HCCI performance and combustion

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SAE TECHNICALPAPER SERIES 2000-01-0332

A Four-Stroke Homogeneous ChargeCompression Ignition Engine Simulation for

Combustion and Performance Studies

Scott B. Fiveland and Dennis N. Assanis

W. E. Lay Automotive Laboratory, University of Michigan

Reprinted From: Compression Ignition Combustion Processes(SP1530)

SAE 2000 World CongressDetroit, Michigan

March 6-9, 2000

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2000-01-0332

A Four-Stroke Homogeneous Charge Compression Ignition

Engine Simulation for Combustion and Performance Studies

Scott B. Fiveland and Dennis N. AssanisW. E. Lay Automotive Laboratory, University of Michigan

Copyright 2000 Society of Automotive Engineers, Inc.

ABSTRACT

A computer simulation of the Homogenous ChargeCompression Ignition (HCCI) four-stroke engine hasbeen developed for combustion and performance studies.The simulation couples models for mass, species, and

energy within a zero-dimensional framework. Thecombustion process is described via a user-definedchemical kinetic mechanism. The CHEMKIN librarieshave been used to formulate a stiff chemical kineticsolver suitable for integration within a complete enginecycle simulation, featuring models of gas exchange,turbulence and wall heat transfer. For illustration, twochemical kinetics schemes describing hydrogen andnatural gas chemistry have been implemented in thecode. The hydrogen scheme is a reduced one,consisting of 11 species and 23 reactions. The naturalgas chemistry is described via the GRI-mechanism 3.0that considers 53 species and 325 reactions, including

NOx chemistry. Computations are first carried out in avariable volume bomb to demonstrate variations inignition with temperature, pressure, equivalence ratio,and composition. Subsequently, the complete cyclesimulation is exercised to demonstrate the variation inoutput parameters to charge inlet temperature andeffective compression ratio. Overall, this studydemonstrates the importance of coupling detailedchemistry descriptions with physical models of the HCCIengine processes.

INTRODUCTION

Homogeneous Charge Compression Ignition (HCCI) iscurrently under widespread investigation due to itspotential to lower NOx and particulate emissions whilemaintaining high thermal efficiency [1, 2, 3, 4, 5].Throughout the years it has endured many names in theliterature: ATAC (Active Thermo-AtmosphereCombustion), LHC (Lean Homogeneous Combustion),CIHC (Compression Ignited Homogeneous ChargeCombustion), AR (Active Radical Combustion), HCDC(Homogeneous Charge Compression Ignition DieselCombustion, Diesel Fumigation, MULDIC (MultipleStaged Diesel Combustion, PREDIC (Premixed Direct-

Injection Combustion), and PCIC (PremixedCompression Ignited Combustion). Engine types haveranged from two-stroke to four stroke configurations witha variety of fuels such as diesel, gasoline, methanolnatural gas, and hydrogen. The HCCI process essentiallyinvolves a premixed fuel/air mixture that is inducted into

the cylinder at equivalence ratios that can vary from leanto stoichiometric [6, 7]. Once within the cylinder, thehomogeneous fuel/air charge is then compressed untiignition commences. Ignition leads to a very rapidcombustion phase where all heat is releasedapproximately in 10-35.

The premixed charge, compression ignition HCCI engineconcept promises to combine the advantages of both theDirect Injection, Compression Ignition (DICI) engine andthe premixed charge, Spark-Ignited (SI) engine, whileeliminating their drawbacks. One benefit that can begained over the current heterogeneous DICI engines isthe elimination of the fuel rich zones that are directlyresponsible for pollutant formation, especiallyparticulates [7]. In addition, the homogeneous lean burnoperation will yield lower gas temperatures and henceNOx, as compared to both SI and DICI counterpartsFurthermore, unthrottled part load operation eliminatespumping losses leading to improved fuel economy ovethe SI engine [1]. In addition, compression ignitioneliminates SI knock associated with autoignition of fuelair mixture in the end zone, ahead of the advancing flamefront. As the HCCI concept operates on the premise oautoignition, this allows the use of elevated compressionratios (approximately 20-25:1), unlike in SI engines. The

combination of lean burning, which is thermodynamicallyattractive and reduces heat transfer losses, and highcompression and thus expansion ratios contributes toindicated thermal efficiencies that can approach 55%Residual gas fraction is also reduced with HCCoperation, thus improving volumetric efficiencycombustion and performance [5]. Overall, HCCI promisesto deliver both high thermal efficiencies and reducedemissions.

Although recent investigations into HCCI combustionappear promising [4, 5, 6, 7, 8, 9, 10, 11, 12] severaproblems with the HCCI combustion concept reappea





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throughout the literature. Most stem from the fact that theHCCI concept gives up two combustion control aspects.First, the timing of ignition is not controlled, neitherindirectly by fuel injection as in a DICI engine, nor directlyby the spark as in an SI engine. Second, the rate of heatrelease is also not controlled neither by the rate of fuelinjection as in a DICI engine, nor by finite turbulent flamepropagation as in an SI engine. As a result, the nearconstant volume combustion event leads to a very rapidrate of heat release, thus promoting high mechanicalstresses [8]. In particular, controlling the ignition event athigher loads is a widely noted problem.

In an effort to understand how mixture preparation andin-cylinder thermodynamic conditions affect the chemicalkinetics, models of varying resolution have beendeveloped [4, 7, 9, 10, 11, 12]. These modeling effortsprovide a good basis for exploring the HCCI combustionphenomena. However, in published zero-dimensionalsimulations, detailed chemistry has typically beencoupled to models of turbulence, heat transfer, and gasexchange that are not predictive over the full enginecycle. Heat transfer models have been based on data

not relevant to HCCI operation, and in many cases thegas exchange process has been neglected by specifyingconditions at a specific point in the engine cycle. Forinstance, Smith et al. [7], Van Blarigan and Goldborough[5], as well as Kusaka et. al. [12] used the Woshni heattransfer correlation [13] to predict heat transfer. TheWoschni model [13] was developed from a regression ofdirect injection diesel engine data and is essentially alinearized radiative/convective relation that will greatlyover-predict heat transfer in a lean burn, premixed, non-sooting engine. Furthermore, studies by Poulos andHeywood [14] as well as Assanis and Heywood [15] have

shown that, since turbulence intensity has a large cyclicvariation, heat transfer correlations based on multiples ofmean piston speed do not provide a truly predictivecapability. In addition, the computational work of Smith et.al. [7] was only carried out over the closed part of thecycle, starting with bottom dead center of the intakestroke, with no gas exchange. Studies of closed cycleprocesses are useful for evaluating chemistry, but do notcapture intake and exhaust jet flows. Consequently,velocity and length scales of in-cylinder flow structuresare not accurately represented; thus, wall heat transferlosses and pre-ignition boundary conditions can beinaccurate.

The objective of this work is to develop a full cyclesimulation model of the HCCI engine that would integratecomplex chemistry with physical models of the in-cylinderprocesses. Emphasis will be placed on integratingflexible chemical kinetic libraries with models ofturbulence-based heat transfer and gas exchangeprocesses for a four-stroke cycle. The paper is arrangedas follows. The HCCI model formulation will be presentedfirst. This will include a brief development of thegoverning equations and various engine submodels (i.e.gas exchange, combustion, heat transfer) as they apply

to a system with well understood chemistry. Thesensitivity of the combustion submodel to changes inpressure, temperature, equivalence ratio, and natural gascomposition will be demonstrated first in a fundamentaladiabatic reactor. The behavior of the each of thesubmodels will then be detailed over the engine cycleand contrasted with other treatments of cyclic processesreported in the literature. Subsequently, parametricstudies of ignition and engine performance for varyingmanifold temperature, pressure, and geometriccompression are performed for both hydrogen andnatural gas using the full cycle simulation.

MODELING ASSUMPTIONS

The compression ignition engine simulation of Assanisand Heywood [15] has provided a solid zero-dimensionaframework for formulation and implementation of thegoverning equations for the HCCI model. The simulationis currently written in a single cylinder version, primarilybecause fundamental studies lend themselves to thisconfiguration. Thermodynamic properties are assumed

uniform throughout the chamber volume. The enginesimulation is a sequence of four-stroke processes. Thegas exchange process is governed by quasi-steady, one-dimensional flow equations that are used to predict flowpast valves. The compression event is defined fromIntake Valve Closing (IVC) to a transition point prescribedwhen chemical reactions become important. Thecombustion event for the HCCI simulation differs fromthose of the SI and DICI types. As a result of its premixednature and compression ignition principle, the rate ocombustion is strictly limited by the chemical kinetics. Inthe reaction regime, the combustion will follow that ofinite rate kinetics and its heat release will be governed

by detailed chemistry. Hence, the combustion event hasbeen modeled using the Chemkin libraries [16], adaptedfor a variable volume plenum and accounting for heattransfer effects. The evolution of heat release andspecies is governed by a user-defined kinetic schemeLater in expansion, the reacting flow mixture will attainchemical equilibrium in response to the changing incylinder conditions, and eventually the composition wilbecome frozen. At this point the mixture is again nonreacting and the chemical energy source term tends tozero.

CONSERVATION EQUATIONS

To define a thermodynamic state within the cylinder, twoindependent properties and the mixture compositionneed to be known. Consequently, our formulation wiltrack the evolution of mass, species, temperature, andpressure throughout the engine cycle. The generaequations governing mass, species, and energy will bedeveloped for a variable volume reactor, as shown inFigure 1. It is assumed that the working fluid behaves asan ideal gas. The heat transfer is governed by turbulenpipe flow, but derives its characteristic velocity as a





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function of mean piston speed, mean velocity, andturbulent intensity, which is found from a zero-dimensional k- formulation.

Figure 1. Energy Transfers Associated with CylinderControl Volume.

CONSERVATION OF MASS The rate of change ofmass within any open system is the net flux of massacross the system boundaries.

(1)

CONSERVATION OF SPECIES Equations tracking theevolution of species within the combustion chamber willbe developed on a mass basis corresponding to thedefinition in Eq. (2),

(2)

where m denotes the total mass within the controlcylinder. The species equations are deduced from theirmulti-dimensional counterparts by neglecting speciesdiffusion terms, consistent with the zero-dimensionalassumption.

(3)Expanding Eq. (3), and applying the continuity equationyields,

(4)

which is the final form of the species conservationequation.

CONSERVATION OF ENERGY The generalizedenergy equation for an open thermodynamic system maybe written as:

(5

Rewriting the first law equation in terms of rate of changeof enthalpy yields,

(6

At this point in the development, an expression needs tobe developed for that relates it to the change in mixturetemperature. Assuming a single phase, multi-componenmixture of ideal gases, its enthalpy, h, is defined as

(7

and (8

where the subcript j refers to each of the componentspecies present in the mixture. Taking partial differentialswith respect to pressure, temperature and composition:

(9

where the partial change of enthalpy with respect topressure at constant temperature and composition iszero, but pressure effects are accounted for in

determining changes in composition with allowance fordissociation. Hence, Eq. (9) becomes

(10

To relate the pressure gradient to the temperaturechange, the equation of state is used in its differentiaform. By manipulating the thermodynamic equation ostate for an ideal gas, we get:

(11

An expression can now be inserted for the rate of changeof the gas constant R by considering its thermodynamicdependence, i.e.:

(12

Taking partial differentials with respect to pressuretemperature and composition,

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(13)

The first term on the R.H.S. of Eq. (13) can berecognized as the partial expansion of the gas constantfor a pure substance. Under the problem assumptions,this term can be shown to be identically equal to zero,

which is consistent with references [16] and [17].

(14)

Using Eqs. (11) and (14), we can obtain an equation thatrelates the rate of change of pressure to the rate ofchange of temperature, i.e.:

(15)

Susbstituing Eq. (15) into Eq. (6) yields an equation forthe rate of change of temperature for an ideal gas,reacting mixture:

(16)

where and (17)

MODELING OF RECIPROCATOR PROCESSES

GAS EXCHANGE The one-dimensional quasi-steadyflow model is used to model flow through both the intakeand exhaust valves during the gas exchange processes.Equation (18) is a function of discharge coefficient, valvearea, gas properties, and pressure differential acrosseach orifice. The values for both valve lift and dischargecoefficients can be specified or predicted. The values forcylinder pressure are updated by solving the system ofstate differential equations through the cycle [18].

(18)

COMBUSTION The combustion process in ahomogenous charge compression ignition engineexhibits no fundamental mixing or entrainment, whichnormally control the combustion event for direct-injectionas well as spark ignition engines. As a result the rate ofheat release is solely driven by the chemical kineticreaction rates. To appropriately model the combustion

event, it was necessary to describe the evolution of heatrelease via a suitable chemical kinetic mechanism, and awell-matched, stiff chemical kinetic solver. CHEMKINwas selected due to the fact that it is a widely acceptedand used kinetic solver by many researchers in a rangeof combustion studies [10, 19]. Simulation developmenhas been done in Fortran 90 within the VisuaEnvironment on a PC 333 Mhz Pentium II. A sourceprogram for a variable volume reactor was written to drivethe CHEMKIN libraries. The driver model includes theeffects of heat transfer, as well as options for differenengine configurations. The program can also operate inan adiabatic standalone mode for fundamental studies.

The rates of creation/destruction of chemical species aremodeled using mass-action kinetics, where the specificreaction rate constants exhibit a strong temperaturedependence. An elementary reaction that involves Kchemical species in I reactions can be represented inthe form,

(19

(20

(21

The specific reaction rate constant, k follow theArrhenius dependence where,

(22

THERMODYNAMIC PROPERTY TREATMENT Thefluid in the cylinder is constantly undergoing a change inmixture composition. Once the composition isdetermined, the partial mixture properties can besummed and appropriately weighted in accordance withtheir mass or mole fractions. This method very simplyallows thermodynamic property calculation for mixturescontaining residual gas, exhaust gas recirculationunburned gaseous fuel, air etc. The thermodynamicproperty treatment will employ the NASA curve-fits fo

specific heat, enthalpy, and entropy:

(23

where the constant of integration is the enthalpyof formation at 0 K.

(24

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7/195

(25)

HEAT TRANSFER MODEL The heat transfer modelthat is utilized was described by Assanis and Heywood[15]. This model is a subset of the k- model under theisotropic turbulence assumption. It uses an energycascade between the mean kinetic energy and the

turbulent kinetic energy. The eddy dissipation is modeledvia dimensional argument. This model accounts for theeffects of mean piston speed, mean charge motion, andturbulence intensity on the heat transfer coefficient.

The definitions for both mean and turbulent kinetic energyare given by Eq. (26).

(26)

The equations governing the cascade model are shownin Eqs. (27)-(29). It is noted that these equations take onthe classical form for transport phenomena, having anunsteady term, a convective term, and a source/sinkterm. The mean flow kinetic energy equation describesthe history of the initial kinetic energy. Over time, thismean kinetic energy is dissipated to large scaleturbulence, whose evolution is described by the turbulentkinetic energy equation. Finally the eddy dissipation,which occurs on a molecular level, dissipates theturbulent kinetic energy. It is modeled via dimensionalconsiderations.

(27)

(28)

(29)

The turbulence production is modeled based on

; (30)

where C=0.09(Universal) and C is an adjustableconstant.

Large scale eddies can be represented by a geometriclength scale defined as

(31

Rewriting Eq. (30) yields

(32

Rapid distortion theory (RDT) is employed to account forapid changes in density. This amplification source termwas added to the transport equation for turbulent kineticenergy

(33

Since RDT assumes negligible dissipation of theturbulent kinetic energy, conservation of mass andangular momentum of each eddy is assumed [15, 20].

(34

(35

(36

The rate of change of density is computed at eachtimestep. From the above definitions, the characteristic

velocity is computed as

(37

The heat transfer coefficient can be computed fromcorrelations of Nusselt number as a function of Reynoldsnumber. The interested reader can refer to Assanis andHeywood [15] for details.

METHOD OF SOLUTION

The mathematical model results in a set of ordinarydifferential equations. These equations must beintegrated simultaneously. Due to the incorporation ofchemical kinetics, two different integrators are used toensure computational efficiency. When the chemicareactions are frozen, the set of equations is integratedusing the standardized, non-stiff, predictor-correctotechnique, ODERT [21], based on a modified form of theAdams Pece method. When pre-combustion reactionsbegin to occur late in the compression stroke, the

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8/196

program switches integrators. A "stiff" ODE integrator,DVODE, is then used to deal with the kinetic equationsthat involve a wide range of reaction time scales, until thespecies concentrations are no longer changing. Aflowchart showing the structure of the engine simulationis shown in Figure 2.

After initializing the state variables, the program proceedsthrough each engine process. The main programcontrols the solution process. During each timestep, the

main program calls the pertinent integrator, which thencalls one of the four main process routines. Each ofthese process routines determines what equations areimportant and integrate them over each timestep. Thecomputer model has been implemented in the FORTRAN77/90 computer language.

BEHAVIOR OF SIMULATION SUBMODELS

COMBUSTION The combustion process in the HCCIengine is driven by chemical kinetics. For demonstrationpurposes both hydrogen and natural gas were used asthe fuels within the full single cycle simulation. Manyresearchers consider hydrogen as the ultimate fuel of thefuture [22]. It has been under study for hybrid applicationsdue to its high energy content and low emissionspotential. In addition, hydrogen exhibits unique fuelcharacteristics, which include high flame speed,ignitibility of lean mixtures, and high effective octane [23].Furthermore, its well understood kinetic scheme, whichinvolves only a moderate number of reactions, makes it agood candidate for modeling studies that can be used for

demonstration purposes. Likewise, natural gas, whichcontains a composition on the order of 85-90% methane2-10% ethane, and 1-4% propane has received attentionin recent years due to its "clean burn" potential. Itschemical kinetic mechanism is reasonably welcharacterized making it a good candidate fofundamental studies that can lead to a working engineconcept. To demonstrate the flexibility of the simulationto accept alternative kinetic schemes, the hydrogenchemistry was described with a reduced schemeavailable in CHEMKIN that considers 11 species and 25reactions [16]. For natural gas, the GRI-Mechanism 3.0was input into CHEMKIN; the detailed mechanismconsiders upwards of 53 species and 325 reactions [24].

Combustion in a Variable Volume Plenum Before usingthe mechanism in the full simulation, it was studied in afundamental variable volume, adiabatic reactor to verifythat it reproduced reported ignition trends for variations ininlet temperature, inlet pressure, and the addition ofhigher order hydrocarbons. Correctly quantifying theonset of ignition is critical to developing a predictive

simulation since it has a large effect on engine outputFor our study, the time of occurrence of the peakconcentration of either OH or HO2 was used as theignition criterion. This is consistent with Glassman [25who notes that that the presence of H, O, OH, and HO2abstract H radicals from methane and hence promotechain propagation. Note however, that other studiesdefine the onset of ignition based on the occurrence othe maximum pressure gradient [25]. Even though

Figure 2. Logic Structure of the Thermo-kinetic HCCI Engine Simulation

Start

Initialize

Read Input

Main Program

IVO --- IVC

IVC --- Ignition

Ignition -- EVO

EVO -- IVO

Integrate Intake 1.Engine Geometry

2. Valve Data

3.Flow Rates

4. Heat Transfer

5.Hydro-Carbon Formation

6. Engine Performance

7. Thermodynamic Properties

8. Transport Properties

CombustionConverge

Finite RateKinetics

SpecifiedROHR

Y NWrite Output

End

YN

Integrate

Integrate

Integrate

Integrate

Compression

Combustion

Expansion

Exhaust





9/197

mixture conditions throughout the compression processchange until the onset of ignition occurs, a usefulquantitative indicator of induction delay (sec) can bedefined as the time that lapses between Bottom DeadCenter (BDC, start of compression) and the onset ofignition. Figure 3 examines the correlation betweeninduction delays based on maximum pressure gradientand peak OH or HO2 concentrations for methanecombustion in a variable volume combustor over a rangeof inlet temperatures and pressures. The compressionratio is 15 and the speed is 1500 rpm. Remarkably, thecorrelation coefficient between the two induction delays is1, suggesting that both ignition criteria are consistent forthe case of HCCI combustion.

Having defined the ignition criterion, induction delaystudies were carried out as the initial conditions of themixture in the variable volume combustor were varied.The first case study varied the initial (BDC) temperatureof a methane-air mixture from 400 K to 800 K. Theequivalence ratio was maintained at 0.5, and the initialmixture pressure at BDC was 1.5 bar. The resultingpressure traces are shown in Fig. 4. As anticipated, the

induction delay decreased with increasing temperature(see Fig. 5). It was further noted that methane did notignite for an intake manifold temperature of 400 K,consistent with previous studies [eg., 10].

Next the initial pressure at BDC was varied from 1 bar to3 bar, while keeping the temperature fixed at 500 K, thecompression ratio fixed at 15, and the mixturecomposition as above. As shown in Fig. 6, ignition time isadvanced with higher inlet boost. This is a result of theincreased trapped air and fuel concentrations at thesehigher inlet pressures generating more free radicals thatin turn accelerate the chemical reactions.

Composition of natural gas can vary widely from state tostate, as well as slightly from cycle to cycle [26]. As aresult many fundamental studies have been carried out tounderstand how ignition behaves under varyingcomposition [9, 10, 27, 28]. Compositions were chosen ina range consistent with Westbrook and Pitz [27]. Amethane-air mixture at 1.5 bar and a temperature of 500K was used as the baseline case. Ethane and thenfinally propane was gradually added to the mixture inincreased quantities. Computed induction delays as afunction of composition are shown in Fig. 7. Consistentwith Westbrook and Pitz [27], Fraser et. al. [28], Naber et.

al. [9], and Agarwal and Assanis [10], increasedconcentrations of ethane and propane reduced inductiondelays, which is a result of increased radical poolspromoting chain branching reactions.

Figure 3. Correlation between induction delays basedon maximum pressure gradient and peak OHor HO2 concentrations for methanecombustion.

Figure 4. Effect of initial mixture temperature on ignitionand pressure profile in a variable volumechamber. Methane combustion withequivalence ratio of 0.5 and initial mixturepressure of 1.5 bar.

0.015

0.016

0.017

0.018

0.019

0.02

0.021

0.022

0.015 0.016 0.017 0.018 0.019 0.02 0.021 0.022

InductionDelayBasedondp/dt(Sec)

Induction Delay based on OH or Ho2

(Sec)

Correlation Coefficient (R) = 1.

0

2 107

4 107

6 107

8 107

1 108

1.2 108

0.012 0.014 0.016 0.018 0.02 0.022 0.024 0.026

400 K

450 K

500 K

600 K

800 K

Pressure(Pa)

Time (Sec)





10/198

Figure 5. Correlation of induction delay as a function ofinitial mixture temperature. Methanecombustion in a variable volume chamber

with equivalence ratio of 0.5 and initialmixture pressure of 1.5 bar.

Figure 6. Variation of onset of ignition with changinginitial pressure of mixture. Methanecombustion in a variable volume chamber

with equivalence ratio of 0.5 and initialmixture temperature of 500 K.

Combustion in Engine Simulation The behavior of thecombustion submodel was further explored within a fullengine cycle simulation. Primary engine specificationsare summarized in Table 1. Unless otherwise noted, forthe engine combustion study reported in this section, aswell as for subsequent studies illustrating the behavior ofother submodels: (i) hydrogen was used as the fuel, (ii)

the engine geometric compression ratio was fixed at 15(iii) intake manifold pressure was 1.5 bar, intake manifoldtemperature was 425 K, mixture equivalence ratio was0.3. and (iv) engine speed was 1500 rpm.

Figure 7. Variation in induction delay under differentcompositions of methane, ethane andpropane. Natural gas combustion in avariable volume chamber with initial mixtureconditions of 1.5 bar and 500 K, and anequivalence ratio of 0.5.

A pressure-volume diagram, over the full engine cycle, isshown in Fig. 8a. Clearly, the combustion is very close toconstant volume. This rapid rate of heat release is inagreement with the work of Van Blarigan andGoldsborough [5]. The pressure and temperature historyis shown in Fig. 8b. As will be shown in subsequensections, the rapid variation of ignition as well as rapidheat release greatly effect engine performance. Thespecies formed during hydrogen combustion are trackedas a function of time, and are shown in Fig. 9. Peaks inH, OH, and HO2 can be discerned near the ignition pointNote that NO emissions are near zero, a result of leancombustion.

0.015

0.016

0.017

0.018

0.019

0.02

0.021

0.022

0.0012 0.0014 0.0016 0.0018 0.002 0.0022 0.0024

InductionDelay(Sec)

(Temperature)-1

1/K

0

5 107

1 10

8

1.5 108

2 108

2.5 108

0.01 0.015 0.02 0.025 0.03

1 BAR

1.5 BAR

2 BAR

3 BAR

Pressure(Pa)

Time (Sec)

0.0187

0.0188

0.0189

0.019

0.0191

0.0192

0.0193

0 1 2 3 4 5 6

InductionDelay(Sec)

Test Case

Baseline: 1.0 CH4

0.97 CH4

+ 0.03 C2H

6

0.95 CH4

+ 0.05 C2H

6

0.90 CH4

+ 0.10 C2H

6

0.80 CH4

+ 0.10 C2H

6+0.1 C

3H

8

Table 1 Engine Specifications

Engine Type 4-Stroke

Displacement 12.7 Liter

Bore 13cm

Stroke 16cmCon. Rod Length 26.93cm

Intake Valve Opening 685 deg *

Intake Valve Closing 220 deg *

Exhaust Valve Opening 485 deg *

Exhaust Valve Closing 795 deg *

* Relative to TDC Intake





11/199

Figure 8. Cyclic characteristics of a four-cycle enginerunning on hydrogen at an inlet manifoldtemperature of 425K, an inlet pressure of 1.5bar, and an equivalence ratio of 0.3. (a)pressure-volume diagram; (b) variation of gaspressure and temperature during combustion.

GAS EXCHANGE Profiles detailing the rate of massflow through the engine intake/exhaust valves, computedusing the one-dimensional, quasi-steady, compressibleflow model, are shown in Fig. 10a. It is noted that theintake valve mass flow rate follows the path of the pistonmotion, in agreement with Assanis and Heywood [15]. Onthe other hand, the exhaust flow exhibits two peaks

Figure 9. Variation of major and minor species duringhydrogen combustion in an engine operatingwith an inlet manifold temperature of 425 K,

an inlet pressure of 1.5 bar, and anequivalence ratio of 0.3.

The first is due to the rapid blowdown event that occurswhen the exhaust valve opens, and the second isconsistent with the piston motion. Figure 10b shows thecalculated mean flux velocities for both the intake andexhaust velocities. The large flow velocities have a direceffect on the in-cylinder flow field and resulting heattransfer coefficient. Note that negative velocitiescorrespond to reverse flows due to an adverse pressuregradient across the valve.

Including the gas exchange processes within the cyclesimulation, as opposed to specifying thermodynamicconditions at BDC, allows the simulation to be used forvariable valve timing studies. For demonstrationpurposes, the exhaust valve closing time was varied from725 to 795 ca-deg to demonstrate the effect that it hason ignition and the resulting engine performance. Thetest case was run using hydrogen fuel, an inlet pressureof 1.5 bar, and an inlet manifold temperature of 500Kalthough the results are qualitatively the same at otheoperating conditions. Figure 11 shows results for ignitiontiming, indicated thermal efficiency, and volumetricefficiency. Closing the exhaust valves early results in a

reduction in scavenging efficiency. This causes moreinternal EGR to reside in the cylinder after the exhaustvalves close, and leads to higher mixture temperaturesafter induction. The elevated temperatures and activeradicals in the internal residual gas cause earlier ignitiontiming. Since ignition timing is advanced into thecompression stroke, the thermal efficiency is greatlyreduced with earlier ignition.

0.1

1

10

100

1000

0.0001 0.001 0.01

LNP

ressure(bar)

LN Volume (m3)

(a)

0

20

40

60

80

100

120

0

500

1000

1500

2000

2500

200 250 300 350 400 450 500

Pressure(bar)

Temperature(K)

Crank Angle (deg)

Temperature

Pressure

Ignition Point

(b)

-0.0005

0

0.0005

0.001

0.0015

-0.0001

0

0.0001

0.0002

0.0003

0.0004

0.0005

0.0006

0.0

08

0.0

085

0.0

09

0.0

095

0.0

1

0.0

105

0.0

11

0.0

115

0.0

12

OH

HO2

O

H2O2

N

NO

MajorSpeciesMoleFractions(-)

p

()

Time (Sec)

HO2, H, and OH, productionprecursor for ignition

NO Production





12/1910

Figure 10. (a) Mass flows and (b) velocities through theintake and exhaust valves for hydrogen-fueled

engine operating at 1500 rpm.

Figure 11. Variation of volumetric efficiency, thermalefficiency, and ignition timing for differentexhaust valve closing times. Hydrogen-fuelledengine operating with intake manifoldpressure of 1.5 bar and intake manifoldtemperature of 500 K.

TURBULENT FLOW MODEL The eddy length scale ofthe turbulence is governed as a function of the distancebetween cylinder head and piston, as noted in Eq. (33)Its variation through the engine cycle is shown in Figure12a. More details can be found in Assanis and Heywood[15]. Figure 12b shows the variation in the mean flowvelocity, turbulent intensity, and the overall characteristicvelocity. It is noted during the intake stroke that the meankinetic energy is initially large and then decreases inresponse to piston motion. It decreases more slowlyduring the compression stroke. The mean flow thenincreases during the exhaust blowdown, a result of thelarge pressure potential across the valve. The largescale turbulence levels are shown to be very high duringthe intake flow. Later in the intake event, turbulencedecays since the rate of dissipation, which is governed bya scaling argument, due to viscous shear stresses islarger than the rate of turbulence production. The use oRDT to account for rapid density changes results in aslight amplification of the turbulence intensity duringcombustion. The characteristic velocity is a scalingargument made up of mean piston speed, mean gas

velocity, turbulent intensity, and during the exhaust, ablowdown velocity. The variation of both velocity flowscales as well as the piston speed, has a direct effect onthe characteristic velocity as well as the subsequent heatransfer process.

HEAT TRANSFER MODEL Once the characteristicvelocity is found, the film coefficient for heat loss can becalculated. The heat transfer coefficient and heat transfeper unit area are shown in Figure 13. It is noted that iexperiences a strong variation over the cycle. This is dueto a combination of variation in gas properties (i.ethermal conductivity and viscosity) as well as the

characteristic velocity and the macroscale of turbulenceThe second peak in the film coefficient is due to the largeexhaust blowdown.

A heat transfer study was performed to examine theeffect of different heat transfer models on ignitionSpecifically, the Woschni heat transfer model [13], thahas been used in the majority of published HCCI studieswas compared to the zero-dimensional k- based heatransfer model by Assanis and Heywood [15]. Theadiabatic combustion case was used as a baseline in thestudy. The comparisons were run at a compression ratioof 17, an inlet pressure of 3 bar, and an inlet temperature

of 450 K, although similar trends follow at other operatingconditions. The pressure and temperature profiles areshown in Figs. 14a and 14b. Clearly, the heat transfemodel has a pronounced effect on the ignition point. Asanticipated, the adiabatic case ignites first because thelack of heat transfer promotes a higher mean cylinder gastemperatures and hence earlier ignition. But what ismost surprising is that the Woschni model predicts theignition point 8 degrees later than that by Assanis andHeywood [15]. This discrepancy is attributed to the facthat the Woschni model was developed for DICI dieseengines by fitting data to the correlation form. In other

-0.2

0

0.2

0.4

0.6

0.8

1

1.2

-100 0 100 200 300 400 500 600 700

Crank Angle (deg)

Mass

Flow/MassFlow

maximum

Exhaust Blowdown

Intake Exhaust

(a)

-200

-100

0

100

200

300

400

-100 0 100 200 300 400 500 600 700

ValveVelocity

(m/s)

Crank Angle (deg)

Reverse Flow To Intake Manifold

Intake Valve Velocity

Exhaust Valve Velocity(b)

20

40

60

80

100

120

335

340

345

350

355

360

720 730 740 750 760 770 780 790 800

Ind.T

hermalEfficiency,

VolumetricEfficiency(%)

IgnitionPoint(deg)

Exhaust Valve Closing (deg)

Ignition Point

Thermal Efficiency

Volumetric Efficiency





13/1911

words, it is a linearized convective/radiative modelspecifically applicable to non-premixed combustion.

Figure 12. Cyclic variation of (a) integral length scale;(b) mean flow velocity, turbulent intensity, andcharacteristic velocity at 1500 rpm.Hydrogen-fuelled engine with an inletpressure of 1.5 bar, an inlet temperature of425 K and an equivalence ration of 0.3.

Figure 14c shows the heat transfer coefficients for theWoschni correlation and that predicted by the k- model.The Woschni coefficient shows very little variation overthe gas exchange event, when heat transfer should behigher due to mean gas flow and the production of large-scale turbulence. By including velocity scales for both themean flow and turbulence intensity, Assanis andHeywood [15] predict an elevated film coefficient duringthe gas exchange processes. Over the combustion eventthe Woschni model predicts a film coefficient on the orderof three times larger than the peak value of the k- model.This is again attributed to radiation being indirectlypresent in the Woschni predictions.

Figure 13. Cyclic variation of heat transfer coefficientand gas-to-wall heat transfer rate for ahydrogen-fuelled engine at 1500 rpm, an inletpressure of 1.5 bar, an inlet temperature of

425 K and an equivalence ratio of 0.3.

PARAMETRIC STUDIES WITH ENGINESIMULATION

Having demonstrated the behavior of the physicasubmodels in the four-stroke HCCI engine cyclesimulation, several parametric studies were performed toexplore the effect of intake mixture conditions, as well asgeometric and effective compression ratios on engineperformance and efficiency. These parametric studieshave been carried-out for the engine specifications listed

in Table 1, but for geometric compression ratios asspecified. Both hydrogen and methane have been usedas the fuels for these engine studies. The operatingspeed was 1500 rpm. Integrated cycle results areinterpreted in the light of cycle-resolved pressure andtemperature profiles, as well as onset of ignition.

TEMPERATURE SWEEP The first study was run withhydrogen at a compression ratio of 15:1, an inlet manifoldpressure of 1.2 bar, and an equivalence ratio of 0.3. Theinlet manifold temperature was varied between 400 K and800 K, and the effect on the ignition point was studied. Asshown in Fig. 15a, when temperature is increased, i

advances the onset of ignition due to increased reactionrates. However, it is also important to note howincreasing inlet manifold temperatures negatively affecindicated thermal efficiency, volumetric efficiency, andtrapped mass. The trends captured in Fig. 15b reflecthat an increase in the inlet temperature reduces trappedmass and volumetric efficiency, which in turn adverselyaffects torque and power output. Advanced ignitionwhich increases compression effort, combined withreduced volumetric efficiencies leads to the observedreduction in net indicated thermal efficiency. It is alsonoted that one of the cases did not

0

0.1

0.2

0.3

0.4

0.5

0.6

-100 0 100 200 300 400 500 600 700

TurbulenceLengthSc

ale/PistonBore

Crank Angle (deg)

TDC

(a)

0

20

40

60

80

100

120

140

-100 0 100 200 300 400 500 600 700

Mean Flow VelocityTurbulent IntensityCharacteristic Velocity

FlowVelocity(m/sec)

Crank Angle (deg)

Turbulence Amplification

Exhaust Blowdown

(b)

0

50 0

1000

1500

2000

0

20

40

60

80

100

100 200 300 400 500 600

HeatTransferCoefficient(w/m

2K)

RateofHeatTrans

fer(kW)

Crank Angle (deg)

Exhaust Blowdown

He at T ra n sfer Co ef fic ien t Ra te o f Hea t T ra n sfer





14/1912

Figure 14. (a) Pressure, (b) temperature, and (c) heattransfer coefficient profiles predicted by k-heat transfer model, Woschni heat transfermodel, and assuming adiabatic operation.HCCI engine with CR=17, operating onnatural gas, with an equivalence ratio of 0.4,intake manifold pressure of 3.0 bar, andintake manifold temperature of 450 K.

Figure 15. Effect of inlet manifold temperature on (a)ignition timing and cyclic pressure profile, and(b) Thermal efficiency, mass trapped andvolumetric efficiency. Hydrogen-fuelledengine with intake manifold pressure of 1.5bar and equivalence ratio of 0.3.

ignite, therefore demonstrating how the simulation can beused to map an engine's flammability limit.

EQUIVALENCE RATIO SWEEP The next study wasconducted with hydrogen fuel for a range of equivalenceratios between 0.15 and 0.4, at an inlet manifold pressureof 1.2 bar and a temperature of 400 K. The enginecompression ratio was maintained at 15. What wasnoted is that increased fuel concentrations promotedradical generation and thus led to shorter inductiondelays. The increased heat release associated withelevated equivalence ratios promotes higher cylindepressures and specific power outputs, up to the point thatits benefits are offset by the increased compression workcaused by overadvanced ignition. The simulationdemonstrates how the lean flammability limit of theengine can be investigated. In this case, note from Fig

0

50

100

150

200

250

300

340 350 360 370 380 390 400

Adiabatic, IGN = 354.5

k- model, IGN = 368.1

Woschni, IGN = 376.1

Pressure(bar)

Crank Angle (deg)

(a)

0

500

1000

1500

2000

2500

340 350 360 370 380 390 400

Adiabatic, IGN = 354.5

k- model, IGN = 368.1

Woschni, IGN = 376.1

Temperature(K)

Crank Angle (deg)

(b)

0

5000

1 1 04

1.5 104

2 1 04

0 100 200 300 400 500 600

k- model

Woschni model

HeatTransferCoefficient(W/m

2K)

Crank Angle (deg)

(c )

0

20

40

60

80

10 0

12 0

300 320 340 360 380 400

400 K

425 K

450 K

500 K

600 K

P

ressure(bar)

Crank Angle (deg)

(a )

-20

0

20

40

60

80

100

120

1.8

1.9

2

2.1

2.2

2.3

2.4

2.5

2.6

350 400 450 500 550 600 650

ThermalEfficiency(%

)MassTrapped(gram)

Temperature (K)

No Ignition

Volumetric Efficiency

Thermal Efficiency

Mass Trapped

(b)





15/1913

16 that hydrogen mixture fails to ignite at an equivalenceratio of 0.15.

Figure 16. Effect of equivalence ratio on ignition timingand cyclic pressure profile. Hydrogen-fuelled

engine with intake manifold pressure of 1.2bar and intake manifold temperature of 410 K.

COMPRESSION RATIO SWEEP One of the criticismsof the lean burn HCCI concept is that it lends itself to avery efficient, but low power output operation. This is inpart due to the fact that extremely high inlet manifoldtemperatures are often needed to promote ignition ofnatural gas at top dead center (TDC). As a result thevolumetric efficiency and overall engine power output areadversely affected. To circumvent this problem, a highcompression ratio can be used to elevate cylinderpressures and temperatures. As a result, the fuel-air

mixture would not have to be heated to extraordinarilyhigh levels to promote ignition near TDC, and theelevated expansion ratios would promote increasedthermal efficiencies.

Realizing that both geometric compression ratio and inletboost level contribute to the effective compression ratiothat the mixture is exposed to, it is important toinvestigate the optimum levels for each. To address thisissue, a simulation case study was performed for differentgeometric compression ratios, but modulated inlet boostpressure levels to yield a constant effective compressionratio. The latter was defined to produce the same peak

motoring pressure under all conditions. The chosengeometric compression ratio set included 20, 28, and35:1. The corresponding inlet pressures were adjustedto values of 3, 2.5, and 1.49 bar, respectively. Thisexercise was carried out for two levels of intake manifoldtemperature, i.e. 420 K and 450 K.

As shown in Fig. 17, the lowest geometric compressionratio resulted in the highest Indicated Mean EffectivePressure. This is a result of the increased trapped masscaused by the higher boost level used in conjunction withthe lowest geometric compression ratio. Notice also that

induction delay increases as geometric compression ratiois decreased. Figure 18, which shows temperatureprofiles for the three cases at 420 K, clearly shows thatemperature rises at a slower rate in the case with thelowest mechanical compression ratio. Since theArrhenius kinetic rates exhibit an exponentiatemperature dependence, a lower temperature profile wilcontribute to a longer induction delay as a result of slowerates of radical generation.

Figure 17. Effect of mechanical compression ratio oninduction delay and mean effective pressure.Methane-fuelled engine, operated at threeintake manifold pressures (3.0, 2.0 and 1.49bar) and two intake manifold temperatures(420 K and 450 K), all at an equivalence ratioof 0.4.

Figure 18. Cyclic gas temperature profiles for threedifferent mechanical compression ratios, butthe same effective compression ratio..Methane-fuelled engine at an intake manifoldtemperature of 420K and an equivalence ratioof 0.4.

0

20

40

60

80

100

300 320 340 360 380 400

= 0 .1 5

= 0 .2

= 0 .3

= 0 .3 5

= 0 .4

Pressure(bar)

Crank Angle (deg)

340

345

350

355

360

365

370

0

2

4

6

8

10

12

14

12 16 20 24 28 32 36 40 44

ID (Tman

= 420 K)

ID Tman

= 450 K

IMEP (420 K)

IMEP (450 K)

InductionDelay(deg)

IndicatedMeanEffectivePressure(bar)

Mechanical Compression Ratio

Pmanifold

= 3.0 bar

Pmanifold

= 2.0 bar

Pmanifold

= 1.49 bar

0

500

1000

1500

2000

2500

320 330 340 350 360 370 380 390 400

35 : 1 Compression Ratio

28 : 1 Compression Ratio20 : 1 Compression Ratio

Temperature(K)

Crank Angle (deg)

20:1 Compression Ratio





16/1914

Since the case with an inlet pressure of 3 bar exhibitedthe highest efficiency, a parametric study was performedto optimize geometric compression ratio at this level ofinlet boost. To promote ignition at lower compressionratios, inlet temperature was elevated to 500 K. Theequivalence ratio was held constant at a lean value of0.4 and the compression ratio was swept from 12 to 24:1.

The resulting pressure profiles, near TDC, are shown inFig. 19a. First, it is noted that the increased geometric

compression ratio promotes a reduction in the inductiondelay, as shown in Fig. 19b. This is a result of moreelevated temperature histories earlier in the cycle, whichpromote chain-initiating reactions earlier. The timing ofoccurrence of ignition greatly affects engine power outputas well as mechanical stresses. Figure 19c correlatesthe timing of ignition with Indicated Mean EffectivePressure (IMEP). Early ignition of the fuel-air charge,

relative to TDC, causes elevated pressures during thepiston upstroke. This increases the compression workwhich in turn reduces the net piston work, as well asincreases peak cylinder pressures dramatically, from 190bar to approximately 330 bar. Consequently, IMEP isreduced from 11.5 bar to under 10 bar, and brake thermaefficiency drops from over 50% to 44% (see Fig. 19d).

Figure 19c also depicts the behavior of IMEP non-dimensionalized with peak cylinder pressure as a

function of the timing of ignition. This normalized ratioessentially a measure of power output per unitmechanical stress, is reduced by a factor of 2 asgeometric compression ratio is increased from 12 to 24This indicates a disproportionate stress penalty for anHCCI engine where the onset of ignition occursprematurely. This case study suggests that a geometriccompression ratio of 17:1 with an inlet boost of 3 bar can

Figure 19. Effect of mechanical compression ratio on (a) cyclic pressure profile, (b) induction delay, (c) ignition timing andIMEP, and (d) BMEP and brake thermal efficiency. Methane-fuelled engine at an intake manifold temperature of 500K, anintake manifold pressure of 3.0 bar and an equivalence ratio of 0.5.fig 19

0

50

100

150

200

250

300

350

320 340 360 380 400

CR 12 :1

CR 15:1

CR 17:1CR 20:1

CR 22:1

CR 24:1

Pressure(bar)

Crank Angle (deg)

(a)

8

8.5

9

9.5

10

10.5

11

11.5

2

3

4

5

6

7

345 350 355 360 365 370

IMEP(bar)

100*IMEP/Pmax(bar)

Ignition Timing (deg)

IMEP/Pmax

IMEP (c)

166

168

170

172

174

176

178

180

182

15 20 25

Ind

uctionDelay(Sec)

Compression Ratio

(b)

9

9.5

10

10.5

11

11.5

43

44

45

46

47

48

49

50

51

14 16 18 20 22 24 26

BM

EP(bar)

BrakeThermalEfficiency

Compression Ratio

Thermal Efficiency

BMEP

(d)





17/1915

produce reasonable power outputs in an HCCIconfiguration, while limiting peak cylinder pressures.This finding is in line with tests that have been performedby Christensen et. al., [29]; and simulation work [7]. Moreimportantly though, this case study has demonstratedhow the four-stroke thermokinetic simulation can be usedto define efficient operating ranges for the HCCI engine.

CONCLUSIONS

A computer simulation of the Homogenous ChargeCompression Ignition (HCCI) engine has been developedfor ignition and performance studies. The simulationcouples models for mass, species, and energy within azero-dimensional framework. A methodology has alsobeen developed for integrating CHEMKIN libraries withphysical models of engine processes in the context of afull cycle simulation. Parametric studies have beenconducted to illustrate the behavior of the combustionmodel, as well as the other physical process models.The cycle simulation has been exercised to study ignitionand performance of natural-gas and hydrogen fuelled

HCCI engines as a function of operating and designconditions. The following conclusions have been drawnfrom our study:

1. It has been demonstrated that the widely availableand used CHEMKIN libraries can be effectivelyintegrated within a full cycle simulation of the HCCIengine. This enables the user to select from differentfuels, and readily describe their chemical kineticsusing reported simplified or complex chemicalschemes, as appropriate. Use of CHEMKIN alsomakes it possible to take advantage of its integratorDVODE to solve stiff chemical reaction systems.

2. For illustration, two chemical kinetics schemesdescribing hydrogen and natural gas chemistry havebeen implemented in the code. The hydrogenscheme is a reduced one, consisting of 11 speciesand 23 reactions. The natural gas chemistry isdescribed via the GRI-Mech 3.0 that considers 53species and 325 reactions, including NOx chemistry.The behavior of the complex chemistry scheme wasfirst studied in a variable volume bomb. Predictedignition trends were consistent with those reported inthe literature as temperature, pressure, equivalenceratio, and composition were varied.

3. It has been demonstrated that predicting heattransfer losses based on velocity and length scalesthat are physically connected to in-cylinderprocesses (gas exchange, mean flow, turbulence,piston motion) are substantially different than thosebased on the empirical, Woschni heat transfercorrelation. It is felt that the experimental databaseon which the latter is based is not appropriate forHCCI engines. It is also shown that differences inheat transfer predictions can shift the predicted startof ignition by up to 10 deg.

4. Several parametric studies of inlet temperaturepressure, and equivalence ratio were carried outusing the complete cycle simulation. It was notedhow mixture preparation as well as the ensuingignition point greatly affect the volumetric efficiencyand the power output. It was also concluded thaHCCI operation with Natural Gas needs highcompression ratios to promote ignition with reducedintake temperatures.

5. A case study was performed for varying geometriccompression ratios, but at constant effectivecompression ratios achieved by modulating inlemanifold pressure. It was noted that the lowesmechanical compression ratio and highest inlet boosproduced the highest efficiency and power outputThis is a result of ignition timing occurring after TDCdue to a lag in the gas temperature.

6. A follow-up study was then performed using an inlepressure of 3 bar over varying geometriccompression ratios. The tradeoffs associated withpower and mechanical stress suggest that a 17:1compression ratio with 3 bar of boost lends itself to

reasonable power outputs for the HCCI, Natural Gasconcept.

Overall, this study demonstrates the importance ofcoupling detailed chemistry descriptions with physicamodels of the rest of the HCCI engine processes, notablygas exchange and turbulent heat transfer.

ACKNOWLEDGEMENTS

The authors gratefully acknowledge a gift fromCaterpillar, Inc. to the University of Michigan that wasused to partially support this study.

REFERENCES

1. Onishi, S., S. J. Hong, K. Shoda, P. Do Jo, and SKato, 1979, Active Thermo AtmosphericCombustion (ATAC) A New Combustion Process forInternal Combustion Engines, SAE Paper 7905011979.

2. Najt, P.M. and D.E. Foster, Compression-IgnitedHomogeneous Charge Combustion, SAE Pape830264, 1983.

3. Thring, R.H., Homogeneous Charge CompressionIgnition (HCCI) Engines, SAE Paper 892068, 1989.

4. Van Blarigan, P., N. Paradiso, and S. GoldsboroughHomogeneous Charge Compression Ignition with aFree Piston: A New Approach to Ideal Otto CyclePerformance, SAE Paper 982484, 1998.

5. Van Blarigan, P., and S. Goldsborough, A NumericaStudy of a Free Piston Engine Operating onHomogeneous Charge Compression IgnitionCombustion, SAE Paper 990619, 1999.

6. Ryan, T.W., and T. Callahan, Homogeneous ChargeCompression Ignition of Diesel Fuel, SAE Pape961160, 1996.





18/1916

7. Smith, J.R., S.M. Aceves, C. Westbrook, and W. Pitz,Modeling of Homogeneous Charge CompressionIgnition (HCCI) of Methane, Proceedings of the1997 ASME Internal Combustion Engine FallTechnical Conference, ASME Paper No. 97-ICE-68,ICE-VOL. 29-3, pp. 85-90, 1997.

8. Gray, A.W., and T.W. Ryan, Homogeneous ChargeCompression Ignition of Diesel Fuel, SAE Paper971676, 1997.

9. Naber J.D., D.L. Siebers, J.A. Caton, Westbrook C.K.,Di Julio S.S., Natural Gas Autoignition Under DieselConditions: Experiments and Chemical KineticModeling, SAE Paper 942034, 1994.

10. Agarwal, A. and D.N. Assanis, Modeling the Effectof Natural Gas Composition on Ignition Delay UnderCompression Ignition Conditions, SAE Paper971711, 1997.

11. Aceves, S.M., Smith, J.R., Westbrook, C.K., Pitz,W.J., Compression Ratio Effect on Methane HCCICombustion, Journal of Engineering for GasTurbines and Power, Vol. 121, pp. 569-574.

12. Kusaka, J., T. Yamamoto, Y. Daisho, R. Kihara, T.Saito, and O. Shinjuku, Predicting HomogeneousCharge Compression Ignition Characteristics ofVarious Hydrocarbons, Proceedings of the 15thInternal Combustion Engine Symposium(International), Seoul, Korea, 1999.

13. Woschni, G., A Universally Applicable Equation forthe Instantaneous Heat Transfer Coefficient in theInternal Combustion Engine, SAE Paper 670931,1967.

14. Poulos, S.G., and Heywood J.B., The Effect ofChamber Geometry on Spark-Ignition EngineCombustion, SAE Paper 830334, SAE Trans., Vol.92, 1983.

15. Assanis, D. N., and Heywood, J. B., Developmentand Use of Computer Simulation of theTurbocompounded Diesel System for EnginePerformance and Component Heat Transfer Studies,SAE Paper 860329, 1986.

16. Kee, R.J., F.M. Rupley, and J.A. Miller, Chemkin-II: AFORTRAN Chemical Kinetics Package for theAnalysis of Gas-Phase Chemical Kinetics, SandiaNational Labs Report SAND89-8009B, 1991.

17. Van Wylen, G.J., R. Sonntag, R., and C. Borgnakke,Fundamentals of Classical Thermodynamics, 4thedition, John Wiley & Sons, 1994.

18. Assanis, D.N., and Polishak, M., Valve Event

Optimization in a Spark-Ignition Engine, ASMETrans., Journal of Eng. for Gas Turbines and Power,Vol. 112, pp. 341-347, 1990.

19. Jessee, J.P., R.F. Gansman, and W.F. Fiveland,Multi-Dimensional Analysis of Turbulent Natural GasFlames Using Detailed Chemical Kinetics,Combustion Science and Technology, Vol.129, pp.113-140, 1997.

20. Agarwal, A., Z. Filipi, D. N. Assanis, and D. Baker,Assessment of Single-and Two-Zone TurbulenceFormulations for Quasi-Dimensional Modeling ofSpark-Ignition Engine Combustion, CombustionScience and Technology, Vol.136, pp. 13-39, 1998.

21. Shampine, L.F., and M.K. Gordon, ComputeSolution of Ordinary Differential Equations: The InitiaValue Problem, Freeman, 1974.

22. Maxwell T.T. and J.C. Jones, Alternative FuelsEmissions, Economics, and Performance, Society oAutomotive Engineers, Inc., Warrendale, PA, 1990.

23. Blarigan P.V., Development of a Hydrogen FueledInternal Combustion Engine Designed for SingleSpeed/Power Operation, SAE Paper 961690, 1996.

24. Smith, G.P., G.M. Golden, M. Frenklach, N.WMoriarty, B. Eiteneer, M. Goldenberg, T. Bowman, RHanson, S. Song, G.C. Gardiner Jr., V. Lissianskiand Z. Qin, GRI-MECH 3.0, http:/www.me.berkeley.edu/gri_mech/

25. Glassman, I., Combustion, Academic Press, Inc.San Diego, Ca, 1996.

26. Liss, W.E., and W.H. Thrasher 1991, Natural Gas asa Stationary Engine And Vehicle Fuel, SAE Pape912364.

27. Westbrook C.K., and Pitz W., 1983, Effects oPropane and Ignition of Methane-Ethane-AiMixtures, Combustion Science and Technology, Vol

33, pp. 315-319.28. Fraser, A.F., Siebers, D.L., Edwards, C.F.

Autoignition of Methane and Natural Gas in aSimulated Diesel Environment, SAE Paper 9102271991.

29. Christensen, M. and Johansson B., Influence oMixture Quality on Homogeneous ChargeCompression Ignition, SAE Paper 982454, 1998.

NOMENCLATURE

Symbol Definition

Thermodynamic PropertiesP Cylinder Gas Pressure

T Temperature

u Internal Energy

h Enthalpy

SH Specific Heat

Cv SH at Const. Volume

Cp SH at Const. Pressure

R Gas Constant

Densityv Specific Volume

Equivalence Ratio

Yj Mass Frac. Species jWmw Molecular Weight

Energy Transfers

W Displacement Work

Qht Heat Transfer

Engine Geometry

Ab Piston Bore Area

Vd Displacement Volume





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Chemical Kinetic Parameters

Ea Activation Energy

kf,r F/R Reaction Rates

Mass Rate of Prod.

i Molar Rate of Prodi,ii Stoichiometric Coeff.qi Progress Variable

Heat Transfer Parameters

U Mean Cylinder Velocityu Turbulent Velocity

KE Kinetic Energy

K Mean KE

k Turbulent KE

P Turbulent Energy Prod.

t Turbulent ViscosityIntegral Length Scale

A Amplification (RDT)

Eddy DissipationVch Characteristic Velocity

kg Thermal Conductivity

hg Film Coefficient

jw!

'




HCCI performance and combustion

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Transcript of HCCI performance and combustion