deviation and the relative standard deviation of theindicated work at different loads. This information reflects the cycle-by-cycle combustionvariation which is important to note in interpreting the mass-burned fractions resulting fromthe corresponding pressure data.^2^3^3.8BMEP ( bar) =>Figure 6.9 : Standard Deviation and Relative S.D. of the Brake Work.The cyclic variation of the high load case (BMEP = 3.8 bar) of the gas-dieselengine operating with 60 bar gas injection pressure and 25% diesel ratio is shown in Fig.6.10 which displays the pressure-volume data for ten successive cycles. At this load, therelative standard deviation of indicated work for 100 cycles (as shown in Fig. 6.9) is 5.6%.At low load (BMEP = 1 bar), the relative standard deviation is 36.7% which indicates‘pa,fa•a•49841^41^41^4^4S^as^41Ln ( Ay! ) r==',>Figure 6.10 : Superposition of 10 Successive Cycles.6.11 Mass-burned Fractions.We can defined the ignition time delay period as the crank angle difference betweenthe first appearance of burning and the actual BOI which is about 3° greater than BOI.Fig.6.11 shows an example of measured cylinder pressure, estimated cylinderpressure, and normalized mass-burned fraction for the operating condition at 1200 RPM,BMEP = 3.8 bar, 20% diesel(CN62) ratio, 60 bar gas injection pressure, 10° poppet angle,50% shrouding. Here the electronic BOI is 32°BTDC, actual BOI is approximatelyTDCUnburned-gm^;Temp. (K/1000)-^-^-MEP:4 MR1 OMMass - bum*-1--FesitimParm-)9929°BTDC, and the apparent ignition delay is about 31 CA degree (4.3 ms). The unburnedgas temperatures during ignition delay period is estimated to be less than 1000 K. Thisignition delay time is approximately an order of magnitude less than those of methanealone presented by Fraser, Siebers and Edwards [36].^lap^200delay-- _combustionperiod ---140 220Crank Angle ('ABDC ) r==>Figure 6.11 : Cylinder Pressure, Temperature and Mass-burned Fraction Distribution.Fig. 6.12 shows mass-burned fraction curves at different loads while the engineworked with the same injection pressure and diesel ratio at best BOI. We see from Fig.6.12 that the ignition delay period 30-35° and is not strongly dependent on load (except forRPAI^- 1200• 00 berMt Rob - 20%Shicidna^50%100the 1 bar case). The combustion period' is roughly twice as long for the 2 bar case as forthe 3.8 bar case. The mass-burned fraction curves do not show any sign of two distinctstages of burning (for the pilot liquid diesel and the natural gas).1.0A 0.80.8LL. 0.4E-9 0.220140 180 180 200 220 240 280 280 300Crank Angle ( °ABDC ) ===>Figure 6.12 : Combustion Pattern at different Loads.Fig. 6.13 shows the normalized mass-burned fraction curve for 3 different gasinjection pressures with the engine operated at about 4 bar. For the highest pressure theignition (i.e., pressure rise) delay period is about 4° less than with 50 bar and the1 The combustion period is defined as the crank angle difference between the x = 0.05 andthe x = 0.95 points.101combustion period not much affected by gas pressure. The shape of the burning curved ismuch the same in all 3 cases; no distinct burning period for the pilot fuel can be observed.LLf1.110.9 -0.8 -0.7 -0.6-0.5  -0.4-0.3 -2 0.2-0.1-0140Crank Angle ( °ABDC ) ^Fig 6.13 : Combustion Pattern for different Gas Injection Pressures.6.12 Summary.The combustion rate analysis utilized to estimate the mass-burned fraction from thecylinder pressure data has been elaborated. The analysis used a two-zone combustion modelwhich solved the mass and the energy conservation equations.Estimation of the amounts of fresh air trapped in the cylinder, and the remaining102combustion products from the previous cycle in the scavenging process followed theprocedure recommended by Heywood [31] has been employed. The burned gas propertiestable used in the computation is generated using STANJAN [34].An indirect method to account the effect of the heat transfer by correcting andnormalized the mass-burned fraction curve resulted from an adiabatic calculation has beenadopted. The unburned hydrocarbons in the exhaust gas is considered in normalizing thecurve.The cylinder pressure data have been measured and recorded for selected engineoperating conditions. The indicated work calculated from the pressure data has beencompared with the brake work to check the validity of the pressure data acquired.A high cyclic variation in the cylinder pressure data indicates poor combustion atlow load as does the high concentration of unburned hydrocarbons in the exhaust gas.The calculated unburned gas temperatures at TDC ranged from 900 K at low loadto 1000K at high load; these are much lower than reported autoignition values for methane(approaches 1300 K).The calculated mass-burned fractions suggest that the amount of heat transfer islarge.Combustion rate analysis shows that the burning rate increases as the load increases.Longer ignition delay time and longer burning time are associated with low load operation.The burning rate seems to be increased by increasing the gas injection pressure.7. CONCLUSIONS AND RECOMMENDATIONS.7.1 Conclusions.Performance, emissions and cylinder pressure data of a diesel-pilot gas injectionengine have been investigated. A combustion rate analysis was employed with the pressuredata to study the burning rate pattern. The following may be concluded :1. With high pressure injection of natural gas and about 20% diesel-pilotenergy ratio the thermal efficiency exceeds that of the conventional dieselengine efficiency at high load. It is less than that of the conventional dieselengine efficiency at low load in the present configuration.2. With the same pilot-diesel ratio, lower gas injection pressure provides betterthermal efficiency at low load than that higher injection pressure. Theopposite is true at high load.3.^Low-load thermal efficiency of the gas-diesel engine depends strongly onthe pilot diesel energy ratio; this indicates that the pilot diesel ratio plays animportant role in helping the natural gas to ignite.1031044. Using a higher cetane number pilot-diesel decreases the ignition delay timeand improves thermal efficiency at low load.5. Lower pilot-diesel ratio produces less oxides of nitrogen (NO) emissionswhich shows the use of natural gas lowered the NO concentration in theexhaust gas.6. Although methane concentration in the exhaust gas was not measureddirectly, high unburned hydrocarbons emissions suggest that a considerableamount of injected gaseous fuel survived the combustion at low load.7. High cyclic variation in the cylinder pressure data is one evidence of poorcombustion at low load. Another is the high rate of unburned hydrocarbonsemission.8.^The calculated mass-burned fractions indicated a large amount of heattransfer in this low-compression-ratio low-speed test engine. They also serveto indicate, especially at low load, long pressure rise (ignition) delay timesand long burning times. No distinct burning curve for the pilot-diesel wasobserved.1059. Combustion rate analysis shows that the burning rate depends on load; thehigher load the higher burning rate than that of the lower load.10. Calculated temperatures of the unburned gas at TDC ranged from 900 K atlow load to 1000 K at high load. These low temperature (associated withlow compression ratio and heat transfer) are probably the main reason forthe low quality of combustion, particularly at high load.7.2 Recommendations.1. Continue experimental observation using a better unburned hydrocarbonsanalyzer which is capable of measuring the methane concentration in theexhaust gas. Determination of the unburned methane helps to provide abetter understanding of the combustion process.2. Consider the possibility to supply the engine with a higher air temperatureto study the effect of the unburned gas temperature on the ignition delaytime and the combustion quality.3.^If analytical work on the combustion rate analysis of the cylinder pressuredata is to continue, an engine heat transfer correlation model should beconsidered in the computation. This is to give a more accurate estimationof the mass-burned fraction.1068 REFERENCES1^Boyer, R.L. and Crooks, W.R., "The Modem Gas Engines," ASME PaperNo51-0GP-4, 1951.2^Karim, G.A., "The Dual Fuel Engine of the Compression Ignition Type - Prospects,Problems and Solutions - A Review," SAE Paper No.831073, SAE Trans. vol.92pp 569-577, 1983.3^Song, S. and Hill, P.G., "Dual Fueling of a Pre-Chamber Diesel Engine, withNatural Gas," J.Eng for Turbines and Power, Trans. ASME, Vol.107 pp 914-921,October 1985.4^Ding, X. and Hill, P.G., "Emissions and Fuel Economy of a Pre-Chamber DieselEngine with Natural Gas Dual Fueling," SAE Paper 860069, SAE Trans Vol. 95pp.612-625, 1986.5^Gettle, L.E., Perry, G.C., Boisvert, J. and O'Sullivan., P.J., "Dual Fuel EngineControl Systems for Transportation Applications," J.of Eng. for Gas Turbines andPower, Vol 109, Oct. 1987, pp. 435-438.6^Gettle, L.E., Perry, G.C., Boisvert, J. and O'Sullivan., P.J., "Microprocessor DualFuel Diesel Engine Control System, SAE Paper No.861577, Oct. 1986.7^Beck, N.J., Johnson, W.P., George, A.F., Petersen, P.W., van der Lee, B., andKlopp, G. "Electronic Fuel Injection for Dual Fuel Diesel Methane", SAE TechnicalPaper 891652, Aug. 1989.8^Miyake, M., Biwa, T., Endoh, Y., Shimotsu, M., Murakami and S., Komoda, T.,"The Development of High Output, Highly Efficient Gas Burning Diesel Engines,"CIMAC Paper D112, Conference Proceeding, Paris-France, Jun. 1983.1071089^Einang, P.M., Koren, S., Kvamsdal, R., Hansen, T. and^Sarsten, A.,"High-Pressure, Digitally Controlled Injection of Gaseous Fuel in a Diesel Engine,with Special Reference to Boil-Off from LNG Tankers," Proceeding CIMACConference, Paris-France, Jun. 1983.10^Wakenell, J.F., O'Neal, G.B. and Baker, Q.A., "High Pressure Late Cycle DirectInjection of Natural Gas in a Rail Medium Speed Diesel Engine," SAE TechnicalPaper 872041, Nov. 1987.11^Mc.Cuiston, F.D.,Jr., Lavoie, G.A. and Kaufmann, "Validation of a Turbulent FlamePropagation Model for a Spark Ignition Engine," SAE Trans Vol 86, pp.200-221,1977.12^Marvin, C.F.,Jr.,"Combustion Time in the Engine Cylinder and its Effects onEngine Performance," NACA Tech Report 276, 1927.13^Rassweiler, G.M. and Withrow,L., "Motion Pictures of Engine Flames Correlatedwith Pressure Card," SAE Journal (Trans.), Vol 42 pp 185-204, May 1938.14^Shayler, P.J., Wiseman, M.W. and Ma, T., "Improving the Determination of MassFraction Burnt," SAE Paper 900351, 1990.15^Amann, C.A., "Cylinder-Pressure Measurement and Its Use in Engine Research",SAE Paper 852067, 1985.16^Krieger, R.B. and Borman, G.L., "The Computation of Apparent Heat Release forInternal Combustion Engines," ASME Paper 66-WAIDGP-4, 1966.17^Borman, G. and Nishiwaki, K., "Internal Combustion Engine Heat Transfer," ProgEnergy Combust Sci, Vol.13 pp.1_46, 1987.18^Annand, W.J.D., "Heat Transfer in the Cylinders of Reciprocating InternalCombustion Engines," Proc Instn Mech Engrs, Vol.177 No.36, 1963.10919^Woschni, G., "A Universally Applicable Equation for the Instantaneous HeatTransfer Coefficient in the Internal Combustion Engine," SAE Paper 670931, 1967.20   " 1990 SAE Handbook 1990 Volume 3," Society of AutomotiveEngineers, Inc., 1990.21^Hames, R.J., Straub, R.D. and Amann, R.W., " DDEC - Detroit Diesel ElectronicControl, " SAE Paper No. 850542, 1985.22^Hill, P.G., Pierik, R.J. and Hodgins, K.B., " Intensifier-Injector Technology," USPatent No. 5 067 467, 1991.23^Ding, X., " Operation Manual for Exhaust Analysis System," Report No. AFL-85-04, Mech Eng Dept - The University of British Columbia, Dec.1985.24^Rohling, N.R., " Operation and Performance of Emissions Console," UnpublishedReport, Mech Eng Dept - The University of British Columbia, Aug.1990.25^Randolph, A.L., "Cylinder-Pressure-Transducer Mounting Techniques to MaximizeData Accuracy," SAE Paper 900171, 1990.26^Lancaster, D.R., Krieger, R.B. and Lienesch, J.H., "Measurement and Analysis ofEngine Pressure Data," SAE Paper 750026, SAE Trans., Vol. 84, pp 155-172, 1975.27^Kach, R.A. and Adamczky, A.A., "Effects of Thermal Loading on PressureMeasurement in a Combustion Bomb," Rev. Sci. Instrum. 56, American Institute ofPhysics, 1985.28^Yuen, D., Hodgins, K.B., " Data Aquisition System for Alternate Fuels EngineTesting," Unpublished Report, Mech Eng Dept - The University of BritishColumbia, May 1991.11029^Ouellette, P., "High Pressure Injection of Natural Gas for Diesel Engine Fueling,"M A.Sc Thesis, Mech Eng Dept, The University of British Columbia, 1992.30 ^,"Installation and Operation Instructions - Meriam LFE Laminar FlowElement," Meriam Instrument, 1981.31^Heywood, J.B., "Internal Combustion Engine Fundamental,"  McGraw-Hill Inc., NewYork, 1988.32^Benson, R.S. and Whitehouse, N.D., "Internal Combustion Engines,"  PergamonPress, Oxford, 1979.33^Van Wylen, G.J. and Sonntag, R.E., "Fundamental of Classical Thermodynamics, Third Edition," John Wiley & Sons, New York, 1987.34^Reynold, Wm.C., "Chemical Equilibrium Solver Ver.3.60 - An Application Software," Mech Eng Dept, Stanford University, 1987.35 ^, "Single Cylinder Two-cycle Diesel Engine Power Plant - Model 1-71, Description and Operating Manual," Diesel Engine Div - General Motor, 1939.36^Fraser, R.A., Siebers, D.L. and Edwards, C.F., "Autoignition of Methane andNatural Gas in a Simulated Diesel Environment," SAE Paper 910227, 1991.9. APPENDICES9.1 Appendix ADIESEL ENGINE EMISSION STANDARDS'Model Year Nox HC2 CO PM1990 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 0.0531998 4.0 1.3 15.5 0.05Table A.1 : Urban Bus Heavy-duty Engine Emission Standardsg/ measured during EPA heavy duty engine test2 Non Methane Unburned Hydrocarbon Gas.3 Proposed level. However, may be relaxed to 0.007 g/  if technology is not availableto meet the proposed level.111Model Year NOx HC2 CO PM1990 6.0 1.3 15.5 0.601991 5.0 1.3 15.5 0.251994 5.0 1.3 15.5 0.101998 5.0 1.3 15.5 0.10112Table A.2 : Heavy-duty Truck Engine Emission Standards9.2 Appendix B113CALIBRATION CURVES1. Engine Speed.ENGINE SPEED CALIBRATION CURVE(MAGNETIC TYPE SENSOR)1^1400ciww 1200i1^1000211 NO800400400^800^800^1000^1200^1400imi44 ENGINE SPEED, RPM(AS OF DIGITAL HAND TACHOMETER)Figure B.1 : Speed Calibration CurveE 140iur 120D0CC MO0t-ow 80CCDQ eo<w2I200402. Load Sensor.114TORQUE SENSOR CALIBRATION(STRAIN-GAGE LOAD CELL)0^20^40 60^80 100 120 140+ IDADIPE^• UNLIICON100 APPLIED TORQUE, N.mFigure B.2 : Torque Calibration Curve3. Diesel Fuel Mass-flow.DIESEL  FLOW CALIBRATIONAVL FUEL BALANCE11565Lm^4■oY3LL2^2I0AVL output cm Ilbratton,4.994 C ko/nr)/Volt0^ 2^ 4APPLIED FLOW, ko/hr+ Calibration Data ---- Error-free ReadingFigure B3 : Diesel Mass-flow Calibration CurveScale : 200 psi/VoltI^1^1^1^1^1^1^1^1^1^1^1^1^1^1^1^1^1^1^14. Cylinder Pressure Sensor.CALIBRATION OF PRESSURE TRANSDUCERPC8 112A sin 10118, Sens 1 17 mVipst1162 1002.0001.9001.6001.7001.6001 500▪ 1.400ui^1.300D 0^1.20013m c^1.100w°Ica.  o 1.000o▪^0.900Wix LI^0.600• 0.700• 0.6000.5000.4000.300^ 2000.1000.0000.000 0.200 0.400 0.600 0.800 1.000 1.200 1.400 1.600 1.600 2.000(Thousands)APPLIED PRESSURE, PSI^ MEASURED PRESSURE - IDEAL READINGFigure B.4 : Cylinder Pressure Sensor Calibration Curve5. Intake Air Flow.117Figure B.5 : Laminar Flow Element Calibration Curve [30].9.3 Appendix CCNG' DF22 CN623Lower Heating Value, kJ/kg : 49,051 45,220 45,220Density, kg/m' 0.7114 860 836Cetane Numbers n/a ---45 62.2Table C.1 : Properties of Test Fuels118Natural Gas (BC Gas). Averaged data in the period of April 1989 up to December 1991.Diesel Fuel No.2 (Chevron).High Cetane Number Diesel Fuel.At STP (25°C,101.3kPa).The cetane number of a diesel fuel is defined (SAE J313 JUN89 [20]) as the percentageby volume of normal cetane in a blend with heptamethylnonane required to match theignition quality of the test fuel.9.4 Appendix DConstituent Percent by VolumeMethane^,CH4 95.5Ethane^,C2H6 3.0Propane^,C3H8 0.5Butanes plus' 0.2Nitrogen 0.6Carbon Dioxide,CO2 0.2Total^100.0Table D.1 : Composition of Natural Gast119'This includes butane (C4H io) and all heavier hydrocarbons.2Nominal gas analysis for 1989, BC Gas.05' TUBESTAINLESS STEELS VAG ELOK FITTINGSSTAINLESS STEEL0.5' SUIAG.-NPTSS UNION3/4-1/2' REDUCER"- (BRASS)2-3/4' REDUCER(GAL V.IR ON)^ 2' TEE JOINT(GAL V. IRON)TO EMISSION CONSOLE9.5 Appendix E<=I TO EXHAUST STACK ^ FROM^ ENGINE120Figure E.1 : Exhaust Sampling Probe in the Engine Exhaust PipeDIESEL SUPPLYUNEACCESSORY SHAFTDRIVEN ACTUATORDDECSOLENOIDfrii*1^rI•t• I^eI•illirf^ 1• An r7,,,Ordrir Sir^44.02\1....^.rezo i 21/ A ■ '^Vi ‘^11^&^I ..—^\ IliVALVE^s,.1k \^SPRING::^ % ■,, \ Ns": -4 CHECKRETURN^CHECK\ s.,4k^OPPETNIXING^\N *^PNOZZLERESERVOIRSPOOLVALVEUNITINJECTORmetering valveRELIEFVALVEPUSHRODBYPASSCHECK VALVECNG Storage(20-200 bar)INTENSIFIER9.6 Appendix F.Pol;Pot^ U.S. Patent • M067A67 Hoverter 1991Figure F.1 : Schematic of the Two-fuel Unit Injector'.121VALVEImplementation of the "Intensifier-Injector Technology" [12]. Courtesy of K.B. Hodgins.1229.7 Appendix GPOWER CORRECTION FACTOR CALCULATION METHOD(in accordance with SAE J1349)1. The correction is made against the standard inlet air conditions, i.e.,Inlet Air Pressure (Absolute)^: 100 kPaInlet Air Temperature : 25°C (298K)Dry Inlet Air Pressure (Absolute)^: 99 kPa2. The correction factor fc,,,.,. applied to the observed brake power depends on theatmospheric factor t and engine factor f„, which is calculated using the empiricalrelationship :^fc. = ( fa ) fmThe atmospheric factor is calculated based on dry inlet air pressure Bdo andinlet air temperature t,t = [ 99 / Bdo ] [(t+273) / 298 ] °.7The engine factor depends on the fuel flow F(g/s), the engine displacementD(dm3), the engine speed N(RPM), and the pressure ratio r of inlet manifoldpressure Po to inlet air pressure Bo . It has the following value :to = (0.036 q/r) - 1.14^if 40 < (q/r) < 65fm = 0.3^if (q/r) < 40fn, = 1.2 if (q/r) > 65where q = (60,000 F) / (D N)since the engine is a two-stroke cycle engine.3.^The correction factor used in the calculation is within the range of 0.90 to 1.10.1239.8 Appendix HDETERMINATION OF THE CYLINDER VOLUMEThe volume of the cylinder is a function of the crank angle. The reference positionis taken when the piston at its farthest postion from the cylinder head, i.e. when piston atits bottom dead centre or BDC. The nearest the piston could approach the cylinder headis the top dead centre or TDC. The distance in between BDC and TDC is the stroke ofthe engine, s.Figure A6.1 : Piston Top Surface Profile.The top of the piston has a shallow bowl shape as shown schematically in Figurea6.1. Its volume was measured by pouring oil into the bowl (60 cm3). Adding the volumeabove the piston (17 cm3) to the piston-bowl volume gives us the clearance volume of thecylinder V, (77 cm 3). Subsequently the compression ratio rc is found to be 16.02 whichis determined as the ratio of the maximum cylinder volume ( displacement volume plusclearance volume) to the minimum cylinder volume (clearance volume).124s/2 cosFigure A6.2 : Geometry of the cylinder.The geometry of the cylinder, piston, and connecting rod is shown in Fig.A6.2where B(0.10795 m) is cylinder bore, s(0.127m) is stroke, 1(0.254m) is connecting rodlength, r is crank radius ( r = s/2 ), e is crank angle. The clearance height c is defined2as the height in the cylinder for the clearance volume, while p is the distance between125t2he piston pin axis to the piston top surface.The cylinder volume at any crank position V(0) isV(8) =4 B2 z^ (H.1)where z is the distance between the piston top and the cylinder head. Referring to Fig.A6.1,z is determined in the following way.z = c + —s + (p+0 + 2—s cose -^t2 - 2sine?®+ p ]2or^z = c +2 + 1 +2 cose - l2 -(.sine?^(H.2)Since the geometrical properties of the engine is known, the cylinder volume at any crankposition V(0) is determined from Eq. (H.1) and Eq. (H.2).The accuracy of the total cylinder volume calculation is limited by the accuracy ofboth the clearance volume measurement and the determination of crank angle.Determination of crank angle depends on the accuracy of determining the reference pointthat is the BDC. It is assigned by using a dial indicator to locate the lowest point of pistonposition. In addition, to minimized the effects of changes in mechanical clearancesdetermine the BDC, procedure as suggested by Lancaster et al [26] was followed.1269.9 Appendix IDETERMINATION OF FUEL-AIR EQUIVALENCE RATIO1. Definitions■^The diesel to gas ratio r is defined as the mass ratio of the diesel fuel mom, to thenatural gas nigas injected into the combustion chamber at the beginning of injection(BOI).r = dist ■^The combined-fuel hydrogen to carbon atomic ratio, n , is defined as the ratio ofthe total hydrogen to the total carbon atoms of the two fuels in a proportionalburning assumption.We can replace^( CH4 + 16/14 r CH2 )^by^CH„where164+ — r(2)H ^14^ 28 + 16 rn =^- -C^1 + 16T^7 + 8 r— 14(L1)Using the above definitions, the complete combustion equation for the two fuels canbe written as,1+^ 1+n11^It4 I 1.1^n pr^4 4CH. +^16..2 + 3.76 N2 )^CO, •^AA2.... • I,^1 ) 02 +2 41^2From which the stoichiometric fuel-air ratio is,127FAsa - (12 + n) ( 1 + 4) (32 + 3.7648)(L2)2. Calculation procedureGiven the experimental data as follows,Pamb Tamb Pabox 9 mad 9 Inds' mg.where Pamb is the ambient pressure, Taro is the ambient temperature, Paw, is the air-boxtemperature, and assuming the scavenged-blower efficiency is li mo , the air gas constant isR. and its specific heat ratio is y , and the residual gas temperature is T, , , the fuel-airequivalence ratio calculation procedure is :1. Compute the the air temperature in the airbox, Tabor , i.e. at scavenged-blower exit.y -1pT^= T [1+ --^6---#—°x)Y -1}]^lio ^P aid)2. Assume Tit..3.^Calculate the trapped mass.P VRaabox ipcTo,4.^Determine the delivery ratio.— mad1285. Obtain the delivery ratio, determine the degree of purity DP using the fit curve ofthe typical uniflow-scavenged data as shown in Figure 6.4.DP = 1.1527 A3 - 0.4094 A2 + 0.0251 A6. Recalculate the T ipc ,Tip: = Tab°, DP +^(1 - DP)7. If^I Tip: - T1 ^5_ 0.5 , proceed to the next step.If not, revise Tipc and repeat step 2 up to 7.8. Compute the mass of air trapped.inau. = DP mu.9. Calculate the residual fractionf,.es = 1 - DP10. Obtain n from the diesel to gas ratio using Eq. (I.1), and the calculate thestoichiometric fuel-air ratio FAswich using Eq. (I.2).11. Compute the fuel-air ratio.m, + mFA - ^ gas12. Determine the fuel -air equivalence ratio.- FA • FA,1299.10 Appendix JITERATION PROCEDUREThe two governing equations, i.e. Eq. (6.3) and Eq. (6.4), can be rewritten asfollowsVr = X' Vb + (1—X)* V ag^ (J.1)um = .x• ub + (1—x)• u.^ (J.2)Rearranging,Vat — Va = X' ( Vb — Vs )— Um = X' ( Ub — Us )From which,XVm — Va — UM — UN— ^V b — Vu^Ub — UN^ — V^b vuLet^A = in^a = V — V^ (J.3)u — u^ub — UNm^ahence^Vb — Vii = A• (ub — u.)or Vb = A ub — A u. + v.If^B = — A u1 + vm^ (J.4)then^Vb = A ub + B (J.5)130In the beginning of the iteration, the following parameters are given,T2 9 P2 9 Vm^Vu 9 Um 9 Vuwhere vm is the system specific volume, um is its specific internal energy , and v. is theunburned gas specific volume, uu its specific energy. Then the iteration procedure is asfollows :1. Calculate^A = v - v— uY2. Calculate^B = — A u. +3. Assume Tb4. Look up the value of the burned gas specific volume v b and specific energyub the burned gas table for the assumed value of Tb and the measuredpressure P2 .5. Check whether the equation Eq. (J.5), which is derived from the twoconservation equations, is satisfied.vb = A ub + B6. If the above equation is satisfied then proceed to the next step. If not, reviseTb and repeat step 4 and 5.7.^Calculate the mass-burned fraction using Eq. (J.1)x = Vas — Vatvb — v.8.^Iteration completed.9.11 Appendix KLISTING OF THE MASS-BURNED FRACTION PROGRAMC This is program xpgdsl.for which takes engine pressure dataC at regular crank angle increments DCA and determines mass-C burned fraction.C modified for gas-diesel engine by H.Gunawan (June19,1992)C^from xpresse.for.IMPLICIT REAL*8(A-H 2O-Z)REAL*8 MAIR,MDSL,MGASREAL*8 MREF,MTRAP,MATRAP,MTOTREAL*8 CA(360),P(360),XMB(360),T(360)REAL*8 PAVG(360),XAVG(360),TAVG(360)REAL*8 WRK(360),WAVG(360)REAL*8 QWL(360),QAVG(360)REAL*8 TU(360),TUAVG(360)REAL*8 TBRN(360),TBAVG(360)CHARACTER*50 PDATCOMMON/PROPS/RDG,RHC,FRES,EQVR,RUCOMMON/STATS/N,NCYC,NCA,CABOI,RPMCOMMON/GEOM/BORE,STROKE,ROD,CLRHCOMMON/MASS/MTOT,MA1R,MDSL,MGASCOMMON/PORT/PABOX,CAIPC,PIPC,T1PC,TRESCOMMON/AMBNT/PAMB,TAMBCOMMON/BURN/UBCC Specify the cylinder geometryC BORE is cylinder bore(m), STROKE(m), ROD is corm rod length(m),C CLRH is clearance height(m).STROKE = 0.1270D0BORE = 0.10795D0ROD = 0.2540130CLRH = STROKE/15.0D0Cc*******************Initialize***********************C^EQVR is the fuel-air equivalence ratio; subscriptsC^ipc and boi refer to intake port closing and beginningC^of injection, respectively.C^CA is crank angle and subscript PR1C^refers to the first pressure record. RU is the gas constantC^for the unburned gas per kg and CVU is the specific heat.C^NPR is the number of lines of pressure records per cycle.C^NCYC is the number of engine cycles to be analyzed;C^NCA is the no. of CA intervals after spark to be analyzedC^for each cycle.C^TIPC is the cylinder contents temp at ipc after mixingC^with residual gas. The residual gas mass fractionC^is determined using a scavenging data typical ofC^two-stroke diesels.OPEN(UNIT=2,FILE='PDATA.DAT',STATUS='OLD')OPEN(UNIT=10,FILE= XPGDSL. OUT' ,STATUS='NEW' )OPEN(UNTT=11,FILE='CH4EQ20.DAT',STATUS=' OLD')OPEN(UNIT=12,FTLE='CH4EQ40.DAT' ,STATUS='OLD')OPEN(UNTT=13 ,FILE='CH4EQ60. DAT' ,STATUS=' OLD')131OPEN(UNIT=14,FILE='CH4EQ80.DAT',STATUS='OLD')OPEN(UNIT=15,FTLE='CH4EQ90.DAT',STATUS='OLD')OPEN(UNTT=16,FILE='CH4EQ10.DAT',STATUS='OLD')OPEN(UNTT=21,FILE='CH2EQ20.DAT',STATUS='OLD')OPEN(UNIT=22,FILE='CH2EQ40.DAT',STATUS='OLD')OPEN(UNIT=23,FILE='CH2EQ60.DAT',STATUS='OLD')OPEN(UNIT=24,FILE='CH2EQ80.DAT',STATUS='OLD')OPEN(UNIT=25,FILE='CH2EQ90.DAT',STATUS='OLD')OPEN(UNTT=26,FILE='CH2EQ10.DAT',STATUS='OLD')READ(2,*)RPM,PAMB,TAMB,PABOX,TRESREAD(2,*)MAIR,MDSL,MGAS,UHCRATREAD(2,*)CAPR1,CAIPC,CABOI,DCAREAD(2,*)NPR,NCA,NCYCWRITE(6,40)READ(6,41)PDATWRITE(1 0,42)PDAT40 FORMAT(1X,'PDATA.DAT = ?')41 FORMAT(A50)42 FORMAT(1X/IX,'Pressure data : ',A50)CC Estimate the mass of air in trapped cylinder charge MATRAPC ,and the equivalence ratio EQVRCALL INTAKE(MATRAP)MTRAP = MATRAP/(1.D0-FRES)MTOT = MTRAP+MDSL+MGASCWRII'E(10,50)NCYCWRITE(10,51)RPM,EQVR,CABOIWRITE(10,52)MAIR,TIPC,CAIPCWRITE(10,53)FRES,MATRAP,MTOTWRITE(10,54)RDG,RHC,UHCRAT50 FORMAT(/,1X,14,'cycles of reduced by XPGDSL.FOR')51^FORMAT(/,1X,'RPM',F7.1,' Equiv Ratio ',F6.3,' CAboi',F8.3)52^FORMAT(1X,'Mair kg ',D10.4,' Tipc K',F6.1,' CAipc',F9.3)53^FORMAT(1X,'Fres',F8.3,' Matrap kg ',E10.4,' Mtot kg ',E10.4)54 FORMAT(1X,'Mdsl/Mgas ',F7.4,' RHC, Ruhc = ',2(1X,F5.3))WRITE(10,103)print*,Init.ializing subroutines'CALL UNBURNED(Q1,Q2,Q3,Q4,1)CALL BURNED(Q1,Q2,Q3,Q4,Q5,Q6,Q7,Q8,1)CALL CYCSTATS(Q1,Q2,Q3,1)PBOIAV = 0.D0DO 90 I = 1,NCAPAVG(I) = 0.D0TAVG(I) = 0.D0TUAVG(I) = 0.D0TBAVG(I) = 0.D0QAVG(I) = 0.D0WAVG(I) = 0.D090 XAVG(I) = 0.D0print*,'reading pressure data'DO 1000 N = 1,NCYCDO 100 I = 1,NPR100 READ(2,*)CA(I),P(I)ICIPC = DINT((CAIPC - CAPR1)/DCA) + 11001 = DINT((CABOI - CAPR1)/DCA) + 1VIPC = VCYL(CAIPC)VBOI = VCYL(CABOI)PBOI = P(KBO1)CA(1) = CABOI + DCA132DO 125 I = 1,NPR-KBOI.GT. 1)CA(I) = CA(I-1) + DCA125 P(I) = P(I+KBOI)PBOI1 = P(1)CABOI1 = CABOI+1.D0VBOI1 = VCYL(CABOI1)TUBOI1 = PBOIl*VB0I1/RU/MTOTPRINT*,'TUBOI1 = ',TUBOI1CALL UNBURNED(TUBOILUU,CVU,VISC,2)ETOT = UU*MTOTV1 = VBOI1P1 = PBOI1TU1 = TUBOI1Ti = TU1XMB1 = 0.D0C WRITE(10,104)CABOILPBOILTUBOI1103 FORMAT(8X,'CA',10X,'P kPa',8X,'Tu K',10X,'Tb K',11X,'X')104 FORMAT(1X,5(2X,E12.6))C****calculate conditions at end of each crank angle interval******XMAXP = 0.D0XMAX = O.DOXMIN = 0.D0CAXMIN = CABOICAXMAX = CABOIDO 200 I = 2,NCAWR1TE(6,*)'calculating step ',I,' of cycle ',NCALL UNBURNED(TU1,UU,CVU,VISC,3)G = 1 .D0/(1.DO+CVU/RU)GAMMA = LDO + RU/CVUTU2 = TU1*(P(I)/P1)**GCALL UNBURNED(TU2,UU,CVU,VISC,2)VU = RU*TU2/P(I)V2 = VCYL(CA(I))ASURF = ACYL(CA(I))CALL QWALL(T1,V2,XMBI,ASURF,DQWL,1)DWRK = (Pl+P(I))/2.D0*(V2-V1)ETOT = ETOT - DWRK + DQWLVM = V2/MTOTUM = ETOT/MTOTCALL BURNED(P(I),TB,VU,VM,UU,UM,VB,XMB(I),2)IF (N .NE. 1) GO TO 1111WRITE(10,104)CA(I),P(I),TU2,TB,XMB(I)C^WRITE(10,1104)CA(I),P(I),XMB(I),T1,GAMMA,TBC WRITE(10,1999)V2,VM,VU,VB,UM,UU,UB1104 FORMAT(1X,6(E11.5,1X))1999 FORMAT(1X,'V2,VM,VU,VB,UM,UU,UB=',7(1X,D9.3))1111 CONTINUEIF(XMB(I) .GT. XMAXP) GO TO 170IF(XMB(I) .GT. )(MIN) GO TO 180XMIN = XMB(I)CAXMIN = CA(I)GO TO 180170 XMAXP = XMB(I)CAXMAX = CA(I)180 CONTINUETBRN(I) = TBTU(I) = TU2T(I) = TB*XMB(I) + (1-XMB(I))*TU2WRK(I) = WRK1 + DWRKQWL(I) = QWL1 + DQWLC*************prepare for next step133CINHdOIS(r£11AIXVVD)SIVISDAD TIVDHfINILNOD 0051HIA111:(1)DAVX(1)DAVX(1)VD(170I'01)a1111A1 005I3((XI `5'I I H)VXIAVIM111103 5011(I)DAV (1)DAVIIIUDAVIADDAVMDAVX(1)VD(50 I I '0 OanxtikbAatsavold/Wonvd = (1)DAVd(DADMIVOINWDAVELL = (1)DAVEII(DADN).LVOINWDAVIlL = (1)0AVf1I(DADN),LVOINWDAVI = WDAVI(DADN)IV0INWDAVA1 = (T)DAVM(DADIN)IVO'13/WDAVO = (1)DAV6(DADN).LVOINWDAVX = (I)DAVX1A1d11/0(19/(10EIVD - (I)VD) = HVELVDNI = I 0051 00AVIOHXIOHVD(170 I '0(DADN).LVOIVAVIOEld = AVIOHdCs m1111:Xe.X`XII`API d.`X0r.VD:X8/LVW2103 IOI(IOVOI)HIIHMCsuope.g paumq-ssum ptre samssaid p8As-aicumasuo: xiAvmod zoi(ZOsapko Ae 10j spiCretm popstimsHIINILNOD 0001(z`HIAIXVVD)SIVISDAD TIVDI0Hd + AVIOHd = AVIOHd(I)d + (1)DAVd = (1)DAVd 00£WfIL + (1)DAVIlL = (1)DAVIII(1)NIIELL + WDAVELL = WDAVHIWI + (1)DM/I = (I)DAVI(1)'1M) + (I)DAVO = (I)DAVO(1)3RINt + (I)DAVM = WDAVAA+ (I)DAVX = (I)DAVXXVIAIX/WELIAIX = WHIAIXVDNI = I 00£ ouHIINILNOD 08ZXVIAIX = WHIAIX08Z OI ODMOD + WHIAIX = WmArxxvinuu*avorrauKisaixvp - (Dva) + raysix- = mop OZZ08Z 01 OD+ (Dam = Wasix(NalAix-) * IIAIONRCIA 'MEND (I)vD) =ozz oI OD (CHAIXVD^(Dva)OtiZ oI oD (XVIAIXVD^(I)VD)VDNI = I 08Z OCINLIAIXVDNIIAIX.= NHAIXVD'NIIAIX.'*.LNIHdXVIAIXVD`dXVIADC.= XVIAIXVD'dXVIAIX.'*INIadNHAIX + dXVIAIX - XVIAIX = XVIAIACINIFAIXVD - XVIAIXVD = ZIAIONHCIII0EIVD - NIIAIXVD =- OCT = XVIAIX21343 11"9 1°3 S°TISFWIS °P*********DHfINILNOD 00Z(I)MAIX = tam TI=(!)}BIM = 13111M(1)1Mt) = MAO=(Da =ZA = IAPEIc****************************************************************DOUBLE PRECISION FUNCTION VCYL(CA)IMPLICIT REAL*8 (A-H 2O-Z)COMMON/GEOM/BORE,STROKE,ROD,CLRHC CA is crank angle degrees ABDC.PI = 3.14159D0APSTON = PI/4.DO*BORE**2CAR = CA*P1/180.D0Z = (1.D0 + 2.D0*ROD/STROICE + DCOS(CAR)1^- DSQRT((2.D0*ROD/STROKE)**2 - (DSIN(CAR))**2))2^* STROKF-2.D0 + CLRHVCYL = Z*APSTONRETURNENDC**********************************************************SUBROUTINE INTAKE(MATRAP)C**********************************************************IMPLICIT REAL*8(A-H 2O-Z)REAL*8 MAIR,MDSL,MGASREAL*8 MREF,MTRAP,MATRAP,MTOTREAL*8 D(3)COMMON/AMBNT/PAMB,TAMBCOMMON/PORT/PABOX,CAIPC,PIPC,TIPC,TRESCOMMON/MASS/MTOT,MAIR,MDSL,MGASCOMMON/PROPS/RDG,RHC,FRES,EQVR,RUC Calculate the blower-exit air temperature, TXBLOEFFBLO = 0.751)0GSTAR = 0.2857D0TABOX = TAMB*(1.D0+((PABOX/PAMB)**GSTAR-1.D0)/EFFBLO)C Estimate the residual gas mass fraction, Fres=Mres/(Mres+Matrap)C and the mass of air in trapped cylinder charge, MATRAPRA = 0.2871)0PIPC = PABOXPIPC = 200.D0VTRAP = VCYL(CAIPC)C Start iteration to find Degree of Purity20 DEGP = 0.6030 D(1) = 0.D0D(2) = 0.D0D(3) = 0.D0M = 140 DEGP = DEGP + 0.05D050 TIPC = TABOX*DEGP + TRES*(1.D0-DEGP)MTRAP = PABOX*VTRAP/RA/TIPCRDELIV = MAIR/MTRAPDEGPUR = 0.173611D0*RDELIV**3.D0-0.95982D0*RDELIV**2.D01^+1.774305*RDELIV - 0.19642D0Y = TIPC - (TABOX*DEGPUR+I.RFS*(1.D0-DEGPUR))IF (Y .GT. 0.D0) GO TO 52D(1) = DEGPGO TO 5352 D(2) = DEGP53 D(3) = D(1)*D(2)IF (D(3) .EQ. 0.1)0) GO TO 40M=M+1DEGP = 0.5D0*(D(1)+D(2))PRINT*,'Deg.of Purity = ',DEGPIF (DABS(Y) .LE. 0.5130) GO TO 100IF (M .LT. 50) GO TO 50100 CONTINUEFRES = 1.D0 - DEGPUR135TIPC = (1.DO-FRES)*TABOX + FRES*TRESMATRAP = (1.DO-FRES) * MTRAPWRITE(10,60)TIPC,RDELIV,MTRAP60 FORMAT(1X,'Tipc,Rdeliv,Mtrap=',3(4X,E10.4))C RDG is mass ratio of diesel-fuel(CH2) to gas(CH4)C Replace CH4 + (16/14)*RDG CH2 by CHnC where n = RHC = (4+16/14*RDG*2)/(1+16/14/*RDG)CC CHn + (l+n/4)/EQVR ( 02 + 3.76 N2 )C ^> CO2 + n2 H2O + (l+n/4)(1-EQVR)/EQVR 02 + (l+n/4)/EQVR*3.76 N2CC RFASTO is the stoichiometric fuel-air ratioC RFA is the fuel-air ratioRDG = MDSLAVIGASRHC = (28.D0+16.D0*RDG)/(7.D0+8.D0*RDG)RFASTO = (12.D0+RHC)/((l.D0+RHC/4.D0)*(32.D0+3.76D0*28.D0))RFA = (MDSL+MGAS)/MATRAPEQVR = RFA/RFASTOPRINT*,'FRES,EQVR,RHC =',FRES,EQVR,RHCRETURNENDCc****************************************************************SUBROUTINE UNBURNED(TU,UU,CVU,VISC,L)c****************************************************************IMPLICIT REAL*8(A-H 2O-Z)REAL*8 MWO2,MWN2,MWCH4,MWCH2,MWH2O,MWCO2REAL*8 HFO2,HFN2,HFCH4,HFCH2,HFH2O,HFCO2REAL*8 DHO2,DHN2,DHCH4,DHCH2,DHH2O,DHCO2REAL*8 CP02,CPN2,CPCH4,CPCH2,CPH20,CPCO2COMMON/PROPS/RDG,RHC,FRES,EQVR,RUIF(L .NE. 1) GO TO 10MWO2 = 31.999D0MWN2 = 28.013D0MWCH4 = 16.043D0MWCH2 = 14.026D0MWH2O = 18.015D0MWCO2 = 44.011DOCc********methane_diesehair mixtures*****************************C Residual gas fraction affect properties of unburned gas onlyC CH4 + (16/14)RDG CH2 + (2+(32)(16/14)RDG)/EQVR (02+3.76N2)C ^> (1+(16/14)RDG) CO2 + (2+(16/14)RDG) H2OC^+ (2+(32)(16/14)RDG) (1-EQVR)/EQVR 02C^+ (2+(32)(16/14)RDG) 3.76 N2CRDG1 = (16.D0/14.D0)*RDGYAIR = (2.D0+(3.D02.D0)*RDG1)/EQVRFSTAR = FRES/(1.D0-FRES)BOTTOM = LDO + RDG1 + YAlR*4.76D0 + FSTAR*1^(3.D0+2.D0*RDG1+YAIR*(1.D0-EQVR)/EQVR+YA1R*3.76D0)XCH4 = 1/BOTTOMXCH2 = RDG1/BOTTOMX02 = YAIR*(1.D0+FSTAR*(1.D0-EQVR)/EQVR)/BOTTOMXN2 = YAIR*3.76D0*(1.DO+FSTAR)/BOTTOMXCO2 = (1.DO+RDG1)*FSTAR/BOTTOMXH2O = (2.DO+RDG1)*FSTAR/BOTTOMTOP = MWCH4 + RDG1*MWCH21 + X02*BOTTOM*MWO2 + XN2*BOTTOM*MWN22 + XCO2*BOTTOM*MWCO2 + XH20*BOTTOM*MWH20WTMOL = TOP/BOTTOM136RU = 8.3143D0/WTMOLHFO2 = 0.0D0HFN2 = 0.0D0HFCH4 = -74873.0D0HFCH2 = -32059.0D0HFH2O = -241827.D0HFCO2 = -393522.D0PR1NT*,'WTMOL,RU =',WTMOL,RUC WRITE(9,45) WITVIOL,RUC WRITE(9,46) XCH4,XCH2,X02,XN2,XCO2,XH2045 FORMAT(1X,'WTMOL, RU = ',2(1X,F8.3))46 FORMAT(1X,'X CH4,CH2,02, N2, CO2, H20 =',6(1X,F8.4))RETURN10 CONTINUETDIM = TU/100.0D0TDIMSQ = DSQRT(TDIM)TDIM2 = TDIM*TDIMTDIM3 = TDIM2*TDIMTDIM4 = TDIM3*TDIIVITDIM14 = TDIM**0.25D0TDIIvL54 = TDIM**1.25D0TDIM74 = TDIIVI**1.75D0TDIM32 = TDIM**1.5D0TDIM52 = TDIM**2.5D0TDIM34 = TDIM**0.75D0CIF(L .NE. 2) GO TO 20C *****Calculate molar enthalpy differences between 298K and TU,C^then calculate the internal energy UUDI102 = 3743.2D0*TDIM+0.80408D0*TDIM52+35714.0D0/1^TDIMSQ-23688.0DO/IDIM-23906.63D0DHN2 = 3906.0DO*TDIM+102558.0DO/TDIMSQ-107270.0D0/1^TDIM+41020.0DO/TDIIV12-39673.0D0DHCH4 = -67287.0D0*TDIM+35179.2D0*TDIM54-1421.43D0*1^TDIM74+64776.0DO*TDIMSQ-39436.89D0DHCH2 = 10418D0*TDIM+2327.6DO*TDIM2-51714.3D0DHCH2 = DHCH2/12.D0DHH2O = 14305.0D0*TDIM-14683.2DO*TDEV154+5516.73D0*1^TDIM32-184.95DO*TDIM2-11876.23D0DHCO2 = 6914.5D0*TDIM-40.265D0*TDIM74-40154.0D0*1^TDIIVISQ+70704.0D0*TDIM14-43912.73D0UU = (X02*(HF02+DH02)+XN2*(HFN2+DHN2)+1 XCH4*(HFCH4+DHCH4)+XCH2*(HFCH2+DHCH2)+XH20*(HFH2O+DHH20)+2 XCO2*(HFCO2+DHCO2))/WTMOL - RU*TURETURN20 CONTINUEIF(L .NE. 3) GO TO 30C ******Calculate specific heat CVU **********CP02=37.432D0+0.020D0*TDIM32-178.57D0/TDIM32+236.88DO/TDIM2CPN2=39.060D0-512.79D0/TDIM32+1072.7D0/TDIM2-820.40D0/1DIM3CPCH4 = -672.87D0+439.74DO*TDIM14-24.875D0*TDIM34+1^323.88D0/TDIMSQCPCH2 = 104.18D0+46.55D0*TDIMCPCH2 = CPCH2/12.D0CPH2O = 143.05D0-183.54DO*TDIM14+82.751D0*TDIMSQ-3.6989D0*TDIMCPCO2 = 69.145D0-.70463D0*TDIM34-200.77DO/TDIMSQ+1^176.76D0/TDIM34CVU = (X02*CP02+XN2*CPN2+XCH4*CPCH4+XCH2*CPCH2+XH20*CPH20+1^XCO2*CPCO2)/WTMOL - RUC WRITE(9,55)CPCH4,CPCH2,CP02,CPN2,CPCO2,CPH20C WRITE(9,56)CVU13755 FORMAT(1X,'CP CH4,CH2, 02, N2, CO2, H2O =',6(1X,F8.4))56 FORMAT(1X,'CVU = ',F8.4)RETURN30 CONTINUEC ********Estimate the mean viscosity of gas mixturesTM = TU**0.645D0VISC = (X02 *MW02*5.09D0 + XN2*MWN2*4.57D0 +1^XCH4*MWCH4*3.35D0 + XCH2*MWCH2*1.33D0 +2^XH2O*MWH2O*3.26D0 + XCO2*MWCO2*3.71D0)3^/WTMOL*10.D0**( -7.D0)*TMRETURNENDc****************************************************************SUBROUTINE BURNED(P,TB,VU,VM,UU,UM,VB,XIVIB,II)c*****************************************************************************IIVIPLICTT REAL*8(A-H 2O-Z)REAL*8 D(3)COMMON/PROPS/RDG,RHC,FRES,EQVR,RUCOMMON/BURN/UBCIF(H .NE. 1) GO TO 10CALL TABLE(Q1,Q2,Q3,Q4,Q5,1)TBM1 = 1000.D0XMl = 0.D0XMB = 0RETURN10 CONTINUEC Find the linear relationship between UB and VB at the flame frontIF( UM .NE. UU)GO TO 210A = -1.DO/PB = VU - A*UUGO TO 220210 A=(VM-VU )/(UM-UU )B = VU - A*UUC SOLVE FOR T,V,U JUST BEHIND FLAME AND MASS FRACTION XC220 TB = TBM1 - 400.D0220 TB = 1000.D0230 D(1) = O.DOD(2) = 0.D0D(3) = 0.D0M = 1240 TB = TB + 100.D0250 CALL TABLE(TB,P,SB,UB,VB,3)Y = VB - ( A*UB + B)IF( Y .GT. 0.D0) GO TO 252D(1) = TBGO TO 253252 D(2) = TB253 D(3) = D(1)*D(2)IF(TB .GT. 6000.D0) WRITE(10,261)IF(TB .GT. 6000.D0) go to 300IF( D(3) .EQ. 0.D0) GO TO 240M= M+1TB = 0.5D0*( D(1) + D(2) )IF(DABS(Y) .LE. 0.1D-7 ) GO TO 300IF( M .LT. 50) GO TO 250WRITE(10,260)260 FORMAT(1X,'NO CONVERGENCE ON FLAME TEMPERATURE IN 50 TRIES')261 FORMAT(1X,'FLAME TEMPERATURE EXCEEDS 6000 K')RETURN300 CONTINUE138C CALCULATE MASS FRACTION XMB CORRESPONDING TO ASSUMED PRESSUREXMB = (VM - VU )/(VB - VU)TBM1 = TBXM1 = XMB321 RETURNENDc****************************************************************CSUBROUTINE TABLE(TBX,PBX,SBX,UBX,VBX,L)Cc****************************************************************C^L =1 INITIALIZATIONC L =2 GIVEN P AND SBC L =3 GIVEN P AND TBCIMPLICIT REAL*8(A-H 2O-Z)REAL*8 PTAB(10),TBTAB(10)REAL*8 SBTAB(10,10),UBTAB(10,10),AMTAB(10,10)REAL*8 SBTAB1(10,10),UBTAB1(10,10),AMTAB1(10,10)REAL*8 SBTAB2(10,10),UBTAB2(10,10),AMTAB2(10,10)REAL*8 SP(10),UP(10),AMP(10),X(10),Y(10)REAL*8 AMREAD(6),UBREAD(6),SBREAD(6),EQREAD(6)REAL*8 AMREAD1(6),UBREAD1(6),SBREAD1(6),EQREAD1(6)REAL*8 AMREAD2(6),UBREAD2(6),SBREAD2(6),EQREAD2(6)COMMON/PROPS/RDG,RHC,FRES,EQVR,RUCIF(L .NE. 1) GO TO 20PSTORE = -1000.D0DO 12 I = 1,10PTAB(I) = 0.D0SP(I) = O.DOUP(I) = 0.D0AMP(I) = 0.D0TBTAB(I) = 0.D0DO 11 J = 1,10SBTAB(I,J) = 0.D0UBTAB(I,J) = 0.D0AMTAB(I,J) = 0.D011 CONTINUE12 CONTINUEc****************************************************************************C READ TABLE OF BURNED GAS PROPERTIESC Reads files for different equivalence ratios; unit 11(0.2),C unit 12(0.4), unit 13(0.6), unitl4(0.8), unitl5(0.9),16(1)DO 120 J. = 1,10DO 110 I = 1,10DO 103 K = 1,6READ(K+10,108)TBTAB(I),PTAB(J),AMREAD1(K),VB,UBREAD1(K),1^H,SBREAD1(K)READ(K+20,108)TBTAB(I),PTAB(J),AMREAD2(K),VB,UBREAD2(K),1^H,SBREAD2(K)103 CONTINUEEQREAD(1) = 0.2D0EQREAD(2) = 0.4D0EQREAD(3) = 0.6D0EQREAD(4) = 0.8D0EQREAD(5) = 0.9D0EQREAD(6) = LODOC Set a burned-gas properties table for a given H to C ratio RHCC by linearly interpolate the CH4 and CH2 tables at the sameC equivalence ratio EQVR.139CALL CUBICS(6,EQREAD,AMREAD1,EQVR,AMTABla,n)CALL CUBICS(6,EQREAD,AMREAD2,EQVR,AMTAB2(I,J))AMTAB(I,J)=A/vITAB2(1,1)+(RHC-2)/2*(AMTAB1(1,J)-AMTAB2(1J))CALL CUBICS(6,EQREAD,UBREAD1 ,EQVR,UBTAB1 am)CALL CUBICS(6,EQREAD,UBREAD2,EQVR,UBTAB2(I,J))UBTAB(1,J)=UBTAB2(I,J)±(RHC-2)/2*(UBTAB1(1,J)-UBTAB2(1,J))CALL CUBICS(6,EQREAD,SBREADLEQVR,SBTAB1(1,J))CALL CUBICS(6,EQREAD,SBREAD2,EQVR,SBTAB2(I,J))SBTAB(1,J)=SBTAB2(1,1)+(RHC-2)2*(SBTAB1(1,J)-SBTAB2(1,J))C READ(1,108)TBTAB(I),PTAB(J),AMTAB(I,J),VB,UBTAB(I,J),H,C 1^SBTAB(I,J)108 FORMAT(1X,2(1X,F6.1),1X,F8.4,4(1X,D11.3))CPTAB(J) = PTAB(J)* 101.325D0UBTAB(I,J) = UBTAB(I,J)*0.001D0C WRITE(10,109)TBTAB(1),PTAB(J),AMTAB(1,1),UBTAB(1,J),SBTAB(I,j)109 FORMAT(1X,'TAB TB,P,AM,UB,SB=',5(1X,D10.4))110 CONTINUE120 CONTINUEPRINT*,'A table of proportional-model burned gas properties'PRINT'*,'^for a given EQVR is formed'RETURNc******************************************************************************C GIVEN P AND SB OR TB20 IF(PBX .EQ. PSTORE) GO TO 160PSTORE = PBXNP = 10XSET = PBXDO 131 I = 1,10DO 130 J. = 1,10X(J) = PTAB(J)130 Y(J) = UBTAB(I,J)CALL CUBICS(NP,X,Y,XSET,UP(I))131 CONTINUEDO 141 I = 1,10DO 140 7 = 1,10140 Y(J) = AMTAB(I,J)CALL CUBICS(NP,X,Y,XSET,AMP(I))141 CONTINUEDO 151 I = 1,10DO 150 J = 1,10150 Y(J) = SBTAB(I,J)CALL CUBICS(NP,X,Y,XSET,SP(I))151 CONTINUE160 CONTINUEIF( L .NE. 2) GO TO 300C**************************************XSET = SBXDO 230 I =1,10X(I) = SP(I)230 Y(I) = TBTAB(I)CALL CUBICS(NP,X,Y,XSET,TBX)DO 240 1 = 1,10240 Y(I) = UP(I)CALL CUBICS(NP,X,Y,XSET,UBX)DO 250 1 = 1,10250 Y(I) = AMP(I)CALL CUBICS(NP,X,Y,XSET,AMX)VBX = 8.3143D0/AMX*TBX/PBXRETURNC**************************************140300 XSET = TBXDO 330I = 1,10X(I) = TBTAB(I)330 Y(I) = SP(I)CALL CUBICS(NP,X,Y,XSET,SBX)DO 340 I = 1,10340 Y(I) = UP(I)CALL CUBICS(NP,X,Y,XSET,UBX)DO 350 I = 1,10350 y(I)= AMP(I)CALL CUBICS(NP,X,Y,XSET,AMX)VBX = 8.3143D0/AMX*TBX/PBXRETURNENDC*************************************************SUBROUTINE CUBICS(NP,X,Y,XSET,YCALC)C NP IS NUMBER OF X,Y DATA PAIRS (I RUNS FROM 1 TO N)IMPLICIT REAL*8(A-H2O-Z)REAL*8 X(10),Y(10),D(10),E(10),F(10),G(10)M = NP -1MM = NP -2C CALCULATION OF SECOND DERIVATIVES G(I)G(1) = 0.D0G(NP) = 0.D0DO 100 I = 2,MD(I) = X(1) - X(I-1)E(I) = 2.D0*( X(I+1) - X(I-1) )RD= x(1-F1) - X(I)100 Gm= 6.D0/FaNya-m-y(I))+6.Domor(Y(I-i) - Y(1))DO 1040 I = 2,MMFA = D(I+i)/E(1)E(I+1) = E(I+1) - FA*F(I)G(I+1) = G(I+1) - FA*G(1)1040 CONTINUEDO 1070 I = 2,MG(NP+1 -1)=(G(NP+1 -1)-F(NP+1-1)*G(NP+2-I))/E(NP+1 -I)1070 CONTINUEC CALCULATION OF INTERPOLATED VALUE YCALC AT X=XSETD(NP) = X(NP) - X(NP-1)I = 1200 1=1+ 1IF( XSET .GE. X(I) .AND. I .LT. NP) GO TO 200DELM = XSET - X(I-1)DELP = X(I) - XSETYCALC = G(I-1)/6.D0/D(I)*DELP**3 + G(1)/6.D0/D(1)*DELM**31^+(Y(I-1)/D(I) -G(1-1)*D(I)/6.D0)*DELP2^+(Y(I)/D(1) -G(1)*D(1)/6.D0 )*DELMRETURNENDc****************************************************************SUBROUTINE CYCSTATS(CA,P,X,L)IMPLICIT REAL*8(A-H 2O-Z)REAL*8 CA(360),P(360),X(360),CAPMAX(200)REAL*8 T10(200),T20(200),T30(200),T40(200),T50(200),PMAX(200)COMMON/STATS/N,NCYC,NCA,CABOI,RPMC^L = 1 : InitializeC^L = 2 : Compute times/secs from spark to 10%, 20%, 30%,40%,50%C of mass-fraction burned for each cycleC^L = 3 : Compute averages and std deviations for all cyclesIF( L .NE. 1) GO TO 10DO 5 I = 1,NCYC141T10(1) = 0.D0T20(I) = 0.D0T30(1) = 0.D0T40(1) = 0.D0T50(1) = 0.D05 PMAX(1)= 0.D0RETURN10 IF(L .NE. 2) GO TO 200CA10 = 0.130CA20 = 0.D0CA30 = O.DOCA40 = 0.D0CA50 = 0.D0DO 50 I = 2,NCACAI = CA(I)CAIM1 = CA(I-1)XI = X(1)XIM1 = X(I-1)DENOM = XI - XEM1IF(DENOM .EQ. 0.D0) GO TO 50DCADX = (CM - CAIM1)/DENOMIF(XI .LT. 0.1D0) GO TO 50IF(CA10 .NE. 0.D0) GO TO 20CA10 = CM - DCADX*(XI - 0.1D0) - CABOIGO TO 5020 IF(XI .LT. 0.21)0) GO TO 50IF(CA20 .NE. 0.D0) GO TO 30CA20 = CM - DCADX*(XI - 0.2D0) - CABOIGO TO 5030 IF(XI .LT. 0.3D0) GO TO 501F(CA30 .NE. O.DO) GO TO 40CA30 = CM - DCADX*(XI - 0.3130) - CABOIGO TO 5040 IF(XI .LT. 0.4D0) GO TO 50IF(CA40 .NE. 0.130) GO TO 45CA40 = CM - DCADX*(XI - 0.4D0) - CABOIGO TO 5045 IF(XI .LT. 0.5130) GO TO 50IF(CA50 .NE. 0.130) GO TO 50CA50 = CM - DCADX*(XI - 0.5D0) - CABOI50 CONTINUET10(N) = CA10/6.D0/RPMT20(N) = CA20/6.DO/RPMT30(N) = CA30/6.DO/RPMT40(N) = CA40/6.DO/RPMT50(N) = CA50/6.D0/RPMC^Determine Pmax for each cyclePMAX(N) = 0.130DO 175 I = 2,NCAIF(P(1) .GT. PMAX(N) ) CAPMAX(N)=CA(I)175 IF(P(I) .GT. PMAX(N) ) PMAX(N) = P(I)C^write time elapsed(sec) for 10%,20%,30%,40%,50% mass burnedWRITE(10,51)T10(N),120(N),T30(N),T40(N),T50(N),PMAX(N),CAPMAX(N)51 FORMAT(1X,'T10,T20,T30,T40,T50,Pmax,CApmax=',7(1X,D12.4))RETURN200 CONTINUEC^Determine averages for all cycles: 10%,20% etc burnedTlOME = 0.D0T2OME = 0.130T3OME = 0.130T4OME = 0.130142T5OME = 0.130DO 300 N = 2,NCYCTlOME = TlOME + T10(N)T2OME = T2OME + T20(N)T30ME = T30ME + T30(N)T4OME = T4OME + T40(N)T5OME = T5OME + T50(N)300 CONTINUETOTALN = FLOAT(NCYC-1)TlOME = T1OME/TOTALNT2OME = T20ME/TOTALNT30ME = 'T30ME/TOTALNT4OME = T40ME/TOTALNT5OME = T5OME/TOTALNC^Determine standard deviation of times to 10%,20% etc burnedSIG10 = 0.D0SIG20 = 0.D0SIG30 = 0.D0SIG40 = 0.D0SIG50 = 0.D0DO 400 N = 2,NCYCSIG10 = SIG10 + (T10(N) - TlOME)**2SIG20 = SIG20 + (T20(N) - T20ME)**2SIG30 = SIG30 + (T30(N) - T30ME)**2SIG40 = SIG40 + (T40(N) - T4OME)**2SIG50 = SIG50 + (T50(N) - T5OME)**2400 CONTINUESIGIO = DSQRT(SIG10/TOTALN)SIG20 = DSQRT(SIG20/TOTALN)SIG30 = DSQRT(SIG30/TOTALN)SIG40 = DSQRT(SIG40/TOTALN)SIG50 = DSQRT(SIG50/TOTALN)C^Determine normalized covariances for 10%,20% etc burnedCR1020 = 0.D0CR1030 = 0.130CR1040 = 0.130CR1050 = 0.D0CR2030 = 0.D0CR2040 = 0.130CR2050 = 0.130C^Determine Std Dev of 10% to 50% burning timesSIG5M1 = 0.130DO 500 N = 2,NCYCCR1020 = CR1020 + (T10(N)-T10ME)*(T20(N)-T20ME)CR1030 = CR1030 + (T10(N)-T10ME)*(T30(N)-T30ME)CR1040 = CR1040 + (T10(N)-T10ME)*(T40(N)-T40ME)CR1050 = CR1050 + (T10(N)-T10ME)*(T50(N)-T50ME)CR2030 = CR2030 + (T20(N)-T20ME)*(T30(N)-T30ME)CR2040 = CR2040 + (T20(N)-T20ME)*(T40(N)-T40ME)CR2050 = CR2050 + (T20(N)-T20ME)*(T50(N)-T50ME)SIG5M1 = SIG5M1 + (T50(N) - T10(N))**2500 CONTINUECR1020 = CR1020/SIG10/SIG20/TOTALNCR1030 = CR1030/SIGIO/SIG30/TOTALNCR1040 = CR1040/SIG10/SIG40/TOTALNCR1050 = CR1050/SIGIO/SIG50/TOTALNCR2030 = CR2030/SIG20/SIG30/TOTALNCR2040 = CR2040/SIG20/SIG40/TOTALNCR2050 = CR2050/SIG20/SIG50/TOTALNSIG5M1 = DSQRT(SIG5MI/TOTALN)143WRITE(10,71)WRITE(10,711) NCYCVVRITE(10,72)WRITE(10,73)T1OME,T20ME,T3OME,T40ME,T5OMEWRITE(10,74)SIGIO,SIG20,SIG30,SIG40,SIG50WRITE(10,76)WRITE(10,77)CR1020WRITE(10,78)CR1030WRITE(10,79)CR1040WRITE(10,80)CR1050WRITE(10,81)CR2030WRITE(10,82)CR2040WRITE(10,83)CR2050WRITE(10,831)SIG5M1CA10 = T1 OME*6.DO*RPMCA20 = T2OME*6.D0*RPMCA50 = T5OME*6.D0*RPMWRITE(10,84)CA10,CA20,CA5071 FORMAT(1X,'Statistics of Time from Spark to Given X')711 FORMAT(1X,I4,' Cycles from Data Set^ ')72 FORMAT(1X,'X = ',18X,'10%',9X;20%',9X,'30%',9X;40%',9X,'50%')73 FORMAT(1X,'Mean Time (sec)',3X,5(2X,D10.4))74 FORMAT(1X,'Std Dev (sec)',5X,5(2x,D10.4))76 FORMAT(1X,'Cross Correlation Factors')77 FORMAT(1X,'10% - 20% Times',F10.3)78 FORMAT(1X,'10% - 30% Times',F10.3)79 FORMAT(1X,'10% - 40% Times',F10.3)80 FORMAT(1X,'10% - 50% Times',F10.3)81 FORMAT(1X,'20% - 30% Times',F10.3)82 FORMAT(1X,'20% - 40% Times',F10.3)83 FORMAT(1X,'20% - 50% Times',F10.3)831 FORMAT(1X,'Std Dev of 10% to 50% burning time =',F10.3)84 FORMAT(1X,'Crank Angles to X3.1,0.2,0.5=',3(1x,f10.2))C^Determine mean value of PmaxPMAXME = 0.D0DO 600 N = 2,NCYC600 PMAXME = PMAXME + PMAX(N)/TOTALNC.....Determine standard deviation of PmaxSIGP = 0.D0DO 700 N = 2,NCYC700 SIGP = SIGP + (PMAX(N) - PMAXME)**2SIGP = DSQRT(SIGP/TOTALN)SPREL = SIGP/PMAXMEWRITE(10,85)PMAXME,SIGP,SPREL85 FORMAT(///,1X,'Mean Pmax= ',D12.6,3x,'Std Dev Pmax= ',D12.6,1 3x,'Relative Std Dev Pmax= ',D12.6)C^Sort in order of 10% times and write times for all cyclesNSORT = NCYCCALL PIKSR2(NSORT,T10,Pmax)DO 1100 I = 2,NCYCWRITE(10,56)I,T10(l),Pmax(I)56 FORMAT(1X,'Ncycle =',I4,3X,'10% time/s = ',D12.6,1 ' Pmax/lcPa = ',D12.6)1100 CONTINUE144C^Sort in order of Pmax and print for all cyclesNSORT = NCYCCALL PIKSR2(NSORT,Pmax,T10)DO 1200 I = 2,NCYCWRITE(10,87) I,Pmax(I),T10(I)87 FORMAT(1X,'Ncycle = ',14,2X,'Pmax = ',D12.6,1 ' 10% time/s = ',D12.6)1200 CONTINUEC^Determine normalized covariance for 10% time and PmaxCRT1Pm = O.DODO 1300 I = 2,NCYCCRT1Pm = CRT1Pm + (T10(l)-T10ME)*(Pmax(I)-Pmaxme)1300 CONTINUECRT1Pm = CRT1Pm/SIG10/SigP/TOTALNWRITE(15,88)CRT1Pm88 FORMAT(1X,'Normalized Covariance for 10% burning time1 and maximum pressure = ',D10.4)RETURNENDC*************************************************************SUBROUTINE PIKSR2(N,ARR,BRR)C*************************************************************C Sorts an array ARR of length N into ascending numerical orderC by straight insertion. N is input; ARR is replaced on outputC making the corresponding rearrangement of array BRRCREAL*8 ARR(N),BRR(N)DO 12 J = 2,NA = ARR(J)B = BRR(J)DO 11 I = J-1,1,-1IF(ARR(I) .LE. A) GO TO 10ARR(I+1)=ARR(I)BRRa+1 BRR(I)11^CONTINUEI = 010^ARR(1+1) = ABRR(I+1) = B12^CONTINUERETURNENDCC********************************************************DOUBLE PRECISION FUNCTION ACYL(CA)C Calculates the cylinder surface area for a given degree CAIMPLICIT REAL*8 (A-H 2O-Z)COM.MON/GEOM/BORE,STROKE,ROD,CLRHCPI = 3.14159D0APSTON = PI/4.D0*BORE**2CAR = CA*PI/180.D0Z = ( LDO + 2.D0*ROD/STROKE + DCOS(CAR)1 -DSQRTa2.D0*ROD/STROKE)**2+(DSIN(CAR))**2))*STROKE/2.D02 + CLRHACYL = Z*PI*BORE + 2.D0*APSTONRETURNENDC*********************************************************SUBROUTINE QWALL(TM1,V2,XMI3,ASLTRF,DQWL,L)C*********************************************************C Calculates the heat transfer from the gas to the cylinder wall145C using Annand's and Woschni's correlation.CIMPLICIT REAL*8 (A-H 2O-Z)REAL*8 MTOTCOMMON/GEOM/BORE,STROKE,ROD,CLRHCOMMON/MASS/MTOT,MAIR,MDSLCOMMON/STATSNNCYC,NCA,CABOLRPMCIF (L .NE. 1) GO TO 20DQWL = 0RETURN20 CONTINUEAAA = 0.38D0BBB = 0.75CCC = 1.6E-12PISVEL = RPM * STROKE / 30.0DENS = MTOT / V2CALL UNBURNED(TM1,UU,CVU,VISC,4)RENUM = DENS * PISVEL * BORE / VISCCALL UNBURNED(TM1,UU,CVU,VISC,3)TRMLCO = CPG * VISC / 0.7D0C^The wall temperature is assumed to be constantTW = 450.0D0QCONV = ASURF * AAA * TRMLCO / BORE * RENUM**(BBB) * (TM1-TW)QRAD = ASURF * CCC * ( TM1**4 - TW**4 )PRINT*,'QCONV,QRAD=',QCONV,QRADDQWL = (QCONV+QRAD)DQWL = DQWL * (60./RPM/360.)RETURNENDCC*****************************************146147Appendix 6.5VERIFICATION OF THE COMPUTATION PROCEDUREFOR THE CONSTANT VOLUME CASE1. Run the program for an arbitrary pressure data.Obtain the unburned gas properties at the beginning of the first calculating stepwhich become the initial condition of the combustion process.The composition of the unburned gas is found to be:Constituent Mole fractionCH4 0.0436CH2 0.013302 0.1845N2 0.7345CO2 0.0087H2O 0.0154(1.0000)Its specific volume, v. = 0.182067 m3/kg ;specific energy,^u„ = - 275.282 kJ/kg.molecular weight, M. = 28.028 kg/kmol.2. The above unburned gas mole fractions give the atom relative populations of C ,H , 0 , N : 0.0656, 0.2318, 0.4018, 1.469, respectively.1483. Run STANJAN using the above initial condition, i.e. atom relative populations,specific volume v. , and specific energy u. , to determine the pressure P.,„, andtemperature T^of the mixture at the end of combustion.It was found that :^= 3328.3 kPa, T, , = 2055.68 K.While the burned gas molecular weight, Mb = 28.205 kg/kmol; which is 0.6%higher than that of the unburned gas.4. In the mass-burned fraction program, consider the first step calculation as a constantvolume combustion (V2 = V2) and specify the pressure at the end of the step P2 =Pte .Ideally, the output of the program will show that the burned gas temperature isequal to that of STANJAN, Tb = T, , . Moreover, the program result shoulddemonsrates that all the fuel is burned at the end of the process, x = 1.0.5.^Output of the program are :(i) Burned Temperature, Tb = 2062.87 K. This result is within 0.35% compareto the one we get from STANJAN.(ii) Mass-burned fraction, x = 0.995260, which is within 0.47% to the idealcase.


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