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Theoretical and experimental investigations

on the performance of dual fuel diesel engine with

hydrogen and LPG as secondary fuels

D.B. Lata*, Ashok Misra

Department of Mechanical Engineering, Birla Institute of Technology, Mesra, Ranchi 835215, India

a r t i c l e i n f o

Article history:

Received 4 March 2010

Received in revised form

9 August 2010

Accepted 9 August 2010

Available online 16 September 2010

Keywords:

Dual fuel engine

Combustion

Modeling

Alternative fuelsHydrogen

a b s t r a c t

The mathematical models to predict pressure, net heat release rate, mean gas tempera-

ture, and brake thermal efficiency for dual fuel diesel engine operated on hydrogen, LPG

and mixture of LPG and hydrogen as secondary fuels are developed. In these models

emphasis have been given on spray mixing characteristics, flame propagation, equilibrium

combustion products and in-cylinder processes, which were computed using empirical

equations and compared with experimental results. This combustion model predicts

results which are in close agreement with the results of experiments conducted on a multi

cylinder turbocharged, intercooled gen-set diesel engine. The predictions are also in close

agreement with the results on single cylinder diesel engine obtained by other researchers.

A reasonable agreement between the predicted and experimental results reveals that the

presented model gives quantitatively and qualitatively realistic prediction of in-cylinder

processes and engine performances during combustion. 2010 Professor T. Nejat Veziroglu. Published by Elsevier Ltd. All rights reserved.

1. Introduction

Due to fast depletion of fossil fuels and increase in demand of

energy with clean environment, it is not only essential to use

liquefied petroleum fuel efficiently, but also to explore other

resources of energy. Hence, several gaseous fuels such as CNG

(CH4) [1], LPG (C3H8) [2], hydrogen [3], biogas [4], producer gas

[5], etc. are being experimented as alternative fuels for dualfuel internal combustion engines. Prakash et al. [6] modified

stationary diesel engine to work on biogas/diesel dual fuel

engine.

The two zone quasi-dimensional model for the simulation of

combustion process in spark ignition engines fueled with

hydrogen, methane, or hydrogenemethane blends were devel-

oped by Federico et al. [7]. Roy et al. [8] investigated the effect of

hydrogen content in the producer gas on the performance and

emissions of a supercharged dual fuel diesel engine fueled at

constant injection pressure and injection quantity. Saravanan

et al. [9] had experimentally analyzed the combustion of

hydrogen with diesel and hydrogen with diethyl ether (DEE) and

observedan increase in brake thermal efficiency. Lambe et al.[10]

converted conventional diesel engine into hydrogen operated

dual fuel engine.

Choi et al. [11] developed the heavy-duty variablecompression engine to investigate the performance and

emission characteristics for hydrogen enriched LPG fueled

engine.Pooniaetal.[12] investigatedthe effectof intake charge

temperature, pilot fuel quantity, exhaust gas recirculation and

throttling of the intake to improve the performance of LPG-

diesel dual fuel engine.

Karim et al. [13] investigated the performance of dual fuel

diesel engine by using different proportions of CH4/H2 mixture

* Corresponding author. Tel.: 91 9431382608; fax: 91 6512275401.E-mail address: [email protected] (D.B. Lata).

A v a i l a b l e a t w w w . s c i e n c e d i r e c t . c o m

j o u r n a l h o m e p a g e : w w w . e l s e v i e r . c o m / l o c a t e / h e

i n t e r n a t i o n a l j o u r n a l o f h y d r o g e n e n e r g y 3 5 ( 2 0 1 0 ) 1 1 9 1 8 e1 1 9 3 1

0360-3199/$ e see front matter 2010 Professor T. Nejat Veziroglu. Published by Elsevier Ltd. All rights reserved.

doi:10.1016/j.ijhydene.2010.08.039
mailto:[email protected]://www.sciencedirect.com/http://www.elsevier.com/locate/hehttp://dx.doi.org/10.1016/j.ijhydene.2010.08.039http://dx.doi.org/10.1016/j.ijhydene.2010.08.039http://dx.doi.org/10.1016/j.ijhydene.2010.08.039http://dx.doi.org/10.1016/j.ijhydene.2010.08.039http://dx.doi.org/10.1016/j.ijhydene.2010.08.039http://dx.doi.org/10.1016/j.ijhydene.2010.08.039http://www.elsevier.com/locate/hehttp://www.sciencedirect.com/mailto:[email protected]

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and equivalence ratio. Ma et al. [14] developed computer

simulation to predict the performance of a hydrogen vehicle

engine. Janabi et al. [15] developed a quasi-dimensional model

to study the effect of hydrogen blending on fuel consumption

and pollutant concentrations. Liu et al. [16] developed thermo-

dynamic multi-zone model to predict combustion process in

dual fuel engine. Wang et al. [17] developed a combustion model

on the basis of CFD and reaction kinetics. Wong et al. [18]developed a model for the reaction rate development in

a motored engine working on fuel mixtures of hydrogen,

methane and propane with air in the presence of exhaust gas

recirculation. Metghalchi and Keck [19] measured the laminar

burning velocities of the stoichiometric hydrogenepropaneeair

flames. Huang et al. [20] studied the laminar burning charac-

teristics of propaneehydrogeneair flames.

The concept of multi-fuels engine system is definitely

attractive but it requires a thorough investigation at both

theoretical and experimental level. With this aim in view, this

paper presents a theoretical analysis and experimental results

on theperformanceof a dual fuel diesel engine with hydrogen

and LPG as secondary fuels.

2. Combustion analysis

2.1. Combustion process

In a dual fuel engine, much of the energy is released from the

combustion of gaseous fuel, while a small amount of diesel

liquid fuel provides ignition. It was reported that addition of

hydrogen in gasoline engine not only reduces the mass of

gasoline but also increases its thermal efficiency [21]. A

gaseous fuel induction in hydrogen-diesel dual fuel engine

also gives higher efficiency [9]. However, the combustionprocess in a diesel engine becomes more complex with the

addition of gaseous fuel [22].

Following Liu and Karim [16] the combustion process

within a hydrogen-LPG-diesel operated dual fuel engine can

be visualized as shown in Fig. 1.

A mixture of hydrogen, LPG, and air is inducted into the

dual fuel engine. At the ideal situation near the end of

compression stroke, diesel is injected into the engine cylinder

as a pilot fuel. As a result the combustion space is divided into

three regions before ignition. The rich and unvaporized pilot

diesel at the core of theinjection forms the first region. During

the end of compression stroke, some part of the vaporized

diesel diffuse into the approaching charge of hydro-

geneLPGeair mixture, and forms the second region, which

becomesflammable due to the high temperature and pressure

of the mixture. The gaseous-air charge, which is away from

the injected pilot fuel, forms the third region of lean fuel

character.

The core of the spray on diesel fuel injection, which

contains rich liquefied pilot diesel, constitutes the unburntdiesel fuel region. As reaction rates of gaseous-air charge with

vaporized pilot diesel in flammable region reach at stoichio-

metric condition, the ignition takes place, and forms the

premixed diffusion combustion region. Now the flame prop-

agates through the gaseous-air charge due to its homogeneity

across the flammable region and forms the flame propagation

region of the gaseous fuel. The remainder gaseous-air charge

far away from the injection pilot zone forms the region of

unburnt lean gaseous-air mixture. This lean gaseous-air

mixture is compressed and heated by the combination of the

movement of the piston and flame front. Once the fuel charge

is entrained from the unburnt zone to the burnt zone, its

energy is released at the burnt zone periphery.During the ignition delay period, a part of the injected

fuel mixes with the gaseous fueleair mixture and forms

a combustible mixture. Thus, ignition starts at premixed

zone. At the time of ignition, the premixed combustible

mixture has been considered to be bounded by the lean

limit of combustion at the outer edge of spray and rich limit

near the core. After ignition more fuel from the spray core

becomes combustible because of continuous gaseous

fueleair entrainment. The concentration of reactant mixture

decreases across the flame front, and the temperature

increases as shown in Fig. 2 [23].

Thus, the gaseous fueleair charge within the engine

cylinder undergoes different combustion processes producingvarying temperature and combustion products.

2.2. Spray mixing model

When the liquid fuel at high pressure is injected into

a combustion chamber, which contains high pressure and

high temperature air, it breaks up into fine droplets. The

variation in velocity between injected fuel and gaseous

fueleair causes deceleration of the spray and growth in the

spray width. This results in non uniform distribution of

velocity, temperature and fueleair ratio.

Rich pilot fuel

Reacting Area

Unburnt Gaseous

Fuel air ChargePropagation of Gaseous Fuel-air

Charge

Premixed Diffusion

Region

Unburnt Pilot

fuel

Fig. 1e

Spray zone.

Temperature

Reactant

Un-burntReactant

Zone

Pre-Combustion

zone

Reaction

Zone

Product

Zone

Temperature

Flame Front

Concentration

Fig. 2e

Temperature and concentration profile.

i n t e r n a t i o n a l j o u r n a l o f h y d r o g e n e n e r g y 3 5 ( 2 0 1 0 ) 1 1 9 1 8 e1 1 9 3 1 11919
http://dx.doi.org/10.1016/j.ijhydene.2010.08.039http://dx.doi.org/10.1016/j.ijhydene.2010.08.039http://dx.doi.org/10.1016/j.ijhydene.2010.08.039http://dx.doi.org/10.1016/j.ijhydene.2010.08.039

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In running engine, properties of fuels and air vary with

time, but for simplicity, these values are treated as average

during the fuel injection period. The instantaneous injection

velocity, Vf[16], fuel mass injection rate, mf[16], spray angle of

pilot fuel [24], break-up period of the pilot fuel [24], spray

penetration of pilot fuel [24], shape of spray cross section

[16,25e27] and rate of air entrainment [28] all were considered

in spray mixing model.

2.3. Flame propagation model

An important intrinsic property of a combustible gaseous fuel

and air mixture is its laminar burning velocity [29]. The

laminar burning velocity as a function of equivalence ratio of

LPGeair mixture at a temperature of 298 K and pressure of

1 atm is given as [19].

SuLPG 4:407f3 150:69f2 308:62f 122:7 (1)

and for hydrogeneair mixture combustion, the equation is

[29]

SuH2 51:902f3 394:46f2 835:14f 267:07 (2)

where SuLPG and SuH2 are the laminar burning velocities of LPG

and hydrogen in cm/sec respectively.

The laminar burning velocity of hydrogeneLPGeair

mixture can be predicted by Le Chateliers Rule [20]

Suf;H2

24 1

xH2SuH2 f

xLPG

SuLPGf

35 (3)

where XH2 and XLPG are the mole fraction of hydrogen and LPG

respectively. The turbulent flame velocity may be obtained by

multiplyingSuf;H2 with a factor K, known as turbulent velocityfactor [24].

2.4. Heat transfer

The heat transfer Qexpressed as

Q AShT TS

has the convective heat transfer coefficient, h, based on

Woschni correlation [30], as

h 110$

d0:2p0:8

C1cm C2

VST1P1V1

P P0

!0:8T0:53

!(4)

where d cylinder diameter, p instantaneous cylinder

pressure (bar), T instantaneous mean gas temperature (K ),

TS is surface temperature assumed to be 550 K [31], C1 2.28,

cm mean piston speed, and

AS Apistoncrown A0:3pistontopland

2.5. Combustion modeling

Because of differences in the reaction rates, reaction constants,

and energy release rates of three different kinds of fuels, i.e.,

diesel, LPG and hydrogen, modeling is confined to single zone

combustion. For manypurposes in diesel engine simulation the

assumption of no dissociation with a single zone model is

acceptable. The assumptions reduce the computational time

significantly without a serious loss of accuracy. As the

combustion should always be weakof stoichiometric, this leads

to temperature at which dissociationdoes not havemuch effect

on the thermodynamic performances of the engine. Hence, for

diesel engines, the combustion model is frequently modeled as

a single zone [32].

It is assumed that the mixture comprising of ideal gases(including gaseous fuels and high temperature vapors) obeys

the following[33]:

The mixture as a whole obeys the equation of state

pV MRmolT, where M is the total number of moles of all

kinds, Rmol the universal gas constant, kJ/kg-mol K. The

system is assumed to undergo quasi static processes.

The unburnt mixture of hydrogen, LPG, air and residual

gases forms a homogenous non reactive mixture.

The instantaneous heat transfer coefficient is same for all

metallic surfaces.

The dissociation takes place whenever temperature exceeds

1600 K. Combustion is assumed to occur due to the entrainment of

fuels and air in stoichiometric proportion during premixed

combustion phase [34].

When a fuel CaHbOgNd burns with air in an equivalence

ratio 4 and the products are subjected to temperature and

pressure which attain equilibrium [35], the equation in the

premixed zone (Fig. 1) for the mixture of diesel, LPG, hydrogen

and air may be written as

xCaHb yC3:36H8:72 zH2 aSf

O2 3:76N2 0:044$Ar/n1CO2

n2H2O n3N2 n4H n5O n6N n7H2 n8O2 n9OH n10CO n11NO n12Ar

(5)

where

aS x

a

b

4

5y

z

2; and x y z 1: (5a)

x, y and z are the mole fractions of diesel, LPG and hydrogen

respectively. The chemical composition of fuels is shown in

Table 1.

Consider an elemental time step dt after some time t t1during the combustion process. Let Ri be the symbols for the

coefficients of the constituents at the beginning of the process

of the time step dt, and Pi the symbols for the coefficients of

the constituents at the end of time step dt. Then following

species are present during the beginning of process at time t1,

Table 1 e Average composition of fuels.

Sr. No. Fuel Composition Mass (%)

1. LPG (Equivalent Chemical

Symbol C3.36H8.72)

C2H6 1.0

C3H8 62

C4H10 37

2. Diesel (C14.4H24.9) C (By weight) 84.8

H (By weight) 15.2


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R1CO2 R2H2O R3N2 R4H R5O R6N R7H2 R8O2

R9OH R10CO R11NO R12Ar R13CaHb R14C3:36H8:72

R15H2

A; say

(6)and at time t2 t1 dt,

P1CO2 P2H2O P3N2 P4H P5O P6N P7H2 P8O2 P9OH

P10CO P11NO P12Ar P13CaHb P14C3:36H8:72 P15H2

B; say

(6a)The absolute internal energy E(T ) of the cylinder contents at

time t1 can be expressed as

ET X

MieiT X

Mie0i T (7)

where Mie0i T is the internal energy at absolute zero of the ith

specie in the mixture. Mi is the number of moles of ith specie

in the gas mixture. Now, the specific internal energy ei(T ) may

be expressed as [33],

eiT Rmol

240@Xj5

j1

uijTj

1A T

35 (8)

where Rmol is the universal gas constant.

The coefficients uij ( j 1e5) are referred from [33]. From

Eqs. (7) and (8), the absolute internal energy ER of the mixture

at time t1 can be written as

ER X15i;j0 1

Rie0j0 X15i;j0 1

RiT1j0 (9)

where i 1e15, and j 1e15 implies that 1 CO2, 2 H2O,

3 N2, 4 H, 5 O, 6 N, 7 H2, 8 O2, 9 OH, 10 CO,

11 NO, 12 Ar, 13 CnHm, 14 C3.36H8.72, 15 H2; T1 is thetemperature at time t1, and the absolute internal energy EP at

time t2 as

EP X15i;j0 1

Pie0j0 X15i;j0 1

PiT2j0 (10)

The first law of thermodynamics for the process is

dQ dW dE

where dE EPER change in internal energy during the time

interval dt; dQ and dW are the corresponding heat and work

transfer respectively. Therefore,

dQ dWMPi e0Pi MPi ePi TP

MRi e0Ri MRi eRi TR

(11)

dQ dWMPi e0Pi MRi e0Ri

MPi ePi TP MRi eRi TR

(12)

where MPi e0Pi MRi e0Ri is the heat of reaction. The QVS, at

absolute zero can be predicted from heat of reaction at some

reference temperature TS; QVS is negative at exothermic

reaction. This lower heat of reaction may be replaced by lower

calorific value qvs, which is positive during exothermic

reaction.

Thus QVS qvsThen the generalized form of first law of thermodynamics

for the process during the step dt becomes

dQ dWMPi e0Pi MPi ePi TS

MRi e0Ri MRi eRi TS

dmfqvs

dQh

(13)

where dmf, the mass of fuels, is burnt during time interval dt,

and dQh is the heat transfer to the metallic surfaces.

2.6. Cycle analysis

During the period of ignition delay, the gaseous-air mixture is

entrained into the core of pilot diesel spray and forms pre-

mixed diffusion burn zone, which is assumed to burn stoi-

chiometrically [15]. The fresh mixture of diesel, gaseous fuels

and air is entrained from the surroundings to this burn zone

and its energy is released immediately. The flame propagates

towards the flammable region of the gaseous-air mixture and

forms a propagation burn zone. Hence, the combustion is

assumed to complete by two combustion processes as the

amount of gaseous fuele

airmixture which is entrained duringthe time of ignition delay period along with premixed pilot

diesel is burnt stoichiometrically and instantaneously: a part

is assumed to burn at constant volume process due to pres-

ence of hydrogen and the rest of the diffused pilot fuel and

gaseous fueleair mixture is assumed to burn at constant

pressure process.

The cycle consists of the following processes, depicted in

Fig. 3,

(i) Polytropic compression of the mixture of air, LPG and

hydrogen from A to B,

(ii) Adiabatic instantaneous constant volume combustion

process from B to C,(iii) Adiabatic instantaneous constant pressure combustion

process from C to D,

(iv) Polytropic expansion of the products of the combustion

from D to E; and finally,

(v) Constant volume exhaust process from E to A.

P

V

A

B

C D

E

Fig. 3e

Cycle analysis.


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2.6.1. Polytropic compression of the mixture of air, LPG and

hydrogen

A mixture of air, LPG and hydrogen is inducted into the

cylinder and compressed polytropically from A to B. The

process calculation is based on the first law of thermody-

namics by dividing the stroke volume into large number of

elemental volume changes. Let i and i 1 represent the

states before and after some time step dt during the processAB. For isentropic compression dQ 0, and as there is no

combustion dmf$qvs 0; the first law for the polytropic

compression process, Eq. (13), reduces to

ETi01 ETi0 dW dQh 0 (14)

The work term, dW, equals pdVfor an infinitesimal change.

If the change is sufficiently small the work term can be

approximated by [32]

dWPi0 Pi0 1

2Vi0 1 Vi0 (15)

where Vi1 is given by [23]

Vi1VC

1 12

rC 1h

R 1 cosq

R2 sin2q1=2i

(16)

where R l/a; l length of connecting rod; a length of crank

radius.

Substitution of Eq. (15) into Eq. (14) yields

ETi0 1 ETi0 Pi0 Pi0 1

2Vi0 1 Vi0 0 (17)

As there is no combustion during this process, the pressure

and temperature relation may be written as

Pi0 1 Vi0

Vi0 1

Ti0 1

Ti0 Pi (18)

The first value of (Ti1) is thus obtained as

Ti0 1 Ti

V1V2

k1

where k is polytropic index, while further values of (Ti1) at

different states are

Ti0 1n1 Ti0 1n

fEi0 1n1

M$CVTi01n1

!(19)

where M$CVTi0 1 P

MiCVi Ti0 1 and

fEi0 1n1 ETi0 1 ETi0 Pi0 Pi0 1

2Vi0 1 Vi0 (19a)

Mi is the number of moles of gas i in the mixture; Cv specificheat.

For an ideal cycle, no products of combustion are left as

residuals; the amount of various gases forming the mixture in

the cylinder depends upon overall equivalence ratio4. If the

actual fueleair ratio is (F/A) and the stoichiometric fueleair

ratio is (F/A)ST, and the number of moles of the mixture in the

cylinder at the beginning of the compression stroke at A is

given by Eq. (6), i.e.

R1 R2 R4 to R13 0; R8 aS

foverall;

R3 3:76 R8; R14 y; R15 z;

and since there is no combustion, the number of moles at

state A and B remains the same. The overall equivalence ratio

(4overall) for the mixture of three fuels is given as

foverall

0B@ md(

ma

"mH2

CcH2Ca

st

mLPG

CcLPGCa

st

#)1CA

CdCa

st

Therefore,

foverall 14:3md

ma

34:01mH2 15:57mLPG (20)

where ma, md, mH2 , and mLPG are the mass of air, diesel,

hydrogen, and LPG in kg respectively.

2.6.2. Adiabatic instantaneous constant volume combustion

process

If during the entire combustion, x moles of diesel, y mol ofLPG

and z moles of hydrogen are burnt then total moles of fuels

burnt are M x y z. It is now assumed that only a fraction

Xf of fuels is burnt during the constant volume combustion

process. Thus, the moles of fuels burnt are XfM during process

BC.

Let1 and 2 represent thestates before and after time step dt

within the process BC. Then Eq. (13) is written as

dQ dW f1E

E2T2 E2TS E1T1 E1TS XfM$qvs dQh

(21)

The first value of T2 is obtained from

T2 T1 XfM$qvs

M1$CVT1(22)

Further calculations ofT2 are made from the expression

T2n T2n1f1E

MRi CVT2n1and P2 P1

T2T1

M2M1

(23)

where M1 and M2 are the number of moles of products before

and after combustion, and qVS is the sum of lower calorific

value of diesel, LPG and hydrogen.

2.6.3. Adiabatic instantaneous constant pressure combustion

process

Now, (1 Xf) portion of pilot diesel fuel is burnt during the

combustion at constant pressure. Thus, the mass of total fuels

burnt during this process is (1 Xf)M.Let 1 and 2 represent the state change during the time step

dt within the process CD. Thus, the first law of thermody-

namics from Eq. (13) assuming no heat losses hence dQh 0;

dW P1(V2 V1) becomes

dQ P1V2 V1 h

MP00ie0P00

i MP00

ieP00

iTS

ih

MR00ie0R00

i

MR00ieR00

iTS

i

1 Xf

M$qvs (24)

MRi isthenumberofmolesatthebeginningofprocessisgivenas

Ri RiBC Xf$M; R8 R8BC aS

Xf$M

;R3 3:76$R8; R13 R13BC

Xf$M

$x; R14 R14BC

Xf$M

y;

R15 R15BC

Xf$M

$z;

where R1,R2,.....

R15, are the moles at the start of


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constant pressure combustion process at C.

P1,P2,.....P15, are the moles of products after combus-

tion at constant pressure within the process CD.

While,

Pi Ri M; i 1 to 12 except 3 and 8; P3 R3;

P8

P8BC aSM; P13 P14 P15 0;Hence, M1

PRi; M2

PPi; then Eq. (13) becomes

0 E2T2 E1T1 P1V2 V1

1 Xf

M$qvs dQ (25)

The first values ofT2 are predicted as

T2 T2BC

1 Xf

$M$qvs

M1CPT1; V2

M2M1

T2T1

$V1;

Further values ofT2 are given as

T2n T2n1f2En1

M2CPT2n1; where CP CVT2 Rmol; f2En1

E2T2 E1T1 P1V2 V1

1 Xf

M$qvs dQ

2.6.4. Polytropic expansion process from D to E

The composition of mixture within the cylinder after

combustion remains the same during the expansion process

DE. Thus, the calculations of the state changes are same as

those of polytropic compression process from A to B with the

number of moles equal to the molesat the endof theadiabatic

instantaneous constant pressure combustion process, D.

The logic for compression and expansion processes is the

same with change in the nomenclature of their respective

variables.

The cycle closes at A by considering constant volume

process from E to A.

The thermal efficiency of the cycle is predicted as [36]

hThermal 1 TE TA

TC TB kTD TC(26)

3. Heat release rate

The ignition delay in millisecond (ms) is given as [37]

sid 0:01$P2:5f1:04exp

6000

Tg

(27)

where P and Tg are the mean cylinder pressure (atm) and

charge temperature (K) during the ignition delay period and fis the equivalence ratio of the fuel vapor-air mixture. Miya-

moto et al. [38] showed two Wiebes functions for heat release

as

dQ

dq a1

QPqP

MP 1$

q

qP

MPexp

a1

q

qP

MP1!

a2Qdqd

Md 1$

q

qd

Mdexp

a2

q

qP

Md1!(28)

where the subscripts, p and d refer to the premixed and

diffusive combustion parts, respectively, Qp and Qd are the

quantity of premixed and diffusion combustion parameter

respectively, Mp and Md are shape factors, qp and qd are the

duration of energy release, a1 and a2 are constants.

QP b$f1t

mdqvsd mLPGqvsLPG mH2 qvsH2

;

Qd 1 bf2t

mdqvsd mLPGqvsLPG mH2 qvsH2

;

4. Mean gas temperature

The mean gas temperature, which is required for the calcu-

lation of heat transfer equation, is predicted by considering

polytropic process, pVk constant

T2 T1

V1V2

k1 T1

p2p1

k1=k(29)

Therefore, at a known reference position, i.e., at inlet valve

closure or crank angle at the start of injection:

pref$Vref k$R$Tref

By assumingk (polytropic index) and R to be constant;

Tcalc pcalc$Vcalc$Tref

pref$Vref(30)

Based on the above equations, a computer programme was

developed to analyze the theoretical results. The main inputs

to the model are: engine geometry, engine speed, mass flow

rate of diesel and gaseous fuels, density of fuels, polytropic

index (calculatedfrom logP vs logVcurve), injection pressure,

nozzle orifice diameter, inlet temperature and pressure.

The gaseous-air mixture properties were computed from

inlet valve closer to start of injection with respect to variation

in cylinder volume. The inlet temperature and pressure were

considered to be initial conditions at inlet valve closer. The

fuel preparation was considered from the end of the brake-upspray penetration length. The flow chart of the computer

model is shown in Fig. 4, where the suffices ivc, inj, id, inj. dur.,

evo, p and d represent inlet valve closer, injection timing,

ignition delay, injection duration, exhaust valve opening,

premixed combustion phase and diffusion combustion phase

respectively. The incremental step of 0.1 and 1.0 crank angle

is exhibited in the flow chart (Fig. 4).

5. Experimental

A test diesel engine setup was developed to carry out the study

on dual fuel engines.A four stroke compression ignition engine, model Ashok

Leyland ALUWO 4CT,turbocharged with inter-cooler andgen-

set was used for the experimental investigation which is

designated as Engine A. Table 2 shows the engine geometry

and operating parameter for the present work. The diesel

engine was modified to work on dual fuel mode by attaching

a hydrogen and LPG gas cylinder connection to the intake

manifold through flame traps, andmass flowmeters, followed

byaonewaynonreturnvalveandcommonflamearrestor.The

engine was coupled to a D.C. generator of 62.5 kW. Theloadon

the engine was varied by introducing five water pump and

twelve 3 kW industrial water heaters in a set of four each. The

engine was run at constant speed of 1500 RPM. The amount of


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pilot diesel fuel was automatically controlled with the help of

governor while the flow of gaseous fuels was controlled

manually. The predetermined amount of gaseous fuel wasinducted into the intake manifold through the gaseous fuel

supply system. High precision optical crank angle encoder

(makerof Kistler) wasused to determine the location of thetop

dead center (TDC) position precisely and then correlate

cylinder pressure data with cylinder volume.

Since the engine was modified to run simultaneously with

liquid and gaseous fuels, two separate fuel induction, meter-

ing and measurement systems were used. The liquid fuel

measurement system for the test rig was based on the gravi-

metric principle. For this purpose a 1000 cc glass bulb appa-

ratus with a control valve was placed in between the fuel tank

and engine fuel supply system. For gaseous fuels two separate

flame traps and mass flow meters were used. The percentage

uncertainty of the measuring instruments is as follows: mass

flow rate of gaseous fuels 4%, mass flow rate of diesel 4%,

load 3% and speed 3 revolutions.The results obtained were compared with the results of

Saravanan et al. [9] on the engine, designated as Engine B in

Table 2.

The experiments were performed on the Engine A under

the following four conditions.

(i) Case I: Engine run on diesel only.

(ii) Case II: Engine run on diesel as pilot fuel and hydrogen as

secondary fuel.

(iii) Case III: Engine run on diesel as pilot fuel and LPG as

secondary fuel.

(iv) Case IV: Engine run on diesel as pilot fuel and LPG plus

hydrogen as secondary fuel.

Fig. 4 e Flow chart for computer model.


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The theoretical results reported in the present paper are

compared with the experimental results of the engine run at

80% load condition for all the four cases described above. The

constants and coefficients, which are used in semi-empirical

equations, are shown in Table 3. The various mass distribu-

tions ofthe fuelsin the four cases of experimentsare shown in

Table 4. Further, for the analysis of brake thermal efficiency

the experiments were conducted for all the four cases (Cases

IeIV) at 9, 16, 40, 64 and 80% of load condition.

6. Result and discussion

The rated speed 1500 rpm, injection pressure 260 bar, injection

timing16BTDC andthe total amountof fuelconsumedat selected

loads were taken for model calibration. The experimental data

analyzed and presented here are based on average values of 100

cycles. The shape factor for heat release profile, convective heat

transfer and gaseous fueleair entrainment rate were determined

by measured cylinder pressure and heat release profile.

Selecting the constant a3 of the air entrainment equation

from 0.016 to 0.11 [28], the rate of gaseous fueleair mixture

entrainment into the spray zone was simulated. The physical

processes such as injection of fuel, atomization of fuel into

droplets, fuel spray penetration, vaporization of fuel and

mixing of diesel fuel with air and gaseous fuels in spray zone,which are collectively known as preparation of fuel, were

evaluated from spray mixing model [16,24e28]. Based on the

combustion modeldescribedabove underSection 2.5 andcycle

analysis under Section 2.6, the temperature at points B, C, D

andE were predicted. Theoveralllaminarflame velocityduring

every phase of combustion was predicted by considering local

Table 2 e Engine specifications.

Sr. No. Engine Aspecification

Engine Bspecification [9]

1. Make and model Ashok Leyland ALU

WO4CT Turbocharged,

inter-cooler, Gen-Set

Kirloskar, AV1 Make

2. General details Four stroke, compression ignition,constant speed, vertical, water-cooled,

direct injection, turbo charger, Intercooler,

Gen-Set

Four stroke, compression ignition,constant Speed, vertical, water-cooled,

direct injection

3. No. of cylinder 4 1

4. Bore (mm) 104 80

5. Stroke (mm) 113 110

6. Rated Speed (rpm) 1500 1500

7. Swept volume (cc) 3839.67 553

8. Clearance volume (cc) 84.90 36.87

9. Compression ratio 17.5:1 16.5:1

10. Injection pressure (bar) 260 205

11. Injection timing (BTDC) 16 23

12. Rated power kW at

1500 rpm

62.5 3.7

13. Inlet Pressure (bar) 1.06 114. Inlet temperature (K) 313 e

15. Nozzle diameter (mm) 0.285 e

Table 3 e Constants/coefficient for semi-empirical equations of Engine A and B.

Case I II III IV

Coefficient for heat transfer for

Engine A & B

C1 during

compression & expansion

2.28 2.28 2.28 2.28

C2 during

Combustion process (m/s K)

3.34 103 3.34 103 3.34 103 3.34 103

Constant (a3) for gaseous

fueleair entrainment

Engine A 0.0915 0.079 0.087 0.095

Engine B e 0.088 e e

Polytropic index k Engine A 1.324 1.32 1.315 1.3


Constant for heat release Engine A q0p 5 5 5 5

q0d 13 8 5 5

Engine B q0p e 7 e e

q0d e 7 e e

Constant k3 Engine A 0.87 0.94 0.97 0.956


Shape factor for heat release rate Engine A mp 4.0 4.06 4.09 4.03

md 1.5 1.57 &1.54

(3rd phase)

1.57 &1.51

(3rd phase)

1.58


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mass fraction of remaining LPG and hydrogen in unburnt

gaseous fueleair mixture. It was further assumed that the

flame propagates in a conicalshape into combustionchamber.

By considering the experimental data available during the

study, a turbulent flame velocity factor of 1.55 was obtained.

The standard range of the turbulent flame velocity factor lies

between 1.2 and 1.6 [24].

In the simulation, an incremental step of 0.1 crank angles

(CA) yielded results within a variation of 3e4% between the

theoretical and experimental values.

6.1. Ignition delay

Ignition delay period was found to be 9 , 13, 11 and 10 for

Cases IeIV respectively. Ignition delay period increased

slightly with the addition of hydrogen and LPG [2].Thismaybe

because the addition of hydrogen or LPG or a mixture of LPG

and hydrogen in the charge reduces the air intake and hence

oxygen in the cylinder. The other reason for increase in delay

period is formation of intermediate compounds during

compression due to partial oxidation of gaseous fueleair

mixture. This is caused by the loss of the very reactive OH

radical in the reaction with molecular hydrogen, which givesless reactive species that are not capable of accelerating the

reaction at the same rate as that of OH radicals [10]. At higher

concentration of hydrogen plus LPG in the mixture, ignition

delay decreases due to addition of significant amounts of

energy and species. The measured and predicted ignition

delay period for all the four cases is shown in Table 5.

6.2. Cylinder pressure and mean gas temperature

The predicted and experimental results of variation in

cylinder pressure with crank angle (CA) under addition of

hydrogen, LPG and mixture of hydrogen and LPG are shown in

Fig. 5. The pilot fuel quantity varied due to addition of gaseousfuel into the cylinder. The quantity of pilot fuel injected affects

the mean diameter of the spray. The peak pressures for Cases

IeIV are 80.70, 68.37, 80.73, and 84.41 bar at 367, 367, 365,

and366 CA respectively. The addition of gaseous fuels in Case

IV increases peak pressure due to higher energy release during

their combustion. The presence of turbocharger with inter-

cooler increases charge/air velocity and densityof the gaseous

charge in the cylinder [39]. This higher air velocity and

gaseous-air entrainment lead to increase in rate of evapora-

tion of the liquid fuel and gives higher rate of heat release

resulting in higher peak pressure in the cylinder.

The predicted peak cylinder pressure was found to be

76.66, 62.98, 75.55 and 78.79 bar for Cases IeIV respectively. A

number of runs were performed for different values of the

diffusion burning factor k3. This value ofk3 varies from 0.055

to 1.10 [39]. The satisfactory results were obtained within

5e6% variation between present combustion model and

experimental results, with k3 equal to 0.87, 0.94, 0.97 and 0.956

for Cases IeIV respectively. The measured and predicted

cylinder peak pressure for all the four cases is shown in

Table 6.

Fig. 6 shows the variation of mean gas temperature with

crank angle for the experimental and predicted results for

Case IV.A dropin temperature(around 20 K) before the start of

combustion is observed in all the four cases, which is due to

the vaporization of pilot fuel. A sharp rise in temperature

during the combustion is observed in Case I as compared to

Cases II to IV for other gaseous fuels. The measured and pre-

dicted mean gas temperature for Cases IeIV is shown in

Table 7. This minor discrepancy of about 4% in the predicted

and measured values may be due to unable to predict correct

air/charge velocity inside the combustion chamber, turbu-

lence factor or gaseous fueleair entrainment into the pilot

diesel spray zone.

6.3. Brake thermal efficiency

Fig. 7 shows experimental result for the brake thermal effi-

ciency at different load conditions. The brake thermal effi-

ciency of the engine when working on dual fuels (Cases IIeIV)

is found to be 17.34%, 18.47% and 18.25% respectively, as

compared to Case I of 19.57% at 9% load. This may be due to

lower charge temperature at the end of compression process,

Table 5 e Ignition delay.

Case I II III IV

Ignition Delay (CA) Measured 9 13 11 10

Predicted 7 10 8 9

Table 4 e Mass of different fuels.

Case no. Mass of dieselkg/min

Mass of hydrogenkg/min

Mass of LPGkg/min

I 0.1434 e e

II 1.148 104 0.04897 e

III 1.066 104 e 0.143072

IV 6.56 105

6.396 103

0.1055Engine B [9] 2.95 103 0.6285 (7.5 lpm) e

0 100 200 300 400 500 600 700

0

10

20

30

40

50

60

70

80

90

100

Pressure

(ba

r)

Crank Angle (Degree)

Measured

Predicted

Cylinder Pressure Curve (Diesel + LPG + H2)

Fig. 5 e Cylinder pressure (bar) vs crank angle (CA) for

diesel and mixture of LPG and hydrogen.


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low flame velocity of the lean gaseous fuels-air mixture andenough time available in transferring heat to the adjacent

cylinder walls. While in Case I at light load, the penetration of

spray may be such that it does not reach the cylinder walls

and combustion is confined to the piston combustion

chamber (bowl) only. The combustion zone is surrounded by

air that works as semi-insulator in between the burned gases

and the cylinder walls. This may reduce the heat losses to the

cylinder walls and thereby increase the thermal efficiency in

Case I operation as compared to Cases IIeIV [10].

The brake thermal efficiency of dual fuel engine when

operating on Cases IIeIV is still lower up to 60% of load i.e.,

23.22%, 24.20% and 25.47% respectively as compared to 25.55%

in Case I. The addition of hydrogen or/and LPG causes higherdiesel consumption. Hydrogen and LPG have higher flame

velocity and diffusivity than diesel fuel; therefore, these

gaseous fuels consume most part of the oxygen from the

entrainedair during main part of combustion. Hence, a part of

injected diesel goes into the exhaust without taking part in

combustion. This leads to lower brake thermal efficiency.

At higher load i.e., at 80% load, the brake thermal effi-

ciencies were 32.64%, 31.3% and 33.56% for the Cases II, III and

IV respectively as compared to 31.5% of Case I, due to presence

of rich gaseous fueleair mixture. The higher flame velocity of

these gaseous fuels takes less time to reach the cylinder walls

and hence less time is available for heat transfer to the

cylinder wall. Therefore, heat losses are less, resulting inhigher brake thermal efficiencies. The brake thermal effi-

ciency of Case IV (mixture of hydrogen and LPG) was more

than brake thermal efficiency of Case II and III, because it was

improved by the presence of small amount of LPG in the

mixture. The LPG reduces the laminar burning velocity of

hydrogen and suppresses the propensity of onset of both

diffusional-thermal and hydrodynamic cellular instabilities in

hydrogen-air flames. It also retards the reaction intensity and

increases the critical radius [38]. While the LPG has unstable

flame at lean gaseous fueleair mixture, but due to the pres-

ence of hydrogen, the flame becomes more stable and hence

results in higher brake thermal efficiency.

The theoretically predicted brake thermal efficiency for

Cases II to IV of Engine A were found to be 31.66%, 27.97%, and

32.3% respectively as compared to 31.03% of Case I at 80%load.Similarly, by present combustion modeling the brake thermal

efficiency of Engine B [9] was predicted on dieselehydrogen

dual fuel mode and was found to be 14.83% as compared to

their measured values of 15.29% at 80% load.

The predicted brake thermal efficiency was 3e4% less than

the measured brake thermal efficiency due to over prediction

of the heat transfer and frictional losses. The assumption that

the process is quasi static is not true in actual sense. As the

liquid fuel is injected into the atmosphere of gaseous fueleair

mixture, it gets vaporized and mixes with the gaseous fuels

that makes non-uniform fueleair ratio in the combustion

zone. The polytropic index k also varies during combustion

process, while in the present combustion model it wasassumed to be constant. The comparison between measured

and predicted brake thermal efficiency at 80% load condition

in all the four cases with respect to Engine A and B are shown

in Table 8.

Table 6 e Measured and predicted cylinder peak pressure.

Case I II III IV

Measured cylinder peak pressure (bar) 80.49 68.37 79.11 83.82

Predicted cylinder peak pressure (bar) 76.66 62.98 75.55 78.79

300 320 340 360 380 400 420 440 460

0

300

600

900

1200

1500

1800

Temperature(K)


Measured

Predicted

Mean Gas Temperature (Diesel + LPG + H2)

Fig. 6 e Mean gas temperature (K) vs crank angle (CA) for

diesel and mixture of LPG and hydrogen.

Table 7 e Measured and predicted cylinder meantemperature.

Case I II III IV

Measured cylinder

mean temperature

(K)

1863.42 at

382 CA

1751.08 at

372 CA

1828.94 at

381 CA

1678.26 at

378 CA

Predicted cylindermean temperature

(K)

1798.13 at378 CA 1698.57 at365 CA 1746.64 at374 CA 1619.39 at370 CA

Fig. 7 e Brake thermal efficiency (h) vs Load (%) with

different gaseous fuels subst.


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6.4. Net heat release analysis

Fig. 8aed, show predicted and measured net heat release rate

with crank angle of Cases I to IV. The net heat release rate of

Case I (Fig. 8a) by first phase of combustion is 94.26 J/CA at

TDC. It shows two phases of combustion. The first phase is

due to premixed combustion while second phase by diffused

combustion of diesel fuel, while Cases IIeIV (Fig. 8bed) show

one more phase of combustion. In dual fuel operation, heat

release is mainly due to three phases of combustion; first, by

premixed burning of pilot diesel fuel and combustion of part

of gaseous fueleair mixture that is entrained during the

ignition delay period. In the second phase of combustion, it is

due to auto ignition of gaseous-air mixture in the close vicinity

of pilot spray zone and diffusive burning of remaining pilot

diesel fuel. In the third phase of combustion, heat is released

by flame propagation from spray zone into the gaseous

fuelseair mixture [12].

At lean gaseous fueleair mixture as in Case II, Fig. 8b shows

lower peak of first phase of combustion 60.57 J/CA at 355 CAthan Case I because heat release is mainly controlled by pre-

mixed burning of part or complete pilot diesel fuel plus small

amount of gaseous fuel entrained into the spray zone. The net

heat released rates in second and third phases of combustion

are 44.86 J/CA at 361 CA and 29.9 J/CA at 366 CA respectively.

The heat released during compression process is due to

preignition chemical reaction of hydrogen with the air.

In Case III, Fig. 8c the heat release rate during compression

process is due to partial oxidation of LPG gas. The heat release

rate during first phase of combustion is 77.57 J/CA at 356 CA.

This is due to energy released by part of pilot diesel fuel

accumulated during ignition delay period and part of LPG gas

entrained in to the spray zone. The net heat release rate by

Table 8 e Measured and predicted brake thermalefficiency.

Case I II III IV

Brake thermal efficiency

Engine A

Measured 31.5 32.64 31.3 33.82

Predicted 31.03 31.66 30.35 32.30

Brake thermal efficiency

Engine B

Measured e 15.29 e e

Predictede

14.83e e

340 350 360 370

0

20

40

60

80

100

Measured

Predicted

Net Heat Release Rate (Diesel)

N

etHeatReleaseRate(J/CA)


340 350 360 370

-20

0

20

40

60

80

Net Heat Release Rate (Diesel+H2)

III Phase of

Combustion

II Phase of

CombustionI Phase of

Combustion

NetHeatR

eleaseRate(J/CA)


Measured

Predicted

340 350 360 370

-20

0

20

40

60

80

Net Heat Release Rate (Diesel + LPG)

III Phase ofCombustion

II Phase of

Combustion

I Phase of

Combustion

NetHeatReleaseRa

te(J/CA)


Measured

Predicted

340 350 360 370

-20

0

20

40

60

80

Net Heat Release Rate (Diesel + LPG + H2)

I Phase of

Combustion

II Phase of

Combustion

III Phase of

Combustion

NetHeatReleaseRate(J/CA)


Measured

Predicted

b

dc

a

Fig. 8 e a. Net heat release rate (J/CA) vs crank angle (CA) for diesel. b. Net heat release rate (J/CA) vs crank angle (CA) for

diesel and hydrogen. c. Net heat release rate (J/CA) vs crank angle (CA) for diesel and LPG. d. Net heat release rate (J/CA) vs

crank angle (

CA) for diesel and mixture of LPG and hydrogen.


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second phase of combustion i.e., diffusion controlled combus-

tion is 39 J/CA at TDC. It was observed that theentire pilotdiesel

quantity does not burn in the first phase of combustion while

some of the pilot fuel quantity along with the induced gaseous

fuel burnt in thesecond phaseof combustion and hence second

peakis distinctin heat release curve.The endof secondphaseof

combustion is shown as second dip in the heat release curve.

The third peak of 35 J/CA was observed at 364 CA, which ispossibly due to energy release by flame propagation of

remaining unburnt gaseous fueleair mixture present in the

combustion chamber. This third peak may also be due to the

presence of turbocharger that makes higher density of

unburned gaseous fueleair mixture. The negative loop during

this third phase of combustion during expansion process indi-

cates larger heat transfer to the cylinder wall and combustion

chamber. The nature of the curve is very much similar to the

findings of Poonia et al. [2].

The net heat release curve of Case IV is shown in Fig. 8d. It

shows two distinct peaks of combustion phases. The heat

release rates in first and second phases of combustion are

73.88 J/CA at 355 CA and 69.04 J/CA at 360 CA respectively.This indicates that due to presence of LPG in the mixture of

hydrogen and air, the propagating flame probably becomes

more stable and significant amount of these fuels are burnt

along with diffusion burning of pilot diesel fuel, during second

phase of combustion.

At higher load (rich gaseous fueleair mixture), both the

phases of combustion are equally important. The peak of first

phase of combustion was affected due to higher intake

gaseous fueleair mixture. It shows that heat release during

premixed combustion phase depends not only on the amount

of pilot diesel fuel but also on the amount of gaseous fuel

entrainment. The end of first phase of combustion is shown

by first dip in the rate of heat release. The second peak is dueto diffusion phase of combustion. It appears that gaseous

fueleair mixture nearer to pilot spray auto ignites at higher

temperature. The remaining combustion is the third phase of

combustion due to flame propagation.

For Cases II, III and IV, total combustion duration was

observed to be increased as compared to Case I. This may be

due to the entrainment of gaseous fuels in the pilot spray zone

and flame propagation [12]. The magnitude of first phase of

combustion decreased in Cases IIeIV (Fig. 8bed), which may

be due to reduction in pilot diesel fuel quantity during pre-

mixed combustion phase and due to sluggish combustion [12].

In all the above three cases, net heat release rate peak occurs

earlier than Case I operation, due to gaseous fueleair mixture

surrounding the pilot fuel, which promoted faster initialcombustion rate, and further caused rise in temperature.

Hence, reaction zone widened and more and more gaseous

fuel then burned in the second phase of combustion [22].

In diesel engine operation, after the injection of pilot diesel

fuel, net heat release rate is reduced due to vaporization of

fuel, prior to combustion. The start of combustion is often

defined as when the net heat release rate becomes positive

[32]. The netheatrelease rate is positive before injection of the

pilot fuel as shown in Fig. 8bed (Cases IIeIV) due to heat

release by preignition energy reaction of gaseous fuels during

compression process. The amount of heat release depends

upon the type of gaseous fuel and its concentration. The

negative net heat release rate during expansion processindicates that there was heat transfer to the cylinder walls,

and the ignition was close to the minimum of net heat release

rate. The heat transfer increased as the temperature and

flame velocity increased during combustion. But due to limi-

tation of present model it is difficult to predict correctnet heat

release rate by preignition reaction of these gaseous fuels

during compression process.

Fig. 9 shows the mass fraction burnt curve for Cases IeIV.

Ignition delay from Fig. 9 for Cases IeIV is found to be 9, 13,

11 and 10 respectively. It is evident from Fig. 9 that about

34%, 27% and 31% of the total mass of fuels is burnt within 1

crank angles for the Cases II, III and IV respectively. Hence the

assumption of an additional constant volume process appearsreasonably justified.

7. Conclusions

The combustion models have been developed to calculate

variations in pressure and net heat release rate with respectto

crank angle. These theoretical models can also be used to

predict ignition delay, brake thermal efficiency, mean gas

temperature, and net heat transfer rate.

These results have been verified by performing experi-

ments on 4 cylinder turbocharged, intercooled with 62.5 kWgen-set diesel engine and with the results on single cylinder

diesel engine obtained by other researchers. The experiments

were performed to measure brake thermal efficiency at

different load conditions, pressure and net heat release rate

with respect to crank angle on following four cases.

(i) Case I: Engine runs on diesel only.

(ii) Case II: Engine runs on diesel as pilot fuel and hydrogen as

secondary fuel.

(iii) Case III: Engine runs on diesel as pilot fuel and LPG as

secondary fuel.

(iv) Case IV: Engine runs on diesel as pilot fuel and LPG plus

hydrogen as secondary fuel.

350 355 360 365 370

0

10

20

30

40

50

60

70

80

90

Mass Fraction Burnt Curve

Crank Angle

Diesel

Diesel + H2

Diesel + LPG

Diesel + H2+ LPG

Fig. 9e

Mass fraction burnt (%) vs crank angle (CA).


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A reasonable agreement between the predicted and

experimental results reveals that this model gives quantita-

tive and qualitative realistic prediction of in-cylinder

processes and engine performances during combustion.

The predicted brake thermal efficiency is found to be 3e4%

less than the experimentally measured brake thermal effi-

ciency due to consideration of quasi static processes.

The predicted maximum pressure rise and net heat releaserate are 6e7% less than experimental results due to under

prediction of charge/air velocity, turbulence and may be due

to over prediction of heat transfer to the cylinder walls and

combustion chamber.

Acknowledgement

The authors are grateful to Professor Pramod S. Mehta,

Internal Combustion Engine Laboratory, Indian Institute of

Technology, Madras, Chennai and Dr. Bimal Kumar Mishra,

Department of Applied Mathematics, Birla Institute of Tech-nology, Mesra, Ranchi, for inspiration and helpful discussion.

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