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Effects of diluents on cycle-by-cycle variations in a spark ignition engine fueled with methanol
Energy 182 (2019) 1132e1140
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Effects of diluents on cycle-by-cycle variations in a spark ignition engine fueled with methanol Miaomiao Zhang a, b, Wei Hong a, b, Fangxi Xie a, b, Yu Liu a, b, *, Yan Su a, b, Xiaoping Li a, b, Haifeng Liu c, Kangning Fang d, Xinbo Zhu d a
State Key Laboratory of Automotive Simulation and Control, Jilin University, Changchun, 130025, People’s Republic of China College of Automotive Engineering, Jilin University, Changchun, 130025, People’s Republic of China State Key Laboratory of Engines, Tianjin University, Tianjin, 300072, People’s Republic of China d Xuzhou XCMG Mining Machinery Co., LTD, Xuzhou, 221000, People’s Republic of China b c
a r t i c l e i n f o
a b s t r a c t
Article history: Received 12 April 2018 Received in revised form 23 May 2019 Accepted 16 June 2019 Available online 19 June 2019
Effects of three diluents on cycle-by-cycle variations of a spark ignition engine fueled with methanol were studied in this paper. Peak in-cylinder pressure, maximum rate of pressure rise and indicated mean effective pressure were investigated experimentally to assess these influences. The results showed that when addition ratios of carbon dioxide and nitrogen increased, the fluctuations of peak pressure and maximum rate of pressure rise increased first and then decrease, the dispersion of indicated mean effective pressure was enlarged, the mean values of these parameters all decreased gradually while carbon dioxide dropped faster. These three parameters did not present clear changes in fluctuation for Argon, only that their mean values showed a slight difference. Linearities between peak pressure and maximum rate of pressure rise corresponding to their crank angles were observed at low addition ratio of carbon dioxide and nitrogen, while it would disappear if the mixture was further diluted. Cycle-by-cycle variations of tested engine would be deteriorated by carbon dioxide and nitrogen, while it kept unchanged and indicated mean effective pressure was increased for argon dilution. The diluent tolerance of this methanol engine is higher than that of gasoline engine. © 2019 Elsevier Ltd. All rights reserved.
Keywords: Cycle-by-cycle variations Spark ignition Methanol Diluent
1. Introduction Because of the scarcity of conventional fossil fuels and stringent emissions demand, alternative fuels are increasingly becoming the study focus in energy field. Among numerous alternative fuels, biofuel is considered as the most favorable one for engines because of its biodegradability, non-toxicity and easy storage and transportation. Moreover, the most important advantage of its usage is attributed to the lower exhaust emissions (carbon monoxide, particular matter and polycyclic aromatic hydrocarbon) when burning in engines [1]. Coal-derived methanol is one of the most commonly used alternative fuels in China for pursuing energy independence. Therefore, methanol engine is becoming a popular power unit. Efforts to enhance its thermal efficiency and to reduce emissions have been received more and more attention [2].
* Corresponding author. State Key Laboratory of Automotive Simulation and Control, Jilin University, Changchun, 130025, People's Republic of China. E-mail address: [email protected] (Y. Liu). https://doi.org/10.1016/j.energy.2019.06.110 0360-5442/© 2019 Elsevier Ltd. All rights reserved.
For spark ignition engines, downsizing is considered as promising route for fuel economy improvement [3]. Within this trend, increasing compression ratio [4] and dilution combustion [5] are two main methods that widely adopted to improve performance. While the former will be limited due to the propensity of knock. Methanol owns low cetane number and high-octane number that can combust at higher compression ratio without knock [6]. In addition, the latent heat evaporation of methanol is 3e5 times higher than that of gasoline, which can improve the volumetric efficiency [7], and the diluent tolerance can be enhanced due to its faster combustion speed and higher laminar flame speed [8]. Dilution combustion has been widely used in gasoline [9] or diesel engine [10], its application in methanol engines has also been a hotspot of research in recent years. Exhaust gas dilution (exhaust gas recirculation) [11] and air dilution (lean burn) [12] are the two main forms of dilution combustion. They are both able to reduce NOx emissions significantly [13] and increase engine efficiency [14]. In Xie's study [15], when the dilution coefficient was about 1.4e1.5 for both cooled EGR and air dilution, the spark-ignition methanol
M. Zhang et al. / Energy 182 (2019) 1132e1140
engine outputted the highest torque and emitted relatively lower NOx emissions. Exhaust gas and air are composed of some individual components, like nitrogen (N2), oxygen (O2), carbon dioxide (CO2), argon (Ar), water vapors (H2O), etc. Argon is the third most common gas in air, and it can be easily obtained as a byproduct in the production of liquid N2 and O2. The above-mentioned influences of exhaust gas and air on combustion are the combined effects of these individual components. Ladommatos et al. studied the influences of EGR components (O2 [16], CO2 [17], and H2O [18]) on the emissions of a diesel engine. They pointed out that the diluent effect of EGR addition was the leading contributor for NOx reduction and particle increase [19]. Li and coworkers [20] conducted experiments on a natural gas spark-ignition engine to study the diluent effect of EGR components. They showed that the contribution of diluent effect on NOx reduction was 50e60% and 41e53% for N2 and CO2 respectively, and the contribution to thermal effect was 40e50% and 41e53% respectively. The influence of diluents on the performance of gasoline direct injection (GDI) engine was examined by Zhang [21]. The results indicated that under the same addition ratio, Ar dilution gave the best fuel economy, while CO2 dilution could more effectively suppress NOx, HC and particulate matter emissions. These studies provide better understanding of dilution combustion on diesel, gasoline and nature gas engines, whereas relevant researches in methanol engine were not mentioned in previous works. Thus, something must be done to fill this gap. An unavoidable phenomenon that exists in combustion of spark ignition engines is cycle-by-cycle variations, which is a big concern that affects engine performance. It was reported that the elimination of cyclic variability would give rise to 10% increase in the power output [22]. Therefore, it is important to deeply study cyclic variations and to minimize it in terms of emissions and efficiency. Cyclic variations in the combustion process are caused by many factors, such as mixture motion, air/fuel ratio, mixture formation and diluent gases. Several investigations have been carried out to study cycle-by-cycle variations in spark-ignition engines. Galloni [23] investigated the combustion stability through experimental tests and numerical analyses in a spark-ignition engine, the results showed that the calculated mean values of both the laminar flame speed and the turbulence intensity seemed correlated with the measured COVIMEP. Gong et al. [24] revealed that factors, like injection and ignition timing, engine speed and load, compression ratio and injector configuration, significantly affect the cycle-bycycle variation of a direct-injection spark ignition (DISI) engine fueled with methanol. Researches [25] also involved the dynamics of cyclic variations and demonstrated that cyclic combustion variability in spark-ignited engine exhibits noisy nonlinear dynamical process [26]. From the literature, dilution combustion in gasoline/diesel/nature gas engines have been well investigated, especially the effects of some diluent components on combustion. Methanol is considered as a long-term renewable fuel for internal combustion engine. Compared with traditional fossil fuels, methanol has totally dissimilar physiochemical properties. This makes a significant influence on dilution combustion of methanol engine. Previous studies revealed that exhaust gas dilution [27] and air dilution [15] will improve emissions and efficiency performance of methanol engine at the proper diluent condition. At the same time, they also deteriorate combustion process due to the prolonged ignition delay and combustion duration. Therefore, it is necessary to comprehensively investigate (using single diluent components) characteristics of dilution combustion in methanol engine, and to weigh its pros and cons for methanol engine applications. The present work fills this gap by using the entry point of cycleby-cycle variations in combustion and by investigating how diluent
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gases (CO2, N2, Ar) affect combustion process of a high compression ratio spark-ignition methanol engine. This work presented a comprehensive understanding of dilution combustion on methanol engine to readers, and will also provide foundation for optimal control of diluent components in dilution combustion. 2. Experimental setup and methods 2.1. Engine and experimental apparatus In this study, a four-stroke, naturally aspirated, water-cooled, spark-ignition methanol engine was employed, which was modified from a direct injection diesel engine. The spark plug was installed in the place of diesel injector and methanol injector was installed in the intake manifold after the throttle. Methanol was supplied by a methanol pump from tank to a methanol rail, and then injected into the intake manifold. The detailed engine specifications are given in Table 1. More information of this engine can be found in [15]. The tested fuel in this paper was refined industrial methanol with a purity of 99%. The physical and chemical features of the methanol fuel can be seen in [28]. Fig. 1 presents the schematic of the methanol engine experimental setup. The engine loads and speed were controlled by an eddy current dynamometer with model of Nan-Feng 160. The combustion pressure in the first cylinder was measured by a Kistler quartz crystal pressure sensor (6125B) with sensitivity of 21.23 pC/ bar, and an AVL charge amplifier was employed to process and amplify pressure signal. The crankshaft angle was acquired using an Kistler encoder (2614B) with resolution of 1. The pressure and crankshaft angle signals were transmitted to an ONO SOKKI DS9110 combustion pressure, thus combustion parameters, such as in-cylinder pressure, heat release rate and in-cylinder temperature, were calculated. The intake air and fuel flow were measured by an air flow meter of SENSYCON Sensy flow P and a fuel flow meter of ONO SOKKI DF-2420, respectively. An independently developed control system based on Intel 80C196 chip was used as the electronic control unit (ECU), which could flexibly control ignition and injection operations. In this research, the three diluent gases (CO2, N2 and Ar) were supplied by compressed gas bottles with purity level of 99%. Their properties at 25 C and 101 kPa are shown in Table 2. The gas from bottles was maintained at 0.8 MPa by a pressure controller before fed into the engine intake manifold. A mass flow controller (MFC) was used to adjust the mass flow of diluents, and their mass flow rate was detected by a Toceil CMF025 dynamical gas mass flow meter with an accuracy of ±0.35%. The ambient temperature of experiments was about 25 C. Meanwhile, because of the pressure drop and kinetic energy rise, according to law of thermodynamics, the gas out of bottle was cooled during the process. Therefore, the temperature of the intake diluent was controlled at about 25 C by heater. Table 3 shows the measurement uncertainty for devices used in this work. Table 1 Experimental engine specification. Engine parameters
Specifications
Engine type Combustion chamber type Bore Stroke Displacement Compression ratio Connecting rod length Intake valve closing Exhaust valve opening
in-line, 4-stroke, 4-cylinder u shape 85 mm 88 mm 1.997L 18.0 175 mm 147 CA BTDC 136 CA ATDC
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Fig. 1. Schematic diagram of the experimental apparatus.
Table 2 Properties of Ar, N2 and CO2 at 25 C and 101 kPa. Gas
Air
Ar
N
Molar mass (kg$kmol1) Gas constant (kJ$kg1$K1) Density (kg$m3) Specific heat capacity (kJ$kg1$K1) Specific heat ratio
28.97 0.2871 1.293 1.004 1.4
39.94 0.2081 1.784 0.524 1.667
28.016 0.2968 1.025 1.038 1.4
2
CO
2
44.01 0.1889 1.977 0.85 1.285
Table 3 Measurement uncertainty analysis.
lambda meter and BOSCH LSU 4.9 lambda sensor at all cases. Engine coolant was controlled at 85 ± 3 C by a closed loop temperature control system. All instruments had been calibrated before they were used. In this test, 200 consecutive combustion cycles were acquired to evaluate cycle-by-cycle variation. In order to intuitively compare and analyze the dilution effects of different gases at the same addition level, addition ratio (AR) was introduced in this paper. It was defined as the ratio of gas charge to the air charge. The formula is as follows:
0
Equipment
Accuracy
Dynamometer Pressure sensor Charge amplifier Crank angle encoder Air flow meter Fuel flow meter
Torque: ±2Nm Speed: ±1 rpm ±0.3% ±0.6% ±0.5% ±0.5% ±0.2%
2.2. Experimental methodology For all cases, the engine speed was maintained at 1400 rpm. By adjusting methanol quantity, the output torque was kept at 43Nm at stoichiometry mixture combustion condition without dilution, which reflects 25% load of this engine at 1400 rpm. The spark advance was 22 CA BTDC (before top dead center) which was the maximum brake torque (MBT) spark timing for the original engine at this operating condition. It is well known that the start of ignition affects the start of combustion [29], which then affects combustion variation [30]. Although the constant spark advance was not realistic for diluent conditions, it would highlight the combustion differences that caused by different physicochemical properties and diluent degree of added gases. This is what this paper wanted to focus. Furthermore, the excess air coefficient was kept at 1 by ETAS
@ ARgas
1 mgas ¼ 100# A mair
(1)
where mair is the practical air mass flow for stoichiometric combustion, and it can be measured by the air flow meter; mgas is the mass flow rates (kg/h) of dilution gas, and it was directly acquired by the gas mass flow meter. The coefficient of variation in indicated mean effective pressure (COVIMEP) is defined as the standard deviation in IMEP divided by the mean IMEP, and is usually expressed in percent [31]:
COVIMEP ¼
sIMEP IMEP
100
(2)
where
IMEP ¼
sIMEP
N 1 X IMEP N i¼1
sffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi PN i¼1 IMEP IMEP ¼ N
(3)
(4)
N is the number of combustion cycle, IMEP is indicated mean effective pressure of individual cycle.
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3. Results and discussions Fig. 2 presents the cycle-by-cycle variations of peak in-cylinder pressure at different addition ratios for the tested three dilutions. For CO2 and N2, the mean peak in-cylinder pressure value was decreased, the variational fluctuation enlarged first and then shrunk with the increase of addition ratio (from 0 to 26.2%). While Ar didn't show clear change in both mean value and variation. Noting that the different values of the mean peak in-cylinder pressure for the three diluents at 0% addition ratio was due to the dispersion in engine behavior. The information of pmax (peak incylinder pressure) reflects the combustion completeness and the amount of heat transfer to the wall [32]. Due to the constant air quantity, the addition of diluent increased in-cylinder mixture heat capacity and decreased O2 concentration, the combustion temperature and flame propagation speed were reduced. Thus, pmax presented a decreasing trend as the addition ratio was increased. Meanwhile, for a small addition ratio, the suppressed chemical reactions also brought about the instability of combustion process, which caused the enlarged fluctuation of peak in-cylinder pressure. While for large addition ratio, pmax was decreased sharply near to the limit of compression terminal pressure, that is why the fluctuation of peak pressure showed a relative compactness distribution. The results also revealed that the decline level of mean incylinder pressure was shrunken successively when adding the same diluent inflow according to the order of CO2, N2 and Ar. Compared with CO2 and N2, Ar-air mixture owns the lowest heat capacity [33] which results in the highest compression temperature. The initial flame kernel formation and development will be improved. Besides, the high thermal diffusivity of Ar [18] will also enhance the chemical reactions and promote combustion process. Similarly, the mixture heat capacity of N2 dilution is lower than that of CO2 dilution, that is why the descent degree of mean in-cylinder pressure of N2 is smaller than that of CO2. Fig. 3 shows the cycle-by-cycle variations of (dp/d4)max (the maximum rate of in-cylinder pressure rise) at different addition ratios for the tested three diluents. The rate of pressure rise reflects burning velocity of mixture and reveals whether the combustion process is timely or not. For all dilutions, with increasing the addition ratio of diluents, the mean value of (dp/d4)max was decreased. This was mainly because that the addition of diluents slowed the burning velocity and prolonged combustion duration, thus the heat release rate was decreased. It was also observed that the fluctuation range of (dp/d4)max didn't show any changes, although it tended to be more centered on the mean maximum pressure rise at large addition ratio (18.6% and 26.2%) of CO2 and N2 dilutions. The reason is similar with the change of peak in-cylinder pressure, and we no longer explain it. The slight decreasing trend in the mean (dp/d4)max of Ar dilution could be attributed to the especial thermophysical parameters of Ar and will be further discussed in the following section. Another result can be seen was that diluent combustion has more impact on the maximum rate of pressure rise than on peak in-cylinder pressure at Ar dilution. Fig. 4 gives the relationship of pmax and 4pmax (the corresponding crank angle at which pmax occurs) when the diluent was added. The magnitudes of the changes in pmax and 4pmax depend on the extent of the cycle-by-cycle variation, and also depend on whether the average burn process is fast or slow. Empirically, the location of the appearance of peak in-cylinder pressure plays an important role in evaluating combustion process, and it is expected to be around 16 CA ATDC at which the best engine dynamic performance occurs [31]. In order to provide a better and deep understanding about this cyclic variation, we chose 5 typical cycles named case 1 to 5 and have pointed out their positions in Fig. 4. The standards of selection are as follows. Case 1, 2, 4 and 5 are cycles
Fig. 2. Cycle-by-cycle variations of peak in-cylinder pressure at different addition ratios; (a) CO2 dilution, (b) N2 dilution, (c) Ar dilution.
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Fig. 3. Cycle-by-cycle variations of the maximum rate of pressure rise at different addition ratios; (a) CO2 dilution, (b) N2 dilution, (c) Ar dilution.
Fig. 4. Peak in-cylinder pressure versus crank angle at which it occurs; (a) CO2 dilution, (b) N2 dilution, (c) Ar dilution.
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with maximum pmax of each diluent condition, case 3 is the peak point with 4pmax equals 0 CA at the same diluent condition as case 4. For example, ‘CO2-26.2%-#3’ means the third cycle with the addition ratio of 26.2% at CO2 dilution. The in-cylinder pressure and heat release rate of cases are given in Fig. 5. A linearity between pmax and 4pmax was observed at each condition (except high addition ratio of CO2 and N2 dilution), revealing that high pmax appeared at small 4pmax. This phenomenon was mainly due to the methanol-charging variations, cycle-by-cycle. The results also showed that for CO2 and N2, the value of peak incylinder pressure decreased and its corresponding crank angle moved away from top dead center when increasing addition ratio. As discussed above, the addition of N2 and CO2 slowed the flame propagation and burning rate of mixture, the combustion process was retarded from the optimum when the spark timing was fixed, thus pmax decreased and 4pmax increased. That is, fast burning cycles corresponded to advanced combustion phasing. When further increasing diluents (AR18.6%), the burn process was excessively retarded, the rate of increase of pressure due to combustion became so slow that it was more than offset by the pressure decrease due to volume increase, thus the first peak (around the compression top dead center) of pressure would exceed the second peak (Fig. 5, case 3). Even worse, misfiring cycles started to occur, and partial and non-burning cycles increased rapidly if the mixture was made even more dilute. Faster burning will bring about less cyclic combustion variations on engine performance [34], while for slow burning, where a large fraction of methanol combusted well after top dead center, the effect of volume change became significant and strengthened the combustion variations. For Ar dilution, the distribution of pmax was centered and the value of pmax increased slightly as addition ratio was increased. Although the dilution of Ar would slower flame propagation, the improved initial flame kernel formation and development on account of the low mixture heat capacity offset this disadvantage and made a less variational combustion phasing. Moreover, the promoted chemical reaction and combustion due to the high thermal diffusivity of Ar made pmax increased slightly with increasing amount of diluent gas (Fig. 5, from case 1 to case 2). At addition ratio of 26.2%, the proportions of cycles with 4pmax located away from top dead center in the total cycles for CO2, N2, Ar were 16.5%, 78%, 100%, respectively. It means that at this extend of dilution, CO2 owned the largest number of cycles with slow burning and long combustion phasing, followed by N2 and then Ar. This can
Fig. 5. In-cylinder pressure and heat release rate of the selected cases.
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be attributed to the highest mixture heat capacity of N2 dilution and the highest thermal diffusivity of Ar dilution. In addition, the maximum pmax of Ar (Fig. 5, case 2) dilution was the highest, followed by N2 (Fig. 5, case 5) and then CO2 (Fig. 5, case 4). Fig. 6 presents the relationship of (dp/d4)max and 4(dp/d4)max (the crank angle at which (dp/d4)max occurs) when the diluent was added. The results disclosed that the interdependency of (dp/ d4)max and 4(dp/d4)max was approximately similar with that of pmax and 4pmax. At each condition (except high addition ratio of CO2 and N2), the data showed a relatively vertical change and was centered around 10 CA. The change was mainly because of the variations of the amount of methanol injecting into cylinder each cycle. For CO2 and N2, the (dp/d4)max was decreased and 4(dp/d4)max moved away from top dead center when adding small amount of dilutions (AR<11.7%). This resulted from the slowed flame propagation and lengthening of combustion duration. With increasing addition ratio, as burning lengthened further, slow burn or partial burn cycles were encountered and the maximum rate of pressure rise occurred at compression stroke. As a consequence, the engine became rough and unstable. A modest decrease of (dp/d4)max was also observed as the mixture was diluted by Ar and the distribution of data was relatively stabilized. As discussed above, Ar dilution would increase in-cylinder temperature at compression stroke, hence the incylinder pressure was relatively high (Fig. 5, case 1 and 2) when mixture burned at this period. Thus, the pressure line showed a less steep rise style. The impact on IMEP (indicated mean effective pressure), due to increased addition ratio of diluents, is shown in Fig. 7. As increasing addition ratio from 0 to 26.2%, CO2 and N2 dilutions showed the same changing trend: the descending mean value of IMEP and the gradually enlarged fluctuation of IMEP data, only that the percentage decline of mean IMEP for N2 dilution was lower than that of CO2 dilution (4.6% for N2, 14.6% for CO2). Increasing addition ratio of mixture slowed burning velocity and lengthened combustion phasing, thus the distribution of IMEP data was widened and IMEP was lowered either due to partial burning. While for Ar dilution, the mean value of IMEP showed an inverse increasing trend (increased by 3.6%) and the dispersion of IMEP data didn't presented clear change. This phenomenon resulted mainly from the positive impact of Ar on mixture, given that Ar has a comparatively lower specific heat capacity and higher thermal diffusivity. The method of return maps has widely used to examine the interrelationships between consecutive cycles of combustion process [35]. Note that we only use it to evaluate cyclic variation here. Fig. 8 depicts the return maps of indicated mean effective pressure when increasing addition ratio of diluents, IMEP(i) (IMEP at the current cycle) versus IMEP(iþ1) (IMEP at the next consecutive cycle). It can be easily found from the figure that no misfiring cycles (IMEP <0) had happened during the diluent process for all diluent gases. For specific condition, centralized IMEP data means that the combustion process is stable while scattered IMEP data means that the combustion process is unstable. In detail, data point which locates away from the center stands for low value of IMEP due to partial burning or high value of IMEP caused by the variational amount of methanol injected. For Ar dilution, the IMEP data was always centered in a circular region, only that the cloud of points moved to the upper right area gradually when the addition ratio was increased. While for CO2 and N2 dilutions, as addition ratio increased, the cloud of points moved to lower left area and the cluster of IMEP data expanded, which definitely resulted in high cycle-by-cycle variations. Differences in CO2 and N2 dilutions are also visible, compared with the operating condition with no diluent, the data distribution of CO2 was more scattered and further away than that of N2 at the same addition ratio. COVIMEP (the coefficient of variation in indicated mean effective
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Fig. 6. Maximum rate of in-cylinder pressure rise versus crank angle at which it occurs; (a) CO2 dilution, (b) N2 dilution, (c) Ar dilution.
Fig. 7. Cycle-by-cycle variations of the indicated mean effective pressure at different addition ratios; (a) CO2 dilution, (b) N2 dilution, (c) Ar dilution.
M. Zhang et al. / Energy 182 (2019) 1132e1140
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Fig. 9. Changes of the coefficient of variation in indicated mean effective pressure with the addition ratio of diluents.
pressure) is one important index to measure cyclic variability in indicated work per cycle. The impact of addition ratio on COVIMEP is given in Fig. 9. It can be found that COVIMEP of CO2 and N2 dilutions increased up to 6.2% and 5.8% respectively when elevating addition ratio to 34.8%, while COVIMEP of Ar almost hold steady. Under the same high addition ratio, the COVIMEP of CO2 is the highest followed by N2 and then Ar. Comprehensive analysis of Figs. 7e9 can draw a conclusion that Ar dilution will enhance IMEP of methanol engine without deteriorating engine cyclic variation. In our previous study [21], a 1.39l four-cylinder GDI engine was employed to study the effects of diluents on combustion and emission characteristics. During that experiment campaign, the same three gases were used. Detailed engine and operation information could be found in the mentioned work. A parallel comparison reveals that at this small-speed and small-load condition, the corresponding addition ratio value of limit COVIMEP (more than 5%) is higher in methanol engine (about 30% at CO2) than that in GDI engine (about 20% at CO2). This suggests that the diluent tolerance of methanol engine is higher than that of GDI engine. The reason can be explained by the faster flame propagation of methanol charge than gasoline charge. 4. Conclusions An experimental investigation about the dilution combustion was conducted on a spark ignition engine fueled with methanol. The different effects of diluent gases on cyclic variation were found. The conclusions drawn from this work are listed as follows:
Fig. 8. Return maps of indicated mean effective pressure at different addition ratios; (a) CO2 dilution, (b) N2 dilution, (c) Ar dilution.
(1) With increasing addition ratios of CO2 and N2, the fluctuations of pmax and (dp/d4)max increase first and then decrease due to the limit of compression terminal pressure, the dispersion of IMEP enlarges. While for Ar, the fluctuations of pmax, (dp/d4)max and IMEP changed slightly. (2) Linearities were found between pmax and 4pmax, (dp/d4)max and 4(dp/d4)max at low addition ratio for CO2 and N2, while they will disappear at high addition ratio due to slow burning. (3) The proportion of slow burning cycles of CO2 dilution is higher than that of N2 dilution at high addition ratio.
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(4) Drawn from the IMEP return map and COVIMEP, CO2 and N2 dilution will deteriorate cyclic variation of this methanol engine while Ar dilution will enhance IMEP with slight changes in cyclic variations. (5) Compared with our previous researches, the diluent tolerance of spark ignition methanol engine is higher than that of gasoline engine.
Declaration of interests The authors declare that they have no known competing financial interests or personal relationships that could have appeared to influence the work reported in this paper. The authors declare the following financial interests/personal relationships which may be considered as potential competing interests: Acknowledgements This work was financially supported by National Natural Science Foundation of China (Grant No. 51876079), funded by the special fund for basic scientific research in central colleges and universities, Natural Science Foundation of Jilin Province (20170101137JC, 20170101132JC), Science and technology research projects of Jilin Province education department for 13th Five Year Plan(JJKH20180141KJ), Project of Development and Reform Commission of Jilin Province (2019C058-4), Excellent Youth Talent Fund Project (20190103049JH), Open Fund of state key laboratory of automotive simulation and control, Jilin University (20181110), Graduate Innovation Fund of Jilin University (101832018C199). References [1] Huang YH, Hong G, Huang GH. Investigation to charge cooling effect and combustion characteristics of ethanol direct injection in a gasoline port injection engine. Appl Energy 2015;160:244e54. https://doi.org/10.1016/j. apenergy.2015.09.059. [2] Verhelst S, Turner JW, Sileghem L, Jeroen Vancoillie. Methanol as a fuel for internal combustion engines. Prog Energy Combust Sci 2019;70:43e88. https://doi.org/10.1016/j.pecs.2018.10.001. [3] Fraser N, Blaxill H, Lumsden G, Bassett M. Challenges for increased efficiency through gasoline engine downsizing. SAE Int J Engines 2009;2:991e1008. [4] Su JY, Xu M, Li T, Gao Y, Wang JS. Combined effects of cooled EGR and a higher geometric compression ratio on thermal efficiency improvement of a downsized boosted spark-ignition direct-injection engine. Energy Convers Manag 2014;78:65e73. https://doi.org/10.1016/j.enconman.2013.10.041. [5] Xie FX, Hong W, Su Y, Zhang MM, Jiang BP. Effect of external hot EGR dilution on combustion, performance and particulate emissions of a GDI engine. Energy Convers Manag 2017;142:69e81. https://doi.org/10.1016/j.enconman. 2017.03.045. [6] Brinkman ND. Effect of compression ratio on exhaust emissions and performance of a methanol-fueled single-cylinder engine. SAE paper no. 770791. 1977. [7] Mohanan P, Babu MKG. A simulation model for a methanol-fueled turbocharged multi-cylinder automotive spark ignition engine. SAE paper no. 912417. 1991. [8] Ryan T, Lestz S. The laminar burning velocity of isooctane, N-heptane, methanol, methane and propane at elevated temperatures and pressures in the presence of a diluent. SAE paper no. 800103. 1980. [9] Xie FX, Hong W, Su Y, Zhang MM, Jiang BP. Effect of external hot EGR dilution on combustion, performance and particulate emissions of a GDI engine. Energy Convers Manag 2017;142:69e81. https://doi.org/10.1016/j.enconman. 2017.03.045. tet JF. Experimental study of various effects of [10] Maiboom A, Tauzia X, He exhaust gas recirculation (EGR) on combustion and emissions of an automotive direct injection diesel engine. Energy 2008;33:22e34. [11] Zhang ZJ, Zhang HY, Wang TY, Jia M. Effects of tumble combined with EGR (exhaust gas recirculation) on the combustion and emissions in a spark ignition engine at part loads. Energy 2014;65:18e24. https://doi.org/10.1016/ j.energy.2013.11.062.
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The dilution, chemical, and thermal effects of exhaust gas recirculation on diesel engine emissions e part 1: effect of reducing inlet charge oxygen. SAE paper no. 961165. Society of Automotive Engineers; 1996. [17] Ladommatos N, Abdelhalim S, Zhao H, Hu Z. The dilution, chemical and thermal effects of exhaust gas recirculation on diesel emissions e part 2: effects of carbon dioxide. SAE paper no. 961167. Society of Automotive Engineers; 1996. [18] Ladommatos N, Abdelhalim S, Zhao H, Hu Z. The dilution, chemical, and thermal effects of exhaust gas recirculation on diesel engine emissions e part 3: effects of water vapour. SAE paper no. 971659. Society of Automotive Engineers; 1997. [19] Ladommatos N, Abdelhalim S, Zhao H, Hu Z. The dilution, chemical, and thermal effects of exhaust gas recirculation on diesel engine emissions e part 4: effects of carbon dioxide and water vapour. SAE paper no. 971660. Society of Automotive Engineers; 1997. [20] Li WF, Liu ZC, Wang ZS, Xu Y. Experimental investigation of the thermal and diluent effects of EGR components on combustion and NOx emissions of a turbocharged natural gas SI engine. Energy Convers Manag 2014;88:1041e50. https://doi.org/10.1016/j.enconman.2014.09.051. [21] Zhang MM, Hong W, Xie FX, Su Y, Han LW, Wu B. Influence of diluents on combustion and emission characteristics of a GDI engine. Appl Therm Eng 2017;124:746e55. https://doi.org/10.1016/j.applthermaleng.2017.06.043. [22] Ozdor N, Dulger M, Sher E. Cyclic variability in spark ignition engines a literature survey. SAE paper no. 940987. 1994. [23] Galloni E. Analyses about parameters that affect cyclic variation in a spark ignition engine. Appl Therm Eng 2009;29:1131e7. https://doi:10.1016/j. applthermaleng.2008.06.001. [24] Gong CM, Huang K, Chen YL, Jia JL, Su Y, Liu XJ. Cycle-by-cycle combustion variation in a DISI engine fueled with methanol. Fuel 2011;90:2817e9. https://doi:10.1016/j.fuel.2011.04.010. [25] Green JrJr, Daw CS, Armfield JS, Finney CEA, Wagner RM, Drallmeier JA, Kennel MB, Durbetaki P. Time irreversibility and comparison of cyclicvariability models. SAE paper no.1999-01-0221. 1999. [26] Yang LP, Song EZ, Ding SL, Brown RJ, Marwan N, Ma XZ. Analysis of the dynamic characteristics of combustion instabilities in a pre-mixed lean-burn natural gas engine. Appl Energy 2016;183:746e59. https://doi.org/10.1016/j. apenergy.2016.09.037. [27] Xie FX, Li XP, Wang XC, Su Y, Hong W. Research on using EGR and ignition timing to control load of a spark-ignition engine fueled with methanol. Appl Therm Eng 2013;50:1084e91. https://doi.org/10.1016/j.applthermaleng. 2012.08.003. [28] Gong CM, Huang K, Jia JL, Su Y, Gao Q, Liu XJ. Regulated emissions from a direct-injection spark-ignition methanol engine. Energy 2011;36:3379e87. [29] Ceviz MA, Yüksel F. Effects of ethanol-unleaded gasoline blends on cyclic variability and emissions in an SI engine. Appl Therm Eng 2005;25:917e25. https://doi:10.1016/j.applthermaleng.2004.07.019. [30] Blasio GD, Belgiorno G, Beatrice C. Effects on performances, emissions and particle size distributions of a dual fuel (methane-diesel) light-duty engine varying the compression ratio. Appl Energy 2017;204:726e40. https://doi. org/10.1016/j.apenergy.2017.07.103. [31] Heywood JB. Internal combustion engine fundamentals. New York: McGrawHill; 1988. [32] Wang JH, Chen H, Liu B, Huang ZH. Study of cycle-by-cycle variations of a spark ignition engine fueled with natural gasehydrogen blends. Int J Hydrogen Energy 2008;33:4876e7883. https://doi:10.1016/j.ijhydene.2008. 06.062. [33] Li WF, Liu ZC, Wang ZS, Dou HL. Experimental and theoretical analysis of effects of atomic, diatomic and polyatomic inert gases in air and EGR on mixture properties, combustion, thermal efficiency and NOx emissions of a pilot-ignited NG engine. Energy Convers Manag 2015;105:1082e95. https:// doi.org/10.1016/j.enconman.2015.08.052. 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Effects of diluents on cycle-by-cycle variations in a spark ignition engine fueled with methanol Miaomiao Zhang a, b, Wei Hong a, b, Fangxi Xie a, b, Yu Liu a, b, *, Yan Su a, b, Xiaoping Li a, b, Haifeng Liu c, Kangning Fang d, Xinbo Zhu d a
State Key Laboratory of Automotive Simulation and Control, Jilin University, Changchun, 130025, People’s Republic of China College of Automotive Engineering, Jilin University, Changchun, 130025, People’s Republic of China State Key Laboratory of Engines, Tianjin University, Tianjin, 300072, People’s Republic of China d Xuzhou XCMG Mining Machinery Co., LTD, Xuzhou, 221000, People’s Republic of China b c
a r t i c l e i n f o
a b s t r a c t
Article history: Received 12 April 2018 Received in revised form 23 May 2019 Accepted 16 June 2019 Available online 19 June 2019
Effects of three diluents on cycle-by-cycle variations of a spark ignition engine fueled with methanol were studied in this paper. Peak in-cylinder pressure, maximum rate of pressure rise and indicated mean effective pressure were investigated experimentally to assess these influences. The results showed that when addition ratios of carbon dioxide and nitrogen increased, the fluctuations of peak pressure and maximum rate of pressure rise increased first and then decrease, the dispersion of indicated mean effective pressure was enlarged, the mean values of these parameters all decreased gradually while carbon dioxide dropped faster. These three parameters did not present clear changes in fluctuation for Argon, only that their mean values showed a slight difference. Linearities between peak pressure and maximum rate of pressure rise corresponding to their crank angles were observed at low addition ratio of carbon dioxide and nitrogen, while it would disappear if the mixture was further diluted. Cycle-by-cycle variations of tested engine would be deteriorated by carbon dioxide and nitrogen, while it kept unchanged and indicated mean effective pressure was increased for argon dilution. The diluent tolerance of this methanol engine is higher than that of gasoline engine. © 2019 Elsevier Ltd. All rights reserved.
Keywords: Cycle-by-cycle variations Spark ignition Methanol Diluent
1. Introduction Because of the scarcity of conventional fossil fuels and stringent emissions demand, alternative fuels are increasingly becoming the study focus in energy field. Among numerous alternative fuels, biofuel is considered as the most favorable one for engines because of its biodegradability, non-toxicity and easy storage and transportation. Moreover, the most important advantage of its usage is attributed to the lower exhaust emissions (carbon monoxide, particular matter and polycyclic aromatic hydrocarbon) when burning in engines [1]. Coal-derived methanol is one of the most commonly used alternative fuels in China for pursuing energy independence. Therefore, methanol engine is becoming a popular power unit. Efforts to enhance its thermal efficiency and to reduce emissions have been received more and more attention [2].
* Corresponding author. State Key Laboratory of Automotive Simulation and Control, Jilin University, Changchun, 130025, People's Republic of China. E-mail address: [email protected] (Y. Liu). https://doi.org/10.1016/j.energy.2019.06.110 0360-5442/© 2019 Elsevier Ltd. All rights reserved.
For spark ignition engines, downsizing is considered as promising route for fuel economy improvement [3]. Within this trend, increasing compression ratio [4] and dilution combustion [5] are two main methods that widely adopted to improve performance. While the former will be limited due to the propensity of knock. Methanol owns low cetane number and high-octane number that can combust at higher compression ratio without knock [6]. In addition, the latent heat evaporation of methanol is 3e5 times higher than that of gasoline, which can improve the volumetric efficiency [7], and the diluent tolerance can be enhanced due to its faster combustion speed and higher laminar flame speed [8]. Dilution combustion has been widely used in gasoline [9] or diesel engine [10], its application in methanol engines has also been a hotspot of research in recent years. Exhaust gas dilution (exhaust gas recirculation) [11] and air dilution (lean burn) [12] are the two main forms of dilution combustion. They are both able to reduce NOx emissions significantly [13] and increase engine efficiency [14]. In Xie's study [15], when the dilution coefficient was about 1.4e1.5 for both cooled EGR and air dilution, the spark-ignition methanol
M. Zhang et al. / Energy 182 (2019) 1132e1140
engine outputted the highest torque and emitted relatively lower NOx emissions. Exhaust gas and air are composed of some individual components, like nitrogen (N2), oxygen (O2), carbon dioxide (CO2), argon (Ar), water vapors (H2O), etc. Argon is the third most common gas in air, and it can be easily obtained as a byproduct in the production of liquid N2 and O2. The above-mentioned influences of exhaust gas and air on combustion are the combined effects of these individual components. Ladommatos et al. studied the influences of EGR components (O2 [16], CO2 [17], and H2O [18]) on the emissions of a diesel engine. They pointed out that the diluent effect of EGR addition was the leading contributor for NOx reduction and particle increase [19]. Li and coworkers [20] conducted experiments on a natural gas spark-ignition engine to study the diluent effect of EGR components. They showed that the contribution of diluent effect on NOx reduction was 50e60% and 41e53% for N2 and CO2 respectively, and the contribution to thermal effect was 40e50% and 41e53% respectively. The influence of diluents on the performance of gasoline direct injection (GDI) engine was examined by Zhang [21]. The results indicated that under the same addition ratio, Ar dilution gave the best fuel economy, while CO2 dilution could more effectively suppress NOx, HC and particulate matter emissions. These studies provide better understanding of dilution combustion on diesel, gasoline and nature gas engines, whereas relevant researches in methanol engine were not mentioned in previous works. Thus, something must be done to fill this gap. An unavoidable phenomenon that exists in combustion of spark ignition engines is cycle-by-cycle variations, which is a big concern that affects engine performance. It was reported that the elimination of cyclic variability would give rise to 10% increase in the power output [22]. Therefore, it is important to deeply study cyclic variations and to minimize it in terms of emissions and efficiency. Cyclic variations in the combustion process are caused by many factors, such as mixture motion, air/fuel ratio, mixture formation and diluent gases. Several investigations have been carried out to study cycle-by-cycle variations in spark-ignition engines. Galloni [23] investigated the combustion stability through experimental tests and numerical analyses in a spark-ignition engine, the results showed that the calculated mean values of both the laminar flame speed and the turbulence intensity seemed correlated with the measured COVIMEP. Gong et al. [24] revealed that factors, like injection and ignition timing, engine speed and load, compression ratio and injector configuration, significantly affect the cycle-bycycle variation of a direct-injection spark ignition (DISI) engine fueled with methanol. Researches [25] also involved the dynamics of cyclic variations and demonstrated that cyclic combustion variability in spark-ignited engine exhibits noisy nonlinear dynamical process [26]. From the literature, dilution combustion in gasoline/diesel/nature gas engines have been well investigated, especially the effects of some diluent components on combustion. Methanol is considered as a long-term renewable fuel for internal combustion engine. Compared with traditional fossil fuels, methanol has totally dissimilar physiochemical properties. This makes a significant influence on dilution combustion of methanol engine. Previous studies revealed that exhaust gas dilution [27] and air dilution [15] will improve emissions and efficiency performance of methanol engine at the proper diluent condition. At the same time, they also deteriorate combustion process due to the prolonged ignition delay and combustion duration. Therefore, it is necessary to comprehensively investigate (using single diluent components) characteristics of dilution combustion in methanol engine, and to weigh its pros and cons for methanol engine applications. The present work fills this gap by using the entry point of cycleby-cycle variations in combustion and by investigating how diluent
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gases (CO2, N2, Ar) affect combustion process of a high compression ratio spark-ignition methanol engine. This work presented a comprehensive understanding of dilution combustion on methanol engine to readers, and will also provide foundation for optimal control of diluent components in dilution combustion. 2. Experimental setup and methods 2.1. Engine and experimental apparatus In this study, a four-stroke, naturally aspirated, water-cooled, spark-ignition methanol engine was employed, which was modified from a direct injection diesel engine. The spark plug was installed in the place of diesel injector and methanol injector was installed in the intake manifold after the throttle. Methanol was supplied by a methanol pump from tank to a methanol rail, and then injected into the intake manifold. The detailed engine specifications are given in Table 1. More information of this engine can be found in [15]. The tested fuel in this paper was refined industrial methanol with a purity of 99%. The physical and chemical features of the methanol fuel can be seen in [28]. Fig. 1 presents the schematic of the methanol engine experimental setup. The engine loads and speed were controlled by an eddy current dynamometer with model of Nan-Feng 160. The combustion pressure in the first cylinder was measured by a Kistler quartz crystal pressure sensor (6125B) with sensitivity of 21.23 pC/ bar, and an AVL charge amplifier was employed to process and amplify pressure signal. The crankshaft angle was acquired using an Kistler encoder (2614B) with resolution of 1. The pressure and crankshaft angle signals were transmitted to an ONO SOKKI DS9110 combustion pressure, thus combustion parameters, such as in-cylinder pressure, heat release rate and in-cylinder temperature, were calculated. The intake air and fuel flow were measured by an air flow meter of SENSYCON Sensy flow P and a fuel flow meter of ONO SOKKI DF-2420, respectively. An independently developed control system based on Intel 80C196 chip was used as the electronic control unit (ECU), which could flexibly control ignition and injection operations. In this research, the three diluent gases (CO2, N2 and Ar) were supplied by compressed gas bottles with purity level of 99%. Their properties at 25 C and 101 kPa are shown in Table 2. The gas from bottles was maintained at 0.8 MPa by a pressure controller before fed into the engine intake manifold. A mass flow controller (MFC) was used to adjust the mass flow of diluents, and their mass flow rate was detected by a Toceil CMF025 dynamical gas mass flow meter with an accuracy of ±0.35%. The ambient temperature of experiments was about 25 C. Meanwhile, because of the pressure drop and kinetic energy rise, according to law of thermodynamics, the gas out of bottle was cooled during the process. Therefore, the temperature of the intake diluent was controlled at about 25 C by heater. Table 3 shows the measurement uncertainty for devices used in this work. Table 1 Experimental engine specification. Engine parameters
Specifications
Engine type Combustion chamber type Bore Stroke Displacement Compression ratio Connecting rod length Intake valve closing Exhaust valve opening
in-line, 4-stroke, 4-cylinder u shape 85 mm 88 mm 1.997L 18.0 175 mm 147 CA BTDC 136 CA ATDC
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Fig. 1. Schematic diagram of the experimental apparatus.
Table 2 Properties of Ar, N2 and CO2 at 25 C and 101 kPa. Gas
Air
Ar
N
Molar mass (kg$kmol1) Gas constant (kJ$kg1$K1) Density (kg$m3) Specific heat capacity (kJ$kg1$K1) Specific heat ratio
28.97 0.2871 1.293 1.004 1.4
39.94 0.2081 1.784 0.524 1.667
28.016 0.2968 1.025 1.038 1.4
2
CO
2
44.01 0.1889 1.977 0.85 1.285
Table 3 Measurement uncertainty analysis.
lambda meter and BOSCH LSU 4.9 lambda sensor at all cases. Engine coolant was controlled at 85 ± 3 C by a closed loop temperature control system. All instruments had been calibrated before they were used. In this test, 200 consecutive combustion cycles were acquired to evaluate cycle-by-cycle variation. In order to intuitively compare and analyze the dilution effects of different gases at the same addition level, addition ratio (AR) was introduced in this paper. It was defined as the ratio of gas charge to the air charge. The formula is as follows:
0
Equipment
Accuracy
Dynamometer Pressure sensor Charge amplifier Crank angle encoder Air flow meter Fuel flow meter
Torque: ±2Nm Speed: ±1 rpm ±0.3% ±0.6% ±0.5% ±0.5% ±0.2%
2.2. Experimental methodology For all cases, the engine speed was maintained at 1400 rpm. By adjusting methanol quantity, the output torque was kept at 43Nm at stoichiometry mixture combustion condition without dilution, which reflects 25% load of this engine at 1400 rpm. The spark advance was 22 CA BTDC (before top dead center) which was the maximum brake torque (MBT) spark timing for the original engine at this operating condition. It is well known that the start of ignition affects the start of combustion [29], which then affects combustion variation [30]. Although the constant spark advance was not realistic for diluent conditions, it would highlight the combustion differences that caused by different physicochemical properties and diluent degree of added gases. This is what this paper wanted to focus. Furthermore, the excess air coefficient was kept at 1 by ETAS
@ ARgas
1 mgas ¼ 100# A mair
(1)
where mair is the practical air mass flow for stoichiometric combustion, and it can be measured by the air flow meter; mgas is the mass flow rates (kg/h) of dilution gas, and it was directly acquired by the gas mass flow meter. The coefficient of variation in indicated mean effective pressure (COVIMEP) is defined as the standard deviation in IMEP divided by the mean IMEP, and is usually expressed in percent [31]:
COVIMEP ¼
sIMEP IMEP
100
(2)
where
IMEP ¼
sIMEP
N 1 X IMEP N i¼1
sffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi PN i¼1 IMEP IMEP ¼ N
(3)
(4)
N is the number of combustion cycle, IMEP is indicated mean effective pressure of individual cycle.
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3. Results and discussions Fig. 2 presents the cycle-by-cycle variations of peak in-cylinder pressure at different addition ratios for the tested three dilutions. For CO2 and N2, the mean peak in-cylinder pressure value was decreased, the variational fluctuation enlarged first and then shrunk with the increase of addition ratio (from 0 to 26.2%). While Ar didn't show clear change in both mean value and variation. Noting that the different values of the mean peak in-cylinder pressure for the three diluents at 0% addition ratio was due to the dispersion in engine behavior. The information of pmax (peak incylinder pressure) reflects the combustion completeness and the amount of heat transfer to the wall [32]. Due to the constant air quantity, the addition of diluent increased in-cylinder mixture heat capacity and decreased O2 concentration, the combustion temperature and flame propagation speed were reduced. Thus, pmax presented a decreasing trend as the addition ratio was increased. Meanwhile, for a small addition ratio, the suppressed chemical reactions also brought about the instability of combustion process, which caused the enlarged fluctuation of peak in-cylinder pressure. While for large addition ratio, pmax was decreased sharply near to the limit of compression terminal pressure, that is why the fluctuation of peak pressure showed a relative compactness distribution. The results also revealed that the decline level of mean incylinder pressure was shrunken successively when adding the same diluent inflow according to the order of CO2, N2 and Ar. Compared with CO2 and N2, Ar-air mixture owns the lowest heat capacity [33] which results in the highest compression temperature. The initial flame kernel formation and development will be improved. Besides, the high thermal diffusivity of Ar [18] will also enhance the chemical reactions and promote combustion process. Similarly, the mixture heat capacity of N2 dilution is lower than that of CO2 dilution, that is why the descent degree of mean in-cylinder pressure of N2 is smaller than that of CO2. Fig. 3 shows the cycle-by-cycle variations of (dp/d4)max (the maximum rate of in-cylinder pressure rise) at different addition ratios for the tested three diluents. The rate of pressure rise reflects burning velocity of mixture and reveals whether the combustion process is timely or not. For all dilutions, with increasing the addition ratio of diluents, the mean value of (dp/d4)max was decreased. This was mainly because that the addition of diluents slowed the burning velocity and prolonged combustion duration, thus the heat release rate was decreased. It was also observed that the fluctuation range of (dp/d4)max didn't show any changes, although it tended to be more centered on the mean maximum pressure rise at large addition ratio (18.6% and 26.2%) of CO2 and N2 dilutions. The reason is similar with the change of peak in-cylinder pressure, and we no longer explain it. The slight decreasing trend in the mean (dp/d4)max of Ar dilution could be attributed to the especial thermophysical parameters of Ar and will be further discussed in the following section. Another result can be seen was that diluent combustion has more impact on the maximum rate of pressure rise than on peak in-cylinder pressure at Ar dilution. Fig. 4 gives the relationship of pmax and 4pmax (the corresponding crank angle at which pmax occurs) when the diluent was added. The magnitudes of the changes in pmax and 4pmax depend on the extent of the cycle-by-cycle variation, and also depend on whether the average burn process is fast or slow. Empirically, the location of the appearance of peak in-cylinder pressure plays an important role in evaluating combustion process, and it is expected to be around 16 CA ATDC at which the best engine dynamic performance occurs [31]. In order to provide a better and deep understanding about this cyclic variation, we chose 5 typical cycles named case 1 to 5 and have pointed out their positions in Fig. 4. The standards of selection are as follows. Case 1, 2, 4 and 5 are cycles
Fig. 2. Cycle-by-cycle variations of peak in-cylinder pressure at different addition ratios; (a) CO2 dilution, (b) N2 dilution, (c) Ar dilution.
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Fig. 3. Cycle-by-cycle variations of the maximum rate of pressure rise at different addition ratios; (a) CO2 dilution, (b) N2 dilution, (c) Ar dilution.
Fig. 4. Peak in-cylinder pressure versus crank angle at which it occurs; (a) CO2 dilution, (b) N2 dilution, (c) Ar dilution.
M. Zhang et al. / Energy 182 (2019) 1132e1140
with maximum pmax of each diluent condition, case 3 is the peak point with 4pmax equals 0 CA at the same diluent condition as case 4. For example, ‘CO2-26.2%-#3’ means the third cycle with the addition ratio of 26.2% at CO2 dilution. The in-cylinder pressure and heat release rate of cases are given in Fig. 5. A linearity between pmax and 4pmax was observed at each condition (except high addition ratio of CO2 and N2 dilution), revealing that high pmax appeared at small 4pmax. This phenomenon was mainly due to the methanol-charging variations, cycle-by-cycle. The results also showed that for CO2 and N2, the value of peak incylinder pressure decreased and its corresponding crank angle moved away from top dead center when increasing addition ratio. As discussed above, the addition of N2 and CO2 slowed the flame propagation and burning rate of mixture, the combustion process was retarded from the optimum when the spark timing was fixed, thus pmax decreased and 4pmax increased. That is, fast burning cycles corresponded to advanced combustion phasing. When further increasing diluents (AR18.6%), the burn process was excessively retarded, the rate of increase of pressure due to combustion became so slow that it was more than offset by the pressure decrease due to volume increase, thus the first peak (around the compression top dead center) of pressure would exceed the second peak (Fig. 5, case 3). Even worse, misfiring cycles started to occur, and partial and non-burning cycles increased rapidly if the mixture was made even more dilute. Faster burning will bring about less cyclic combustion variations on engine performance [34], while for slow burning, where a large fraction of methanol combusted well after top dead center, the effect of volume change became significant and strengthened the combustion variations. For Ar dilution, the distribution of pmax was centered and the value of pmax increased slightly as addition ratio was increased. Although the dilution of Ar would slower flame propagation, the improved initial flame kernel formation and development on account of the low mixture heat capacity offset this disadvantage and made a less variational combustion phasing. Moreover, the promoted chemical reaction and combustion due to the high thermal diffusivity of Ar made pmax increased slightly with increasing amount of diluent gas (Fig. 5, from case 1 to case 2). At addition ratio of 26.2%, the proportions of cycles with 4pmax located away from top dead center in the total cycles for CO2, N2, Ar were 16.5%, 78%, 100%, respectively. It means that at this extend of dilution, CO2 owned the largest number of cycles with slow burning and long combustion phasing, followed by N2 and then Ar. This can
Fig. 5. In-cylinder pressure and heat release rate of the selected cases.
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be attributed to the highest mixture heat capacity of N2 dilution and the highest thermal diffusivity of Ar dilution. In addition, the maximum pmax of Ar (Fig. 5, case 2) dilution was the highest, followed by N2 (Fig. 5, case 5) and then CO2 (Fig. 5, case 4). Fig. 6 presents the relationship of (dp/d4)max and 4(dp/d4)max (the crank angle at which (dp/d4)max occurs) when the diluent was added. The results disclosed that the interdependency of (dp/ d4)max and 4(dp/d4)max was approximately similar with that of pmax and 4pmax. At each condition (except high addition ratio of CO2 and N2), the data showed a relatively vertical change and was centered around 10 CA. The change was mainly because of the variations of the amount of methanol injecting into cylinder each cycle. For CO2 and N2, the (dp/d4)max was decreased and 4(dp/d4)max moved away from top dead center when adding small amount of dilutions (AR<11.7%). This resulted from the slowed flame propagation and lengthening of combustion duration. With increasing addition ratio, as burning lengthened further, slow burn or partial burn cycles were encountered and the maximum rate of pressure rise occurred at compression stroke. As a consequence, the engine became rough and unstable. A modest decrease of (dp/d4)max was also observed as the mixture was diluted by Ar and the distribution of data was relatively stabilized. As discussed above, Ar dilution would increase in-cylinder temperature at compression stroke, hence the incylinder pressure was relatively high (Fig. 5, case 1 and 2) when mixture burned at this period. Thus, the pressure line showed a less steep rise style. The impact on IMEP (indicated mean effective pressure), due to increased addition ratio of diluents, is shown in Fig. 7. As increasing addition ratio from 0 to 26.2%, CO2 and N2 dilutions showed the same changing trend: the descending mean value of IMEP and the gradually enlarged fluctuation of IMEP data, only that the percentage decline of mean IMEP for N2 dilution was lower than that of CO2 dilution (4.6% for N2, 14.6% for CO2). Increasing addition ratio of mixture slowed burning velocity and lengthened combustion phasing, thus the distribution of IMEP data was widened and IMEP was lowered either due to partial burning. While for Ar dilution, the mean value of IMEP showed an inverse increasing trend (increased by 3.6%) and the dispersion of IMEP data didn't presented clear change. This phenomenon resulted mainly from the positive impact of Ar on mixture, given that Ar has a comparatively lower specific heat capacity and higher thermal diffusivity. The method of return maps has widely used to examine the interrelationships between consecutive cycles of combustion process [35]. Note that we only use it to evaluate cyclic variation here. Fig. 8 depicts the return maps of indicated mean effective pressure when increasing addition ratio of diluents, IMEP(i) (IMEP at the current cycle) versus IMEP(iþ1) (IMEP at the next consecutive cycle). It can be easily found from the figure that no misfiring cycles (IMEP <0) had happened during the diluent process for all diluent gases. For specific condition, centralized IMEP data means that the combustion process is stable while scattered IMEP data means that the combustion process is unstable. In detail, data point which locates away from the center stands for low value of IMEP due to partial burning or high value of IMEP caused by the variational amount of methanol injected. For Ar dilution, the IMEP data was always centered in a circular region, only that the cloud of points moved to the upper right area gradually when the addition ratio was increased. While for CO2 and N2 dilutions, as addition ratio increased, the cloud of points moved to lower left area and the cluster of IMEP data expanded, which definitely resulted in high cycle-by-cycle variations. Differences in CO2 and N2 dilutions are also visible, compared with the operating condition with no diluent, the data distribution of CO2 was more scattered and further away than that of N2 at the same addition ratio. COVIMEP (the coefficient of variation in indicated mean effective
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Fig. 6. Maximum rate of in-cylinder pressure rise versus crank angle at which it occurs; (a) CO2 dilution, (b) N2 dilution, (c) Ar dilution.
Fig. 7. Cycle-by-cycle variations of the indicated mean effective pressure at different addition ratios; (a) CO2 dilution, (b) N2 dilution, (c) Ar dilution.
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Fig. 9. Changes of the coefficient of variation in indicated mean effective pressure with the addition ratio of diluents.
pressure) is one important index to measure cyclic variability in indicated work per cycle. The impact of addition ratio on COVIMEP is given in Fig. 9. It can be found that COVIMEP of CO2 and N2 dilutions increased up to 6.2% and 5.8% respectively when elevating addition ratio to 34.8%, while COVIMEP of Ar almost hold steady. Under the same high addition ratio, the COVIMEP of CO2 is the highest followed by N2 and then Ar. Comprehensive analysis of Figs. 7e9 can draw a conclusion that Ar dilution will enhance IMEP of methanol engine without deteriorating engine cyclic variation. In our previous study [21], a 1.39l four-cylinder GDI engine was employed to study the effects of diluents on combustion and emission characteristics. During that experiment campaign, the same three gases were used. Detailed engine and operation information could be found in the mentioned work. A parallel comparison reveals that at this small-speed and small-load condition, the corresponding addition ratio value of limit COVIMEP (more than 5%) is higher in methanol engine (about 30% at CO2) than that in GDI engine (about 20% at CO2). This suggests that the diluent tolerance of methanol engine is higher than that of GDI engine. The reason can be explained by the faster flame propagation of methanol charge than gasoline charge. 4. Conclusions An experimental investigation about the dilution combustion was conducted on a spark ignition engine fueled with methanol. The different effects of diluent gases on cyclic variation were found. The conclusions drawn from this work are listed as follows:
Fig. 8. Return maps of indicated mean effective pressure at different addition ratios; (a) CO2 dilution, (b) N2 dilution, (c) Ar dilution.
(1) With increasing addition ratios of CO2 and N2, the fluctuations of pmax and (dp/d4)max increase first and then decrease due to the limit of compression terminal pressure, the dispersion of IMEP enlarges. While for Ar, the fluctuations of pmax, (dp/d4)max and IMEP changed slightly. (2) Linearities were found between pmax and 4pmax, (dp/d4)max and 4(dp/d4)max at low addition ratio for CO2 and N2, while they will disappear at high addition ratio due to slow burning. (3) The proportion of slow burning cycles of CO2 dilution is higher than that of N2 dilution at high addition ratio.
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(4) Drawn from the IMEP return map and COVIMEP, CO2 and N2 dilution will deteriorate cyclic variation of this methanol engine while Ar dilution will enhance IMEP with slight changes in cyclic variations. (5) Compared with our previous researches, the diluent tolerance of spark ignition methanol engine is higher than that of gasoline engine.
Declaration of interests The authors declare that they have no known competing financial interests or personal relationships that could have appeared to influence the work reported in this paper. The authors declare the following financial interests/personal relationships which may be considered as potential competing interests: Acknowledgements This work was financially supported by National Natural Science Foundation of China (Grant No. 51876079), funded by the special fund for basic scientific research in central colleges and universities, Natural Science Foundation of Jilin Province (20170101137JC, 20170101132JC), Science and technology research projects of Jilin Province education department for 13th Five Year Plan(JJKH20180141KJ), Project of Development and Reform Commission of Jilin Province (2019C058-4), Excellent Youth Talent Fund Project (20190103049JH), Open Fund of state key laboratory of automotive simulation and control, Jilin University (20181110), Graduate Innovation Fund of Jilin University (101832018C199). References [1] Huang YH, Hong G, Huang GH. Investigation to charge cooling effect and combustion characteristics of ethanol direct injection in a gasoline port injection engine. Appl Energy 2015;160:244e54. https://doi.org/10.1016/j. apenergy.2015.09.059. [2] Verhelst S, Turner JW, Sileghem L, Jeroen Vancoillie. Methanol as a fuel for internal combustion engines. Prog Energy Combust Sci 2019;70:43e88. https://doi.org/10.1016/j.pecs.2018.10.001. [3] Fraser N, Blaxill H, Lumsden G, Bassett M. Challenges for increased efficiency through gasoline engine downsizing. SAE Int J Engines 2009;2:991e1008. [4] Su JY, Xu M, Li T, Gao Y, Wang JS. Combined effects of cooled EGR and a higher geometric compression ratio on thermal efficiency improvement of a downsized boosted spark-ignition direct-injection engine. Energy Convers Manag 2014;78:65e73. https://doi.org/10.1016/j.enconman.2013.10.041. [5] Xie FX, Hong W, Su Y, Zhang MM, Jiang BP. Effect of external hot EGR dilution on combustion, performance and particulate emissions of a GDI engine. Energy Convers Manag 2017;142:69e81. https://doi.org/10.1016/j.enconman. 2017.03.045. [6] Brinkman ND. Effect of compression ratio on exhaust emissions and performance of a methanol-fueled single-cylinder engine. SAE paper no. 770791. 1977. [7] Mohanan P, Babu MKG. A simulation model for a methanol-fueled turbocharged multi-cylinder automotive spark ignition engine. SAE paper no. 912417. 1991. [8] Ryan T, Lestz S. The laminar burning velocity of isooctane, N-heptane, methanol, methane and propane at elevated temperatures and pressures in the presence of a diluent. SAE paper no. 800103. 1980. [9] Xie FX, Hong W, Su Y, Zhang MM, Jiang BP. Effect of external hot EGR dilution on combustion, performance and particulate emissions of a GDI engine. Energy Convers Manag 2017;142:69e81. https://doi.org/10.1016/j.enconman. 2017.03.045. tet JF. Experimental study of various effects of [10] Maiboom A, Tauzia X, He exhaust gas recirculation (EGR) on combustion and emissions of an automotive direct injection diesel engine. Energy 2008;33:22e34. [11] Zhang ZJ, Zhang HY, Wang TY, Jia M. Effects of tumble combined with EGR (exhaust gas recirculation) on the combustion and emissions in a spark ignition engine at part loads. Energy 2014;65:18e24. https://doi.org/10.1016/ j.energy.2013.11.062.
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