Lean Burn Performance of a Spark Ignition Engine with an Ethanol

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Experimental study on lean burn performance of a spark ignition engine with ethanol-gasoline dual injection system Yuan Zhuang, Yejian Qian, and Guang Hong Energy Fuels, Just Accepted Manuscript • DOI: 10.1021/acs.energyfuels.7b03028 • Publication Date (Web): 13 Feb 2018 Downloaded from http://pubs.acs.org on February 22, 2018

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Experimental study on lean burn performance of a spark ignition engine with ethanol-gasoline dual injection system Order of Authors: Yuan Zhuang, Yejian Qian*, Guang Hong First Author: Yuan ZHUANG, PhD Affiliation: School of Automotive and Transportation Engineering, Hefei University of Technology Address: No.93 Tunxi Road, Hefei, China Phone: +86 551 62901787 E-mail address: [email protected] Second and Corresponding Author: Yejian Qian, PhD Affiliation: School of Automotive and Transportation Engineering, Hefei University of Technology Address: No.93 Tunxi Road, Hefei, China Phone: +86 551 62901787 E-mail address: [email protected] Third Author: Guang HONG, PhD Affiliation: School of Electrical, Mechanical and Mechatronic Systems, Faculty of Engineering and Information Technology, University of Technology, Sydney Address: PO Box 123, NSW 2007, Australia Phone: +61 02 95142678 E-mail address: [email protected]

Abstract A spark ignition (SI) engine with ethanol direct injection (EDI) and gasoline port injection (GPI) system has been studied quite recently. In order to investigate its application, experiments were carried out on a single-cylinder 250cc short stroke motorcycle SI engine, which is supplied with ethanol-gasoline dual injection system. Experimental findings have revealed that EDI helped increase the lean combustion limit owing to the combination effects of the dual injection and ethanol fuel. At the identical engine speed and load, the lean burn limit in the dual injection model was, on average, 0.2 times higher than that in the corresponding single GPI model. Increasing ethanol energy ratio (EER) gave rise to the increase in the lean burn limit; nevertheless, the increase rate began reducing when the EER was above 48% and 42% at 4000rpm and 3500rpm, correspondingly. Subsequently, the ethanol direct injection (LEDI), following the intake valve closing slightly extended the lean burn limit, probably, owing to the unsatisfactory mixture quality. Early ethanol direct injection (EEDI) prior to the intake valve closing was more helpful in increasing lean burn limit as compared with the LEDI. Nonetheless, further advance of spark timing earlier than 35 CAD BTDC in LEDI situation, in addition 30 CAD BTDC in EEDI situation, might 1

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lead to the reduction of lean burn limit. Keywords: Ethanol; Gasoline; Dual injection system; Lean burn

1. Introduction: Owing to the increasing attention towards the energy safety, together with the issue of atmospheric pollution, a bigger demand asks for enhancing engine efficiency, in addition to minimizing the pollutant emissions. Stratified or lean combustion is a technology, which possesses the application to meet these needs [2]. Lean combustion might be beneficial for engine thermal efficiency with the help of a reduction of pumping losses, heat losses and endothermic dissociation losses [3]. Operating with excess oxygen might result into a complete combustion, accordingly, generating more energy for each mole of fuel [4]. Moreover, the mixture’s ratio of specific heats exhibits an increase with the growth of air-fuel ratio (AFR); additionally, it might theoretically lead to better engine thermal efficiency [4]. In the emission part, the emissions in lean circumstances are different from their stoichiometrically fueled counterparts. Hydrocarbon (HC) and carbon monoxide (CO) emissions typically decrease more owing to thorough combustion. Nevertheless, as the lean burn approaches the misfires, the combustion exhibits inconsistency with the increase in the production of HC and CO emissions [2]. Nitric oxide (NOX) emissions normally reach the highest point in slightly lean circumstances (1.05≦λ≦1.1, λ or lambda is defined as excess air coefficient) [2]. Further leaning the mixture above this range might bring down the NOX due to low peak cylinder temperature. In lean circumstances, the three-way catalyst is no longer capable of converting NOX with high efficiency. Accordingly, the NOX emissions require being sufficiently low in order to offset this disadvantage. NOX less than 100ppm is typically termed as optimum [4]. Despite the merits, lean burn is not easy to control owing to the fact that the extreme lean fuel mixture can give rise to an inconsistent combustion. In the lean burn condition, the mixture near the spark plug might become hard to be ignited; moreover, the flame propagation might be hampered until the misfire takes place 2


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[6]. This unstable combustion confines the upper AFR limit of lean burn [8]. Several technologies have been in development for the purpose of making sure a stable ignition and complete combustion, for instance, high energy ignition, which adds flammable gases and the use of alternative fuels like ethanol [10]. Ethanol fuel, as a promising alternative fuel, has been investigated and proved that it is capable of improving the lean burn constancy, together with extending the lean burn limit owing to its wider flammability limit (λ of 0.4~1.7 in ethanol as compared with λ of 0.5~1.3 in gasoline) and higher laminar flame rate (39 cm/s as compared with 33 cm/s of gasoline). It was discovered that the laminar flame rate of the blend fuel E10 (10% ethanol and 90% gasoline) did not exhibit reduction until it reached λ value of 1.2. Nevertheless, the laminar flame rate of pure gasoline started decreasing when the λ was just above 1.1 [12]. Correspondingly, investigations involving the port fuel injection (PFI) engines have discovered the fact that using ethanol/gasoline blends could more largely extend lean burn limit in comparison with using pure gasoline. As revealed by Alexandrian et al., the highest lean burn limit in pure gasoline circumstances was 1.2. It was extended to 1.4 when making use of 40% ethanol/gasoline blends. Also, it was discovered that light engine load and low engine speed both were more suitable for lean burn, owing to the fact that the lean burn limit in these situations was higher than the limit in the heavy load and high rate circumstances [14]. The combustion stability represented by the coefficient of variation of IMEP (COVIMEP) had been reported that it lowered when the ethanol/gasoline blends were put to use. In general, approximately 20% improvement in the COVIMEP was achieved; moreover, the improvement becomes more obvious when the ethanol/gasoline ratio is increased [16]. The advent of DI technology has made stratified combustion possible in the SI engines through the production of a rather rich local mixture, which is adjacent to the spark plug in the compression stroke while keeping a highly lean global AFR (λ>>1)[17]. The engine load, in this method, was governed by the fuel amount, together with the SOI timing. The latest stratified combustion direct injection spark ignition (DISI) 3


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engine has realized a stable operation at AFR≈40:1 that is way more than the normal lean burn limit of AFR≈18:1 in homogeneous lean burn [18]. Nevertheless, the stratified combustion is not conveniently achievable as it asks for attentive optimization of the end of injection timing regarding the spark timing, in addition to a slight fluctuation that might lead to the misfires [19]. Moreover, late injection with stratification could give rise to noticeable fuel impingement on the piston crown. When the fuel is unsatisfactorily vaporized prior to the commencement of ignition, the engine combustion will be negatively affected and lean burn limit would be reduced. For ethanol, its volatility is greater than that of gasoline if the ambient temperature is higher than 410K [20]. This might help lower the length of period for the fuel evaporation, together with producing more homogeneous mixture in the lean circumstance. Therefore, the recent developments in the lean burn technology have thrown more emphasis on making use of ethanol for the purpose of realizing the stable stratified combustion [19]. However, as the latent heat of vaporization of ethanol is twice bigger than that of gasoline (Table 2), the stronger charge cooling impact of the ethanol vaporization might substantially lower the in-cylinder mixture temperature that is found as a crucial parameter in lean combustion [22]. Typically, the unburned mixture temperature below 1000K is termed as might be leading to the obvious reduction in the lean mixture burning rate and causing the unstable combustion [21]. Accordingly, the stratified combustion might be negatively affected by the ethanol fuel when the DI ethanol amount is great enough for substantially bringing down the unburned mixture temperature. The SI engine with dual injection system (EDI+GPI) has the application to avoid this disadvantage of ethanol in the stratified combustion. In the EDI+GPI mode, the EDI amount could be lowered, together with being adjusted to a level that would not significantly bring down the unburned mixture temperature, meanwhile, still maintaining the quick formation of sufficient amount of locally rich mixture surrounding the spark plug. The total energy input would not be influenced by the decreased EDI amount all through the procedure, owing to the fact that the GPI might be adjusted in order to maintain the 4


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total energy input. Accordingly, the lean burn operation load window would not be affected. Furthermore, the homogeneous mixture provided by the GPI might further improve the flame propagation in the stratified combustion, together with offsetting the impact of inhomogeneous mixture caused by the direct injection technology [24]. Therefore, the EDI+GPI engine has the application for the improvement of the stratified combustion. Recently, Oh et al.[25] and Hemdal et al.[26] discovered the fact that the fast vaporization rate of the ethanol fuel spray could help form more homogeneous fuel mixture in a DI engine. This property was inclined to improving the mixture quality in the stratified combustion and substantially improving the engine lean burn performance. Daniel et al.[27] investigated the lean combustion of ethanol fuel in a DISI engine. The empirical findings demonstrated that making use of ethanol more largely extended the lean burn limit as compared with using gasoline. NOX and COVIMEP decreased in the DI ethanol circumstances. Wu et al.[28] investigated the impact of AFR on the engine operations and emissions in a SI engine fueled with the ethanol/gasoline blends. They discovered that making use of the ethanol/gasoline blends could improve the engine torque output. CO emissions decreased with the increase of AFR and HC emissions lowered primarily, followed by a slight rise with the growth of AFR. As suggested by the review provided above, it might be realized that the combination of ethanol fuel and DI technology in a SI engine with dual injection system has great application prospects in increasing the lean burn limit. That is why, as a novel approach, the lean burn performance in a SI engine with ethanol-gasoline dual injection system (EDI+GPI) is worth investigation. This paper is throws focus on this issue. 2. Experimental setup 2.1 Engine and instrumentation As the experimental setup was elaborated in [29], only brief information about the rig, together with the 5


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details associated with this work, was introduced here. Figure 1 sheds light on the systematic sketch of the engine bench. An adjusted Yamaha YBR250 motorcycle engine was utilized that featured pent-roof, side-mounted spark plug and flat topped piston crown. The engine stroke was 58mm and the displacement was 250cc. Table 1 presents the engine’s key features. As presented in the Figure 1, the research engine comprised an electronic control unit (ECU), in addition to a direct injection system (9, 10, 11, 14) for the ethanol fuel as well as a port injection system (16) for the gasoline fuel. The engine incorporated two injection systems. The first was the original gasoline PI system, whereas, the other one was the ethanol DI system. The ethanol DI system concluded a six holes injector as well as a Bosch returnless high pressure pump. The injector had a 34° spray angle and 17° bent axis; moreover, the fuel pressure could be maintained at a steady range between 30 bar and 130 bar. It requires being taken into notice that as the engine used in this test figured with short stroke and small displacement which are normally less than conventional multi-cylinder production passenger engines in the market, the empirical findings from this engine may be affected, which give rise to the issues, for instance, high engine speed, high COV and wall-wetting, which is likely to limit the universality of the results. The ethanol injector was installed on the left side of cylinder head with a 15° from the under surface of the cylinder head, and with a 12° from the vertical surface owing to the packaging constraint. It requires being observed that the electrode of spark plug is positioned adjacent to the edge of the sprays. In this case, there is a locally affluent mixture around the area close to the spark plug. Figure 2 reveals the relative positions of the ethanol injector and spark plug. AFR was monitored with the help of a Bosch wide-band lambda sensor (12) when the SI engine was fueled using the pure gasoline. In the ethanol-gasoline dual injection model, it was measured by the means of a Horiba MEXA-584L gas analyzer. When the engine was in dual-injection mode, the fuel’s chemical composition of H/C (hydrogen to carbon ratio) and O/C (oxygen to carbon ratio) were primarily calculated 6


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from two fuels flow with the use of a self-developed Matlab code. Subsequent to that, these values were input to the gas analyzer for the real-time lambda monitoring and controlling. The fuels’ flows were calculated by the means of the measurement of the injection pulse width. The relation between injection pulse width and fuel mass flow was calibrated while setting up the test rig. During the tests, the same engine circumstances were recorded three times, followed by taking an average for the purpose of minimizing the empirical uncertainty. Error bars were employed for lean burn limit in order to exhibit the significance of the trends owing to the unstable nature of the inlet air flow in a single cylinder engine. Measurements of the accuracies of Total HC,CO, CO2 and NOx emissions were carried out by the Horiba MEXA-584L gas analyzer, which were within±12 ppm, ±0.06%, 0.5% and ±30 ppm, correspondingly. For more details of the test, refer to [29]. The measurement uncertainty can be divided in A and B types of uncertainty. The value of the type A uncertainty is obtained with the help of an array of experimental measures by U A = s2 =

1 j ( x j − x)2 ∑ j =1 j ( j − 1)

where s is the mean global variance, j indicates the number of observations, ‫ ݆̅ݔ‬represents the mean values acquired for each point and ‫ ̅ݔ‬depicts the mean of all the acquired data. The type B uncertainty is dependent on the quantity measured, being associated with the previous measurements or calibration data of the sensors and is presented by

UB =

B k

where B is associated with the instrument variable leading to the uncertainty (smaller scale, value available by the manufacturer) and k is a representation of a coefficient that takes into account the probability that the measured value can be found in a determined interval (probability distribution). The combined measurement uncertainty is provided by 7


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U C = U C2 + 2B

For the variables derived from the directly measured quantities through some mathematical operation with the primary variables, the derivative approach was employed for the determination of the combined uncertainty as N

U C2 = ∑ i =1 [

∂f 2 ] U C ( xi ) ∂xi

where ݂ indicates the variable function of the directly measured variables and ܷܿ(‫ )݅ݔ‬is the combined uncertainty of each of the variables ‫݅ݔ‬. In a bid to ensure a 95% confidence level, the combined uncertainty values were multiplied by a factor expanding the probability level within the t-student distribution. The sensors as well as the instruments used, together with their measurement uncertainties have been presented in the Table 2.

2.2 Test fuels Table 3 lists the specifications of both the ethanol and gasoline fuels. These data were provided by their corresponding suppliers. The gasoline fuel was provided by the BP Australia, whereas, the ethanol fuel was supplied by the Manildra Group. The purity of the ethanol fuel is 99.9%.

2.3 Experimental procedures The engine was initiated and heated using pure gasoline fuel. When the lubricate oil temperature rose to 90±5°C, the amount of the gasoline was lowered, whereas, that of ethanol with the same amount of energy was supplied in order to keep the total energy unchanged. During the tests, the engine parameters, for instance, rate, load, EER, ethanol fuel SOI timing and spark timing were first reached designated level at stoichiometric AFR. Subsequently, the AFR was progressively increased through the enlargement of the opening angle of the throttle. When mixture became increasingly lean, the IMEP first raised and achieved the highest value; subsequent to that, it began decreasing with further growth of AFR. During this process, 8


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when the COVIMEP was below 10% and the IMEP had a 2% reduction from its maximal value, the AFR was defined as lean burn limit. However, if the COVIMEP was greater than 10%, irrespective of the increase of IMEP, the AFR was adjusted until the COVIMEP was less than 10% again. The lean burn limit in this study is defined as the highest achievable λ. The method of defining 2% reduction in IMEP as a marking for lean burn limit was put forward by Albert et al.[30]. Making use of the combustion stability less than of 10% COVIMEP as another marking for lean combustion limit deals with making sure the stable combustion and that the drive comfortable capability would not be affected by the engine cyclic fluctuation[31]. Equation 1 puts forward the definition of EER. & & & Ethanol / gasoline energy ratio ( EER ) = HE Ethanol ( HE Ethanol + HEGasoline )

Where

& , rate of heating energy, kJ / s ) = ( HE

(1)

fuel mass flow rate ( g / s ) × LHV.

The denominator in

Equation 1 represents the total heat output of the gasoline and ethanol fuels. COVIMEP is defined as: k

COVIMEP = [

∑ ( IMEP − IMEP ) i

2

/ (k − 1) / IMEP]

i =1

where IMEP represents the indicated mean effective pressure of the ith engine cycle (Bar), IMEP indicates the average value of the indicated mean effective pressures of k cycles (Bar), and k depicts the number of engine consecutive cycles. Experiments were, initially, carried out for the purpose of studying the impact of engine conditions on the lean burn performance of the engine with ethanol-gasoline dual injection system. Test involving the impact of engine speed was carried out in the range between 3500rpm and 5000rpm at medium load (5.5 Bar -7.5 Bar IMEP). Moreover, the test for the impact of engine load (IMEP) was carried out at IMEP range from 4.5 Bar to 8.5 Bar with 1 Bar interval and rate of 4000rpm. The reason for the selection of 3500rpm as the starting rate was because the engine operation below 3500rpm was unstable (COV>10%). Spark timing for the test regarding the impact of engine load was at 30 CAD BTDC for 4.5 Bar IMEP, which was retarded 3 9


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CAD for each 1 Bar IMEP increment, in accordance with the map of engine original control unit. The reason for the retard spark timing meant avoiding the knocking, together with ensuring the combustion stability. EERs of 24% and 48% were selected as the baseline ratios. The engine fueled with the pure gasoline (EER of 0%) was also measured for the purpose of comparison. Secondly, the experiments were carried out for the investigation of the impact of EER and ethanol SOI timing on lean burn performance. In the test regarding the impact of EER on lean burn, the EER level was varied from 0% to 60% with engine speed at 4000rpm and load at medium level. The minimum achievable EER in this study was 24% that was owing to the limitation of minimum injector opening pulse. Injection pressure was fixed at 40 Bar all through this part of experiments. The ethanol SOI timing was varied in two ranges 50 CAD BTDC ~ 110 CAD BTDC (defined as LEDI) and 270 CAD BTDC ~ 330 CAD BTDC (defined as EEDI) as the tests were carried out for studying the impact of SOI timing on the lean combustion. The interval for each SOI timing sweep step was 20 CAD. Accordingly, the impact of EDI on homogeneous lean burn (early fuel direct injection) and stratified lean burn (injecting fuel later in the compression stroke) might be compared. When the ethanol fuel SOI timing was in the scope of 120 CAD BTDC and 250 CAD BTDC, the engine tests were not carried out owing to the fluctuation of the IMEP on the bases of the conclusions published in [32]. The spark timing was kept at 25 CAD BTDC in order to guarantee sufficient time for fuel atomization, mixture formation and stable combustion. Engine speed was set at 4000rpm with the load at medium level. The EDI pressure for EER 24% was set at 40 Bar that was subsequently raised to 90 Bar when the EER was greater than 48%. It could ensure the similar injection (around 60 CAD) duration when different amount of ethanol was injected. Finally, the experiments were carried out for studying the impact of spark timing on the lean combustion. The engine load was varied from medium range to heavy range (IMEP 5.5 -8.5 Bar) at two EERs of 24% and 48%. The reason for raising engine load range to high level (8.5 Bar IMEP) revolved around putting 10


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efforts for the combination of the impact of lean burn and knock mitigation that was EDI+GPI engine’s another advantage. Suppressing engine knock is another approach for the improvement of the engine efficiency owing to the fact that the engine knock usually limits the SI engine efficiency at the high load. If the efficiency improvement caused by the knock mitigation and lean burn could be combined, the engine efficiency might reach a higher level. Nevertheless, this combination is obvious merely at the high load where the knock is likely to happen, in addition to being constrained by the adoption of optimal spark timing. Accordingly, the engine load was increased to heavy range in order to figure out how much engine efficiency could be gained from these joint impacts in this investigation. Table 4 sheds light on the test matrix; moreover, more details of the testing conditions might be found in the same table.

3. Results and discussion 3.1 Impact of speed and load on EDI+GPI engine lean burn In the lean burn, the variation of engine speed exerts impact on the available mixture surrounding the spark plug for ignition [2]. Figure 3 reveals the impact of engine speed on EDI+GPI engine lean burn limit at the medium load. It might be observed that the lean burn limit at three EERs exhibits a gradual decline as the engine speed is increased from 3500 rpm to 5000 rpm. The reason suggests that the high engine speed is capable of enhancing the in-cylinder turbulence that, in turn, improves the DI fuel vaporization as well as flame propagation velocity. Conversely, the high turbulence in the cylinder can increase the risk of ignition blown away from the spark plug [9], which prolongs the ignition delay, together with lowering the combustion stability (Figure 4), eventually, limiting further increase of the lean burn limit. It can be observed that the impact of ignition blowing away might play a crucial role in exerting impact on the lean burn limit owing to the fact that the engine speed is the only variable. Low engine speed appears to be more appropriate for reaching higher lean burn limit. It can also be observed 11


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that the lean burn limit in the EDI condition is greater than that in the GPI condition from the Figure 3. The lean burn limit enlarges if the EER is increased from 24% to 48%. The variation of lean burn limit with the engine speed for different EERs is analogous. The lean burn limit reaches the highest when engine speed is 3500 rpm. The impact of engine speed on COVIMEP at the lean burn limit has been presented in the Figure 4. The combustion becomes unstable as the fuel mixture is leaned. Accordingly, it becomes quite imperative to monitor COVIMEP in a bid to maintain the combustion stability in an acceptable range. It might be observed that the COVIMEP stays stable at three EERs when the speed is in the range from 3500rpm to 4000rpm. At more than 4000rpm, COVIMEP increases sharply with the increase of engine speed. COVIMEP exhibits the highest increases in EER 0%. It increases from 2.1% to 3.7% when the speed is raised from 4000rpm to 5000rpm. The increase of COVIMEP with speed in EDI condition is quite moderate. When the speed is raised from 4000rpm to 5000rpm, COVIMEP at EERs of 24% and 48% increase only 0.8% and 0.5%, correspondingly. It is indicated that the combustion is negatively affected by the increase of engine speed in the lean burn condition. Ignition blowing away might be one of the key factors causing the increase of COVIMEP[5]. In order to further analyze the possibility of ignition blowing away, which gives rise to the decrease of the lean burn limit, the combustion initiation duration, CA0-10%, is defined as the crank angle degrees starting from the spark timing to the timing of 10% of the fuel mass burnt, which has been presented in the Figure 5. This result has been presented here because it is closely associated with the formation of ignition spot and initial flame growth. As evident from the Figure 5, the CA0-10%, exhibits a slight increase at first with the increase of speed when the engine speed is less than 4000rpm. At more than 4000rpm, the CA0-10% exhibits a substantial increase with further increase of the engine speed. This result reveals the fact that the increase of engine speed is harmful for the formation of ignition spot and initial flame growth both. The 12


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increased propensity of the ignition blowing might be one of the key reasons for this result, owing to the fact that the unstable ignition directly affects the initial flame propagation as well as combustion initiation duration. It can also be observed that CA0-10% at EERs of 24% and 48% is shorter than at 0% EER from the Figure 5. This could be explained by the fact that the ethanol exhibits a better low temperature combustion stability and faster laminar flame rate. Previous investigations have shed light on the fact that there was a correlation between COVIMEP and CA0-10%. It was figured out that COVIMEP started rising greatly when the CA0-10% was over certain scope [32]. In this paper, when the CA0-10% is equal to 25 CAD, it is termed as a turning point. Figures 4 and 5 reveal that COVIMEP increases sharply at all the EERs, if CA0-10% is greater than 25 CAD. Nevertheless, the combustion mechanism keeps stable and COVIMEP is around 1.8% when the CA0-10% is lower than 25 CAD. Despite the fact that increasing the engine speed is harmful for the extension of the lean burn limit, the indicated specific hydrocarbon (ISHC) emission decreases with the increase of engine speed, as evident from the Figure 6. The decrease of ISHC with the engine speed might be explained by two mechanisms. At first, the increase in the engine speed might improve the DI fuel vaporization, flame propagation and combustion efficiency. Accordingly, the mixture quality and combustion process both are improved, which result into lower HC emissions. Secondly, the lean burn limit decreases with the increase in the engine speed. Richer mixture can minimize the risk of miss fire in the lean condition; therefore, the ISHC minimizes with the growth of engine speed. The engine speed of 4000rpm was chosen as the baseline speed for the rest of the tests on the bases of the analysis of lean burn limit and HC emissions. Figure 7 illustrates the variation of lean burn limit with the engine load at 4000rpm. As evident from this figure, there is a symmetrical decay of lean burn limit at all the three EERs when the IMEP is above or 13


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below 6.5 Bar. The lean burn limit decays when the IMEP is lower than 6.5 Bar that might be explained by the increased throttling loss and reduced charge efficiency caused by the decreased throttle opening. On the decrease in the throttle opening, the reduced charge efficiency results into more residual of exhaust gases trapped in the cylinder, which eventually leads to the spark lag and unstable combustion. The lean burn limit, accordingly, should be lowered in order to guarantee the combustion stability. The lean burn limit decay when the IMEP is more than 6.5 Bar might be caused by the reduction of the combustion speed. Low combustion speed might give rise to low engine efficiency, which is especially obvious at the heavy load owing to the fact that the fuel injection amount is large at the high engine load; additionally, the low combustion speed might significantly elongate the combustion duration of such a large amount of fuel, together with augmenting the heat losses [2]. Accordingly, the lean burn limit requires being lowered for the purpose of ensuring the fast and stable combustion for these heavy loads. It can also be observed that the rate of lean burn limit decay at EERs of 24% and 48% is smaller than that in the GPI condition (EER of 0%) at either side of 6.5 bar IMEP, which is evident from the Figure 7. The ethanol-gasoline dual injection leads to a higher lean burn limit at all the tested loads. The impact of engine load on the lean burn limit at different EERs is similar in this test. Figure 7 reveals the fact that the highest lean burn limit is achieved at IMEP around 6.5 Bar. Accordingly, the medium engine load ranging between 5.5 bar and 7.5 bar might be selected as the baseline range for the rest of the tests. The reason for not choosing a specific load, for instance, 6.5 bar IMEP, as the baseline condition, suggest that a specific engine load has no engineering implications when it is in real-world application. In a bid to further analyze the feasibility of this load range (IMEP 5.5 to 7.5 bar), the variations of COVIMEP, NO and HC emissions with engine load at the lean burn limit have been presented in the Figures 8, 9 and 10, correspondingly. As evident from the Figure 8, the COVIMEP at all the three EERs generally decrease with the increase in the engine load. If the load is in the range between 5.5 bar and 7.5 bar, the 14


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COVIMEP at the three tested EERs are almost constant, irrespective of the relatively high lean burn limit in this load range (Figure 7). Figures 9 and 10 shed light on the fact that the ISHC and ISNO first decrease with the increase in the engine load until the IMEP reaches around 6.5 bar at the three EERs. When the IMEP is greater than 6.5 bar, ISHC and ISNO increase progressively with the further increase of IMEP. The highest ISNO among the three EERs is at IMEP of 8.5 bar and the highest ISHC is at IMEP of 4.5 bar. In the above analysis, the lean burn limit initially increases with the increase in the engine load if the IMEP is lower than 6.5 bar (see in Figure 7). The decrease of ISNO and ISHC in this load range might be explained by the increased lean burn limit, decreasing the in-cylinder temperature, in addition to providing more amount of fresh air for HC oxidization. Lower in-cylinder temperature might reduce the NO formation and more amount of fresh air might supply more oxygen for the HC oxidization. When the engine load is further raised from 6.5 bar to 8.5 bar, the lean burn limit decreases. Accordingly, the excess air reduces and in-cylinder temperature prior to the ignition increases, which gives rise to more complete combustion and higher combustion temperature. ISHC and ISNO increase subsequently. The low ISHC at the load range from 5.5 bar to 7.5 bar suggest the fact that the lean burn limits chosen for these conditions are appropriate owing to the fact that there is no serious increase in the HC emissions, which might be caused by the misfire or incomplete combustion due to too lean mixture. Briefly, the medium engine load range (5.5 bar to 7.5 bar) is regarded as an appropriate range for the lean burn on the bases of the above analysis. These conditions were chosen as the baseline conditions in the following tests.

3.2 Impact of EER on lean burn The ethanol fuel is termed as one of the key factors influencing the lean burn performance in a SI engine with ethanol-gasoline dual injection system. The effects of EER on the lean burn limit, efficiency and emissions are going to be discussed in this section for further analysis. 15


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The variations of the lean burn limit with EER at speed of 3500rpm and 4000rpm have been presented in the Figure 11. As evident from the figure, the lean burn limit increases progressively with the increase of EER. There is a slowdown of the increasing trend when the EER reaches 42% at 3500rpm and 48% at 4000rpm. As reviewed, ethanol exhibits a wider flammability limit, faster laminar flame rate and better combustion stability as compared with gasoline. The wider flammability limit permits ethanol/air mixture to have a higher AFR in order to avoid misfire or incomplete combustion. Faster laminar flame rate and favorable low temperature combustion stability might ensure stable flame propagation and combustion when the mixture is leaned. Accordingly, the EDI technology can extend the lean burn limit. Despite the fact that the SOI timing of ethanol fuel is at 300 CAD BTDC, yet the adjacent deployment of spark plug and injector tip might still contribute towards the creation of the rich mixture close to the spark plug when the ethanol is directly injected into the cylinder chamber. Nevertheless, when the EER is at high level (over 42% at 3500rpm and 48% at 4000rpm), the serious charge cooling effect and lean burn could lead to low in-cylinder temperature that prolongs the combustion duration, together with lowering the combustion stability and, accordingly, constraining further extension of the lean burn limit [4]. This might be a part of the reasons for the slowdown of the lean burn limit increase trend when the EER is greater than 42% at 3500rpm and 48% at 4000rpm. Figure 12 presents the variations of COVIMEP with EER at the lean burn limit. It is quite evident that the COVIMEP at both the engine speeds stay at low level (˂2%) when the EER is less than 48 %. At the EER above 48%, the COVIMEP increases substantially with the increase in the EER. The leap of COVIMEP when EER is greater than 48% might be explained by the low in-cylinder temperature. When EER is over certain level (EER of 48%), the charge cooling effect of the ethanol fuel vaporization has a great influence on the in-cylinder temperature. It lowers the combustion rate, together with giving rise to the unstable combustion. The great charge cooling effect causing the deterioration of combustion stability at stoichiometric AFR was 16


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discovered by author’s previous study [33]. Hanabusa et al.[21] and Zeldovich et al.[22] also reported the similar results suggesting that the low in-cylinder temperature reduced the lean combustion stability. Another factor that might lead to high COVIMEP is the inhomogeneous mixture quality at high EER levels. For the same engine conditions (including engine speed, spark timing, injection pressure and SOI timing), the time period between SOI and ignition is constant. When the EER is increased, the injection duration prolongs and the time for the fuel vaporization reduces, resulting into an unsatisfactory mixture quality, followed by deteriorating the combustion process and increasing COVIMEP. The variations of indicated thermal efficiency with EER have been presented in the Figure 13. The indicated thermal efficiency increases monotonously with the increase of EER. The reason suggests involvement of the increased lean burn limit, mole multiplier impact and ethanol's high combustion speed [35].The highest indicated thermal efficiency is 37.9% for 3500rpm and 37.5% for 4000rpm at EER of 60%. When the EER is greater than 48%, the increase rate of indicated thermal efficiency begins slowing down. The slowdown of increase trend might be associated with the reduced increase rate of the lean burn limit (seen in Figure 11), great charge cooling effect and bad mixture quality. The variations of ISNO and ISHC with EER at the lean burn limit have been manifested in the Figures 14 and 15. The ISNO increases with the increase of EER until it reaches 48% at 4000rpm and 42% at 3500rpm, as evident from the Figure 14. When the EER is greater than these two ratios, the ISNO declines with the further increase of the EER. This result is different from the result at stoichiometric AFR where ISNO decreases linearly with the increase of the EER [33]. This variation tendency is believed to be a combination of lean burn limit and charge cooling effect. The increase of ISNO when the EER is less than 48% at 4000rpm and 42% at 3500rpm might be owing to the raised lean burn limit that provides more oxygen for the oxidization process. The increased locally rich mixture spot caused by the raised DI amount (EER) might be another reason leading to the increased NO emissions [37]. The decrease of ISNO when the EER is above 17


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48% at 4000rpm and 42% at 3500rpm might be resulted from the great charge cooling effect caused by the high EER level. As regards the ISHC, as evident from the Figure 15, it first decreases when the EER is less than 24%. Thereafter, it stays stable in the EER range from 24% to 48%. Subsequent to that, it increases sharply with the further increase of the EER. The leap of ISHC, when the EER is over 48%, can affirm the previous discussion regarding the great charge cooling deteriorates the lean combustion, together with slowing down the increase of the lean burn limit (seen in Figure 11). This is owing to the formation of HC emissions is primarily from the incomplete combustion in the lean burn condition [4]. Too lean mixture or low in-cylinder temperature might increase the tendency of incomplete combustion. The lean burn limit in the EER range from 48% to 60% only slightly increases (Figure 11). Accordingly, the low in-cylinder temperature is likely to be the major factor for the high HC emissions.

3.3 Impact of injection timing on lean burn The SOI timing is a critical controlling parameter because it severely disrupts the formation of ignitable mixture in an SI engine with direct injection system. In the stratified combustion, the influence of SOI timing becomes more critical. The SOI timing and spark timing require being accurately controlled for ensuring the stable combustion where slight fluctuation in flow field can result into misfire. Accordingly, it is quite essential to figure out the relationship between SOI timing and lean burn limit. It ought to be pointed that the similar study has been carried out previously by the author and the conclusions might be observed in [39]. Nonetheless, the engine operation condition of this test is different from the previous test; moreover, the result presented here is meant to ensure the content integrity of this report, together with providing a comprehensive analysis of the impact of EDI+GPI on lean burn. The variation of the lean burn limit with ethanol fuel SOI timing has been presented in the Figure 16. As evident from it, the EEDI demonstrates an obvious advantage over LEDI in extending the lean burn limit. 18


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Additionally, the lean burn limit initially grows with the advance of SOI timing, followed by reducing with the further advance of SOI timing for the EEDI and LEDI condition. In the EEDI and LEDI condition, the peak value of the lean burn limit occurs at 290 CAD BTDC and 90 CAD BTDC, correspondingly. The maximal lean burn limit (λ) achieved in this test is 1.33 where the SOI timing sets at 290 CAD BTDC and the EER is 48%. The impact of the EER on the lean burn limit is different in EEDI and LEDI condition. The lean burn limit at EER of 24% is lower than that at EER of 48% and the maximum value is 1.33 in EEDI model. Nevertheless, the lean burn limit (λ) at EER of 24% is higher than that at EER of 48% and the λ is 1.1in LEDI condition. Prior investigation figured out the fact that the LEDI can produce stratified charge in the area close to the spark plug [22,32]. Theoretically, the stratified charge is more beneficial for extending the lean burn limit. However, the LEDI exhibits harmful influence on the lean burn and the lean burn limit (λ) in the LEDI model is much smaller than that in the EEDI model. This result might be because of the severe fuel impingement caused by the LEDI, low in-cylinder flow rate when the inlet valve closed and the lack of time for the fuel vaporization before spark discharge. Previous study dealing with the same engine has discovered this phenomenon [18]. Detailed mechanism for this phenomenon was explained with the help of numerical simulation, which has been reported in [32]. These negative factors might negatively affect the combustion, decreasing the combustion stability, and prolonging the combustion duration. That is why the lean burn limit decreases. The λ reaches the maximum value at SOI timing of 290 CAD BTDC and EER of 48% because of the homogenous quality of the fuel-air mixture. The similar conclusions were also discovered in [32]. Figure 17 illustrates the variation of the indicated thermal efficiency with SOI timing at two EERs. The indicated thermal efficiency at both the tested EERs in the LEDI model is lower than that in the EEDI model. The SOI timing sheds light on a weak effect on the indicated thermal efficiency in the EEDI model as it stays around 36.7% at EER of 24% and 37.2% at EER of 48%, correspondingly. The indicated thermal 19


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efficiency exhibits an increase trend with the advance of SOI timing and stays stable when the SOI timing is greater than 70 CAD BTDC in the LEDI model. In comparison with the trend presented in the Figure 16, the EER of 48% demonstrates a positive impact on the indicated thermal efficiency in the EEDI model but it is inversed in the LEDI model. The indicated thermal efficiency result presented in the Figure 17 might be associated with the formation of air and fuels mixture. In the EEDI condition, the time for fuel evaporation is longer than that in the LEDI condition, accordingly, leading to more homogeneous mixture that enhances combustion quality and finally increases the indicated thermal efficiency. In the LEDI condition, the DI ethanol fuel might easily intrude on the piston crown as the piston is near the top dead center (TDC) [20]. This wall-wetting might result into the unsatisfactory mixture homogenous quality and combustion process, accordingly, lowering the indicated thermal efficiency. When SOI timing is advanced in the LEDI circumstances, the time available for fuel evaporation and mixture formation is extended [17] that improve the mixture quality, in addition to generating the higher in-cylinder temperature. The combustion quality is therefore improved subsequently and the efficiency increases. Figure 18 highlights the fluctuation of COVIMEP with SOI timing. As evident from the figure, COVIMEP enlarges with the advance of SOI timing at the tested EER in the LEDI and EEDI condition. In the EEDI condition, COVIMEP achieves the highest at SOI timing of 330 CAD BTDC, where it reaches 4.82% at EER of 24% and 3.97% at EER of 48%. In the LEDI condition, COVIMEP at SOI timing of 110 CAD BTDC is higher than that at other timings, which is 9.1% and 7.21% at EER of 24% and 48%, correspondingly. COVIMEP at SOI timing of 50 CAD BTDC reaches the minimum in the LEDI condition. The reason is the low lean burn limit at this timing (seen in the Figure 16), because richer mixture might improve the flame propagation process, in addition to contributing to the improvement of the combustion stability. Author’s previous tests have suggested that the raise of EER might improve the engine combustion 20


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stability at the stoichiometric AFR [33]. Nevertheless, the increase of EER exhibits an opposite impact on COVIMEP in this test. As likely to be evident from the Figure 18, the COVIMEP in the LEDI condition at the EER of 48% is bigger than its counterpart at the EER of 24%. This result might be because of the fact that the engine is running at leaner circumstances at EER of48% than that at EER of 24%, which has been presented in the Figure 16. The variations of ISNO and ISHC emissions with SOI timing have been illustrated in the Figures 19 and 20, correspondingly. As evident from the Figure 19, ISNO at EER of 24% in the EEDI model remains unchanged when SOI timing varies. It enlarges incrementally with the advance of SOI timing when the EER is raised to 48%. The increase of ISNO at EER of 48% might be because of the fact that the engine is running at leaner condition at the EER of 48% than that at EER of 24% (seen in Figure 16). Excessive amount of fresh air might provide more oxygen in the formation of NO emissions. In addition, the high EER level (48%) might result into more locally rich mixture spots; furthermore, these spots lead to the increase of NO emissions [37]. The ISNO at both the EERs stays at low level with slight increase when the SOI timing is advanced in the LEDI model. Figure 20 suggests that the effect of SOI timing on the ISHC is small at EER of 24% in both the EEDI and LEDI models. At EER of 48%, the ISHC in EEDI model increases progressively with the delay of SOI timing. Conversely, ISHC in the LEDI condition gradually decreases with the delay of SOI timing. The decrease of ISHC with the delay of SOI timing at EER of 48% in the LEDI condition might be explained by the decreased lean burn limit (seen in Figure 16) that lowers the available oxygen for HC oxidization. The decrease of ISHC in the EEDI condition might be caused by the increased time for the DI fuel vaporization, enhancing the mixture feature and increasing the heat recovery from cylinder wall to the charge.

3.4 Impact of spark timing on lean burn Spark timing is another parameter that might significantly affect the lean burn performance. As the 21


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mixture is leaned, the mixture becomes hard to be ignited and the combustion duration is prolonged. Accordingly, the spark timing ought to be carefully adjusted in order to make sure stable ignition and proper combustion duration [2]. Figure 21 highlights the variation of lean burn limit with spark timing. It ought to be observed that the spark timing sweep for EERs of 0% (GPI) and 24% is stopped at the knock limited spark advance (KLSA) found in the stoichiometric AFR in order to prevent knocking as well as other abnormal combustion despite the fact that it is known that the lean burn might decrease the propensity of knocking [39]. At the EER of 48%, no knocking is found at the stoichiometric AFR when the spark timing is advanced to 45 CAD BTDC. That is why the spark timing for the EER of 48% is swept up to 45 CAD BTDC, which is the upper limit in this investigation. As it might be observed in the figure, the lean burn limit generally decreases with the delay of spark timing. It begins decreasing if the spark timing is earlier than 35 CAD BTDC in the LEDI model and 30 CAD BTDC in EEDI model at EER of 48%. The increase of lean burn limit with the spark timing can be explained by the fact that the advanced spark timing can lead to the combustion process closer to TDC. More combustion process is, accordingly, occurring at a smaller in-cylinder volume, resulting into greater combustion pressure and temperature. The increased combustion pressure and temperature shorten the combustion duration that makes the combustion less affected by the piston movement and cylinder flow. Accordingly, the combustion stability is improved that contributes to the extension of the lean burn limit. Further advancing of spark timing over certain range might lower the mixture quality owing to the less time for ethanol vaporization. Furthermore, excessively advanced spark timing decreases the charge pressure and temperature prior to the ignition. The low pressure and temperature would negatively affect the initiation of flame kernel, accordingly, elongating the flame progress phase [21]. That is why the lean burn limit decreases. The variation of IMEP with spark timing has been presented in the Figure 22. As evident from the figure, 22


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IMEP initially increases with the advance of spark timing at all the EERs. The increase of IMEP might be explained by the advanced combustion phasing that is a result of the advance of spark timing, and more complete combustion, caused by the raised lean burn limit (seen in the Figure 21). At the EERs of 0% (GPI) and 24%, the spark timing sweep stops on reaching the KLSA. Spark timing at the EER of 48%, however, might be advanced to 45 CAD BTDC owing to the low knock tendency at this EER level. IMEP at this EER manifests a decline trend when the spark timing is earlier than 40 CAD BTDC in the LEDI model and 35 CAD BTDC in the EEDI model. Spark timing of 35 CAD BTDC and 40 CAD BTDC are the timings that achieve the highest brake torque (MBT). The MBT timing is set as the spark timing that produces the maximal IMEP at unchanged total energy flow rates. At the EER of 48%, the charge cooling effect of the direct injection and ethanol’s high octane rating, suppressing the knock tendency, are the factors permitting the spark timing to be advanced such early. The impacts of EDI on the knock mitigation and the lean burn are well combined when the EER is 48%. In the EEDI condition, IMEP at 35 CAD BTDC is the greatest. IMEP at the EER of 0% (GPI) and 24% is higher than that at EER of 48% in the LEDI condition. The reason is the unsatisfactory mixture homogeneous quality that deteriorates the combustion process. There is a liner relationship between the IMEP and lean burn limit, which can also be observed in the Figure 22. By analyzing the Figures 21 and 22, it might be observed that the MBT timing for the EEDI and LEDI conditions is 35 CAD BTDC and 40 CAD BTDC at the EER of 48%, which has 5 CAD retard as compared with the timings for the highest lean burn limit. This result suggests that the peak IMEP does not occur at the timing for the highest lean burn limit; moreover, the similar conclusions were reported in [40]. The indicated thermal efficiency for the KLSA or MBT timing has been illustrated in the Figure 23. As it is presented, the highest indicated thermal efficiency is 38.7% that is at the EER of 48% in the EEDI condition. This value is greater than the indicated thermal efficiencies found in the tests presented in the previous section. It is associated with the combined benefits of the lean burn and knock mitigation, which 23


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are caused by the EDI. In comparison with the findings presented in the section 3.3, the indicated thermal efficiency in the EEDI model at both the EERs is higher than that in the other models. It is termed as the lowest (33.5%) in the GPI model. The variation of COVIMEP with spark timing has been presented in the Figure 24. It might be observed that COVIMEP increases with the delay of spark timing at three EERs, except in the LEDI condition at which the COVIMEP slightly increases when the spark timing is earlier than 25 CAD BTDC at EER of 24% and 40 CAD BTDC at EER of 48%. The decrease of COVIMEP might be explained by the advanced spark timing that lowers the combustion duration, together with advancing the combustion phasing, accordingly, making the combustion less affected by the piston motion and turbulence [27]. The slight increase of COVIMEP in the LEDI condition is ascribable to the reduced mixing time. It can also be observed in the Figure 24 that the COVIMEP in the EEDI condition is lower than that in the other conditions. The LEDI produces the highest COVIMEP value among the other injection strategies (GPI and EEDI). The low COVIMEP in the EEDI condition might be explained by the improved mixture homogenous quality and enhanced low temperature combustion stability of the ethanol fuel. The variation of ISNO with the spark timing has been presented in the Figure 25. As suggested by the figure, the ISNO decreases with the delay of spark timing. ISNO in GPI and LEDI condition is at a low level which is less than 2.0 g/kW·hr. When the injection strategy is shifted from LEDI to EEDI, the ISNO decreases rapidly with the delay of the spark timing. ISNO in the EEDI model is about 0.8 g/kW·hr, which is higher than that in the GPI model at the EER of 24%. There is the highest ISNO when the EER is 48% in the EEDI model. It might result from the higher lean burn limit (seen in Figure 21) that provides more oxygen for nitrogen oxidization. The variation of ISHC with spark timing has been presented in the Figure 26. Generally, ISHC increases with the advance of spark timing at the lean burn limit, which can be explained by the fact that when the 24


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spark timing is advanced, the highest in-cylinder pressure increases, making more hydrocarbons to be trapped in the crevice volumes. It can also be observed that the HC emissions at the EER of 23% are less than that at EERs of 0% (GPI) and 48% in both the LEDI and EEDI conditions. ISHC at the EER of 48% is the highest among the other EERs. It might be associated with the high in-cylinder pressure (IMEP, Figure 22). The high ISHC at the EER of 48% in the LEDI condition might be ascribable to the low cylinder temperature which is caused by the late fuel injection, and unsatisfactory mixture homogeneous quality that is caused by the reduced time for vaporization. Misfire and partial combustion ought to not be the causes for the high ISHC at the EER of 48% in the EEDI model. Nonetheless, these might be the causes for high ISHC in the LEDI condition, owing to the fact that the COVIMEP at the EER of 48% remains low level (less than 4%) in the EEDI model, whereas the COVIMEP in the LEDI condition is high. The high COVIMEP in the LEDI condition suggests the unstable combustion, for instance, misfire and partial combustion might occur.

4. Conclusions In order to study the combination effects of dual fuel injection and using ethanol fuel on the lean combustion, the detailed experiments had been carried out on a short stroke small (250cc) motorcycle single cylinder SI engine that was modified with the ethanol-gasoline dual fuel injection system. Different engine conditions (speeds and loads), EERs, SOI timings and spark timings have been tested for the evaluation of the lean burn performance. The key conclusions have been summarized as hereunder: 1. The lean burn limit in the EDI+GPI condition is less susceptible to the changes of engine speed and load than that in the GPI condition. In general, the lean burn limit decreases with the increase of engine speed, and first increases then decreases with the increase of engine load at the three tested EERs. Low rate (4000rpm) and medium load range (5.5-7.5 Bar IMEP) are termed as the suitable condition for the achievement of higher lean burn limit, in addition to being chosen as the baseline conditions. 2. The combination of DI and ethanol fuel (EDI) is capable of effectively extending the lean burn limit in a 25


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certain range. The lean burn limit increases with the increase of EER and the increase rate begins slowing down when the EER is greater than 48% at 4000rpm and 42% at 3500rpm. When the EER is less than 48%, ISHC and COVIMEP remain at the low level. Nevertheless, when the EER ratio is greater than 48%, they increase sharply, which might be resulted from the great charge cooling effect at the high EER level (EER˃48%). ISNO exhibits an opposite trend in comparison with the ISHC. It first incrementally enlarges with the increase of EER, and then begins decreasing when the EER is greater than 48% at 4000rpm and 42% at 3500rpm. 3. EEDI is more influential than the LEDI on increasing the lean burn limit. The maximal λ reached in this model is 1.33 at EER of 48%. The engine efficiency, combustion and combustion stability (COVIMEP) are improved in the EEDI condition. The lean burn limit in the LEDI condition at the EER of 48% is just above the stoichiometric value (λ=1) whereas the lean burn limits in other conditions are all above 1.1.COVIMEP and indicated thermal efficiency generally decreases with the delay of SOI timing in both EEDI and LEDI models. 4. Both the EDI+GPI and GPI only conditions are sensitive to the variation of the spark timing. When the spark timing is delayed than 35 CAD BTDC in the LEDI model and 30 CAD BTDC in the EEDI model, the lean burn limit is increased with the advance of spark timing. Further advance of the spark timing leads to the decrease of lean burn limit. When the spark timing is advanced in a certain range in the high load condition, knock could still happen in the lean combustion. As EDI is also an effective approach in the knock suppression, the combined effects of knock mitigation and lean burn yield to higher indicated thermal efficiency. The highest indicated thermal efficiency is 38.7% when the spark timing is 35 CAD BTDC and the EER is 48% in the EEDI condition. In terms of emissions, ISNO and ISHC generally increase with the advance of spark timing. In this study, the EDI+GPI exhibit its effectiveness on extending the lean burn. However, as the engine 26


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employed is a single cylinder motorcycle engine figured with short stroke (58mm), and small displacement air cooled and side DI system, some engine test conditions are accordingly confined, for instance, making use of high engine speed for the purpose of avoiding the high COV owing to its short stroke, decreasing DI pressure to reduce wall-wetting in small engine displacement and relative high COVIMEP up-limit as compared with the multi-cylinder vehicle engines. All these factors make the conclusion obtained in this investigation; moreover, this very engine exhibits some limitations to be translated to more conventional engines (multi-cylinder passenger engines). If this test is carried out at a production DI engine in which the stroke is normally at a range between 80-90mm (suggesting that the engine can operate steadily in the low speed and COV can also be improved) and the DI fuel formation process is fully optimized, the conclusion is likely to be different. The LEDI is likely to exhibit more benefits as compared with the EEDI in extending the lean burn limit owing to the fact that the stratified mixture abound the spark plug and higher EER (>42%) may also continuously sustain the further improvement of the lean burn limit rather than decreasing it because of deteriorated mixture quality. Emissions are also likely to benefit from the reduced testing speed and optimized fuel formation process. Nevertheless, whether these predictions are true or require not being proved on the real engine test, fortunately, the author is currently modifying a 1.5L turbo charged DI engine with dual injection system. Further investigations on the lean burn in multi-cylinder commercialized engine will be carried out once the experimental facility is ready for the use.

Acknowledgments The work in this study received the financial support from the National Natural Science Foundation of China (Grant No.51606056 and 51676062), Anhui Provincial Natural Science Foundation with Grant No. 1708085QE106 and 1708085ME102, State Key laboratory Open Foundation (Grant No.K2016-06). The 27


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authors would like to express their great appreciation to Manildra Group in Sydney for the provision of the ethanol fuel, in addition to the technical assistance provided by the UTS Engineering Workshop.

Nomenclature ATDC = After top dead center BTDC = Before top dead center CO = Carbon monoxide DI = Direct fuel injection EDI = Ethanol fuel direct injection EEDI = Early ethanol direct injection ERR = Ethanol/gasoline energy ratio ECU = Electronic control unit ISCO = Indicated specific carbon monoxide ISHC = Indicated specific hydrocarbon ISNO = Indicated specific nitric oxide ISEC = Indicated specific energy consumption IMEP = Indicated mean effective pressure PFI = Port fuel injection LHV = Low heat value LEDI = Late ethanol direct injection SOI = Start of injection HE = Heating energy HC = Hydrocarbon 28


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NO = Nitric oxygen PFI = Port fuel injection Lambda (λ) = air/fuel equivalence ratio

Reference [1] Bradley D. Fundamentals of lean combustion. Lean Combustion: Technology and Control, 2008; 19-53. [2] Dunn-Rankin D. Lean combustion: technology and control. Academic Press. 2011. [3] Pannone G, Johnson R. Methanol as a Fuel for aLean Turbocharged Spark ignition Engine. SAE Tech. Pap. Ser. 1989,Paper No. 890435. [4] Germane G, Wood C, Hess C. LeanCombustion in Spark-Ignited Internal Combustion Engines -A Review. SAE Tech. Pap. Ser. 1983,Paper No. 831694. [5] Rousseau S, Lemoult B, Tazerout M. Combustion characterization of natural gas in a lean burn spark-ignition engine. Proceedings of the Institution of Mechanical Engineers, Part D: Journal of Automobile Engineering 1999; 213(5): 481-489. [6] Piock W, Weyand P, Wolf E, Heise V. Ignition Systems for Spray-Guided Stratified Combustion. SAE Int. J. Engines 2010; 3(1): 389-401. [7] Toulson E, Schock H, Attard W. A Review of Pre-Chamber Initiated Jet Ignition Combustion Systems. SAE Tech. Pap. Ser. 2010,Paper No. 2010-01-2263. [8] Hassaneen A, Varde K, Bawady A, Morgan A. A Study of The Flame Development and Rapid Burn Durations In A Lean-Burn Fuel Injected Natural Gas S.I. Engine. SAE Tech. Pap. Ser. 1998,Paper No. 981384.

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[9] Li L, Wang Z, Wang H, Deng B. A Study of LPG Lean Burn for a Small SI Engine. SAE Tech. Pap. Ser. 2002,Paper No. 2002-01-2844. [10] Saanum I, Bysveen M, Tunestål P, Johansson B. Lean Burn Versus Stoichiometric Operation with EGR and 3-Way Catalyst of an Engine Fueled with Natural Gas and Hydrogen Enriched Natural Gas. SAE Tech. Pap. Ser. 2007,Paper No. 2007-01-0015. [11] Beeckmann J, Kruse S, Peters N. impact of Ethanol and n-Butanol on Standard Gasoline Regarding Laminar Burning Velocities. SAE Tech. Pap. Ser. 2010,Paper No. 2010-01-1452. [12] Beeckmann J, Röhl O, Peters N. Experimental and Numerical Investigation of Iso-Octane, Methanol and Ethanol Regarding Laminar Burning Velocity at Elevated Pressure and Temperature. SAE Tech. Pap. Ser. 2009,Paper No. 2009-01-1774. [13] Wu CW, Chen RH, Pu JY. The influence of air–fuel ratio on engine performance and pollutant emission of an SI engine using ethanol–gasoline-blended fuels. Atmospheric Environment 2004; 38(40): 7093-7100. [14] Alexandrian M, Schwalm M. Comparision of ethanol and gasoline as automotive fuels. ASME papers 1992, 92-WA/DE-15 [15] Warnberg J, Boehmer M, Denbratt I. Optimised Neat Ethanol Engine with Stratified Combustion at Part-load; Particle Emissions, Efficiency and Performance. SAE Tech. Pap. Ser. 2013,Paper No. 2013-01-0254. [16] Liang C. Investigation on the performance of a spark-ignited ethanol engine with DME enrichment. Energy Conversion and Management 2012; 58: 19-25. [17] Heywood JB. Internal combustion engine fundamentals. New York: McGraw-Hill, 1988.

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[18] Zhuang Y, Hong G. The impact of Direct Injection Timing and Pressure on Engine Performance in an Ethanol Direct Injection Plus Gasoline Port Injection (EDI+GPI) SI Engine. SAE Tech. Pap. Ser. 2013, Paper No. 2013-01-0892. [19] Adomeit P, Lang O, Pischinger S, Aymanns R. Analysis of Cyclic Fluctuations of Charge Motion and Mixture Formation in a DISI Engine in Stratified Operation. SAE Tech. Pap. Ser. 2007,Paper No. 2007-01-1412. [20] Oh H, Bae C, Min K. Spray and combustion characteristics of ethanol blended gasoline in a spray guided DISI engine under lean stratified operation. SAE International Journal of Engines 2010; 3(2): 213-222. [21] Hanabusa H, Kondo T, Hashimoto K, Sono H. Study on Homogeneous Lean Charge Spark Ignition Combustion. SAE Tech. Pap. Ser. 2013,Paper No. 2013-01-2562. [22] Zeldovich YB. Regime classification of an exothermic reaction with nonuniform initial circumstances. Combustion and Flame 1980;39 (2): 211-214. [23] Shayler P, Jones S, Horn G, Eade D. Characterisation of DISI Emissions and Fuel Economy in Homogeneous and Stratified Charge Modes of Operation. SAE Tech. Pap. Ser. 2001,Paper No. 2001-01-3671. [24] Shiga S. A study of the combustion and emission characteristics of compressed-natural-gas direct-injection stratified combustion using a rapid-compression-machine. Combustion and Flame 2002; 129 (1): 1-10. [25] Oh H, Bae C, Min K. Spray and Combustion Characteristics of Ethanol Blended Gasoline in a Spray Guided DISI Engine under Lean Stratified Operation. SAE Int. J. Engines 2010; 3(2): 213-222.

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[26] Hemdal S, Denbratt I, Dahlander P, Warnberg J. Stratified Cold Start Sprays of Gasoline-Ethanol Blends. SAE Int. J. Fuels Lubr. 2009; 2(1): 683-696. [27] Daniel R, Wang C, Xu H, Tian, G. impacts of Combustion Phasing, Injection Timing, Relative Air-Fuel Ratio and Variable Valve Timing on SI Engine Performance and Emissions using 2,5-Dimethylfuran. SAE Int. J. Fuels Lubr. 2012; 5(2): 855-866. [28] Chan WW, Chen RH, Pu JU, Lin TH. The influence of air–fuelratio on engineperformance and pollutantemission of an SIengine using ethanol–gasoline-blended fuels. Atmospheric Environment 2004; 38(40): 7093–7100. [29] Zhuang Y, Wang J, Hong G. Development of a research engine for investigating ethanol fuel direct injection plus gasoline fuel port injection (EDI+GPI). Proceedings of the 18th Australasian Fluid Mechanics Conference, Launceston, Australia, 3-7 December 2012. [30] Albert TJ, Karl H. A method of lean air-fuel ratio control using combustion pressure measurement. JSAE review 2001; 22(4): 389-393. [31] Wiesinghe JS, Hong G. Impact of Spark Assistance on Auto-ignition Combustion in a Small Two-Stroke Engine. Proceedings of the Institution of Mechanical Engineers. Part D: Journal of Automobile Engineering 2011; 1 (225): 115-126. [32] Huang Y, Hong G, Huang R. Effect of injection timing on mixture formation and combustion in an ethanol direct injection plus gasoline port injection (EDI+GPI) engine[J]. Energy, 2016, 111:92-103. [33] Ayala F, Gerty M, Heywood J. impacts of Combustion Phasing, Relative Air-fuel Ratio, Compression Ratio, and Load on SI Engine Efficiency. SAE Tech. Pap. Ser. 2006,Paper No. 2006-01-0229. [34] Zhuang Y, Hong G. Preliminary investigation to combustion in a SI engine with direct ethanol injection and port gasoline injection (EDI+GPI). Proceedings of the 18th Australasian Fluid Mechanics Conference, Launceston, Australia, 3-7 December 2012. 32


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[35] Daniel R, Wang C, Xu H, Tian G. Dual-Injection as a Knock Mitigation Strategy Using Pure Ethanol and Methanol. SAE Int. J. Fuels Lubr. 2012; 5(2): 772-784. [36] Szybist J, Foster M, Moore W, Confer K. Investigation of Knock Limited Compression Ratio of Ethanol Gasoline Blends. SAE Tech. Pap. Ser. 2010,Paper No. 2010-01-0619. [37] Kiyota Y, Akishino K, Ando H. Concept of LeanCombustion by Barrel-Stratification. SAE Tech. Pap. Ser. 1992,Paper No. 920678. [38] Hardalupas Y, Taylor A, Whitelaw J, Ishii K. Influence of Injection Timing on In-Cylinder FuelDistribution in a Honda VTEC-E Engine. SAE Tech. Pap. Ser. 1995,Paper No. 950507. [39] Zhuang Y, Hong G. Effects of direct injection timing of ethanol fuel on engine knock and lean burn in a port injection gasoline engine. Fuel 2014; 135(11):27-37. [40] Topinka JA, Gerty MD, Heywood JB. Knock Behavior of a Lean-Burn, H~2 and CO Enhanced, SI Gasoline Engine Concept. SAE SP 2004: 39-52. [41] Yang J, Anderson R. Fuel Injection Strategies to Increase Full-Load Torque Output of a Direct-Injection SI Engine. SAE Tech. Pap. Ser. 1998,Paper No. 980495.

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Table captions Table 1-Engine specifications Table 2- Sensors and their measurement uncertainties Table 3-Test fuel properties Table4-Test matrix 34


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Table 1-Engine specifications Engine type Displacement Bore × stroke Compression ratio

Single cylinder, air cooled, 4-stroke, SOHC. 249.0 cm3 74.0 mm x 58.0 mm 9.80:1 35


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Opening: 45° BTDC Intake valve timing Closing: 60° ABDC Opening: 87° ATDC Exhaust valve timing Closing: 21° ATDC

Table 2- Sensors and their measurement uncertainties Associated measurement Device uncertainty Engine speed 0.3 % of full scale Torque 0.03 of the output voltage Air volumetric flow 4.0 % of the measurement rate Gasoline mass flow 1.8% of the measurement 36


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rate Ethanol mass flow rate Bosch Lambda sensor Cylinder pressure HC concentration NOx concentration K type thermocouples

2.1% of the measurement 1.7% of the measurement 0.5% of full scale 3.2% of the measurement 3.8% of the measurement 0.75% of the measurement

Table 3-Test fuel specifications Ethanol Gasoline Chemical formula C2H6O C8.67H16.01 H/C ratio 3 1.795 O/C ratio 0.5 0 Gravimetric oxygen 34.78 0 content (%) 790.9 744.6 Density@20℃(Kg/m3) Research Octane count 106 95 Stoichiometric air/fuel 9.0:1 14.79:1 ratio 37


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LHV(MJ/kg) LHV(MJ/L) Enthalpy of vaporization (kJ/kg) Temperature at initial boiling point(℃)

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26.9 21.3

42.9 31.9

840

373

78.4

32.8

Table4-Test matrix

Tests circumstances EER (%) rate (rpm) IMEP (Bar)

Engine speed

Engine load

EER

Spark timing

SOI timing

24%, 48% 3500-5000

24%, 48% 4000 4.5-8.5 (1 Bar interval)

0-60% 4000

24%, 48% 4000

24%, 48% 4000

5.5-7.5

5.5-8.5

5.5-7.5

5.5-7.5

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Spark timing (CAD BTDC) SOI timing (CAD BTDC)

25

30-18

25

15-45

25

300

300

300

300

50-110, 290-330

Figure captions Figure 1– Schematic of the engine system Figure 2– Relative position of injector and spark plug in cylinder head Figure 3– Fluctuation of lean burn limit with engine speed at medium load Figure 4–Fluctuation of COVIMEP with engine speed at lean burn limit and medium load Figure 5– Fluctuation of CA0-10% with engine speed at lean burn limit and medium load Figure 6– Fluctuation of ISHC with engine speed at lean burn limit and medium load Figure 7– Fluctuation of lean burn limit with load at 4000rpm Figure 8– Fluctuation of COVIMEP with load at lean burn limit and 4000rpm Figure 9– Fluctuation of ISNO with load at lean burn limit and 4000rpm Figure 10– Fluctuation of ISHC with load at lean burn limit and 4000rpm Figure 11– Fluctuation of lean burn limit with EER at 3500rpm and 4000rpm Figure 12– Fluctuation of COVIMEP with EER at lean burn limit at 3500rpm and 4000rpm Figure 13– Fluctuation of indicated thermal efficiency with EER at lean burn limit at 3500rpm and 4000rpm Figure 14– Fluctuation of ISNO with EER at lean burn limit at 3500rpm and 4000rpm Figure 15– Fluctuation of ISHC with EER at lean burn limit at 3500rpm and 4000rpm Figure 16– Fluctuation of lean burn limit with SOI timing at 4000rpm Figure 17– Fluctuation of indicated thermal efficiency with SOI timing at lean burn limit and 4000rpm Figure 18– Fluctuation of COVIMEP with SOI timing at lean burn limit and 4000rpm 39


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Figure 19– Fluctuation of ISNO with SOI timing at lean burn limit and 4000rpm Figure 20– Fluctuation of ISHC with SOI timing at lean burn limit and 4000rpm Figure 21– Fluctuation of lean burn limit with spark timing at 4000rpm Figure 22– Fluctuation of IMEP with spark timing at lean burn limit and 4000rpm Figure 23– The highest indicated thermal efficiency Figure 24– Fluctuation of COVIMEP with spark timing at lean burn limit and 4000rpm Figure 25– Fluctuation of ISNO with spark timing at lean burn limit and 4000rpm Figure 26– Fluctuation of ISHC with spark timing at lean burn limit and 4000rpm

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Figure 1– Schematic of the engine system

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Figure 2– Relative location of injector and spark plug in cylinder head

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Figure 3–Fluctuation of lean burn limit with engine speed at medium load

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Figure 4–Fluctuation of COVIMEP with engine speed at lean burn limit and medium load

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Figure 5– Fluctuation of CA0-10% with engine speed at lean burn limit and medium load

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Figure 6 – Fluctuation of ISHC with engine speed at lean burn limit and medium load

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Figure 7– Fluctuation of lean burn limit with load at 4000rpm

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Figure 8– Fluctuation of COVIMEP with load at lean burn limit and 4000rpm

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Figure 9– Fluctuation of ISNO with load at lean burn limit and 4000rpm

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Figure 10– Fluctuation of ISHC with load at lean burn limit and 4000rpm

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Figure 11– Fluctuation of lean burn limit with EER at 3500rpm and 4000rpm

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Figure 12– Fluctuation of COVIMEP with EER at lean burn limit at 3500rpm and 4000rpm

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Figure 13– Fluctuation of indicated thermal efficiency with EER at lean burn limit at 3500rpm and 4000rpm

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Figure 14– Fluctuation of ISNO with EER at lean burn limit at 3500rpm and 4000rpm

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Figure 15– Fluctuation of ISHC with EER at lean burn limit at 3500rpm and 4000rpm

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Figure 16– Fluctuation of lean burn limit with SOI timing at 4000rpm

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Figure 17– Fluctuation of indicated thermal efficiency with SOI timing at lean burn limit and 4000rpm

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Figure 18– Fluctuation of COVIMEP with SOI timing at lean burn limit and 4000rpm

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Figure 19– Fluctuation of ISNO with SOI timing at lean burn limit and 4000rpm

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Figure 20– Fluctuation of ISHC with SOI timing at lean burn limit and 4000rpm

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Figure 21– Fluctuation of lean burn limit with spark timing at 4000rpm

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Figure 22– Fluctuation of IMEP with spark timing at lean burn limit and 4000rpm

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Figure 23– The highest indicated thermal efficiency

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Figure 24– Fluctuation of COVIMEP with spark timing at lean burn limit and 4000rpm

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Figure 25– Fluctuation of ISNO with spark timing at lean burn limit and 4000rpm

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Figure 26– Fluctuation of ISHC with spark timing at lean burn limit and 4000rpm

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Lean Burn Performance of a Spark Ignition Engine with an Ethanol

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