Numerical and Experimental Research on the Effects of Hydrogen Injection Timing on the Performance of Hydrogen/N-Butanol Dual-Fuel Engine with Hydrogen Direct Injection

Abstract

Hydrogen injection timing (HIT) plays a crucial role in the combustion and emission characteristics of a hydrogen/n-butanol dual-fuel engine with hydrogen direct injection. This study employed an integrated approach combining three-dimensional simulation modeling and engine test bench experiments to investigate the effects of HIT on engine performance. In order to have a more intuitive understanding of the physical and chemical change processes, such as the stratification state and combustion status of hydrogen in the cylinder, and to essentially explore the internal mechanism and fundamental reasons for the improvement in performance of n-butanol engines by hydrogen addition, a numerical study was conducted to examine the effects of HIT on hydrogen stratification and combustion behavior. The simulation results demonstrated that within the investigated range, at 100 °CA BTDC hydrogen injection time, hydrogen forms an ideal hydrogen stratification state in the cylinder; that is, a locally enriched hydrogen zone near the spark plug, while there is a certain distribution of hydrogen in the cylinder. Meanwhile, the combustion state also reaches the optimal level at this hydrogen injection moment. Consequently, 100 °CA BTDC is identified as the optimal HIT for a hydrogen/n-butanol dual-fuel engine. At the same time, an experimental study was performed to capture the actual complex processes and comprehensively evaluate combustion and emission characteristics. The experimental results indicate that both dynamic performance (torque) and combustion characteristics (cylinder pressure, flame development period, etc.) achieve optimal values at the HIT of 100 °CA BTDC. Notably, under lean-burn conditions, the combustion parameters exhibit greater sensitivity to HIT. Regarding emissions, the CO and HC emissions initially decreased slightly, then gradually increased with advanced injection timing. The 100 °CA BTDC timing effectively reduced the CO emissions at λ = 0.9 and 1.0. CO emissions at λ = 1.2, and showed minimal sensitivity to the injection timing variations. Therefore, optimized HIT facilitates enhanced combustion efficiency and emission performance in hydrogen-direct-injection n-butanol engines.

1. Introduction

Facing energy shortages, environmental pollution, and increasingly stringent emission regulations, the pursuit of renewable and clean alternative energy sources has become a critical pathway to improve energy structures, alleviate the energy crisis, promote green development, and achieve the strategic goals of “carbon peak and carbon neutrality” [1,2]. Among alternative fuels, alcohol-based fuels—as renewable oxygenated biofuels—have garnered extensive attention due to their wide availability and clean combustion characteristics [3,4]. They are widely regarded as one of the most promising alternative fuels for spark-ignition engines [5]. Compared to methanol and ethanol, n-butanol exhibits a higher lower-heating value, lower latent heat of vaporization, and higher energy density [6]. Additionally, n-butanol is less water-soluble and less corrosive, demonstrating superior compatibility with existing power equipment, and it does not generate nitrogen oxide or sulfur compounds after combustion [7]. Furthermore, advancements in production technology enable n-butanol to be produced via bio-fermentation using cost-effective raw materials such as Scenedesmus obliquusalgae [8], wild plants, cereal corn [9], and lignocellulose [10]. These low-cost production methods strengthen the feasibility of n-butanol as an engine fuel. Consequently, research on n-butanol as an engine fuel has attracted widespread global scholarly interest and is recognized as one of the most promising alternative fuels for spark-ignition engines [11].

Extensive research has confirmed the technical feasibility of utilizing n-butanol as an alternative fuel for gasoline engines [12]. However, compared to gasoline, n-butanol exhibits inferior evaporation characteristics and a higher latent heat of vaporization [13]. Consequently, engines fueled with pure n-butanol exhibit a reduced power output that is relative to conventional gasoline engines [14].

Hydrogen, as a zero-emission renewable fuel, demonstrates excellent combustion properties, such as low ignition energy and rapid flame propagation, positioning it as a promising fuel for internal combustion engines [15]. The comparative properties of these fuels are summarized in

Table 1. The properties of fuels.

	Butanol	Gasoline	Hydrogen
Molecular formula	C4H9OH	C5-C12	H2
Cetane number	25	5–25	-
Research octane number	96	80–99	-
Density/kgL−1	0.808	0.72–0.78 c	0.09 a,b
Viscosity (Pa·s) at 20 °C	3.64	0.28–0.59	-
Lower caloric value/MJkg−1	33.1	42.7	119.7
Latent heat of evaporation/MJkg−1 (25 °C)	0.58	0.38–0.50	-
Saturated vapor pressure/kPa (38 °C)	2.27	31.01	-
Stoichiometric air–fuel ratio	11.21	14.7	34.5
Flammability limits /(Vol%)	1.4–11.2	0.6–8.0	4.0–76.0
Oxygen content/(Mass%)	21.6	-	-
Laminar flame speed/cms−1 (25 °C)	48–53	37–43	185
Autoignition temperature/°C	385	~300	585
Minimum ignition energy/mJ	-	0.24	0.02

^a at 1 bar, ^b at 273 K, ^c at 20 °C.

[16]. Nevertheless, current hydrogen production technologies face economic challenges due to high costs, low volumetric energy density, and storage complexities [17]. Therefore, an optimal application involves using hydrogen as an auxiliary fuel in limited proportions to enhance engine performance [18,19]. Recent studies have investigated the combustion and emission characteristics of butanol–hydrogen blended engines. Iyer et al. [20] examined hydrogen enrichment in a direct-injected methanol-fueled SI engine, finding that hydrogen extends the late injection timing limit by improving fuel–air mixing, thereby optimizing combustion control. Georgescu et al. [21] demonstrated that a hydrogen addition enhances combustion in SI engines, increasing the heat release rate, peak cylinder pressure, and thermal efficiency while shortening the combustion duration. This configuration reduces energy consumption by 4.8% and significantly cuts emissions as follows: HC (−11.1%), CO (−12.5%), NOx (−63.2%), and CO₂ (−33.7%). These and other studies confirm that hydrogen supplementation improves the combustion characteristics and reduces CO and HC emissions in SI engines [22].

SI engines fueled by butanol and hydrogen utilize two primary injection strategies. In the first mode, both fuels are injected into the intake port, forming a homogeneous mixture [23]. In the second mode, n-butanol is port-injected while hydrogen is directly injected into the cylinder, creating a stratified mixture [16]. Compared to direct hydrogen injection, an intake port injection of hydrogen reduces the mass of inducted air due to displacement effects and increases the risk of backfire in the intake manifold [24].

Su et al. [25,26,27,28] conducted comprehensive research on hydrogen/n-butanol dual-fuel rotary engines utilizing intake port injection. Their experimental results demonstrate that hydrogen enrichment significantly increases brake thermal efficiency and the peak in-cylinder temperature while shortening the flame development period and flame propagation duration. A hydrogen addition also effectively reduces combustion instability and decreases HC and CO emissions compared to pure n-butanol operation. However, NO_x emissions increase with rising hydrogen-blending ratios due to elevated combustion temperatures. Raviteja S et al. [29] investigated the impact of minor hydrogen additions (injected into the intake airstream) on butanol-fueled SI engines. The results indicate that hydrogen enrichment improves indicated thermal efficiency by reducing ignition delay, shortening the combustion duration, and elevating the cylinder pressure and temperature. HC and CO emissions decrease, while NOx emissions exhibit a proportional increase with the hydrogen concentration.

Our team has conducted in-depth research on the combustion and emission performance of a hydrogen/n-butanol dual-fuel engine with hydrogen direct injection [16,30]. We found that the hydrogen-blending ratio can enhance the combustion characteristics and lean-burn stability of the n-butanol engine, especially under lean-burn conditions. After implementing the hydrogen addition, CO emissions and HC emissions decreased significantly, but NOx emissions increased. A certain EGR rate can effectively reduce the increase in NOx caused by a hydrogen addition. By adjusting the EGR rate and hydrogen-blending ratio, NOx emissions can be effectively reduced while ensuring the engine’s power performance and stability.

In addition, other researchers have also conducted numerous numerical and experimental studies on SI engines with hydrogen direct injection. Shang Z et al. [31] conducted numerical research on an n-butanol ignition engine with hydrogen direct injection, aiming to observe the effect of the hydrogen-blending ratio on idling performance. Hydrogen blending can accelerate the process of heat release, reduce the emissions of HC and CO, and improve the combustion performance of n-butanol engines at an idle speed. Meng F et al. [32] experimentally studied the combustion and emission characteristics of n-butanol engines with hydrogen direct injection under lean combustion conditions. They found that hydrogen direct injection could significantly improve the power performance and fuel economy of n-butanol engines and reduce the emissions of some harmful gases; however, the NOx emissions rose sharply after hydrogen was added. Meng F et al. experimentally studied the influence of ignition timing and the hydrogen-blending ratio under lean combustion conditions on the combustion and emission characteristics of n-butanol engines with hydrogen direct injection. They found that hydrogen direct injection could significantly improve the power performance and fuel economy of n-butanol engines and reduce the CO and HC emissions; however, it increased NOx emissions. Li G et al. [33,34,35] studied the effect of the injection strategy on the combustion and emissions of hydrogen/gasoline dual-fuel engines through experiments and numerical methods. The results are as follows. Firstly, HIT directly affects the hydrogen concentration near the sparking plug. Secondly, the formation of an enriched hydrogen zone near the spark plug could enhance the ignition and combustion speed of the engine. Thirdly, the relatively uniform distribution of hydrogen throughout the cylinder could reduce emissions.

Although extensive research has been conducted on hydrogen direct injection technology, studies specifically addressing the influence of hydrogen injection timing (HIT) under different λs on hydrogen/n-butanol dual-fuel engine performance remain scarce. The distribution of hydrogen within the cylinder significantly affects combustion efficiency and emission characteristics [36], while adjusting HIT serves as the primary method to optimize mixture stratification [37].

This study systematically investigates the effects of HIT on the performance of an n-butanol/hydrogen dual-fuel engine, emphasizing the novel exploration of the synergistic influence of HIT and lambda variations—a topic that has not been thoroughly examined in previous research. Addressing the current knowledge gap regarding mixture stratification mechanisms and control strategies in hydrogen/n-butanol dual-fuel engines, a combined numerical and experimental approach was employed. Numerical simulations were conducted to elucidate the stratification patterns of hydrogen under different direct injection strategies and their impact on in-cylinder combustion processes. Experimentally, the study evaluates how HIT affects the combustion and emission characteristics under realistic engine operating conditions, enabling the proposal of an optimized hydrogen injection strategy. This work provides new insights into the coupled role of injection timing and the equivalence ratio in dual-fuel combustion using n-butanol and hydrogen, contributing to the advancement of low-carbon engine technologies.

2. Experimental Apparatus and Modeling Methodology

2.1. Experimental Apparatus

The experimental engine platform was modified from a gasoline direct injection (GDI) engine, integrating an independent high-pressure hydrogen direct injection (DI) system to achieve a dual-fuel combustion mode with hydrogen cylinder direct injection and n-butanol port fuel injection. Key technical specifications are summarized in

Table 2. Engine specifications.

Engine Parameter	Parameter Values
Engine Type	four cylinders; dual injection; naturally aspirated; water-cooled; spark-ignited
Compression ratio	9.6:1
Bore × Stroke/mm	82.5 × 92.8
Displaced volume/L	1.984
Maximum power/kW	132 (5000–6000 rpm)
Maximum torque/N m	320 (1600–4000 rpm)

[30]. An electronic control unit (ECU) regulates key operational parameters via closed-loop algorithms, including hydrogen injection timing, injection pulse width, and throttle opening. The schematic configuration of the hydrogen/n-butanol dual-fuel test engine is illustrated in Figure 1.

The measurement instruments and their resolutions are detailed in

Table 3. Test equipment of the engine.

Parameters	Type	Precision	Measurement Range
Speed	CW160	≤±1 rpm	0~6000 rpm
Torque	CW160	≤±0.28 N·m	0~600 N·m
Excess air coefficient	Lambda Meter LA4	≤±1.5%	0.700~32.767
N-butanol mass flow rate	ONO SOKKI DF-2420	±0.01 g/s	0.2~82 kg/h
Hydrogen mass flow meter	DMF-1-1A/B	±0.2%	0.2~2 kg/h
Crank angle	Kistler-2614B	≤±0.5°	0~720°
Cylinder pressure	AVL-GU13Z-24	≤±0.3%	0~20 MPa
Carbon monoxide (CO)	AVL DICOM 4000	≤±0.01% vol	0~15% vol
Hydrocarbon (HC)	AVL DICOM 4000	≤±1 ppm	0~30,000 ppm vol
Nitrogen oxides (NOx)	AVL DICOM 4000	≤±1 ppm	0~5000 ppm vol

. The hydrogen mass flow meter and the n-butanol mass flow rate were measured using a DMF-1-1A/B flow meter and an ONO SOKKI DF-2420 fuel flow meter, respectively. One end of the crankshaft was connected to a CW160 eddy current dynamometer to control the engine speed and measure torque. The other end was connected to a Kistler 2614B4 crankshaft angle encoder, which transmitted the measured speed signal to the Dewesoft SRIUSi combustion analyzer. An AVL GU13Z-24 pressure sensor was installed in cylinder No.2 to measure the in-cylinder pressure signal. This signal was transmitted to the combustion analyzer for combustion analysis. An Bosch LSU4.2 wideband oxygen sensor was installed in the exhaust pipe to measure the excess air coefficient (λ). Emissions of CO, HC, and NOx were measured using an AVL DICOM 4000 five-component tail gas analyzer.

2.2. Modeling Methodology

2.2.1. Engine Model and Initial Parameters

• Engine model

In this paper, a three-dimensional computational (3D) domain composed of the intake port passage, combustion chamber, and exhaust passage is established. The combustion chamber also includes intake and exhaust ports, intake and exhaust valves, a piston top and cylinder head, as shown in Figure 2a. A three-dimensional model of the test engine was constructed using CONVERGE 3.0 software, and the mesh model of the engine was obtained through meshing. The engine mesh model, along with the relative positional relationships of the port fuel injector, in-cylinder hydrogen injection nozzle, and spark plug within the combustion chamber, is shown in Figure 2b. The basic size of the mesh was set as 4 mm with an adaptive mesh refinement (AMR). To ensure the accuracy of the simulation results, two and five refinement levels were set, respectively, for the cylinder and the boundary of the hydrogen inflow. The sectional views of the dynamic mesh model at BDC during the intake stroke and TDC during the compressed stroke are shown in Figure 2c,d.

2.

Initial parameter

The setting of boundaries and initial conditions directly affects the accuracy and reliability of the simulation results. Based on the test bench results, the boundary conditions and initial conditions of the engine are set in this paper. The specific parameters are shown in

Table 4. Boundary and initial condition parameter values.

Parameters	Values
Combustion chamber top surface (K)	550
Piston (K)	600
Cylinder wall (K)	450
Intake port wall (K)	313
Exhaust port wall (K)	500
Intake air (K)	313
Cylinder inside (K)	800
Exhaust back pressure (kPa)	100

.

2.2.2. Mathematical Model

The mass conservation equation, momentum conservation equation, and energy conservation equation are the foundation for simulation calculations. CONVERGE simulation software, based on these three fundamental conservation equations and combined with its built-in mathematical models for turbulence, spray, ignition, combustion, and other processes, enables simulation calculations for different combustion mechanisms: turbulence, spray, ignition, and combustion.

The turbulent model in this study employs the Renormalization Group (RNG) k-ε equations [38]. This model offers superior stability and convergence, making it well-suited for the detailed simulation of in-cylinder airflow motion across a wide range of Reynolds numbers. N-butanol intake port injection can generally be considered as a homogeneous mixture before entering the cylinder [39]. For hydrogen direct injection, six intake orifices matching the actual hydrogen injector’s nozzle size, geometry, and spatial position were incorporated into the geometric model to simulate hydrogen injection. The injected hydrogen mass and injection timing were controlled by adjusting the injection flow rate, start of injection (SOI) phase, and end of injection (EOI) phase. The ignition process was simulated using a source/sink modeling approach to represent the spark plug. Based on the experimental measurements, an energy deposition zone with a 1.5 mm spherical radius was constructed at the spark plug gap. The ignition energy was set to 0.02 J, with the start time corresponding to the ignition timing and with a duration of 0.5 °CA. The combustion model employed for the hydrogen/n-butanol dual-fuel engine investigated in this paper is the SAGE-MZ combustion model [39]. The chemical kinetics mechanism adopted for the hydrogen/n-butanol dual-fuel combustion was developed by integrating the detailed n-butanol chemical kinetic mechanism proposed by Sarathy et al. [40] (878 reactions involving 118 species) with the detailed hydrogen combustion mechanism from the Lawrence Livermore National Laboratory (21 reactions involving 10 species) [41].

2.3. Model Validation

To validate the reliability of the simulation model, this study compares the cylinder pressure and heat release rate data obtained from simulation calculations against data acquired from engine bench tests.

The selected operating conditions are as follows: Engine speed: 1500 rpm; intake manifold pressure: 43 kPa; excess air ratio (λ): 1.0; ignition timing: 10 °CA BTDC; hydrogen injection pressure: 5 Mpa; HIT: 100 °CA BTDC; hydrogen-blending ratio: 5%. The validation results are shown in Figure 3. As can be seen from the figure, the error between the simulated and experimental cylinder pressure data at each crank angle position is less than 5%. The simulation results for the cylinder pressure, heat release rate, and emissions correspond well with the experimental test results. The validation results demonstrate that the simulation model is reliable and can accurately simulate the engine performance for this study, and it can also be used for subsequent analyses of other results.

3. Results and Analysis

3.1. Numerical Analysis

3.1.1. Effects of the HIT on Hydrogen Mixture Distribution in the Cylinder

Figure 4 illustrates the hydrogen stratification state within the cylinder at spark timing for different hydrogen injection timings (HITs). The results demonstrate that HIT significantly influences hydrogen stratification. When hydrogen is injected relatively late, such as at 60 °CA BTDC and 80 °CA BTDC, the hydrogen exhibits insufficient diffusion time. The hydrogen distribution within the cylinder is highly uneven, mainly concentrated on one side of the cylinder. The hydrogen concentration near the spark plug is low, and there is no hydrogen distribution in the far end of the cylinder. This uneven hydrogen distribution state is unfavorable for the ignition process and improving engine performance. This is because hydrogen has a high flame propagation speed and low ignition temperature; an uneven hydrogen distribution leads to uneven flame propagation during combustion. That is, regions with a high hydrogen concentration burn faster and at higher temperatures, while regions without hydrogen burn slower and at lower temperatures. Therefore, this uneven combustion will lead to a decrease in the engine’s power performance. As the HIT is advanced, the hydrogen concentration near the spark plug first increases and then decreases, reaching its maximum at 100 °CA BTDC.

At this injection timing, a favorable stratified hydrogen–air mixture is formed at spark timing, characterized by a certain enrichment of hydrogen near the spark plug and a certain distribution of hydrogen throughout the cylinder. Due to hydrogen’s low ignition energy and high flame propagation speed, the locally enriched hydrogen zone near the spark plug can rapidly ignite the mixture and increase the combustion speed. The broad distribution of hydrogen concentration decreasing progressively outward from the spark plug is beneficial for increasing the combustion speed and completeness in the latter combustion process. But when the injection timing is further advanced to 120 °CA BTDC and 140 °CA BTDC, hydrogen has sufficient time to spread and diffuse within the cylinder, becoming evenly distributed throughout the cylinder. The hydrogen concentration near the spark plug decreases, while the concentration in the far end of the combustion chamber increases. The accelerating effect of hydrogen on the ignition process weakens at this HIT. In summary, the analysis indicates that 100 °CA BTDC is the optimal HIT within the investigated range for the hydrogen DI n-butanol engine.

3.1.2. Effects of the HIT on the Cylinder Temperature

Figure 5 shows the cylinder temperature at different HITs. It can be seen from the figure that at a 10 °CA BTDC crank angle, the spark plug has just fired. At all the HITs, a noticeable small range of temperature rise appears near the spark plug, indicating stable engine ignition. At the −8 °CA ATDC crank angle, the spark plug completes ignition and distinct flame kernels form at all injection timings. At the 0 °CA ATDC crank angle, the flame begins to spread. The piston moves upward, further compressing the cylinder space, and hydrogen diffuses further. The flame propagation situations for the five HITs show little difference. At the 10 °CA ATDC crank angle, the flame has propagated past the middle. The flame propagation is significantly faster at the 100 °CA BTDC injection timing compared to the others, while the flame propagation speed is slowest at 140 °CA BTDC.

As analyzed in Section 3.1.1, this is related to the favorable hydrogen distribution state at the 100 °CA BTDC. At 60 °CA BTDC and 80 °CA BTDC, the hydrogen diffusion time is short and is mainly distributed on one side of the cylinder. The hydrogen concentration in the middle and far ends of the cylinder is very low, which cannot promote combustion well. When the HIT is advanced to 140 °CA BTDC, the hydrogen distribution within the cylinder is relatively uniform, resulting in less-than-ideal acceleration effects on both the ignition and combustion processes. At the 20 °CA ATDC crank angle, flame propagation in the axial view is already close to the cylinder wall. From the axial diagram, although there is not much difference among the five hydrogen injection moments, it can still be seen that a slightly higher cylinder temperature is obtained at the 100 °CA BTDC. However, looking at the radial views from 20 °CA ATDC to 30 °CA ATDC, the difference in flame propagation speed is more apparent. The flame propagation speed at the 100 °CA BTDC is significantly higher than at the other timings. When the crank angle is 50 °CA ATDC, flame propagation is essentially complete. The radial view shows that at the 60 °CA BTDC HIT, the area of low-temperature regions near the combustion chamber walls is larger. At 100 °CA BTDC, the flame has almost propagated to the entire cylinder wall. This is still determined by the hydrogen distribution state. The earlier the HIT, the less uniform the hydrogen distribution. The favorable stratification state at 100 °CA BTDC can balance the acceleration of ignition and combustion rates. Hydrogen has a short quenching distance, and its favorable distribution helps reduce wall quenching, allowing the flame to propagate closer to the walls. Therefore, the 100 °CA BTDC injection timing is the most ideal.

3.2. Experimental Results Analysis

To observe the actual working process of the hydrogen/n-butanol dual-fuel engine, this section experimentally investigates the effects of HIT on the combustion and emission performance of the engine. To clarify the optimization potential of hydrogen addition under different mixture concentrations, this chapter selects excess air coefficients (λ) of 0.9, 1.0, and 1.2 to represent rich mixture, stoichiometric, and lean-burn conditions, respectively. A typical urban operating condition is represented by selecting 1500 rpm and MAP = 43 kPa. The specific engine experimental conditions are shown in

Table 5. Engine experimental conditions.

Operating condition	1500 rpm, MAP = 43 kPa
Ignition advance angle/°CA BTDC	MBT
Hydrogen injection timing/°CA BTDC	60, 80, 100, 120, 140
Hydrogen blending ratio/%	5
Excess air ratio	0.9, 1.0, 1.2
N-butanol injection timing/°CA BTDC	300

.

3.2.1. Effects of the HIT on Combustion Characteristics

In this paper, the flame development duration (θ0–10) and the rapid combustion duration (θ10–90) were respectively defined as the crank angles for which 0–10% and θ10–90% of the fuel mass have been burned. Both are important parameters for evaluating the engine combustion status [42].

Figure 6 shows the variation of θ0–10 with HIT for λ = 0.9, 1.0, and 1.2. It can be seen that as the HIT is advanced, the θ0–10 first shortens slightly and then gradually lengthens. According to the research in 3.1.1, when the HIT is 60 °CA BTDC and 80 °CA BTDC, the hydrogen concentration near the spark plug is relatively low, which cannot effectively accelerate ignition or promote the rapid spread of the initial flame, resulting in a longer θ0–10. At the injection timing of 100 °CA BTDC, a locally enriched hydrogen is used at the initial stage of combustion, thus reducing θ0–10. When the hydrogen injection time reaches 120 °CA BTDC and 140 °CA BTDC, the hydrogen distribution in the cylinder is relatively uniform, and the hydrogen concentration near the spark plug decreases due to increased diffusion, slowing flame propagation, and lengthening the θ0–10.

Additionally, it can be observed that compared to the rich and stoichiometric conditions, under lean-burn conditions, the θ0–10 is more significantly affected by HIT. This is because in the lean-burn conditions, the mixture is thin, the initial temperature in the cylinder is lower, n-butanol evaporation and atomization become more difficult, the ignition of the combustible mixture is more difficult, and it is more likely to cause a fire. Therefore, ignition of the combustible mixture is more difficult and prone to misfire. Hydrogen, as a gaseous fuel with its low ignition energy, ensures ignition stability and effectively mitigates the impact of n-butanol’s high latent heat of vaporization on the θ0–10. Therefore, a locally enriched hydrogen distribution near the spark plug is even more critical under lean-burn conditions, and the θ0–10 is more sensitive to injection timing under lean-burn conditions.

Figure 7 analyzes the variation in the rapid combustion duration (θ10–90) with HIT for λ = 0.9, 1.0, and 1.2. The figure shows that the θ10–90 gradually shortens as the injection timing is advanced. At HITs of 60 °CA BTDC and 80 °CA BTDC, hydrogen is injected late, with insufficient time to spread and diffuse to the far end of the combustion chamber. Hydrogen is concentrated on one side of the cylinder, resulting in a highly uneven distribution. This stratification state is unfavorable for stable and rapid flame propagation, hence the θ10–90 is longer. As the injection timing is advanced, at 100 °CA BTDC, a locally enriched hydrogen zone near the spark plug is formed, with a concentration gradient from the center outward, and a certain hydrogen also exists in the far end of the combustion chamber. Compared with the uneven hydrogen distribution state, the stratified hydrogen distribution state is more conducive to improving the combustion speed. When the hydrogen injection time is postponed to 120 °CA BTDC and 140 °CA BTDC, hydrogen is evenly distributed in the cylinder, which is conducive to the flame spread in the later stages of combustion and shortens the θ10–90. In summary, the θ10–90 gradually shortens as the HIT is advanced. It is worth noting that under let-burn conditions, as the hydrogen injection time gradually advances from 60 °CA BTDC to 100 °CA BTDC, the θ10–90 shortens more significantly. This is because under lean burn conditions, the mixture concentration in the cylinder is low, and flame propagation is slow; the advantage of hydrogen’s high flame speed is better utilized. Therefore, the influence of injection timing on the θ10–90 is more pronounced under lean-burn conditions.

Figure 8 shows the cylinder pressure curves under different HITs for λ = 0.9, 1.0, and 1.2. As can be seen from the figure, for the five HITs within the experimental range, as the HIT is advanced, the cylinder pressure first increases and then decreases, and the position of peak cylinder pressure (P_max) also advances first and then retards. The maximum P_max is obtained at the 100 °CA BTDC injection timing, followed by 140 °CA BTDC, with 60 °CA BTDC yielding the smallest peak pressure.

Combined with the analysis of the effect of injection timing on θ0–10 and θ10–90 in Figure 6 and Figure 7, it is understood that the hydrogen stratification state formed at 100 °CA BTDC is the most ideal. It ensures reliable ignition while rapidly igniting the in-cylinder mixture, increases the combustion speed, concentrates heat release, results in higher cylinder pressure, and advances the position of P_max. At the 140 °CA BTDC injection timing, the hydrogen is evenly distributed in the cylinder. Although there is no locally enriched hydrogen zone near the spark plug to assist ignition, due to the uniform filling of the entire cylinder with hydrogen, the flame propagation speed and combustion speed are increased. Therefore, the pressure inside the cylinder at this time is second only to the optimal HIT but higher than the others.

Furthermore, it can be observed that the larger the λ, the higher the sensitivity of cylinder pressure to injection timing. At λ = 0.9, 1.0, and 1.2, the P_max at 100 °CA BTDC increases by 2.77%, 3.06%, and 5.41% respectively, compared to that at 60 °CA BTDC. This is because under lean-burn conditions, the in-cylinder thermal environment is poor, and a favorable hydrogen stratified mixture state is more crucial for stable ignition and increasing in-cylinder combustion speed. Therefore, under lean-burn conditions, a reasonable hydrogen injection strategy is of greater significance for the combustion performance of the hydrogen/n-butanol dual-fuel engine.

Figure 9 shows the variations in the engine torque with HIT. As seen from the figure, the engine torque first increases and then decreases as the HIT is advanced, reaching its optimal dynamic performance at 100 °CA BTDC within the experimental range. When the injection is late (60 °CA BTDC and 80 °CA BTDC), the stratification effect near the spark plug is poor. Most of the hydrogen is concentrated on one side of the combustion chamber, and there is almost no hydrogen at the far end, which is not conducive to flame propagation. When the injection is too early (120 °CA BTDC and 140 °CA BTDC), hydrogen is evenly distributed, and there is no local enriched hydrogen stratification near the spark plug, which weakens the acceleration effect on the ignition process, especially under lean-burn conditions. At the HIT = 100 °CA BTDC, the stratified mixture state has a locally enriched hydrogen zone near the spark plug. This stratified state of hydrogen is beneficial for both stable ignition and ensuring rapid flame propagation, resulting in more concentrated and complete combustion, thereby enhancing engine power performance. Consequently, the torque first increases and then decreases as HIT is advanced.

3.2.2. Effects of the HIT on Gas Emission Characteristics

Figure 10 shows the variations in the CO emissions with HIT under different λs. Overall, as the λ increases, CO emissions show a decreasing trend, with a significant drop especially from λ = 0.9 to λ = 1.0. This is because CO is primarily produced under low-temperature, oxygen-deficient conditions and is greatly influenced by the mixture concentration. At λ = 0.9, the mixture concentration is high, and oxygen deficiency leads to incomplete combustion, hence high CO emissions. As λ increases, oxygen becomes more sufficient in the cylinder, and CO emissions gradually decrease. Furthermore, under lean-burn conditions, the CO emissions hardly change with injection timing; this is because the CO emissions remain at an extremely low level under this condition, below 0.11% overall. At λ = 0.9 and 1.0, the CO emissions first gradually decrease and then increase as the injection timing is advanced, reaching their lowest level at 100 °CA BTDC. CO emissions at λ = 0.9 are higher, so HIT has the most significant impact on them. CO emissions at 120 °CA BTDC are 16% higher than at 100 °CA BTDC. This is mainly because an appropriate HIT can effectively accelerate the combustion speed, increase the cylinder temperature, and favor CO oxidation. Moreover, as analyzed earlier, the 100 °CA BTDC timing balances ignition stability and overall combustion completeness, reducing CO emissions.

Figure 11 shows the variations in the HC emissions with HIT under different excess air coefficients. As seen from the figure, HC emissions under rich conditions are much higher than those under the other λ values. HC emissions under lean-burn conditions are slightly higher than under stoichiometric conditions. This is mainly because HC emissions are products of incomplete combustion in internal combustion engines, influenced by factors such as mixture concentration, wall quenching, and crevice effects. At λ = 0.9, the fuel concentration is high, and oxygen-deficient conditions cause incomplete fuel combustion, resulting in higher HC emissions. Under lean combustion conditions, although there is sufficient oxygen, due to the thin mixture in the cylinder, ignition is difficult and the combustion speed is slow, leading to incomplete combustion, which, in turn, leads to an increase in HC emissions. Further, it can be found that the HC emissions slightly decrease initially and then slowly increase as the injection timing is advanced at λ = 1.0 and 1.2. This is because an appropriate HIT can balance ignition stability and hydrogen’s promoting effect on the entire combustion process. At λ = 0.9, the earlier the HIT, the lower the HC emissions. This is because under rich conditions, ignition stability is relatively high. The earlier HIT leads to a more uniform and wider hydrogen distribution within the cylinder, better promoting the entire combustion process and making combustion more complete, and thus lowering the HC emissions.

Figure 12 shows the variations in the NO_x emissions with HIT under λ. Overall, as the HIT is advanced, the NO_x emissions first increase and then decrease. The reason for this result mainly depends on the three conditions for NO_x formation as follows: a high temperature, an oxygen-rich environment, and the duration of high-temperature conditions. The HIT directly determines the hydrogen distribution state within the cylinder, thereby affecting the combustion process and cylinder temperature. Based on the previous analysis, at the 100 °CA BTDC injection timing, the hydrogen distribution state leads to the optimal combustion state, resulting in the highest temperature and the longest duration of high-temperature conditions, hence the highest NO_x emissions. Additionally, within the experimental range, as the λ increases, NO_x emissions first rise and then fall, peaking at λ = 1.0. This is because, compared to λ = 1.0, the oxygen-deficient conditions under rich operation thus limit NO_x formation. Under lean-burn conditions, although oxygen is abundant, the lower combustion temperature suppresses NO_x formation.

4. Conclusions

This paper studies the influence of hydrogen injection timing on the mixture distribution, combustion, and emissions of hydrogen/n-butanol engines under different λs through numerical simulation and experiments. The main conclusions of this article are as follows:

• If hydrogen injection occurs too early, the hydrogen becomes evenly distributed throughout the cylinder. If hydrogen injection occurs too late, the hydrogen lacks sufficient time to spread and diffuse, resulting in its concentration primarily on one side of the cylinder. Both overly advanced and overly retarded injection timings fail to create a locally hydrogen-enriched stratified mixture distribution. This is detrimental to accelerating the ignition process and the rapid propagation of the initial flame.
• At the 100 °CA BTDC injection timing, an ideal hydrogen stratification state is formed. This state features a locally hydrogen-enriched region near the spark plug while maintaining a certain level of hydrogen distribution throughout the cylinder. Therefore, 100 °CA BTDC is the optimal injection timing for achieving the desired hydrogen stratification and obtaining the best combustion state in a hydrogen/n-butanol engine.
• Hydrogen injection timing affects the distribution of the in-cylinder mixture, which, in turn, influences the combustion process. Key combustion performance parameters, such as the torque, flame development period, and cylinder pressure, all reach their optimal values at the 100 °CA BTDC injection timing. Furthermore, under lean-burn conditions, the combustion parameters exhibit greater sensitivity to variations in injection timing.
• CO emissions and HC emissions show a trend of slightly decreasing initially and then slowly increasing as the injection timing is advanced. The 100 °CA BTDC injection timing effectively reduces CO emissions at λ = 0.9 (slightly rich) and λ = 1.0 (stoichiometric). However, CO emissions at λ = 1.2 (lean) are less affected by the injection timing changes. Therefore, selecting a reasonable hydrogen injection timing is beneficial for improving both the combustion performance and emission characteristics of the hydrogen/n-butanol engine.