Effects of Lean Burn on Combustion and Emissions of a DISI Engine Fueled with Methanol–Gasoline Blends

Abstract

Methanol has significant potential as an alternative fuel for internal combustion engines. Using methanol–gasoline blends with lean-burn technology in traditional spark-ignition engines can enhance fuel economy and reduce emissions. This paper investigates the effects of lean burn on the combustion and emissions in a commercial direct-injection gasoline engine fueled with methanol–gasoline blends. The lean-burn mode is adjusted by controlling the injection strategy. The results show that homogeneous lean burn (HLB) has earlier combustion phase and better power performance when the excess air ratio (λ) is less than 1.3, while its combustion phase extends more than stratified lean burn (SLB) when λ exceeds 1.4. Both lean-burn modes achieve optimal fuel economy at λ = 1.2–1.3. Under stable conditions, BSFC decreases with higher methanol blending ratios, with SLB being more economical at low blending ratios and HLB at higher ratios. The lowest HC and particulate matter emissions for both modes are achieved around λ = 1.3. SLB has lower NO_X emissions when λ < 1.3, while HLB shows lower NO_X emissions when λ > 1.3. The particulate size distribution is bimodal for blending lean-burn conditions, with SLB having the highest nucleation mode peak and HLB the highest accumulation mode peak. M20 (20% volume of methanol) corresponds to the highest particle emissions under lean-burn conditions. This study can provide a deeper understanding of methanol–gasoline blending lean burn, and provide a reference for emission control of spark-ignition engines.

1. Introduction

Methanol, a hydrogen-based fuel, plays a unique role in the transformation of energy structure and carbon reduction in the transportation industry [1]. To achieve the “1.5 °C” goal by 2050, hydrogen and its derivative fuels (including methanol and ammonia) need to account for more than 14% of end-use energy consumption [2]. Methanol has been extensively researched as a fuel for internal combustion engines in recent years [3,4], primarily due to its several advantages over traditional fossil fuels. Methanol contains oxygen and has a higher octane number than gasoline [5], allowing for more efficient and cleaner combustion [6]. Methanol is a liquid at normal temperature and pressure, which allows it to be transported and stored like gasoline and diesel, with fewer changes to existing engine structures. Its definite molecular structure, unlike gasoline’s mixture of hydrocarbons, makes its combustion process easier to simulate and predict [7]. Additionally, methanol can be produced from coal, natural gas, and other sources, reducing petroleum dependence and achieving energy independence.

The high octane number of methanol enables it to burn at higher compression ratios, enhancing anti-knock properties and fuel economy. Balki [8] compared the effects of burning gasoline, methanol, and ethanol on engine performance and emissions in a single-cylinder engine, finding that the effective thermal efficiency of pure methanol was higher than that of gasoline across a speed range of 1600 to 3600 rpm. When engine speed, spark timing, and compression ratio were optimized, the effective output power with pure methanol increased by 12.25% compared to gasoline [9]. Güdden [10] tested pure methanol port injection in a 5L single-cylinder large bore engine, demonstrating that methanol’s benefits are equally favorable for large bore engines, with higher effective thermal efficiency than the most advanced natural gas engines of the same displacement. However, using methanol in internal combustion engines also presents challenges, such as producing unconventional emissions like unburned methanol and formaldehyde [11]. Additionally, methanol’s low flash point and high latent heat of vaporization can lead to poor mixture atomization during cold start in low-temperature environments [12]. Blending or dual-fuel applications can mitigate this issue. In spark-ignition engine applications, methanol–gasoline blending has been widely studied, with findings showing that methanol–gasoline blends can accelerate combustion, shorten the flame development and rapid combustion periods, and promote complete mixture combustion [13].

Yang et al. [14] studied the performance of a methanol–gasoline dual-fuel engine under different injection modes, showing that methanol direct injection combined with gasoline port injection resulted in higher mean effective pressure, faster heat release rates, and more stable engine operation. Agarwal et al. [15] found that methanol–gasoline blending improved effective thermal efficiency in gasoline engines, with the M20 blend achieving higher efficiency than M10. Methanol–gasoline blending also resulted in lower NO, CO, and particulate emissions but higher HC emissions. Tian et al. [16] compared the effects of blending methanol, ethanol, and butanol with gasoline on engine combustion and emissions, finding that all three blends improved engine torque and heat release rates, with methanol’s advantages being the most pronounced. Mishra et al. [17] showed that three different methanol–gasoline blends (5%, 10%, 15%) exhibited better emission performance and knock resistance than pure gasoline, with 10% volume of methanol yielding the lowest emissions. They recommended using 5–10% volume of methanol blends to enhance gasoline engine performance and reduce emissions.

Lean burn is an effective method to improve thermal efficiency and reduce emissions in spark-ignition engines. Methanol’s fuel properties provide a wider lean-burn limit compared to gasoline, making methanol combined with lean burn a practical application of methanol in internal combustion engines [18,19,20]. Wu et al. [21] compared the lean-burn characteristics of methanol and gasoline under idle conditions, finding that methanol demonstrated better lean-burn performance, with higher indicated thermal efficiency and fewer conventional emissions under three lean-burn conditions (λ = 1.0, 1.2, 1.4). Gong [22] simulated the lean-burn characteristics of methanol engines at different high compression ratios, showing that at larger equivalence ratios, the indicated thermal efficiency decreased with decreasing compression ratio, while at smaller equivalence ratios, it increased with decreasing compression ratio. A compression ratio of 14 achieved the best compromise between pressure rise control and lean-burn limit expansion. They [23] also experimented with twin-spark plug synchronous ignition in methanol engines, finding that when fuel injection occurred at 110° CA BTDC and ignition at 21° CA BTDC, twin-spark plug ignition allowed stable lean burn at medium compression ratios. Zhou et al. [24] studied the impact of valve lift control strategies on methanol–gasoline engines, showing that lean burn achieved through intake valve lift control exhibited higher fuel economy, and methanol–gasoline blending further extended the lean-burn limit while improving fuel economy and emission performance. In addition to studying pure methanol or methanol–gasoline blends in traditional spark-ignition engines, researchers are also exploring various combustion modes to expand the lean-burn limit of methanol combustion. For instance, Gong et al. [25] investigated the impact of ignition timing on engine combustion and emission performance at an excess air ratio of 2.0 using hydrogen port injection combined with methanol direct injection. Wang et al. [26] studied the effects of different pre-chamber ignition modes on methanol/air mixture combustion. Chen et al. [27] researched the combustion characteristics of natural gas–methanol dual-fuel engines under different excess air ratios. Zhao et al. [28] examined the characteristics and particulate emissions of butanol–methanol–gasoline blends under lean combustion in spark-ignition engines.

Most existing research on methanol lean burn focuses on dual-fuel combustion, twin-spark plug, or higher performance ignition systems (e.g., jet ignition). In comparison, gasoline mixed with methanol directly burned in a spark-ignition engine has certain advantages in convenience and practicality. Therefore, this article hopes to examine the lean burn characteristics of methanol–gasoline blended fuel by studying different lean burn modes. The lean-burn characteristics of methanol–gasoline blends in traditional gasoline engines require more comprehensive study. From the existing research, few literatures have conducted comprehensive comparative studies on homogeneous lean burn and stratified lean burn of methanol–gasoline blends in spark-ignition engines. Therefore, this paper investigates the effects of methanol addition on lean-burn combustion characteristics and conventional emissions in a commercial direct-injection gasoline engine. The study focuses on the differences between homogeneous and stratified lean-burn characteristics under various methanol addition levels by controlling fuel injection. This study can provide a deeper understanding of methanol–gasoline blending lean burn and provide a reference for combustion and emission control of spark-ignition engines.

2. Experimental Setup and Procedure

2.1. Engine and Experimental Apparatus

In this study, the injector is centrally located, and the spark plug is side mounted. The specific parameters and configurations of the engine are listed in

Table 1. Detailed technical specifications of the test engine.

Engine Parameters	Specifications
Engine type	In-line, 4-Cylinder, 4-Stroke
Combustion System	Spray-Guided GDI
Displacement	1.39 L
Bore × Stroke	76.5 mm × 75.6 mm
Compression Ratio	10:1

. The experimental setup used in this study is shown in Figure 1.

The cylinder pressure was monitored by a quartz crystal pressure sensor model ZF42 (Kistler, Winterthur, Switzerland), and the crank angle was measured by an angle encoder. A combustion analyzer (model DS-9110, ONO SOKKI, Yokohama, Japan) displayed the cylinder pressure curve in real time based on the cylinder pressure and crank angle signals and calculated heat release rate, cylinder temperature, cumulative heat release, and other combustion characteristic data as needed. The intake airflow was measured by an air flow meter, and the fuel flow rate was monitored in real time by an DF-2420 (ONO SOKKI, Yokohama, Japan) flow meter.

Control parameters were managed by a self-designed electronic control unit based on a 16-bit Freescale microcontroller. Emissions of gases such as hydrocarbons and nitrogen oxides were measured by DICOM 4000 (AVL, Graz, Austria) gas analyzer. Particulate emissions were measured by a DMS500 (Cambustion, Cambridge, UK) fast particulate analyzer with a sampling frequency of 10 Hz, collecting data for 30 s per condition and calculating the average concentration values. The excess air ratio during the experiments was monitored and controlled by a lambda meter and an LSU 4.9 (BOSCH, Leinfelden, Germany) lambda sensor. The engine coolant temperature was maintained at 76 ± 2 °C. Engine speed and load were controlled by an eddy current dynamometer. All experimental instruments were calibrated before use, and the accuracy of tyingghe main instruments is listed in

Table 2. Measurement uncertainties of devices.

Device	Model	Uncertainty
Dynamometer	CW160	Torque: ±2 Nm; Speed: ±1 r/min
Pressure sensor	ZF42	±0.3%
Charge amplifier	FLEXIFEM PIEZO	±0.6%
Crank angle encoder	LF-72BM-C05E	±0.5%
Air flow meter	SENSYCON	±0.5%
Fuel flow meter	DF-2420	±0.2%
Gas analyzer	DICOM 4000	HC: ±30 ppm; CO: ±0.01%; NO: ±20 ppm; CO2: ±0.1%

.

2.2. Experimental Methodology

During the experiments, the engine speed was maintained at 1400 rpm, and the output torque under stoichiometric conditions was controlled at 40 Nm for each test group. The fuel injection quantity remained constant while adjusting the excess air ratio. The study focuses on the effects of the excess air ratio and methanol blending ratio on the combustion process and emissions. To maintain a consistent research focus, the ignition timing was set at 16° CA BTDC under all working conditions, minimizing the influence of ignition timing on the combustion process, even though a fixed ignition advance is not ideal for most conditions [29]. Four different methanol–gasoline blends were used in the experiments: M0, M10, M20, and M40, where M0 represents pure gasoline, and M10 indicates a methanol volume fraction of 10% in the blended fuel. Each fuel was tested under two lean-burn modes: homogeneous lean burn (HLB) and stratified lean burn (SLB). In HLB mode, there was a single injection at 300° CA before top dead center (BTDC). In SLB mode, there were two injections at 300° CA BTDC and 70° CA BTDC. Before conducting this study, we conducted a preliminary test on the ratio of the two injections and found that when the ratio of the first and second injections was 8:2, the particulate emissions were the lowest. Therefore, in SLB mode, the ratio of the two injections was 8:2. The excess air ratio was varied from 1 to 1.6 for each lean-burn mode. Data recording began after the engine reached relatively stable operation for each condition, with combustion characteristics calculated from 200 consecutive combustion cycles. Engine stability was indicated by the coefficient of variation of the mean indicated effective pressure, defined as follows:

$${COV}_{IMEP}={\displaystyle \frac{{\sigma }_{IMEP}}{\overline{IMEP}}}\times 100\%$$

(1)

where

$$\overline{IMEP}={\displaystyle \frac{1}{N}}\sum _{i=1}^{N}IMEP$$

(2)

$${\sigma }_{IMEP}=\sqrt{{\displaystyle \frac{\sum _{i=1}^N\left(IMEP-\overline{IMEP}\right)}{N}}}$$

(3)

$N$ represents the combustion cycle number, and $IMEP$ represents the indicated mean effective pressure value of each cycle.

3. Results and Discussion

3.1. Effect of Lean Burn on Combustion and Engine Performance

Figure 2 illustrates the effects of different methanol blending ratios and excess air ratios on the flame development period and combustion duration. In these figures, “H” represents homogeneous lean burn (HLB) and “S” represents stratified lean burn (SLB), denoted by solid and dashed lines, respectively. This notation is used consistently throughout the text. In this study, the flame development period is defined as the crank angle degrees from the start of ignition to 10% cumulative heat release, while the combustion duration is defined as the crank angle degrees from 10% to 90% cumulative heat release. The results indicate that as the lean-burn degree increases, both the flame development period and combustion duration lengthen under all conditions. This is primarily due to the increased heat capacity of the mixture from added air, which lowers the end of compression temperature, slows the initial flame development and propagation, and inhibits chemical reactions. Furthermore, for the same lean-burn mode, the variation trends for different methanol blending ratios are consistent, and the magnitude of change is similar. When the excess air ratio is less than 1.3, the flame development period and combustion duration are shorter for HLB. However, when the excess air ratio exceeds 1.4, these periods deteriorate more significantly for HLB. This is because, at high excess air ratios, the overall mixture in the cylinder becomes excessively lean in HLB mode due to early injection, leading to potential misfires. Conversely, a small amount of late injection forms a locally rich zone near the spark plug, facilitating stable combustion.

Figure 3 shows the effects of different methanol blending ratios and excess air ratios (λ) on the combustion center (CA50) and the coefficient of variation of indicated mean effective pressure (COV_IMEP). In this study, the combustion center is represented by the crank angle corresponding to 50% cumulative heat release, denoted as CA50. The definition of the coefficient of variation of cycle-by-cycle variation is referred to as described earlier. Generally, the engine performs best when the combustion center is around 10 crank angle degrees. An advanced combustion center indicates more fuel is burned before top dead center, and a delayed center indicates more fuel is burned during the piston downstroke, both of which are unfavorable for efficient energy conversion. As the excess air ratio increases, the flame development and combustion duration lengthen, causing more fuel to burn during the power stroke, and the combustion phase is gradually delayed. Similarly, the locally rich mixture in SLB stabilizes initial combustion temperatures, causing the combustion center to advance compared to HLB under globally lean conditions, i.e., closer to top dead center. With increasing excess air ratio, the instability of combustion increases under all conditions. At λ = 1.3, the COV_IMEP approaches or exceeds 5%, and the instability worsens with larger λ.

To analyze the effects of methanol addition and minimize interference from combustion instability, Figure 4 shows the results for combustion center and cycle-by-cycle variation for excess air ratios from 1 to 1.3. Each graph consists of two parts, the upper part is the HLB condition, and the vertical axis corresponds to the red line scale; the lower part is the SLB condition, and the vertical axis corresponds to the black line scale. Generally, at the same excess air ratio, for the increases in methanol blending ratio, the combustion center initially delays slightly and then gradually advances. This can be explained as follows: the high oxygen content in methanol increases the laminar flame speed compared to gasoline, resulting in faster combustion with higher methanol content. However, due to methanol’s lower heating value, a higher methanol blending ratio requires a larger fuel injection quantity at a fixed load, leading to a longer injection duration. This results in more fuel evaporating near ignition, causing insufficient mixing and an extended combustion phase. At lower blending ratios, the latter effect predominates, while at higher blending ratios, the former advantage becomes more evident. Notably, at λ = 1.2, SLB with M10 fuel achieves the most advanced combustion center, likely because the 16° BTDC ignition timing aligns with its MBT point. Furthermore, as shown in Figure 4 and combined with Figure 3, HLB exhibits more stable combustion at lower excess air ratios compared to SLB. At higher excess air ratios, the cycle-by-cycle variation coefficient for SLB is relatively smaller, reflecting the benefit of locally rich mixtures in promoting combustion under globally lean conditions.

Figure 5 illustrates the effects of excess air ratio on in-cylinder pressure and heat release rate. Solid lines represent in-cylinder pressure curves, and dashed lines represent heat release rate curves. The results show that with increasing excess air ratio, the peak cylinder pressure gradually decreases and shifts away from top dead center, with similar trends observed in the heat release rate curves. Figure 6 shows the relationship between the average peak cylinder pressure of 200 combustion cycles and the corresponding crank angle. The excess air ratio increases in the direction indicated by the arrow. The peak cylinder pressure gradually decreases, and the corresponding crank angle initially moves away from top dead center and then returns closer to top dead center. This is because as the combustion phase extends, more fuel burns during the piston downstroke, resulting in a flatter pressure curve. Especially when the excess air coefficient is too large, the cylinder pressure formed by compression and initial combustion becomes higher. In comparison, the decrease in peak cylinder pressure is smaller for SLB, and the corresponding crank angle changes more gradually.

The effect of the excess air ratio on exhaust temperature is shown in Figure 7. As the excess air ratio increases from 1 to 1.5, the exhaust temperature for SLB gradually decreases, while for HLB, it first decreases and then increases. On the one hand, a higher degree of air dilution increases the heat capacity of the mixture, resulting in lower combustion temperatures. On the other hand, an extended combustion phase increases the proportion of late combustion, causing a rise in exhaust temperature. For HLB, when λ does not exceed 1.3, the increase in combustion duration is relatively slow, and the mixture’s heat capacity effect predominates, lowering the combustion temperature. As λ continues to increase, the combustion phase delays significantly, leading to a dominant late combustion part and a subsequent rise in exhaust temperature. For SLB, the presence of locally rich mixtures causes the combustion duration to change slowly, even when λ exceeds 1.3, keeping the reduction in combustion temperature dominant and resulting in a gradual decrease in exhaust temperature. Additionally, when λ = 1.6, the exhaust temperature for HLB decreases again, possibly due to high cycle variation coefficients and severe misfires. The figure also shows that under the same excess air ratio, a higher methanol blending ratio results in a higher exhaust temperature, explained by the higher combustion rate of methanol.

Figure 8 shows the effect of the excess air ratio on the brake-specific fuel consumption (BSFC). In this study, the equivalent gasoline BSFC for blended fuels is calculated based on equal heating value. The BSFC decreases first and then increases with the increase in the excess air ratio under all conditions, reaching a minimum around λ = 1.2–1.3. With the increase in lean combustion, the combustion temperature decreases, reducing the heat transferred to the cylinder walls and piston, thereby lowering heat losses. Additionally, as the excess air ratio increases from 1 to 1.3, both HLB and SLB show a decrease in exhaust temperature, indicating less heat is carried away by the exhaust gases, thereby improving fuel conversion efficiency and reducing BSFC to the minimum value. Although the combustion duration gradually extends during this process, the increase is slow. When λ exceeds 1.3, the combustion duration increases sharply, causing more fuel to burn during the piston downstroke, which cannot form effective cylinder pressure. At this point, the negative impact of the delayed combustion phase on combustion predominates, and BSFC begins to increase.

The experimental results also show that under relatively stable combustion conditions (λ = 1–1.4), as the methanol blending ratio increases, the BSFC gradually decreases. This is mainly because higher methanol content in the fuel increases the constant volume combustion degree and reduces pumping losses. Additionally, at low methanol blending ratios, SLB is more economical, while at higher methanol blending ratios, HLB has lower BSFC, likely due to much lower exhaust temperatures compared to SLB, resulting in less heat being expelled.

3.2. Effect of Lean Burn on Emissions

Figure 9 shows the changes in HC emissions in the exhaust under different conditions. As the excess air ratio increases, HC emissions first decrease and then sharply increase, with the minimum value appearing around λ = 1.3. HC emissions are mainly affected by unburned fuel and late oxidation. Under lean combustion conditions, sufficient oxygen is present, and the degree of late oxidation is mainly determined by temperature. At lower λ, as air dilution increases, the combustion phase delays, the proportion of late combustion increases, and the temperature rises, promoting the oxidation of unburned HC. As the excess air ratio continues to increase, the mixture becomes excessively lean, combustion temperature drops sharply, flame quenching intensifies, incomplete combustion increases, and unburned HC increases. This is especially true for SLB, where the concentration gradient is larger, resulting in relatively higher unburned HC emissions.

The effect of the excess air ratio on NO_X emissions in the exhaust is shown in Figure 10. For both lean-burn modes, the volume fraction of NO_X in the exhaust first increases and then gradually decreases, reaching a maximum at λ = 1.1. The NO_X produced by methanol–gasoline blended direct-injection combustion is primarily thermal NO_X. According to the Zeldovich theory, the formation of thermal NO_X is mainly determined by the oxygen concentration and temperature within the cylinder. As combustion transitions from stoichiometric to lean, the combustion temperature decreases, inhibiting NO_X formation. However, during lean combustion, the excess air increases the oxygen concentration in the cylinder, promoting NO_X formation. At this stage, the effect of excess oxygen predominates, slightly increasing the volume fraction of NO_X emissions. When air dilution continues to increase, the impact of combustion temperature surpasses that of oxygen concentration, and the formation of thermal NO_X gradually decreases.

When the excess air ratio is less than 1.3, SLB forms a situation where the area around the spark plug is locally rich while the surrounding area is lean, resulting in a relatively lower oxygen concentration during the rapid combustion period compared to HLB, thereby reducing the production of thermal NO_X. Consequently, NO_X emissions are lower for SLB at this stage. When the excess air ratio exceeds 1.3, the influence of combustion temperature during the rapid combustion period starts to dominate, producing more NO_X, while NO_X emissions for HLB are lower. Additionally, under the same air dilution conditions, higher methanol blending ratios result in lower NO_X emissions due to faster combustion rates, shortening the high-temperature duration during combustion and reducing the formation of thermal NO_X.

Figure 11 shows particulate emissions under different conditions, where (a) and (b) illustrate the nucleation mode and accumulation mode particulate concentrations, respectively. Nucleation mode particulates represent small-sized particles with diameters ranging from 0 to 30 nm, primarily formed by the condensation of volatile substances, including hydrocarbons and sulfates. Additionally, primary soot particles, which have diameters around 10–80 nm, are also part of the nucleation mode particulates. Accumulation mode particulates represent larger particles with diameters ranging from 20 to 500 nm, primarily produced by the aggregation of soot resulting from the incomplete combustion of locally rich mixtures [30,31].

For all conditions, as the excess air ratio increases, both nucleation mode and accumulation mode particulates exhibit similar trends. The total particulate concentration first decreases and then increases, achieving the minimum value within the range of λ = 1.2–1.4. Compared to stoichiometric combustion, lean combustion results in more thorough fuel mixing, reduced high-temperature hypoxia, and decreased soot production and HC formation, leading to lower concentrations of both nucleation mode and accumulation mode particulates. With further increases in air dilution, the rapid rise in HC emissions leads to increased condensation of hydrocarbons in the exhaust, raising the nucleation mode particulate concentration. Additionally, the deterioration of combustion stability and severe misfires result in locally rich mixtures or individual cycles of incomplete combustion within the cylinder, worsening the emission of accumulation mode particulates.

Figure 12 presents the particle size distribution under stable lean-burn conditions (λ = 1.1–1.4). The distributions for all conditions are bimodal, with particles predominantly in the 5–300 nm range. The first peak around 20 nm indicates nucleation mode particulates, while the second peak around 100 nm represents accumulation mode particulates. For pure gasoline, only when SLB and the excess air coefficient is 1.1, there are more nucleation mode particulates, and the peak values of accumulated particulates are higher in other conditions. For methanol–gasoline blends, SLB is more likely to have the highest peak at the nucleation mode particulates, while HLB is more likely to have the highest peak at the accumulation mode particulates. This is because the combustion speed of the blended fuel is faster and the high-temperature duration is shorter, which shortens the agglomeration and adsorption time of volatile substances and reduces the number of large-size particles. Compared with HLB, SLB corresponds to a lower combustion temperature, so the process of generating primary carbonaceous particles by H-abstraction and thermal pyrolysis is suppressed. In addition, the exhaust temperature of SLB is higher, which also promotes the oxidation of particles in the later stage of combustion and reduces the particle size.

Since the total particulate concentration in the exhaust is lowest around λ = 1.3, and combustion instability increases sharply beyond this point, this study examined the effects of methanol blending and lean combustion modes on particle size distribution under this condition. Figure 13 shows the particle size distribution at an excess air ratio of 1.3 for different methanol blending ratios. For both SLB and HLB, M20 fuel produces the highest number of particles, with SLB showing a dominance of nucleation mode particulates and HLB showing accumulation mode particulates. Increasing the methanol blending ratio raises combustion speed and temperature, enhancing the dehydrogenation reaction and increasing primary soot particles. However, higher methanol ratios absorb more heat during vaporization, lowering initial mixture temperature, reducing air volume, and increasing oxygen concentration, which leads to more complete combustion and lower accumulation mode particulate concentrations. Furthermore, the particles in the exhaust are also affected by the degree of oxidation in the late combustion phase. Thus, for HLB, the exhaust temperature of M10 is higher than that of M20, resulting in more complete late-stage oxidation of the particles, and the overall particle concentration is lower for M10 compared to M20.

4. Conclusions

In this study, we investigated the effects of lean combustion on the combustion and emission characteristics of a direct-injection spark-ignition engine using methanol–gasoline blended fuel. The main conclusions of this study are as follows:

(1)

For excess air ratios less than 1.3, HLB shows shorter flame development and combustion durations, a combustion center closer to top dead center, and smaller cyclic variations. For excess air ratios greater than 1.4, the HLB combustion phase extends more than that of SLB, with larger cyclic variations. Under stable conditions, the combustion center initially delays and then advances with increasing methanol blending ratios.

(2)

Optimal fuel economy is achieved by both HLB and SLB within the excess air ratio range of 1.2–1.3. Under the same stable conditions, BSFC decreases with higher methanol blending ratios, with SLB being more economical at lower blending ratios and HLB at higher ratios.

(3)

HLB and SLB achieve the lowest HC emissions at λ = 1.3 and the highest NO_X emissions at λ= 1.1, respectively. SLB has lower NO_X emissions when λ < 1.3, while HLB shows lower NO_X emissions when λ > 1.3.

(4)

Both lean-burn modes show the lowest particulate concentrations around the excess air ratio of 1.3. Under stable blending lean-burn conditions, the particulate size distribution is bimodal, with SLB having the highest nucleation mode peak and HLB the highest accumulation mode peak. M20 corresponds to the highest particle emissions under lean-burn conditions.