The Japan Society
TheJapanSociety
of Mechanical
Mechanical Engineers
Engineers
of
Proceed
±ngs
of
the
Power
Engineering-03
Japan
9-13. 2003. Kobe,
International
Conferenee
{:COPE-03} November
A311
PERFORMANCE
STUDIES OF A BIOGAS FUELED DIESEL
OPERATING IN A DUAL FUEL MODE
M, Z.Haq', M. H. Rahman "
and
ENGINE
Z.A. Bhllttot
University
ofEngineering
'DepartrnentofMechanicalEngineering,Bangladesh
and
TechnologM Dhaka
CentralWbrk$hop, EME, Bangladesh Army, Dhaka Cantonment
'"901
t
on
Wbrkshep, EME, Bangladesh
Army, Dhaka Cantonment
Adhoc Field
Thepresentpaper reports the experimenta1 and modeling studies of a 4-cylinder,
4-streke,
direct
iniection
diesel
engine. Studyiscanied out fortwo modes
ofoperatiens - one with straight dieselfuelingand
with a fixed
amount
another in dual-fuel
mode using natural gas and biogasof two compositions
of pilotdiesel
ABSTRACT
iniectionfor the ignitionof the charge.
Both experiments
and
simulations
are
carried
for different
engine
out
speeds at different
loading
conditions, The results obtained fromthe simulations are inreasonable agreement
with the experimental results; however,neglect
of detailed
combustion in the model
resulted
in higherrated
powerestimate ofthe engine, Lower cylinder gas pressureand higherbulk gas temperatures are obtained in case
of fue!ingby natural gas,and these quantities
decreasewith increaseinthe carbon dioxidecontent ofbiogas.
Keyvvords: Dieselengine, Alternativefuels,Heat release, Naturalgas,Biogas.
1.INTRODUCTION
Existingstationary dieselengines can be retrofitted
fairlyeasily foroperation with alternative gaseous fuels,
such
as natural
gas and biogas.Natural gas is new being
widely
used to fuelcombustion
engines,
however itsreserve
is limitedin many areas ofthe world,
Hence, biogas is a
in nature and
potentialaltemative fuel that is renewable
thereby does not contribute
to the net atmospheric
concentration
ofthe
greenhouse gascarbon dioxide.Itisa
colorless
combustible
gas produced by the fermentationof
cellulose
materials,
manure
or cow-dung
[1].Metharieand
carbon
dioxideconstitute about 95% of biogas and rests are
trace
erganiclnon-organic
elements
of
variab]e
compositions.
Moreover,the amount of methane and
carbon
dioxidein biogasvaries with the sources of origin
and
the variatien
affects
combustion
process and heat
release
rates
be used
[2].Fortunately,same
reaction
te model bothmethane
and
modeled methane
combustion
the available experimental
and
with
coTnprehensive
where
engine
the various
modeling
and
engine-operating
mechanism
biogascombustion
data is very
results
The present paperreports
both experimental
and modeling,
engine
fortwo
modes
dieselfuelingand
of
another
[3],
[4]. Hence,
is possible
parameters can be
The modeled
data
[5].
simulation
inter-
- one
with
straight
using
natural
biogasof two compositions
(onvolumetric basis,
70% methane
+ 30% carbon
diexide isdenoted by BG30,
gas and
RELEASE
ln general,
literature
reports
heatavaiiable fromstoichiQmetric
the maximum
combustion
arnount
ofunit
of
mass
fuel, knewn as heating value,
Hewever, practical
processesare not essentially stoichiometric, and
heatrelease depends en the equivalence
ratio, ip.
Therefbre,
two usefu1 definitions
of heatrelease
quantityat standard
condition ofO.1MPa
pressureand 25eC are:
1. Specificenergy, SE - net heat available from charge of
of
combustion
can
in dual-fueimode
operations
2, HE,ff
unit mass
the perforrnance
studies,
clirect
iajectiondiesel
ofa
analyzed,
and
(kJlkg-charge),
2. Energydensity,
ED - net heatavailable fromcharge
consistent
to obtain optimum
condition
also providesvaluable
insightintothe complex
relationship characterizing diesel
engine operation.
studied
and
50% rnethane + 50% carbon dioxideis denoted by
BGSO), Study iscarried out for fourdifferent
engine
speeds
of 2100, 2200, 2250 and
2300 rpm at differentloading
conditions, and various engine performance parametersare
unit volume
of
(kJ/m3-charge).
Shown in Fig. 1 are the heating values of methane,
Itisseen
gasoline,dieseland biogas of two compositions.
that the heating
values
ofbiogas are sigriificantlylowerthan
those of hydrocarbon
fuelsand the deviation
ishigherfor
highercarbon dioxidecontent ofbiogas.
Shown in Figs.2
and
3 are the variations of specific energy and energy
density,
respectively,
of equivalence
plottedas a function
ratio forthe same fuels.
Itisobservecl thattheheat release
fora given arnount ef charge dependson bothequivalence
ratio and the fuelitselgand
the heat releases frornbiegas
are
significantly
lower than those of the
premixtures
hydrocarbon based fuelsand the deviation
increaseswith
increasein itscarbon dioxidecontent.
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by regulating the gas flow rates. The pilotfuel
ibjectienin case of gaseous fuelsare required becauseof
the auto-ignition ternperature
of methane
which is
(540eC)
significantly
higherthan diesel
and the temperature
(2600C)
generatedfromthe compression of the air insidethe engine
cylinder
is not sufficient for the ignitionof the charge [6].
A schernatic diagram ofthe engine setup isshown inFig.4.
AII measuring
instrurnentsare calibrated using relevant
standard
procedures. Experimental
proceduresreported in
[7]are used as general guideline for experimentation.
BritishStandardBS-5514, equivalent to ISO Standard3046
and SAE StandardJ1349 [8],
has been used for
the engine brake power and fuelconsurnption
rates. Details
ofthe
experimental
in[9].
procedure are described
An engine model, with similar configuration
to the
varied
7060i'.m
5oJt
4oe]
3oga
20'p[g
to=
`de-rating'
oGasollneDieseiMethaneBG30BG50
Fig.1. Heating values ofdiesel,
biogasoftwo compositions,
gasoline,methane
and
experimental
55++
5+x 5+x 5+ 8+x5+x
5+
×
5+x ×
×
×
gex
×
1.5co1,Oee5
GasolineB
O
BG30
+
O.5
O.6
Die$et
×
A
Methane
BG50
O.7 O.8 O.9 1.0 f.1 t.2 f.3 1.4
e
Fig,2. Specificenergy
of fue1-airpremixtures.
4.0
3,5Gg
4. RESUIJTS AND
8g:,:.ooA+x
oob+x DoA+x
se8'
DoA+
3.o6"o-
2.5mE)
×
×
Gasollne
oB30
O.5
O,6
O.7
O.8
Diesel AB50 Methane
O,9 t.O 1,1 a.2 1.3 1.4
e
Fig.3. Energy densityof fue1-airpremixtures.
3. EXPERIMENTAL
SETUP AND ENGINE
commercial
DISCUSSIONS
and
x
+
using
[10],A schernatic
Shown inFigs,6 and 7 are the cylinder gas pressures
bulk gas temperatures,
respectively,
of the dieselengine
amount
same
of output
generatingapproximately
power
when
itis fueledby diesel,natural gas and biogas of two
compositions.
It is seen that, peak cylinder gaspressure,
P...,isthe highest
fordieselfuelthat isfo11owed
by natural
gas, BG30 and BG50, However, in case of bulk gas
temperatures, natural
gas produced the highest peak
temperature and the diese!
operations produced the lowest
orre,
ln cases of beth temperatures and pressures,
biogas
produced lowerva]ues than the corresponding
natural gas
values
and the deviations
increasewith increasein carbon
dioxidecontent of biogas. The Iowerpressureand higher
temperatures generatedin case of natural gas operations
are
in line with the experimental
observations
made
in the
present study
[9], Shown in Fig. 8 are the indicator
diagramsforthesame conditions as shown in Figs.6 and 7.
It is seen that,indicatordiagramsare very close to each
other
because of the factthat similar amount
of heat
addition with differentfue]sproduced similar amount of
brake power, and thereby estimated
bTake thermal
×
2.0lQ
1,5tu1.0E88eQ o
simulated
GT-Power
diagram ofthe importantcomponents
of the mode] isshewn
in Fig,5. In the presentmodeling,
directirijection
Weibe
combustion
model
isused to iniposethe combustion
rates
using
a threeterm Weibe function
as outlined
in [10],
and
the superposition of three functionsmake
it possibleto
model
pre-ignitionand largertail presentin the typica]
dieselengine heat release profile[11].
Chen-Flynnfriction
model
[12]isused to calculate engine frictionand Wbschni
heat transfer correlation [13]
is used to estimate the heat
transfer from the cylinder, while Colburn analogy
[14]is
used to predictconvective
heattransfercoethcient in ducts
and
pipes forming the intake and exhaust manifolds.
Variousparametersrequired in the modeling
are selected as
per guidelinespresented in [10].In all cases of the
modeling,
ambient air isassumed
to be at BS-55 14 standard
condition
300 K temperature and 60%
(O.1MPa pressure,
relative humidity) and modeling
iscarried out forthe same
conditions
as those efthe
experiments.
Detailsof the model
and
the simulations
are reported
in [15].
3.0G.ep'os2.5:\an
2.0gu
is also
engine,
engine sirnulation software
MODEL
Experiments are carried out in a 4-stroke,4-eylinder
directiniectiondiesel engine, specification of which is
presented in Appendix-I. Diesel has been feecl
to the
irijector
and gaseous fuels
are supplied
pump under gravity,
to the intakemanifbld
efthe
engine
by fumigationmethod,
In case of fueledby gaseous
fuels,
a fixedamount
of diesel
is iTijected
directly
intothe engine cylinder as a
(2kg/hr)
pilot fuel to ignitethe charge and the engine Ioads are
eenciencies are also
close
to each
othez
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,Air
Exhaust
N.izlC
A'essureGauge
-
-,e
A
i
ys=-ugdi
,tsi
J
t'
EEn.tr<
y:Egspo 4Hl･1.fi･=tlE
r11LtIt[t1.t
gees1efiop-
-g
-"-iE ntneT
NGSupPiY
r-!
L
Engine
Elli][lynHrnometerDisp,
Dyno
Fig.4,Experimental
setup showing
themajor instmmentatlon,
lntakeAlr
Exhaes[
{3avK,O.IMPa,60fiR
IIIi
E/igincCimk"aln
Fig.5.GT-Power model
TDC
8r.7g6!5mS4a.3tu92dielL.oo
BDC
ofthe
engine,
I800y-
TDC
BDC
16oovg
14oogo
ri2008Sl
10oDut8
sooMm]
600
-30
O
30
60
90
' "20
CrankshaftAngte
Fig.6.Cylindergas pressuresforsimilar
different
fuelsused, engine
speed
150
-30
180
(deg)
output
is2250rpm,
O
30
6e
90
CrankshaftAngle
power for
"20
150
180
(deg)
Figfor
a',i,e9',//?d.e:,ga.S,gxrn,p,:,r:.tu
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88･g6gi4Ie,!to"oo,o
O.1
02
DisplacementVo[ume
O.3
O.4
(ILtre)
in indicating
engine
be misleading
performance.Therefore,
in the analyses of
brake thermaJ efficiency is employed
the engine is fue!ed by
engine
perfbrmancewhen
alternative
fuelssuch as natural gas and biogas. In the
analysis,
brake mean effective pressure,bmep, is
present
the
used insteadof brake power,Pb. The bmep removes
effect of the engine
size, and this quantity is roughly
as diffbrent
comparable
even in very differentengines,
under
fueL, necessarily
engines
burri the same
the same
conditions
and
hence generate
approximately
isbmep are known to
similar pressures,
and the differences
and
are not
represent
engine
design
differences
genuine
For spark ignition
irrelevant
differencessuch as size [16].
of
values
of bmep
are in the range
engines,
maximurn
O,85 te !.05 MPa at the engine speed where maxirnum
rated power,bmep
torqueisobtained and at the maximum
values
are 10 to 15% lower,For naturally aspirated diesel
the bmep isin the O.7 to O.9 MPa range, with the
engines,
rated power of about
O.7MPa [11].
bmep at the maximum
's
fuelsused.
diagramsfbrdifferent
Fig.8. indicator
Shown in Fig.9 are the variations of brake specific
with brake power, Pb, at different
fuelconsumption,
bsy2),
only.
it
the engine isfueledby diesel
testconditions
when
is high at part ioadingand its value
is seen that bsy?]
decreaseswith increasein the btake power until the brake
power reaches the rated value where the bsyZris minimum
and
any furtherincreasein brake power results in higher
bsti:.The values of bEfbfor the slower engine operation
shows
lower values at partload condition, and the situation
It is alse
isfoundto be reversed at overload condition,
300
290
280?-g
27qE.
26oS
with engine
speed,
observed that the rated power increases
within
the speed range considered in the present study,
2508
240
This behavior ef bofbvs brakepower can be explained by
components
invelvedin frictienal
considering
the different
loss is
losses: a significant
fractionof total frictional
ofthe
total is
related to the engine
speed, a fraction
directly
while
a
directly
related to the peakengine cylinder
pressure
fractionof the tetal tossremains essentially constant [15],
At iow load conditions, the efiect of engine cylinder
lossesare
pressureis not significant, rather the frictional
the shafts,
in moving
dictatedby the energy consumed
valves
and
pumps, and these lossesincreasewith speed.
islow fbr low speed
at low lead conditions,
bsy2r
Therefore,
operations.
However, at higher load, the effect of peak
significant
and peak
cylinder
pressure becomes more
engine
pressureforslower engine operations are higherfor
the same power generation and hence highe! frictional
Moreoverl at higherenglne speed lessheatis lost
losses.
through the cooling
system.
Therefbre, at high-speed
are
lower fbr rateci power
of boflr's
operations, values
conditions.
However, beyond the rated power, suppiy ofair
burning because of the
isnot suthcient te ensure complete
heterogeneousmixture of air and fuel,and incase of overall
stoichiometric
mixture some portionof the charge starve
Therefbre,at
from oxygen to complete the combustion,
mixture
rated load, average
.isalways lean and beyond the
is not completed
and
rated
load conditions,
combustion
some
of the energy
inputislostin the form of incomplete
and results in highervalues
ofbstIr's,
combustion
products
boflr,is also a
Brake specific fuel consumption,
measure of engine overall efficiency, generallyexpressed
as brake thermal eenciency,
nb, and theses quantifiesare
inverselyrelated, so that the IDwer the bsy?,thehigherthe
fuels
of the engine.
However, for different
ecaciency
diffefent
heating
values,
the
values
of
bofle's
might
having
230
22e5
10
15
P,
Fig. 9, Brake specific
different
engine speeds.
2e
25
(kW)
fuei consumption
of
diesel for
Figure 10 shows the brake thermal efficiency plotted
operation.
Brake
functionof bmep forstraight diese!
as shown
in Figs. I1-13,have trend
therrnal eenciencies,
as shown in Fig. 1O, when
the
similar to the diesel
fueling
dieselis substituted partiallyby natural gas, BG30 and
and
modeled
results are
BG50. Both the experimental
plottedin these figures.It is seen that, in case of pure
dieseloperation
as shown
in Fig. 10, the experimental
values
of the brake thermal
ethciency
increases
with
increasein bmap unti! itreaches itsmaximurn value at the
to that
rateci brake power of the engine
corresponding
speed.
Beyond the rated power, brake thermal eenciency is
of combustion
efficiency
reduced
because of the decrease
resulting
fromthe shortage of oxygen required to burn the
fuelcompletely. In case ofnatural gas fueling,as shown in
Fig. 11, the efficiency continues to increaseeyen beyond
the rated power forthe dieseloperatien.
Itisbecause ofthe
increased combustion
ethciency
resulted frorn the
isinductedinto
combustion
of premixed natural gas which
Itispessibleto
the intake
manifold by fumigationmethod.
fueledby natural
significantly overload the engine when
gas,however with a potentialto damaged engine caused by
oveFheating.
Engine efficiency is slightly reduced in case
as a
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in Figs.12 and 13, Moreover,
as shown
of biogas fueling,
of power forhigher
the modeled
results show the reduction
of
carbon
dioxide in biogaswhich is due to the presence
any energy but absorbs
in the forrn of raised
that does not supply
dioxide
carbon
heat when
some
temperature.
In the results
modeled
up
403836su32302B26242220
exhausted
-*v
in Fig.10-13,itisseen that the
plotted
2-3% of the experirnental results
predicts
power and the modeling
VAA
to therated
engine
in all cases considered.
higher thermal eenciency
Moreover,in case of straight dieselfuelingthe modeling
decreasein thermal
also
failedto show any noticeable
eenciency beyond therated power because ef itsIimitation
to be
that causes the eenciency
to model
real combustion
of
the
the
rated
because
reduced
beyond
power
heterogeneousnature ofthe charge. However, for gaseous
fueloperations as shown in Figs.11-13,slight reduction m
beyond some higher rated powerscan
thermal eenciencies
negligible
engine
speed effect is
be anticipated. Practically
observed
inthe results shown in Figs, 10-13,where brake
mean
thermal efficiencies are plottedas a functionofbrake
emphasize
the
Hence,
these
effective pressure,
plots
rather than brake power in
of the uses ofbmep
irnportance
which
engine
data representation
provide a rnore usefu1
of different
sizes
the
results
for
engines
basisforextending
and
Fig. 12.Brake thermal
speeds
la-
34A
ev
.2300
22
20
O.4
O.2
tzgege{}
iA
e.4
2250
i
A
2300.
!
v
1,O
(MPa)O.8
.Expt.
NCrpm}
Mode[ ExpL
2tOO
2200
2250
2300
t
o
.
i
o
A
!
v
o.6
bmep
1,O
(MPa)O.8
o
fordifferentengine
Fig.13, Brake thermal ernciencies
by biogasBG50 and pilot
speeds
when
the engine isfueled
6
dieselfuel.
D
v
y
1.0･
O.6
bmep
D
o
"l .pm'ex;e-
32S
30.a
28
26
24
22
20
O.2
NCrpm) Mode[
2aoo
.
2200
i
225e
e
for differentengine
ethciencies
isfueledby biogasBG30 and pilot
o
.a
2B
-
40
38
36
-etttftei
.di'Rgg9keag
-
the engine
2aoo
2200
diesel
fuel.
38
sch
32e
3o.o
when
Model ExpL
N(rpm)
O,6
bmep
40
36
O.4
O.2
speeds.
26
24
ygets
nr
are within
values
twV.b"n
.
{MPa)e,s
5. CONCLUSIONS
Fig. 10,Brake thermal ethci,encies fordifferentengine
when
the engine isfueled
by straight diesel.
speeds
L
Majorconclusions ofthe presentstudy are:
Dieselfuelcan be pardallysubstituted by natural gas
The conversion to dualand biogasin dieselengines.
by fumigatienrnethod demands only simple
However,
of the engine intake system,
is
required
to
igriite
the
charge as
diesel
irijection
pilot
methane
hashigherauto-ignition temperaturethan that
fuelmode
modification
40
3B
gnMe.s"e-,i.
36
suA
32S
3o."
28
26
24
22
20
sEgeegkR
D,bgeag
VOAwoA
O,2
Fig, 11. Brake thermal
when
the engine
2. With modest
Medel
2100
1
a
2200
.
o
2250
23eo
i
A
!
v
effbrts,
itis possibleto use GT-Power
engines
with
to model
diesel
gaseous altemative
fuels such as natural gas and biogas. Results thus
agreement with the
obtained
are also in reasonable
Expt,
N(rpm)
O,6
O,4
bmep
speeds
ofdiesel.
vv
.
code
experimentai
results.
1.0
3. Modering and simulation using GT-Power provide
inter-relationship
valuable
insight
intothe complex
Once the
characterizing
dieselengine operations.
model
isverified against the experimental resultsl it
engine
can be used to stucly the effects of various
optirnum
engine
operating
for
perforrnance.
parameters
(MPa}O.8
for differentengine
efficiencies
is fueled'bynatural gasand pilot
4. Modeling with the neglect of detailsef
processesresults in higher rated capacity
dieselfuel.
cornbustion
estimate
of
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the dieselengine forthe fuelsconsidered
in the present
study.
5, Lower
gas pressure and higher bulk gas
generated in case of Diesel engine
fueled by primarily natural gas. These quantities
decreasewith increasein carbon dioxidecontent ofthe
biogas.
cylinder
temperature
are
Engine with Biogas/DieselDual Fuelingfbr Optimized
Performance, M.Sc. thesis (2003),
Bangladesh
University
ofEngineering and Technology.
16. Lurnley,J. L., Engines - An Introduction,(1999),
CambridgeUniversity
Press.
8.APPENDIX-I
6.ACKT-IOWLEDGEMENTS
of
En ineSecifications
The authors are gratefu1to Dr.Syed Wahiduzzaman
Gamma Technologies fbr providing the GT-Power
software
to Department ofMechanical
Engineering,BUET,
ModelType
4-cylinder,
diesel
engine
Compressionratio
Ratedpower
FueliTijection
purnp
1. Goodger, E.M,, Alternative
Fuels- ChemicalEnergy
Resources,(19gO),
McMillian PressLtd.,UK.
2. Haq, M. Z, arid Mizanuzzaman, M., Burning, Heat
Release and Exhaust Products of Biogas-Air
Combustion, J. Instn. of Engrs, Banglaclesh
Vbl.27, (2002),
(Multidisciplinary),
pp. 74-79,
KubotaV1502B,
: Vertical,
water coered,
Bore x Stroke
Totaldispracement
7.REFERENCES
:
Fuel iTijection
Fuei irijectien
pressure
Combustionchamber
Fuel
:76
mmx82
mm
: 1487 cc:21:
17 kW
at 225O rpm
: Bosch K
type
12.5bTDC
: 13,73 MPa
: Spherical
: DieselfuelNo. 2-D
:at
3. Mizanuzzaman,M.,ModelingofFIame?ropagationin 9, NOMENCLATURE
LaminarBiogas-AirPremixture, M.Sc. thesis,(2001),
Bangladesh
University
ofEngineering
and Technology,
4, Gu, X., Haq, M. Z.,Lawes, M. and WbolleM R.,
USblBDCBG30BG50
Laminar Burning Nlelocity
Bottom Dead Centre
and
MarksteinLengths of
Methane-Air Mixtures, Combustion and Flame, iFbl.
Biogascontaining 309i6carbon dioxide
and 70% methane,
121,(2000),pp.41-58.
by volurne
5. Morel, T.,Keribar,
Biogas containing 50% carbon dioxide
R. Silvertri,
J.and Wahiduzzaman,
S,,IntegratedEnginefVehicleSimulation
and 50% methane,
by volume
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