dee+biogas hcci

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Energy 35 (2010) 3614e3622

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Energy

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Diethyl ether as an ignition improver for biogas homogeneous chargecompression ignition (HCCI) operation - An experimental investigation

K. Sudheesh, J.M. Mallikarjuna*

Internal Combustion Engines Laboratory, Department of Mechanical Engineering, Indian Institute of Technology Madras, Chennai 600 036, India

a r t i c l e i n f o

Article history:Received 24 October 2009Received in revised form28 April 2010Accepted 29 April 2010Available online 14 June 2010

Keywords:BiogasDiethyl etherHCCIIgnition improver

* Corresponding author. Tel.: þ91 44 22574698; faxE-mail address: [email protected] (J.M. Mallikar

0360-5442/$ e see front matter � 2010 Elsevier Ltd.doi:10.1016/j.energy.2010.04.052

a b s t r a c t

This paper deals with experimental investigations of a homogeneous charge compression ignition (HCCI)engine using biogas as a primary fuel and diethyl ether (DEE) as an ignition improver. The biogas isinducted and DEE is injected into a single-cylinder engine. For each load condition, best brake thermalefficiency DEE flow rate is determined. The results obtained in this study are also compared with those ofthe available biogas-diesel dual-fuel and biogas spark ignition (SI) modes. From the results, it is foundthat biogas-DEE HCCI mode shows wider operating load range and higher brake thermal efficiency (BTE)at all loads as compared to those of biogas-diesel dual-fuel and biogas SI modes. In HCCI mode, at4.52 bar BMEP, as compared to dual-fuel and SI modes, BTE shows an improvement of about 3.48 and9.21% respectively. Also, nitric oxide (NO) and smoke emissions are extremely low, and carbon monoxide(CO) emission is below 0.4% by volume at best brake thermal efficiency points. Also, in general, in HCCImode, hydrocarbon (HC) emissions are lower than that of biogas SI mode. Therefore, it is beneficial to usebiogas-DEE HCCI mode while using biogas in internal combustion engines.

� 2010 Elsevier Ltd. All rights reserved.

1. Introduction

Generally, spark ignition (SI) and compression ignition (CI)engines are most commonly used for power generation and intransportation vehicles. The SI engines have low smoke emissionsand brake thermal efficiency with high HC, CO and NOx emissions.However, the CI engines show comparatively higher brake thermalefficiency, but they emit high amounts of smoke and NOx. In theseengines, achieving low exhaust emissions with low fuelconsumption is a challenge. Also, today’s emission legislations areforcing engine manufacturers to search for engines having lowexhaust emissions and higher fuel economy. Today, HCCI isemerging as an effective alternative combustion process for CImode. With certain fuels, it can provide higher brake thermalefficiency like CI engines with ultra-lowNOx and particulate matter(PM) emissions. In these engines, the premixed nature of a chargeeffectively eliminates the PM, whereas combustion with leanmixture without definite flame propagation reduces cylinder gastemperatures leading to ultra low NOx emissions simultaneously.

Biogas is an attractive source of energy for rural areas especiallyin countries like India. It is generated by anaerobic digestion of cowdung, other animal wastes and plant matters such as leaves and

: þ91 44 22574652.juna).

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water hyacinth. All the above mentioned sources are renewable innature and abundantly available. Biogas contains approximatelytwo-thirds (by volume) of methane (CH4) and the rest is mostlycarbon dioxide (CO2) with traces of hydrogen sulphide (H2S). Biogashas a low energy density due to the presence of CO2. The propertiesof biogas are given in Table 1. Usage of biogas in SI engines enhancesknock resistance due to the presence of CO2. However, it reducesbrake thermal efficiency and increases HC emissions [1,2]. It is notpossible to use biogas directly in CI mode due to its higher auto-ignition temperature. Dual-fuel mode of operation is a feasible wayof using biogas in CI engines with diesel as a secondary fuel [3,4].But, dual-fuel mode emits more HC and CO emissions with lowerbrake thermal efficiency. However, smoke and NOx emissions arelower as compared to the conventional CI mode due to the gaseousnature of biogas and the presence of CO2.

Diethyl ether is considered as a renewable fuel because it can beproduced from ethanol through the dehydration process [5]. Lowautoignition and boiling temperature of DEE are reasons forselecting it as an ignition improver along with lethargic fuels likebiogas. Also, DEE has a higher energy density than ethanol. Theimportant properties of DEE are given in Table 2. Generally, DEE isused as a cold starting aid in CI engines. DEE is also blended withdiesel for improving the brake thermal efficiency and reducingemissions in CI engines [6].

Onishi et al. (1979) introduced controlled autoignitioncombustion in a two-stroke engine in order to reduce instability at

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Table 2Properties of DEE.

Calorific value 33.9 MJ/kgDensity 713 kg/m3

Boiling point 34.4 �CStoichiometric air fuel ratio (mass basis) 11.1Autoignition temperature 160 �CCetane number >125

Nomenclature

lBG biogas excess air ratiolDEE DEE excess air ratiolT total excess air ratiosMair mass flow rate of the airMBTE best brake thermal efficiencyNOP nozzle opening pressureMBG mass flow rate of the biogasMDEE mass flow rate of the DEEAFBG biogas stoichiometric airefuel ratioAFDEE DEE stoichiometric airefuel ratioCAD crank angle degree

K. Sudheesh, J.M. Mallikarjuna / Energy 35 (2010) 3614e3622 3615

part-loads and they also achieved good reduction of emissions andfuel consumption [7]. Najt et al. (1983) extended the HCCIcombustion into a four-stroke engine using primary reference fuels[8]. A four-stroke gasoline HCCI engine was tested by Thring et al.(1989) and the important parameters required for successfulgasoline HCCI operation at part-loads were investigated [9]. Ryanet al. (1996) used a port fuel injection (PFI) injector to supply dieselinto the intake air stream at various inlet air temperatures andcompression ratios [10]. This resulted in early heat release duringcompression stroke itself, and they concluded that low compres-sion ratio is most suitable for port injected diesel fuelled HCCIengine. Garcıa et al. (2009) investigated the effect of inlet chargetemperature, cool EGR, injection timings and equivalence ratio onthe performance of a diesel fuelled HCCI engine and achievedcomparatively higher loads by using cool EGR [11]. Shi et al. (2006)studied the effect of internal and externally cooled EGR on theperformance of a diesel fuelled HCCI engine. They concluded thatinternal EGR benefited the formation of homogeneous mixture andreduced smoke emission, whereas externally cooled EGR couldhelp extend upper load limit of HCCI operation [12]. Due to theearly heat release characteristics with low volatility of diesel,researchers tried to use various alternative fuels with high auto-ignition temperature, viz. natural gas, methanol, ethanol, liquefiedpetroleum gas (LPG) without many modifications to the originalengine. Inlet charge heating, variable compression ratio (VCR) anduse of secondary low octane fuels as an ignition improver weremethods tried for achieving HCCI combustion with the above fuels.Christensen et al. (1997) investigated the HCCI combustion char-acteristics using natural gas with inlet air heating [13]. SwamiNathan et al. (2008) adopted acetylene as a fuel for HCCI enginebecause of its moderate autoignition temperature and high flam-mability limits. They used inlet charge temperature to controlcombustion phasing [14,15]. Chen et al. (2000) introduced dual-fuelHCCI combustion of natural gas with dimethyl either (DME). They

Table 1Properties and composition of biogas.

Calorific value 17 MJ/kgDensity (1 atm and 15 �C) 1.2 kg/m3

Flame speed 0.25 m/sStoichiometric air fuel ratio (mass basis) 5.7Autoignition temperature 650 �CFlammability limits with air (%) 7.5e14Research octane number 130

Typical biogas composition in % volumeMethane 57.37Carbon dioxide 42.1Carbon monoxide 0.08

found that by optimizing a proportion of DME and natural gas, NOxemissions could be lowered to near zero levels. The dual-fueloperation gave higher brake thermal efficiency than that of CI mode[16]. Zheng et al. (2004) used DME as an ignition controller ina methanol fuelled HCCI engine [17]. Mack et al. (2009) experi-mentally proved the suitability of wet ethanol in HCCI combustion.They used inlet charge heating to control the combustion phasing[18]. Swami Nathan et al. (2008) investigated biogas HCCIcombustion with manifold injection of diesel. They compared theresults of biogas-diesel HCCI mode with that of biogas-diesel duel-fuel mode of operation. Due to the low volatility and high boilingtemperature of diesel, inlet heating was used for proper mixing ofdiesel with manifold inducted biogas. They concluded that biogas-diesel HCCI operation is superior to dual-fuel mode of operation ina BMEP range of 2.5 to 4 bar [19]. However, they couldn’t operate anengine in HCCI mode below 2.5 bar and above 4 bar BMEPs due tosystem limitations and difficulty in controlling inlet charge heating.The motivation for the present work is to operate a single-cylinderengine in biogas fuelled HCCI mode in a wide load range by usingDEE as an ignition improver. These types of studies are limited inliterature. Therefore, a detailed study would help not only toevaluate and compare the HCCI mode with conventional biogas SIand biogas diesel-dual-fuel modes, but also to find the feasibility ofusing biogas in HCCI mode.

In this study, the effect of DEE flow rate on the performance,emissions and operating load range of a biogas fuelled HCCI modeare studied at a constant engine speed of 1500 rev/min., and coolingwater outlet temperature of 50 �C. Finally, DEE mass flow rate forbest brake thermal efficiency point at each load condition has beenfound out. In addition, the performance and emission characteris-tics of biogas-DEE HCCI mode are compared with the availableresults of biogas SI [1,2] and biogas-diesel dual-fuel [3] modes.

2. Experimental setup

A single-cylinder, water-cooled, direct injection CI engine is usedfor conducting experiments. The engine is coupled to an eddycurrent dynamometer for loading and measurement purposes. Theengine specifications are shown in Table 3. The DEE is stored in anaccumulator and is injected into the intake manifold using aninjector at a linepressureof 2 bar. An in-housebuilt electronic circuitis used for controlling DEE flow rate. The biogas is directly inductedthrough the intake manifold. The biogas is generated in a nearbyplant, which uses cowdung andwater to produce it. It is collected ina flexible bag at the plant and transported to the place of usage. It isactually sent through a floating drum in order to maintain the

Table 3Engine specifications.

Bore� stroke 80� 110 mmConnecting rod length 231 mmCompression ratio 16:1Rated power output 3.7 kW @ 1500 rpmDisplacement volume 553 cm3

Injector NOP 220 bar

1. Air flow measurement

2. DEE supply unit

3. DEE pressure gauge

4. DEE pulse width controller

5. DEE fuel injector

6. Biogas flow meter

7. Biogas flow control valve

8. Biogas floating drum

9. Pressure transducer

10. Crank angle encoder

11. Data acquisition system

12. Personal computer

13. Cooling water inlet

14. Cooling water flow regulator

15. Cooling water out let temperature sensor

16. Cooling water outlet

17. Engine

18. Exhaust gas temperature sensor

19. Eddy current dynamometer

20. Electromagnetic clutch

21. Electric motor

22 to 25. Exhaust gas analyzers

26. Engine exhaust

Fig. 1. Schematic of experimental setup.

K. Sudheesh, J.M. Mallikarjuna / Energy 35 (2010) 3614e36223616

required constant pressure as in the plant. The biogas composition ismeasured every time using a MRU make non-dispersive infrared(NDRI) analyzermeant for this purpose. A typical composition of thebiogas is shown in Table 1. The flow rate of biogas is controlled bya fine-control valve and is measured with the help of a turbine type

Fig. 2. Photograph of experimental setup.

gas flow meter before being admitted to intake manifold. Thecooling water outlet temperature is measured by a resistancetemperature detector (RTD) and the air flow rate is measured byusing a dry turbine type flowmeter. Exhaust gas analyzers workingon the principles of FID for HC, NDIR for CO and CLD for NO,respectively, areused tomeasure the emission levels. A Bosch smokemeter is used to measure smoke emissions. Figs. 1 and 2 showschematic and photographic views respectively of the experimentalsetup developed and used in this study.

3. Experimental procedure

First, the engine is motored at about a speed of 600 rev/min.,and simultaneously DEE is injected into the intake manifold. Thus,the engine is started in HCCI mode with DEE as the only fuel and

Table 4Mixture strengths used in biogas-DEE HCCI mode.

lDEE¼ 6.1 lDEE¼ 7.6 lDEE¼ 9.8

BMEP lBG lT BMEP lBG lT BMEP lBG lT

2.27 4.74 2.66 2.80 3.75 2.52 4.54 2.17 1.771.73 6.64 3.21 2.27 4.53 2.86 4.33 2.27 1.841.30 9.30 3.75 1.73 5.26 3.15 3.89 2.45 1.960.86 15.27 4.49 1.29 6.58 3.60 3.35 2.51 2.010.54 20.0 4.84 0.86 8.23 4.05 2.80 2.84 2.21

0

10

20

30

40

50

60

70

80

340 350 360 370 380

Pre

ssur

e (b

ar)

Crank angle (degree)

Cooling water temperature = 50 CelsiusEngine speed = 1500 rev/min

DEE = 6.1

BMEP = 2.27 BarBMEP = 1.73 BarBMEP = 1.29 BarBMEP = 0.86 BarBMEP = 0.54 Bar

Fig. 3. Variation of cylinder gas pressure with load for lDEE of 6.1.

0

10

20

30

40

50

60

70

80

350 360 370 380 390

Pre

ssur

e (b

ar)



DEE = 9.8




the engine speed is allowed to reach about 1000 rev/min., byadjusting DEE flow rate. Then, biogas is inducted through intakemanifold and its flow rate is adjusted to reach rated engine speedof 1500 rev/min. At this condition, the engine is allowed to rununtil the coolant water temperature reached 50 �C. Afterwards,DEE flow rate is gradually reduced, because after the warm-upperiod, the engine knocks due to a higher DEE flow rate. At a givenDEE flow rate, the possible operating load range is found out byvarying biogas flow rate. In fact, the operating load range isdecided by the misfiring and knocking limits of the engine. Theabove procedure is repeated with various DEE flow rates. Everytime, the possible operating load range is determined by varyingbiogas flow rates based on the misfiring and knocking limits. Here,the knocking limit is considered as the rate of pressure rise ofmore than 10 bar/CAD. Finally, the combination of DEE and biogasflow rates for best brake thermal efficiency at each load conditionis determined. Uncertainty and error analysis of the measured andcalculated parameters is done, and the values are shown in theAppendix [20].

4. Results and discussion

Both biogas and DEE flow rates affect combustion characteristicsof biogas-DEE HCCI mode. In order to present their effect on engineperformance and emissions, in the following discussion, biogasexcess air ratio (lBG), DEE excess air ratio (lDEE) and total excess airratios (lT) are defined as follows and used.

0

10

20

30

40

50

60

70

80

330 340 350 360 370 380 390

Pre

ssur

e (b

ar)



DEE = 7.6



lBG ¼ MairMBG*AFBG

; lDEE ¼ MairMDEE*AFDEE

lT ¼ MairðMBG*AFBG þMDEE*AFDEEÞ

In the following discussion, biogas-DEE HCCI, biogas SI and biogas-diesel dual-fuel modes are referred to as HCCI, SI and dual-fuelmodes, respectively. Here, the present results of biogas-DEE HCCImode are compared with those of biogas SI mode [2] and biogas-diesel dual-fuel mode [3] at 25, 50, 75 and 100% of the maximumpossible load (4.52 bar BMEP) of biogas-DEE HCCI mode. Where theexact values of parameters are not available, interpolated values areused for comparison. However, such comparison may not be veryaccurate because the composition of biogas used in each case isdifferent. The composition of the biogas in biogas-diesel dual-fuelmode is 19% CO2 and 73% CH4 by volume [3]. In biogas SI mode [2]and in the present study, the composition of biogas is 41% CO2 and58% CH4 by volume.

4.1. Cylinder gas pressure

Table 4 shows various excess air ratios used to operate theengine in biogas-DEE HCCI mode under different load conditions.The DEE excess air ratios shown in Table 4 correspond to the DEEflow rates, which have themaximum number of best brake thermalefficiency points in the possible load range.

0

10

20

30

40

50

60

70

80

330 340 350 360 370 380 390

Pre

ssur

e (b

ar)


DEE = 5.2DEE = 6 1DEE =7.6DEE =8.5

Cooling water temperature = 50 CelsiusLoad = 1.8 bar BMEPEngine speed = 1500 rev/min

Fig. 6. Variation of cylinder gas pressure with DEE flow rate.

0

200

400

600

800

1000

1200

1400

1600

320 340 360 380 400

Tem

pera

ture

(ke

lvin

)



DEE = 6.1


Fig. 7. Variation of cylinder gas temperature with load for lDEE of 6.1.

0

200

400

600

800

1000

1200

1400

1600

320 330 340 350 360 370 380 390 400

Tem

pera

ture

(ke

lvin

)



DEE = 9.8


Fig. 9. Variation of cylinder gas temperature with loads for lDEE of 9.8.


Figs. 3e5 show variations of cylinder pressures with loads atdifferent DEE excess air ratios. From Figs. 3e5, it can be seen thatthe occurrence of peak pressure advances with respect to the topdead center (TDC) with an increase in load. Also, the occurrence ofpeak pressure retards with an increase in DEE excess air ratio. Atlower DEE excess air ratios, energy liberated fromDEE is higher andthus autoignition occurs at an early stage. In general, peak pressurevaries from about 51 to 72 bar for the entire load range considered.

Fig. 6 shows the variation of cylinder pressure for various DEEexcess air ratios at 1.8 bar BMEP. From Fig. 6, it is seen that at lowerDEE excess air ratio, the occurrence of peak pressure advances andpeak cylinder pressure increases due to the excess amount of DEEenergy supply. This leads to an increased rate of pressure rise andengine noise, whereas, at high DEE excess air ratio, the cylinderpressure reduces and the occurrence of peak pressure retards dueto a lowDEE flow rate. It is observed that during experiments if lDEEincreases above 8.5, the engine misfires. In this condition, theenergy liberated from DEE may not be sufficient to autoignite thebiogas.

4.2. Cylinder gas temperature

Figs. 7e9 show the variations of cylinder gas temperatures withloads at various DEE excess air ratios. Here, the cylinder gastemperature is calculated from the measured cylinder pressure andgeometric volume at a given instant. The cylinder gas temperaturefollows a similar trend as that of the cylinder gas pressure. Fora given load condition, the cylinder gas temperature is higher for

0

200

400

600

800

1000

1200

1400

320 340 360 380 400

Tem

pera

ture

(ke

lvin

)



DEE = 7.6


Fig. 8. Variation of cylinder gas temperature with loads for lDEE of 7.6.

lower DEE excess air ratio due to advanced combustion phasing.From Figs. 7e9, it can be seen that after 320 CAD, there is a slightincrease in the cylinder gas temperature. It may be due to the lowtemperature reactions or cool flames (see Figs. 10e12). However,the cylinder gas temperature continues to increase even after thecool flames due to the compression of cylinder gases by upwardmovement of the piston. Occurrence of the peak cylinder gastemperature also follows a similar trend as that of the cylinder gaspressure. Peak cylinder gas temperatures are obviously higher athigher loads. In general, with a reduction in DEE excess air ratio,peak cylinder gas temperature increase due to advanced combus-tion phasing.

4.3. Heat release rate

Figs. 10e12 show heat release rate patterns with respect to loadsand various DEE excess air ratios. From Figs. 10e12, it is observedthat with an increase in the DEE excess air ratio, the start of lowtemperature reactions (cool flames) is delayed and peaks of themare reduced. This is because of increased biogas flow rate, whichreduces the reaction rate. It may be due to the higher dilution effectof CO2 present in biogas, which can even delay the start of the lowtemperature reactions. At high load conditions, the peak of coolflame reduces due to a reduction in the DEE flow rate. Generally,a low DEE flow rate is required for high load conditions because ofhigher overall wall temperatures. From Figs. 10e12, it is also seenthat the peak of the main combustion phase increases and thecombustion duration reduces with an increase in load. It may be

-10

0

10

20

30

40

50

60

70

320 330 340 350 360 370 380

Hea

t re

leas

e ra

te (

J/C

A)



DEE = 6.1


Fig. 10. Patterns of heat release rate at various load for lDEE of 6.1.

-10

0

10

20

30

40

50

60

70

320 330 340 350 360 370 380

Hea

t re

leas

e ra

te (

J/C

A)



DEE = 7.6


Fig. 11. Patterns of heat release rate at various loads for lDEE of 7.6.

-10

0

10

20

30

40

50

60

70

320 340 360 380 400

Hea

t re

leas

e ra

te (

J/C

A)


DEE = 5.2DEE = 6.1DEE = 7.6DEE = 8.5

Cooling water temperature = 50 CelsiusLoad = 1.8 bar BMEPEngine speed = 1500 rev/min

Fig. 13. Patterns of heat release rate at various DEE flow rate.


due to lower total excess air ratio and higher wall temperatures athigher load conditions. The occurrence of peak heat release rateand the start of main combustion advance as the DEE excess airratio reduces. This is because of higher energy release by DEE.Overall, for the entire load range, the maximum heat release rate isabout 80 J/CAD.

Fig. 13 shows the effect of DEE flow rate on heat release rate.From Fig. 13, it is seen that at a low DEE excess air ratio, combustionphasing advances due to the excess amount of energy suppliedfrom DEE. This reduces the indicated power and leads to a lowerbrake thermal efficiency, whereas a very low DEE flow rate retardscombustion phasing and increases combustion durationwith lowercylinder pressures, and it again reduces the brake thermal effi-ciency. At 1.8 bar BMEP, lDEE of 6.1 gives the best brake thermalefficiency operating condition.

4.4. Maximum rate of pressure rise

Fig. 14 shows a variation of the maximum rate of pressure risewith DEE excess air ratio and load conditions. At all the loadconditions, the maximum rate of pressure rise is below 9 bar/CAD.It may be due to the presence of CO2 in the biogas, which has a gooddilution effect. Also, the higher heat capacity of CO2, which acts asa diluent, reduces the rate of reactions leading to lower combustionrates. At all the best brake thermal efficiency points, the maximumrate of pressure rise is below 7.5 bar/CAD. It helps smooth engineoperations without knocking, and also extends the operating loadrange.

-10

0

10

20

30

40

50

60

70

80

90

320 340 360 380 400

Hea

t re

leas

e ra

te (

J/C

A)



DEE = 9.8 BMEP = 4.54 BarBMEP = 4.33 BarBMEP = 3.89 BarBMEP = 3.35 BarBMEP = 2.80 Bar

Fig. 12. Patterns of heat release rate at various loads for lDEE of 9.8.

4.5. Operating load range

Fig. 15 shows the possible operating load range for various DEEexcess air ratios. From Fig. 15, it is observed that as the DEE excessair ratio increases, the load limits are extended. It means that theoperating load range shifts towards the right-hand side. At low loadconditions, the engine temperature is low and therefore higher DEEflow rate is required to autoignite the biogas. But, as the loadincreases, the engine temperature increases, and, along withenergy supplied by the DEE, leads to engine knock. This restrictsupper load limit for a given DEE flow rate. In Fig. 15, the dotted lineshows the best brake thermal efficiency DEE excess air ratio for theentire load range. The biogas-DEE HCCI mode has awide load rangeof zero to 4.52 bar BMEP.

4.6. Energy share of biogas and DEE

Fig. 16 shows the percentage of energy supplied by DEE andbiogas at various load conditions. From Fig. 16, it can be seen that atlow loads, the energy supplied by the biogas is low. At low loads,biogas cannot autoignite easily due to a low overall enginetemperature; therefore, a larger amount of DEE is required forautoignition. At 0.54 bar BMEP, the energy supplied by biogas isabout 23.5% of total energy (at the best BTE point). As loadincreases, the energy contribution of biogas increases. At 4.52 barBMEP, the energy contribution from biogas is around 82%. In Fig. 16,the dotted line shows energy share of the biogas at the best brakethermal efficiency points.

0

1

2

3

4

5

6

7

8

9

0.0 1.0 2.0 3.0 4.0 5.0

Rat

e of

pre

ssur

e ri

se (

bar/

o CA

)

Brake mean effective pressure (bar)

DEE = 5.2DEE = 6.1DEE = 7.6DEE = 8.5DEE = 9.8


Fig. 14. Variation of maximum rate of pressure rise with DEE excess air ratio and load.

5

6

7

8

9

10

11

0.0 0.5 1.0 1.5 2.0 2.5 3.0 3.5 4.0 4.5 5.0

DE

E a

ir e

xces

s ra

tio



Fig. 15. Possible operating load range with DEE excess air ratio.

0

5

10

15

20

25

30

0.0 0.5 1.0 1.5 2.0 2.5 3.0 3.5 4.0 4.5 5.0

Bra

ke t

herm

al e

ffic

ienc

y (

%)



MBTE points

Cooling water temperature = 50 Celsius Engine speed = 1500 rev/min

Fig. 17. Variation of brake thermal efficiency with load and DEE flow rate.


4.7. Brake thermal efficiency

Fig. 17 shows a variation of brake thermal efficiency (BTE) withload and DEE excess air ratios in HCCI mode. In Fig. 17, the dottedline represents the best BTE points for different load conditions inHCCI mode. From Fig. 17, it is observed that in HCCI mode, the bestBTE point for each load condition depends upon mass flow rate ofDEE. It may be because the optimum mass flow rates of DEE andbiogas provide a better combustion phasing (Fig. 13).

Fig. 18 shows the comparison of BTE in HCCI, SI [2] and dual-fuel[3] modes of operation. Only the best BTE points in HCCI mode areconsidered for comparison. In general, from Fig. 18, it is observedthat the HCCI mode shows better BTE at all loads as compared tothe other two modes. It may be because of full-throttle operationand higher compression ratio used in HCCI mode as compared to SImode. Also, it may be due to the reduced combustion duration inHCCI mode as compared to that of dual-fuel mode, which is similarto conventional CI engine combustion. From Fig. 18, it is observedthat at 25, 50, 75 and 100% loads, with HCCI mode, an improvementin BTE of about 29.5, 15.7, 19.3 and 9.2% respectively is observed ascompared to that in SI mode. Further, at 50, 75 and 100% loads, animprovement of about 27.3, 8.81 and 3.48% respectively is observedas compared to the dual-fuel mode.

4.8. NO and smoke emissions

Fig. 19 shows a variation of NO emissions with DEE excess airratios and loads in HCCI mode. Fig. 20 shows a comparison of NOemissions in HCCI (at the best BTE points) and SI modes at four

0

10

20

30

40

50

60

70

80

90

0.0 0.5 1.0 1.5 2.0 2.5 3.0 3.5 4.0 4.5 5.0

Bio

gas

ene

rgy

(%)


DEE = 5.2DEE = 6.1DEE =7.6DEE =8.5DEE =9.8

MBTE points


Fig. 16. Energy share of biogas and DEE at various loads.

percentages of loads considered. From Fig. 20, it is observed that atall load conditions considered, NO emissions are very much lowerin HCCImode as compared to those of SI mode. Thismay be becauseof instantaneous and low temperature combustion in HCCI mode.Due to instantaneous combustion, the compression effect onburned gas by burning mixture may be eliminated. Thereby, localhigh temperature regions in burned gas products are eliminated,which, in turn, reduce the formation of thermal NO [21,22]. As thereis no flame propagation in case of HCCI combustion, prompt NOformation may be negligible. From Fig. 20, it is found that at 25, 50,75 and 100% loads, there is a reduction in NO emissions by about93.61, 99.6, 97.2 and 99.7% respectively in HCCI mode as comparedto SI mode.

During the experiments, under all the operating conditions, itwas observed that smoke emissions are very much lower (less than0.1 BSU) in HCCI mode. It may be mainly due to the absence of fuel-rich zones during HCCI combustion because of pre-mixed leangaseous phase HCCI mode, unlike the conventional CI mode ofoperation.

4.9. CO emissions

Carbon monoxide (CO) emission is generally an indication ofincomplete oxidation of fuel. It is found that gaseous fuelled HCCIengine shows lower CO emissions compared to liquid fuels due tohigher miscibility of gaseous fuel with air [11]. Fig. 21 showsa variation of CO emission with DEE excess air ratio and loads forHCCI mode. At low load conditions, CO emission is higher may bedue to low combustion temperatures. At higher loads, higher

0

5

10

15

20

25

30

0.0 0.5 1.0 1.5 2.0 2.5 3.0 3.5 4.0 4.5 5.0 5.5

Bra

ke t

herm

al e

ffic

ienc

y (

%)


HCCI

Porpatham et.al (SI,25% throttle)

Porpatham et.al (SI,100% throttle)

Duc et.al (Dual fuel)


Fig. 18. Comparison of brake thermal efficiency in HCCI, SI and dual-fuel modes.

0

5

10

15

20

25

30

35

0.0 0.5 1.0 1.5 2.0 2.5 3.0 3.5 4.0 4.5 5.0

Nit

ric

oxid

es (

pp

m)



MBTE points


Fig. 19. Variation of NO emissions with DEE excess air ratio and load.

0

200

400

600

800

1000

1200

1400

25 50 75 100

Nit

ric

oxid

es (

ppm

)

Percentage of maximum possible HCCI load

HCCI modeSI mode


Fig. 20. Comparison of NO emission in HCCI and SI modes.

0

0.5

1

1.5

2

2.5

0.0 0.5 1.0 1.5 2.0 2.5 3.0 3.5 4.0 4.5 5.0 5.5

Car

bon

mon

xide

s (%

vol

)


HCCI

Porpatham et. al ( SI, 25% throttle)Porpatham et. al ( SI,100% throttle)


Fig. 22. Comparison of CO emission in HCCI and SI modes.


combustion temperatures may reduce the CO emissions compara-tively. Fig. 22 shows a comparison of CO emission in HCCI and SImodes. From Fig. 22, it is seen that except at 2.7 bar BMEP, COemission in HCCI mode is higher than that of SI mode. It may be dueto low combustion temperature in HCCI mode, which may lead toflame quenching. At 25, 50, 75 and 100% loads, CO emissions are4.2, 1.85, 1.73 and 0.64 times higher, respectively, than those of SImode. At 2.7 bar BMEP, CO emission in SI mode is about 20 times

0.00

0.10

0.20

0.30

0.40

0.50

0.60

0.0 0.5 1.0 1.5 2.0 2.5 3.0 3.5 4.0 4.5 5.0

Car

bon

moo

xid

es (

% v

ol)


DEE = 5.2DEE = 6.1DEE = 7.6DEE = 8.5DEE =9.8

MBTE points


Fig. 21. Variation of CO with DEE excess air ratio and load.

higher than that of HCCI mode. It may be due to the very richmixture used at that load condition at 25% throttle in SI mode.

4.10. Hydrocarbon emissions

Fig. 23 shows a variation of hydrocarbon (HC) emissions withDEE excess air ratio and load. From Fig. 23, it can be seen that at lowload conditions, HC emissions are comparatively lower because ofthe higher DEE flow rate. Due to this, the start of combustionadvances (Fig. 10) and leads to comparatively higher cylinder gastemperatures (about 1100 to 1300 K with respect to loads as shownin Fig. 7), which improve DEE oxidization. Also, during the expan-sion stroke, the left out DEE can easily oxidize due to the hightemperature of hot cylinder gas because of its low autoignitiontemperature. At intermediate loads, the HC emission increases,which may be due to an increase in the biogas flow rate with load.At higher biogas flow rates, more biogas could occupy crevices andwould come out into the cylinder during expansion stroke, as thecylinder gas temperature is lower than the autoignition tempera-ture of methane. This may inhibit oxidization of HC duringexpansion stroke, leading to higher HC emissions, whereas, athigher loads, a higher cylinder gas temperature helps oxidization ofbiogas and thereby the HC emissions are comparatively lower athigher loads.

Fig. 24 shows a variation of HC emissions in HCCI and SI modesat four percentages of loads considered. In HCCI mode, at 25, 50 and75% loads, the HC emissions are lower by about 83.9, 29.6 and

0

500

1000

1500

2000

2500

3000

3500

4000

4500

0.0 0.5 1.0 1.5 2.0 2.5 3.0 3.5 4.0 4.5 5.0

Hyd

ro c

arbo

ns (

ppm

)


DEE = 5.2DEE = 6.1DEE = 7.6DEE = 8.5DEE =9.8

MBTE points


Fig. 23. Variation of HC emissions with DEE excess air ratio and load.

0

1000

2000

3000

4000

5000

6000

7000

8000

9000

25 50 75 100

Hyd

ro c

arbo

ns (

ppm

)

Percentage of maximum possible HCCI load

HCCI

SI


Fig. 24. Comparison of HC emissions in HCCI and SI modes.

Uncertainty and error analysis.

Parameter Uncertainty Error of instrument

Speed �0.06% 1 rpmAir flow rate �0.90% Volume¼ 0.001 m3

Time¼ 0.1 sFuel flow rate �0.40% Volume¼ 0.0001 m3

Time¼ 0.1 sIntake charge temperature �0.06% Temperature¼ 0.1 �CNO emission �3.10% 1 ppmUBHC (FID) �2.10% 1 ppmBMEP �0.10% Speed¼ 1 rpm

Torque¼ 0.0485 nmBrake power �0.40% Torque¼ 0.0485 nm

Speed¼ 1 rpmBrake thermal efficiency �0.60% Torque¼ 0.0485 nm

Speed¼ 1 rpmVolume (fuel)¼ 0.0001 m3


45.57% respectively, as compared to SI mode. This may be due toleanmixtures in HCCI mode, unlike near stoichiometric mixtures inthe SI mode. However, at 100% load, the HC emissions at bothmodes (HCCI and SI) are almost at equal levels.

5. Conclusions

From the experimental investigations in biogas-DEE HCCI mode,the following conclusions are drawn:

Biogas can be used more effectively in a HCCI mode whencompared to other modes of operation. The engine can be easilystarted in HCCI mode by using DEE as an ignition improver unlikeother fuels and methods used for HCCI operation. In this work,biogas-DEE HCCI mode of operation is possible in a load range ofzero to 4.52 bar BMEP. The biogas-DEE HCCI mode shows higherBTE at all loads as compared to biogas-diesel dual-fuel and biogas SImodes. In addition, it is found that HC, NO and smoke emissions arelower in biogas-DEE HCCI mode as compared to biogas SI mode.However, CO emissions are higher at all loads. At 25, 50, 75 and100% loads in biogas-DEE HCCI mode, BTE shows an improvementof 29.5, 15.7, 19.3 and 9.2% respectively as compared to SI mode. Inaddition, compared to dual-fuel mode, at 50, 75 and 100% loads,there is an improvement in BTE by about 27.3, 8.81 and 3.48%respectively. In biogas-DEE HCCI mode, at 25, 50 and 75% loads, HCemissions are reduced by about 83.9, 29.6, and 45.57% respectivelyas compared to the SI mode. At 100% load, HC emissions arecomparable in both modes. In biogas-DEE HCCI mode, at 25, 50, 75and 100% loads, CO emissions are 4.2, 1.85, 1.73 and 0.64 timesrespectively more than those of SI mode. The percentage reductionof NO emission in biogas-DEE HCCI mode at 25, 50, 75 and 100%loads are 93.61, 99.6, 97.2 and 99.7% respectively as compared tothose in SI mode. At all load conditions, it is observed that, smokeemissions are very much lower (less than 0.1 BSU) in biogas-DEEHCCI mode. Therefore, it is feasible to operate biogas fuelled enginein HCCI mode for better performance and emission characteristicswith DEE as an ignition improver compared to other modes ofoperation.

Acknowledgments

The authors wish to thank Suresh Kumar N., Cyril Mathew,Nagarajan K., Michael John Bose M., Subramanian M.K., and Babu S.for their help in carrying out the experiments. Their contribution isgreatly appreciated.

Appendix

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