Biomass Derived Producer Gas as a Reciprocating

8/8/2019 Biomass Derived Producer Gas as a Reciprocating

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Biomass and Bioenergy 21 (2001) 6172

Biomass derived producer gas as a reciprocatingengine fuelan experimental analysis

G. Sridhar , P.J. Paul, H.S. MukundaCombustion Gasification and Propulsion Laboratory,

Department of Aerospace Engineering,Indian Institute of Science, Bangalore 560 012, India

Received 28 June 2000; accepted 13 December 2000

ABSTRACT

This paper uncovers some of the misconceptions associated with the usage of producer gas, alower calorific gas as a reciprocating engine fuel. This paper particularly addresses the use ofproducer gas in reciprocating engines at high compression ratio (17 : 1), which hitherto hadbeen restricted to lower compression ratio (up to 12 : 1). This restriction in compression ratiohas been mainly attributed to the auto-ignition tendency of the fuel, which appears to besimply a matter of presumption rather than fact. The current work clearly indicates thebreakdown of this compression ratio barrier and it is shown that the engine runs smoothly atcompression ratio of 17 : 1 without any tendency of auto-ignition. Experiments have beenconducted on multi-cylinder spark ignition engine modified from a production diesel engine atvarying compression ratios from 11:5 : 1 to 17 : 1 by retaining the combustion chamber design.As expected, working at a higher compression ratio turned out to be more efficient and alsoyielded higher brake power. A maximum brake power of 17:5 kWe was obtained at an overallefficiency of 21% at the highest compression ratio. The maximum de-rating of power in gasmode was 16% as compared to the normal diesel mode of operation at comparable compressionratio, whereas, the overall efficiency declined by 32.5%. A careful analysis of energy balancerevealed excess energy loss to the coolant due to the existing combustion chamber design.Addressing the combustion chamber design for producer gas fuel should form a part of futurework in improving the overall efficiency. c_ 2001 Elsevier Science Ltd. All rights reserved.

Keywords: Biomass; Compression ratio; De-rating; Producer gas; Spark ignition engine

1. Introduction

With the renewed interest in biomassenergy by necessity, biomass-basedtechnologies are achieving prominence notonly as rural energy devices but also asindustrial power plants. Gasification is onesuch processCorresponding author. Tel.: +91-80-360-0536; fax:

+91-80-3601692.E-mail address: [email protected]

(G. Sridhar).

where clean gas could be generatedusing a wide variety of bio-residues as thefeed stock and in turn use the fuel gas forpower generation purposes. These arebeing used in standard diesel engines indual-fuel mode of operation so as to obtaindiesel savings up to 85%. Operation ofengines on gas alone has been explored insome limited sense by a number ofresearchers ever since World War II. In the


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present times, adopting these technologieshas immense economic benefits, a routepursued by a number of researchers is in

[15].

Development of gas engines using producergas has been explored ever since World WarII. It is estimated that over seven millionvehicles in Europe, Australia, SouthAmerica and Pacific Islands were convertedto run on producer gas during World War II.These engines were spark ignited engines,mostly in the lower compression ratiobracket operating either on charcoal orbiomass derived gas. Extensive fieldworkhas been carried at National SwedishTesting Institute of Agricultural Machinery[1] by mounting gas generator and engineset on trucks and tractors. There have alsobeen sporadic installations at Paraguay andSri Lanka [1] for power generationapplication.

The question of power generation usingproducer gas has been addressed in recenttimes by a few researchers [1,2,6] andattempts have been made to convertstandard compression ignition engine to agas engine with the relaxation imposed onthe compression ratio (CR) and others [3]operating a supercharged SI engine to

realise the rated output. Also, oneresearcher [4] has reported working onproducer gas fuelled engine at high CR(16:5 : 1) for water pumping applicationwithout any sign of knock. There appearsno earlier work on a systematic study onthe engine behaviour using producer gasfuel. The other important reason thatappears responsible is the non-availability

of standard and proven gasificationsystems, which could generate gas ofconsistent quality on a continuous basis forengine applications.

Systematic studies are essential from theviewpoint of establishing the highest usefulcompression ratio (HUCR) for producer gasfuel and also indirectly establish the octanerating for the fuel. This paper reports workon a producer gas fuelled spark ignitionengine converted from a production dieselengine.

A well researched, tested and anindustrially proven gasifier system capableof generating consistent quality wasemployed as the gas generator for testingpurpose [7,8]. The engine has been testedand verified at the highest compressionratio of 17 : 1 in order to establish knock-less performance by capturing thepressure-crank angle trace. Subsequentlythe engine has been tested at varying CRsso as to arrive at an optimum CR formaximum brake power and effiency.The

overall energy balance has been analysedand the shortcomings identified. Also theemission levels in terms of CO and NO havebeen examined.


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2. Misconceptions and clarification

Prior to this development there have beentwo misconceptions regarding producer gasfuel and they are identified as follows: (1)auto-ignition tendency at higher CR whenused in reciprocating engine, (2) large de-rating in power due to calorific value of thefuel being low.

It was thought that these perceptions hadno reasonable basis. Indeed, the basis forthe contrary seemed to exist. Firstly,producer gas being a mixture of many gasspecies with large fraction being inertshould have higher octane rating whencompared to natural gas and biogas. Thegas contains a large fraction of inert gaseslike CO2 and N2 accounting to 1215% and4850%, respectively, and these could actas knock suppressors [9]. However, so farthere has not been any research of octanerating test conducted on producer gas fuel.Moreover, it is not clear if any establishedtest procedure exists for producer gas likethe Methane number test for natural gasand biogas. One crude way of assessment isto test the fuel gas in standard engines andplace them accordingly in the octane ratingtable.

Secondly, there is a general thinking thatproducer gas being a lower calorific fuel,

the extent of de-rating would be largewhen compared to high calorific value fuelslike natural gas (NG) and liquefiedpetroleum gas (LPG). The de-rating if anycould be due to two possible reasons.Firstly, with the lower energy density fuelsthere is a net decrease in number ofmolecules when compared to high-energyfuels like diesel, gasoline, NG or LPG. Thiscontributes to some de-rating in case oflow energy density fuels [10]. De-rating ofpower on account of calorific value will besmall because of marginal differences in

the energy release per unit mixture(air+fuel gas) [11]. This can be explained asfollows. The calorific value of producer gasvaries between 4.7 and 5:0 MJ N m3 asagainst 30 MJ N m3 for NG. The energydensity per unit (producer gas +air) mixtureis only 1520% lower than NG and airmixture even though the calorific value ofproducer gas is one-eighth of NG. This is

because the stoichiometric air=fuel ratiofor producer gas is 1.2 as compared to 17for NG. Hence the extent of de-rating withproducer gas would not be marginalcompared to NG fuelled operation atcomparable operating conditions. This gapcould be nullified by working producer gasat higher CR when compared to NG. Theupper limit of CR for NG has beenidentified to be around 15:8 : 1 based on arecent work [10].

3. Earlier work

Shashikantha et al. [2] has reported relatedwork on a converted diesel engine at CR of11:5 : 1. In addition to the change in CR,the combustion chamber of the originalengine i.e. bowl-in piston (hemispherical)was modified to Hesselman (shallow W)with an aim of achieving a higher level ofturbulence by squish rather than swirl.With the above modification a poweroutput of 16 kWe has been reported in gasmode with efficiency in the range of 2124%against a rated output of 17 kWe in dieselmode. However, the same authors [6]subsequently claim a lower output andefficiency of 11:2 kWe and 15%. Theseauthors do discuss the knock tendencies athigher CR but no experimental evidenceseems to be provided in support.Measurements have been reported of

various parameters including that ofexhaust emissions, however nomeasurements have been made withrespect to gas composition, which isconsidered essential from the viewpoint ofestablishment of input energy.

The first experimental work in the higherCR range has been reported byRamachandra [4] on a single cylinder dieselengine (16:5 : 1 CR) coupled to a waterpump. A power de-rating of 20% has beenreported at an overall efficiency of 19%

without any signs of detonation. This workdoes not report any other measurementlike the pressure-crank angle diagram inorder to rationalise some of the results.Work on gasoline engine operation onproducer gas has been reported by Parke[12] with de-rating claims of 34%,compared to gasoline operation. The sameauthors [3] suggest supercharging to


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enhance the engine output. Martin andWauters [5] have reported work usingcharcoal gas and producer gas on an SIengine with a de-rating of 50% and 40%,respectively, at a CR of 7 : 1. However, thesame authors claim 20% de-rating whenworked with producer gas at a CR of 11 : 1.The authors present a CR barrier of 14 : 1and 11 : 1 for charcoal and producer gas,respectively, with inadequate experimentaljustification. From the literature survey itappears that no experimental evidence isavailable to support the phenomenon ofknock in producer gas engines, eventhough it is believed knock would occur athigher CR.

4. Current investigations

It is a well-acknowledged fact that it isdesirable to operate an internal combustionengine at the highest possible CR so as toattain higher overall efficiencies. But thegain in efficiency beyond a certain CR canbe expected to be marginal due to otherinfluencing factors such as heat loss andfriction. In the case of an SI engine thelimitation of CR comes from the knocksensitivity of the fuel. It has beenexperimentally investigated that the upperlimit for compression ratio for SI engineoperation is 17 : 1 beyond which there is afall in efficiency [13]. The above conclusion

is based on extensive tests with iso-octaneas the fuel also doped with anti-knockagent. If one were to consider this as theupper limit and since no other work hasbeen conducted at higher CR for SI engines,choosing a production engine in the aboverange for the current investigation seemedvery appropriate.The current investigationwas conducted on a commercially availablediesel engine so as to explore thepossibility of working at the existing CR of17 : 1 and optimising the same if required.At the onset of investigation, it was

perceived that increase in CR could haveconflicting effects on the power output ofthe engine.This could be explained asfollows. It has been universally recognisedthat turbulent flame speed [9] plays a vitalrole in the heat release rate during thecombustion process in an engine cylinder.The turbulent flame speed can be treatedas an enhanced form of laminar flame

under the influence of time varyingturbulence [9] within the combustion

chamber of the engine. The laminar flamespeed is again a function of initial pressure,temperature and the mixture composition.An earlier computational work by Mishra[14] indicated that the laminar flame speedfor stoichiometric producer gas and airmixture could decrease by one-tenth as theinitial pressure is enhanced by a factor of40. However, these calculations were madeat an initial temperature of 300K, and theinitial temperature at which combustionstarts is high in the case of internalCombustion engines. The influence ofinitial pressure and temperature on laminarflame speed can be explained in simpleterms as follows. The increase in theunburned gas temperature results inincrease in adiabatic flame temperatureand hence the average reaction rates. Theincrease in the reaction rate is a result ofthe increase in the number of radicalsreleasedthus contributing to increase inthe flame speed, whereas the rise inpressure can result in reduction in theamount of radicals released thus retardingthe flame speed. Therefore, the conflictingnature of the effects of initial pressure andtemperature needs to be recognised. Theeffect of these at varying CR is anadditional feature that needs to berecognised in order to arrive at the

optimum CR. Consequently, the presentinvestigation was started with anassumption that the optimum compressionratio would be between 12 : 1 and 17 : 1for maximum power output and overalleffciency.

In the current investigation, CR was theparameter that was varied. The influenceof CR on power, effciency and emissionshas been studied in some detail. Minimumignition advance for best torque (MBT) hasbeen determined at different Rs. The

variation of cylinder pressure with time hasbeen captured using a piezo-basedtransducer. The overall energy balance hasbeen projected.


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5. Conversion methodologies

A three cylinder, direct injection dieselengine of 3:3 l capacity, with a CR of 17 : 1was converted into a spark ignition engine

to drive a 25-kVA alternator. The salientfeatures of the engine are given in Table 1.Modifications attempted on the engine forconversion are as follows:

1. Insertion of spark plug in place of fuelinjectors without changing its location(centrally located).

2. Adaptation of a distributor type batterybased ignition system with a provision toadvance=retard ignition timing. The setignition timing was checked using astroboscope.

3. The combustion chamber designcomprising a flat cylinder head and bowl-inpiston was retained. No attempts weremade to change the combustion chamberdesign except that the thickness of thecylinder head gasket was varied toaccomplish different CRs of 17 : 1, 14:5 : 1,13:5 : 1 and 11:5 : 1.

4. For in-cylinder pressure measurement,provision was made on one cylinder headby drilling a 1:5 mm diameter hole for

pressure measurement and fitting anoptical sensor on the crankshaft for crankangle measurement.

6. Experimental set-up and measurementscheme

The well-researched, tested and industrialversion of IIScs-open top down draft, twinair entry 75 kg h1 solid bio-residue gasifiersystem [7] formed the gas generator. Thisstate-of-the art technology has undergoneextensive testing both in India [8] andoverseas [15] and proven to be a world-class system. The system has qualified for

long hours of continuous operation inmeeting the industrial requirements interms of generation of consistent qualitygas. The overall details of the gasifiersystem are presented in Fig. 1. As shown inthe figure, the system had the provision totest the quality of the gas prior to supply to

engine. At the engine intake, a carburetor isprovided


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the engine. At the engine intake, acarburetor is provided for proportioning airand fuel flow. As there were no carburetorscommercially available to cater to producergas, a locally made carburetion system andmanually controlled valve were used forproportioning.

Measurements were made with respect tothe following parameters:

(A) Producer gas compositions using on-linegas analysers. The gases analysed were CO,CO2, CH4, O2 and H2. The N2 concentrationwas deduced by difference. The CO, CO2,CH4 components were determined usinginfrared gas analysers and the H2component using a thermal conductivity-based analyser. The O2 measurementsystem was based on chemical cell.

(B) In-cylinder pressure variation datasynchronized with the crank anglemeasurement was acquired

on a computer for every one-degree crankangle.The pressure measurement wasaccomplished usinga pre-calibrated Piezo based pressuretransducer(M=s PCB make).(C) Measurement of voltage and currentacrossthree phases and frequency for poweroutputcalculationsthe load bank constituted ofresistors.(D) Air and gas Kow to the engine using pre-calibratedventurimeters.(E) Engine exhaust analysisO2, CO2; CO,and NOand temperature.

with the crank angle measurement wasacquiredon a computer for every one-degree crank

on a computer for every one-degree crankangle. The pressure measurement was

accomplished using a pre-calibrated Piezobased pressure transducer (M=s PCB make).

(C) Measurement of voltage and currentacross three phases and frequency forpower output calculationsthe load bankconstituted of resistors.

(D) Air and gas FLow to the engine usingpre-calibrated venturimeters.

(E) Engine exhaust analysisO2, CO2; CO,and NOand temperature.


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7. Experimental procedure

Experiments were initiated on the engineonly after the gasifier system stabilised i.e.attained steady state operation in terms ofgeneration of consistent quality gas. Thetypical time scale for attaining steady stateof operation from the cold start was 23 h.During this period the gas was flared in aburner. The gas composition wasdetermined using on-line gas analysers,pre-calibrated using a known producer gasmixture. The calibrations of these analyserswere checked at random time intervals soas to minimise errors in long durationoperation. Typically gas composition at thetime of start of the engine test was 19 1%H2; 191% CO; 2% CH4; 121% CO2; 20:5%H2O and rest, N2. The mean calorific valueof gas varied around 4:650:15 MJ N m3.The feedstock used for gasification isCausurina species wood with moisturecontent between 12% and 15% on dry basis(sun-dried wood).

Once the gas composition stabilised, theengine was operated for a few minutes at1500 RPM at no-load condition. All the testson the engine were conducted around aconstant speed of 150050 RPM. Thethrottling for speed control and air and fuelproportioning was achieved using manuallyoperated valves.

Experiments were conducted at CRs of 17 :1, 14:5 : 1, 13:5 : 1 and 11:5 : 1 and theseCRs were achieved by varying the thicknessof the cylinder head gasket. Thecompression ratio values are based on thecylinders geometric measurements andwere verified by matching the motoringcurve with an engine simulation curve. Theengine was tested at different ignitiontiming settings to determine the MBT atdifferent CRs. With a set ignition timing,the air and fuel were tuned to achieve

maximum power. Measurements wereinitiated 1015 min after attaining stableoperation. The in-cylinder pressure datawith a resolution of 1 crank angle wasacquired on a computer in excess of 100150 consecutive cycles. Prior to the start ofthese tests, the TDC was accuratelydetermined using a dial gauge and

synchronized with the optical crank angle

measuring system with an accuracy of 0:5.

8. Results and observations

8.1. Performance

The first and the foremost result of thesetests is that the engine worked smoothlywithout any sign of knock at a high CR of17: 1. There was no sign of audible knockduring the entire load range. Moreover, theabsence of knock was clear from thepressure-crank angle recordings both at fullload and part load.

The engine delivered a maximum output of17:5 kWe (20 kW shaft power) at a CR of 17: 1 with an overall efficiency of 21%

compared to 21 kWe (24 kW shaft power)output at 31% efficiency in diesel mode.The overall efficiency calculated is basedon the ratio of the shaft output deliveredto the energy content in the biomass. Theoverall efficiency in gas mode is based onthe ratio of mechanical shaft output to theenergy content in the biomass. The usefuloutput and efficiency decreased with thelowering of CR. A maximum output of 15:3kWe (17:6 kW shaft power) at an overallefficiency of 18% was obtained at a CR of11:5 : 1. The variation of brake power with

CR is shown in Table 2. The power outputat an intermediate CR of 14:5 : 1 and 13:5 :1 were 16.4 and 16:2 kWe, respectively,and with overall efficiencies being 20%. Theoverall efficiency at 13.5 CR is the same asthat at 14:5 : 1 on account of leaneroperation.


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The extent of de-rating in brake power isabout 16.7% at a CR of 17 : 1 andincreased as high as 26% at 11:5 : 1 whencompared to diesel mode of operation. Thegain in overall efficiencies from CR 11.5 to17 : 1 works out to be 16.6%, which meansan increase of 3% per unit CR increment.However, the incremental gain per unit CRfrom 14.5 to 17 : 1 is 2%. These figures arewell within the range of 1 to 3% gains perunit incremental of CR [9]. The fuelairequivalence ratio (PHI) at which themaximum power was derived was around1.001.06 with the exception of 0.86 at CRof 13:5 : 1. The air to fuel ratio is tunedfrom the viewpoint of deriving maximumoutput and therefore the efficiency figuresare necessarily not the maximum that canbe obtained. It may be possible to achievehigher overall efficiencies by operating atleaner conditions.

The peak shaft output at varying CR wasfound to be sensitive to the producer gascomposition. The hydrogen fraction in thefuel gas dictated the ignition timingsetting. Therefore the minimum advancefor brake torque (MBT) varied with thechange in hydrogen content in the fuel gas.A higher fraction of hydrogen means thatthe ignition timing has to be retarded inorder to benefit from the increase in theflame speed. With a faster burn rate the

optimum spark timing has to be closer tothe TDC, the mixture temperature andpressure at the time of initiation of sparkwill be higher and hence the laminar flamespeed at the start of combustion will alsobe higher. Therefore optimizing theignition timing based on hydrogen fractionis vital from the viewpoint of derivingmaximum shaft output.

Since the mixture flame speed is a strongfunction of hydrogen content in the gas,the MBT will differ based on the actual fuel

gas composition. For a gas compositioncontaining 20:5 0:5% H2 and 19:5 0:5%CO, the MBTs have been identified asshown in Table 2, the measurementaccuracy of MBT is within 3 crank angle.The MBT in the present case has turned outto be about 610BTDC at a CR of 17 : 1 andhas gone up to as high as 1416BTDC for aCR of 11:5 : 1. These values are much

lower compared to 30 to 40BTDC at a CRof 11.5 based on earlier work [2,6,12] butmatches with 10BTDC [4] at a CR of 17 : 1.The mechanical efficiency of the engine ata CR of 17 : 1 is about 80% and increases toas high as 87% at a CR of 11:5 : 1. Theincrease in mechanical efficiency isattributed to the reduction in rubbingfriction [16] due to lower cylinder pressuresencountered at lower CRs. The mechanicalefficiency values are based on indicatedpower measurement (based on integrationof pressurevolume diagram) and thesewere found to be identical with the Morsetest results.

8.2. Pressurecrank angle data

The pressurecrank angle recording at allthe CRs did not show any trace of knock forall ranges of load including that of peakload and this is visible from the pressurecrank angle diagrams as shown in Fig. 2.Faster burn rate due to presence ofhydrogen in the fuel gas may be one factorfor the non-knocking performance at highercompression ratio. The faster burn rateaccompanied by retarded ignition timingsetting obviates any auto-ignition tendencyof the end gas. Increasing the flame speedor retarding the ignition timing setting isone possible way of reducing knock

tendency and this is well acknowledged inthe literature [9]. The maximum indicatedmean effective pressure (IMEP) wasobtained at ignition timing correspondingto the maximum shaft output. The IMEPobtained at varying CR as a function ofignition timing is shown in Fig. 3. The maxIMEP recorded was 595 kPa at a CR of 17 : 1and declined to about 485 kPa at a CR of11:5 : 1. The maximum IMEP was obtainedat a PHI of 1.05 and this falls well withinthe anticipated PHI of 1.0 to 1.1 [9]. Thepoint of occurrence of peak pressure at all

CRs occurred at 1819ATDC, except that at13:5 : 1 which occurred at 17ATDC, whichis close to the generally acknowledgedvalue of 17ATDC for MBT [9,17]. Therefore,for CR of 17 : 1, 14.5 and 11:5 : 1 the MBTshould be more advanced


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(lie within 23) from what was actuallymeasured. The variation of IMEP or the netuseful output within this close range wouldbe marginal. The peak cylinder pressuresand their point of occurrence is shown inTable 3. The peak pressure at 14:5 : 1 CR islower than 13:5 : 1 probably due to a slightdeparture from MBT. The coefficient ofvariation of the IMEP at all CRs and ignitionsettings occurred well within 33.5%,implying low cycle-to-cycle variation. Thereason for low cyclic variation is the fasterrate of combustion occurring inside theengine cylinder. The faster rate ofcombustion is attributed to higher flamespeeds due to the presence of hydrogen inthe gas and also to the bowl-in pistoncombustion chamber design with increasedsquish effect.

8.3. Energy balance

Fig. 4 represents the overall energybalance at a CR of 17 : 1. The energybalance at MBT showed that about 32.5% ofthe energy was realised as useful output(indicated power), about 27% was lostthrough the exhaust (including the CO inthe exhaust) and the remaining 40% to thecooling water (inclusive of radiative losses,etc). As expected with the advancement ofignition timing, the loss through exhaustmode reduced and increased to the coolantroute. The energy balance in gas modeshowed that a large fraction of

the energy was lost to the cooling waterwhen compared to the diesel mode. Fig. 5compares the energy balance in gas anddiesel mode (at rated output of 21 kWe) at

a CR of 17 : 1, the energy loss at maximumpower delivered to the coolant andmiscellaneous is about 40% compared to33% in diesel and whereas, the energy lossthrough exhaust in gas mode decreased by5%. The indicated power and thereby thethermal efficiency being higher in dieselmode is evident from the pressuretimecurve shown in Fig. 2. The energy balance

at varying CRs is shown in Fig. 6. There wasan increase in the loss of energy throughexhaust (which includes the energy in theform of CO) with the reduction in the CR,

whereas, the loss through the coolant andmiscellaneous was higher at higher CR. Theuseful energy is 32.5% at a CR of 17 : 1 andhas declined to 25.7% at a CR of 11:5 : 1.The useful energy at a CR of 13:5 : 1 hasstood about the same as that at 14:5 : 1due to a relatively leaner operation. Theincreased amount of heat loss to thecooling water as a whole in gas operation


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could be attributed to the enginecombustion chamber design. It has beenquoted in the literature [9] that withengine geometries such as bowl-in-pistonthere will be 10% higher heat transfer. Theheat transfer to the coolant in the currentcase falls well within this range. With theincrease in compression ratio the overall

conversion efficiencies must improvethermodynamically, similarly, the heat lossto the coolant and exhaust should havereduced. However, increased energy loss tothe coolant at a higher compressionratio is probably due to increase in heattransfer coefficient.


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accounted for because it forms a small part( 5%) of NOx generated [9]. The NO levelhas been represented in milligram per unitMJ of input energy. These results havebeen compared to the Swiss norms becauseof some earlier collaborative work with theSwiss scientists; it was indicated that Swissnorms were stringent with respect to theemission levels. The NO level reduced withthe retardation of ignition timing and thisfeature was observed for all CRs. The

NO level was observed to be maximum atthe highest compression ratio withadvanced ignition timings, whereas for theMBT range of 620 BTDC the NO wasroughly about the same in almost all thecases. However, there was one exceptionof NO being higher at MBT for a CR of 13.5due to a leaner operation.

It is a well-known fact that NO generationis strongly dependent on the temperature

and also residence time in the combustionchamber. With the flame speed of the gasmixture being high, the ignition setting isretarded whereby the residence time in thehigh temperature combustion chamber isautomatically reduced. Therefore, the lowNO levels at retarded ignition setting is anexpected and consistent behaviour. Theabove results match well with those quotedin the literature [9], which show small tomodest variation of NO with CR. Thevariation of carbon monoxide (CO) with PHIis shown in Fig. 8. The CO levels have been

represented in grams per MJ of inputenergy. The trend of CO with PHI is clearfrom the figure. The CO levels were lowerat the highest CR, and this could beattributed to higher temperatures, leadingto relatively complete combustion.

. Conclusions

erformance of the engine at higher CR has

9

Pbeen smooth and it has been establishedbeyond doubt that the operating enginesusing producer gas in SI mode at highercompression ratios is technically feasible.The cylinder pressurecrank angle trace hasshown smooth pressure variations duringthe entire combustion process without anysign of abnormal pressure raise. A shorterduration of combustion has been observedwith producer gas fuel, requiringretardation of the ignition timing toachieve MBT. These faster burning cycles

are corroborated by low cyclic pressurefluctuations with a coefficient of variation3%. The faster burning process has beenidentified to be due to the higher flamespeed of the fuel gas mixture and the same


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has been attributed to the hydrogen

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content in the gas. So, increased quantityof hydrogen in the fuel gas mixture isdesired from the viewpoint of approachingan ideal cycle operation. The perceivednegative influence of pressure on flamespeed at higher CR does not seem to existbased on the above results. The maximumde-rating in power is observed to be 16% ingas mode when compared to dieseloperation at comparable CR. The extent ofde-rating was much lower when comparedto any of the previous studies [36]. Thisnumber matches with a similar kind of de-rating reported with NG operation [10].However, the overall efficiency drops downby almost 32.5% compared to normal dieselmode of operation. A careful analysis of theenergy balance revealed excessive heat lossto the coolant at all CRs and this resultedin engine overheating within 3040 min ofoperation at full load. This phenomenon ofexcessive heat loss could be attributed tobowl-in piston combustion chamber design.Higher heat loss to the cylinder walls canbe expected because the combustionchamber is originally meant for dieseloperation with inherent swirl. Hencesuitable modification of the combustionchamber is essential from the view point ofreducing the energy loss to the coolant.Reduction in 10% heat loss to the coolantwould amount to improvement in overall

efficiencies by 3%. The study of the enginecombustion chamber and modifications canbe looked upon as future work.

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