Fuel 89 (2010) 1721–1730

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An experimental investigation on the use of EGR in a supercharged natural gasSI engine

Amr Ibrahim *, Saiful BariSustainable Energy Centre, School of Advanced Manufacturing and Mechanical Engineering, University of South Australia, SA 5095, Australia

a r t i c l e i n f o a b s t r a c t

Article history:Received 4 May 2009Received in revised form 3 July 2009Accepted 7 July 2009Available online 22 July 2009

Keywords:EGRSI engineNatural gasInlet pressureNO

0016-2361/$ - see front matter � 2009 Elsevier Ltd. Adoi:10.1016/j.fuel.2009.07.005

* Corresponding author. Address: Sustainable EnergManufacturing and Mechanical Engineering, UniversitLakes, SA 5095, Australia. Tel.: +618 830 25123; fax:

E-mail address: amralihi@yahoo.com (A. Ibrahim).

The use of lean burn technology in spark-ignition engines has been dominant; however, lean burn tech-nique can not economically satisfy the increasingly restricted future emission standards. Consequently,alternative combustion techniques need to be investigated and developed. In this paper, the use of thestoichiometric air–fuel mixture with exhaust gas recirculation (EGR) technique in a spark-ignition natu-ral gas engine was experimentally investigated. Engine performance and NO emissions were studied forboth atmospheric and supercharged inlet conditions. It was found that the use of EGR has a significanteffect on NO emissions. NO emissions decreased by about 50% when EGR dilution increased from zerowith an inlet pressure of 101 kPa to close to the misfire limit with an inlet pressure of 113 kPa. In addi-tion, the use of EGR effectively suppressed abnormal combustion which occurred at higher inlet pressure.The use of higher inlet pressure in the presence of EGR improved engine performance significantly.Engine brake power increased by about 20% and engine fuel consumption decreased by about 7% whileNO emissions decreased by about 12% when 5% of EGR dilution was employed with an inlet pressure of113 kPa compared to using undiluted stoichiometric inlet mixture with an inlet pressure of 101 kPa.

� 2009 Elsevier Ltd. All rights reserved.

1. Introduction impact on both environment and human health [4]. In addition,

The World Health Organization has stated recently that urbanair pollution is the most significant environment risk factor andthat it is among the most important causes of death in Europe.According to the World Health Organization in 2002, exposure toparticulate matter (PM) emissions was estimated to be responsiblefor about 100,000 premature deaths per year [1]. Automotive emis-sions greatly contribute to urban atmospheric pollution. Diesel en-gine emissions currently represent a main source of the urban PMemissions. Diesel exhaust is currently classified by the Interna-tional Agency for Research on Cancer (IARC) as probably carcino-genic to humans. According to the US Environmental ProtectionAgency, EPA, motor vehicle exhaust was responsible for about56% of carbon monoxide and 49% of nitrogen oxides emissionsemitted into the atmosphere in USA [2].

The conventional fuels which have dominated the internal com-bustion engine chemical energy supply are basically petrol anddiesel. These fuels were preferred to be used because of their avail-ability, competitive price, and high energy density [3]. However,the use of both petrol and diesel has resulted in a serious harmful

ll rights reserved.

y Centre, School of Advancedy of South Australia, Mawson+618 830 23380.

these fuels are derived from petroleum oil which is non renewableand it is expected to be totally consumed within the next few dec-ades [4,5]. As the petroleum oil reserve declines, its cost is ex-pected to increase dramatically in the future especially in thecountries that have poor oil reserves and depend on imported oilfor their energy supply.

Recently, these environmental and economical concerns havemotivated governments and research organizations to investigatedifferent types of engine fuels that can be widely used as an alter-native to the dominant conventional fuels of both petrol and diesel.Alternative fuels are expected to be friendlier to the environmentand more sustainable than the conventional fuels.

Natural gas might be considered as the cleanest fossil fuel. Asnatural gas consists basically of methane, the hydrogen to carbonmole ratio (H/C) in most natural gas compositions is close to 3.8,which is the highest hydrogen to carbon ratio compared to anyother hydrocarbon fuel. The carbon mass percentage in naturalgas is close to 75% compared to 86–88% for both petrol and diesel,which makes natural gas produces less carbon dioxide per unit ofenergy released [6]. Although natural gas is not renewable, naturalgas is more sustainable than petroleum oil as the world reserves ofnatural gas is bigger than the oil. In addition, some of natural gasreserves exist in parts of the world that have poor oil reservesincluding Australia. Using natural gas as an affordable energysource in these countries can reduce their dependence on importedoil.

Table 2

1722 A. Ibrahim, S. Bari / Fuel 89 (2010) 1721–1730

Until a clean renewable energy source becomes widely avail-able to the public with an affordable price, natural gas would re-main as one of the main clean alternative fuel options that canbe economically available in many parts of the world at least forsome time in the future [7].

One of the natural gas spark-ignition engine combustion tech-nologies, which might use up to almost 100% excess air [8], begunin the early 1980s and became known as ‘‘lean burn” combustion.Lean burn combustion resulted in relatively lower in-cylinder tem-perature, and consequently lower thermal stresses and lowerknocking tendency. This permitted the use of turbocharging, rela-tively higher compression ratio, and optimum spark advance tim-ing to achieve high engine efficiency with an acceptable durability.In addition, this technique satisfied the previous governmentalemission standards without the need of exhaust gas after-treat-ment which resulted in relatively lower engine costs. As a result,most of the research and investment were directed towards thedevelopment of lean burn natural gas engines.

Currently, increasingly stringent ambient air quality standardsare forcing the production of engines that emit lower emissions;see Table 1 [9]. In order for the engine under the lean burn modeto produce lower nitrogen oxides (NOx) emissions, it has to operatewith a leaner mixture. However, the use of excessive air dilutioncan deteriorate engine stability, increase hydrocarbon (HC) andCO emissions, and decrease engine efficiency. Another way to con-trol NOx emissions is to retard the spark timing, which also leads toa decrease in engine efficiency and an increase in HC emissions[10]. Therefore, it seems that any efforts towards further decreasein NOx emission would lead to an increase in HC emission and a de-crease in engine thermal efficiency [11]. It has become obvious thatit would be difficult for the conventional gas engine operating onlean burn mode to meet the stringent future emission standards(see Table 1) especially for NOx emissions without using exhaustgas after-treatment [12].

The current technologies used for NOx emission after-treatmentin lean burn engines such as the selective catalytic reduction (SCR)devices are expensive and add some complexity to the engine use.For example, the SCR technique consists of ammonia storage, feed,and injection system and a catalyst. In this system, the ammonia isinjected in the exhaust gases upstream of the catalyst. In order forthis system to operate properly, a certain exhaust gas temperaturerange must be maintained [13]. In addition, an oxidation catalystwould be necessary to reduce both the HC and CO emissions.

On the other hand, the three-way catalyst (TWC) can be used toreduce NOx, HC and CO emissions in spark-ignition engines. TheTWC is capable of reducing the three emissions at the same timeand it is much less expensive than the SCR devices used in leanburn engines. However, in order for the TWC to operate efficiently,the engine must operate with almost a stoichiometric air–fuel mix-ture (i.e. without excess air). When the engine operates near thestoichiometric mixture, the in-cylinder temperature increases,and consequently, the thermal stresses and the knocking tendencyincrease. In order to reduce the in-cylinder temperature, an inletcharge dilution must be employed. One of the methods used to di-lute the inlet charge is to recycle some of the exhaust gases back

Table 1The European emission standards, g/kW h [9].

Year Standard CO HC NOx PM

1996 Euro2 4 1.1 7 0.152000 Euro3 2.1 0.66 5 0.12005 Euro4 1.5 0.46 3.5 0.022008 Euro5 1.5 0.46 2 0.022013 Euro6a 1.5 0.13 0.4 0.01

a Proposal only.

into the cylinder intake with the inlet mixture. This method iscalled exhaust gas recirculation (EGR). Adding EGR to the inletmixture will reduce the oxygen partial pressure in the inlet mix-ture, and consequently the in-cylinder NOx production will de-crease [14,15]. Also, the use of EGR with a stoichiometric air–fuelmixture can economically reduce engine emissions by allowingthe use of a three-way catalyst.

Several investigations which compared the effects of the use oflean combustion and EGR strategies on natural gas engine fuel con-sumption indicated that the use of the stoichiometric air–fuel mix-ture with EGR strategy resulted in higher fuel consumptioncompared to the lean burn strategy [11,16,17]. For instance, in2007, Saanum and co-workers [11] concluded that a penalty in en-gine thermal efficiency must be accepted when EGR is used as analternative to lean burn. However, these results can indicate thatthe use of a stoichiometric air–fuel mixture with EGR in naturalgas spark-ignition engines has not been fully optimised yet. Fur-ther research is still needed in the current time and in the futureto optimise the natural gas engine operation for the EGR strategyin order to achieve extremely low emissions accompanied withhigh engine power and efficiency. The aim of the research de-scribed in this paper is to experimentally investigate the effect ofusing EGR at both atmospheric and supercharged inlet conditionson natural gas engine performance and NO emissions. Although ni-tric oxide (NO) and nitrogen dioxide (NO2) are usually grouped to-gether as NOx, nitric oxide is the predominant oxide of nitrogenproduced inside an engine cylinder [18]. Consequently, only NOemission was experimentally measured and analyzed in this paper.

2. Experimental setup

The experimental research was carried out using a single cylin-der spark-ignition Ricardo engine whose specifications are shownin Table 2. The Ricardo engine was coupled to an electrical dyna-mometer in order to provide the load for the engine. The engineused to operate using petrol fuel before it was converted to operateon natural gas.

Fig. 1 shows a schematic of the experimental setup. The recy-cled exhaust gas was taken from a hole located on the exhaust pipewith the help of a small suction pump. The hot exhaust gas wascooled by passing it through a water-cooled heat exchanger. Botha regulating valve and an orifice flow meter were installed downstream from the heat exchanger in order to regulate and measurethe exhaust gas flow respectively. The increase of the percentage ofEGR in the inlet mixture was done by increasing the amount of theexhaust gas flowing back to the engine intake.

A supercharger was installed in order to provide the enginewith inlet charge at high pressure. The supercharger was drivenby an electric motor via a belt. Air, natural gas (from a pipe linesupply), and cooled exhaust gas were mixed in the superchargerintake before they were delivered at high pressure to anintercooler.

Ricardo engine specifications.

Item Value

Serial No. 120/73No. of cylinders 1Bore (mm) 76.2Stroke (mm) 111.125Capacity (cc) 507Maximum speed (rpm) 3000Inlet valve opens (� BTDC) 9Inlet valve closes (� ABDC) 34Exhaust valve opens (� BBDC) 43Exhaust valve closes (� ATDC) 8

Fig. 1. A schematic of experimental set-up.

A. Ibrahim, S. Bari / Fuel 89 (2010) 1721–1730 1723

The intercooler cooled down the air–fuel-exhaust gas mixturebefore it was delivered to the engine intake. A by pass valve wasinstalled in order to control the inlet mixture pressure as shownin Fig. 1. A pressure gauge was used to measure the inlet pressurewith an accuracy of ±2 kPa while K-type thermocouples were usedto measure the EGR cooler inlet and outlet temperatures in addi-tion to the intercooler inlet and outlet temperatures with an accu-racy of ±1 �C. Engine speed was measured with an accuracy of±10 rpm using a mechanical tachometer. All the research describedin the paper was carried out at the wide open throttle conditionwhile the inlet mixture pressure increased from 101 to 113 kPaby a supercharger.

An Alcock Model 450 V viscous flow air meter was used tomeasure the air flow rate flowing to the engine with an accuracyof ± 0.09 m3/h while a Dwyer Model RMC gas flow meter wasused to measure the natural gas flow rate flowing into the enginewith an accuracy of ±0.045 m3/h. The fuel flow rate was con-trolled by a regulating valve so that the air–fuel ratio was fixedat the stoichiometric condition with an accuracy of ±4%. Also, a6 mm square orifice meter was used to measure the flow rateof the exhaust gases recycled back to engine intake. The pressuredifference across the orifice meter was measured with an accu-racy of ±0.01 kPa using a U tube manometer. The percentage ofexhaust gases recycled back to the engine intake (%EGR) was cal-culated as a percentage of the total inlet mass flow rate asfollows:

%EGR ¼_mEGR

_ma þ _mf þ _mEGR� 100 ð1Þ

where _mEGR; _ma; _mf are the flow rates of EGR, air, and fuel respec-tively in kg/s.

NO emissions were measured using Beckman Industrial Model951A NO emission analyzer. This emission analyzer uses thechemiluminescence technique, which depends on the emission oflight, for measuring NO emissions.

The setup for collecting the in-cylinder pressure crank-angledata is also shown in Fig. 1. The in-cylinder pressure crank-angledata was determined using water-cooled piezoelectric Kistler7061B pressure sensor, top dead centre (TDC) position optical sen-sor, and magnetic pickup shaft encoder. A Kistler 5007 chargeamplifier was used to convert the output electrical charge fromthe piezoelectric pressure sensor into DC voltage. The output sig-nals from the charge amplifier, TDC position sensor, and shaft en-coder were received by an analogue digital converter in order toconvert the continuous signals into digital which were suitableto be handled by a laptop via a data acquisition card as indicatedin Fig. 1. Lab View 7.1 computer software was used to monitorand save the outflow data from the charge amplifier, shaft encoder,and TDC position sensor.

3. Results and discussion

In this section, the effect of varying the percentage of EGR in theinlet mixture at both atmospheric and supercharged inlet condi-tions on engine performance and NO emissions is discussed. Table3 shows the engine operating conditions which were used duringthis study. The average inlet temperature of the air–gas–EGR

Table 3Engine operating conditions.

Item Value

Speed (rpm) 1500Spark timing (� BTDC) 47Compression ratio 8Cooled EGR temperature (�C) 65Atmospheric inlet pressure (kPa) 101Supercharged inlet pressure (kPa) 113Air–fuel ratio Stoichiometric

1724 A. Ibrahim, S. Bari / Fuel 89 (2010) 1721–1730

mixture was about 27 and 35 �C for the atmospheric and super-charged inlet conditions, respectively.

Fig. 3. Air flow rate variations with EGR dilution for atmospheric and superchargedinlet conditions.

Fig. 4. Fuel and EGR flow rate variations with EGR dilution for atmospheric andsupercharged inlet conditions.

3.1. Engine power

Fig. 2 shows the effect of varying the percentage of EGR in theinlet mixture on engine brake power at both atmospheric andsupercharged inlet pressures. When the inlet mixture was dilutedwith EGR at constant inlet pressure, the EGR replaced some of theinlet air, and consequently, the air flow rate decreased as shown inFig. 3. Subsequently, the fuel flow rate was decreased in order tomaintain a stoichiometric air fuel ratio as shown in Fig. 4. As a re-sult, engine power generally decreased with the increase of per-centage of EGR dilution for both atmospheric and superchargedinlet conditions. The increase of the inlet pressure from 101 to113 kPa increased the inlet air and fuel flow rates as shown in Figs.3 and 4, respectively, and consequently, engine power significantlyincreased. The use of undiluted air–fuel mixture (no EGR) at an in-let pressure of 113 kPa increased the in-cylinder temperature sub-stantially and led to abnormal combustion occurrence as describedin the next section. This led to some loss in engine power com-pared to the use of 5% of EGR at the same inlet pressure of 113 kPa.

Fig. 2 indicates that increasing the inlet pressure from 101 to113 kPa without EGR dilution increased engine power by about16% as it increased from about 3.4 to 3.95 kW. However, the useof an EGR dilution of 5% with an inlet pressure of 113 kPa increasedengine power by about 20% compared to the use of undiluted inletmixture with an inlet pressure of 101 kPa. On the other hand,employing the maximum EGR dilution limit of about 12% withan inlet pressure of 113 kPa increased engine power by only 5%compared to the use of an inlet pressure of 101 kPa without EGR.

Fig. 2. Brake power change with varying the percentage of EGR dilution foratmospheric and supercharged inlet conditions.

3.2. Misfire and abnormal combustion

The higher increase in EGR dilution resulted in engine misfirewhich was identified by the measured in-cylinder pressure dataas shown in Fig. 5. The misfire pressure cycles were similar tothe pressure cycles measured during engine motoring process atthe same conditions.

Misfire started to occur with an EGR dilution above 10% foratmospheric inlet conditions and 12% for supercharged inlet condi-tions. The increase of EGR dilution misfire limit from about 10% to12% with the increase of inlet pressure could be explained by theeffect of inlet pressure on residual exhaust gas fraction. The resid-ual exhaust gas fraction decreases with the increase of inlet pres-sure, which makes the engine more tolerant to EGR dilution.Also, the increase in the in-cylinder unburned mixture densitywith the increase of inlet pressure slightly increased the combus-tion rate as described in Section 3.6. The increase of the combus-tion rate with inlet pressure could also contribute in the increaseof the engine tolerance to EGR dilution from 10% to 12%. Misfireoccurrence resulted in high cycle to cycle variations and engineinstability.

Fig. 5. In-cylinder pressure data for supercharged inlet conditions and an EGRdilution of 13%.

Fig. 7. Pressure crank-angle data for an abnormal combustion cycle at super-charged inlet conditions without EGR.

Fig. 8. Heat release rate calculated for an abnormal combustion cycle at super-charged inlet conditions without EGR.

A. Ibrahim, S. Bari / Fuel 89 (2010) 1721–1730 1725

On the other hand, when the engine was operated with a stoi-chiometric air–fuel mixture without EGR dilution at a super-charged inlet pressure of 113 kPa, some abnormal combustionpressure cycles were recorded. Abnormal combustion occurrencewas identified by the excessively high maximum pressure of somecombustion cycles compared to the maximum pressure of the nor-mal combustion cycles for the same operating conditions as shownin Fig. 6.

In order to further investigate these abnormal combustion cy-cles, one abnormal combustion cycle was isolated and used forcombustion analysis. Fig. 7 shows the isolated abnormal combus-tion pressure cycle data, which shows the maximum pressureoccurring almost at TDC. This could suggest that either the com-bustion was too fast or the combustion started too early. Boththe heat release rate and the heat release were calculated usingthis pressure data and shown in Figs. 8 and 9, respectively. BothFigs. 8 and 9 illustrate that the combustion started at about 80�BTDC while the normal spark timing was set at 47� BTDC. Thiscan explain that the air–fuel mixture was pre-ignited before theoccurrence of a normal spark. When the engine was operated witha stoichiometric air–fuel mixture without EGR dilution at a super-charged inlet pressure of 113 kPa, the combustion pressure and

Fig. 6. In-cylinder pressure data for supercharged inlet conditions without EGR.Fig. 9. Heat release calculated for an abnormal combustion cycle at superchargedinlet conditions without EGR.

Fig. 10. Pressure crank-angle data at supercharged inlet conditions without EGRand with a retarded spark timing of 35� BTDC.

1726 A. Ibrahim, S. Bari / Fuel 89 (2010) 1721–1730

temperature excessively increased and caused some overheatedspots inside the cylinder which ignited the air–fuel stoichiometricmixture earlier during some of the combustion cycles. This phe-nomenon is called pre-ignition or surface ignition. Surface ignitionis the ignition of the air–fuel mixture by any means other than thespark discharge such as an overheated valve or spark plug. Follow-ing surface ignition, a turbulent flame develops at each surfaceignition location and starts to propagate across the combustionchamber in an analogues manner to what occurs with normalspark-ignition [18].

It is worth mentioning that surface ignition differs from engineknock. Engine knock occurs when essentially a portion of the air–fuel mixture ahead of the propagating flame (the end gas) isauto-ignited due to high in-cylinder temperature and pressure.When knock occurs, much of the chemical energy of the end gasis released rapidly causing high local pressures and the propaga-tion of pressure waves of substantial amplitude across the combus-tion chamber. Both surface ignition and knock can lead to engineinstability and may cause engine damage.

Engine knock can be identified using the recorded cylinder pres-sure data. Knock causes in-cylinder pressure to fluctuate with dif-ferent amplitudes and frequencies according to knock severityduring a combustion cycle. Light knock leads to pressure fluctua-tions with smaller amplitudes and frequencies compared to heavyknock. The surface ignition combustion pressure signal shown inFig. 7 has no fluctuations, which indicates that the surface ignitionwas not accompanied with knock occurrence.

Fig. 8 shows that there is a noticeable large portion of negativenet heat release rate near the end of combustion which indicatesthat large amount of cylinder heat transfer occurred due to theexcessive in-cylinder temperature caused by pre-ignition. In addi-tion, the too advanced start of combustion for the pre-ignitioncombustion cycle resulted in a substantial increase in compressionwork, and consequently, the indicated power calculated for thepre-ignition combustion cycle was essentially lower compared tothe indicated power calculated for normal combustion cycles. Thisincreased cycle to cycle variations and led to engine instability.

In order to show if the setting of the advanced spark timing of47� BTDC was the reason behind the overheated hot spots insidethe cylinder which caused surface ignition, the spark timing wasretarded from 47� to 42�, 37�, and 35� BTDC, respectively for theundiluted stoichiometric air–fuel mixture and the in-cylinder pres-sure was measured for each spark timing condition. It was foundthat pre-ignition occurred at all the retarded spark timing condi-tions. In addition, both pre-ignition and misfire took place with aspark timing of 35� BTDC as shown in Fig. 10.

On the other hand, when the stoichiometric air–fuel mixturewas diluted with cooled EGR at a high inlet pressure of 113 kPaand a spark timing of 47� BTDC, the pre-ignition combustion cyclesentirely disappeared with a percentage of EGR dilution in the inletmixture of about 5%. This indicates the positive effect of EGR dilu-tion on reducing in-cylinder combustion temperature and thermalstresses, which can eliminate abnormal combustion occurrencesuch as surface ignition. The presence of high specific heat gasessuch as water vapor in the EGR in addition to the increase of thetotal inlet mass with the increase of the percentage of EGR dilutionin the inlet mixture help reduce the combustion temperature. Also,the increase of percentage of EGR dilution in the inlet mixture de-creases oxygen concentration and slows down the combustion ratewhich leads to a decrease in both maximum cylinder pressure andtemperature.

3.3. Engine stability

One important measure of cyclic variability, which can be calcu-lated from the measured cylinder pressure data, is the coefficient

of variation in indicated mean effective pressure, COV (%), whichcan be calculated as follows [18]:

COV ¼ rimep

imep� 100 ð2Þ

where imep (kPa) is the average indicated mean effective pressurecalculated for a number of cycles, n, while rimep (kPa) is the standarddeviation in indicated mean effective pressure. Both parameters canbe calculated as follows [19]:

imep ¼Xi¼n

i¼1

imepðiÞ,

n ð3Þ

rimep ¼

ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi1

n� 1

Xi¼n

i¼1

ðimepðiÞ � imepÞ2vuut ð4Þ

The COV defines the cycle to cycle variations in indicated workper cycle and it can be considered as a good indicator for enginestability determination. Heywood [18] demonstrated that vehicledrivability problems usually result when the coefficient of varia-tion in indicated mean effective pressure, COV, exceeds about10%. However, other studies showed that engine stability startsto deteriorate when COV increases above 5% [20].

Fig. 11 shows the effect of varying the percentage of EGR dilu-tion in the inlet mixture on the coefficient of variation in indicatedmean effective pressure, COV, for both atmospheric and super-charged inlet conditions. The dilution of the inlet mixture withEGR up to 10% at atmospheric inlet conditions resulted in a modestincrease in COV. However, when the percentage of EGR dilution in-creased above 10%, COV started to increase rapidly due to misfireoccurrence, which was identified for some of the combustion pres-sure cycles.

On the other hand, when the inlet mixture was diluted withEGR at higher inlet pressure, the COV was improved compared tothe undiluted inlet mixture condition as EGR dilution reducedthermal stresses and eliminated the pre-ignition occurrence. TheCOV decreased from 7% with no EGR dilution to about 1% with10% EGR dilution. However, when the percentage of EGR increasedabove about 12%, the COV started to increase rapidly due to misfireoccurrence. Fig. 11 shows that the increase in inlet pressure in-creased engine tolerance to EGR dilution from about 10% to 12%.

The maximum tolerable EGR dilution limit, which is about10% for the atmospheric inlet condition and about 12% for the

Fig. 11. The variation of coefficient of variation in indicated mean effective pressurewith the change of percentage of EGR dilution at both atmospheric and super-charged inlet conditions.

Fig. 13. The effect of varying the percentage of EGR dilution in the inlet mixture onin-cylinder pressure at inlet pressure of 113 kPa and spark timing of 47� BTDC.

A. Ibrahim, S. Bari / Fuel 89 (2010) 1721–1730 1727

supercharged inlet condition and for the operating conditionsshown in Table 3, might be considered low compared to themaximum tolerable limit for some modern engines which mighttolerate EGR dilution limits of more than 15%. This is because theRicardo engine had an essentially flat type combustion chamberwith the spark plug located on its side. This type of combustionchamber does not support the creation of a significant flow motionsuch as swirl during combustion and consequently the combustionrate is not rapid enough to tolerate substantially high percentagesof EGR dilution conditions.

3.4. In-cylinder pressure

Both Figs. 12 and 13 show the effects of varying the percentageof EGR dilution in the inlet mixture on in-cylinder pressure forboth atmospheric and supercharged inlet pressures, respectively,with a spark timing of 47� BTDC.

The maximum cylinder pressure decreases while the crank an-gle at which the maximum cylinder pressure occurs is furthershifted away from TDC with the increase of percentage of EGR. This

Fig. 12. The effect of varying the percentage of EGR dilution in the inlet mixture onin-cylinder pressure at atmospheric inlet conditions and spark timing of 47� BTDC.

can confirm that EGR dilution is capable of reducing in-cylinderstresses, and consequently, it can be used to suppress abnormalcombustion occurrence such as surface ignition and knock.

Fig. 14 shows the effect of the increase in inlet pressure on max-imum cylinder pressure at different percentages of EGR dilutionand spark timing of 47� BTDC. The increase in inlet pressure in-creased the maximum cylinder pressure while the dilution of theinlet mixture with EGR reduced it to lower levels.

The increase of the inlet pressure of the undiluted air–fuel mix-ture from 101 to 113 kPa increased the maximum cylinder pres-sure by about 16.6% as it increased from 44.6 to 52 bar. However,when the inlet mixture was diluted with EGR with a percentagedilution of 12%, the maximum cylinder pressure decreased byabout 31.5% as it decreased from 52 to 35.6 bar as shown in Fig. 14.

3.5. Heat release rate and heat release

In order to show the effect of the increase of percentage of EGRdilution in the inlet mixture on both the net heat release rate andnet heat release at both atmospheric and supercharged inlet condi-tions, the pressure signals shown in Figs. 12 and 13 were analyzed

Fig. 14. The effect of the increase of inlet pressure on maximum cylinder pressureat different percentages of EGR dilution and spark timing of 47� BTDC.

Fig. 17. The effect of varying the percentage of EGR dilution in the inlet mixture onthe net heat release at atmospheric inlet condition and spark timing of 47� BTDC.

1728 A. Ibrahim, S. Bari / Fuel 89 (2010) 1721–1730

for combustion analysis. Both Figs. 15 and 16 show the effect ofvarying the percentage of EGR dilution in the inlet mixture onthe net heat release rate for both atmospheric and superchargedinlet conditions, respectively.

It can be noted from both Figs. 15 and 16 that the maximumheat release rate decreases with the increase of percentage ofEGR dilution. Also, the crank angle at which the maximum heat re-lease rate occurs is shifted away further from the TDC with the in-crease of percentage of EGR dilution. The increase of EGR dilutionin the inlet mixture decreases the in-cylinder oxygen concentra-tion, and consequently, it reduces the heat release rate.

Both the heat release rate in Figs. 15 and 16 were integrated tocalculate the net heat release. Both Figs. 17 and 18 show the effectof the change of percentage of EGR dilution in the inlet mixture onthe heat release for both atmospheric and supercharged inlet con-ditions, respectively.

The heat release crank-angle relationship has the characteristicS-shape. The heat release starts with almost zero at the time ofspark which was fixed at 313� (or 47� BTDC), and then it startsto increase 10–20� after the spark timing. This angle is usuallycalled flame development angle or sometimes ignition delay. Fol-

Fig. 15. The effect of varying EGR dilution in the inlet mixture on the net heatrelease rate at atmospheric inlet condition and spark timing of 47� BTDC.

Fig. 16. The effect of varying EGR dilution in the inlet mixture on the net heatrelease rate at an inlet pressure of 113 kPa and spark timing of 47� BTDC.

Fig. 18. The effect of varying the percentage of EGR dilution in the inlet mixture onthe net heat release at an inlet pressure of 113 kPa and spark timing of 47� BTDC.

lowing the flame development duration, the heat release notice-ably increases with crank angle until it reaches its maximumvalue where essentially almost all of the fuel chemical energyhas been released at basically the end of combustion.

Both Figs. 17 and 18 indicate that the maximum heat release,which essentially identifies the end of combustion, occurs lateras the percentage of EGR dilution increases.

3.6. Combustion duration

Fig. 19 shows the effect of varying the percentage of EGR dilu-tion in the inlet mixture on the total combustion duration for bothatmospheric and supercharged inlet conditions. The total combus-tion duration was calculated as the crank angle interval from thespark timing to essentially the end of combustion where the heatrelease reaches its maximum value. Fig. 19 indicates that the in-crease of percentage of EGR dilution in the inlet mixture increasesthe total combustion duration for both atmospheric and super-charged inlet conditions. The increase of EGR dilution decreasedthe oxygen concentration during combustion which slowed downthe combustion rate and increased the total combustion duration.

Fig. 19. The variation of the total combustion duration with the percentage of EGRdilution at atmospheric and supercharged inlet conditions.

Fig. 21. The effect of the increase of percentage of EGR dilution in the inlet mixtureon NO emissions at two different inlet pressures.

A. Ibrahim, S. Bari / Fuel 89 (2010) 1721–1730 1729

However, the combustion duration slightly decreased with the in-crease of inlet pressure. The decrease of combustion duration withthe increase of inlet pressure could be due to the increase of theunburned mixture density during combustion which slightly in-creased the burning rate.

3.7. Engine fuel consumption

Fig. 20 shows the effect of varying the percentage of EGR dilu-tion in the inlet mixture on brake specific fuel consumption (bsfc)for both atmospheric and supercharged inlet conditions. EGR dilu-tion had little effect on fuel consumption when it was added to theinlet mixture at atmospheric inlet pressure with lower percentageof dilution of up to about 8%. However, when the EGR dilution wasincreased above 8%, engine fuel consumption started to increaserapidly as the EGR misfire limit approached.

On the other hand, the use of an undiluted stoichiometric air–fuel mixture at an inlet pressure of 113 kPa increased the in-cylin-der pressure and temperature significantly which led to surfaceignition occurrence in addition to a substantial increase in heattransfer, which resulted in an increase in engine fuel consumption

Fig. 20. The variation of brake specific fuel consumption with the percentage ofEGR dilution at atmospheric and supercharged inlet conditions.

compared to the use of diluted inlet mixture at the same inlet pres-sure of 113 kPa. The dilution of the inlet mixture with EGR dilutionof 5% reduced in-cylinder thermal and mechanical stresses whichprevented surface ignition occurrence and decreased cylinder heattransfer, and consequently, this resulted in a significant decrease inengine fuel consumption. Fig. 20 shows that the increase of EGRdilution from 5% to 10% has less significant effect on engine fuelconsumption, however, the engine fuel consumption started to in-crease rapidly as the EGR misfire limit approached.

Increasing the inlet pressure increased engine power as demon-strated in Fig. 2 and significantly decreased fuel consumption asshown in Fig. 20. Engine fuel consumption decreased by about7% when 5% of EGR dilution was employed at an inlet pressure of113 kPa compared to the use of undiluted inlet mixture with an in-let pressure of 101 kPa. However, the use of the maximum EGRdilution limit of 12% decreased engine fuel consumption by only1% compared to using an inlet pressure of 101 kPa without EGR.The supercharger work was neglected during brake power and bsfccalculations as the mechanical supercharger used with the Ricardosingle cylinder engine was assumed to simulate the use of a turbo-charger in a multi-cylinder engine.

3.8. NO emissions

Fig. 21 shows the effect of the increase of percentage of EGR inthe inlet mixture on NO emissions at both atmospheric and super-charged inlet conditions. Fig. 21 indicates that the use of EGR dilu-tion has a significant capability on reducing NO emissions. Theincrease of EGR dilution in the inlet mixture decreased both thecombustion temperature and oxygen concentration which led toa significant decrease in NO emissions.

NO emissions decreased by about 50% when the EGR dilutionwas employed at a rate close to the misfire limit with an inlet pres-sure of 113 kPa compared to the use of undiluted inlet mixturewith an inlet pressure of 101 kPa. On the other hand, using 5% ofEGR dilution with an inlet pressure of 113 kPa reduced NO emis-sions by about 12% compared to using undiluted inlet mixture withan inlet pressure of 101 kPa.

4. Conclusions

In this paper, the use of EGR strategy was experimentally inves-tigated in a natural gas spark-ignition engine. Engine performance

1730 A. Ibrahim, S. Bari / Fuel 89 (2010) 1721–1730

and NO emissions were studied for both atmospheric and super-charged inlet conditions and the following conclusions have beenobtained:

1. The increase of the percentage of EGR dilution in the inlet mix-ture decreases the oxygen concentration, and consequently, itdecreases the combustion rate significantly. For instance, theincrease of EGR dilution from 0% to 10% at atmospheric inletconditions increased the total combustion duration from about74� to 95�.

2. The decrease in combustion rate with the increase of EGR dilu-tion in the inlet mixture helps reduce both the maximum cylin-der pressure and temperature, and consequently, preventsabnormal combustion occurrence such as surface ignition,which occurred at higher inlet pressure.

3. The increase of EGR dilution in the inlet mixture decreases boththe maximum cylinder temperature and oxygen concentrationwhich leads to a significant reduction in NO emissions. Forinstance, NO emissions decreased by about 50% when EGR dilu-tion increased from zero with an inlet pressure of 101 kPa toclose to the misfire limit with an inlet pressure of 113 kPa.

4. The engine tolerance to EGR increases with the increase of inletpressure. The maximum tolerable EGR dilution limit increasedfrom about 10% to 12% when the inlet pressure increased from101 to 113 kPa.

5. The use of an undiluted stoichiometric air–fuel mixture athigher inlet pressure deteriorates engine stability. However,engine stability was improved when the inlet mixture was suf-ficiently diluted with EGR.

6. Engine performance can be significantly improved when theEGR dilution strategy is employed at higher inlet pressure.Engine brake power increased by about 20% and engine fuelconsumption decreased by about 7% while NO emissionsdecreased by about 12% when 5% of EGR dilution was employedwith an inlet pressure of 113 kPa compared to using undilutedstoichiometric inlet mixture with an inlet pressure of 101 kPa.

References

[1] Baldassarri LT, Battistelli CL, Conti L, Crebelli R, De Berardis B, Iamiceli AL, et al.Compressed natural gas and comparison with liquid fuels. Sci Total Environ2006;355(1–3):64–77.

[2] US Environmental Protection Agency, EPA, web information: <http://www.epa.gov/air/urbanair/co/>, <http://www.epa.gov/air/urbanair/nox/noxfldr.pdf>; 2008 [accessed in August 2008].

[3] Demirbas A. Biodiesel production from vegetable oils via catalytic and non-catalytic supercritical methanol transesterification methods. Progr EnergyCombust Sci 2005;31:466–87.

[4] Agarwal AK. Biofuels (alcohols and biodiesel) applications as fuels for internalcombustion engines. Progr Energy Combust Sci 2007;33:233–71.

[5] Akansu SO, Dulger Z, Kahraman N, Veziroglu TN. Internal combustion enginesfuelled by natural gas–hydrogen mixtures. Int J Hydrogen Energy2004;29:1527–39.

[6] Maclean HL, Lave LB. Evaluating automobile fuel/propulsion systemtechnologies. Progr Energy Combust Sci 2003;29:1–69.

[7] International Association for Natural Gas Vehicles, <http://www.iangv.org/>;2009 [accessed 23.06.09].

[8] The Engine Manufacturers Association, USA, The Use of Exhaust GasRecirculation Systems in Stationary Natural Gas Engines, technical report,web information: <www.enginemanufacturers.org/admin/library/upload/830.pdf>; 2004 [accessed in August 2008].

[9] Rabl A. Environmental benefits of natural gas for buses. J Transport Res, Part D2002;7:391–405.

[10] Bhargava S, Clark NN, Hildebrand MW. Exhaust Gas Recirculation in a Lean-Burn Natural Gas Engine. SAE paper no. 981395, 1998.

[11] Saanum I, Bysveen M, Tunestal P, Johansson B.Lean Burn versus StoichiometricOperation with EGR and 3-Way Catalyst of an Engine Fueled with Natural Gasand Hydrogen Enriched Natural Gas. SAE paper no. 2007-01-0015, 2007.

[12] Nellen C, Boulouchos K. Natural Gas Engines for Cogeneration: HighestEfficiency and Near-Zero-Emissions through Turbocharging, EGR and 3-WayCatalytic Converter. SAE paper no. 2000-01-2825, 2000.

[13] Control Technologies for Hazardous Air Pollutants. Handbook, EPA 625/6-91-014, 1991.

[14] Ibrahim A, Bari S. Optimization of a natural gas SI engine employing EGRstrategy using a two zone combustion model. J Fuel 2008;87:1824–34.

[15] Ibrahim A, Bari S, Bruno F. A Study on EGR Utilization in Natural Gas SI EnginesUsing a Two-zone Combustion Model. SAE paper no. 2007-01-2041, 2007.

[16] Einewall P, Tunestal P, Johansson B. Lean Burn Natural Gas Operation vs.Stoichiometric Operation with EGR and a Three Way Catalyst. SAE paper no.2005-01-0250, 2005.

[17] Reppert T, Chiu J. Heavy Duty Waste Hauler with Chemically Correct NaturalGas Engine Diluted with EGR and Using a Three-Way Catalyst, NationalRenewable Energy Laboratory Report, report no. NREL/SR-540-38222, USA,2005.

[18] Heywood JB. Internal combustion engine fundamentals. USA: McGraw-Hill;1988. ISBN:0-07-028637-X.

[19] Czarnigowski J. A Simple Method of Analysis and Modeling of Cycle-to-CycleVariation of Engine Work – The Example of Indicated Pressure. SAE paper no.2007-01-2079, 2007.

[20] Allenby S, Chang WC, Megaritis A, Wyszynski ML. Hydrogen enrichment: away to maintain combustion stability in a natural gas fuelled engine withexhaust gas recirculation, the potential of fuel forming. Automob Eng2001;215:405–18.