Y LIU PACKET MODEL

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Development of a multizone model fordirect injection diesel combustion

Y Liu and K C Midkiff

Department of Mechanical Engineering,

The University of Alabama, Tuscaloosa, Alabama,

USA

S R Bell

School of Engineering, University of Kansas,

Lawrence, Kansas, USA

Accepted 26 August 2003

Abstract: Diesel engines have attracted considerable engines are high thermal efficiencies and reduced

attention in recent years because of the increasingly emissions of hydrocarbons and oxides of nitrogen

restrictive `engine-out emission standards being adopted (NOx ). However, diesel engines still emit high levels

by regulatory agencies. The cutting-edge technologies of of particulates and achieving acceptable NOx levels

emissions reduction in engines fall into three categories: is becoming a challenge as more restrictive emissions

preprocessing, improved combustion processing and post- restrictions are implemented. Consequently, motiv-

processing. An engine cycle simulation was developed to ation for developing cleaner-burning diesel engines

investigate and, thus, find possible avenues of reducing has escalated.

emissions through modifying the combustion process. This Three categories of emission reduction strategies

simulation includes models for fresh air charging, fuel and are commonly applied to engines: preprocessing,

air mixing, wall heat transfer, diesel droplet evaporation, combustion processing and postprocessing. Here,

ignition delay and mixture combustion with species equi- `preprocessing refers to all technologies used to treat

librium reactions. These models, together with a thermo- the fresh air and/or fuel before its admission into

dynamic analysis of the cylinder gas, yield instantaneous the cylinder and `postprocessing refers to the tech-

cylinder conditions, overall indicated engine performance nologies applied to the exhaust gas after it leaves the

and a prediction of the engine-out NOx and soot emissions. engine cylinder. The application of preprocessing

The engine parameters and operating conditions used technologies generally increases the production cost

in the work presented here were chosen to be representative of the fuel and reduces emissions by influencing the

of a Caterpillar 3401 single-cylinder diesel engine. Experi- combustion process. The obvious shortcomings ofmental investigations were also conducted with the engine, postprocessing methods are the extra cost in main-

and the combustion model has been verified by comparing tenance and replacement of these devices. Because

the experiment results to the simulation results. numerical and experimental studies play a unique

role in combustion improvement for engines, further

Key words: multizone model, engine cycle simulation, numerical and experimental studies are needed.

emission modelling Engine models of varying complexity are com-

monly employed, including relatively simple zero-

dimensional thermodynamic combustion models as

1. Introduction well as complex three-dimensional models. Zero-

dimensional models of diesel engines are typically

used to analyse the heat release and fuel-mass burn-Diesel fuel is an attractive fuel for reciprocatinginternal combustion engines due to its relatively low ing rates based on the solution of a system of ordi-

nary differential equations for pressure, temperatureprice and its availability. Compared to spark-ignited

gasoline engines, advantages inherent in diesel and mass. However, they do not explicitly model the

71Int. J. Engine Res. Vol. 5 No. 1JER 02601 2004 IMechE


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Y Liu, K C Midkiff and S R Bell

detailed phenomena, such as diesel fuel droplet

vaporization, air entrainment, local temperatures

and local equivalence ratio, that vary temporally and

spatially. Three-dimensional models use time-

averaged and turbulence-correlated conservation

equations of mass, momentum, energy and atomic

species. In addition to the possible drawbacks ofmesh-dependence and mathematically inaccurate

approaches to modelling viscous flow and boundary

conditions, the time cost to obtain solutions for the Fig. 2 Schematic of packet development and air entrainmentgoverning equations is often burdensome. Quasi- during fuel injection.dimensional models can be used to provide some

details of the temporal and spatial phenomena of include mass conservation, energy conservation, thecombustion with the potential to reduce the compu- ideal gas equation of state and the cylinder gastational time significantly. volume constraint. The thermodynamic state of the

The engine cycle simulation developed in this gas in each packet is determined from a differentialwork was used to model a Caterpillar 3401 engine, equation for the packet gas energy as shown in Fig. 3:which is a single-cylinder, four-stroke, water-cooled,

turbocharged direct injection diesel engine. dEgdt

=Q combustion W +Q heat transfer m lhl +m ea hea

(1)

2. Cycle Simulation Descriptionwhere Q combustion is the chemical energy released due

to diesel combustion, W is the boundary work occur-A schematic describing the major processes in a typi-ring on the packet, Q heat transfer represents the radiativecal engine cylinder, which include combustion, heatand convective heat losses from the packets and thetransfer and piston work, is shown in Fig. 1. As fuellast two terms represent enthalpies of the vaporizedis injected into the cylinder, several fuel packets ofdiesel fuel and entrained air respectively. Theequal mass are assumed to be formed for each crank

internal energy of each packet may be written asangle. Packets formed at different crank angle timesare subject to varying evaporation rates and air dEg

dt=mg Cv T

g +Cv Tg m g (2)entrainment rates. The packet development is sche-

matically described in Fig. 2. The thermodynamicCombining equations (1) and (2), an expression forstate of the cylinder gas in each of the packets isthe packet temperature can be obtained asevaluated continuously throughout the engine simu-

lation, and the combustion process details within theTg =

1

mg Cv(Q combustion W

+Q heat transferpacket are provided at each time step.

The governing equations used in the simulation m l hl +m ea hea m g CvTg ) (3)

RungeKutta integration of equation (3) is performed

on each packet, yielding instantaneous local tem-

perature across the cylinder. An average cylinder

Fig. 1 Schematic of a d irect injection (DI) engine with thermo-

Fig. 3 Schematic of the thermodynamic state in local packets.dynamic processes identified.

72 Int. J. Engine Res. Vol. 5 No.1JER 02601 2004 IMechE


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Development of a multizone model for DI diesel combustion

temperature is then calculated at each instant and, Equation (4) suggests three critical features. Firstly,

the entrained air mass should be related to the massusing the ideal gas equation, the cylinder gas press-

ure is determined. From the close of the intake valve of the surrounding fluid. Secondly, the overall

entrained air mass should increase as the fuel injec-to the start of fuel injection (FIS), a single zone is

used to simulate the compression process, with tion continues. Finally, at each time step there should

be stratification of entrainment into different packets.property values updated each time step.

To calculate Q combustion in equation (3), several sub- Each packet is described by an entry index, I, and apacket classification index, J. It is the total number ofroutines are used for modelling the processes of fuel

evaporation, air entrainment, ignition delay, diesel entry indices and Jt is the total number of packet

classification indices. Kis an arbitrary air entrain-chemical energy release and combustion product for-

mation. The boundary work rate (W ) is obtained by ment constant. The parameter mea refers to the

entrained mass of the local packet with entry indexthe simple calculation of the product of pressure, P,

and volume change, DV, across each calculation time and classification index Jand Irespectively. The par-

ameter mub is the total mass of the unburned zone.step. The heat loss rate (Q

heat transfer ) includes the heat

exchange processes with the surroundings. The last The parameters h and hend are the instantaneous cal-

culation crank angle and the fuel-injection-end crankthree terms on the right-hand side of equation (3)

describe diesel droplet enthalpy, entrained air angle respectively. The parameters Dhcalc and Dhinjare the calculation time step and overall fuel injectionenthalpy and the internal energy change of a packet.

duration respectively used in the calculation. The

parameter hs is the time required to completely3. Subroutine developmententrain the surrounding fluid into the spray jet or

the time required for the spray jet to engulf theDuring the diesel injection process, fuel is sprayed

combustion chamber volume.into a relatively high-temperature, high-pressure cyl-

For simplicity, diesel droplet evaporation wasinder charge where the liquid begins vaporizing. As

modelled by the isolated, steady state, single-dropletdescribed earlier, the non-uniform mixture in the cyl-

approach of Spalding [3, 4]. To summarize, the solu-inder is modelled using multiple zones and the

tion of mass flux out of the diesel droplet surface isinjected fuel is divided evenly into the packets. In

obtained after solving the energy and species equa-developing the simulation for non-uniform mixtures, tions for the droplet through application of suitablethe desire was to introduce air/fuel stratification in

boundary conditions. The mass flux can be writtena realistic yet simple manner. As fuel is injected, air

asfrom the surrounding unburned zone is entrained

into the various burning packets. Each packet is m =g* ln(B +1) (6)identified by the time of entry into the cylinder and

where g* =rg /Rand B is the mass transfer number.classified by a simple numbering method to permit

During evaporation prior to combustion, the masseasy identification (the packet numbering method istransfer number is evaluated by assuming the Lewisshown in Fig. 2). No geometrical characteristics arenumber equal to one (the thermal transfer equivalentassigned to the spray jet and local packets. Theto diffusion transfer). The thermal and diffusionentrained air mass is determined by a simple

transfer numbers prior to combustion may beentrainment function. As the spray jet develops, theexpressed respectively asentrained air mass increases. As the end of injection

approaches, the spray jet dissipates and the airBT =

Cg(T2 Tw )

L +Cl (Tw TR )(7)entrainment decays. A simple mathematical equation

[1, 2] for the air mass entrained, mea , was used:

and

mea (I, J) =Kmub

It Jt (Y+1) ADhcalchs B exp A

I

It

J

JtB (4)BD =

YF2 YFWYFW YFR

(8)

In equation (4), the value of injected fraction, Y, is

determined as a function of crank angle: where the subscripts 2 , w, F and R refer to the sur-

rounding ambient, the surface wall of the dieseldroplet, the fuel side and the droplet inner side

Y= Ghend h

Dhinjwhen hhend

(5) respectively. After ignition and in the presence of

combustion, modification must be made on the mass



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transfer numbers, i.e. equations (7) and (8). The yields the cylinder gas temperature, expressed as

updated equations are written as

Tg =1

mg Cv( W +Q heat transfer m ex hex

BT =DH f Y02 +Cg (T2 Tw)

L +Cl (Tw TR )(9)

+m in hin m g CvTg ) (16)

andThe mass flowrates of exhaust and intake gas can

be calculated using a simple, one-dimensional, adia-BD =

f Y02 +YFWYFR YFW

(10)batic, quasi-steady, incompressible flow equation for

an ideal gas:where frefers to the stoichiometric fueloxygen ratio

and DHis the enthalpy of combustion for diesel fuel.

An ignition delay model was used to account for m in,ex=AVP CfS2RTc

c 1 CAP2P1B

2/c

AP2P1B

(c+1)/c

D0.5

the physical and chemical preparation of the first(17)

group of flamelets in the cylinder after fuel injection.

After evaluating several models with the operating where P1 is the upstream pressure, P2 is the down-conditions of the test engine, the Hardenberg and stream pressure, AVP is the open valve port areaHase equation [5] was selected for ignition delay: depending on the valve lift, valve diameter and

design of the intake system or exhaust system, Cfistid=Kconst exp CEA A

1

RT

1

17190B+A21.2

P 12.4B0.63

D the discharge coefficient for the valve port, c is theratio of specific heats and P2 /P1 is the downstream(11)upstream pressure ratio. For choked flow, the press-

where ure ratio, P2 /P1 , in equation (17) can be replaced by

Kconst =(0.36+0.22Sp ) (12) P2P1

=A2

c +1Bc/(c 1)

(18)

EA =618840

CN+25(13)

Combining equations (16), (17) and (18) and theThe variables appearing in equations (11) to (13) are

necessary geometric parameters, solutions describ-temperature Tin Kelvin, mean piston speed Sp ing the gas exchange processes can be obtained within m/s, pressure P in bars absolute, activation energy

acceptable accuracy.EA in units of kJ/kmol, cetane number CN and The oxides of nitrogen, NOx , are formed mainly atignition delay tid in crank angle (CA) degrees. high temperature [7] in the presence of nitrogen and

A simple model was used [1, 2] for the heatoxygen. Nitrogen oxide emissions in combustion

exchange between the cylinder wall and the localresult from (a) thermal NO, oxidation of molecular

combusting packets. The heat transfer rates from thenitrogen in the post-flame zone; (b) prompt NO, for-

burning packets and the unburned zone can bemation of NO in the flame zone, and (3) fuel-bound

expressed respectively asNO, oxidation of nitrogen containing compounds in

the fuel [8]. Thermal NO is the dominant mechanismQ

i

=Qtot

V2/3i (Ti Tm )

S V2/3i (Ti Tm )+V2/3m (Tm Tw )(14)

in internal combustion engine combustion [7].Combustion products of eleven species [9] were con-

andsidered in the NOx model: H, O, N, H2 , OH, CO, NO,

O2 , H2O, CO2 and N2 . The extended Zeldovich mech-Q m =Q

tot

V2/3m (Tm Tw )

S V2/3i (Ti Tm) +V2/3m (Tm Tw )

(15)anism [8, 10], assuming that N remains in the steady

state concentration [11] and all other species concen-In equations (14) and (15), Q tot is the overall cylinder trations achieve chemical equilibrium, is used toheat transfer calculated from Woschnis correlation

determined NO kinetics. The extended Zeldovich[6] and the subscript i refers to the burning packets,

mechanism consists of the following three reactions:m to the unburning zone and w to the cylinder wall.

The summation ofQ m and all Q

i yields Q

tot .

N2+

Ou

KA

KB NO

+N (19)A thermodynamic model of the cylinder gas is

employed to simulate the exhaust and intake pro-

cesses under a transient system analysis of the N+O2uK

C

KD

NO+O (20)boundary work and heat loss. Energy conservation



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formation process usually considers the simpleN+OHu

KE

KF

NO+H (21) stages of particle generation and particle growth. The

soot formed during the combustion process also sim-

Once the concentrations of O2 , N2 , H, O, and OH in ultaneously undergoes oxidation, which can occur at

equations (19), (20) and (21) are known from equilib- precursor, nuclei and particle growth stages. Soot

rium analysis, the formation rate of N (in terms of oxidation, which is a heterogeneous reaction,

concentration) can be obtained as depends on the diffusion of reactantsto and productsfrom the soot surface as well as the kinetics of the

dCNdt

=KA CN2CO KB CNO CN reaction.

Soot formation is primarily controlled by the tem-KC CNCO

2+KD CNO CO perature, pressure and equivalence ratio of the local

KE CNCOH +KF CNO CH (22) packets. Based on the work of Hiroyasu and Kadota

[14], the following model for soot formation wasThe forward and reverse reaction rate constants KA , adopted:KB , KC , KD , KE and KF, which are functions of tem-

perature, are shown in Table 1 [12]. After substituting dmsoot,formdh

=Aform mdiesel,gas wB

1P0.5 exp A1.25104

RT Bthe steady state N concentration obtained above, thenon-linear equation for the rate of change of CNO (24)concentration (dCNO /dt) can be expressed as

where mdiesel,gas is the mass of the diesel vapour in

the local combustion packet, w is the equivalencedCNOdt

=KA CN2CO KB CNO CN

ratio of local packet, the cylinder pressure P is in

+KC CNCO2

KD CNO CO units of MPa, Ris the universal gas constant

(kJ/kmol K), Tis the temperature in local packet (K)+KE CN COH KF CNO CH (23)

and Aform and B1 are constants.

The NO concentration is obtained by solving equa- Soot oxidation is also governed by the tempera-

tion (23). ture, pressure and equivalence ratio, and the follow-

For diesel combustion, particulate emissions stem ing equation was used to model the process of soot

primarily from the combustion of locally rich mix-oxidation [14]:tures. Carbonaceous particulates formed from gas-

phase processes are generally referred to as soot. Thedmsoot,oxidation

dh=Aoxidationmsoot

PO2

PP1.8 exp A

1.4104

RT Bbalance between the formation and subsequent

(25)destruction by combustion governs the soot concen-

tration in the exhaust during engine combustion. where msoot is the mass of the soot particulate in theDespite much recent progress towards the under- local packet, PO

2is the oxygen partial pressure in the

standing of soot formation mechanisms [13], the local packet (MPa), P is the cylinder pressure (MPa),exact processes are still unclear. A study of the soot Tis the temperature in the local packet (K) and

Aoxidation is a constant.

The net soot formation rate is the summation ofRate constants (cm3/mol s) and temperature (K)

the soot formation and oxidation rates. Once the localdiesel vapour mass, soot mass, equivalence ratio,KA =1.8 10

14 exp A38370

T Boxygen mass fraction, cylinder pressure and tem-

KB = 3.81013 exp A

425

T B perature are known, the net soot formation in localpackets can be achieved by integrating over time the

KC =1.8 1010T exp A

4680

T B summation of soot formation and oxidation rates.Summation of the net soot formation for all packets

KD =3.8 109T exp A

20820

T B across the cylinder yields the overall instantaneoussoot mass in the chamber.

KE =7.11013 exp A

450

T BKF =1.7 10

14

exp

A24560

T

B 4. Results and DiscussionAs mentioned earlier, the cycle simulation has beenTable 1 Rate constants for the NO formation

mechanism [11]. developed for modelling the Caterpillar 3401 engine,



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and the simulation results are compared with the

experimental results. The engine specifications and

baseline operating conditions for the modelling

study were selected to be representative of the test

diesel engine. The major engine specifications and

baseline operating conditions are summarized in

Table 2. Experimental measurements of cylinderpressure, injection timing, NOx emissions and per-

formance parameters were conducted on the engine

at half-load and 1700 r/min. Cylinder pressure was

measured with a piezoelectric pressure sensor

coupled with an angular position encoder. A chemi-

luminescent analyser was used to measure exhaust

gas NOx concentration. Injection timing was meas-

ured using a needle lift indicator. The simulationFig. 4 Instantaneous cylinder pressures and needle lift as amodel was used to predict combustion phenomena

function of crank angle for baseline firing and motoringat the same conditions for which measurements

operations with conditions: 1700 r/min, half-load,were made.CR= 15.1, DD = 40 mm, FIS = 19 CA.For the modelling cases investigated, it was

assumed that no wall impingement of the fuel sprayFigure 5 shows the experimental and predictedoccurred and the formation of combustion products

results for the heat release rate with the experimentalwas calculated for each packet. Summation of theseneedle lift for the baseline firing operating condition.products across the cylinder yielded the concen-The ignition delay can be obtained from the rise oftration of chemical species in the exhaust gas. Thethe needle to the start of the positive heat release ratepredicted specific emissions were obtained throughin the experimental results. During the ignitiondividing the emission mass by the measured brakedelay, the heat release rate is slightly negative. Thepower to compare with the experimental brakemodel prediction of the heat release rate of premixedspecific emissions.combustion is slightly less than the experimentalFigure 4 shows that there is good agreementresult. The peaks of the heat release rate curves inbetween the experimental and predicted cylinderthe premixed combustion period were pronouncedpressures under motoring and baseline firingin both the experimental result and simulationoperating conditions. The baseline engine operatingresults. The predicted heat release rate also shows aparameters include a speed of 1700 r/min, inlet air

temperature of 348 K, inlet pressure of 1.82 bar, start

of fuel injection (FIS) at 19 CA BTDC and injection

duration of 20 CA. In the modelling work, an

assumed mean diameter of 40 mm for diesel droplets

(DD) was used.

Cylinders 1Cycle 4Bore (mm) 137Stroke (mm) 165Displacement volume (cm3 ) 2442Compr ession r atio (CR) 15.1Normal injection timing (BTDC) 20 0.5Injection duration (CA) 200.5I nlet air temperature (K ) 349Inlet air p ressur e (kPa) 182Exhaust p ressure (kPa) 131Speed (r/min) 1700Power (kW) 20.9

Fig. 5 Instantaneous heat release rates and needle lift as aTorque (N m) 117function of crank angle for the baseline firing operation

with conditions: 1700 r/min, half-load, CR= 15.1,Table 2 Caterpillar 3401 engine specifications andbaseline operating conditions. DD =40 mm, FIS = 19 CA.



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slightly longer period of late diffusion combustion

than the experimental results. Overall, good agree-

ment is seen between the experimental and predicted

heat release results.

Results of the predicted cylinder average tempera-

ture and specific NOx and soot emissions are shown

for the baseline conditions in Fig. 6. Experimentalsoot measurements have not been made and, there-

fore, are not available for verification. In equations

(24) and (25), constants for soot formation, Aform , B1and Aoxidation , were tentatively chosen as 2 10

4, 0.6

and 5102 respectively, according to the experimen-

tal results by others [15] using the same engine.

Comparison between the simulation results and

experimental soot results from elsewhere [15] sug-

gests that the order of magnitude predicted appears

reasonable. As shown in Fig. 6, the specific NOx and

soot emissions peaked and levelled off late in the Fig. 7 Instantaneous cylinder pressures as a function of

combustion process as the gas temperatures dropped crank angle for varying engine loads with conditions:

to relatively cool levels. full-load: Pin = 202 kPa, Pex =187 kPa, Tin =55 C;

Figure 7 shows the influence of the engine running quarter- and half-loads: P in =182 kPa, Pex =170 kPa,

load on the cylinder pressure. The best pressure Tin = 75 C.

match between the simulation and experiment is

seen at the half-load condition. Table 3 presents theengine emissions are given in Table 4. Reasonably

influence of the engine running load on the enginegood agreement of specific NOx emissions between

performance for both experiment and simulation.the experimental and simulation results occurred for

The results show that increased loads yielded lowerall loads, but in all cases the model underpredicts

indicated specific fuel consumption (i.s.f.c.). TheNOx emissions. This underprediction of NOx is simi-results also show good agreement between thelar to many previous modelling efforts. This prob-

experimental and predicted values for power andably occurs because the small but very hot regions

fuel consumption.in the flame sheets surrounding burning fuel pro-

The influence of the engine running load on theduce high levels of NOx emissions but are not rep-

resented well by relatively coarse models. Predicted

soot emissions increased with load, as would be

expected. Figure 8 shows the soot formation in the

cylinder as a function of crank angle for quarter-,

half- and full-load conditions. Although the pre-

dicted rate of soot formation is lower for quarter-

load, the soot formation rate is nearly the same for

both half- and full-loads. Due to the increased equiv-

alence ratio for full-load, the lower soot oxidation

yields a higher specific engine-out soot for the

prediction.

Figure 9 presents the influence of varying engine

speeds (1300, 1500 and 1700 r/min) on the cylinder

pressures. With the same inlet air temperature, inlet

pressure, exhaust pressure and engine load, decreas-

Fig. 6 Instantaneous gas average temperature and in-cylinder ing the engine speed increases the cylinder pressure

in both t he experimental and simulation results.specific NOx and soot emissions as a function of crankangle for the baseline firing operation with conditions: Table 5 presents the influence of engine speed on the

engine performance at half-load with increasing1700 r/min, half-load, CR=15.1, DD=40 mm, FIS=

19 CA. engine speed, leading to increased engine indicated



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Indicated power (kW) I.s.f.c. (g/kW h)

Load Experiment Simulation Experiment Simulation

1/4 15.1 17.6 193 1962/4 26.0 28.5 183 1864/4 47.3 50.3 181 184

Table 3 Influence of the engine load on the engine performance.

Specific NOx(g/kW h) Specific soot (g/kW h)

Load Experiment Simulation Simulation

1/4 14.8 9.2 0.082/4 14.1 12.1 0.584/4 10.2 7.2 2.0

Table 4 Influence of the engine load on the engine emissions.

Fig. 9 Instantaneous cylinder pressures as a function ofFig. 8 Instantaneous in-cylinder soot emissions as a function

crank angle for varying engine speeds with conditions:of crank angle for v arying engine loads with conditions:

Pin =182 kPa, Pex =170 kPa, Tin =75 C, half-load.full-load: Pin =202 kPa, Pex =187 kPa, Tin =55 C;

quarter- and half-loads: Pin= 182 kPa, Pex =170 kPa,cylinder pressure in both the experimental and simu-

Tin =75 C.lation results. Increasing the inlet air temperatures

slightly improves the fuel efficiency, as shown inpower and essentially constant fuel consumption.

Table 7, but leads to higher levels of specific NOxTable 6 shows that increasing engine speed slightly

emissions, as shown in Table 8, and predicted sootincreases the specific NOx emissions for both experi- emissions increase with temperature increase due toment and simulation but decreases specific soot

the higher equivalence ratio resulting from theemissions in the simulation results. Again, the model

higher inlet air temperature.slightly underpredicts NOx levels for all engine

speeds.

The cylinder pressures around top dead centre 5. Conclusions(TDC) are presented in Fig. 10, showing the influence

of varying inlet air temperatures on the combustion A phenomenological engine cycle simulation has

been developed for evaluating the use of diesel fuelprocess. Increasing inlet air temperatures lowers the



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Speed (r/min) Experiment Simulation Experiment Simulation

1300 24.7 26.6 183 1871500 25.2 27.1 183 1871700 26.0 28.5 182 186

Table 5 Influence of the engine speed on the engine performance at half-load.


Speed (r/min) Experiment Simulation Simulation

1300 13.4 11.2 0.841500 13.4 11.8 0.861700 14.1 12.1 0.58

Table 6 Influence of the engine speed on the engine emissions at half-load.

combustion and emission formation processes were

investigated along with overall engine performance.

Reasonable agreements with experimental data for

cylinder pressure and NOx were obtained using the

model. The major conclusions of the study include:

1. The multizone model for direct injection (DI)

diesel combustion was developed and verified

with limited experimental data. Coupled with the

chemical equilibrium reactions, the extended

Zeldovich mechanism and soot formation oxi-

dation were successfully integrated in the engine

simulation for calculating NOx and soot emissions.

Predicted soot emissions are yet to be verified by

comparison to experiment, although literature

data [15] suggest that the order of magnitude

predicted is reasonable.

2. The difference between predicted and experimen-Fig. 10 Instantaneous cylinder pressures as a function of

tal peak cylinder pressures in the baseline andcrank angle for varying inlet air temperatures with

parametric studies is less than 5 per cent for allconditions: Pin = 182 kPa, Pex =170 kPa, half-load,

cases. The predicted indicated specific fuel con-1700 r/min.

sumptions are very close to the experimental

results. The calculated specific NOx emissions forin a reciprocating, compression ignition engine. A

Caterpillar 3401 engine was modelled and tested varying loads, engine speeds and inlet air tem-

peratures showed reasonable agreements with theexperimentally and results from the model and

experiments have been compared. The details of the experimental results, although model predictions


Tin (C) Experiment Simulation Experiment Simulation

75 26.0 28.5 183 18695 25.8 28.2 182 186115 25.5 28.1 181 183

Table 7 Influence of the inlet air temperature on the engine performance at half-load.



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Tin (C) Experiment Simulation Simulation

75 14.1 12.1 0.5895 15.4 14.1 0.77

115 16.6 16.2 0.92

Table 8 Influence of the inlet air temperature on the engine emissions at half-load.

were somewhat lower than the measured NOx Subscripts

ea entrained airemissions.

3. The influence of engine operating parameters ex exhaust

g gas in local packet(such as engine load, engine speed and inlet air

temperature) on the engine combustion process in intake

l liquid phaseand emissions has been investigated. Models for

NOx and soot emissions provide the potential to

assist in the design and operating parameterAcknowledgementsselection for optimized operation.

The authors gratefully acknowledge the financial

support from Caterpillar, Inc. and The Center for

Advanced Vehicle Technologies at The UniversityNotation

of Alabama, which receives partial funding from

the Alabama Department of Transportation underBTDC before top dead centreProject HPP-1602 (526).Ci concentration of species i (i=H, O, N,

H2 , OH, CO, NO, O2 , H2O, CO2 or N2 )

CA crank angleReferencesCN cetane number

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DD mean diameter for diesel dropletsfueled, direct-injected, internal combustion engine. PhD

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1984 (Gordon and Breach Science Publishers, Newm mass flux out of the diesel dropletYork).surface

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