Fuel 105 (2013) 425–431

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Fuel

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Primary investigation to leveraging effect of using ethanol fuel on reducinggasoline fuel consumption

Yuan Zhuang ⇑, Guang HongSchool of Electrical, Mechanical and Mechatronic Systems, Faculty of Engineering and Information Technology, University of Technology, Sydney, Australia

h i g h l i g h t s

" Gasoline fuel could be leveraged by ethanol fuel using ethanol direct injection plus gasoline port injection (EDI + GPI)." Brake mean effective pressure and volumetric efficiency increased with the increase of ethanol/gasoline energy percentage." The NO emission was decreased by enhanced charge cooling effect and reduced in-cylinder peak temperature in EDI + GPI." CO and HC emissions increased with the increase of ethanol/gasoline energy percentage.

a r t i c l e i n f o

Article history:Received 20 February 2012Received in revised form 27 August 2012Accepted 5 September 2012Available online 27 September 2012

Keywords:Ethanol direct injectionGasoline port injectionSingle cylinder spark ignition engineLeveraging effect

0016-2361/$ - see front matter � 2012 Elsevier Ltd. Ahttp://dx.doi.org/10.1016/j.fuel.2012.09.013

Abbreviations: ATDC, after top dead center; AFR, adead center; BMEP, brake mean effective pressurconsumption; BSEC, brake specific energy consumptiCA50, crank angle at which the mass burnt fraction isethanol fuel direct injection; ECU, electronic controenergy ratio; GPI, gasoline port injection; IMEP, indicaHE, heating energy; PI, port injection.⇑ Corresponding author. Address: PO Box 123, NS

0433020876.E-mail addresses: Yuan.Zhuang-1@student.uts.edu

uts.edu.au (G. Hong).

a b s t r a c t

Ethanol has been used as an alternative fuel or fuel addicts in spark ignition (SI) engines for years. How-ever, the existing methods of using ethanol fuel, such as blending gasoline and ethanol, pure ethanol andby-fuel of ethanol or gasoline do not make the best use of ethanol’s potentials in improving engine per-formance. Compared with gasoline fuel, ethanol fuel possesses greater octane number and latent heat ofvaporization, which allow higher compression ratio and consequently lead to the increased thermal effi-ciency. Ethanol fuel’s higher combustion velocity could also help increase the combustion efficiency andminimize the energy loss. This paper reports our preliminary investigation to the leveraging effect ofusing ethanol direct injection plus gasoline port injection (EDI + GPI) on reducing the consumption of gas-oline fuel. Experiments were conducted on a YBR250 engine which was a single cylinder SI engine mod-ified to be equipped with EDI + GPI. At each of the four designated engine speeds, the engine load was setto be either medium or light and the ethanol/gasoline energy ratio (EER) was varied from 0% to 60.1%. Therate of the total heating energy of two fuels was kept constant in one of the two engine load conditions.Experimental results were analyzed and discussed in terms of engine performance, in-cylinder combus-tion characteristics and engine emissions. They showed certain leveraging effect of using ethanol fuel bythe increased BMEP, volumetric efficiency and thermal efficiency and reduced NO with the increase ofEER.

� 2012 Elsevier Ltd. All rights reserved.

1. Introduction driving force to searching alternative fuels that are sustainable

The degradation of the global environment and the foreseeablefuture depletion of worldwide fossil fuel reserves have been the

ll rights reserved.

ir fuel ratio; BTDC, before tope; BSFC, brake specific fuelon; CAD, crank angle degree;50%; DI, direct injeciton; EDI,l unit; EER, ethanol/gasolineted mean effective pressure;

W 2007, Australia. Tel.: +61

.au (Y. Zhuang), guang.hong@

and environmental friendly. Ethanol fuel is one of the renewablefuels for addressing these issues. The potential of ethanol fuel inimproving the performance of internal combustion engines hasbeen extensively investigated. Park et al. and Huang et al. foundthat adding ethanol fuel to gasoline could improve the mixtureburnt rate and combustion efficiency due to its high combustionvelocity [1,2]. Nakata et al. and Szybist et al. pointed out thatcharge cooling effect, high heating value of a stoichiometric mix-ture for ethanol blends (per unit mass of air), additional thermody-namic effects on the ratio of specific heats (c) and mole multipliereffect could all attribute to the increase of thermal efficiency whena SI engine was fueled with ethanol or ethanol/gasoline blended[3,4]. Caton et al. and Ayala et al. showed that by taking the advan-tages of ethanol’s anti-knock ability enhanced by its high octane

426 Y. Zhuang, G. Hong / Fuel 105 (2013) 425–431

number and high latent heat of vaporization, engine compressionratio was increased and this consequently improved the enginethermal efficiency [5,6].

However, when ethanol fuel is blended with the gasoline fuelprior to fueling in a spark ignition (SI) engine, the engine uses aconstant ratio of ethanol to gasoline, no matter how engine operat-ing condition changes. In this case, ethanol’s greater ability to sup-press engine knock and to reduce pollutant emissions has nochance to show, but its disadvantages such as low heating valueand low flashing temperature may degrade the engine perfor-mance. The advantages of the ethanol fuel should be implementedand the problems associated with using ethanol as a renewablefuel should be resolved. As the octane number of the ethanol fuelis greater than that of gasoline fuel, this allows the engine to havea greater compression ratio without engine knock. With increasedcompression ratio, the engine thermodynamic efficiency will be in-creased. It was predicted that, due to its latent heat greater thanthat of gasoline fuel, ethanol fuel’s evaporation would reduce thepeak temperature during combustion and consequently allow theincrease of compression ratio. This has led to a new idea of ethanoldirect injection plus gasoline port injection (EDI + GPI) which is ex-pected to use the ethanol fuel in a more effective and efficient wayand to reduce or avoid the problems associated with the currentlyused blended ethanol and gasoline fuel.

Aiming to leverage the effect of the available ethanol in reduc-ing the consumption of gasoline fuel, Cohen et al. first proposed di-rect injection of ethanol fuel in developing a small turbocharged SIengine which should match the performance of a much largerexisting engine [7]. In their report, they estimated ethanol’s energyvalue increased by the leveraging effect on increasing the effi-ciency of using gasoline fuel. This leveraging effect was assessedexperimentally on a Ford ‘EcoBoost’ which was a 3.5 L gasoline tur-bocharged direct injection (GTDI) engine with direct fuel injectionof E85 and port fuel injection of gasoline (E85 DI + Gasoline PI) [8].The compression ratio of the GTDI engine was 9.8:1. It was in-creased to 12:1 for E85 DI + Gasoline PI. Their experimental resultsshowed that the engine thermal efficiency could be improved andthe ethanol fuel could be used to conserve gasoline usage. This ver-ified the estimation of leveraging effect of ethanol fuel in reducingthe consumption of gasoline fuel proposed by Bromberg et al. [9].

Investigation to dual-injection strategies applied to SI engineshas been reported in the past 2 years [12,13]. To investigate theflexibility of dual-injection strategies, Wu et al. [12] conductedexperiments on a single-cylinder SI engine with port injection ofgasoline fuel and direct injection of one of the three fuels: gasoline,ethanol and 2.5-dimethylfuran (DMF). In their experiments, theport injected and the directly injected fuels were tested at five dif-ferent ratios. The spark timing was fixed at the knock-limited max-imum brake toque. Their results showed that the indicated meaneffective pressure (IMEP) increased with the decreased PI massfraction, independent which fuel was directly injected. When eth-anol fuel was directly injected, the indicated efficiency increasedand the HC, NOx and CO2 emissions reduced. Zhu et al. [13] inves-tigated the combustion characteristics of the dual-injection systemon a single-cylinder SI engine. They conducted experiments in lightand heavy load conditions only but with three different combina-tions of PI and DI injections of gasoline and ethanol fuels. Their re-sults showed that, at light load, the IMEP increased by 2% when theethanol fuel was directly injected and the gasoline fuel was port in-jected but decreased in other combinations of dual-fuel injection.Also in light load condition, the percentage of the ethanol fuel af-fected the combustion characteristics significantly. However, atheavy load, the percentage of the fuel directly injected played amore important role in affecting the combustion characteristics.

The leveraged effect of ethanol fuel blended with gasoline fuelat different ratios on the performance of DI gasoline engines has

also been demonstrated in the investigation to blended ethanol/gasoline fuel. To study the feasibility of using ethanol fuel in directinjection (DI) gasoline engines, E0 (100% gasoline) and E100 (100%ethanol) fuels were tested on a V6 3-L DI gasoline engine [10]. Thecompression ratio for E100 was 13:1 and for E0 11.5:1. The en-gine’s full load performance with these two fuels was compared.The engine torque with E100 was well above that with E0 overthe full range of engine speed. The maximum torque with E100was 7.6% greater than that with E0. Their results showed that theE100 fuel was leveraged effectively on this DI gasoline engine withincreased engine torque when the amount of the ethanol fuel in-jected was equivalent to that of the gasoline fuel based on heatingvalues. The torque increment was partially explained as due to theincrease of burned gas mole fraction. The effect of different blend-ing ratios of ethanol/gasoline was also investigated through exper-iments conducted on a single-cylinder 4-stroke DI gasoline enginewith varied spark timing [11]. It was analyzed based on combus-tion performance, regulated emissions and engine efficiency. Theirresults showed that, with the ethanol/gasoline blend ratio in-creased from 0% to 100%, the engine efficiency increased and theemissions of CO and NOx reduced or remained.

This paper reports our preliminary results of investigating theleverage effect of using ethanol fuel on reducing the consumptionof gasoline fuel and on improving engine performance of a singlecylinder research engine equipped with EDI + GPI. The resultsinclude the effect of ethanol fuel energy ratio on the engine perfor-mance such as BMEP, volumetric efficiency, fuel consumption,in-cylinder pressure indicated thermal efficiency and emissions.

2. Experimental setup

2.1. Test engine and instrumentation

The experiments were performed on a research engine whichwas modified from a four-stroke single-cylinder SI engine forYamaha motorcycle YBR250. The specifications of the engine areshown in Table 1. Fig. 1 is the schematic diagram of the enginetesting set up. As shown in Fig. 1, the research engine is equippedwith an electronic control unit (ECU), a direct injection system(9,10,14) for ethanol fuel and a port injection system (16) for gas-oline fuel. The port fuel injection pressure is 250 kPa and the pres-sure in the common rail for direct ethanol fuel injection can beadjusted to be a fixed value in a range of 3–13 MPa. The ethanol di-rect injector is mounted on the same side as the spark plug oppo-site the sprocket of camshaft to avoid interference. There is a slopeangle of 15� between the axis of the injector and the horizontalsurface which is the interior surface of the cylinder head and 12�between the axis of the injector and the vertical surface. The tipof the injector is placed between the intake valve seat and thespark plug, attempted to use the tumble flow to form richer mix-ture adjacent to the spark plug. Both port and direct fuel injectionsare controlled by the ECU.

As shown in Fig. 1, the engine is coupled to a DC dynamometer(2). The in-cylinder pressure was measured using a Kistler 6115Bmeasuring spark plug pressure transducer (15). During the enginetesting, the temperature of the engine body was between 250 �Cand 270 �C. The temperature of the lubricating oil was maintainedbetween 85 �C and 95 �C. The temperature was measured throughK-type thermocouples (8,13,18). A 80 L intake buffer tank (19),with a volume approximately 320 times the engine‘s displacementvolume, was used to stabilize the intake flow. A Bosch wide-bandlambda (12) sensor was mounted in the exhaust pipe. It measuresthe equivalence ratio (k), when engine is operated with gasolinefuel only. The exhaust gas emissions were measured using a HoribaMEXA-584 L gas analyzer (5). Exhaust gas samples were taken at a

Table 1Engine specifications.

Engine type Single cylinder, air cooled, 4-stroke,SOHC

Displacement 249.0 cm3

Bore � stroke 74.0 mm � 58.0 mmCompression ratio 9.80:1Tested engine speed 3500 rpm, 4000 rpmIntake valve opening and closing

timingOpening: 405� BTDCClosing: 120� BTDC

Exhaust valve opening and closingtiming

Opening: 87� ATDCClosing: 381� ATDC

Table 2Test fuel properties.

Ethanol Gasoline

Chemical formula C2H6O C2–C14H/C ratio 3 1.795O/C ratio 0.5 0Gravimetric oxygen content (%) 34.78 0Density@20 �C (kg/m3) 790.9 744.6Research octane number 106 95Stoichiometric air/fuel ratio 9.0:1 14.79:1LHV (MJ/kg) 26.9 42.9LHV (MJ/L) 21.3 31.9Enthalpy of vaporization (kJ/kg) 840 373Temperature at boiling point (�C) 78.4 32.8

Y. Zhuang, G. Hong / Fuel 105 (2013) 425–431 427

position 0.4 m from the exhaust valve and before the three-waycatalyst converter.

2.2. Test fuel

The gasoline used in this test was supplied by BP Australia withan octane number of 95. It represents the most favorable charac-teristics offered by the current market and provides a benchmarkto the ethanol fuel. The ethanol fuel was provided by ManildraGroup. Properties of both ethanol and gasoline fuels were listedin Table 2.

2.3. Experimental procedures

The experiments were conducted at two selected engine loads(light load and medium load) with engine speed varied from3500 rpm to 5000 rpm (500 rpm interval). At each engine load con-dition, the total energy input value (ethanol + gasoline) was fixedand the EER (defined in Section 3) was varied from 0% (gasolineonly) to 60.1%. The ethanol direct injection fuel pressure was fixedat 4 MPa when EER was less than or equal to 48.3% and 6 MPawhen EER was greater than 48.3%. Spark timing and ethanol injec-tion timing were fixed at 15 crank angle degrees (CAD) BTDC and300� BTDC respectively in the experiments. 15 CAD BTDC wasthe spark timing in the original engine control system. 300 CADBTDC injection timing was for providing sufficient time for ethanolfuel and fresh charge from the intake port to mix, so that the mix-ture before combustion was homogenous. The inlet air flow ratesin different EER conditions were calculated prior to the experi-ments to achieve stoichiometric air/fuel ratio with dual fuels inthe experiments.

Fig. 1. Schematic of th

3. Results and discussion

This section presents and discusses the experimental results inthree subsections: engine performance, combustion characteristicsand engine emissions. In each subsection, results at differentethanol/gasoline energy ratios were studied. As described in Eq.(1), the ethanol/gasoline energy ratio (EER) in this paper is definedas the rate of the heating energy (HE) of the ethanol fuel divided bythe rate of the total heating energy of ethanol and gasoline fuels.The rate of heating energy is equal to the fuel mass flow ratemultiplied by the lower heating value (LHV) of the fuel.

Ethanol=gasoline energy ratioðEERÞ ¼_HEEthanol

_HEEthanol þ _HEGasoline

ð1Þ

where _HE (rate of heating energy) = fuel mass flow rate � LHV. Thedenominator in Eq. (1) is the rate of the total heating energy of thetwo fuels. This total heating energy rate was kept unchanged ineach of the two engine load conditions, and the EER was variedby changing the mass flow rates of both ethanol and gasoline fuels.

3.1. Engine performance

To examine the leveraging effect of using ethanol fuel on reduc-ing the gasoline fuel consumption, the variation of engine perfor-mance with different EER is presented and discussed in thissubsection. The engine performance is assessed with parametersincluding the brake mean effective pressure (BMEP), volumetricefficiency, brake specific fuel consumption (BSFC) and brake spe-cific energy consumption (BSEC).

e engine system.

50%

55%

60%

65%

70%

75%

0% 10% 20% 30% 40% 50% 60% 70%

Fig. 3. Variation of volumetric efficiency with EER.

300.0

350.0

400.0

450.0

500.0

550.0

600.0

0% 10% 20% 30% 40% 50% 60% 70%

Fig. 4. Variation of BSFC with EER.

428 Y. Zhuang, G. Hong / Fuel 105 (2013) 425–431

Fig. 2 shows the variation of BMEP with EER. It can be seen thatthe engine BMEP increases with the increase of EER at all the testconditions. This shows that the direct injection of ethanol fuelcould help increase the engine power output of a port-injectiongasoline engine when the energy of the ethanol fuel was equiva-lent to that of the replaced gasoline fuel. Therefore, the ethanol fuelcould be leveraged effectively to reduce the gasoline consumption.Factors contributing to the increase of BMEP may include chargecooling effect associated with fuel injection and ethanol’s high la-tent heat of vaporization, ethanol fuel’s high combustion velocity,high energy content of stoichiometric mixture per unit mass ofair, and mole multiplier effect. They will be further discussed inthe following sections.

Volumetric efficiency is an important parameter to evaluate theleveraging effect of using ethanol fuel. As shown in Fig. 3, with theincrease of EER, the volumetric efficiency is increased in both en-gine load conditions. This could be due to the charge-cooling effectand the reduced partial pressure associated with port fuel injec-tion. When the fuel is injected into the intake port, the charge cool-ing effect is enhanced by the heat taken by the injected fuel. Inaddition, the stoichiometric air/fuel ratio (AFR) of the ethanol fuelis smaller than that of the gasoline fuel. To maintain a stoichiom-etric AFR of the mixture and a constant total energy input, moremass of ethanol fuel than that of gasoline fuel is needed. On theother hand, the partial pressure of the port injected gasoline fueldecreases with the reduced quantity of gasoline fuel when theethanol fuel is directly injected into the combustion chamber. Allof these factors help increase the volumetric efficiency.

The increase of the air flow equates to the increase of fuel en-ergy flux in a stoichiometric mixture, and the lower heating value(LHV) of a stoichiometric mixture per unit mass of air increaseswith the increased ratio of ethanol fuel to gasoline fuel [4]. Thismay be explained as follows. When the EER ratio increases, moremass of ethanol fuel than that of gasoline fuel is required to main-tain the input energy value unchanged. Thus more air is required tokeep mixture at a stoichiometric value. The additional air flow dueto the increase of ethanol usage and volumetric efficiency increasesthe mixture’s LHV and consequently increases the engine poweroutput.

Despite of the improved BMEP and volumetric efficiency, thelow LHV of ethanol, compared with gasoline, still dominates thefuel consumption. Fig. 4 shows that the BSFC increases when theEER increases. This should be due to the ethanol fuel’s lowerLHV, so that more fuel is needed to maintain the same engine brakepower. To evaluate the effect of EER based on the fuel’s energy raterather than mass flow rate, BSEC is used to evaluate the fuel con-sumption in terms of total fuels energy rather than the quantityof the fuels consumed. BSEC, as described by Eq. (2), was definedin [14].

0.7

0.6

0.5

0.4

0.3

0.20% 10% 20% 30% 40% 50% 60% 70%

Fig. 2. Variation of BMEP with EER.

BSEC ¼_HEEthanol þ _HEGasoline

P

¼_mEEthanol � LHVEthanol þ _mGasoline � LHVGasoline

Pð2Þ

where _HEEthanol and _HEGasoline are the heating energy rates, as definedin Eq. (1), of the ethanol fuel and gasoline fuel respectively. _mEthanol

and _mGasoline are the mass flow rates (g/s) of the ethanol and gasolinefuels. LHVEthanol and LHVGasoline are the lower heating values of theethanol and gasoline fuels. P is the engine power.

Fig. 5 shows the variation of BSEC with EER. It can be seen that,with the increase of EER, less energy input is required to achievethe same engine power output as that with gasoline fuel only.

The improvement of engine power output with the same energyinput value may also be due to another factor, the mole multipliereffect. The mole multiplier effect is defined as the ratio of the num-ber of product moles (n products) to the number of reactant (nreactants), as described by Szybist et al. in [4]. They used Eq. (3)to describe how the mole multiplier effect affected the engineperformance.

W ¼ n � �R � T1

c� 1� 1� P2

P1

� �c�1=c" #

ð3Þ

where W is the work done by the fuel mixture in the expansion pro-cess in one engine cycle, n is the number of moles in the combustionproducts, �R is the universal gas constant, T1 is the initial tempera-ture, P1 and P2 are pressure values in initial and finial states ofthe expansion process, and c is the ratio of constant-pressure andconstant-volume specific heats.

The mole multiplier effect is described in Eq. (3). Assuming thatother parameters be kept unchanged in the expansion process,

1200014000160001800020000220002400026000

0% 10% 20% 30% 40% 50% 60% 70%

Fig. 5. Variation of BSEC with EER.

Fig. 6. In-cylinder pressure at different EER at 3500 rpm and light load.

Fig. 7. Maximum in-cylinder pressure at different EER at 3500 rpm and light load.

Hea

t rel

ease

rate

(j/C

AD)

CAD ( ATDC)

Fig. 8. Heat release rates at different EER at 3500 rpm and light load.

Y. Zhuang, G. Hong / Fuel 105 (2013) 425–431 429

increased ratio of ethanol fuel to gasoline fuel would enhance themole multiplier effect because of the increased ratio of numberof moles in the combustion products to the number of moles inthe reactants according to the calculations in [4]. This means thatmore moles of combustion products would be available to covertmore heat energy to work during the expansion process whenthe percentage of ethanol fuel is increased. The expansion workis also a function of c. Greater c will result in greater expansionwork based on Eq. (3). Considering ethanol fuel’s c which is greaterthan the gasoline fuel’s in the standard condition, further increaseof W is possible with the c increased by increasing EER.

3.2. In-cylinder combustion characteristics

To understand the combustion in EDI + GPI, the results of in-cylinder pressure, heat release rate, CA50 (Crank Angle of 50% massburnt fraction) and indicated thermal efficiency at various EER at3500 rpm and light load condition are discussed in this subsection.As indicated in the previous section, the total energy input valuewas fixed in this test condition.

Fig. 6 shows the variation of cylinder pressure with the crankangle for four different ERR values of 0%, 24.3%, 48.4% and 60.1%.The corresponding peak pressure varying with EER is shown inFig. 7. As shown in both figures, the increase of EER results in thepeak pressure increased from 1.428 MPa to 2.08 MPa. The occur-rence of the peak pressure is slightly more advanced with the in-crease of EER (E0% at 24�BTDC, E24.3% at 23�BTDC, E48.4% at23�BTDC and E60.1% at 18�BTDC). These could be attributed tothe increase of volumetric efficiency and mole multiplier effectand ethanol fuel’s higher combustion velocity, compared withthe gasoline fuel’s combustion velocity.

Figs. 8 and 9 illustrate the heat release rate and CA50 derivedfrom the same pressure data as that for the results shown in Figs. 6and 7. As shown in Fig. 8, the heat release rate at the beginning of thecombustion process is almost independent of the EER. However,after that, the heat release rate is increased rapidly with the increaseof EER. The peak value of the heat release rate also increases with theincreased EER. CA50 describes the crank angle position where theaccumulated heat release reaches 50% of the total released heat.The CA50 of an effective engine should occur after the TDC. SmallerCA50 means that work was done more effectively by the combustionproduct on the piston which is at a position closer to but just after theTDC. As shown in Fig. 9, the CA50� is 36� after the TDC with 0%ethanol fuel and reduced to 32� when EER is increased to be 24.3%.It is further reduced to be 27� when the EER is 60.1%. The reducedCA50 with increased EER indicates that the timing for 50% of the fuelburnt is getting closer to the TDC, so that the work done by the com-bustion product on the piston is increased with the increased EER.

The indicated thermal efficiency is also used to assess the lever-age effectiveness of the direct injection of ethanol fuel. Fig. 10

shows the variation of indicated engine thermal efficiency withEER at 3500 rpm and light engine load condition. As it can be seen,with the increase of EER, the indicated thermal efficiency graduallyincreases from 24.3% with EER of 0% to 28.6% with EER of 60.1%.This could be resulted from the high laminar flame velocity andlow flame temperature of ethanol fuel, which reduced the heatlosses through the cylinder wall. More efficient fuel energy conver-sion and stronger charge-cooling effect of the ethanol fuel thanthat of the gasoline fuel may also contribute to the increase of indi-cated thermal efficiency [15].

3.3. Engine emissions

Figs. 11–13 show the variation of engine emissions with EER, inthe same test conditions as that for the results shown in Figs. 2 and4. As shown in the figures, the brake specific nitric oxide (BSNO)decreases but the brake specific hydrocarbon (BSHC) and brake

CA5

0 (

ATD

C)

Fig. 9. Variation of CA50 with EER at 3500 rpm and light load.

Fig. 11. Variation of BSCO with EER.

Fig. 13. Variation of BSHC with EER.

Fig. 12. Variation of BSNO with EER.

430 Y. Zhuang, G. Hong / Fuel 105 (2013) 425–431

specific carbon monoxide (BSCO) increase with the increase of EER.Fig. 11 shows that the BSCO measurements at light load conditionsare less than that at medium load conditions up to EER of 48.4%.When the EER is over 48.4%, more CO is produced at light load con-ditions than that at medium load conditions. The BSHC results attwo engine load conditions also cross at EER of 48.4%, as shownin Fig. 13. As illustrated in Fig. 12, with the increase of EER, theBSNO in most of tested conditions (except at 5000 rpm and lightload) first increases until EER reaches 24.3%, then it reduces withthe increase of EER. As sufficiently high temperature lasting for acertain time period is the necessary condition to form the NOproduct in combustion, the increase of BSNO may be because theethanol fuel improves the engine combustion and increase thein-cylinder temperature. The decrease of NO with the increasedEER may be resulted from the in-cylinder temperature decreasedby the increased percentage of ethanol fuel in combustion. Thereduction of BSNO with the increase of EER validates the studyconducted by Nakata et al. [3]. In their results, NOx was reducedwith the increased volumetric percentage of ethanol fuel in E10,E50 and E85. As they analyzed, the reduction of NOx was due tothe combustion temperature reduced by the ethanol fuel. TheNOx was reduced with the increased volumetric ratio of the ethanolfuel blended with the gasoline fuel. The decrease of in-cylindertemperature may be attributed to two factors. One is the high la-tent heat of vaporization of the ethanol fuel, which decreases thein-cylinder temperature when it vaporizes. The other factor is thatthere are more triatomic molecules in the combustion products ofethanol fuel than that of the gasoline fuel. The more the triatomicmolecules are produced, the higher the gas heat capacity and thelower combustion gas temperature will be. However the low in-cylinder temperature can also lead to the increment of unburned

Fig. 10. Variation of indicative thermal efficiency with EER at 3500 rpm and lightload.

combustion products. Although the oxygen atom contained inthe ethanol fuel can enhance the oxidation process and in turn re-duce CO and HC emissions, the pre-mentioned two factors maymake the in-cylinder temperature too low to complete the com-bustion. As a result, the CO and HC emissions may increase withthe increase of EER. This may explain why the CO and HC emissionsare greater in medium load conditions than that in light load con-ditions when the EER is over 48.3%, as shown in Figs. 11 and 13. Itis because the in-cylinder temperature increases with the in-creased engine load and the ethanol fuel’s cooling has become lessdominating to reducing the cylinder temperature.

4. Conclusions

Aimed to investigate the leveraging effect of using ethanol fuelon reducing the consumption of gasoline fuel in SI engines, exper-iments were conducted on a single-cylinder 4-stroke air-cooled

Y. Zhuang, G. Hong / Fuel 105 (2013) 425–431 431

research engine equipped with EDI + GPI. In the experiments, thedefined ethanol/gasoline energy ratio (EER) was varied from 0%(gasoline only) to 60.1% at two engine loads and four engine speedconditions between 3500 rpm and 5000 rpm. The total energy in-put of ethanol and gasoline fuels was kept constant in each ofthe two engine load conditions. The variations of engine perfor-mance, emissions and combustion characteristics with the EERwere analyzed and discussed. Conclusions can be drawn as follows.

1. The BMEP and volumetric efficiency were increased with theincrease of EER. This indicates that to achieve the comparableengine brake power, less energy input would be required in aSI engine equipped with EDI + GPI. Hence the gasoline fuel con-sumption could be reduced by the leveraging use of ethanolfuel.

2. The maximum in-cylinder pressure was increased and the CA50was reduced with the increase of EER. This shows that theengine expansion work increased when the ratio of ethanol fuelto gasoline fuel was increased. The indicated thermal efficiencyincreased consequently.

3. The NO emissions decreased and the CO and HC emissionsincreased with the increase of EER. The reduced NO could bedue to the in-cylinder temperature reduced by directly injectedethanol fuel greater than that of port injected ethanol gasoline.The increase of CO and HC emissions could also be attributed tothe relatively low in-cylinder temperature which led to incom-plete combustion.

4. The leveraging effect of ethanol fuel on improving engine per-formance could be attributed to factors such as the coolingeffect of the ethanol fuel directly injected into the combustionchamber, the LHV of the stoichiometric mixture per unit massof air increased with EER, mole multiplier effect and ethanolfuel’s high combustion velocity.

Acknowledgments

The scholarship provided by the China Scholarship Council isacknowledged. The authors would like to express their great

appreciation to Manildra Group in Sydney for providing theethanol fuel and to the FEIT Workshop at UTS for their technicalassistance and support.

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