di dual fuel
TRANSCRIPT
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SAE TECHNICALPAPER SERIES 2000-01-0286
Development of a Simulation Model for DirectInjection Dual Fuel Diesel-Natural Gas Engines
D. T. Hountalas and R. G. PapagiannakisNational Technical University of Athens
Reprinted From: Vehicle and Engine Systems Models(SP1527)
SAE 2000 World CongressDetroit, Michigan
March 6-9, 2000
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ISSN 0148-7191Copyright 2000 Society of Automotive Engineers, Inc.
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2000-01-0286
Development of a Simulation Model for Direct Injection Dual
Fuel Diesel-Natural Gas Engines
D. T. Hountalas and R. G. PapagiannakisNational Technical University of Athens
Copyright 2000 Society of Automotive Engineers, Inc.
ABSTRACT
During the last years a great deal of effort has beenmade for the reduction of pollutant emissions from directinjection Diesel Engines. Towards these efforts engineershave proposed various solutions, one of which is the use
of gaseous fuels as a supplement for liquid diesel fuel.These engines are referred to as dual combustionengines i.e. they use conventional diesel fuel andgaseous fuel as well. The ignition of the gaseous fuel isaccomplished through the liquid fuel, which is auto-ignited in the same way as in common diesel engines.One of the fuels used is natural gas, which has arelatively high auto-ignition temperature. This isextremely important since the CR of most conventionaldiesel engines can be maintained. In these engines thereleased energy is produced partially from thecombustion of natural gas and from the combustion ofliquid diesel fuel. The aim for the usage of dual-fuel
combustion systems is mainly to reduce the particulateemissions (soot) by replacing diesel fuel with natural gaspartially or entirely. For this reason in the present workare given preliminary results of a theoretical investigationusing a simulation model developed for dual fuel engines.The model is a two-zone combustion one, taking intoaccount, on a zonal basis, details of diesel fuel sprayformation and the mixing with the surrounding gas, whichis a mixture of air and natural gas. The main differencefrom a conventional diesel engine combustion system isthat the natural gas is already mixed and ready forcombustion. The natural gas burning initiates after theignition of the diesel fuel and its rate depends on the rate
of entrainment of surrounding gas inside the fuel jetformed. A soot model has been used to estimate theformation of soot while a detailed equilibrium model hasbeen used to determine the concentration of chemicalspecies. For nitric oxide the extended Zeldovichmechanism is used. The model is applied on a singlecylinder test engine located at the authors laboratory atvarious operating conditions of the engine. The amountof liquid fuel supplemented by natural gas has beenvaried and its affect on engine performance andemissions has been examined. From these preliminaryresults it is revealed a serious effect on the heat release
rate inside the engine cylinder and a reduction oparticulate emissions when compared to experimentadata obtained from the engine using diesel fuel only.
INTRODUCTION
The use of alternative gaseous fuels in engines for theproduction of power has been increasing worldwide. Thishas been prompted by the cleaner nature of theircombustion compared to conventional liquid fuels as welas their relative increased availability at attractive pricesDiesel engines can be made to operate on gaseous fuelsefficiently [1-4]1*. These engines usually have thegaseous fuel mixed with air at the start of thecompression stroke. This mixture does not auto-ignitedue to its higher self ignition temperature. The resultingmixture after the compression stroke is ignited throughthe ignition of the amount of diesel fuel injected. Diesefuel can auto-ignite readily creating ignition sources for
the surrounding gaseous fuel mixture. Most current duafuel engines are made to operate either on gaseous fuelswith diesel ignition or only on liquid fuel injection asnormal diesel engines [1-4].
A computer based mathematical model can provide anadequate way for describing details of the complicatedmixing, combustion and pollutants emission processes inthese engines. The present contribution describes amodel that simulates dual-fuel combustion using a twozone approach to simulate the combustion of premixedgas/air charge.
Its main purpose is to describe the operation of existing
diesel engines, where part of the liquid fuel is replaced bygaseous fuel. These engines are usually referred to asfumigated ones and the replacement of liquid fuel withgaseous aims mainly to the reduction of soot emissions.
Furthermore due to the partial replacement of liquid fuewith gaseous one no specific modifications are requiredand thus the technique can be applied on existingengines.
1. Numbers in brackets designate references at the end of
the paper.
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The main aim of the present model is to estimate, apartfrom power and efficiency, the concentration of sootemissions and nitrogen oxides at the engine exhaust(NO), by replacing diesel fuel with natural gas partially atvarious amounts ranging from 0-40%. For this reasonpreliminary results from the modeling are provided forvarious engine operating conditions. To validate themodel under normal diesel operations and to comparethe findings under dual-fuel operation an extendedexperimental is conducted on a single cylinder testengine. From the analysis of results obtained it isrevealed that the simulation model developed predictsadequately engine performance and pollutants undernormal diesel engine operations. Furthermore comparingthe findings for dual-fuel operation with the standarddiesel one, a serious effect of the gaseous fuel isobserved on engine performance and soot, NOemissions. To validate the findings of the dual-fuelmodeling an experimental investigation is currently underway and results will be presented in the near future.
GENERAL DESCRIPTION OF THE MODEL
The cylinder charge at the start of compression stroke isa homogeneous mixture of gaseous fuel and air, that hasdifferent thermodynamic properties from the air charge innormal diesel engines. We replace a part of liquid fuelwith gaseous fuel in order to have the same poweroutput.
For combustion a two-zone model is considered [5]. Ineach zone there is a uniformity in space of pressure,temperature and composition at each instant of time,neglecting heat transfer between the zones. The firstzone consists of air/gaseous fuel (unburned zone) and
the second zone consists of combustion productsunburned mixture and evaporated liquid fuel (burningzone). These zones are separated by the area of theconical jet, which is formed during the injection of thediesel fuel [6]. Steady state jet theory, including wallimpingement, is used to describe the fuel-air mixingprocess. Ignition of the gaseous fuel, which is alreadymixed and ready for combustion, occurs as a result of thediesel jet auto-ignition after the ignition delay period ofthe diesel fuel [4]. Due to the lean nature of thesurrounding mixture no flame is considered, i.e. thecombustion rate of gaseous fuel depends on itsentrainment rate inside the burning zone. Of course this
consideration will have to be modified in the case ofnormal dual-fuel engines where the liquid fuel is a verysmall percentage of the total and used for the ignition ofthe main charge.
For the heat transfer calculation between the charge andcylinder wall, the Annand formula is employed [7]. Forprediction of soot emissions, the amount of net soot is
calculated considering the difference between the ratesof soot formation and soot oxidation inside the burningzone [6-14].
Dissociation of combustion products is taken into accounby incorporating the Vickland et al [15] method includingeleven species. For the formation of nitric oxides theextended Zeldovich chain reaction mechanism isconsidered [6-10,13-16,19].
The liquid fuel used is dodecane (C12
H26
) representingwidely the commercial diesel fuel and the gas fuel used isa mixture of methane (CH4) and propane (C3H8) inproportion (90%-10%) in (v/v), which is a typicarepresentative of natural gas fuel.
MATHEMATICAL TREATMENT
CONSERVATION OF ENERGY As already statedinside the combustion chamber there are considered twozones, a burning one consisting of air, evaporated fuelgaseous fuel plus burned products and an unburned oneconsisting of air and gaseous fuel. Each zone posses its
own temperature and composition. The pressure is thesame for the both zones. The first low equation for eachzone may be written as [6,9]
(1
where dmfbhf is the total enthalpy addition to theburning zone from the liquid diesel fuel and dmbuhbu isthe total enthalpy addition to the burning zone from theunburned zone due the entrainment rate of unburnedmixture. The internal energies and enthalpies in the
above equation considered are total, so that the heat ofcombustion is taken into account implicitly. Thetemperature, pressure and volume of each zone isobtained from the integrating of a set of three ordinaryfirst order differential equations with unknowns Tb, Tuand P. These equations have been derived after somemathematical elaboration from the first law and perfecgas state equations for both zones and the volumebalance. The system of the three equations has asfollows
(2
(3
( ) ffbbubu
u,bu,bu,b
hdmhdm
dVPdQdU
+
+=
( )
puu
uuuuuuuuuu
Cm
/dmudmhdPVdmTRdQdT ++=
pbb
bspeciesk
kkbbbfevfev
ububbbbbbbb
b
C/m
udymdmudmh
dmhdPVdRTmdmTRdQ
dT
+
++
= =
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(4)
The values of Tb, Tu and P must satisfy thecontinuity equation and volume balance as well as, that
(5)
HEAT TRANSFER MODEL At each time step, wecalculate the total area of the cylinder as
(6)
Once the area is known and employing the Annandexpression [14], we have
(7)
where ,b,c are constants and is the thermalconductivity. The only problem existing is the applicationof the above equation during the combustion stroke,
since the burning zone is not entirely in contact with thecylinder wall surfaces. [8] For this reason a bulk averagetemperature of both zones is used as follows
(8)
The thermal conductivity and dynamic viscosity arecalculated using this bulk average temperature. The totalheat exchange rate is then distributed among the twozones according to their mass, temperature and specific
heat capacity as follows
(9)
ESTIMATION OF INJECTION RATE The volume flowrate of diesel fuel and its injection rate are calculatedusing the following equations :
(10
where Pnoz is the pressure difference across theinjector hole as
(11
where f is the liquid fuel density, dnoz is the diameteof the nozzle and CDinj is the nozzle dischargecoefficient. The fuel line pressure Pfl is an input to theprogram together with the dynamic injection timing. Thetotal mass of injected liquid fuel per cylinder cycle isdetermined from the integration of EQ. (10) [16].
MASS ENTRAINMENT RATE After initiation of fueinjection the burning zone begins to form and penetratesinto the combustion chamber. The present model, whilebased on simplified relations, represents the abovemechanism with satisfying accuracy. For jet penetration
the proposal of Hiroyasu et al [13] has been adopted.
Inside the jet we assume uniformity of densitytemperature and composition. The resulting fuel jet isconsidered to have the shape of a cone. The volume ofunburned mixture entrained by the jet is derived from itsvolume change. The jet is considered to be steady andconsisting of a homogenous gas mixture and itspenetration length is given as [10,13]
(12
From the previous equation we obtain, after derivationwith respect to time t, the jet velocity as
(13
If S2 and S1 represent the jet tip position at timest2=t1+dt and t1 respectively, then the massentrainment rate before impingement is given by
(14
where is the jet cone angle given by
(15
In most models the assumption is made that the burningzone after impingement follows a path parallel to thecylinders walls. In the present work the existing weltested wall jet theory of Glauert [17] is used to
[ ]
++
+++
++
=
=
b
pb
bu
pu
ucyl
bspeciesk
kkbbbfevfev
ububbbbbbb
pb
b
bbbbbbuuu
uuuuuuuu
pu
u
VC
RV
C
RV
/
]udymdmudmh
dmhdRTmdmTRdQ[C
R
PdVdRTmdmTRdmTR
dmudmhdmTRdQC
R
dP
cylbu
totalbu
VVV
mmm
=+=+
Head.cyllinerpistontot AAAA ++=
( ) ( )
+= 4w
4gwg
b
tot TTcTTD
ReaAdQ
=
=
=
u,bi
vii
u,bi
ivii
gCm
TCm
T
=
=
u,bi
ivii
iviiu,b
TCm
TCmdQdQ
fff
f
noz2noz
Dinjf
qm
P2
4
dCq
!!
!
=
=
cylflnoz PPP =
tdP
95.2S inj
25.0
u
noz
=
tdP
475.1U inj
25.0
u
noz
=
( )31322u SStan3
dm
=
o
u2 001.075.0tan
=
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determine the burning zone history after wallimpingement upon the cylinder walls. At this point theunburned mixture is entrained only from the wall jet front.
The entrainment rate into the wall jet is estimated from itsvolume change as follows,
(16)
where to is the transient time for the change of the freejet into a wall jet, Rc is the cylinder radius and tw is thewall travel time.
IGNITION DELAY MODEL Before the injected liquidfuel ignites, it undergoes a delay period. The ignitiondelay period is estimated from the following correlation[18]
(17)
Ignition initiates when the previous integral becomesequal to one, the integration having started from theinitiation of injection. At this point we must state that dueto its lower cetane number, auto-ignition of the gaseousfuel is avoided, since in the opposite case seriousproblems would arise [11]. Thus the liquid fuel is theignition source for the gaseous fuel.
BURNING MODEL FOR THE LIQUID FUEL The semi-empirical model of Whitehouse-Way is used forcalculating the rate of combustion of Diesel fuel. Thepreparation rate of injected fuel is given by [5]
(18)
where mfi is the total amount of fuel injected up to theconsidered time, mfupr is the total unprepared fuel,PO2 is the partial pressure of available oxygen andk,x,m are constants.
The burning rate of the diesel fuel is effected by anArrhenious type equation. This burning rate is given by
(19)
where mfav is the total amount of unburned fuel, maavis the total amount of unburned air and st is thestoichiometric equivalence ratio of the liquid fuel.
At the beginning of the combustion period, combustion iscontrolled by the reaction rate. In a short time, after theprepared fuel during the delay period is consumed,combustion is controlled by the preparation rate.
ESTIMATION OF GASEOUS FUEL BURNING RATE As already stated in the present work, we replace a parof liquid fuel with gaseous fuel in order to have the samepower output. The working media at the start of thecompression stroke is a mixture of air and gaseous fuelThus the temperature level of the mixture during thecompression stroke is very important for establishingwhether auto-ignition takes place. Experiments incombustion bombs have shown that auto-ignition ogaseous fuel under diesel-like conditions requires highetemperatures from those achieved during thecompression stroke [1,2,3], for this reason auto-ignition isavoided.
During the combustion stroke the mass of the gaseousfuel, entrained inside the burning zone, is proportionathe entrainment rate of unburned mass. The combustionrate of gaseous fuel is controlled by the entrainment rateand the local conditions inside the burning zone. Thereaction rate is determined using an Arrhenious typeequation as follows, [10]
(20
where mgifav is the available mass of gaseous fuelmO2av is the available mass of oxygen in the burningzone, "Egif" is the activation energy of gaseous fuel andA,z1,z2 are constants. The total rate of heat release isobviously the sum of the two rates, the liquid fuel hearelease and the gaseous one.
CHEMICAL SPECIES FORMULATION Combustionproducts are defined by dissociation considerations. Fothe C-H-O system the complete chemical equilibriumscheme proposed by Vickland et al. (1962) is used [15]For the combustion zone, given its volume, temperaturemass of fuel burned and mass of air entrained, theconcentration of each one of the eleven species can becalculated by solving a system of 11 equationscontaining: 4 atom balance equations (one for eachelement C,H,O,N) and 7 equilibrium equations. Aparfrom soot, at any instant of time, the following gasspecies are considered to be present inside each burnedregion, all in chemical equilibrium,
(1) H2O 2) H2 (3) OH (4) H(5) N2 (6) NO (7) N (8) CO2(9) CO (10) O2 (11) O
where the species are referred to by the number inparenthesis, in front of their name, and expressed bymeans of kmole fraction X. The equilibriumconcentrations of the eleven species can be described bythe following seven equilibrium reactions (B=X1/X2):
5.0o
5.11w
5.12w22
cut5.4
tttanUR
3dm
=
dt
T
5500expPa
1S
t
0
b
757.0del
pr
=
mO
xfupr
x1fifpr 2
Pmmkdm =
( )staavfav
b
757.0fb
m,mmin
T
5500expP'kdm
=
=
=
bm
fg
z
b
avO
z
b
favg
HC,CHi
fbg
TR
Eexp
m
m
m
mdm
i
2
2
1
i
834
i
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(21)
where Kp are the reactions equilibrium constants.
The solution of the above equations is obtained byincorporating the method proposed by Vickland et al(1962), with some slight modifications to ensure fastconversion.
NITRIC OXIDE CHEMICAL KINETICS Since theformation of nitric oxides is a kinetically controlledmechanism, the extended Zeldovich chain reactionmechanism is used. The three following reactions areconsidered
(22)
After some transformations the kmoles of NO areexpressed as follows:
(23)
where (NO) denotes concentration, = (NO)/(NO)e andsubscript "e" denotes equilibrium. "V" is the burned gasvolume and RI (i=1,2,3) is the one way equilibrium rate,for the "i" reaction, defined as follows:
(24)
and kif is the forward reaction rate constant for the ireaction.
SOOT FORMATION MODEL The modeling of sootformation is a very difficult task in common internalcombustion engine simulation models. For this reasonresearchers usually use semi-empirical models which
have been derived as correlation from the analysis ofexperimental data. In the present work a semi-empiricamodel, that has been widely tested, is used to predict therate of soot formation. [6,8-10,13]
The soot formation and oxidation rates are givenrespectively by,
(25
(26
where ms is the net soot formed, PO2 is the partiapressure of oxygen and Af and Ab are constants.
The net soot formation rate is then obtained from theexpression:
(27
EXPERIMENTAL FACILITIES AND PROCEDURE
To calibrate the present model an experimentainvestigation has been conducted on a single cylinderLister LV1, direct injection diesel engine located at theauthors laboratory. The results of this investigation areused to calibrate and evaluate the model under normadiesel fuel operation. These results are also used asbasis to evaluate the findings when using gaseous fuel asa supplement for liquid fuel. The technical data of theengine are given in Table 1. This is a naturally aspiratedair-cooled, four-stroke engine, with a bowl-in-pistoncombustion chamber. The normal speed range is 1000
3000 rpm. A three hole injector nozzle (each hole havinga diameter of 0.24mm) is located in the middle of thecombustion chamber head. The injector nozzle openingpressure is 190 bar. The main parts of the test installationused are:
Lister LV1 diesel engine.
Heenan & Froude hydraulic dynamometer.
Tank and flow meter for diesel fuel.
TDC marker (magnetic pick-up) and rpm indicator.
K-type thermocouples for measuring thetemperatures of the exhaust gas and engine oil.
Kistler 6001 miniature piezoelectric transducer fomeasuring the pressure in the cylinder, flushmounted to the cylinder head and carefullycalibrated.
Kistler 7063 piezoelectric transducer for measuringfuel-line pressure before the injector.
Kistler 5007 charge amplifiers.
( )
( )567P222
896P222
235P22
2104P222
573p2
10112p2
241p2
XBXPK,NOHN2
1OH
XBXK,COOHHCO
XBXPK,H
2
1OHOH
BPXK,OH2OH2
XXPK,NN2
1
XXPK,OO2
1
XXPK,HH2
1
=++
=++
=+
=+
=
=
=
)(
10f3
T/31256f22
10f12
102.4k,HNOOHN
eT104.6k,ONOON
106.1k,ONNON
=++
=++
=++
( )[ ] ( )
+
+
=
32
1
12
RR
R1
R12
dt
VNOd
V
1
( ) ( )
( ) ( )
( ) ( )eef33
e2ef22
eef11
OHNkR
ONkR
NONOkR
=
=
=
=
bm
sf5.0favfsf
TR
EexpPmAdm
=
bm
sb8.1O
sbsbTR
EexpP
P
PmAdm 2
sbsfs dmdmdm =
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The output signals from the magnetic pick-up andpiezoelectric transducers are connected to the input of aKEITHLEY DAS-1801ST A/D board, installed on a IBMcompatible Pentium II PC.
The exhaust gas analysis system consists of a group ofanalyzers for measuring all the main gaseous pollutantstogether with soot concentration. The nitric oxideconcentration (NO) was measured by a standard Signalchemiluminiscent analyzer having a NO2 to NO converterswitch and fitted with a heated line, while the sootconcentration was measured using a Bosch RTT100analyzer.
The series of experiments conducted for this preliminarystudy, involved as already mentioned only the diesel fuel(no gas was added) in a variety of engine operatingconditions. The following discussion refers to the resultsfor an engine speed of 1500 rpm and four different loads,namely 20%, 40%, 60% and 80% of full engine load.Currently efforts are given to conduct experiments usinggaseous fuel as a supplement for liquid fuel taking intoaccount the observations obtained from the modeling.These results will be used to evaluate the findings of thenewly developed dual-fuel model.
RESULTS AND DISCUSSION
Figures 1 to 4 present the comparison betweentheoretical and experimental pressure and heat releasetraces in the case of the 100% Diesel fuel, for 1500 rpmengine speed and four different loads namely 20, 40, 60
and 80% of full load respectively. It can be observed thatthe agreement in all cases is good, a fact that ispromising for the utilization of the model to predict engineperformance for dual fuel engines. It must be stated thathe values of the two-zone models constants are held thesame for the entire range of operating parametersvariation and for all fuel mixtures considered in thepresent study. Results for the experimental heat releaserate curves in the above figures were obtained from theanalysis of the corresponding experimental cylindepressure traces using a diagnostic code [20].
Figure 1. Comparison between experimental andcalculated pressures and heat release tracesat 1500 rpm engine speed and 20% loadunder 100% diesel fuel operation.
Figure 2. Comparison between experimental andcalculated pressures and heat release tracesat 1500 rpm engine speed and 40% loadunder 100% diesel fuel operation.
Table 1. Engine basic design data, Lister LV1-Dieselhigh speed engine
Bore 85.73mm
Stroke 82.55mm
Connecting Rod Length 148.59mm
Compression Ratio 18
Cylinder Dead Volume 28.03cm3
Inlet Valve Opening 15CA before TDC
Inlet Valve Closure 41CA after BDC
Exhaust Valve Opening 41CA before BDC
Exhaust Valve Closure 15CA after TDC
Inlet Valve Diameter 34.5mm
Exhaust Valve Diameter 31.5mm
Static Injection Timing 28CA before TDC
100 120 14 0 1 60 18 0 20 0 220 2 40
Crank Angle (deg CA)
0
10
20
30
40
50
60
70
80
P
ressure
(bar)
0
50
1 00
1 50
2 00
2 50
H
ea
tR
elease
R
ate
J/de
20% Load-1500 rpmFuel:Diesel
Calculated
Exp. Cyl. Pressure
Exp. Heat Release
100 120 14 0 1 60 18 0 20 0 220 2 40
Crank Angle (deg CA)
0
10
20
30
40
50
60
70
80
P
ressure
(bar)
0
50
1 00
1 50
2 00
2 50
H
eatR
elease
R
ate
J/d
e
40% Load-1500 rpmFuel:Diesel
Calculated
Exp. Cyl. Pressure
Exp. Heat Release
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Figure 3. Comparison between experimental andcalculated pressures and heat release tracesat 1500 rpm engine speed and 60% loadunder 100% diesel fuel operation.
Figure 4. Comparison between experimental andcalculated pressures and heat release tracesat 1500 rpm engine speed and 80% loadunder 100% diesel fuel operation.
The variation of maximum cylinder pressure with load byvarious percentages of gaseous fuel is given in Figure 5.As shown the maximum combustion pressure is affectedby the presence of the gaseous fuel in the correspondingmixtures. Combustion of the diesel-gas fuel mixtureresults to higher rates of heat release and this in turn
results to higher combustion pressures compared to purediesel fuel under the same conditions. Differencesbecome higher as engine load and percentage of liquidfuel replacement by gaseous fuel increase.
An explanation for the previous, is given by Figure 6,which presents the corresponding heat release ratediagrams for the two extreme mixtures considered at 40and 80% load.
Figure 5. Maximum combustion pressure vs load forvarious percentages of Diesel/Gaseous fuel.
It is observed that during the premixed-controlledcombustion phase of the liquid fuel, a sudden increase inheat release rate occurs, at higher load conditions. This
is due to the sharp rate of combustion caused bypresence of the gaseous fuel, which is already premixedwith air and ready for combustion.
Figure 6. Heat release rate at 1500 rpm and 80%, 40%load for 100/0 and 60/40 Diesel/GaseousFuel operation
Figures 7 to 9 present the comparison betweentheoretical pressure traces in the case of 100%, 90%80%, 70%, and 60% Diesel fuel for 1500 rpm and fo
1 0 0 1 2 0 1 4 0 1 6 0 1 8 0 2 0 0 2 2 0 2 4 0
Crank Angle (deg CA)
0
1 0
2 0
3 0
4 0
5 0
6 0
7 0
8 0
P
ressure
(bar)
0
50
10 0
15 0
20 0
25 0
HeatR
elease
Rate
(J/deg)60% Load -1500 rpm
Fuel :Diese l
C alcu lated
E x p . Cy l . P r es s u r e
E x p . H ea t R e leas e
1 0 0 1 2 0 1 4 0 1 6 0 1 8 0 2 0 0 2 2 0 2 4 0
Crank Angle (deg CA)
0
1 0
2 0
3 0
4 0
5 0
6 0
7 0
8 0
Pressure
(bar)
0
50
10 0
15 0
20 0
25 0
HeatRelease
Rate
(J/deg
)80% Load -1500 rpm
Fuel :Diese l
C alcu lated
E x p . Cy l . P r es s u r e
E x p . H ea t R e leas e
1 2 3 4 5 6
B.M.E.P. (bar)
50
60
70
80
90
M
axim
um
Pressure
(bar) 100/0-D iese l /G as
90 /10-D iese l /G as
80 /20-D iese l /G as
70 /30-D iese l /G as
60 /40-D iese l /G as
1 4 0 1 6 0 1 8 0 2 0 0 2 2 0 2 4 0
Crank Angle (deg CA)
0
5 0
10 0
15 0
20 0
25 0
30 0
TotalHeatRelease
Rate
(J/deg)
0
50
10 0
15 0
20 0
40% l oad -1500 rpm
100 /0-D iese l /G as
60/4 0-D iese l /G as
80% l oad -1500 rpm
100 /0-D iese l /G as
60/40 -D iese l /G as
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three different loads. It is observed that under part loadconditions the presence of gaseous fuel affect slightly thevalues of the pressure. This is due to the fact that duringthe premixed-controlled combustion phase the heatrelease rate at part load when using a mixture of Diesel-Gas fuel is similar to the one when using 100% dieselfuel. Under high load conditions the increase of thepercentage of gaseous fuel replacing the diesel fuel,leads to a sharper rate of heat release during thepremixed-controlled phase, as already shown in Figure 6.
Figure 7. Comparison between calculated Pressuretraces for various percentages of Diesel/Gaseous fuel at 1500 rpm and 40%load.
Figure 8. Comparison between calculated Pressuretraces for various percentages of Diesel/Gaseous fuel at 1500 rpm and 60%load.
Figure 10 presents the variation of brake specific fuelconsumption with engine load. It is observed that in thecase of 100% diesel fuel operation the model predictsaccurately the measured bsfc values for all loadsexamined.
Figure 9. Comparison between calculated Pressuretraces for various percentages of Diesel/Gaseous fuel at 1500 rpm and 80%load.
It is observed that the increment of gas percentage in thefuel results in an analogue increment of engine efficiencywhich becomes higher as load increases.
Figure 10. Brake specific fuel consumption at variousengine loads for various percentages ofDiesel/Gaseous fuel.
The variation of Nitric Oxide concentration with engineload is presented in Figure 11. Comparing the calculatedvalues of NO with the measured ones at 100% diesel fue
operation, it is revealed a good coincidence. The modeover-predicts slightly absolute values, which is normal foa two-zone model, but it manages to predict the trendwith load. This enables us to use it for NO predictionunder dual fuel operation. The formation of Nitric Oxide inthe fuel jet is directly related to the local oxygenconcentration and the gas temperature. Formation oNitric Oxide takes place mainly during the premixed andinitial part of mixing-controlled combustion, where high
1 0 0 1 2 0 1 4 0 1 6 0 1 8 0 2 0 0 2 2 0 2 4 0
C r a n k A n g l e (d e g C A )
0
10
20
30
40
50
60
70
80
Pressure
(bar)
40% Load -1500 rpm
100/0-Dies e l /G as
90 /10-Dies e l /G as
80 /20-Dies e l /G as
70 /30-Dies e l /G as
60 /40-Dies e l /G as
1 0 0 1 2 0 1 4 0 1 6 0 1 8 0 2 0 0 2 2 0 2 4 0
Crank Angle (deg CA)
0
1 0
2 0
3 0
40
5 0
6 0
7 0
80
Pressure
(bar)
60% Load -1500 rpm
100 /0-D iese l /G as
90/1 0-D iesel /G as
80/2 0-D iesel /G as
70/3 0-D iesel /G as
60/4 0-D iesel /G as
1 0 0 1 2 0 1 4 0 1 6 0 1 8 0 2 0 0 2 2 0 2 4 0
Crank Angle (deg CA)
0
1 0
2 0
3 0
4 0
5 0
6 0
7 0
8 0
90
P
ressure
(bar)
80% Load -1500 rpm
100 /0-D iese l /G as
90/10 -D iese l /G as
80/20 -D iese l /G as
70/30 -D iese l /G as
60/40 -D iese l /G as
1 2 3 4 5 6
B.M.E.P. (bar)
90
1 20
1 50
18 0
2 10
2 4 0
2 70
3 0 0
B
.S.F.C.
(gr/PS/hr)
1 0 0 / 0 -Di es el/ Gas
9 0 / 1 0 -Di es el/ Gas
8 0 / 2 0 -Di es el/ Gas
7 0 / 3 0 -Di es el/ Gas
6 0 / 4 0 -Di es el/ Gas
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local temperatures exist. At part load conditions the NOconcentrations differ slightly for various percentages ofDiesel/Gas mixtures due to the previously mentionedeffect on the heat release rate. As load increases nitricoxide concentration increase with the percentage ofliquid fuel replaced by the gaseous one.
Figure 11. Nitric Oxide emissions at 1500 rpm andvarious loads for various percentages ofDiesel/Gaseous fuel
Figure 12. Soot emissions at 1500 rpm and variousloads for various percentages of Diesel/Gaseous fuel
Figure 12 presents the variation of the soot concentration
in the engine exhaust with load for various proportions ofDiesel/gas mixture. Observing Figure 12 we can see thatthe current model predicts adequately the experimentallymeasured soot emissions at 100% diesel fuel operation.This is encouraging and makes the results obtainedunder dual fuel operation more reliable. A difference isobserved at 60% load, which is due to the differentbehaviour of the engine fuel injection system, asobserved from the measured values of fuel injectionpressure. As this percentage of gaseous/liquid fuelincrease, the soot concentration is decreased since less
liquid fuel is injected and thus less soot is formed. Alsodue to the higher temperatures existing at highpercentages of liquid fuel replacement, the soot oxidationrate is higher contributing to a further decrease of emittedsoot. The decrease of soot is more intense at high loads.
CONCLUSIONS
Under the present work a new model has been
developed to simulate the operation of dual-fuel dieseengines for the prediction of performance and pollutanemissions. This preliminary investigation focuses onexisting diesel engines where a part of the liquid fuel isreplaced with a gaseous one equivalent to natural gasThe main purpose is to examine the effect of liquid fuepercentage replaced by gaseous fuel on performanceand mainly soot and nitric oxide emissions.
For this reason a two-zone model has been developedcapable of operating with various percentages of liquidand gaseous fuel. Combustion is initiated by the ignitionof the liquid fuel, while the burning rate of gaseous fuel iscontrolled by its entrainment rate into the burning zoneDue to the lean nature of the surrounding unburned air-gaseous fuel mixture no flame front is considered insidethe unburned zone. To validate the model before using ito predict engine performance and pollutants emissionsunder dual-fuel operation an extended experimentainvestigation has been conducted on a high speed Dsimple cylinder test engine located at our laboratoryMeasurements have been taken for both performanceand emissions at various operating conditions usingdiesel fuel only. These are used for model validation andalso serve as comparison between normal diesel fueoperation and operation when using various percentages
of liquid and gaseous fuel.Comparing calculated and measured values undenormal diesel operation a good coincidence is observedfor both performance and pollutant emissions. Thisenables us to use the model to predict engine behaviourunder dual-fuel operation. From the analysis ocomputational data it is revealed that dual-fuel operationresults to higher combustion pressures. These pressuresincrease with increasing percentages of gaseous/liquidfuel. The effect at high engine loads is more intenseConcerning engine efficiency it is revealed in general thathe replacement of liquid fuel with gaseous results to animprovement of engine efficiency. This improvement ismore intense at higher engine loads and higher values ofthe gaseous/liquid fuel ratio.
As far as pollutant emissions are concerned the use ofgaseous fuel has a negative effect on NO emissions anda positive on soot. Specifically an increase of NOemissions is observed which is serious at high engineloads and gaseous/liquid fuel ratios. On the other handsoot is seriously decreased when using gaseous fuel asa supplement for liquid fuel. Again this effect is stronge
1 2 3 4 5 6
B.M.E.P. (bar)
3 00
4 50
60 0
7 50
9 00
1 0 5 0
1 2 0 0
1 3 5 0
1 5 0 0
1 6 5 0
1 8 0 0
1 9 5 0
NitricOxide
(ppm
)
100/0-Dies e l /G as
90 /10-Dies e l /G as
80 /20-Dies e l /G as
70 /30-Dies e l /G as
60 /40-Dies e l /G as
E xp . D i e se l
1 2 3 4 5 6
B.M.E.P. (bar)
0 . 0 0
0 . 0 2
0 . 0 4
0 . 0 6
0 . 0 8
0 . 1 0
0 . 1 2
0 . 1 4
0 . 1 6
0 . 1 8
0 . 2 0
S
oot(m
g/lt)
100/0-D iese l /G as
90 /10-D iese l /G as
80 /20-D iese l /G as
70 /30-D iese l /G as
60 /40-D iese l /G as
E xp . D i e se l
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at higher engine loads and gaseous/liquid fuel ratios. Butthe most important is that when using gaseous fuel wealways have a reduction of soot emission regardless ofthe engine operating conditions.
The results of this preliminary investigation areencouraging and urge us to conduct an experimentalinvestigation to verify the findings. This is currently underprogress and results will be given in the near future. Eventhough it is difficult to generalize the findings of the
current preliminary investigation, we believe that they areimportant since the reduction of soot emissions onexisting DI diesel engines is extremely important.
REFERENCES
1. Bahr,O., Karim, G.A., Liu,B., An examination of theflame spread limits in a dual fuel engine, Appl.Therm. Engng., Vol 19, pp.1071-1080, 1999
2. Karim, G.A., Khan, M.O., Examination of effectiverates of combustion heat release in a dual-fuelengine, J.S.M.E., Vol 10, No 1, 1968
3. Agarwal, A. Assanis, D.N., Multidimensionalmodeling of natural gas ignition under compressionignition conditions using detailed chemistry SAEpaper, No 980136, 1998
4. Pirouzpanah, V., Kashani, B.O., Prediction of majorpollutants emission in direct-injection dual-fuel dieseland natural-gas engines, SAE paper, No 990841,1999
5. Whitenhouse,N.D., Sareen,B.K., Prediction of heatrelease in quiescent chamber Diesel Engine allowingfor fuel/air mixing, SAE paper, No 740084, 1974
6. Kouremenos,D.A. ,Racopoulos,C.D. andHountalas,D.T. A computer simulation of combustion
process in Diesel Engines with no-swirl for thepurpose of heat release and nitric oxide prediction,Proc. Int. A.M.S.E., Vol. 3.3,207-218,1986
7. Annand, W.J.D., Heat transfer in the cylinders ofreciprocating internal combustion engines, Proc.Inst. Mech. Engrs., 177, 973-990, 1963
8. Kouremenos,D.A. ,Racopoulos,C.D. andHountalas,D.T. `Multi zone combustion modeling forthe prediction of pollutants emissions andperformance of DI Diesel engines, SAE paper, No970635, 1977.
9. Kouremenos,D.A. ,Racopoulos,C.D. andHountalas,D.T., Computer simulation with
experimental validation of the exhaust nitric oxideand soot emissions in divided chamber Dieselengines, Trans. ASME, WA meeting, San FranciscoCalifornia, Vol.10-1, pp.15-28, 1989.
10. Ramos, J.I., Internal Combustion Engine Modeling,Hemisphere, New York, 1989
11. Heywood, J.B., Internal Combustion EngineFundamentals, McGrawHill, New York, 1988
12. Benson, R.S. and Whitehouse, N.D., InternalCombustion Engines, Pergamon, Oxford, 1979
13. Hiroyasu, H., Kadota, T. and Arai, M., Developmenand use of a spray combustion modeling to predicdiesel engine efficiency and pollutant emissionsBulletin, J.S.M.E., 26, 569-576, 1983
14. Racopoulos, C.D., Hountalas, D.T., Tzanos, E.I.andTaklis, G.N.,A fast algorithm for calculating thecomposition of diesel combustion products using aneleven species chemical equilibrium schemeAdvances in Engng Software, 19, 109-119, 1994
15. Vickland, C.W., Strange, F.M., Bell, R.A. andStarkman, E.S., A consideration of the hightemperature thermodynamics of internal combustionengines, Trans. SAE, 70, 785-793, 1962GlauertM.B., The wall jet, J.Fluid Mech., 1, 625-643, 1956
16. Bazari, Z.,A DI Diesel combustion and emissionpredictive capability for use in cycle simulation, SAEpaper, No 920462, 1992
17. Glauert, M.B., The wall jet, J. Fluid Mech., 625643,1956
18. Kadota, T., Hiroyasu, H. and Oya, H., Spontaneousignition delay of a fuel droplet in high pressure andhigh temperature gaseous environments, BulletinJ.S.M.E., 19(130), 1976
19. Lavoie, G.A., Heywood, J.B. and Keck, J.C.Experimental and theoretical study of nitric oxideformation in internal combustion engines, CombustSci. and Technol.,1,313-326, 1970
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NOMENCLATURE
, b, c = constants
adel = constantA = area, m2
Af = constant in soot formation mechanism
Ab = constant in soot oxidation mechanism
Cv = specific heat capacity under constant volumeJ/kgK
Cp = specific heat capacity under constantpressure, J/kgK
CDinj = injector hole discharge coefficient
dnoz = njector hole diameter, m
D = cylinder bore, m
E = activation energy, J/Kmole
k = constant for liquid fuel preparation ratek = constant for liquid fuel reaction ratek = constant for gaseous fuel reaction ratem = mass, kg, or constant for the liquid fuel
preparation rate
= mass flow rate, kg/sec
P = pressure, Pa
PO2 = partial pressure of oxygen, bar
= volumetric flow rate, m3/sec
m!
q!
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Greek
Subscripts
Dimensionless number
Abbreviations
Q = heat transfer to walls, J
R = radius, m
Rm = universal gas constant, J/kmoleK
S = penetration length, m
t = time, sec
T = absolute temperature, K
U = jet velocity, m/sec
u = specific internal energy, J/kg
V = volume, m3
X = Kmole fraction on species
y = mass fraction on species
x = constant
zi = constants
= initial jet angle, radP = pressure difference, Pa = thermal conductivity, W/mK = dynamic viscosity, kg/ms = kinematic viscosity, m2/sec = density, kg/m3 = equivalence ratio
a = air
av = available quantity
b = burned zone
del = delay
ev = evaporated liquid fuel
e = equilibrium
f = fuel
fpr = liquid fuel prepared
fi = liquid fuel injected
fupr = liquid fuel unprepared
fb = fuel burned
g = gaseous fuel or gas mixture
i = index denoting control volume or the kind ofgaseous fuel
= index denoting the species of the gaseousmixture
o = initial value
tot = total
s = soot
sf = soot formed
sb = soot oxidated
st = stoichiometric
u = unburned zone
w = wall
Re = Reynolds
DI = direct injection
bmep = brake mean effective pressure
bsfc = brake specific fuel consumption
NO = nitric oxideppm = parts per million (volume)