Study on the Effects of the Hydrogen Substitution Rate on the Performance of a Hydrogen–Diesel Dual-Fuel Engine under Different Loads

Abstract

Due to having zero carbon emissions and renewable advantages, hydrogen has great prospects as a renewable form of alternate energy. Engine load and hydrogen substitution rate have a considerable influence on a hydrogen–diesel dual-fuel engine’s efficiency. This experiment’s objective is to study the influence of hydrogen substitution rate on engine combustion and emission under different loads and to study the impact of exhaust gas recirculation (EGR) technology or main injection timing on the engine’s capability under high load and high hydrogen substitution rate. The range of the maximum hydrogen substitution rate was determined under different loads (30%~90%) at 1800 rpm and, then, the effects of the EGR rate (0%~15%) and main injection timing (−8 °CA ATDC~0 °CA ATDC) on the engine performance under 90% high load were studied. The research results show that the larger the load, the smaller the maximum hydrogen substitution rate that can be added to the dual-fuel engine. Under each load, with the increase of the hydrogen substitution rate, the cylinder pressure and the peak heat release rate (HRR) increase, the equivalent brake-specific fuel consumption (BSFC_equ) decreases, the thermal efficiency increases, the maximum thermal efficiency is 43.1%, the carbon dioxide (CO₂) emission is effectively reduced by 35.2%, and the nitrogen oxide (NOx) emission decreases at medium and low loads, and the maximum increase rate is 20.1% at 90% load. Under high load, with the increase of EGR rate or the delay of main injection timing, the problem of NOx emission increases after hydrogen doping can be effectively solved. As the EGR rate rises from 0% to 15%, the maximum reduction of NOx is 63.1% and, with the delay of main injection timing from −8 °CA ATDC to 0 °CA ATDC, the maximum reduction of NOx is 44.5%.

1. Introduction

Under the double troubles of energy shortage and global warming, many countries have put forward carbon-reduction targets and carried out research on alternative fuels for engines. In order to minimize internal combustion engine carbon emissions and achieve clean combustion, it is crucial to employ low- or zero-carbon, clean, and renewable alternative fuels [1].

Researchers constantly explore various alternative fuels, such as natural gas, H₂, biofuel, alcohol [2,3,4,5,6], etc., as possible alternatives to current transport fuels in order to reduce total carbon emissions during transport. Due to its absence of carbon elements and its advantages, such as fast combustion rate, large diffusion coefficient, and wide ignition limit [7,8], hydrogen is considered the cleanest and ideal alternative energy source [9]. H₂ can be utilized as a single fuel, which is usually called a “Hydrogen fuel internal combustion engine”. Extensive study has been undertaken on the engine’s performance [10,11,12,13,14,15]. Compared with traditional engines, it has high combustion efficiency and no harmful emissions, such as carbon smoke, CO, and CO₂. Hydrogen can be added to the engine as an alternative fuel and the engine can switch between the two fuels according to demand, with high flexibility to adapt to different working loads and driving conditions. Existing engine infrastructure can be utilized without the need for a dedicated high-pressure hydrogen supply system, thereby reducing costs.

Hydrogen has been used extensively as a clean energy source throughout the last few decades. Yuriy Gutarevych et al. [16] found that adding hydrogen improves fuel economy and reduces CO emissions in gasoline engines. Karagoz et al. [17] found that adding hydrogen improved the thermal efficiency and reduced the power of gasoline engines, with CO and THC emissions approaching zero, but NOx emissions increased by up to 99.5%. Masaki Naruke et al. [18] showed that hydrogen doping can effectively extend the lean burn limit of gasoline engines and improve thermal efficiency. Due to their great efficiency, potent power production, longevity, and dependability, as well as combustion economy, diesel engines are widely employed in a variety of industries. However, their CO₂ and particulate emissions are relatively high [19,20] and secondary fuel can be added to reduce emissions while maintaining power. Adding hydrogen to diesel engines has been a new idea in recent years. The higher natural temperature and lower ignition energy of hydrogen enable diesel engines to use diesel as a pilot to ignite hydrogen, which can improve the performance of diesel engines by utilizing the advantages of zero-carbon cleanliness of hydrogen and combining the high efficiency and good economy of diesel engines.

The prior literature shows that hydrogen-doped diesel-engine combustion is an efficient technique to decrease emissions. Szwaja et al. [21] investigated the impact of mixing hydrogen on engine performance and the research results showed that mixing hydrogen in a diesel engine could make the fuel mixture more uniform and burn better. Jafarmadar et al. [22] investigated how the dual-fuel engine’s performance changed and discovered that, as the hydrogen mixture grew, so did the average effective pressure also increase. Menna et al.’s [23] research of a hydrogen–diesel dual-fuel engine revealed that combining hydrogen with diesel increased NOx pollution while lowering CO₂ emissions. Yang Zhenzhong [24] studied the emission characteristics of CO, CO₂, and HC in different proportions of hydrogen–diesel hybrid fuels and found that hydrogen–diesel dual-fuel helped to lower the specific emission of CO, CO₂, and HC. Tsujimura et al. [10] discussed the effect of hydrogen mixing on the performance of single-cylinder diesel engines and the findings indicated that the thermal efficiency decreased at low load and increased at high load. Adnan et al. [25] studied the effect of engine speed on hydrogen doping in diesel engines and the results of the tests indicated that when the engine is running at medium to low speeds, CO emissions are reduced. As the engine speed increases, the highest cylinder temperature increases, leading to an enhancement in CO and NOx. Naber et al. [26,27] studied the spontaneous combustion characteristics of hydrogen gas in diesel engines and the findings suggested that there is a prominent correlation between the natural characteristics of hydrogen gas and gas temperature and ignition time. Liang Li et al. [28] discussed the impact of hydrogen doping on the efficiency of engines and discovered that power, combustion temperature, and cylinder pressure all rose at a 10% hydrogen substitution rate. The higher the temperature, the more NOx emissions. They also found that under high load, hydrogen doping led to a more significant increase in NOx.

A dual-fuel engine’s cylinder combustion can be improved by the addition of a particular quantity of hydrogen, effectively reducing some emissions but worsening NOx emissions. Based on this, the introduction of the EGR technique can lessen NOx emissions [29]. By utilizing EGR technology, diesel engines’ NOx emissions may be successfully reduced and the knocking phenomenon brought on by hydrogen combustion in diesel engines can be prevented [30,31]. In an inquiry into emissions of hydrogen-diesel dual-fuel engines, Dimitriou [32] et al. discovered that EGR technology significantly affects NOx emission control at moderate load. Yu Xiumin et al. [33] analyzed the consequences of collaborative control of hydrogen quantity and EGR on engine NOx emission and it was declared that when the hydrogen substitution rate was high, the effect of EGR on NOx emission reduction was more obvious. Wu [34] reported that the application of a 20% hydrogen blending ratio and 40% EGR in a diesel engine reduces NOx and smoke emissions at the same time, as opposed to the use of pure diesel fuel. The results obtained display that injection parameters also have a big impact on the combustion and emission performance of a dual-fuel engine. Miyamoto et al. [35] found that premature fuel injection results in an important surge in NOx production and a quick rise in combustion chamber pressure. Additionally, Tomita et al. [36] also confirmed that the average effective pressure and NOx will raise due to premature diesel injection. Gaurav, Tripathi et al. [37] examined the impact of the ignition advance angle on a hydrogen–diesel dual-fuel engine, and the conclusions proved that with the advance of ignition timing, the combustion of mixed fuel was more complete, resulting in a reduction in HC and CO, but an increase in NOx. Osama H. Ghazal [38] examined the shape of injection parameters and hydrogen mixing on diesel-engine performance and the results showed that, with the earlier injection time, engine pressure rise rate increased, in-cylinder combustion pressure and temperature raised, and NOx increased. The main injection timing has an enormous effect on the time, directly affecting engine performance and exhaust emissions. When hydrogen is burned in a diesel engine, the adjustment of the main injection timing can further optimize the combustion process to improve combustion efficiency and reduce emissions.

According to the past research findings, many researchers have explained the advantages of hydrogen combustion in traditional engines. However, it is necessary to further analyze the effect of the hydrogen substitution rate on the capability of a dual-fuel engine under different loads, and take EGR technology and fuel-injection control to solve the emission problems caused by the increase in the hydrogen substitution rate under partial loads. In this paper, the dual-fuel integrated controller independently developed by the team can facilitate the precise injection control of diesel and hydrogen, along with the flexible switching of single-fuel and dual-fuel modes, and realize the collaborative integrated control of the fuel injection and air system. By controlling the ratio of two fuels under different loads, the maximum hydrogen substitution rate is determined and the detrimental impact of hydrogen doping on the functional mode of the diesel engine is studied. At the same time, the influence of EGR rate and main minimum injection timing on the performance of hydrogen–diesel dual-fuel engines is studied, which provides a theoretical basis for further optimization and control of engine performance and emissions.

2. Test Device and Test Method

2.1. Test Device

The engine parameters used in the test are shown in

Table 1. High-pressure common rail diesel engine technical parameters.

Projects	Parameter
Rated power (kW)	115
Compression ratio	16.6:1
Rated speed (rpm)	3200
Maximum torque (N·m)	450
Maximum torque speed (rpm)	1800
Cylinder diameter × stroke (mm)	95 × 105
Air intake form	Supercharged intercooling
Engine displacement (L)	2.977
Fuel injection system	High-pressure common rail direct injection
Rail pressure (MPa)	0–200
Cooling mode	Water cooling

. The two fuels used in the test are 0# diesel and 99.5% hydrogen, and the physical and chemical characteristics of the two fuels used in the test are shown in

Table 2. Physicochemical properties of diesel and hydrogen fuels.

Property Fuel	Hydrogen	Diesel Oil
Main components	H2	C10–C21
C Mass fraction %	0	86–89
H Mass fraction %	100	12.6
O Mass fraction %	0	0–0.4
Density (kg·m−3)	0.09	840
Theoretical air–fuel ratio	34.3	14.3
Low heat value (MJ·kg−1)	120.9	42.5
Flame-propagation speed (m·s−1)	3	0.42
Minimum ignition energy (MJ)	0.02	0.25
Fire limit (%)	4–75	0.6~6.5
Diffusion coefficient (cm2·s−1)	0.63	—
Flame quenching distance (mm)	0.64	—

[11,39]. The schematic diagram of the test bench is shown in Figure 1, the physical object is shown in Figure 2, and the main test and control equipment are shown in

Table 3. Main measurement and control equipment.

Device Name	Model Number
Transient fuel consumption meter	FCMA
Measure and control instrument	EIM609
Atmospheric simulation integrated measurement and control system	CEM101
Hydrogen flowmeter	CMFS025MB67N4BPMKZZ
Combustion analyzer	AVL 622
Fourier transform multicomponent gas analyzer	AVL FTIR i60

.

As is shown in Figure 1, the hydrogen from the hydrogen cylinder group (1) is depressurized and reaches the hydrogen common rail (8). It is then injected into the intake manifold by the hydrogen nozzle (9) and mixed with air before entering the combustion chamber. The diesel oil is pressurized by the high-pressure oil pump (13) to reach the diesel high-pressure common rail (14) and, then, directly injected into the combustion chamber by the diesel nozzle (15). After being pressurized by the turbocharger (19), the air flows through the intercooler (20) to the intake manifold. The mixture of air and hydrogen enters the cylinder and is ignited by diesel fuel. After combustion, part of the exhaust gas flows through the exhaust pipe and is discharged through the turbine. The other part of the exhaust gas can flow back to the intake pipe through the EGR cooler (23) and EGR valve (22) and enter the cylinder again to participate in combustion. Adjust the control parameters required by the dual fuel ECU (27) through the upper computer (31), output control signals to control the hydrogen nozzle (9), diesel nozzle (15), and EGR valve (22), and further adjust the amount of hydrogen, diesel, and exhaust gas injected into the cylinder. The AVL combustion analyzer (28) collects cylinder pressure through the cylinder pressure sensor (32) and calculates parameters such as heat release rate and temperature. Finally, the emission analyzer (26), dual fuel ECU (27), and AVL combustion analyzer (28) are read through the data collector (30).

2.2. Test Method

When studying the consequence of hydrogen substitution rate on dual-fuel engine economy, burn and emission properties under various loads, due to the characteristics of hydrogen combustion speed, high calorific value, and large diffusion coefficient under high loads lead to rough in-cylinder combustion, increased heat load, and increased NOx emission, etc. Therefore, in the test process, the maximum exhaust temperature and maximum explosion pressure are mainly used as constraint conditions to obtain the maximum hydrogen substitution rate under each load and record the data, so as to avoid problems such as rough combustion in the cylinder and high emissions.

During the test, the engine cooling water temperature is controlled at (80 ± 0.5) °C, and test data collection is done after the engine is running steadily. The main test method is to conduct a test with a replacement rate of 0 after a fixed speed and load, then open the hydrogen supply valve, adjust the hydrogen supply pressure to 0.4 MPa, switch the dual-fuel mode, gradually increase the hydrogen-injection pulse width, and increase the hydrogen gas supply. With the self-developed dual-fuel ECU as the control core, the fuel-injection quantity of the original engine under the current hydrogen-doping amount is obtained according to the total energy-conservation control strategy, so as to realize the accurate injection control of diesel and hydrogen. When the maximum hydrogen substitution rate of this working condition is increased, three groups of data are collected at each working condition and averaged and the data are recorded. Then, the hydrogen pulse width is gradually reduced to 0 and the hydrogen gas supply valve is closed and switched to the next working condition. The above steps are repeated to carry out the test and record the data, thus obtaining the maximum hydrogen substitution rate under different working conditions.

The impacts of EGR technology and injection timing on the economy, combustion characteristics, and emission characteristics of a diesel engine blended with hydrogen were studied. During the test, the maximum hydrogen substitution rate corresponding to a high load at 1800 rpm was chosen. First, fix the diesel injection strategy and adjust the EGR valve. The data under different EGR rates were recorded. Then the EGR valve is closed, the injection timing of the diesel main injection is gradually adjusted from −8 °CA ATDC to 0 °CA ATDC, and the data of different main injection timings are recorded. The impact of the aforementioned variables on engine performance is analyzed next. During the emission test, the sampling frequency of the AVL FTIR i60 multicomponent gas analyzer is 1 Hz and the sampling time is 10 s.

2.3. Measurement Calculation Formula

The hydrogen substitution rate represents the percentage of the energy delivered by hydrogen to the energy jointly provided by the two fuels at the same operating point, and the EGR rate represents the ratio of the amount of recycled exhaust gas to the total intake of air sucked into the cylinder. The calculation formula for the hydrogen substitution rate is shown in (1) and the calculation formula for the EGR rate of exhaust-gas recirculation is shown in (2). In order to characterize the economy of a hydrogen–diesel dual-fuel engine with the same calorific value, the hydrogen consumption is converted into diesel consumption through the calorific values of hydrogen and diesel, and the converted diesel consumption plus the pilot diesel consumption is the total fuel consumption under the dual-fuel setting. The BSFC_equ is equal to the total fuel consumption compared to the effective work it does and is a member of the key metrics for assessing engine fuel economy. The calculation formula is shown in (3).

$${\gamma }_H=\frac{m_H{h}_H}{m_H{h}_H+m_D{h}_D}\times 100\%$$

(1)

In the formula,$m_D$ is the diesel consumption, kg/h; $m_H$ is the hydrogen consumption, kg/h; $h_H$ = 120.9 MJ/kg is the low calorific value of hydrogen gas; $h_D$ = 42.5 MJ/kg is the low calorific value of diesel fuel; and ${\gamma }_H$ is the hydrogen substitution rate, %.

$${\eta }_{EGR}=\frac{{\phi }_{C{O}_2}{}_{,in}}{{\phi }_{C{O}_2}{}_{,out}}\times 100\%$$

(2)

In the equation, ${\phi }_{C{O}_2,in}$ is the volume fraction of CO₂ in the intake air diluted by EGR; ${\phi }_{C{O}_2,out}$ is the volume fraction of CO₂ in the exhaust gas; and ${\eta }_{EGR}$ is the EGR rate.

$$b_e=\frac{m_D+\frac{h_H{m}_H}{h_D}}{P_e}\times {10}^3$$

(3)

In the formula, $b_e$ is BSFC_equ, g/(kW·h) and $P_e$ is the engine power, kW.

3. Results and Discussions

3.1. Influence of Hydrogen Substitution Rate in Various Loads

Taking the maximum explosion pressure and exhaust temperature as constraints, the range of the hydrogen mixing ratio of a diesel engine was studied under different loads with the speed of 1800 rpm. The maximum hydrogen substitution rate under different loads was obtained.

Table 4. Maximum hydrogen-doping pulse width and hydrogen substitution rate under different loads.

Operation Point	Maximum Hydrogen-Doping Pulse Width (ms)	Maximum Hydrogen Substitution Rate (%)
30% load	12	60
50% load	10	35
70% load	10	28
90% load	8	15

shows the maximum hydrogen-mixing pulse width and hydrogen substitution rate under various loads. Table 4 demonstrates that the higher the load, the smaller the hydrogen substitution rate. When the load is 30%, the maximum hydrogen substitution rate can reach 60%; then, the hydrogen substitution rate gradually decreases with the increase of the load. The primary cause is that under high load, because hydrogen has a high calorific value, it burns quickly and has a high diffusion coefficient; such problems as rough combustion, increased NOx emission, and increased heat load limit the increase of hydrogen mixing amount.

3.1.1. Analysis of Combustion Characteristics

Figure 3 shows the effect of the hydrogen substitution rate on cylinder pressure and HRR under different loads, while Figure 4 shows the effect of the hydrogen substitution rate on maximum cylinder pressure and the maximum average temperature in the cylinder under different loads.

From Figure 3, it is apparent that the peak cylinder pressure and HRR both increase with the growth of the hydrogen substitution rate under various loads, and the corresponding peak phase shifts forward, while they raise with the growth of the engine load. Compared with pure diesel, under the condition of maximum hydrogen substitution rate under each working condition, when the load is 30%, 50%, 70%, and 90%, the peak cylinder pressure increases by 4.81%, 5.08%, 6.77%, and 6.81%, respectively, and, as the load grows, the peak cylinder pressure rises higher. Under the condition of a 90% high load, with the hydrogen substitution rate increasing from 0% to 15%, the cylinder’s peak pressure rose from 12.53 MPa to 13.38 MPa and the peak HRR increased from 197 J/°CA to 228 J/°CA.

As shown in Figure 4, compared with pure diesel engines, the highest cylinder pressure and the average temperature in the cylinder increase along with a rise in the rate of hydrogen substitution; the greater the load, the higher the maximum cylinder pressure and temperature. With the increase of the hydrogen substitution rate, when the load is 30%, 50%, 70%, and 90%, the average increase of the maximum cylinder pressure is 2.70%, 3.4%, 4.25%, and 4.49%, respectively. The average increase in the maximum temperature was 0.15%, 1.28%, 2.17%, and 2.51%, respectively.

When the combustible mixture formed during the combustion delay period catches fire, the hydrogen in the cylinder will be ignited by multiple flame centers in the cylinder in a short time and most of the hydrogen will be burned in the premixed combustion mode. Due to hydrogen’s rapid rate of combustion, when premixed combustion’s pressure and temperature rise, the energy is released quickly after the fuel in the cylinder is burned, which shortens the ignition-delay period and combustion duration, The shorter ignition-delay period is influenced by the high flow flame velocity of hydrogen combustion, leading to stronger turbulence in the engine cylinder, accelerating the diesel’s diffusion combustion, and, thus, accelerating the overall combustion speed of the fuel, increasing the peak pressure, and HRR, while also raising the cylinder temperature.

3.1.2. Economic Analysis

Figure 5 shows the comparison of BSFC_equ and thermal efficiency between pure diesel and dual-fuel modes. Figure 6 shows the effect of different hydrogen substitution rates on BSFC_equ and thermal efficiency.

As shown in Figure 5, the BSFC_equ drops while the thermal efficiency grows at the highest hydrogen substitution rate, compared to a pure diesel engine. In comparison to pure diesel, under the condition of maximum hydrogen substitution rate, when the load is 30%, 50%, 70%, and 90%, the BSFC_equ decreases by 2.7%, 3.4%, 4.5%, and 4.1%, respectively, and the maximum BSFC_equ decreases by 4.5% after adding hydrogen, and the minimum BSFC_equ is 196.57 g/(kW·h). The thermal efficiency increases by 2.8%, 3.5%, 4.6%, and 4.2%, respectively. After hydrogen mixing, the maximum increase in thermal efficiency was 4.6% and the highest thermal efficiency was 43.1%.

In Figure 6, the BSFC_equ drops progressively as the hydrogen substitution rate increases, whereas the thermal efficiency grows gradually as the hydrogen substitution rate increases. The BSFC_equ drops and the thermal efficiency rises as the load grows. As the rate of hydrogen substitution increases, when the load is 30%, 50%, 70%, and 90%, the average reduction range of BSFC_equ is 1.70%, 1.44%, 2.31%, and 1.94%, respectively. The average increase range of thermal efficiency is 1.70%, 1.50%, 2.31%, and 2.06%, respectively.

As the rate of hydrogen substitution grows, the high diffusion rate of hydrogen can significantly improve the mixing rate of air and hydrogen fuel and improve the uniformity of the combustible mixture in the engine combustion chamber. However, hydrogen has a quicker combustion rate and flame-propagation rate. After the air and fuel in the cylinder are mixed, energy is released quickly and the temperature in the combustion chamber increases accordingly. At high combustion temperatures, the fuel burns more fully, and the combustion process is more thorough, thereby improving thermal efficiency. In addition, the inclusion of hydrogen improves the uniformity and efficiency of the combustion, with more hydrogen burns, thereby improving the overall fuel utilization rate and reducing fuel consumption.

3.1.3. Analysis of Emission Characteristics

Figure 7 shows the comparison of CO₂ emissions between pure diesel and dual-fuel modes and Figure 8 shows the effect of different hydrogen substitution rates on CO₂ emissions.

As seen in Figure 7, the CO₂ emission under the maximum hydrogen substitution rate is lower than that of a pure diesel engine under all loads. Compared with pure diesel oil, under the condition of maximum hydrogen substitution rate, when the load is 30%, 50%, 70%, and 90%, the CO₂ decreases by 35.2%, 34.8%, 28.4%, and 15.5%, respectively, after adding hydrogen.

As illustrated in Figure 8, the CO₂ emission under each load decreases gradually with the increase of the hydrogen substitution rate; CO₂ emissions decrease as the load increases. With an increase in hydrogen substitution rate, when the load is 30%, 50%, 70%, and 90%, the average reduction range of CO₂ is 20.2%, 21.9%, 17.7%, and 9.9%, respectively.

Hydrogen itself is a kind of carbonless combustion. When it burns, mixed combustion produces less CO₂, which inhibits the generation of CO₂ from the source. The flame-propagation speed of hydrogen is faster, which can reduce the nonuniformity of diesel spray when introduced into diesel engines. Moreover, the high diffusivity of hydrogen increases the quality and uniformity of the premixed combustible mixture and improves the complete combustion efficiency. Therefore, mixing hydrogen reduces CO₂ emissions. As the load increases, the engine requires more energy to generate additional power. To provide more power, the engine burns more fuel. However, when the engine is under high load, the combustion efficiency is improved and the energy of the fuel can be more fully utilized, meaning that more input energy is converted to output power and the waste of energy is reduced. This results in more power output per unit of energy consumed, reducing additional fuel consumption and corresponding CO₂ emissions.

Figure 9 shows a comparison of NOx emissions between pure diesel and dual-fuel modes and Figure 10 shows the effect of different hydrogen substitution rates on NOx emissions.

Figure 9 illustrates that, under the maximum hydrogen substitution rate, the NOx emission is lower than the pure diesel engine in the middle- and low-load region and the NOx emission is higher compared to the pure diesel engine in the high-load region. In contrast to pure diesel, under the condition of maximum hydrogen substitution rate, when the load is 30% and 50%, the NOx emission decreases by 11.6% and 10.9%, respectively, after adding hydrogen. When the load is 70% and 90%, the NOx emission increases by 7.4% and 20.1%, respectively.

From Figure 10, with an uptick in the rate of hydrogen substitution, NOx emissions decrease in medium- and low-load areas, while NOx emissions increase in high-load areas. As the load increases, NOx emissions decrease. As the hydrogen substitution rate increases, when the load is 70% and 90%, the average increase in NOx is 4.54% and 9.53%, respectively.

The conditions that lead to NOx production are high temperatures and an oxygen-enriched environment and the factors that affect NOx emission mainly include high-temperature duration, maximum combustion temperature in the cylinder, and oxygen content in the cylinder. At low load, the airflow in the cylinder is weak. With a spike in the rate of hydrogen substitution, the amount of pilot diesel oil is reduced, so that the initial ignition point is reduced and the temperature increases slightly. At low loads, the maximum hydrogen substitution rate can be increased and more hydrogen is introduced to reduce the air in the cylinder, which strengthens the oxygen-content reduction. Moreover, the decrease of the pilot fuel injection rate in the later stage makes the maximum combustion temperature in the cylinder not increase much, which is beneficial to the lowering of NOx emissions. At medium and high loads, with the enhancement of airflow movement, the creation of mixed gas and the dispersion of oil are accelerated. In addition, the amount of pilot diesel increases under high load, resulting in more initial ignition points. Hydrogen-assisted diesel combustion can improve the complete combustion efficiency, and complete combustion leads to higher peak pressure, resulting in a higher temperature. Since hydrogen has a high heat value and burns more quickly, the overall combustion speed in the cylinder is accelerated and the maximum temperature rises greatly, which leads to the growth of NOx emission with the increase in the hydrogen substitution rate at high load.

3.2. Impact of EGR Rate

According to the research on the impact of the hydrogen substitution rate on engine performance under different loads mentioned above, it is evident that, as the hydrogen substitution rate increases at high loads, the engine’s cylinder combustion improves and the economy grows, but Nox emissions are increased. Based on this, EGR and fuel-injection timing technology are introduced to explore the extent of their impact on the Nox emissions of a dual-fuel engine at high load.

Table 5. Test operating point.

Research Parameters	Load/%	Maximum Hydrogen Substitution Rate/%	Preinjection Timing/°CA ATDC	Main Injection Timing/°CA ATDC	EGR Rate/%
impact of EGR rate	90	15	−18.6	−4	0, 5, 10, 15
impact of main injection timing	90	15	−18.6	0, −2, −4, −6, −8	0

lists the test operating points.

3.2.1. Analysis of Combustion Characteristics

Figure 11 shows the effect of the EGR rate on the cylinder pressure and HRR of a dual-fuel engine, while Figure 12 shows the effect of the EGR rate on the maximum cylinder pressure and the maximum average temperature in the cylinder of a dual-fuel engine. From Figure 11, as the EGR rate increases, both the peak cylinder pressure and HRR decrease, and the corresponding peak phase gradually shifts back. As the EGR rate rises from 0% to 15%, the peak cylinder pressure falls from 13.47 Mpa to 12.97 Mpa, and the peak HRR decreases from 228 J/°CA to 186 J/°CA. As seen in Figure 12, as the EGR rate increases, the maximum cylinder pressure and average temperature decrease. Compared with the no EGR mode, the maximum cylinder pressure at 5%, 10%, and 15% EGR rates decreases by 0.5%, 1.7%, and 3.7%, respectively, and the maximum average temperature decreases by 0.8%, 1.1%, and 2.5%, respectively.

As the EGR rate has grown, more recycled exhaust gas is introduced into the cylinder, so that polyatomic molecules such as CO₂ and H₂O are added to the mixture, and the inert gas CO₂ increases, which changes the composition of the mixture in the cylinder. The dilution and heat-capacity effects of the exhaust gas are continuously enhanced and the combustion and heat release of the combustible mixture are inhibited, which reduces the temperature of combustion, the ignition time is delayed, the ignition-delay period is prolonged, the flame-propagation speed is reduced, and the combustion end time is gradually delayed, thus increasing the combustion duration. And that addition of recycled exhaust gas diminishes the oxygen content, which will directly lead to the decrease of the chemical reaction rate involving oxygen, and the combustion chemical reaction rate of the mixed gas in the cylinder will be slowed down and the combustion will be inhibited, thus leading to the decrease of the maximum cylinder pressure and HRR.

3.2.2. Economic Analysis

Figure 13 shows the effect of the EGR rate on BSFC_equ and the thermal efficiency of a dual-fuel engine. From Figure 13, the rate of EGR is rising, BSFC_equ increases, and thermal efficiency decreases. Compared with no EGR, after adding the 5%, 10%, and 15% EGR rate, the BSFC_equ increased by 0.39%, 0.84%, and 1.2%, respectively, the thermal efficiency decreased by 0.36%, 0.82%, and 1.3%, respectively, and the highest thermal efficiency was 43.1%.

The duration of combustion is extended by the use of EGR. It lengthens the heat dissipation time, increases the specific heat capacity of the mixture, and increases the heat absorption, which causes the engine heat loss to rise [40]. As the EGR rate improves, some exhaust gas is once again involved in combustion, leading back to a higher smoke value in the exhaust gas returned to the cylinder. The concentration of oxygen in the cylinder decreases, inhibiting fuel combustion, resulting in a decrease in average combustion temperature and cylinder pressure. The decrease in temperature leads to incomplete combustion and insufficient release of heat energy. Moreover, lower cylinder pressure can lead to a decrease in combustion efficiency, requiring more fuel to maintain the same power output. The effects of the aforementioned parameters combined to result in a rise in BSFC_equ and a fall in thermal efficiency.

3.2.3. Analysis of Emission Characteristics

Figure 14 shows the effect of the EGR rate on the NOx and CO₂ emissions of a dual-fuel engine. From Figure 14, CO₂ emissions rise while NOx emissions fall as the EGR rate rises. Compared with no EGR, after adding 5%, 10%, and 15% EGR rates, CO₂ emissions increase by 2.1%, 4.7%, and 6.6%, respectively, and NOx emission decrease by 28.9%, 57.8%, and 63.1%, respectively.

As the EGR rate has increased, CO₂ exists in the exhaust gas itself and the oxygen concentration in the cylinder dilutes, which leads to the decrease of combustion temperature and combustion speed in the cylinder and incomplete fuel combustion, thus increasing CO₂ emissions.

The generation condition of a NOx emission is high temperature and oxygen enrichment. With the increase of the EGR rate, the exhaust gas dilutes the fresh air that can be used for combustion, which slows down the combustion speed and reduces the temperature and pressure. Since EGR introduces high heat-capacity components, such as CO₂ and H₂O, to absorb part of the heat, the maximum temperature decreases and the high-temperature area decreases; so, the NOx generation area decreases, which leads to a large reduction in NOx emission. On the other hand, due to the lack of oxygen, will hinder the contact between N₂ and O₂ to some extent, thus achieving the purpose of reducing NOx emissions.

3.3. Impact of Main Injection Timing

3.3.1. Analysis of Combustion Characteristics

Figure 15 shows the effect of main injection timing on the cylinder pressure and HRR of a dual-fuel engine, while Figure 16 shows the effect of main injection timing on the maximum cylinder pressure and the maximum average temperature in the cylinder of a dual-fuel engine.

From Figure 15, as the main injection timing advances, the peak cylinder pressure and HRR gradually increase and the corresponding peak phase gradually shifts forward. As the main injection timing moves forward from 0 °CA ATDC to −8 °CA ATDC, the pressure grows from 10.97 MPa to 15.96 MPa, and the peak HRR increases from 195 J/°CA to 271 J/°CA. According to Figure 16, with the advance of the main injection timing, the maximum pressure and average temperature both increased. In comparison to the top dead center, when the main injection timing is −2, −4, −6, and −8 (°CA ATDC), the maximum cylinder pressure increases by 9.7%, 21.8%, 33.7%, and 45.5%, respectively, and the maximum average temperature increases by 1.4%, 2.8%, 4.9%, and 6.8%, respectively.

The primary cause is that, as the main injection timing advances, the amount of pilot diesel increases before the piston reaches TDC, which is conducive to the early formation of a mixture of diesel near the TDC. The diesel part of the main injection is premixed with the surrounding air. At this time, with the piston’s upward motion, after the cylinder pressure and temperature reach certain conditions, the fuel begins to burn and ignites the premixed hydrogen mixture in the cylinder, more fuel is involved in combustion, increasing the number and energy of ignition points. The combustion rate of the mixture continues to accelerate, making the combustion process more complete and efficient, resulting in higher thermal power conversion efficiency [41], ultimately resulting in an increased pressure and HRR. In addition, with the advance of the main injection timing, the combustion starting point and center of gravity move forward, causing more fuel to participate in combustion near the TDC, and also causing a sharp increase in peak pressure and HRR. And the stratification degree of the mixture in the cylinder is greatly reduced and ever more diesel ignition cores are distributed in the thin combustible mixture outside the diesel-spray fuel spray, which improves the activity of the combustible mixture outside the fuel spray and increases the temperature and rate of combustion.

3.3.2. Economic Analysis

Figure 17 shows the effect of main injection timing on BSFC_equ and the thermal efficiency of a dual-fuel engine. From Figure 17, as the main injection timing advances, the BSFC_equ decreases while the thermal efficiency increases. Compared with the TDC, when the main injection timing is pushed from −2 °CA ATDC to −8 °CA ATDC, the maximum decrease in BSFC_equ is 8.3%, the maximum increase in thermal efficiency is 8.5%, and the maximum thermal efficiency is 44.5%.

The main injection diesel, as the ignition fuel, will be injected in advance to promote the fuel as a whole to start combustion near TDC. At this time, when the piston goes up, the peak pressure increases significantly. Simultaneously, the combustion start point and the combustion center of gravity move forward, making the combustible mixture near TDC burn rapidly, which is conducive to heightening the combustion thermal efficiency. When the piston goes up, the cylinder content volume becomes smaller, the combustion constant volume is better, and the heat transfer area is reduced, such that there is a decrease in the amount of heat transferred from the combustible mixture through the cylinder wall, increasing thermal efficiency and decreasing fuel consumption rate. Coupled with the characteristics of a large diffusion coefficient of hydrogen itself and fast flame-propagation speed, the temperature is increased, and the fuel combustion quality is improved. As a result, thermal efficiency is improved and the equivalent fuel consumption is reduced.

3.3.3. Analysis of Emission Characteristics

Figure 18 shows the effect of main injection timing on NOx and CO₂ emissions of a dual-fuel engine. From Figure 18, CO₂ emissions fall as the main injection timing advances while NOx emissions rise. Compared with the TDC, when the main injection timing is −2, −4, −6, and −8 (°CA ATDC), the CO₂ emission decreases by 4.6%, 5.8%, 6.5%, and 8.2%, respectively. NOx emissions increased by 14.1%, 25.5%, 36.5%, and 44.5%, respectively.

The mixing and combustion state of the gas mixture during combustion can be improved by moving up the main injection timing. Since the ignition delay of pilot diesel oil is prolonged, and a more combustible mixture is formed, the number and energy of ignition points in the cylinder are increased and the combustion is improved. Moreover, by early injection of fuel for ignition, the mixture is ignited earlier in the cylinder and combustion begins earlier. The mixture has more time to fully burn, reducing unburned fuel residue, and enhancing combustion efficiency lowers the amount of fuel needed to produce the same amount of energy, which lowers CO₂ emissions.

With the main injection timing shift, the fuel injection is earlier, and more mixture is formed, which increases the pressure and temperature and accelerates the generation of NOx emissions. The homogeneous mixture accelerates the combustion rate in the cylinder and shortens the rapid combustion period, thus providing more time for the generation of NOx emissions in the later combustion period, which leads to the increase of NOx emissions. Therefore, the NOx emissions can be effectively reduced by delaying the main injection timing. If the primary injection timing is delayed, combustion will take place during the piston’s descent phase, when the temperature and pressure in the cylinder are low, which is helpful to slow down the rate of NOx emissions generated by the reaction between nitrogen and oxygen.

4. Conclusions

The combustion characteristics, economy, and emission characteristics of a hydrogen–diesel dual-fuel engine under different loads are studied experimentally and the effects of the EGR rate and main injection timing on the performance of a dual-fuel engine under high loads and maximum hydrogen substitution rate are studied. The following are the primary conclusions:

• The larger the load, the smaller the maximum hydrogen substitution rate that can be added. Under each load, compared to pure diesel, as the hydrogen substitution rate increases, the combustion becomes stronger, the flame-propagation speed becomes faster, the peak cylinder pressure and HRR both increase and the corresponding peak phase shifts forward. The BSFC_equ has decreased, with a maximum reduction of 4.5% and a minimum BSFC_equ of 196.57 g/(kW·h). The thermal efficiency has increased, with a maximum increase of 4.6% and a maximum thermal efficiency of 43.1%. CO₂ emissions decreased by a maximum reduction of 35.2%, while NOx emissions decreased at medium to low loads, while at high loads, the maximum increase was 20.1%;
• With the increase of the EGR rate, more exhaust gas is introduced, resulting in a decrease in oxygen content and temperature in the cylinder. The peak cylinder pressure and HRR both decrease and the corresponding peak phase moves backward, the BSFC_equ increases, and the thermal efficiency decreases, with a maximum decrease of 1.3%. CO₂ emissions have increased, with a maximum increase of 6.6%. NOx emissions have decreased, with a maximum reduction of 63.1%;
• With the advance of main injection timing, the mixture formed early near TDC to participate in combustion, the peak cylinder pressure and HRR both increase, and the corresponding peak phases move forward, the BSFC_equ decreases, and the thermal efficiency increases, with the maximum increase of 8.5%. CO₂ emissions decrease and NOx emissions increase, with a maximum increase of 44.5%. Therefore, NOx emissions in a dual-fuel engine can be reduced by increasing the EGR rate and delaying main injection timing.