Effect of Methanol Injector Bore Arrangement on Combustion and Emissions in Dual-Fuel Engines

Abstract

The physical and chemical properties of methanol differ significantly from those of conventional diesel, and its injection strategy plays a critical role in engine performance. In this study, a three-dimensional simulation model of a methanol–diesel dual-fuel engine integrated with chemical reaction kinetics was developed using CONVERGE software. The effects of methanol injection position and angle on combustion characteristics, emission performance, and engine economy were systematically investigated through numerical simulation and theoretical analysis, leading to the optimization of the methanol injection strategy. By varying the distance between the methanol nozzle and the cylinder head as well as the methanol injection angle, changes in temperature, pressure, heat release rate (HRR), and other engine parameters were analyzed. Additionally, the impact on emissions, including soot, HC, CO, and NOx, was evaluated, providing a theoretical foundation for optimizing dual-fuel engine performance and enhancing methanol utilization efficiency. The results indicate that the methanol injection position minimally affects engine performance. When the methanol spray is positioned 3 mm from the cylinder head, it facilitates the formation of a homogeneous mixture, resulting in optimal power output and enhanced environmental performance. In contrast, the injection angle has a more pronounced effect on combustion and emission characteristics. At a methanol injection angle of 65°, the mixture homogeneity reaches its optimal level, leading to a significant enhancement in combustion efficiency and engine power performance. Excessive injection angles may lead to combustion deterioration and reduced engine performance. The primary reason is that an excessive spray angle may cause methanol spray to impinge on the cylinder wall. This leads to wall wetting, which adversely affects mixture formation and combustion.

1. Introduction

With the escalating global environmental challenges, ships, as the primary mode of maritime transportation, exert a substantial environmental impact due to the pollutants emitted by their power systems. Consequently, identifying sustainable alternative energy sources to mitigate the environmental impact of heavy oil consumption has emerged as a critical priority [1,2,3]. Methanol is recognized as a promising alternative fuel [4,5], attracting significant interest due to its wide combustion limits, high oxygen content, established production methods, cost-effectiveness, and widespread availability. Methanol fuel is predominantly utilized in international shipping vessels and has reached the stage of practical application [6]. For new ship construction, methanol propulsion systems offer lower investment costs compared to liquefied natural gas (LNG) systems. The primary methods of methanol application in engines include direct mixing, intake port injection, and direct in-cylinder injection. Direct mixing necessitates the use of costly additives, thereby elevating operational expenses. Intake port injection entails the introduction of a minimal quantity of diesel into the intake port to ignite the directly injected methanol. Direct in-cylinder injection employs high-pressure direct injection of both diesel and methanol, utilizing a small quantity of diesel as the pilot fuel. High-pressure direct injection dual-fuel engines demonstrate significant potential for emission reduction [7]. The use of diesel as the pilot fuel [8] enhances the combustion efficiency and overall performance of methanol–diesel dual-fuel engines. Yin [9] conducted a comparative analysis of intake port injection and direct injection methods, revealing that dual-nozzle combustion achieves a higher methanol substitution rate at the cost of increased NOx and HC emissions. Under full-load conditions [10], the maximum methanol substitution rate in a dual direct injection engine can reach 96.0%. The effects of high methanol substitution rates [11,12] on the combustion and emission characteristics of heavy-duty diesel engines remain insufficiently understood. The alignment of spray, airflow, and chamber design in diesel–methanol dual-fuel engines, along with the position [13] and angle of methanol injection, significantly influences combustion characteristics, mixture homogeneity, and exhaust emissions. Full-load operation is a prevalent scenario for marine engines, particularly in high-power-demand situations such as acceleration and sailing against currents. Optimizing combustion and emission characteristics during full-load operation is crucial for enhancing the efficiency and environmental performance of marine power systems. Cung et al. [14] conducted an experimental study on a dual-fuel combustion mode, revealing that under high-load conditions, an increase in the methanol substitution rate led to an increase in the maximum pressure rise rate. Optimizing combustion and emission characteristics during full-load operation is crucial for enhancing the efficiency and environmental performance of marine power systems. Sun et al. [15] investigated the combustion and emissions of a 95% methanol and 5% diesel mixture, and the present study further explores the impact of combustion and emissions under this specific methanol–diesel ratio.

While previous studies have investigated the use of methanol in dual-fuel engines, the distinct contribution of this research lies in its systematic examination of the effects of methanol injector bore arrangement—specifically, the injection position and angle—on combustion and emissions in a methanol–diesel dual-fuel engine. In contrast to prior research, which typically examines either injection strategies or combustion characteristics in isolation, this study integrates both aspects to offer a comprehensive understanding of how the physical configuration of the methanol injector impacts engine performance.

(1) Injection Position Optimization: Previous research has predominantly concentrated on the effects of methanol substitution rates and injection timing, while the specific influence of the longitudinal position of the methanol injector relative to the cylinder head remains underexplored. This study methodically varies the injector position to determine the optimal distance that facilitates uniform mixture formation and improves combustion efficiency.

(2) Injection Angle Optimization: Although some studies have explored the effects of injection angles, they frequently do so without considering the interaction with other parameters. This research investigates the interaction between the injection angle and the injector position, offering a more integrated perspective on how these parameters collectively affect combustion and emissions. This study determines an optimal injection angle of 65°, which markedly enhances mixture homogeneity and combustion efficiency.

(3) Comprehensive Emission Analysis: In contrast to many previous studies that concentrated on a single type of emission (e.g., NOx or soot), this research conducts a thorough analysis of multiple emissions (soot, NOx, HC, and CO) under various injection configurations. This integrated approach enables a more balanced optimization of engine performance and environmental impact.

By addressing these research gaps, this study provides novel insights into the optimization of methanol–diesel dual-fuel engines, advancing the development of more efficient and environmentally sustainable marine propulsion systems. The findings establish a theoretical foundation for the design and operation of dual-fuel engines, particularly in high-power-demand scenarios such as full-load operation.

2. Establishment and Validation of the Methanol–Diesel Dual-Fuel Engine Model

2.1. Basic Engine Parameters

To establish a three-dimensional simulation model of a methanol–diesel dual-fuel engine, it is essential to define the fundamental technical parameters and boundary conditions of the engine.

Table 1. Main technical parameters and boundary conditions of the diesel engine.

Name	Technical Parameters	Name	Technical Parameters
Model	4190ZLC-2	Total displacement (L)	23.82
Number of cylinders	4	Cylinder turbulence inch length (mm)	3
Average effective pressure (MPa)	11.09	Cylinder arrangement	In-line
Ignition sequence	1-3-4-2	Piston average speed (m/s)	7.0
Rotational speed (rpm)	1000	Initial cylinder turbulent flow energy (m2/s2)	18.375
Stroke (mm)	210	Compression ratio	14
Effective power (kW)	220	The moment the exhaust valve opens	58 °BBDC
Torque (N∙m)	2100	The moment the exhaust valve closes	After 56 °A top dead center
Connecting rod length (mm)	410	The moment when the intake valve starts	66 °B after top dead center
Bore (mm)	190	The moment the intake valve is closed	154 °ABDC

presents the primary technical parameters and boundary conditions of the diesel engine, which serve as the foundation for subsequent numerical simulations. The accuracy and reliability of the simulation model are ensured by precisely defining the engine geometry, operating conditions, and initial state.

2.2. Submodels Used

Combustion within the cylinder encompasses intricate physical processes, including spray breakup and atomization, fuel chemical reactions, heat and mass transfer, and airflow dynamics. Therefore, the selection of suitable models is critical for enhancing the accuracy of the simulation. The simulation sub-models are selected as shown in

Table 2. Submodels used in the simulation.

Technical Parameters	Name
Turbulence model	RNG κ-ε model [16]
Droplet crushing model	KH-RT model [17]
Droplet collision model	NTC collision model [18]
Wall heat transfer model	Han and Reitz model [19]
Combustion model	Sage model [20]
Evaporation model	Frossling model [21]
NOx emission model	Zeldovich NOx model [22]
Soot emission model	Hiroyasu model [23]

.

2.2.1. Turbulence Model

The internal turbulence within the combustion chamber is characterized by highly complex turbulent flow dynamics. To close the governing equations, the corresponding turbulent diffusion terms in the Reynolds stress term and the energy equation must be incorporated into the turbulence model. The RNG k-ε model, a two-equation turbulence model based on viscous flow, incorporates the swirling effects of the fluid in the mean flow and extends the standard k-ε model by including a low-Reynolds number viscous flow analytical equation. Owing to its enhanced accuracy in turbulence calculations, the RNG k-ε model is extensively utilized in the combustion simulation of internal combustion engines [24,25].

The RNG k-ε model computes the turbulent kinetic energy (k) and the dissipation rate (ε) through two transport equations, expressed as follows:

$${\displaystyle \frac{\partial }{{\partial }_t}}\left(\rho \mathrm{k}\right)+{\displaystyle \frac{\partial }{{\partial }_{{\mathrm{x}}_{\mathrm{j}}}}}\left[\rho \mathrm{k}{\mathsf{\mu }}_{\mathrm{j}}-\left(\mathsf{\mu }+{\displaystyle \frac{{\mathsf{\mu }}_{\mathrm{t}}}{{\mathsf{\sigma }}_{\mathrm{k}}}}\right){\displaystyle \frac{{\partial }_{\mathrm{k}}}{{\partial }_{{\mathrm{x}}_{\mathrm{j}}}}}\right]={\mathsf{\mu }}_{\mathrm{t}}\left({\mathrm{P}}_{\mathrm{B}}+\mathrm{P}\right)-\mathsf{\rho }\mathsf{\epsilon }-{\displaystyle \frac{2}{3}}\left({\displaystyle \frac{{\partial }_{\mathrm{i}}}{{\partial }_{{\mathrm{x}}_{\mathrm{j}}}}}+\rho \mathrm{k}\right){\displaystyle \frac{{\mathrm{D}}_{\mathrm{k}}}{{\mathrm{D}}_{\mathrm{t}}}}$$

(1)

The turbulent energy dispersion rate (ε) equation is:

$$\begin{array}{c}{\displaystyle \frac{\partial }{{\partial }_t}}\left(\rho \mathsf{\epsilon }\right)+{\displaystyle \frac{\partial }{{\partial }_{{\mathrm{x}}_{\mathrm{j}}}}}\left[\rho \mathsf{\epsilon }{\mathsf{\mu }}_{\mathrm{j}}-\left(\mathsf{\mu }+{\displaystyle \frac{{\mathsf{\mu }}_{\mathrm{t}}}{{\mathsf{\sigma }}_{\mathrm{k}}}}\right){\displaystyle \frac{{\partial }_{\mathrm{k}}}{{\partial }_{{\mathrm{x}}_{\mathrm{j}}}}}\right]\hfill \\ ={\mathrm{C}}_{\mathsf{\epsilon }1}{\displaystyle \frac{\mathsf{\epsilon }}{\mathrm{k}}}\left[{\mathsf{\mu }}_{\mathrm{r}}\mathrm{P}-{\displaystyle \frac{2}{3}}\left({\mathsf{\mu }}_{\mathrm{r}}{\displaystyle \frac{{\partial \mathsf{\mu }}_{\mathrm{i}}}{{\partial \mathrm{x}}_{\mathrm{i}}}}+\rho \mathrm{k}\right){\displaystyle \frac{{\partial \mathsf{\mu }}_{\mathrm{i}}}{{\partial \mathrm{x}}_{\mathrm{i}}}}\right]+{\mathrm{C}}_{\mathsf{\epsilon }1}{\displaystyle \frac{\mathsf{\epsilon }}{\mathrm{k}}}{\mathsf{\mu }}_{\mathrm{r}}{\mathrm{P}}_{\mathrm{B}}-{\mathrm{C}}_{\mathsf{\epsilon }2}\rho {\displaystyle \frac{{\epsilon }^2}{\mathrm{k}}}\hfill \\ +{\mathrm{C}}_{\mathsf{\epsilon }4}\rho \epsilon {\displaystyle \frac{{\partial \mathsf{\mu }}_{\mathrm{i}}}{{\partial \mathrm{x}}_{\mathrm{i}}}}-{\displaystyle \frac{C_{\mu }{\eta }^3\left(1-\eta /{\eta }_0\right)}{1+\beta {\eta }^3}}\rho {\displaystyle \frac{{\partial }_{\mathsf{\epsilon }}}{{\partial }_{{\mathrm{x}}_{\mathrm{j}}}}}\hfill \end{array}$$

(2)

The turbulence ratio rate (ζ) equation is:

$$\eta =\mathrm{S}{\displaystyle \frac{k}{\epsilon }}$$

(3)

$$\mathrm{p}={\mathrm{S}}_{\mathrm{i}\mathrm{j}}{\displaystyle \frac{{\partial }_{{\mathsf{\mu }}_{\mathrm{i}}}}{{\partial }_{{\mathrm{x}}_{\mathrm{j}}}}}$$

(4)

$${\mathrm{P}}_{\mathrm{B}}=-{\displaystyle \frac{{\mathrm{g}}_{\mathrm{i}}}{{\sigma }_{\mathrm{h}}}}{\displaystyle \frac{{\partial }_{\rho }}{{\rho \partial \mathrm{x}}_{\mathrm{i}}}}$$

(5)

where from-turbulent viscosity is in ${\mathrm{m}}^2/\mathrm{s}$;

${\mathrm{C}}_{\mathsf{\epsilon }1}$, ${\mathrm{C}}_{\mathsf{\epsilon }2}$, ${\mathrm{C}}_{\mathsf{\epsilon }3}$, ${\mathrm{C}}_{\mathsf{\epsilon }4}$ are empirical constants; S is the source term; P is the turbulent kinetic energy generation term.

2.2.2. Droplet Crushing Model

In this study, the KH-RT fuel crushing model [26] is employed, which is well suited for analyzing and simulating high-speed droplet breakup during diesel injection. This process can be divided into two stages: first, the fuel is ejected from the nozzle to form a liquid column, which then undergoes primary breakup (KH breakup) to form droplets. Subsequently, these droplets are subjected to secondary breakup (RT breakup) for further fragmentation.

In the KH model, droplet breakup is described based on the surface wave growth theory. The radius of the droplet after breakup is given by:

$$\mathrm{r}={\mathrm{B}}_0{\mathit{\Lambda }}_{KH}$$

(6)

where ${\mathrm{B}}_0$ is the model constant, which is 0.61 in this case.

The rate of change of the droplet radius is expressed as:

$${\displaystyle \frac{d{r}_0}{dt}}={\displaystyle \frac{\left(r_0-r\right)}{{\tau }_{KH}}},\left(r<r_0\right)$$

(7)

where ${\tau }_{KH}$ is the crushing time, which can be expressed as:

$$\mathrm{r}{\tau }_{KH}={\displaystyle \frac{3.726{\mathrm{B}}_{1}r}{{\mathit{\Omega }}_{KH}{\mathit{\Lambda }}_{KH}}}$$

(8)

where ${\mathit{\Omega }}_{KH}{\mathit{\Lambda }}_{KH}$ is the product of frequency and amplitude.

${\mathrm{B}}_1$ is the crushing time constant, which is related to the initial disturbance level of the spray liquid jet.

When breakage occurs, the child droplets are added to the calculation accordingly, and they are given a velocity component in the normal direction of the parent droplet velocity, which can be expressed as:

$$v_n=C_1{\mathit{\Omega }}_{KH}{\mathit{\Lambda }}_{KH}$$

(9)

where ${\mathrm{C}}_1$ is the model constant.

To improve the accuracy of spray predictions 155 under conditions of high density and high relative velocity, a critical coefficient, 156 known as the KH crushing length, is introduced. This coefficient marks the transition 157 from KH crushing to RT crushing.

$${\mathrm{L}}_{\mathrm{b}}=C_b{d}_0\sqrt{{\displaystyle \frac{{\rho }_{fuel}}{{\rho }_{air}}}}$$

(10)

where $C_b$ is the model constant.

The KH crushing equation is applicable under high-density and high-relative-velocity conditions, whereas RT crushing occurs due to the rapid deceleration of droplets, leading to the development of surface waves at the stagnation point [27]. The RT perturbation process is typically characterized by the frequency of the rapidly growing wave Ω and its corresponding wavenumber K.

$${\mathit{\Lambda }}_{KH}={\displaystyle \frac{{2\pi \mathit{\Omega }}_{RT}}{K_{RT}}}$$

(11)

$${\mathit{\Omega }}_{RT}={\left\{{\displaystyle \frac{2}{3\sqrt{3}\sigma }}{\displaystyle \frac{{\left[-\left(\overrightarrow{g}-\overrightarrow{a}\right)\left({\rho }_f-{\rho }_g\right)\right]}^{\raisebox{1ex}{$3$}\!\left/ \!\raisebox{-1ex}{$2$}\right.}}{{\rho }_f+{\rho }_g}}\right\}}^{{\displaystyle \frac{1}{2}}}$$

(12)

$$K_{RT}={\left[{\displaystyle \frac{-\left(\overrightarrow{g}-\overrightarrow{a}\right)\left({\rho }_f-{\rho }_g\right)}{3\sigma }}\right]}^{1/2}$$

(13)

where $\overrightarrow{g}$ is the acceleration of the Earth’s gravity and $\overrightarrow{a}$ is the acceleration of the resistance experienced in the direction of motion.

The surface wave of the droplet arises from the interaction between the droplet and the surrounding air. When the wave amplitude exceeds the droplet diameter, the droplet undergoes fragmentation. The radius at which smaller droplets are generated is

$$r={\displaystyle \frac{{\pi C}_{RT}}{K_{RT}}}$$

(14)

where $C_{RT}$ is the model constant; the smaller its value, the smaller the droplet radius.

2.2.3. Droplet Collision Model

Two evaporation models are frequently employed in CONVERGE: the Frossling model and the Chiang model. The primary distinction between these models lies in their respective formulations for calculating the Nusselt (Nu) and Sherwood (Sh) numbers of the droplets.

2.3. Model Validation

The simulation process begins at the closure of the intake valve and concludes at the opening of the exhaust valve. The main parameters of the diesel engine are given in

Table 3. Inject parameter.

Technical Parameters	Name
Number of nozzles (pcs)	8
Nozzle diameter (mm)	0.26
Mass (mg)	194
Pressure (bar)	20
Angle (°)	75
SOI/°CA ATDC	20
Mass (mg)	190

and the initial conditions in

Table 4. Initial conditions.

Technical Parameters	Name
Initial temperature (K)	335.15
Initial pressure (MPa)	1.93
Piston head temperature (K)	625.15
Cylinder head temperature (K)	553.15
Inner wall temperature (K)	403.15
Workload	100%

. Following the methodology of Zhang [28], numerical simulations of the diesel engine under full-load conditions were performed in this study. Figure 1 a shows the 3D map of the engine and (b) shows the 3D mesh map.

As shown in Figure 2, under full-load conditions, the predicted expansion pressure curve in the cylinder slightly exceeds the experimental curve after the peak pressure point, with a deviation within 5%. The discrepancies between the simulation results and experimental data primarily arise from wall heat transfer losses, simplifications in the turbulence model, and idealized assumptions in chemical reaction kinetics. Although these discrepancies have a minor effect on combustion and emission characteristics under full-load operation, future research could mitigate these errors by incorporating more precise heat transfer and turbulence models. The simulation results for cylinder pressure and the ignition process align closely with the experimental data reported by Zhang [28], demonstrating the model’s capability to accurately predict the combustion heat release process in the diesel engine.

2.4. Validation of Methanol Combustion Mechanism

Validation of the selected methanol combustion mechanism, comprising 53 species and 176 reactions [29], is essential to ensure its accuracy. The validation of methanol–diesel combustion was performed at a methanol substitution rate of 20%.

As illustrated in Figure 3, at a methanol substitution rate of 20%, the predicted cylinder pressure and ignition process show strong agreement with the experimental data. During the cylinder compression phase, the simulated values are marginally higher than the experimental values. At an approximately −2° crank angle, the simulated values are slightly lower than the experimental values. This discrepancy primarily stems from simplifications in the simulation model, including combustion delay, spray breakup and evaporation, turbulence intensity, wall heat transfer, and chemical reaction kinetics. These factors exhibit greater complexity in experimental conditions, leading to a marginally faster combustion process compared to the simulated results. Future research could further minimize the discrepancies between simulation and experimental results by incorporating more accurate spray models, turbulence models, and chemical reaction kinetics models. The peaks in the simulation are in good agreement with the experimental data of Xuan [30], and the subsequent work process exhibits nearly identical behavior. These findings demonstrate that the model effectively predicts the combustion heat release process in the diesel engine. Particularly under high-temperature and high-pressure conditions, the mechanism accurately captures the oxidation reaction pathways and radical generation processes of methanol.

2.5. Grid Independence Validation

CONVERGE efficiently addresses dynamic grid processing challenges and enables automatic grid generation, thereby enhancing computational efficiency and reducing grid processing time. Additionally, the self-generated high-quality grids ensure high computational accuracy. They support fixed refinement in specific regions and adaptive refinement based on temperature, velocity, and component gradients. Grid independence analysis was performed to account for the complex physical and chemical processes occurring within the combustion chamber. The validation was carried out under the specified conditions, using base grid sizes of 4 mm, 6 mm, 8 mm, and 10 mm. Adaptive mesh refinement (AMR) was employed to enhance the resolution of temperature and velocity fields. Three levels of fixed embedding were applied to the spray, piston, and cylinder wall regions. The impact of varying base grid sizes on cylinder pressure is illustrated in Figure 4.

The cylinder pressures obtained from the four base grids exhibit strong agreement, with minor deviations observed at the peak. The cylinder pressure results for the 4 mm base grid are nearly identical to those for the 6 mm grid. However, as can be seen from

Table 5. Calculation time required for different base grids.

Base Grid Size	4 mm	6 mm	8 mm	10 mm
Time required (h)	100	46	25	14

, the computational time required for a 6 mm grid is only 50% of that required for a 4 mm grid, as indicated by the calculation times for different base grids. Based on a balance between simulation accuracy and computational efficiency, the 6 mm mesh was chosen as the base grid. In subsequent studies, the maximum adaptive mesh refinement (AMR) embedding level was set to 3, yielding a minimum mesh size of 0.75 mm in the simulations.

3. Effect of Methanol Nozzle Position on Engine Combustion Characteristics

3.1. Effect of Methanol Nozzle Longitudinal Position on Combustion

The injection position is determined by the nozzle location, specified as the longitudinal distance h from the cylinder head, as illustrated in Figure 5. The initial parameters of the methanol nozzle and the diesel nozzle are shown in

Table 6. The initial parameters of the methanol nozzle and the diesel nozzle.

Technical Parameters	Data
Diesel injection angle (°CA)	−10
Methanol injection angle (°CA)	−9.5
Duration of diesel injection (°CA)	0.9
Duration of methanol injection (°CA)	9.5
Diesel nozzle pressure (bar)	190
Methanol nozzle pressure (bar)	600

. The computational scheme for methanol injection positions is detailed in

Table 7. Position of methanol injection holes.

Nozzle Parameters	Z–0.002	Z–0.003	Z–0.004
Nozzle to cylinder head distance (m)	0.002	0.003	0.004

.

The average cylinder pressure and heat release rate (HRR) curves are presented in Figure 6. Following an ignition delay of approximately −1 °CA, the diesel fuel undergoes auto-ignition. During the injection process, the diesel fuel experiences shear forces, which entrain the surrounding mixture. This promotes further mixing with the in-cylinder gas during atomization and evaporation. The initial stage of combustion is predominantly driven by diesel combustion, characterized by a rapid heat release rate, as evidenced by the HRR curve. However, methanol near the flame kernel ignites rapidly. The interaction between diesel’s diffusion combustion and methanol’s premixed combustion significantly enhances the combustion rate. The hydroxyl groups in methanol produce a significant quantity of hydroxyl radicals, thereby accelerating the oxidation process of methanol. As Z-0.004 is positioned deeper within the engine cylinder compared to Z-0.003 and Z-0.002, the injected methanol is closer to the pilot flame. Consequently, combustion initiates earlier at Z-0.004, and the reaction becomes highly intense during the subsequent combustion process. As the pilot flame propagates within the cylinder, the combustion rate progressively increases. The heat release rate stabilizes until 7 °CA after top dead center (TDC), at which point combustion is largely complete. At 12 °CA after TDC, the HRR attains its maximum value, after which the curve declines steadily and eventually approaches zero. Figure 6 illustrates the cylinder pressure and HRR curves for different methanol nozzle positions. As the nozzle position moves downward, the data show that the cylinder pressure and HRR initially increase and subsequently decrease. At a nozzle position of 3 mm, the cylinder pressure and HRR achieve their peak values, indicating optimal combustion efficiency at this configuration.

As shown in Figure 7, the average in-cylinder temperature under the three conditions is nearly identical. The longitudinal nozzle distance exhibits minimal influence on the average temperature before top dead center (TDC). As the piston descends, the average temperature associated with different longitudinal nozzle distances progressively rises. After TDC, the average temperature slightly rises with increasing longitudinal nozzle distance, resulting in higher exhaust losses. The increased heat loss reduces thermal efficiency, thereby increasing elevated carbon emissions. At a longitudinal nozzle distance of 3 mm, the temperature reaches a higher level. The variation in cylinder pressure before TDC, driven by piston compression and diesel auto-ignition, is influenced by the engine’s compression ratio and diesel ignition characteristics. At the end of the compression stroke, the formation of a low-pressure zone facilitates rapid evaporation of the methanol spray, promoting the formation of a homogeneous mixture after TDC. Simultaneously, a greater amount of methanol from a longer longitudinal nozzle distance impinges on the piston wall, enhancing spray breakup and fostering the formation of a uniform mixture.

From Figure 8, as the methanol nozzle position is adjusted downward, the ignition delay (CA10) initially decreases and then increases, with the combustion center (CA50) and combustion end (CA90) following a similar pattern. This phenomenon is attributed to the gradual shift in the fuel-rich zone from the piston wall to the piston bowl, enhancing fuel–air mixing and thereby improving combustion efficiency. However, as the nozzle position is moved further downward, the spray penetration distance increases, leading to fuel impingement on the piston wall and subsequent cooling, which reduces mixture uniformity and combustion efficiency. Additionally, the downward adjustment of the methanol nozzle position diminishes the interaction between methanol and diesel sprays. The methanol absorbs the high latent heat during vaporization, initially reducing the cylinder temperature and prolonging the ignition delay. This shifts the combustion process backward, shortening the combustion duration and concentrating heat release. At a methanol injection distance of 3 mm, the interaction between the spray and in-cylinder airflow is optimized, resulting in a uniform mixture. This phenomenon can be attributed to the conservation of momentum and the evaporation rate of the spray. A shorter injection distance reduces spray penetration depth, preventing fuel impingement on the cylinder wall and thereby enhancing mixture uniformity. The 3 mm injection distance corresponds to the earliest CA10 and CA50 as well as the longest CA10-CA90 duration. This advances the combustion phase, extends the combustion time, and increases the temperature during the expansion stroke.

Table 8. The effect of the position of the methanol nozzle on the temperature field in the cylinder.

Show the Moment	−5 (°CA)	−2 (°CA)	0 (°CA)	1 (°CA)	2 (°CA)	5 (°CA)
Temperature
0.002
0.003
0.004

illustrates the influence of the longitudinal position of the methanol nozzle, defined as the distance between the nozzle and the cylinder head, on the temperature field within the cylinder. The table demonstrates the impact of the methanol nozzle position on the combustion process by presenting the temperature distribution at various crankshaft angles (−5 °CA, −2 °CA, 0 °CA, 1 °CA, 2 °CA, 5 °CA). The table includes three longitudinal nozzle positions, Z-0.002 m, Z-0.003 m, and Z-0.004 m, representing the distance between the nozzle and the cylinder head.

During the compression phase, the in-cylinder temperature progressively rises due to piston compression. Different nozzle positions exhibit minimal impact on the temperature field, resulting in a relatively uniform temperature distribution. As the nozzle position is adjusted downward, the temperature increases marginally, though the difference is not substantial. At the ignition stage, spontaneous combustion of diesel fuel and methanol combustion commence, leading to a rapid increase in temperature. The temperature field at the Z-0.003 m position exhibits elevated temperatures, indicating higher combustion efficiency. The temperature field at the Z-0.004 m position also displays higher temperatures. However, as the methanol spray penetrates deeper into the cylinder, some fuel may impinge on the piston wall, causing localized temperature non-uniformity.

During the expansion phase, combustion persists, with the temperature peaking before gradually declining. The temperature field at the Z-0.003 m position reaches a higher peak temperature, signifying more intense combustion and enhanced combustion efficiency at this location. The temperature field at the Z-0.004 m position exhibits lower temperatures, likely due to fuel impingement on the piston wall, leading to incomplete combustion. At the Z-0.002 m position, the nozzle is positioned closer to the cylinder head, resulting in a shallower methanol spray penetration depth and relatively uniform fuel–air mixing. However, the short spray distance may prevent complete combustion of some fuel, leading to a more uniform temperature field distribution but a lower peak temperature. At the Z-0.003 m position, the nozzle distance is optimal, allowing the methanol spray to fully mix with the air. This results in high combustion efficiency and a temperature field with a higher peak temperature, indicating a more intense combustion process. At the Z-0.004 m position, the nozzle is positioned deeper, resulting in greater methanol spray penetration. This may cause some fuel to impinge on the piston wall, leading to localized temperature non-uniformity, reduced combustion efficiency, and a lower peak temperature in the temperature field.

In summary, a methanol injection position of 3 mm from the cylinder head yields the most favorable combustion performance.

3.2. Effect of Methanol Nozzle Longitudinal Position on Emissions

Figure 9, Figure 10, Figure 11 and Figure 12 investigate the impact of the longitudinal position of the methanol nozzle on the emissions of soot, NOx, HC, and CO. The results demonstrate that the longitudinal position of the methanol spray significantly influences the emission characteristics. At the Z-0.003 position, the emissions of soot and NOx reach their peak, primarily due to the proximity of the methanol spray to the ignition flame, which intensifies the combustion process, leading to increased localized high-temperature and oxygen-deficient regions that promote the formation of soot and NOx. In contrast, at the Z-0.004 position, the deeper penetration of the methanol spray into the cylinder enhances the interaction between the spray and airflow, resulting in a more uniform distribution of the air–fuel mixture. This reduces localized high-temperature and oxygen-deficient regions, thereby decreasing the emissions of soot and NOx. HC emissions are higher at the Z-0.002 position, mainly due to the weaker interaction between the spray and airflow, which leads to an uneven distribution of the air–fuel mixture. As a result, some methanol fails to combust completely, forming unburned HC. As the nozzle position moves downward, HC emissions initially increase and then decrease, indicating that an optimal nozzle position can improve the uniformity of the air–fuel mixture and reduce the formation of unburned HC. The variation trend of CO emissions is similar to that of soot and NOx, reaching its peak at the Z-0.003 position, primarily due to the increased generation of incomplete combustion products caused by localized high-temperature and oxygen-deficient regions. At the Z-0.004 position, the more uniform distribution of the air–fuel mixture enhances combustion efficiency, leading to a reduction in CO emissions. In summary, the longitudinal position of the methanol nozzle significantly influences the emission characteristics of the engine. An optimal nozzle position can effectively improve the uniformity of the air–fuel mixture, thereby reducing the generation of emissions, particularly CO and HC.

These findings align with previous studies in the literature. For instance, Zhang [28] demonstrated that soot formation is dependent on high-temperature and oxygen-deficient conditions, whereas Yin [10] indicated that NOx generation is highly sensitive to temperature, with elevated temperatures significantly enhancing NOx emissions. Furthermore, Shi [11] revealed that HC emissions are strongly correlated with the uniformity of the air–fuel mixture and combustion efficiency. The results of this study corroborate these findings and indicate that the non-uniform distribution of methanol spray contributes to localized high temperatures, oxygen-deficient zones, and unburned regions, thereby influencing emission characteristics.

In conclusion, the influence of methanol injection position on combustion is relatively limited. However, an optimal methanol injection position enhances mixture uniformity, consequently reducing overall exhaust emissions. Under full-load conditions, although emissions are not minimized, the methanol nozzle position at h = 3 mm is selected for subsequent research to achieve enhanced power output.

4. Effect of Methanol Injection Angle on Engine Combustion and Emission Characteristics

4.1. Effect of Methanol Spray Angle at 60−67.5° on Combustion

Based on the aforementioned results, the methanol nozzle was positioned at a vertical distance of 3 mm from the cylinder head, and the methanol injection angle was systematically varied. Yang [31] demonstrated that the combustion efficiency of methanol spray improves within the range of 60° to 75°. Consequently, the methanol injection angle calculation scheme is detailed in

Table 9. Methanol nozzle angle experimental scheme.

Methanol Injection Angle
XZ Direction (°)	60	62.5	65	67.5

.

Figure 13, Figure 14 and Figure 15 present the pressure, temperature, and heat release rate (HRR) curves for four conditions, with the methanol nozzle positioned at a longitudinal distance of 3 mm from the cylinder head. The average temperature remains nearly constant during the compression phase and the initial combustion phase. As the methanol injection angle decreases from 67.5° to 60°, the fuel-rich zone of the methanol spray shifts from the cylinder wall toward the piston wall, enhancing spray dispersion. As evident from the figures, the HRR advances progressively, the maximum cylinder pressure increases steadily, and the peak temperature rises, resulting in more intense combustion. This phenomenon occurs because a reduced methanol injection angle enhances the interaction between methanol and pilot diesel, facilitating the mixing of methanol with high-temperature air. As the methanol spray angle is further reduced, the HRR advances, but the cylinder pressure and temperature decline. A smaller methanol spray angle increases the spray penetration distance, causing the concentrated fuel at the spray front to reach the piston bowl first. This results in significant fuel impingement on the piston wall, leading to cooling and reduced mixture uniformity in the piston bowl, thereby degrading combustion performance.

Table 10. The effect of the angle of the methanol nozzle on the temperature field in the cylinder.

Show the Moment	−5 (°CA)	−2 (°CA)	0 (°CA)	1 (°CA)	2 (°CA)	5 (°CA)
Temperature
60°
62.5°
65°
67.5°

illustrates the variation in the temperature field within the cylinder at different methanol injection angles (60°, 62.5°, 65°, and 67.5°). The table presents the temperature distribution at various crankshaft angles (−5 °CA, −2 °CA, 0 °CA, 1 °CA, 2 °CA, and 5 °CA). By comparing the temperature distributions across different injection angles, it is evident that the methanol injection angle significantly influences the temperature field within the cylinder. Specifically, at a 60° injection angle, the methanol spray is predominantly concentrated near the piston wall, leading to elevated local temperatures in the combustion chamber. However, the overall temperature distribution is non-uniform. At the 62.5° injection angle, the methanol spray distribution is more uniform, resulting in a more consistent temperature distribution within the combustion chamber and a more stable combustion process. At a 65° injection angle, the interaction between methanol and diesel sprays is optimized, leading to a uniform temperature distribution within the combustion chamber. This results in the most intense combustion process and the highest peak temperature. At the 67.5° spray angle, the methanol spray distribution is relatively uniform. However, the larger spray angle causes some methanol to concentrate in the upper part of the cylinder, leading to slight fluctuations in the temperature distribution within the combustion chamber.

As depicted in Figure 16, as the methanol spray angle increases, CA10 advances slightly, while the combustion center (CA50) and combustion end (CA90) initially advance and then retard. However, the extent of these shifts varies, with CA90 exhibiting the most significant change. At a methanol spray angle of 65°, the rapid combustion rate of methanol results in a substantial amount of methanol–air mixture being consumed during the premixed combustion phase. This significantly reduces the combustion duration, increases pressure and temperature, and enhances the engine’s power output. Although the 65° injection angle yields optimal performance under full-load conditions, adjustments may be necessary under partial-load or variable operating conditions to accommodate different combustion requirements. Future research should focus on developing injection angle optimization strategies for variable operating conditions.

The literature cited in this study is directly relevant to the research content and substantiates the conclusions drawn. Yang [31] offered theoretical insights into the effect of methanol spray angle on cylinder pressure. Sun [15] validated the relationship between heat release rate (HRR) and methanol spray angle. Wen [13] demonstrated the correlation between in-cylinder temperature and methanol spray angle. Therefore, the citations in this study are well-justified and align with the findings of the referenced literature. Future research should investigate the optimal injection angle under varying operating conditions, such as high and low loads, to enhance combustion efficiency and environmental performance.

4.2. Effect of Methanol Spray Angle at 60−67.5° on Emissions

Figure 17 illustrates the in-cylinder distributions of soot, NOx, HC, and CO under full-load conditions, with the methanol nozzle positioned at a longitudinal distance h of 3 mm. As shown in the figure, soot, NOx, and CO emissions initially increase and subsequently decrease as the methanol injection angle increases. At an injection angle of 62.5°, CO emissions are comparable to those at 65°, likely due to improved mixture uniformity and higher combustion efficiency at 62.5°, which reduces the formation of incomplete combustion products such as CO. However, as the injection angle increases further, the impingement of the spray on the cylinder wall becomes more pronounced, resulting in combustion degradation and elevated CO emissions. In contrast, HC emissions initially decrease and then increase with increasing methanol spray angle.

In dual-fuel engines, soot emissions are primarily influenced by diesel combustion under high-temperature, oxygen-deficient conditions. However, during diesel combustion, oxygen depletion in the core combustion region suppresses NOx generation and reduces soot production. Since the rate of NO formation is slower than the fuel combustion reaction rate, NO generation necessitates sufficient high-temperature residence time. Insufficient high-temperature residence time reduces NO formation, preventing it from reaching equilibrium concentrations. Following diesel ignition, soot and NOx formation commence, with soot being generated at a higher rate. Subsequently, methanol participates in combustion, which is characterized by higher oxygen content and extended high-temperature residence time. This promotes NOx formation in high-temperature regions, while soot is gradually oxidized, except for a small fraction remaining in the central cylinder area where flame propagation is incomplete.

In summary, a methanol nozzle angle of 65° enhances mixture uniformity and reduces HC emissions, significantly improving the engine’s combustion rate and power output. This results in higher cylinder pressure, elevated temperatures, and optimal power performance.

5. Discussion

In summary, this study comprehensively investigates the effects of methanol injection position and angle on the combustion and emission characteristics of dual-fuel engines using numerical simulations and theoretical analysis. The results demonstrate that a methanol nozzle positioned at a longitudinal distance of 0.003 m from the cylinder head, with an injection angle of 65°, achieves optimal combustion efficiency and environmental performance. Future research should focus on further optimizing methanol injection strategies, investigating the combined use of methanol with other alternative fuels, and developing adaptive injection strategies for variable operating conditions to enhance the efficiency and environmental performance of dual-fuel engines.

6. Conclusions

This paper investigates the combustion and emission characteristics of a methanol–diesel dual-fuel engine. We developed the model using CONVERGE 3.0 software, and through numerical simulations, the methanol longitudinal nozzle position and spray injection angle were optimized. Based on the simulation outcomes, the following conclusions are derived:

• An optimal increase in the methanol longitudinal injection position h enhances engine performance. Among the tested methanol longitudinal positions, Z-0.003 exhibits superior combustion characteristics, including the earliest CA10 and CA50, as well as the longest CA10-CA90 duration. This results in higher combustion temperatures and increased heat release rates (HRRs). Although the soot level at Z-0.003 is marginally higher than at Z-0.004, the HC emissions are the lowest, indicating more complete combustion.
• Four spray angles—60°, 62.5°, 65°, and 67.5°—were investigated. At a 65° spray angle, the methanol spray achieves optimal mixing and reaction with the pilot fuel combustion mixture. This significantly reduces the combustion duration, increases cylinder pressure and temperature, and enhances the engine’s power output. Excessive spray angles can degrade combustion and reduce power performance. This is primarily due to methanol spray impingement on the cylinder wall, causing wall wetting and adversely affecting mixture formation and combustion. Additionally, the 65° spray angle results in the lowest HC emissions, indicating more complete combustion and superior power performance.