Experimental Investigation on the Effects of Direct Injection Timing on the Combustion, Performance and Emission Characteristics of Methanol/Gasoline Dual-Fuel Spark Turbocharged Ignition (DFSI) Engine with Different Injection Pressures under High Load

Abstract

The exceptional properties of methanol, such as its high octane number and latent heat of evaporation, make it an advantageous fuel for efficient utilization in dual-fuel combustion techniques. The aim of this study is to investigate the effect of direct methanol injection timing on the combustion, performance and emission characteristics of a dual-fuel spark ignition engine at different injection pressures. We conducted four different direct injection pressure tests ranging from 360° ahead to 30° CA ahead at 30° CA intervals. The experimental results indicate that regardless of the injection pressure, altering the methanol injection timing from −360° to −30° CA ATDC leads to distinct combustion behavior and changes in the combustion phase. Initially, as the injection timing is delayed, the combustion process accelerates, which is followed by a slower combustion phase. Additionally, the combustion phase itself experiences a delay and then advances. Regarding performance characteristics, both the brake thermal efficiency (BTE) and exhaust gas temperature (EGT) exhibit a consistent pattern of first increasing and then decreasing as the injection timing is delayed. This suggests that there is an optimal injection timing window that can enhance both the engine’s efficiency and its ability to manage exhaust temperature. In terms of emissions, there are different trends in this process due to the different conditions under which the individual emissions are produced, with CO and HC showing a decreasing and then increasing trend, and NOx showing the opposite trend. In conclusion, regardless of the injection pressure employed, adopting a thoughtful and well-designed injection strategy can significantly improve the combustion performance and emission characteristics of the engine. The findings of this study shed light on the potential of methanol dual-fuel combustion and provide valuable insights for optimizing engine operation in terms of efficiency and emissions control.

1. Introduction

In the context of environmental trends in today’s world, with the global movement toward achieving carbon neutrality, the demand for emission-reduction technologies is on the rise [1]. Corresponding responses have been proposed in various areas. In the realm of transportation, electric vehicles have emerged as a promising solution to combat carbon emissions [2]. Nevertheless, there are still several hurdles that need to be overcome. But this still has some limitations. Electric vehicles produce zero carbon emissions during operation, and the production and disposal of batteries still contribute to carbon emissions. As a result, internal combustion engines (ICEs) will continue to have a prominent role in the transportation industry for the foreseeable future. Consequently, it is imperative to find ways to make ICEs carbon neutral.

The escalating cost of fuel and the increasingly strict emissions regulations have compelled the automotive sector to focus on enhancing the efficiency and eco-friendliness of ICEs. Despite being well-established power sources, gasoline engines still hold immense potential for further advancement. The prevailing trend in gasoline engine development centers around augmenting power density and compression ratio. This is achieved through innovative strategies such as downsizing engines and implementing turbocharging. However, a notable challenge faced by gasoline engines, especially in turbocharged applications, is the low octane rating of pure gasoline fuel [3]. This low octane rating renders the engine vulnerable to detonation under high loads, ultimately diminishing thermal efficiency. Additionally, the availability of gasoline derived from fossil fuels is limited, posing restrictions on its usage and contributing to severe environmental pollution. There is a need to develop an environmentally friendly fuel for use in internal combustion engines [4].

In order to achieve carbon neutrality, various industries have come up with different response technologies. In the field of internal combustion engines, renewable energy-fueled internal combustion engines hold great promise. In the field of fuels for internal combustion engines, there is a need to develop fuels that are adapted to the trends in the development of internal combustion engines. Higher compression ratios in internal combustion engines and the use of turbocharged engines are increasing the popularity of high-octane fuels [5].

Combining low carbon and high octane characteristics, methanol is a very suitable fuel choice [6]. Methanol has a lower carbon content and a higher octane rating. As far as the source of methanol is concerned, methanol has a well-established production process that allows it to be produced at low cost and in sufficient quantities. In addition, the physicochemical properties of methanol are particularly suitable for engine fuel. In the actual use of methanol, methanol has great advantages. Methanol is a low-priced renewable fuel that is easy to transport and store, making it a good choice for alternative energy. In comparison to gasoline, methanol has a higher octane number, resulting in better thermal efficiency [7]. Methanol is easier to store than fuels such as hydrogen and natural gas. Direct injection engines can also mix methanol into the cooling cylinder because it has a higher latent heat of vaporization. Since methanol contains a large amount of oxygen and has no carbon–carbon bonds, very little particulate emissions are produced during the combustion of methanol [8,9].

In summary, methanol fuel can significantly improve engine performance in terms of combustion and emission. The combustion and emission characteristics of a methanol direct injection engine at different compression ratios were investigated by Gong et al. Compared to gasoline, methanol direct injection engines can adopt high compression ratios, which has a great improvement on engine power and emission characteristics [7]. The reduction in engine particulate matter (PM) was achieved using bio-butanol–methanol–gasoline blends fuel by Zhao et al. [9]. Compared to gasoline fuel, bio-butanol–methanol–gasoline produces particulate emissions at a very low level. Effects on knock intensity and specific fuel consumption were investigated by Breda et al. in a turbocharged GDI engine with port water/methanol injection [10]. All the mixtures analyzed in the study demonstrate an improvement of over 10% in indicated specific fuel consumption (ISFC) when compared to the gasoline-only scenario.

However, in real-life usage, methanol presents challenges. Its low energy density requires a larger fuel tank, affecting a vehicle’s range [11]. And the use of pure methanol fuel in the engine can cause starting difficulties because methanol can result in a longer ignition delay [12,13]. To solve this problem, instead of using methanol fuel alone, methanol is used as a supplementary fuel to gasoline. Dual-fuel engines have a number of advantages over single-fuel engines, with lower emission levels, more energy efficiency, and the ability to combine the benefits of both fuels [14]. In other words, in the combined use of methanol and gasoline, the majority of the energy is provided by gasoline with methanol providing only a small portion of the energy. However, the use of dual fuels as a certain emission characteristic does not guarantee a positive impact [15]. The exact effect of the impact needs to be combined with the injection strategy. In addition, a suitable direct methanol injection strategy is required to maximize the benefits of methanol. This will allow the full utilization of methanol and maximize the use of methanol to improve characteristics. In the mode of methanol direct injection plus gasoline port injection, the methanol injection timing and injection pressure are important parameters of the injection strategy. This paper experimentally investigates the effect of methanol injection moment on the engine at different pressures.

2. Experimental Apparatus and Procedure

2.1. Test Engine and Fuel

Experiments were conducted on a compression ratio (CR) of 9.6 1984cc in-line four-cylinder (EA888, Volkswagen, Berlin, Germany), turbocharged engine that utilized a methanol/gasoline dual fuel system. The enginewas retrofitted with low flow injectors. For more information on the specific engine parameters, please refer to

Table 1. Specifications of the test engine [16,17].

Engine Parameters	Specifications
Engine type	In-line 4-cylinder, turbocharged
Bore × stroke (mm)	82.5 × 92.8
Displacement (L)	1.984
Compression ratio	9.6:1
Maximum power	137 kW (5500 rpm)
Maximum torque	320 N·m (1400–4000 rpm)
Maximum direct pressure	24 MPa

.

In this comprehensive research paper, we conducted an extensive investigation on the utilization of specific fuels for port injection and utilized methanol as a fuel for direct injection. To ensure accurate results, we employed gasoline with a research octane number of 92, a widely accepted standard in the automotive industry, as the fuel for port injection. To facilitate easy reference and comprehension, we compiled a comprehensive table,

Table 2. Fuel properties [11,18].

Fuel Properties	Gasoline	Methanol
Chemical formula	C4–C12	CH3OH
Content of O (%)	0	50
RON	92	110
Density at 293 K (kg/m3)	796	830
Stoichiometric air–fuel ratio	14.3	6.5
Lower heating value (MJ/kg)	43.5	19.7
Latent heat of vaporization (kJ/kg)	301.2	1109
Laminar flame speed (m/s)	0.38	0.523

, which presents the key physical and chemical properties of methanol and gasoline. By including this table, we aim to provide readers with a comprehensive overview of the fundamental properties of methanol and gasoline.

2.2. Experimental Instrumentation

Figure 1 shows the schematic of the experimental apparatus used in this study. Table 3 shows the instrumental parameters for the experimental studies in this paper. The table shows the parameters, type, precision and range of the instrument.

A dynamometer (CW160, Kemai, Tianjin, China) is connected to the test engine to measure the torque and speed. The in-cylinder pressure was measured by a piezoelectric pressure transducer (Kistler 6058A, Winterthur, Switzerland) in conjunction with a charge amplifier (AVL ZI45, Graz, Austria). The crank angle was recorded with a crank angle encoder (ERN420, Heidenhain, Traunreut, Germany), which has a resolution of 0.1° CA. The in-cylinder pressure and crank angle encoder are connected to the combustion analyzer. A lambda sensor (LSU4.9, Bosch, Gerlingen, Germany) was connected to the lambda meter (ETAS, Stuttgart, Germany) to measure the excess air coefficient. Two fuel consumption meters (CFMG025, ToCeil, Shanghai, China) (DF2420, Ono-Sokki, Yokohama, Japan), respectively, recorded the gasoline and methanol mass flow rates. The gaseous emissions, including hydrocarbon (HC), carbon monoxide (CO), and nitrogen oxide (NOx) emissions, were measured by the emission analyzer (MEXA-584L, Horiba, Osaka, Japan).

The filter has a great influence on the intake pressure [19], so the pressure sensor of the intake pressure control system is arranged on the intake manifold, and the pressure sensor can feed back the value of the pressure to the ECU to realize the control of the intake pressure by adjusting the throttle. Meanwhile, thermocouples are arranged on the intake manifold corresponding to the four cylinders of the engine to feed back the collected intake air temperature to the intake air temperature control system to realize the control of intake air temperature.

Table 3. The specifications of the test equipment [20,21].

Parameters	Type
Speed	CW160
Torque	CW160
Pressure in-cylinder	AVL ZI45
Crank angle	Heidenhain ERN420
λ	ETAS Lambda Meter
CO	Horiba MEXA-584L
HC	Horiba MEXA-584L
NOX	Horiba MEXA-584L
Gasoline mass flow rate	ToCeil CMFG025
Methanol mass flow rate	Ono-Sokki DF2420

Table 3. The specifications of the test equipment [20,21].

Parameters	Type
Speed	CW160
Torque	CW160
Pressure in-cylinder	AVL ZI45
Crank angle	Heidenhain ERN420
λ	ETAS Lambda Meter
CO	Horiba MEXA-584L
HC	Horiba MEXA-584L
NOX	Horiba MEXA-584L
Gasoline mass flow rate	ToCeil CMFG025
Methanol mass flow rate	Ono-Sokki DF2420

2.3. Experimental Procedure

The speed was kept constant at 1500 rpm, and the initial torque was approximately 185 N·m. The test conditions were set at λ = 1.0 with a manifold pressure of 120 kPa and an inlet air temperature of 42 °C. The injection parameters could be adjusted by the electronic control unit (ECU). The coolant temperature was maintained at 90 ± 2 °C. The specific parameters of the injection strategy are shown in

Table 4. Operating conditions at various injection timings.

Parameters	Specifications
Injection pressure	4, 6, 8, 10 MPa
Injection timing	−330~−30° CA ATDC (interval 30° CA)
Gasoline port fuel injection pressure	305 kPa
Gasoline port fuel injection timing	−320° ATDC
Fuel injection mass ratio of PFI	7.20 kg/h
Fuel injection mass ratio of DI	0.72 kg/h
Intake air temperature °C	42 °C
Intake air pressure	120 kPa
Engine speed	1500 rpm
Spark timing	−5° CA ATDC

.

To ensure the accuracy of the experiment, in-cylinder pressure data for all conditions were recorded using a combustion analyzer, encompassing 200 cycles.

The heat release rate (HRR) was calculated as follows [22]:

$$HRR=\frac{\gamma }{\gamma -1}p\frac{dV}{d\theta }+\frac{\gamma }{\gamma -1}\frac{dV}{d\theta }V$$

(1)

where p denotes the in-cylinder pressure, θ denotes the crank angle, γ denotes the heat capacity ratio, and V is the cylinder volume.

Throughout this paper, CA50 and CA90 denote the crankshaft angles associated with the 50% and 90% mass combustion portions, respectively. The ignition delay (CA0-10) is determined as the crank angle interval between the start of injection (CA01) and CA10. Moreover, the combustion duration (CA10-90) is evaluated by measuring the crank angle span between CA10 and CA90 [23].

The coefficient of variation (COV) is commonly used to evaluate the combustion stability of an engine. In this paper, the COV of IMEP (COV_IMEP) is used to represent cycle-by-cycle variations. The COV is defined by the following formula [24]:

$$COVx=\frac{\sigma x}{\overline{x}}\times 100\%$$

(2)

where

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

(3)

$$\sigma x=\sqrt{\frac{{\displaystyle \sum _{i=1}^N(xi-\overline{x})}2}{N}}$$

(4)

Brake thermal efficiency (BTE) was calculated as follows [25,26]:

$$BTE=\frac{P_B}{M{F}_{methanol}\times LH{V}_{methanol}+M{F}_{gasoline}\times LH{V}_{gasoline}}\times 100\%$$

(5)

In the numerator, P_B denoted the brake power generated by the engine. The denominator as a whole represents the input energy, which is provided by both methanol and gasoline. Here, MF_methanol and MF_gasoline are the injection mass flow of methanol and gasoline, respectively, and LHV_methanol and LHV_gasoline indicate the corresponding lower heating values.

3. Results and Discussion

3.1. Effect of Injection Timing on Combustion Characteristics

Figure 2 demonstrates the effect of injection timing on in-cylinder pressure and heat release rate with injection pressures of 4, 6, 8, and 10 MPa. The graph depicts the variations in in-cylinder pressure and heat release rate throughout the crank angle for different injection timings, ranging from −360° to −30° CA ATDC. Initially, there is a slight decrease in in-cylinder pressure as the injection timing shifts from −330° to −180° CA ATDC. Subsequently, as the injection timing further retards until reaching −30° CA ATDC, the cylinder pressure shows an increase. In the range of −180° to −150° CA ATDC, the injection timing represents an inflection point in the cylinder pressure trend, with the in-cylinder pressure curve reaching its lowest state. Both the heat release rate and in-cylinder pressure follow similar overall patterns, reaching the same inflection point in this range.

Figure 3a showcases the effect on the maximum in-cylinder pressure (Pmax) and its corresponding phase (APmax). In parallel, Figure 3b illustrates the influence of the maximum heat release rate (HRRmax) and its corresponding phase (AHRRmax). For each injection pressure, Pmax exhibits a consistent pattern of initial decline followed by a subsequent rise with the progression of injection timing. A similar trend is observed in HRRmax across different injection timings. As the injection timing advances, both Pmax and HRRmax initially decrease and then increase. Irrespective of the injection pressure conditions, Pmax, HRRmax, and the minimum values are consistently observed within the range of −180° to −150° CA ATDC. Furthermore, within this range, APmax and AHRRmax exhibit the latest occurrences under each injection pressure condition.

As a result, the presence of methanol direction injection has a pronounced impact on both in-cylinder pressure and heat release rate when injection timing falls within the range of −180 to −150° CA ATDC. This phenomenon is caused by the pressure in the cylinder during methanol direct injection. This is because at an injection timing of −180° ATDC, according to the characteristics of the piston engine, the piston resides at the BDC (bottom dead center), resulting in the lowest in-cylinder pressure state [27]. Consequently, the cooling effect of methanol is maximized in this scenario. Methanol atomization efficiently dissipates a substantial amount of heat prior to combustion initiation, leading to a reduced initial temperature and pressure of compression stroke within the cylinder. This reduction in temperature and pressure subsequently causes a decrease in both Pmax and HRRmax, while also inducing a delay in their respective phases.

Figure 4 depicts the effect of varying injection timing and pressure on the combustion process and phase. A clear pattern of change can be seen in the curve. When the injection timing was from −360° CA to −30° CA ATDC, combustion phases CA50 and CA90 experienced a delay followed by advancement in their respective combustion phases. On the other hand, CA10−90 initially increased and then steadily decreased, while CA0-10 remained relatively constant. In other words, the influence of injection timing on CA0-10 was not as significant as its impact on CA0-90 and CA10-90. The reason for this phenomenon is that they are affected by different conditions. This is because combustion starts in the vicinity of the spark plug, so the duration of CA0-10 is largely influenced by the environment in the vicinity of the spark plug with temperature and mixture concentration being the main factors. In other words, direct methanol injection has little impact on the area. But CA0-90 and CA10-90 rely on the temperature; this is because the activity of the fuel is positively correlated with temperature. The increased reactivity at higher temperatures results in an accelerated burning rate and reduced combustion duration [28]. When methanol is introduced into the cylinder during the combustion process, its remarkable characteristic of possessing a high latent heat of vaporization comes into play. This unique property of methanol has a significant impact on the combustion process by causing a cooling effect on the end mixture within the cylinder. As a result, the combustion duration is extended due to the cooling effect induced by methanol’s high latent heat of vaporization. This phenomenon is worth noting as it has implications for engine performance and efficiency. By understanding the cooling effect brought about by methanol’s high latent heat of vaporization, we can develop strategies to optimize the combustion process to achieve better performance, fuel economy, and reduced emissions.

Figure 5 provides the IMEP and COV_IMEP for various injection timings. Based on the trends indicated in the figure, as the injection timing is delayed, IMEP exhibits a tendency to increase and then decrease with the maximum value of IMEP occurring at the injection timing in the range of −180~−210° CA ATDC. However, COV_IMEP remains almost constant during this process.

At −180° CA ATDC, depending on the characteristics of the engine, the piston is at the bottom dead center. The direct injection of methanol in the cylinder at this point creates a larger coverage area. The more area the methanol covers, the less likely it is that the gasoline mixture will come into direct contact with the walls. At the same time, it is at the end of the intake stroke, and the stronger airflow movement in the cylinder during the latter part of the intake stroke aids in the atomization of the methanol. At this time, the direct injection of methanol produced by the atomization effect is better and can be more fully combusted. On the other hand, the tendency to knock can be suppressed more effectively. Knock can lead to faster combustion: during normal combustion, fuel burns evenly in the cylinder at a moderate rate. Knock will lead to a sudden acceleration of combustion speed so that the fuel cannot be completely burned, wasting part of the energy. In addition to this, this effect can lead to lowering the combustion temperature and reducing the amount of energy dissipated in the form of heat in the combustion chamber and cylinder walls, which is efficiently utilized. As for COV_IMEP, a stoichiometric ratio mixture was used in this experiment, so COV_IMEP was maintained at a relatively constant state. This careful balance of fuel and air ensures that the combustion process is optimized and efficient, resulting in consistent and predictable performance. By maintaining a stoichiometric mixture, the COV_IMEP remains relatively constant throughout the experiment. This is crucial because COV_IMEP is a vital indicator of engine performance and stability. Consistency in COV_IMEP values allows for an accurate comparison and analysis of different factors influencing engine operation, such as ignition timing, fuel composition, or engine modifications. Therefore, the decision to employ a stoichiometric ratio mixture in this experiment ensures a reliable and stable COV_IMEP measurement, enabling an accurate evaluation of the variables under investigation.

3.2. Effect of Injection Timing on Performance Characteristics

The graph in Figure 6 illustrates the relationship between BTE and injection timing. It reveals a pattern where, as the injection timing is delayed, the BTE initially increases and then starts to decrease. This can be attributed to the combustion characteristics within the engine. When the injection is delayed, fuel is injected closer to the end of the intake stroke. As a result, the fuel can be better atomized, resulting in a rise in BTE. However, with a further delay in injection time, the fuel has less time to mix with the air in the combustion chamber, and at the same time, the probability of methanol hitting the walls increases. This results in poorer air–fuel homogeneity and more efficient combustion, which leads to a decrease in BTE. If the injection timing is delayed excessively, the combustion process may not have sufficient time to complete before the piston starts its downward stroke. This can lead to incomplete combustion and a subsequent decrease in BTE. Therefore, the maximum value of BTE is observed within a specific injection timing range of −180~−150° CA ATDC, indicating the optimal timing for achieving the highest thermal efficiency in the engine.

Within this range, the combustion process is appropriately timed, allowing for efficient and complete fuel combustion. The fuel–air mixture is ignited, burned, and expanded to generate the maximum amount of useful work. Consequently, the engine exhibits the highest thermal efficiency at this particular injection timing range. By carefully adjusting the injection timing within the recommended range, we can achieve the desired balance between fuel efficiency and power output. This optimal timing ensures that the fuel is burned efficiently, maximizing the conversion of chemical energy into mechanical work. Therefore, choosing the suitable injection timing within the −180 to −150° CA ATDC range is critical for obtaining the highest level of thermal efficiency in the engine.

At various injection pressures, Figure 7 presents the effects of injection timing on EGT (exhaust gas temperature). At different direct injection pressures, EGT showed the same pattern. At each injection pressure condition, EGT rises and then falls as injection timing is delayed from −360° CA to −30° CA ATDC. The reason for this phenomenon corresponds to the experimental phenomenon shown earlier. EGT is associated with the phase of Pmax. Such experimental results are also certified in Figure 3. APmax is delayed first in advance. Earlier, APmax allowed sufficient cooling of the cylinder gases, which reduced the EGT [28]. The variations in injection timing and its impact on EGT provide valuable insights for engine optimization. By carefully selecting the injection timing within an appropriate range, we can effectively control EGT, thus enhancing the overall thermal efficiency and performance of the engine.

3.3. Effect of Injection Timing on Emission Characteristics

Under different injection pressures, Figure 8 illustrates how injection timing affects emissions. The three parts of the figure correspond to the trends of the three conventional emissions. The experimental results show that the emissions of CO, HC and NOx exhibit different patterns at different injection timing. From −330° CA to −30° CA ATDC, CO and HC show the same trend, decreasing and then increasing, corresponding to the above two components, respectively. This trend is maintained at different injection pressures. However, NOx showed the opposite trend. NOx increases and then decreases when injection timing is from −330° CA to −30° CA ATDC.

The observed phenomena can be attributed to the different conditions under which the engine produces CO, HC and NOx. Inadequate combustion leads to the production of CO and HC [29]. During the injection period from −330° CA to −60° CA ATDC, the cylinder volume becomes larger and then smaller during the corresponding injection. These changes bring about changes in the pressure being applied, which helps to improve the atomization of methanol and enhance its combustion, thereby reducing HC and CO. In addition, flame quenching of the combustion piston surfaces is a major source of emissions. Methanol tends to impact the piston as it approaches the top dead center, increasing the likelihood of flame quenching. On the other hand, NOx emissions are primarily influenced by factors such as cylinder temperature, oxygen concentration, and combustion duration [30]. Therefore, fuller in-cylinder combustion produces higher combustion temperatures and, according to Figure 4, the delayed combustion phase and the longer combustion duration result in longer reaction times at high temperatures. In summary, NOx emissions increase.

4. Conclusions

In this paper, dual-fuel gasoline/methanol experiments were conducted on a DFSI engine to explore how methanol injection timing affects engine combustion, performance, and emission characteristics to significantly optimize the performance of a dual-fuel engine at high loads. The main conclusions are as follows:

The timing of methanol direct injection significantly influences the combustion performance of a DFSI engine, particularly under high loads. Across various injection pressures, the effect of injection timing is consistently notable. Within the injection timing range of −330° to −30° CA ATDC, optimal combustion performance, indicated by the lowest values of parameters Pmax and HRRmax, is achieved between −210° and −150° CA ATDC. This corresponds to the latest phase of injection, leading to an extended combustion duration from CA10 to CA90. Additionally, CA50 and CA90 occur at the latest points in the combustion process. Emissions, specifically HC and CO, are minimized in the −210° to −150° CA ATDC range. However, NOx emissions are highest during this specific injection timing.

The combined approach of gasoline port injection and methanol direct injection proves to be a feasible method for optimizing gasoline/methanol dual-fuel engines. The full utilization of methanol direct injection is achievable, contributing to enhanced engine performance. Despite the advantages, this dual-fuel method is associated with higher levels of NOx and increased HC emissions. However, as the experiment utilized a stoichiometric ratio mixture, the impact on efficient exhaust gas treatment is deemed insignificant, thereby avoiding additional costs related to exhaust gas treatment.

The study highlights a trade-off between optimal combustion performance and emissions in dual-fuel engines. While the late-phase injection timing (−210° to −150° CA ATDC) maximizes combustion efficiency and minimizes HC and CO emissions, it concurrently results in elevated NOx emissions. Researchers and engineers need to carefully consider and balance these trade-offs based on specific application requirements and environmental considerations when designing or optimizing dual-fuel engines.