15. Study on Biogas Premixed Charge Diesel Dual Fuelled Engine - 10 Orang

www.elsevier.com/locate/enconman

Energy Conversion and Management 48 (2007) 2286–2308

Study on biogas premixed charge diesel dual fuelled engine

Phan Minh Duc *, Kanit Wattanavichien

Mechanical Engineering Department, Faculty of Engineering, Chulalongkorn University, Phaya-Thai Road, Patumwan, Bangkok 10330, Thailand

Received 22 July 2006; accepted 29 March 2007Available online 21 May 2007

Abstract

This paper presents an experimental investigation of a small IDI biogas premixed charge diesel dual fuelled CI engine used in agri-cultural applications. Engine performance, diesel fuel substitution, energy consumption and long term use have been concerned. Theattained results show that biogas–diesel dual fuelling of this engine revealed almost no deterioration in engine performance but lowerenergy conversion efficiency which was offset by the reduced fuel cost of biogas over diesel. The long term use of this engine with bio-gas–diesel dual fuelling is feasible with some considerations.� 2007 Elsevier Ltd. All rights reserved.

Keywords: Dual fuel; Biogas; Premixed charge; IDI

1. Introduction

Biogas, produced by the anaerobic fermentation of cel-lulose biomass materials, is a clean fuel for internal com-bustion engines. In oil crisis situations, it may act as apromising alternative fuel, especially for diesel engines,by substituting for a considerable amount of fossil fuels.Diesel engines can be easily converted to fumigated dualfuel engines. This is the most practical and efficient methodto utilize high spontaneous ignition temperature alternativefuels, such as biogas. In the fumigated dual fuel method,biogas mixes with air before this mixture enters the com-bustion chamber, and at the end of the compression stroke,an amount of diesel fuel, called the pilot injection, isinjected to ignite it. This method has the advantage ofthe ability to switch back to diesel operation in case of ashortfall in biogas supply during an important operation.Because of these benefits, dual fuelling of diesel and biogas[1–4], as well as producer gas [5–10], LPG [11–16], NG [17–26] or hydrogen [11,27,28], have been investigated widelyworldwide for some past decades. Karim G.A. et al.

0196-8904/$ - see front matter � 2007 Elsevier Ltd. All rights reserved.

doi:10.1016/j.enconman.2007.03.020

* Corresponding author. Tel.: +66 2 2186607; fax: +66 2 2522889.E-mail address: [email protected] (P.M. Duc).

[11,17,18,29–31] have investigated dual fuel operation withdifferent gaseous fuels (hydrogen, methane, propane,CNG, LPG) with respect to engine performance, combus-tion characteristics, exhaust gas emissions and factorsinfluencing them. These factors include the engine loads,diesel substitution, injection timing, intake air temperatureand EGR. They concluded that the prolonged ignitiondelay caused by the presence of gaseous fuel in the com-pression process, the reduction of oxygen concentrationin the charge and the increase in the polytropic index ofthe charge leads to significant changes in combustion char-acteristics, exhaust gas emission, engine performance andfuel consumption. This was confirmed by other researchersin this field [20,21,24]. A considerable number of pastinvestigations concentrated on engine performance andfuel consumption. While those revealed decreases in engineoutput [32], others reported unchanged [19] or evenincreased [12,14,33] output. A loss in thermal efficiencyhad been reported by some authors [25,34], whereas othersstated comparable or higher efficiencies [15,35–38] or lossat low to medium loads but gains at high to full loads[13,33,39,40]. Solutions to improve dual fuel part load havebeen investigated and proposed, such as throttling theintake air charge [41], increased intake air pressure [42],temperature [11,25,43,44], controlled amount and time of

mailto:[email protected]

Nomenclature

CI compression ignitionDI direct injectionIDI indirect injectionNG natural gasCNG compressed natural gasLPG liquefied petroleum gasEGR exhaust gas recirculationDDF diesel dual fuelTDC top dead centerBDC bottom dead centerbTDC before top dead centerCA crank angleUHC unburned hydrocarbonLHV lower heating valueFTIR fourier transform infrared spectroscopyTBN total base number

SDC specific diesel consumption, g/kWhSTEC specific total energy consumption, MJ/kWhDS diesel substitutionA/V surface to volume ratio of combustion chamber,

m�1

md mass of diesel fuel delivered per engine cycle, kg/cycle, mg/cycle

mair mass of air sucked per engine cycle, kg/cyclem�

d DDF diesel mass flow rate in DDF operation, kg/sm�

d D diesel mass flow rate in straight diesel operation,kg/s

(A/F)s stoichiometric fuel air ratio(A/F)s,d stoichiometric air fuel ratio of diesel fuel(A/F)s,biog stoichiometric air fuel ratio of biogasU fuel air equivalent ratiobmep brake mean effective pressure, kPa

P.M. Duc, K. Wattanavichien / Energy Conversion and Management 48 (2007) 2286–2308 2287

pilot injection [11,12,15,35,45] or controlled EGR flow andtemperature [25,40,46–48]. While more information of thetype and composition of gaseous fuels used were provided,less detailed information about engine geometry was men-tioned. This makes it more difficult to assess/analyze thereported results since the engine performance, thermal effi-ciency, diesel substitution and exhaust gas emissionsdepend not only on the physical/chemical properties ofthe gaseous fuels but also on the engines used. In addition,almost all past investigations were conducted with engineson test benches at which engine cooling water and lube oiltemperatures had been controlled to ensure not exceeding apredetermined value. This is contrary to real operationalconditions at which the temperatures may increase to highlevels. The reported information about long term use withdual fuelling has also not been reported clearly/fully.

It seems that dual fuel operation for IDI engines is lesseffective than for DI engines because of too high surface tovolume ratio of the combustion chamber. In addition, thedifference in combustion chamber geometry of this typewould have an effect on the dual fuel characteristic. In thiswork, a comparative investigation between straight dieseland biogas premixed charge diesel dual fuel CI (henceforthcalled DDF) was conducted to obtain clear information.The following aspects were concerned: engine performance,diesel substitution, energy consumption and the effect oflong term use.

2. Description

2.1. Test system

The test system installation is shown schematically inFig. 1. A small single cylinder IDI CI engine KubotaRT120 with specifications shown in Table 1 was used.There was no engine modification except a gas mixer

designed particularly for it and added to the intake mani-fold as a means to introduce biogas. The engine was cou-pled with an alternator to form a system loaded byvariable resistances. Engine load is the product of alterna-tor current and voltage divided by the mechanic-electricityconversion efficiency of this system. This efficiency hadbeen determined prior to this study to ensure correct engineload setting.

The biogas fuel used has been produced by a biogas pro-ducing system at a pig farm in Ratchaburi province, Thai-land. Biogas, with pressure higher than atmospheric, fromvery big cellars of the producing system is led by the pipesystem and introduced to the engine via the gas mixer toform a homogeneous charge prior to combustion. Its flowrate was controlled manually by a regulator and a valvelocated upstream of the mixer.

The consumed intake air and biogas flow rates weremeasured by means of an orifice plate and inclined manom-eters. The engine speed signal was sensed by a photodiodesensor. A data acquisition system and a computer programwas designed and installed to collect engine speed and load,time to consume a fixed diesel fuel volume (43 ± 0.01 cm3)and the temperatures of the intake air, cooling water, lubri-cant oil and exhaust gas at a frequency of 1 Hz. Thesedata were stored in the computer hard disk for off line cal-culation and analysis. The instruments used are listed inTable 2.

2.2. Fuels properties

Thai commercial diesel fuel, with the main propertiesgiven in Table 3, was used throughout this investigation.The properties of the biogas obtained from very big cellarsof the producing system remained nearly unchanged duringthe test period. Its main properties were determined andshown in Table 4. As observed, methane is the main

TEST ENGINE KUBOTA RT120

Generator

SurgeTank

Air Filter

Valve GasFlow Meter

Gas Mixer

AirFlow Meter

Computer

Temp. CoolingWater

Temp.Lube Oil

Temp.Exhaust

Gas

Diesel FuelTank

DieselFlow Meter

FI

SpeedSensor

Current Voltage

A/D Converter

Regulator

Biogas frommain supply system

VariableResistor

P

Ambient Temperature,Pressure, Humidity

Inlet Air

Fig. 1. Diagram of test system.

2288 P.M. Duc, K. Wattanavichien / Energy Conversion and Management 48 (2007) 2286–2308

constituent (73% by volume) of the biogas used, whichmakes it suitable for engines with high compression ratio.In addition, the carbon content is also low (74.8% by mass

Table 1Test engine specification

Engine type Single horizontal cylinder, naturallyaspirated, water cooled, 4-cycle, IDI

Rating output 7.73 kW at 2400 rpmMaximum torque 42 N m at 1500 rpmBore · stroke 94 · 90 mm

Swept volume 624 mm3

Compression ratio 21:1

Combustion system Kubota TVCSPre-chamber volume 21 mm3

Pre/Main volume ratio 67%/33%

Piston-cylinder head clearance 0.85 mm

Inlet valveOpen/close 20/45 CA degInner seat diameter/Lift 34.3 mm/7.5 mm

Exhaust valveOpen/close 50�/15�CAInner seat diameter 29.1 mmLift 7.5 mm

Fuel pump Bosh PFR1KFuel delivery control Centrifugal typeTiming device None

Injector type/injection pressure Single spring Pintle nozzle/140 barStatic injection timing 19–21� CA bTDCCooling system RadiatorLubricant system Forced feed

of methane, compared to 84.7% of diesel), resulting in asignificant decrease in specific soot/CO2 emission.

2.3. Test procedures, definition and examination

The experimental investigation was conducted in twophases. In the first phase, engine performances and fuelconsumptions for both modes of fuelling were determined.The engine was test at steady state with different enginespeeds and loads. The test speeds were 1000, 1200, 1500,1800, 2000 and 2400 rpm. The engine torque was varieduntil the maximum value available at each test speed wasattained. For each test point, a set of parameters for bothfuelling modes was measured. Ambient pressure, differen-tial pressure between ambient and each of two orifices(for air and biogas flow rate measurement) and humiditywere recorded manually, five times for each point. Enginespeed, power output and temperatures of intake air, bio-gas, diesel, cooling water, lubricant oil and exhaust gaswere recorded by the acquisition system with a frequencyof 1 Hz, and 61 samples for these parameters were usedin the calculations. Fuel consumption was also measuredfive times for each point, and the corresponding time wasdetermined by the computer clock. During the tests, thelubricant oil and cooling water levels were monitored;water and oil was added as necessary. At each operationalpoint, firstly, the engine operated with straight diesel as thebaseline, and then, the amount of diesel fuel delivered tothe engine was decreased to an amount as small as possible,accompanied with increased biogas flow, in the dual fuel

Table 4Biogas properties

Constituent By volume By massCO2 19% 37.38%N2 6.5% 8.14%O2 1.5% 2.15%CH4 73% 52.34%H2S 20 ppm

Density 0.9145 kg/m3 (273 K, 1 at)LHV 26.17 MJ/kg(A/F)s,CH4

17.23

Table 2Instrument specification

Measurement Instrument Accuracy/division Sample size

Ambient pressurea Barometer Accuracy 0.5 mmHg/division 1 mmHg 5Ambient humiditya Psychrometer Accuracy 0.05 �C/division 0.1 �C 5Engine speedb Photodiode/transmitter ±3 rev/min 61Engine load: voltageb Transmitter 0.2%/0.5 V 61Engine load: currentb Transmitter 0.5%/0.01 A 61Diesel consumptionb Liquid level detector 43±0.01 cm3

Time Computer clockLube oil temperatureb Thermal couple + transmitter 1%/division 0.1 �C 61Cooling water temperatureb Thermal couple + transmitter 1%/division 0.1 �C 61Exhaust gas temperatureb Thermal couple + transmitter 1%/division 0.1 �C 61Ambient air temperatureb Thermal couple + transmitter 0.5%/division 0.1 �C 61

Air consumptionSurgetemperatureb Thermal couple/transmitter 0.5% 61Differential pressurea Orifice + inclined manometer 0.14 mmH2O/division 1 mm 5

Biogas consumptionBiogas temperatureb Thermal couple/transmitter 0.5%/division 0.1 �C 61Differential pressurea Orifice + inclined manometer 0.08 mmH2O/division 1 mm 5

a Parameter was manually collected five times for each measurement.b Accuracy includes that of the board of the acquisition system. Parameter was collected at frequency of 1 Hz and stored in a computer. Sixty-one

samples were used in calculation.

Table 3Thai commercial diesel properties

Properties Unit Test method (ASTM) Value

Specific gravity – D1298 0.826Cetane number – D613 47 min.Cetane index – D976 47 min.Viscosity at 40 �C cSt D445 1.8 – 4.1Pour point �C D97 10 max.Cloud point �C D2599 16 max.Carbon residue wt.% D4530 0.05 max.Water and sediment vol.% D2709 0.05 max.Ash wt.% D482 0.01 max.Flash point �C D93 52 min.Lubricity by HFRR lm CEC F-06-A-96 460 max.LHV kJ/kg 42,500 min.(A/F)s – 14.5

Table 6EMA 200-h test cycle

Step Speed (rpm) Torque % Rated power Time (min)

1 Rated – 100 602 85% Max. 95 603 90% 28% 25 304 Idle 0 0 30

Test engines run continuously five cycles before shut-down period of 9 h aday. The cycle is repeated to accumulate 200 working hours.

Table 5Durability test cycle

Step Speed(rpm)

Power(kW)

% Ratedpower

Torque(N m)

Time(min)

1 2400 7.730 100 30.76 1452 2400 6.957 90 27.68 603 2400 6.184 80 24.60 60


mode. The amount of biogas introduced was varied manu-ally to achieve the required engine torque and speed. Thismeans that the engine operated with minimum diesel fuelconsumption at these speeds. It is noted that minimum die-sel consumption does not mean minimum energy consump-tion in the case of dual fuel. In the second phase, the engine

was tested for endurance with DDF. With respect toinspecting engine operation at the rated condition sug-gested for diesel, also as a critical area in DDF, the testengine, after run in, followed a durability test cycle (asshown in Table 5), modified from EMA 200-h test cycle(Table 6). After warm up, the engine ran these three cyclesa day continuously to accumulate approximately 13 h ofengine operation before shut down. This procedure wasrepeated to achieve a total 240 h accumulation of opera-tion. Lubricant oil ‘‘CF grade’’, produced by Siam Kubota,was used for this test. During the test, the temperatures ofcooling water, lubricant oil, and exhaust gas were measured,and the levels of cooling water and oil were monitored andadditions made as necessary. The oil was sampled, analyzedand completely changed after approximately 60, 80 and


100 working hours. The oil test methods are shown inTable 7. Lastly, the engine was disassembled, visuallyinspected, measured and rated.

The engine torque, power output, BMEP and fuel con-sumption presented have been corrected to correspondwith those at standard conditions, ISO 3046. Diesel substi-tution, DS, is defined as the ratio between the equivalentdiesel mass flow rate replaced by biogas in DDF operationand the diesel mass flow rate in diesel operation at the sameengine speed and torque, as below

DS ¼ ð1� m�

d;DDF=m�

d;DÞ � 100ð%ÞBrake total energy conversion efficiency, gf,DDF, is definedas the ratio between the engine brake power output and therate of total fuel energy supplied to the engine, as below

gf ;DDF ¼P

m�

dLHVd þ m�

biogLHVbiog

� 100ð%Þ

The total fuel air equivalent ratio, UDDF, is defined as theratio between the actual fuel air ratio and the stoichiome-tric fuel air ratio. Hence, it is the ratio between the stoichi-ometric mass of air required to burn the diesel and biogasinside the cylinder completely and the actual air introducedto the cylinder, as below

UDDF ¼mdðA=FÞs;d þ mbiogðA=FÞs;biog

mair

The above equations are also adequate with diesel fuelling(mbiog = 0), corresponding with subscript ‘‘D’’ of parame-ters U and gf.

Table 7Lubricant oil test methods

Iron D-6595Chromium D-6595Lead D-6595Copper D-6595Aluminium D-6595Nickel D-6595Silver D-6595Molybdenum D-6595Titanium D-6595Silicon D-6595Sodium D-6595Magnesium D-6595

Calcium D-6595Phosphorus D-6595Zinc D-6595

Oxidation FTIRNitration FTIRSulfation FTIRWater FTIRSoot FTIR

Fuel SAWTBN D-4739Viscosity D-445

2.4. Results and discussion

Parameters including brake torque, brake power, spe-cific diesel consumption, specific total energy consumption,brake total energy conversion efficiency, diesel substitution,total fuel air equivalent ratio and volumetric efficiency werecalculated, and their uncertainties were estimated accord-ing to the method described in Ref. [49], with C95% confi-dence and the following assumptions:

– Neglected covariances;– Heating value and stoichiometric air fuel ratio of diesel

and biogas, diesel density, swept volume and dischargecoefficient of the two orifices are considered constants;

– Neglect the uncertainty generated as above parametersare corrected to the standard condition.

The relative uncertainties of all the mentioned parame-ters are in an acceptable range and their highest valuesare shown in Table 8, giving confidence to thisinvestigation.

2.4.1. Full load operation

As expected, at all test speeds, there was no deteriora-tion in DDF engine performance compared with that withdiesel fuelling. Comparisons in engine performance, dieselfuel consumption and energy conversion efficiency are pre-sented in Figs. 2–8 and Table 9. At full load, the maximumdiesel substitution was about 36% at the lowest speed. Itreached a peak of about 48.8% at 1800 rpm before decreas-ing by 8% at rated speed. Energy conversion efficiencies inboth fuelling modes were comparable, even slightly higherwith dual fuelling at 1800 rpm. This may be a result ofhigher heat release rate and shorter combustion durationwith DDF, making the in cylinder peak pressure closer toTDC, a more effective cycle, leading to lower exhaust gastemperature, especially at low speeds. Another factor thatmay contribute to the high energy conversion efficiency atlow/medium speeds is that there is no heat loss due tothe passage of the fraction of biogas in the main combus-

Table 8Highest relative uncertainties

Parameters Diesel operation (%) DDF operation (%)

Engine speed 0.39 0.33bmep 2.51 2.40Torque 2.51 2.40Power 2.87 2.78gf 2.06 2.13SDC 2.86 2.30STEC 2.86 2.14U 2.18 3.17gv 0.20 0.77Toil 1.00 0.21Tw 0.98 1.28Tex 0.81 1.63Tair 1.02 0.86Ds 3.17

Engine Performance @ Full Load

35.93

40.5941.3641.97

38.6637.51

35.89

40.5241.5141.98

38.6837.54

754777

844832

816755778

844 835815

722

72330.3

33.6

36.8

40.1

43.3

900 1100 1300 1500 1700 1900 2100 2300 2500Engine speed (rev/min.)

Engi

ne T

orqu

e (N

m)

690

740

790

840

890

940

990

bmep

(kPa

)

Torque_Diesel Torque_DDFbmep_Diesel bmep_DDF

Fig. 2. Full load performance in two fuelling modes.

Specific Diesel Consumption & Substitution

290292295297313346

172169151168198223

35.6 36.8

43.5 42.1 40.748.8

0

250

500

750

900 1100 1300 1500 1700 1900 2100 2300 2500Engine Speed (rev/min)

Con

sum

ptio

n (g

/kW

h)

10

30

50

Subs

titut

ion

(%)

SDC_Diesel SDC_DDF Substitution

Fig. 3. Specific diesel consumption and substitution at full load in two fuelling modes.

Brake Total Energy Conversion Efficiency

24.5

27.1

28.5

28.7

29.0 29.2

24.0

26.7

28.4

29.1

28.5 28.5

23

27

31


Effic

ienc

y (%

)

η f, D η f, DDF

Fig. 4. Brake total energy conversion efficiency at full load in two modes.


tion chamber as combustion occurs. With the IDI KubotaTVCS combustion system, the heat loss of this biogas frac-tion while contacting with the chamber wall may be less

since the swirl motion of the fluid in the main chamber isvery low. When engine speed increases, this benefit maybe offset by the heat transfer of that fraction to the cylinder

Full load Fuel Air Equivalent Ratio

0.85

0.90

0.94

0.810.80

0.85

0.90

0.950.95

0.830.83

0.89

0.7

0.8

0.9

1.0


Fuel

-Air

Equi

vale

nt R

atio

Φ DDFΦ Diesel

Fig. 5. Full load total fuel air equivalent ratio in two fuelling modes.

Exhaust Gas Temperature

567570

572551508486

538506

412380326

268

45%

36%31%

28%

11%

5%

100

400

700

1000


Tem

pera

ture

(°C

)

Tem

pera

ture

Cha

nge

(%)Tex_Diesel

Tex_DDFChange (DDF-Diesel)

Fig. 6. Exhaust gas temperatures at full load in two fuelling modes.

Cooling Water Temperature

102101100100

9796

104104

101101

9897

1.92%1.02% 0.99% 0.99%

2.88%1.92%

90

95

100

105

110

115

900 1100 1300 1500 1700 1900 2100 2300 2500Engine speed (rev/min)

Tem

pera

ture

(°C

)

Tem

pera

ture

Cha

nge

(%)

Tw_Diesel Tw_DDF Change (DDF-Diesel)

Fig. 7. Cooling water temperatures at full load in two fuelling modes.


head wall, leading to divergence of the two efficiencies. Themain factor contributing to this divergence would be theeffect of the prolonged ignition delay in dual fuelling. The

presence of gaseous fuel with the air charge, during theintake and compression process leads to: 1 a decrease ofoxygen concentration; 2 a decrease of charge temperature

Lube Oil Temperature

102102

969694

86

104103

979797

88

2.27%3.09%

1.03% 1.03% 0.97%1.92%

85

95

105

115

900 1100 1300 1500 1700 1900 2100 2300 2500Engine Speed (rev/min)

Tem

pera

ture

Cha

nge

(%)

Toil_DieselToil_DDFChange (DDF-Diesel)

Tem

pera

ture

(°C

)

Fig. 8. Lube oil temperatures at full load in two fuelling modes.

Table 9Full load performance with the two cases of fuelling

Diesel fuellingEngine speed (rpm) 1000.6 1200.6 1500.6 1801.4 2001.0 2401.5Ambient air temperature (�C) 33.1 26.1 28.1 33.2 36.2 31.1Brake torque (N m) 37.51 38.66 41.97 41.36 40.59 35.93bmep (kPa) 754 777 844 832 816 723

Dual fuellingEngine speed (rpm) 1000.6 1200.6 1500.6 1801.3 2001.1 2401.5Ambient air temperature (�C) 33.2 26.1 28.1 33.2 36.1 31.2Brake torque (N m) 37.54 38.68 41.98 41.51 40.52 35.89bmep (kPa) 755 778 844 835 815 722


at the time of starting injection due to the lower polytropicindex of biogas (1.305 compared to 1.400 of air); and 3 thepre-ignition reactions of the biogas–air–residual gas mix-ture [11,18]. These factors may cause ignition delay toincrease, leading to a higher pressure rise rate in the pre-mixed combustion phase of diesel fuel and, hence, lowerdiesel substitution due to the end gas knock limit. The pro-longed ignition delay then causes the combustion processto shift some degrees toward BDC, leading to less effectivecycles. While the efficiency, in the diesel mode, increasedwith increased speed, it had a peak of 29.1% at 1800 rpmin the DDF mode.

The high energy conversion efficiency in the dual fuellingmode was also revealed with the decreasing trend in exhaustgas temperature, as shown in Fig. 6. The exhaust gas tem-perature was lower than that in diesel fuelling at all testspeeds, especially at the low speed range. As the combustiontakes place closer to TDC, a larger fraction of the fuelenergy is converted to work and, to some extent, heat trans-fer. This leads to lower exhaust gas temperature, and hence,the loss of energy brought by exhaust gas decreased.

While the exhaust gas temperature revealed a decreasingtrend, the temperature of the cooling water and lube oilrevealed the reverse, always higher than those in diesel fuel-ling, as seen in Figs. 7 and 8. It is noted that these trendsoccur at comparable intake air temperatures of the two

cases of fuelling as in Table 9. The highest increase in cool-ing water temperature, at 2000 rpm, was 3 �C, accountingfor approximately 3%. Similarly, the highest increase inlube oil temperature was 3 �C, accounting for approxi-mately 3%, at 1200 rpm. The increases in cooling waterand oil temperatures might result from the higher maxi-mum temperature in the combustion chamber, hencehigher heat transfer across the combustion chamber wallto the cooling water and engine block. Although dieselhas the higher stoichiometric flame temperature (2300 K)than that of methane (2250 K), the combustion tempera-ture in the case of dual fuelling could be higher than thatin the case of diesel fuelling because of the higher equiva-lent ratio of the homogeneous charge mixture as shownin Fig. 5.

2.4.2. Part load operation

A bird’s eye view of the comparison in engine operationwith the two cases of fuelling is revealed by Figs. 9–21. Thegeneral trends as the engine operated with biogas–dieselfuelling were higher fuel air equivalent ratio, lower volu-metric efficiency and energy conversion efficiency, higherlube oil and cooling water temperatures and lower exhaustgas temperature.

Compared to diesel fuelling, diesel substitution decreasedfrom a high value of about 93–94% at low load to low values

1000 1200 1400 1600 1800 2000 2200 2400

5

10

15

20

25

30

35

40

45

Engine speed (rev/min)

Brak

e To

rque

(Nm

)

Diesel Substitution (%)

4045 45

50

55

60

65

70

75

80

85

9095

Fig. 9. Maximum diesel substitution in dual fuelling.

1000 1200 1400 1600 1800 2000 2200 2400

5

10

15

20

25

30

35

40

45


Bra

ke T

orqu

e (N

m)

Fuel Air Equivalent Ratio (Diesel Fuelling)

0.25

0.90.85

0.8

0.750.7

0.65

0.6

0.550.5

0.45

0.4

0.35

0.3

Fig. 10. Fuel air equivalent ratio (diesel fuelling).


of about 43–49% at high/full loads at constant speeds asobserved in Fig. 9. However, it is noticed that the minimummass of diesel delivered per cycle increased with increasedload at constant speeds, namely 22–25 mg/cycle as observed

in Figs. 22, 26, 30 and 34. The maximum diesel substitutionis limited due to the end gas knock limit [12,19,21,35,50,51].With dual fuelling, the total fuel air equivalent ratio isalways higher than that with diesel fuelling, resulting from

1000 1200 1400 1600 1800 2000 2200 2400

5

10

15

20

25

30

35

40

45


Bra

ke T

orq

ue

(N

m)

Total Fuel Air Equivalent Ratio (Dual Fuelling)

0.85

0.9

0.85

0.8

0.75

0.7

0.650.6

0.55

0.5

0.45

0.4

0.5

0.5

0.7

Fig. 11. Total fuel air equivalent ratio (DDF).

1000 1200 1400 1600 1800 2000 2200 2400

5

10

15

20

25

30

35

40

45


Bra

ke T

orqu

e (N

m)

Brake Energy Conversion Efficiency (Diesel)

131619

2224

26

28

29

29.53030.531

31.5

2928

26

26

24

22

Fig. 12. Brake energy conversion efficiency (diesel fuelling).


the replacement of a fraction of the air charge by biogas.This increased fuel air equivalent ratio has some importanteffects. First, the gaseous fuel replaces a fraction of the dieselliquid fuel and forms a homogeneous mixture, leading to ahigher combustion rate (of the homogeneous mixture, and

the rate increases with increased gaseous fuel air ratio),reduced diffusion diesel combustion and reduced wallimpingement of diesel, hence improving the total fuel con-version efficiency and producing less soot. The soot reduc-tion is proportional to the diesel substitution [19]. Second,

1000 1200 1400 1600 1800 2000 2200 2400

5

10

15

20

25

30

35

40

45


Bra

ke T

orqu

e (N

-m)

Brake Total Energy Conversion Efficiency (Dual Fuelling)

3029.529

28

26

24

22

19

16

13 10

26

29

Fig. 13. Brake total energy conversion efficiency (DDF).

1000 1200 1400 1600 1800 2000 2200 2400

5

10

15

20

25

30

35

40


Bra

ke T

orqu

e (N

m)

Specific Diesel Consumption (g/kWh) (Diesel Fuelling)

700

500

400

350

320

300

290

270

285280

290320

300 290

285

275

273 273

400

350

320

Fig. 14. Specific diesel consumption (diesel fuelling).


increasing diesel substitution leads to a relatively larger frac-tion of gaseous fuel occupying the main chamber. As itburns, there is no heat loss due to the connecting passage.Third, on the opposite side, increasing the equivalent ratioleads to lower fuel conversion efficiency. With too lean bio-gas–air mixture, the flame front can not propagate fast

enough and far enough to consume the entire mixture withinthe time period available [11,25,29–31], leading to higherUHC, carbon monoxide emission and higher energy con-sumption. In IDI engines, this phenomenon is more severesince they have high surface to volume ratio of the combus-tion chamber, especially with the main chamber. As the mix-

1000 1200 1400 1600 1800 2000 2200 2400

5

10

15

20

25

30

35

40

45


Bra

ke to

rque

(N

m)

Specific Diesel Consumption (g/kWh) (Dual Fuelling)

5040

1520 25 30

40

50

607080

100

80

100

120

130150

150

200

130

Fig. 15. Specific diesel consumption (DDF).

1000 1200 1400 1600 1800 2000 2200 2400

5

10

15

20

25

30

35

40

45


Bra

ke T

orqu

e (N

m)

Exhaust Gas Temperature (oC) (Diesel Fuelling)

210

250

300

350 400

450

500

550

Fig. 16. Exhaust gas temperature (diesel fuelling).


ture strength reaches a certain level, depending on the pre-vailing condition in the combustion chamber, the combus-tion duration of the gas becomes shorter, more completesince the flame spread speed increases. Fourth, the pro-longed ignition delay causes the combustion process to last

later with respect to TDC, leading to a less effective cycle.The net effect of the above factors drives the trend in totalenergy conversion efficiency of the DDF engine. It produceslower total energy conversion efficiency in DDF at low/med-ium load ranges, although the reverse trend may be at

1000 1200 1400 1600 1800 2000 2200 2400

5

10

15

20

25

30

35

40

45


Bra

ke T

orqu

e (N

m)

Exhaust Gas Temperature (oC) (Dual Fuelling)

180

210

250

300

350

400

450500

Fig. 17. Exhaust gas temperature (DDF).

1000 1200 1400 1600 1800 2000 2200 2400

5

10

15

20

25

30

35

40

45


Bra

ke T

orqu

e (N

m)

Cooling Water Temperature (oC) (Diesel Fuelling)

99101

979595

95

93

91

88

90

8684

84

82

78 80 80 78 76

99

95

939188869

Fig. 18. Cooling water temperature (diesel fuelling).


higher/full loads. With diesel fuelling, a relatively high effi-ciency of 28% could be reached with almost all engine speedsin the range of 1200–2200 rpm and engine loads from med-

ium value (about 22.5 N m), Fig. 12, but from a higher value(about 30 N m) with dual fuelling, as shown in Fig. 13. Atlevels higher than about 50% of maximum load, the deteri-

1000 1200 1400 1600 1800 2000 2200 2400

5

10

15

20

25

30

35

40

45


Bra

ke T

orqu

e (N

m)

Cooling Water Temperature (oC) (Dual Fuelling)

105

103

97 9910195

93

9190

88

86

91 97959391

84

82807876

82

Fig. 19. Cooling water temperature (DDF).

1000 1200 1400 1600 1800 2000 2200 2400

5

10

15

20

25

30

35

40

45


Bra

ke T

orqu

e (N

m)

Lube Oil Temperature (oC) (Diesel Fuelling)

91

95 97 99101

101101

99

979593

91

91

90

88

86

84

8286

8076

76

78

Fig. 20. Lube oil temperature (diesel fuelling).


oration in energy conversion efficiency of the engine in DDFmode (about 3%) is less than that of other engines of the DItype due to the contribution of the large fraction of the gas-eous fuel in the main chamber. It is also noted that the DDFoperation was established with respect to minimizing the

diesel fuel used. Hence, the deterioration will be lower ifthe target of energy conversion efficiency is concerned.

One important result is that, accompanied with adecrease of about 1.5%, the high efficiency island of DDFmoved to a higher speed and load area compared with that

1000 1200 1400 1600 1800 2000 2200 2400

5

10

15

20

25

30

35

40

45


Bra

ke T

orqu

e (N

m)

Lube Oil Temperature (oC) (Dual Fuelling)

105

103

97 99101

101

9795939190

91

95

93

91

90

88

86

84

8278

80

76

7880 82

97

97

Fig. 21. Lube oil temperature (DDF).


of diesel fuelling. The occurrence of a high efficiency islandin DDF mode at higher engine torque manifests that thepositive effect of increased mechanical efficiency, largerfraction of gaseous fuel in the main chamber and highercombustion efficiency at higher combustion chamber tem-perature dominates, is higher than, the negative effect ofthe importance of heat transfer from the working fluid tothe combustion chamber wall. Similarly, the occurrenceat the higher engine speed area reveals that the positiveeffect of the larger fraction of gaseous fuel in the mainchamber, higher combustion efficiency at higher combus-tion chamber temperature and decreased importance ofheat loss due to decreased cycle time exceed the negativeeffect of increased friction loss.

The comparison in exhaust temperature in the two fuel-ling cases, Figs. 16 and 17, clarifies this point more. At thesame given load and speed, in DDF mode, the exhaust tem-perature was always lower. In addition, from the consid-ered point, the slopes of the contour lines were higherthan those in straight diesel mode. It is then inferred that,in the DDF mode, the positive effect on combustion effi-ciency due to the replacement of a fraction of the liquid die-sel fuel by gaseous fuel and the occupying by a fraction ofgaseous fuel in the main chamber at relatively high com-pression ratio dominates as the operational engine speedincreases since lesser heat energy loss is brought by theexhaust gas. At higher engine speeds and torques, accom-panied with higher work produced, the higher combustiontemperature would cause an increase in heat transfer to thecylinder head and cylinder wall, leading to a higher temper-ature of cooling water and lubricant oil. This would result

in a very high thermal load to the engine. As observed, acritical area at which high oil and cooling water tempera-ture occurred corresponded with approximately 2000 rpmand above of engine speed and about 27.5 Nm of brake tor-que (580 kPa of bmep). This area is marked by the red con-tour line shown in Figs. 19 and 21. In the remaining area ofengine operating condition, except full load at all speedsand medium load with speeds higher than 2000 rpm,DDF produced relatively lower temperatures of coolingwater and lubricant oil. Enveloping the critical area wasthe one revealing very high temperature gradients withrespect to load and speed. With Figs. 22–37, the detailedcomparison between the two cases of fuelling at four speedsof 1500, 1800, 2000 and 2400 rpm is presented. Asobserved, the volumetric efficiency in DDF was alwayslower than that in diesel fuelling due to the increased pres-sure drop caused by the gas mixer. Because of this, lube oilwas sucked into the inlet port since there is no seal for theinlet valve of this engine type, resulting higher lube oil con-sumption. At fixed engine speed, with increased loads,while the volumetric efficiency gradually decreased in thediesel case, it might fluctuate in the DDF case because ofthe gaseous fuel introduced.

2.4.3. Endurance test

The engine endurance test revealed the following results:

– Lubricant oil consumption was very high, at levels unac-ceptable, as shown in Fig. 38. This was due to highercooling water and lube oil temperatures. Moreover,the increase in pressure drop in the intake system with

Diesel Fuel Delivery @ 1500 rev/min

45403431

272320181613 12

864311

1620

25

43%

50%53%

61%

71%73%78%

84%94%

92%

0

20

40

60

80

100

120

5 10 15 20 25 30 35 40 45Brake Torque (Nm)

Die

sel I

nje

ctio

n (

mg

/cyc

le)

0%

10%

20%

30%

40%

50%

60%

70%

80%

90%

100%

Die

sel S

ub

stit

uti

on

(%

)

Diesel Injection (Diesel)

Diesel Injection (DDF)

Diesel Substitution

Fig. 22. Comparison in diesel delivery and substitution at 1500 rpm in two fuelling cases.

Fuel Air Equivalent Ratio @ 1500 rev/min

0.230.27

0.320.36

0.420.48

0.55

0.61

0.72

0.81

0.39 0.39 0.40

0.470.50

0.590.63

0.680.74

0.83

0.00

0.20

0.40

0.60

0.80

1.00

Brake Torque (Nm)

Fu

el A

ir E

qu

ival

ent

Rat

io

Φ Diesel Φ DDF

5 10 15 20 25 30 35 40 45

Fig. 23. Comparison in fuel air equivalent ratio at 1500 rpm in two fuelling cases.

Volumetric & Energy Conversion Efficiency @ 1500 rev/min

0.167

0.2110.243

0.300 0.310 0.319 0.3110.303 0.297 0.285

0.106

0.1630.205

0.2390.263 0.273 0.281 0.296 0.3000.284

1.14 1.171.12 1.11 1.11 1.11 1.11 1.10 1.10 1.09

1.07 1.08 1.07 1.06 1.07 1.06 1.07 1.06 1.06 1.06

0.00

0.70

En

erg

y C

on

vers

ion

Eff

icie

ncy

0.3

0.6

0.8

1.1

1.3

Vo

lum

etri

c E

ffic

ien

cy

ηv,Diesel ηv,DDF

ηf,Diesel ηf,DDF

Brake Torque (Nm)5 10 15 20 25 30 35 40 45

Fig. 24. Comparison in volumetric and energy conversion efficiency at 1500 rpm in two fuelling cases.


Specific Total Energy Consumption @ 1500 rpm

21.69

17.0114.81

11.99 11.70 11.30 11.60 11.37 12.12 12.62

33.82

22.02

17.5515.09

13.70 13.19 12.81 12.18 12.01 12.68

0

10

20

30

40

5 10 15 20 25 30 35 40 45

Brake Torque (Nm)

To

tal E

ner

gy

Co

nsu

mp

tio

n (

MJ/

kWh

)

STEC_Diesel STEC_DDF

Fig. 25. Comparison in specific total energy consumption at 1500 rpm in two fuelling cases.


13 15 16 18 21 2326 29 32 35

39

1

43

43215 8 10 13

15 1822

53%49%

92% 93%

90% 86%83%

76%

68%65%

59% 56%

0

20

40

60

80

100

120

5 10 15 20 25 30 35 40 45

Brake Torque (Nm)

Die

sel I

nje

ctio

n (

mg

/cyc

le)

0%

10%

20%

30%

40%

50%

60%

70%

80%

90%

100%

Die

sel S

ub

stit

uti

on

(%

)

Diesel Injection (Dies el)


Diesel Substitution

Fig. 26. Comparison in diesel delivery and substitution at 1800 rpm in two fuelling cases.


0.260.30

0.350.39

0.440.49

0.560.61

0.680.74

0.84

0.94

0.42 0.440.51 0.54 0.54

0.610.65

0.690.75

0.79

0.87

0.95

0.00

0.40

0.80

1.20

5 10 15 20 25 30 35 40 45

Brake Torque (Nm)

Fu

el A

ir E

qu

ival

ent

Rat

io

Φ Diesel Φ DDF




0.1460.187

0.2220.257

0.277 0.296 0.300 0.310 0.3080.306 0.303

0.0960.139

0.1600.199

0.236 0.2490.268 0.285 0.287 0.298 0.302

0.287

0.291

0.93 0.91 0.91 0.91 0.92 0.91 0.91 0.90 0.92 0.91

0.90 0.89 0.87 0.87 0.87 0.87 0.88 0.88 0.88 0.88 0.89 0.90

0.95 0.96

0.00

0.25

0.50

0.75

5 10 15 20 25 30 35 40 45

Brake Torque (Nm)

En

erg

y C

on

vers

ion

Eff

icie

ncy

0.3

0.5

0.7

0.9

1.1

Vo

lum

etri

c E

ffic

ien

cy

ηv,Diesel ηv,DDF

ηf,Diesel ηf,DDF



24.5

19.3

16.214.0 13.0 12.2 12.1 11.6 11.8 11.6 11.9 12.5

25.9

22.4

18.1

15.3 14.5 13.4 12.6 12.5 12.1 11.9 12.4

37.4

0

10

20

30

40

5 10 15 20 25 30 35 40 45Brake Torque (Nm)

To

tal E

ner

gy

Co

nsu

mp

tio

n (

MJ/

kWh

)


Fig. 29. Comparison in total energy consumption at 1800 rpm in two fuelling cases.


41383532292724222018161413

8742111

10 12 14 16 1924

54%50%

42%

56%59%63%

67%70%

79%

87%91%

93%

91%

0

40

80

120

0 5 10 15 20 25 30 35 40 45Brake Torque (Nm)

Die

sel I

nje

ctio

n (

mg

/cyc

le)

0%

10%

20%

30%

40%

50%

60%

70%

80%

90%

100%

Die

sel S

ub

stit

uti

on

(%

)



Diesel Substitution

Fig. 30. Comparison in diesel delivery and diesel substitution at 2000 rpm in two fuelling cases.



0.27 0.300.33

0.370.41

0.460.51

0.570.62

0.68

0.75

0.82

0.90

0.53 0.52 0.54 0.560.59

0.630.68

0.73 0.730.77

0.840.88

0.95

0.10

0.30

0.50

0.70

0.90

1.10

0 5 10 15 20 25 30 35 40 45

Brake Torque (Nm)

Fuel

Air

Eq

uiva

len

t R

atio

Φ Diesel Φ DDF



0.13

0.17

0.21

0.250.26

0.28 0.29 0.29 0.30 0.30 0.30 0.30

0.10

0.13

0.170.19

0.210.22 0.23

0.27 0.28 0.29 0.29

0.29

0.28

0.07

0.93 0.93 0.93 0.93 0.93 0.93 0.92 0.92

0.91 0.91 0.90 0.89 0.90 0.90 0.88 0.90 0.90 0.910.88 0.89 0.90

0.930.930.94 0.92 0.92

0.05

0.18

0.30

0.43

0.55

0 5 10 15 20 25 30 35 40 45Brake Torque (Nm)

En

erg

y C

onv

ersi

on E

ffic

ien

cy

0.3

0.5

0.7

0.9

1.1

Vo

lum

etri

c E

ffic

ien

cyηv,Diesel ηv,DDF

ηf,Diesel ηf,DDF



28.1

21.0

17.314.5 13.7 13.2 12.6 12.6 12.0 12.0 12.0 12.1 12.4

35.7

26.7

20.718.9

17.3 16.1 15.613.5 13.0 12.6 12.3 12.7

54.0

0

15

30

45

60

0 5 10 15 20 25 30 35 40 45Brake Torque (Nm)

To

tal E

ner

gy

Co

nsu

mp

tio

n (

MJ/

kWh

)





13 15 16 17 19 21 23 24 26 28 30 33 36 37

22191614131110

1 1 1 1 2 49

54%50%

46%41%

93% 94% 94%

92% 90%

81%

62% 60%56% 55%

0

40

80

120

0 5 10 15 20 25 30 35 40Brake Torque (Nm)

Die

sel I

nje

ctio

n (

mg

/cyc

le)

0%

10%

20%

30%

40%

50%

60%

70%

80%

90%

100%

Die

sel S

ub

stit

uti

on

(%

)



Diesel Substitution

Fig. 34. Comparison in diesel delivery and diesel substitution at 2400 rpm in two fuelling cases.


0.290.32

0.360.38

0.430.47

0.510.55

0.590.63

0.680.73

0.800.85

0.49 0.49 0.50 0.52

0.580.63

0.68 0.70 0.70 0.72 0.730.78

0.85 0.90

0.10

0.40

0.70

1.00

0 5 10 15 20 25 30 35 40Brake Torque (Nm)

Fu

el A

ir E

qu

ival

ent

rati

o

Φ Diesel Φ DDF



0.10

0.14

0.17

0.210.23

0.250.26 0.26

0.28 0.28 0.29 0.29 0.29

0.10

0.13

0.160.18

0.19 0.200.21

0.240.26

0.27 0.28 0.29

0.29

0.06

0.28

0.90 0.89 0.90 0.89 0.89 0.89 0.89 0.89 0.89 0.89 0.89 0.89 0.88

0.88 0.86 0.86 0.86 0.86 0.85 0.86 0.87 0.86 0.86 0.87 0.86 0.85

0.87

0.84

0.05

0.25

0.45

0 5 10 15 20 25 30 35 40Brake Torque (Nm)

En

erg

y C

on

vers

ion

Eff

icie

ncy

0.2

0.4

0.6

0.8

1.0

Vo

lum

etri

c E

ffic

ien

cy

ηv,Diesel ηv,DDF

ηf,Diesel ηf,DDF




25.8

21.1

17.1 15.9 14.7 14.1 13.7 13.0 12.7 12.6 12.4 12.3

27.9

21.920.2 18.6 18.1 17.0

14.8 13.9 13.2 12.8 12.6

12.3

34.4

12.6

37.6

56.2

0

15

30

45

60

0 5 10 15 20 25 30 35 40

Brake Torque (Nm)

To

tal E

ner

gy

Co

nsu

mp

tio

n (

MJ/

kWh

)



Lubricant Oil Consumption

0

20

40

60

80

100

120

140

160

180

13 21 33 46 62 70 82 96 108

122

135

144

157

166

179

191

206

217

230

243

Time of Measurement (hour)

Oil

con

sum

pti

on

(m

l/ho

ur)

Fig. 38. Lube oil consumption in endurance test.


a gas mixer installed resulted in the increase in suction oflube oil to the cylinder since there was no seal at theintake valve stem. The increase in lube oil consumptionrequired higher added amount of lube oil after each testcycle.

– Lube oil viscosity increased with oil working time. How-ever, after 100 h, it was in an acceptable range. Thismight be due to a larger amount of oil being added dur-ing the test.

Table 10Oil analysis results

Engine working time (h) Lubricant oil change

First time Second time

3.86 16.41 33.3 61.5 78.5 100.7

Iron – – C C C –Chromium A A A A A AAluminium – C A A A ASilicon C C C C C CViscosity (100 �C) – – – – – –

C: caution (first level warning limit), A: abnormal (second level warning limit

– The factors having high concentration in lube oil arepresented in Table 10. It is noted that not all the oil sam-ple results are in that table. From the beginning, thechromium concentration was at an abnormal level (A);second level warning limit. This revealed that the firstand fourth piston ring wear was very high. Iron concen-tration was at the first level warning (C) after 33.3 h andat abnormal level (A) after 144 h of the test. Aluminiumconcentration was at the first level warning (C) after

Note

Third time

121.5 144 183 226.6 242.5 C A

C A A A A >25 ppm >40 ppmA A A A A >1 ppm >2 ppmA A A A A >4 ppm >6 ppmC C C C C >15 ppm >25 ppm– – – – – <12.7 sCt >13.4 sCt

).

Fig. 39. Destruction of piston crown due to high thermal load.


16 h and at abnormal level (A) after about 33 h of thetest. Aluminium in the lube oil results from destructionof the piston. Fig. 39 shows the piston picture after theendurance test. Since the engine operated at higher fuelair equivalent ratio and combustion took place in thehomogeneous charge, the combustion period might beshorter and combustion temperature might becometoo high for the piston and ring to withstand. Siliconconcentration was always at the first level warning (C).

3. Conclusion

An experimental investigation of an ‘‘unmodified’’ smallIDI biogas premixed charge diesel CI dual fuelling enginewas conducted with the concern on engine performance,maximizing diesel fuel substitution, energy consumptionand long term use. The following results were obtainedand concluded.

a. Biogas premixed charge diesel dual fuelling for theengine produced almost no performance deteriora-tion at all test speeds.

b. The DDF mode produced lower energy conversionefficiency, which was offset by large replacement ofdiesel by biogas that has relatively low cost and isa renewable energy source. The efficiency deteriora-tion reduced when engine load increased. At fullload, the efficiency was comparable with that in die-sel fuelling. It is then inferred that at low/mediumloads, the DDF engine produced higher UHC andless soot.

c. The DDF high efficiency island moved to higherengine speeds and loads, revealing the effect of a frac-tion of gaseous fuel occupying the main combustionchamber.

d. The DDF mode resulted in lower exhaust gas temper-ature regardless of engine load and speed, highercooling water and lube oil temperatures at high loadsand high engine speeds. These changes are thought tobe due to the shorter combustion period broughtabout by DDF. A critical area was observed withvery relatively high temperatures of lubricant oiland cooling water.

e. The endurance test revealed that lube oil consump-tion was high, at unacceptable levels, due to theincreased oil and cooling water temperatures and itssuction to the cylinder as a gas mixer was installed.The engine could not withstand the higher thermalload brought by faster DDF burning at the enginespeeds and loads proposed for diesel fuel.

f. The DDF engine with high diesel substitution shouldavoided operating at the critical area. For safeengine operation at this area, reduced substitutionis needed.

Acknowledgements

The authors would like to express thanks to the SiamKubota Industry Co., Ltd. for their supporting this inves-tigation. Thanks are also to Mr. Kritchai Cojchaplayuk inestablishing and conducting this experiment.

References

[1] Jiang C, Liu T, Zhong J. A study on compressed biogas and itsapplication to the compression ignition dual-fuel engine. Biomass1989;20:53–9.

[2] Leif G, Pal B, Bengt J, Per S. Reducing CO2 emission by substitutingbiomass for fossil fuels. Energy 1995;20:1097–113.

[3] Henham A, Makkar MK. Combustion of simulated biogas in a dualfuel diesel engine. Energy Convers. Mange. 1998;39:2001–9.

[4] Modarres RMR, Karim GA. Examination of the dual-fuel engineperformance using low BTU gaseous fuels. In: Proceedings of the13th international conference on thermal engineering and thermo-grammetry, Budapest, Hungari; 2003.

[5] Kapur T, Kandpal TC, Garg HP. Electricity generation from ricehusk in Indian rice mills: potential and financial viability. BiomassBioenergy 1996;10:393–403.

[6] John GC, Carol RP. Independent power plant using wood waste.Energy Convers Mange 1996;37:1205–9.

[7] Rejeev MJ, Anil KR. Development of a sugarcane leaf gasifier forelectricity generation. Biomass Bioenergy 1995;8:91–8.

[8] Uma R, Kandpal TC, Kishore VVN. Emission characteristics of anelectricity generation system in diesel alone and dual fuel modes.Biomass Bioenergy 2004;27:195–203.

[9] Vyarawalla F, Parikh PP, Dak HC, Jain BC. Utilisation of biomassfor motive power generation – gasifier engine system. Biomass1984;5:227–42.

[10] Parikh PP, Bhave AG, Kapse DV, Shashikantha. Study of thermaland emission performance of small gasifier dual-fuel engine systems.Biomass 1989;19:75–97.


[11] Karim GA. A review of combustion processes in the dual fuel engine– the gas diesel engine. Progr Energy Combust Sci 1980;6:277–85.

[12] Selim MYE. Sensitivity of dual fuel engine combustion and knockinglimits to gaseous fuel composition. Energy Convers Mange2004;45:411–25.

[13] Poonia MP, Ramesh A, Gaur RR. Effect of intake air temperatureand pilot fuel quantity on the combustion characteristics of a LPGdiesel dual fuel engine. SAE 982455; 1998.

[14] Slawomir L. The influence of regulating parameters of dual fuelcompression ignition engine fuelled with LPG on its maximumtorque, overall efficiency and emission. SAE 2001-01-3264; 2001.

[15] Abd Alla GH, Soliman HA, Badr OA, Abd Rabbo MF. Effect ofinjection timing on the performance of a dual fuel engine. EnergyConvers Mange 2002;43:269–77.

[16] Selim MYE.A study of some combustion characterstics of dual fuelengine using EGR. SAE 2003-01-0766; 2003.

[17] Karim GA. Examination of some measures for improving theperformance of gas fuelled diesel engines at light load. SAE Trans1991.

[18] Karim GA, Ito K, Abraham M, Jensen L. Examination of the role offormaldehyde in the ignition processes of a dual fuel engine. SAETrans 1991;100:975–82.

[19] Wong WY, Clark MK, Stuart RB. Performance and emissions of aNG dual-fueled, indirect injected diesel engine. SAE 911766; 1991.

[20] Naber JD, Siebers DL, Di Julio SS, Westbrook CK. Effects of naturalgas composition on ignition delay under diesel conditions. CombustFlame 1994;99:192–200.

[21] Nwafor OMI, Rice G. Combustion characteristics and performanceof natural gas in high speed indirect injection diesel engine.Renewable Energy 1994;5:841–8.

[22] Balasubramanian V, Sridhara K, Ganesan V. Performance evaluationof a small agricultural engine operated on dual fuel NG system. SAE951777; 1995.

[23] Yasuhiro D, Toru Y, Takahisa K, Takeshi S, KiharaEdwin NQ.Combustion and exhaust emissions in a direct-injection diesel enginedual-fueled with natural gas. SAE Paper 950465; 1995.

[24] Mtui PL, Hill PG. Ignition delay and combustion duration withnatural gas fueling of diesel engines. SAE 961933; 1996.

[25] Kusaka J, Daisho Y, Kihara R, Saito T. Combustion and exhaust gasemissions characteristics of a diesel engine dual-fueled with NG. In:Proceedings of the COMODIA, Japan, 1998.

[26] Papagiannakis RG, Hountalas DT. Combustion and exhaust emis-sion characteristics of a dual fuel compression ignition engineoperated with pilot diesel fuel and natural gas. Energy ConversMange 2004;45:2971–87.

[27] Gopal G, Srinivasa RP, Gopalakrishnan KV, Murthy BS. Use ofhydrogen in dual-fuel engines. Int J Hydrogen Energy 1982;7:267–72.

[28] Haragopala RB, Shrivastava KN, Bhakta HN. Hydrogen for dualfuel operation. Int J Hydrogen Energy 1983;8:381–4.

[29] Karim GA. Exhaust emission from dual fuel engines at light load.SAE 932822; 1993.

[30] Karim GA, Badr O, Liu B. An examination of the flame spread limitsin a dual fuel engine. Appl Therm Eng 1999;19:1071–80.

[31] Karim GA, Wong YK. An analytical examination of the effects ofexhaust gas recirculation on the compression ignition process ofengines fuelled with gaseous fuels. SAE 961936; 1996.

[32] Zhang Y, Liu X, Yang Q, Han X, Cheng C. The studies of anelectronically controlled CNG system for dual fuel engines. SAE2001-01-0145; 2001.

[33] Chengji Zuo MY. Operating characteristics and description of a dualfuel engine for diesel–NG heavy-duty operation. SAE Paper 1999-01-3523; 1999.

[34] Hountalas DT, Papagiannakis RG. Development of a simulationmodel for direct injection dual fuel diesel–NG engines. SAE 2000-01-0286; 2000.

[35] Abd Alla GH, Soliman HA, Badr OA, Abd Rabbo MF. Effect ofpilot fuel quantity on the performance of a dual fuel engine. EnergyConvers Mange 2000;41:559–72.

[36] Shinichi G, Hirohide F, Rafael DD. Dual fuel dieselsel engine usingbutane. SAE 920690; 1992.

[37] Sudhir CV, Desai V, Suresh Kumar Y, Mohanan P. Performance andemission studies on the effect of injection timing and diesel replace-ment on a 4-S LPG–diesel dual-fuel engine. SAE 2003-01-3087; 2003.

[38] William PJ, John BN, Van der Lee A, Lovkov O, Koshin VK, PiatovIS. All electronic dual fuel injection system for the Belarus D-144diesel engine. SAE 901502, 1990.

[39] Dong Jian GX, Li Gesheng, Zhang Xintang. Study on diesel–LPGdual fuel engines. SAE Paper 2001-01-3679; 2001.

[40] Ishida M, Amimoto N, Tagai T, Sakaguchi D. Effect of EGR andpreheating on NG combustion assisted with gas–oil in a diesel engine.In: Proceedings of the COMODIA, Nagoya, Japan; 2001.

[41] John K, Diana DB. Analysis of knock in a dual-fuel engine. SAE922367; 1992.

[42] Satbir S, Kong S.-C., Rolf DR, Sundar RK, Midkiff KC. Modelingand experiments of dual-fuel engine combustion and emissions. SAE2004-01-0092; 2004.

[43] Abd-Alla GH, Soliman HA, Badr OA, Abd-Rabbo MF. Effects ofdiluent admissions and intake air temperature in exhaust gasrecirculation on the emissions of an indirect injection dual fuelengine. Energy Convers Mange 2001;42:1033–45.

[44] Kusaka J, Tsuzuki KI, Daisho Y, Saito T.A numerical study oncombustion and exhaust gas emissions characteristics of a dual fuelnatural gas engine using a multi-dimensional model combined withdetailed kinetics. SAE 2002-01-1750; 2002.

[45] Selim MYE. Pressure-time characteristics in diesel engine fuelled withnatural gas. Renewable Energy 2001;22:473–89.

[46] Ishida M, Cho JJ, Yasunaga T. Combustion and exhaust emissions ofa DI diesel engine operated with dual fuel. In: Proceedings of theFISITA World Automotive Congress, Seoul, Korea; 2000.

[47] Ishida M, Tagai T, Ueki H, Sakaguchi D. Ignition and combustioncharacteristics of methanol mixture in a dual fuel diesel engine. In:Proceedings of the COMODIA, Japan; 2004.

[48] Selim MYE. Effect of exhaust gas recirculation on some combustioncharacteristics of dual fuel engine. Energy Convers Mange2003;44:709–23.

[49] Patrick FD. Measurement and data analysis for engineering andscience. New York: McGraw-Hill; 2005.

[50] Liu Z, Karim GA. Knock in dual fuel engines. In: Proceedings of theCOMODIA, Japan; 1994.

[51] Nwafor OMI. Knock characteristics of dual-fuel combustion in dieselengines using natural gas as primary fuel. Sadhana – AcademyProceedings in Engineering Sciences 2002;27:375–82.

Documents

15. Study on Biogas Premixed Charge Diesel Dual Fuelled Engine - 10 Orang