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Study of injection pressure couple with EGR on combustion performance and emissions of natural gas-diesel dual-fuel engine
Fuel 261 (2020) 116409
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Full Length Article
Study of injection pressure couple with EGR on combustion performance and emissions of natural gas-diesel dual-fuel engine
T
Yingjie Chena, Zan Zhub, Yajuan Chena, Haozhong Huanga, , Zhaojun Zhua, Delin Lva, Mingzhang Pana, Xiaoyu Guoa ⁎
a b
College of Mechanical Engineering, Guangxi University, Nanning 530004, China Guangxi Yuchai Machinery Co., Ltd, Yulin 537005, China
ARTICLE INFO
ABSTRACT
Keywords: Dual-fuel engine EGR Injection pressure Natural gas Performance
Natural gas (NG) dual-fuel engines can attain a similar thermal efficiency as that of diesel engines while achieving lower emissions. However, the trade-off relationship between CH4 and NOX emissions limits the development of dual-fuel (DF) engines. In order to resolve this problem, the effects of injection pressure (IP) and exhaust gas recirculation (EGR) ratio on the combustion and emission of diesel/NG dual-fuel engines are investigated in this study. The results show that the diesel/NG dual fuel has a distinct three-stage heat release characteristic in the high-temperature combustion process. As the injection pressure increases, the flame propagation speed of methane and the indicated thermal efficiency (ITE) increase. However, the methane in the crevice region and cylinder wall cannot be ignited because of low temperature; these are the main sources of methane emissions. When the EGR rate increases, the indicated thermal efficiency first increases and there after decreases, and diesel is cleaved through the reaction chain, PC4H9 → C2H4 → CH4, to produce CH4; this becomes one of the reasons for CH4 emissions. When the EGR rate is small (< 10%), CO and CH4 emissions are not sensitive to IP; when the EGR rate is large (> 30%), CO and CH4 emissions can be significantly reduced by increasing the IP. Accordingly, when the injection pressure is 160 MPa and the EGR rate is 20%, the diesel/NG DF engine can achieve higher ITE and lower emissions.
1. Introduction As emission regulations become more stringent and the cost of diesel fuel rises, among the solutions to environmental and economic problems is the use of alternative fuels in diesel engines. Natural gas (NG) is considered as one of the most potential alternative fuels because it offers several benefits, such as clean combustion, competitive price, wide range of sources, and large reserves [1]. Compared with the traditional engine, dual-fuel (DF) engines can operate under lower temperature and pressure conditions. The NG has a relatively high octane number (ON); its introduction can prolong the ignition delay (ID) and facilitate the complete consumption of fuel. However, NG is a gaseous fuel with a high auto-ignition temperature, and the aeration efficiency is reduced when it is injected into the intake port; hence, diesel is required as an ignition source to ignite the NG [2,3]. The use of diesel to ignite natural gas can reduce emissions, such as NOX, soot, and CO2 while ensuring thermal efficiency. However, unburned CH4 emissions from diesel/NG DF engines are significant under low load conditions. It is well known that unburned CH4 is a ⁎
greenhouse gas and its global warming potential is 25 times that of CO2 [4,5]; hence, the reduction of unburned methane emission is a critical problem in diesel/NG DF engines. Increasing the injection pressure (IP) can effectively improve the atomization degree of fuel. This can improve the mixing quality of the mixture, make the combustion more complete, and ultimately improve the economy of the DF engine as well as reduce unburned HC and CO emissions. Yousefi et al. [6] combines the experiment and simulation to explore the effect of fuel injection parameters on the performance and emission of diesel/NG DF engine. They concluded that without split injection, the increased IP had little effect on thermal efficiency, but could significantly reduce the emissions of unburned methane. Chang et al. [7] compared the emissions resulting from a homogeneous charge-induced ignition mode and pure diesel mode with different fuel injection strategies. The results indicate that the dual-fuel model with a low IP (60–100 MPa) can achieve soot emissions similar to that of the pure diesel model with a high IP (120–140 MPa). Song et al. [8] compared the emissions of pure diesel mode and DF mode under low IP conditions and reported that although the IP was 20 MPa, ultra-low
Corresponding author. E-mail address: [email protected] (H. Huang).
https://doi.org/10.1016/j.fuel.2019.116409 Received 21 July 2019; Received in revised form 8 September 2019; Accepted 9 October 2019 0016-2361/ © 2019 Elsevier Ltd. All rights reserved.
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Nomenclature ATDC CA CA10 CA50 CFD CH4 CN CO DF EGR ER HTC HRR Hucng Hudiesel
ICP ICT ID IP ITE LHV LTC MPRR mcng mdiesel mdual NG NO NOX ON TDC
after top dead center crank angle crank angle corresponding to 10% of total heat release crank angle corresponding to 50% of total heat release computational fluid dynamics methane cetane number carbon monoxide dual fuel exhaust gas recirculation equivalence ratio high-temperature combustion heat release rate low heat value of total natural gas low heat value of diesel
NOX and soot emissions were still obtained in the DF model. Ryu et al. [9] used biodiesel as an ignition agent to perform the function of biodiesel/NG DF on combustion and emission of diesel engines at different pre-injection pressures. Results show that as the pre-injection pressure increases, soot emissions decrease, whereas NOX emissions increase. Based on the foregoing, it can be observed that unburned methane and CO emissions can be reduced by increasing the IP; however, the trade-off relationship between unburned methane and NOX still exists. The research conducted by Sanghoon et al. [10] indicates that when the combustion temperature is lower than 1650 K, the formation of an area where nitrogen oxides and soot accumulate is prevented, that is, a practically zero emission of nitrogen oxides and soot is achieved. Accordingly, to resolve the trade-off relationship between unburned CH4 and NOX and concurrently reduce the emissions of NOX and unburned CH4, it is necessary to achieve a homogeneous charge compression ignition (HCCI) low-temperature combustion (LTC) mode in the diesel engine. The LTC has two distinct characteristics [11]: long ID and low combustion temperature. LTC can reduce NOX emissions, and a longer ID can promote fuel and gas mixing; consequently, more NG can be oxidized. Therefore, the LTC technology can simultaneously decrease NOX and unburned methane emissions from diesel/NG DF engines. In this regard, the most effective way to reduce combustion temperature is the introduction of a large proportion of exhaust gas recirculation (EGR). Abdelaal et al. [12] investigated the impact of EGR on diesel/NG DF combustion and emission. They concluded that compared with pure diesel, the ID period is prolonged using dual fuel. Moreover, the ID period further increases as the EGR rate increases, and the maximum pressure rise rate (MPRR) is effectively reduced; NOX emissions are reduced as the EGR rate increases under different loads. Under heavy load conditions, when the EGR rate is 20%, the NOX specific emissions of pure diesel fuel can be reduced from 22 to 5 g/kW·h; CO and HC exhibit opposite trends, but the growth rate is not large. Hosseinzadeh et al. [13] used availability analysis to compare the thermal, chemical, and radical effects of EGR gases. The simulation results indicate that when the EGR rate is low, thermal action is more important than chemical action; the different types of free radicals in the EGR also have a significant impact on enhancing the combustion and emission characteristics of the engine. Selim et al. [14,15] reported that the ID was prolonged by the use of EGR technology at a fixed injection time; this is beneficial to the improvement of mixture quality and enhancement of combustion condition. However, the decrease in oxygen concentration leads to the deterioration of combustion. When the EGR rate is less than 5%, the thermal efficiency increases with the decrease in pressure rise rate. The pressure rise rate and thermal efficiency simultaneously
in-cylinder pressure in-cylinder temperature ignition delay injection pressure indicated thermal efficiency low heat value low-temperature combustion maximum pressure rise rate mass of natural gas mass of diesel total mass of fuel natural gas nitric oxide nitrogen oxide octane number top dead center
decrease with the EGR increase particularly when the EGR rate is more than 5%. The computational fluid dynamics (CFD) software is coupled with the chemical kinetic mechanism for simulation; this is a typical technique for the in-depth exploration of emission characteristics and combustion phenomena in the engine. In the previous work of our research group, the CFD software is coupled with a simplified diesel/NG DF mechanism for calculation, and the effects of injection ratio and multiple injection timing on the combustion and emission of DF engine are investigated [1]. The simulation is performed on the basis of experimental values, and simulation data are combined with experimental data to explain the emission characteristics and combustion phenomena in the DF engine. Mattarelli et al. [16] performed a numerical simulation on a 4-cylinder 2.8-L high-speed diesel engine at three speeds by coupling the KIVA-3V software and diesel/NG DF reaction mechanism, which contains 81 species and 421 reactions. Results show that the mean effective pressure of the original engine can be reached by using a diesel/NG DF engine; the NG substitution ratio and diesel injection time corresponding to different loads are also simulated. Shu et al. [17] coupled a simplified diesel/NG DF mechanism with CONVERGE software to explore the effect of spray cone angle on the emission from and combustion in the DF engine. The cloud map of NOX, CH4, CO, and some active molecules at different spray cone angles are given by simulation results, and the combustion and emission characteristics of diesel/NG DF engine are explained by combining the mechanisms. In summary, solving the trade-off relationship of NOX-CH4 is a hot spot in the research of diesel/NG DF engines. Increasing the IP can increase thermal efficiency and reduce the emission of unburned methane, but higher IP will promote the combustion temperature in the cylinder to increase the NOX emission. The introduction of EGR is an effective means of reducing NOX emissions. However, the coupling effect of IP and EGR is not clear, so this is the focus of this study. In this study, a diesel engine equipped with a natural gas intake system is tested, and the combustion and emission from the diesel/NG Table 1 Main parameters of test engine. Compression ratio Bore × stroke Connecting rod length Number of valves Speed Displacement Start of injection
2
17.5 123 mm × 145 mm 224 mm 4 1720 r/min 10.3 L −14 CA ATDC
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consumption is introduced, and the NG mass consumed is converted into diesel mass. The conversion formula is as follows:
Table 2 Diesel and natural gas properties [20,21]. Properties
Diesel
Air-fuel ratio (kg/kg) CN ON Lower heating value (MJ/kg) Auto-ignition temperature (°C) Boiling point (°C) Rich burn limit Lean burn limit
14.3 52.5 20–30 43.5 250 180–360 7.6 1.4
Natural gas
mdual = mdiesel +
16.4 – 130 50 650 −162 13.9 5.0
Model
Wall interaction Spray breakup Turbulence model Evaporation Turbulence Combustion
Walljet1 [25] WAVE [23] Renormalized group (RNG) k–ε Dukowicz [24] k–zeta–f [26] Chemkin
mcng
(2)
The ITE is calculated by the following formula: it
=
3600 Pi mdiesel × H µdiesel + mcng × Hµcng
(3)
where Pi is the indicated power, Hµdiesel is the low heat value for diesel, Hµcng is the low heat value for NG, and mdiesel and mcng are the masses of diesel and NG, respectively. In this experiment, the effects of EGR rate and IP on the emission and combustion of a DF engine are investigated under the following conditions: speed, 1720r/min; torque, 370 N·m (approximately 22% of load); diesel ignition amount, 27 mg/cycle; NG substitution rate, 50%. Moreover, the EGR rate varies from 0% to 40%, and the IPs are 80, 120, and 160 MPa.
Table 3 Computational models. Description
Hµcng H µdiesel
3. Numerical models 3.1. Spray modeling During the breakup process, the droplets collide and evaporate at the same time as they are atomized; the droplet change on the wall face also performs a certain function [22]. The Euler-Lagrange method is employed to simulate diesel spray, and the WAVE breakup model [23] suitably simulates the nozzles because the injector is a multi-hole type. The Dukowicz evaporation model [24] assumes that the heat and mass transfer processes are completely similar. Moreover, compared with other evaporation models, it does not require iteration; this considerably reduces the calculation time. The phenomenon of droplet wall impact is described by the Walljet1 model [25]. 3.2. Turbulence modeling Combustion in the engine cylinder is classified as turbulent combustion. Turbulence performs the mass and heat transfer functions during the combustion process. In this study, the k–zeta–f turbulence model [26] is used to calculate the in-cylinder flow. The k–zeta–f model has a higher accuracy and stability compared with the k–ε model; its computational time is only 15% more than that of the k–ε two–equation turbulence model.
Fig. 1. Calculated grid of 45° sector combustion chamber model.
DF engine are mainly explored. The CFD software coupling kinetic mechanism is used for the simulation calculation; moreover, from the microscopic point of view, the DF combustion mechanism and formation mechanism of emissions are described. 2. Experimental setup and procedure
3.3. Combustion modeling
The 6-cylinder turbocharged and inter-cooled diesel engine is used in the experiment. The NG supply system is installed on the original machine, and the matching design of the electronic control unit, which can accurately control the gas intake of NG, is performed. The main parameters of the test engine are summarized in Table 1; the main properties of the two fuels are listed in Table 2. In the experiment, the external EGR is used to control the EGR rate by adjusting the EGR valve. In view of the high temperature of the imported exhaust gas, it is cooled by the inter-cooler while preheating the fresh charge. The detailed description of the experimental device and the uncertainty of the instrument can be found in Ref. [18], and that of the test equipment and measuring instrument in Ref. [19]. The EGR rate is calculated by measuring the CO2 mole fractions in the exhaust gas and intake gas. The calculation formula is as follows:
REGR =
in out
× 100%
In order to explain the combustion phenomenon in the diesel/NG DF engine from a small molecule point of view, the calculation of AVLFIRE V2014 coupled with Chemkin is presented in this paper. The mechanism used in this study is a diesel/NG simplification mechanism constructed by our research group that includes 143 species and 746 reactions [19]. Among these, diesel is represented by n-heptane, and natural gas is represented by methane. Table 3 lists the sub-models used in the calculation. 3.4. Computational domain The calculation model in this study is generated by FIRE ESE-Diesel platform. The nozzle diameter of the injector is 0.121 mm, and the number of nozzle holes is eight; the calculation process is from the intake valve closing (−156°CA ATDC) to exhaust valve opening (140°CA ATDC). Because of the symmetry of combustion chamber and the center position of injector on the cylinder head, the number of diesel injection holes is eight. In order to improve the calculation efficiency, only oneeighth of the combustion chamber model is used (see Fig. 1).
(1)
where in and out are the mole fractions of CO2 in the intake and exhaust gases, respectively. Additionally, in this study, the concept of equivalent fuel 3
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Fig. 2. Experimental data of ICP and HRR compared with simulation results.
4. Results and discussion
the figure, experimental data agreed well with simulation results. The experimental and predicted results of CO, CH4, and NOX for diesel/NG DF engine under an 80-MPa IP and 30% EGR rate are compared in Fig. 3. As shown in the figure, the simulation data agree well with
Fig. 2 presents the experimental data of the ICP and HRR compared with the simulation of diesel/NG DF engine. As can be observed from 4
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The experimental data of the variations of HRR and ICP with different EGR rates for NG/diesel DF engine at a fixed IP (80 MPa) are shown in Figs. 6; 7 presents the corresponding simulation results. As shown in Fig. 6, the peak HRR and maximum pressure decrease as the EGR rate increases. This is because of the presence of a large number of polyatomic molecules, such as H2O and CO2 in the waste gas. Their addition decreases the oxygen concentration and increases the heat capacity of the intake charge in the cylinder, inhibits the complete combustion of the fuel, and decreases the HRR peak. The results in a decrease in the in-cylinder temperature (ICT) and leads to a drop in the in-cylinder burst pressure. This effect becomes more pronounced as the EGR increases. Fig. 8 presents the calculation results of the ITC, HRR, and key components as a function of CA at an 80-MPa IP and 10% EGR rate; the injection timings and durations are also shown (green lines). In the LTC stage, n-heptane undergoes a negative temperature coefficient phenomenon and a low temperature reaction because in the LTC zone, the in-cylinder temperature is lower, the oxygen content is higher, and the probability of n-heptane contact with oxygen increases. Therefore, nheptane dehydrogenates with oxygen to form C7H15 and HO2 (see Fig. 8), that is, the reaction, C7H16 + O2 = C7H15 + HO2, occurs. Hightemperature combustion (HTC) is mainly divided into three stages: the first stage is defined as the CA from the peak of the n-heptane concentration to the peak of the HRR; the second is defined as the CA from the peak of the HRR to the complete consumption of C7H16; the third phase is defined as the CA from the complete consumption of C7H16 to the end of in-cylinder combustion. The first phase of HTC is the premixed combustion of NG and diesel. At this stage, n-heptane is rapidly consumed mainly by the reaction, C7H16 + HO2 = C7H15 − 2 + H2O2, a large amount of heat is released, and the temperature increases. When the temperature reaches 1000 K, H2O2 is decomposed into two OH radicals by the reaction, H2O2 + M = OH + OH + M, the OH concentration rises and collides with CH4, and the NG content decreases. The HCO + O2 = CO + HO2 reaction pathway produces CO; its main depletion reaction is CO + OH = CO2 + H. The temperature continues to rise, the H2O2 completely breaks, and a large amount of OH accumulates. The OH reacts with most of the natural gas; this releases a large amount of heat, increases the temperature, and promotes the oxidation of CO (CO + HO2 = CO2 + OH, CO + OH = CO2 + H). In the second stage, as the injection continues, the amount of diesel entering the cylinder continues to increase and replenish the amount of CO in the cylinder; hence, the net CO content increases, but the generation rate decreases. The third stage of HTC is the premixed combustion of NG. The exothermic heat at this stage is all produced by natural gas combustion, and the temperature gradually rises to a higher level, which considerably promotes the oxidation of CO (CO + OH = CO2 + H); moreover, CO and OH concentrations are simultaneously reduced. Hyunhook et al. [27] indicated that the excessive global equivalent ratio (ER) is the main cause of the low thermal efficiency and high emission of the diesel engine. Therefore, a higher local equivalent ratio is used to solve this problem. In other words, the stratified combustion of the mixture can achieve higher thermal efficiency while reducing emissions. The ID period is defined as the CA from the start of injection to the CA10. In DF mode, the ID is closely related to the ER in the cylinder and the activity of reactants [28,29]. Fig. 9 shows the variation of ID under different IPs and EGR rate. It can be seen from that, with the increase of IP, the atomization effect of diesel is better, and a more uniform mixture is formed, the activity of reactants increases and the ID period is shortened. With the increase of EGR rate, the ID period increased first, then decreased and finally increased. This is because under the condition of
Fig. 3. Comparison between experimental and predicted values of CO, CH4, and NOX for diesel/NG DF engine at 80-MPa IP and 30% EGR rate.
Fig. 4. Experimental ICP and HRR of DF engine at different IPs.
experimental data; however, certain errors exist. The main reasons for this are as follows: diesel is replaced by two components, whereas NG is only replaced by methane; additionally, the influence of turbulence on chemical kinetics is not fully taken into consideration in the simulation process; finally, there is an error in the measurement of experimental values [1]. The experimental data of ICP and HRR of the diesel/NG DF engine with an EGR ratio of 30% are shown in Figs. 4; 5 shows the corresponding calculation results. Fig. 4 shows that the peak HRR and maximum ICP increase with the increase in IP when the EGR rate is 30%. On the one hand, a higher IP can improve the atomization and evaporation of fuel beam; consequently, the kinetic energy of fuel beam increases and improves the space utilization rate. On the other hand, the same mass is injected at a higher IP, which is equivalent to shortening the injection duration; consequently, fuel and air can mix more fully.
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Injection pressure (MPa)
Temperature
OH
80
120
160
Fig. 5. In-cylinder temperature and OH radical distribution at different IPs.
efficiency, and increases the ITE. As the EGR rate increases, the ITE first increases and thereafter decreases. This results from the increase in the EGR rate, the prolonged mixing time of diesel and NG/air mixture, and the large premixing ratio of the mixture, which promotes complete fuel combustion and ITE increase. When the EGR rate is further increased, more H2O and CO2 molecules in the waste gas replace the oxygen molecules of the intake gas and reduce the oxygen concentration in the intake charge; the polyatomic molecules have a higher specific heat capacity, which leads to the decrease in the ICT and deterioration of combustion in the cylinder. Additionally, CA50 deviates from the top dead center (TDC) (see Fig. 10(b)), and ITE decreases. Fig. 12 shows the experimental data of MPRR with the change in the EGR ratio of diesel/NG dual-fuel engine at different IPs. As shown in the diagram, the MPRR increases with the increase in IP. This is because the higher the IP, the better the atomization effect of the fuel, the more homogeneous mixture can be produced, the premixed combustion proportion is increased, and CA50 advances (see Fig. 10(b)); consequently, the MPRR increases. As the EGR rate increases, the MPRR initially decreases, thereafter increases, and finally decreases. When the EGR rate is small (< 10%), the ignition delay period does not considerably vary, and combustion is mainly affected by oxygen concentration; accordingly, the MPRR decreases as the EGR increases. At a moderate EGR rate, the MPRR increases as the EGR rate increases. This is because as the EGR rate increases, the ID period is prolonged, the premixing ratio of mixture increases, and CA50 approaches the TDC (see Fig. 10(b)), causing the MPRR to rise. At a large EGR rate, the combustion in the cylinder is severely anoxic, which leads to the deterioration of combustion; accordingly, the MPRR decreases. The main component of nitrogen oxide in an internal combustion engine is NO [30], which is mainly derived from nitrogen in the air involved in combustion. The emissions of NO can be divided into three categories: thermal NO, prompt NO, and fuel NO. The experimental data of NOX with the change in the EGR ratio of diesel/NG dual-fuel
Fig. 6. Experimental ICP and HRR of DF engine at different EGR rates.
small EGR rate (< 10%), the oxygen content in the cylinder decreases, the activity of reactants decreases, and the ID period is prolonged. Under the condition of large EGR rate (< 20%), the combustible mixture is diluted, the equivalent ratio increases, the reaction activity increases [29], and the ID period is shortened. With the further increase of EGR rate, the effect of oxygen content in cylinder is greater than that of equivalent ratio. At this time, the oxygen content in cylinder is very little, the activity of reactant is decreased, and the ID period is prolonged. Fig. 10 presents the indicated thermal efficiency (ITE) and CA50 experimental data of diesel/natural gas DF engine with the change in EGR rate at different IPs. The figure shows that as the IP increases, fuel atomization is promoted, the local concentration area is reduced, the local ER is increased, this promotes the mixing of diesel and NG/air mixture, accelerates the combustion rate of CH4 (see Fig. 11), it is helpful to complete combustion of the fuel, improves combustion
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EGR (%)
Temperature
OH
0
10
20
30
40
Fig. 7. In-cylinder temperature and OH radical distribution at different EGR rates.
engine at different IPs are presented in Fig. 13. The diagram shows that NOX emissions increase as the IP increases. This is because as the IP rises, the atomization effect of diesel is better and the stratification is not obvious, but more combustible mixtures are formed, complete fuel combustion is promoted, and the ICT rises (see Fig. 5); consequently, NOX emissions increase. As the EGR rate increases, the amount of exhaust gas that enters the cylinder increases; this causes the oxygen concentration and in-cylinder combustion temperature to decrease; thus, NOX emission is reduced. Fig. 13 shows that when the EGR rate is small, the IP has a considerable impact on NOX emission. This is because at a small EGR rate, the increase in IP aids in the mixing of diesel and NG; this increases the combustion rate, CA50 advances (see Fig. 10(b)), cylinder temperature rises (see Fig. 5), and NOX emissions increase. When the EGR rate increases, the ignition delay period is prolonged, the fuel and air is fully mixed, and the effect of increasing IP on the mixing process is weakened. Fig. 13 further shows that the effect of IP on NOX emission is less than that of the EGR rate. Fig. 14 shows the main path of CO generation and consumption in diesel/NG dual-fuel mode at 1200 K and different EGR ratios. The path analysis is calculated by a transient closed homogeneous batch reactor of the SENKIN program [31] under homogeneous, adiabatic, and constant volume conditions. During combustion, CO is the main intermediate product of hydrocarbon fuel that is mainly formed by reactions
HCO + O2 = HO2 + CO and HCCO + O2 = OH + 2CO. If the oxygen concentration and temperature of the mixture are sufficiently high and reaction time is adequate, then CO will be oxidized to CO2, and the oxidation reactions of CO are mainly CO + HO2 = CO2 + OH and CO + OH = H + CO2. Fig. 15 presents the experimental data of CO emissions from diesel/ NG DF engines at different IPs and EGR rates. The increase in IP promotes the mixing of diesel and intake charge, and the local ER is increased, which is conducive to the formation of a homogeneous mixture. This causes the advance in CA50 (see Fig. 10(b)) and the increase in in-cylinder combustion temperature (see Fig. 5) to promote CO oxidation. Therefore, CO emissions decrease with the IP increase. When the EGR rate increases, CO emissions exhibit an upward trend. This is because with the increase in the EGR rate, the OH radical concentration decreases (see Fig. 7). This promotes the reaction, HCCO + O2 = OH + 2CO, in the positive direction and suppresses the reaction, CO + OH = H + CO2, in the same reaction direction; accordingly, CO emissions increase. On the other hand, polyatomic molecules, such as CO2 and H2O, have high heat capacities. After the addition of waste gas, the total heat capacity in the cylinder increases, and the combustion rate decreases; this results in a decrease in the ICT. Therefore, the increase in the EGR rate can inhibit the oxidation of CO and lead to the increase in CO emissions. Fig. 15 shows that CO emissions are insensitive to the IP when the
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Fig. 8. Variations in temperature, HRR, and key components in cylinder at different CAs.
emissions decrease with the increase in the IP; this occurs because of the reasons previously outlined. It is noteworthy that the reaction chain, C7H16 → C7H15-2 → PC4H9 → C2H4 → CH4, is the main pathway for methane formation from diesel pyrolysis. As the EGR rate increases, the amount of CH4 that is produced by diesel through the branch chain reaction, PC4H9 → C2H4 → CH4, increases (see Fig. 14) and results in increased methane emissions. Fig. 16 further shows that the injection pressure has a slight effect on methane emissions at low EGR rates. When the EGR rate is low, the oxygen content is relatively sufficient, which aids in the complete combustion of CH4; this decreases CH4 emissions. As the EGR rate increases, the oxygen content in the cylinder decreases, and the injection pressure rises; this aids in the oxidation reaction between the fuel and oxygen molecules; this improves the oxidation rate of methane. Therefore, the IP has a more significant influence on CH4 emissions when the EGR rate is large. Fig. 17 presents the calculation results of methane distribution in the cylinder at different IPs; the corresponding slices of turbulent kinetic energy at different IPs are shown in Fig. 18. The figure shows that when the IP increases, the turbulent kinetic energy in the cylinder and the flame propagation speed of methane increases (see Fig. 11); this contributes to the combustion of methane in the cylinder. However, the mixture near the cylinder wall and the crevice region does not ignite because of its low temperature. Although this part of the gas re-enters the cylinder during the expansion stroke, the in-cylinder temperature is already low at this instance, and the combustible mixture cannot be ignited. Consequently, unburned methane is discharged out of the cylinder along with the exhaust gas during the exhaust stroke. Thus, it is evident that the increase in injection pressure promotes
Fig. 9. Experimental ID at different IPs and EGR rates.
EGR rate is low. This is because at such rate, the oxygen content is high, the CA50 advances (see Fig. 10(b)), and the in-cylinder combustion temperature increases (see Fig. 7), which promotes the oxidation of CO; consequently, CO emission is low. When the EGR rate is higher, there are more waste gases in the cylinder, and the specific heat capacity becomes larger. The increase in the IP promotes the mixing of fuel and intake charge, reduces the formation of a rich zone, and contributes to the oxidation of CO; thus, CO emissions are reduced. Fig. 16 shows the experimental data of CH4 emissions from diesel/ NG DF engines at different EGR rates and IPs. The figure shows that CH4 8
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(a) ITE
(b) CA50
Fig. 10. Experimental ITE and CA50 at different IPs and EGR rates.
Fig. 11. Combustion rate of CH4.
Fig. 13. Experimental NOX emissions at different IPs and EGR rates.
DF engine was carried out. As the injection pressure increased, CO and CH4 emissions decreased, whereas the MPRR, ITE, and NOX emissions increased. When the EGR rate increased, CH4 and CO emissions increased, NOX emissions decreased, the ITE first increased and thereafter decreased, whereas the MPRR first decreased, thereafter increased, and finally decreased. In addition, the increase in injection pressure increased the in-cylinder turbulent kinetic energy and the flame propagation speed of methane; this aided in the re-ignition of unburned methane in the cylinder and consumed most of the unburned methane. At low EGR rates, the injection pressure had a lower impact on CO and CH4 emissions. The CO and CH4 emissions significantly decreased as the IP increased under high EGR rate conditions. It was found by the path analysis that, as the EGR rate increased, the amount of CH4 produced by diesel through the branch chain reaction, PC4H9 → C2H4 → CH4, increased; this became one of the sources of CH4 emissions. The natural gas in the cylinder wall and crevice region was not oxidized because of the low temperature in the cylinder wall. Thus, the NG in the crevice region and cylinder wall became the main source of unburned CH4 emissions. It is worth pointing out, the diesel/NG DF had a distinct three-stage heat-release characteristic in the HTC process, and the diesel/NG dual-fuel engine can obtain higher thermal efficiencies and lower emissions at a 160-MPa IP and 20% EGR ratio. Furthermore, under the condition of introducing large proportion of EGR, the combustion of DF engine is very unstable. Therefore, in order to further understand the effect of fuel injection pressure coupling EGR rate on diesel/NG DF engine, the cycle-to-cycle variation of diesel/NG DF engine should be explored in the future.
Fig. 12. Experimental MPRR at different IPs and EGR rates.
natural gas combustion in the center of the cylinder, but the remaining methane in the crevice region and cylinder wall becomes the main source of unburned methane emissions. 5. Conclusion In this study, the effects of different IPs (80, 120, and 160 MPa) and EGR ratios (0%–40%) on the emissions and combustion of a diesel/NG
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Fig. 14. Main path of CO generation and consumption in diesel/NG dual-fuel mode at different EGR rates.
Fig. 16. Experimental CH4 emissions at different injection pressures and EGR rates.
Fig. 15. Experimental CO emissions at different IPs and EGR rates.
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Crank angle
80 MPa
(°CA ATDC)
120 MPa
160 MPa
20
60
80
Fig. 17. Distribution of methane in cylinder at different IPs.
80 MPa
120 MPa
160 MPa
Turbulent Kinetic Energy
Fig. 18. Slices of equivalent ratio and turbulent kinetic energy at different IPs at TDC.
Acknowledgements
[6] Yousefi A, Guo H, Birouk M, et al. On greenhouse gas emissions and thermal efficiency of natural gas/diesel dual-fuel engine at low load conditions: coupled effect of injector rail pressure and split injection. Appl Energy 2019;242:216–31. [7] Chang Z, Yu C, Zhang H, et al. Experimental investigation of homogeneous charge induced ignition (HCII) with low-pressure injection to reduce Pm emissions in a heavy-duty engine. SAE Int J Engines 2016;9:1559–70. [8] Song H, Liu C, Li Y, et al. An exploration of utilizing low-pressure diesel injection for natural gas dual-fuel low-temperature combustion. Energy 2018;153:248–55. [9] Ryu K. Effects of pilot injection timing on the combustion and emissions characteristics in a diesel engine using biodiesel–CNG dual fuel. Appl Energy 2013;111:721–30. [10] Kook S, Bae C, Miles PC, et al. The influence of charge dilution and injection timing on low-temperature diesel combustion and emissions. SAE Trans 2005:1575–95. [11] Kavuri C, Kokjohn SL. Computational optimization of a reactivity controlled compression ignition (RCCI) combustion system considering performance at multiple modes simultaneously. Fuel 2017;207:702–18. [12] Abdelaal MM, Hegab AH. Combustion and emission characteristics of a natural gasfueled diesel engine with EGR. Energy Convers Manage 2012;64:301–12. [13] Hosseinzadeh A, Saray RK, Mahmoudi SS. Comparison of thermal, radical and chemical effects of EGR gases using availability analysis in dual-fuel engines at part loads. Energy Convers Manage 2010;51:2321–9. [14] Selim MY. Effect of exhaust gas recirculation on some combustion characteristics of dual fuel engine. Energy Convers Manage 2003;44:707–21. [15] Selim MY. A study of some combustion characteristics of dual fuel engine using EGR. SAE Trans 2003:1152–60. [16] Mattarelli E, Rinaldini CA, Golovitchev VI. CFD-3D analysis of a light duty Dual
This work is supported by National Natural Science Foundation of China (51966001) and Scientific Research and Technology Development Program of Guangxi (1598007-44). References [1] Huang H, Lv D, Zhu J, et al. Development of a new reduced diesel/natural gas mechanism for dual-fuel engine combustion and emission prediction. Fuel 2019;236:30–42. [2] Namasivayam AM, Korakianitis T, Crookes RJ, et al. Biodiesel, emulsified biodiesel and dimethyl ether as pilot fuels for natural gas fuelled engines. Appl Energy 2010;87:769–78. [3] Azimov U, Tomita E, Kawahara N, et al. Premixed mixture ignition in the end-gas region (PREMIER) combustion in a natural gas dual-fuel engine: operating range and exhaust emissions. Int J Engine Res 2011;12:484–97. [4] Yousefi A, Birouk M, Guo H. An experimental and numerical study of the effect of diesel injection timing on natural gas/diesel dual-fuel combustion at low load. Fuel 2017;203:642–57. [5] Forster P, Ramaswamy V, Artaxo P, et al. Changes in atmospheric constituents and in radiative forcing. Chapter 2. Climate Change 2007. The Physical Science Basis;2007.
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Y. Chen, et al. Fuel (Diesel/Natural Gas) combustion engine. Energy Procedia 2014;45:929–37. [17] Shu J, Fu J, Liu J, et al. Effects of injector spray angle on combustion and emissions characteristics of a natural gas (NG)-diesel dual fuel engine based on CFD coupled with reduced chemical kinetic model. Appl Energy 2019;233:182–95. [18] Huang H, Wang Q, Shi C, et al. Comparative study of effects of pilot injection and fuel properties on low temperature combustion in diesel engine under a medium EGR rate. Appl Energy 2016;179:1194–208. [19] Huang H, Zhu Z, Chen Y, et al. Experimental and numerical study of multiple injection effects on combustion and emission characteristics of natural gas–diesel dual-fuel engine. Energy Convers Manage 2019;183:84–96. [20] Wei L, Geng P. A review on natural gas/diesel dual fuel combustion, emissions and performance. Fuel Process Technol 2016;142:264–78. [21] Wang Z, Du G, Wang D, et al. Combustion process decoupling of a diesel/natural gas dual-fuel engine at low loads. Fuel 2018;232:550–61. [22] Baumgarten C. Mixture formation in internal combustion engines. Springer Science & Business Media; 2006. [23] Module LM. AVL FIRE® VERSION 2011 2011. [24] Butler TD, Cloutman LD, Dukowicz JK, et al. Multidimensional numerical
[25] [26] [27] [28] [29] [30] [31]
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simulation of reactive flow in internal combustion engines ☆. Prog Energy Combust Sci 1981;7:293–315. Naber J, Reitz RD. Modeling engine spray/wall impingement. 1988. Wilcox D. Turbulence modeling for CFD. La Canada: DCW Industries; 2006. Park H, Shim E, Bae C. Expansion of low-load operating range by mixture stratification in a natural gas-diesel dual-fuel premixed charge compression ignition engine. Energy Convers Manage 2019;194:186–98. Zhang J, Niu S, Zhang Y, et al. Experimental and modeling study of the auto-ignition of n-heptane/n-butanol mixtures. Combust Flame 2013;160:31–9. Jiang X, Deng F, Yang F, et al. Ignition delay characteristics and kinetic investigation of dimethyl ether/n-pentane binary mixtures: interpreting the effect of the equivalence ratio and dimethyl ether blending. Energy Fuel 2018;32:3814–23. Heywood J. Internal combustion engine fundamentals. New York: Mcgraw-Hill; 1988. Kee RJ, Rupley FM, Miller JA. Chemkin-II: a Fortran chemical kinetics package for the analysis of gas-phase chemical kinetics. Livermore, CA (USA): Sandia National Labs; 1989.
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Full Length Article
Study of injection pressure couple with EGR on combustion performance and emissions of natural gas-diesel dual-fuel engine
T
Yingjie Chena, Zan Zhub, Yajuan Chena, Haozhong Huanga, , Zhaojun Zhua, Delin Lva, Mingzhang Pana, Xiaoyu Guoa ⁎
a b
College of Mechanical Engineering, Guangxi University, Nanning 530004, China Guangxi Yuchai Machinery Co., Ltd, Yulin 537005, China
ARTICLE INFO
ABSTRACT
Keywords: Dual-fuel engine EGR Injection pressure Natural gas Performance
Natural gas (NG) dual-fuel engines can attain a similar thermal efficiency as that of diesel engines while achieving lower emissions. However, the trade-off relationship between CH4 and NOX emissions limits the development of dual-fuel (DF) engines. In order to resolve this problem, the effects of injection pressure (IP) and exhaust gas recirculation (EGR) ratio on the combustion and emission of diesel/NG dual-fuel engines are investigated in this study. The results show that the diesel/NG dual fuel has a distinct three-stage heat release characteristic in the high-temperature combustion process. As the injection pressure increases, the flame propagation speed of methane and the indicated thermal efficiency (ITE) increase. However, the methane in the crevice region and cylinder wall cannot be ignited because of low temperature; these are the main sources of methane emissions. When the EGR rate increases, the indicated thermal efficiency first increases and there after decreases, and diesel is cleaved through the reaction chain, PC4H9 → C2H4 → CH4, to produce CH4; this becomes one of the reasons for CH4 emissions. When the EGR rate is small (< 10%), CO and CH4 emissions are not sensitive to IP; when the EGR rate is large (> 30%), CO and CH4 emissions can be significantly reduced by increasing the IP. Accordingly, when the injection pressure is 160 MPa and the EGR rate is 20%, the diesel/NG DF engine can achieve higher ITE and lower emissions.
1. Introduction As emission regulations become more stringent and the cost of diesel fuel rises, among the solutions to environmental and economic problems is the use of alternative fuels in diesel engines. Natural gas (NG) is considered as one of the most potential alternative fuels because it offers several benefits, such as clean combustion, competitive price, wide range of sources, and large reserves [1]. Compared with the traditional engine, dual-fuel (DF) engines can operate under lower temperature and pressure conditions. The NG has a relatively high octane number (ON); its introduction can prolong the ignition delay (ID) and facilitate the complete consumption of fuel. However, NG is a gaseous fuel with a high auto-ignition temperature, and the aeration efficiency is reduced when it is injected into the intake port; hence, diesel is required as an ignition source to ignite the NG [2,3]. The use of diesel to ignite natural gas can reduce emissions, such as NOX, soot, and CO2 while ensuring thermal efficiency. However, unburned CH4 emissions from diesel/NG DF engines are significant under low load conditions. It is well known that unburned CH4 is a ⁎
greenhouse gas and its global warming potential is 25 times that of CO2 [4,5]; hence, the reduction of unburned methane emission is a critical problem in diesel/NG DF engines. Increasing the injection pressure (IP) can effectively improve the atomization degree of fuel. This can improve the mixing quality of the mixture, make the combustion more complete, and ultimately improve the economy of the DF engine as well as reduce unburned HC and CO emissions. Yousefi et al. [6] combines the experiment and simulation to explore the effect of fuel injection parameters on the performance and emission of diesel/NG DF engine. They concluded that without split injection, the increased IP had little effect on thermal efficiency, but could significantly reduce the emissions of unburned methane. Chang et al. [7] compared the emissions resulting from a homogeneous charge-induced ignition mode and pure diesel mode with different fuel injection strategies. The results indicate that the dual-fuel model with a low IP (60–100 MPa) can achieve soot emissions similar to that of the pure diesel model with a high IP (120–140 MPa). Song et al. [8] compared the emissions of pure diesel mode and DF mode under low IP conditions and reported that although the IP was 20 MPa, ultra-low
Corresponding author. E-mail address: [email protected] (H. Huang).
https://doi.org/10.1016/j.fuel.2019.116409 Received 21 July 2019; Received in revised form 8 September 2019; Accepted 9 October 2019 0016-2361/ © 2019 Elsevier Ltd. All rights reserved.
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Nomenclature ATDC CA CA10 CA50 CFD CH4 CN CO DF EGR ER HTC HRR Hucng Hudiesel
ICP ICT ID IP ITE LHV LTC MPRR mcng mdiesel mdual NG NO NOX ON TDC
after top dead center crank angle crank angle corresponding to 10% of total heat release crank angle corresponding to 50% of total heat release computational fluid dynamics methane cetane number carbon monoxide dual fuel exhaust gas recirculation equivalence ratio high-temperature combustion heat release rate low heat value of total natural gas low heat value of diesel
NOX and soot emissions were still obtained in the DF model. Ryu et al. [9] used biodiesel as an ignition agent to perform the function of biodiesel/NG DF on combustion and emission of diesel engines at different pre-injection pressures. Results show that as the pre-injection pressure increases, soot emissions decrease, whereas NOX emissions increase. Based on the foregoing, it can be observed that unburned methane and CO emissions can be reduced by increasing the IP; however, the trade-off relationship between unburned methane and NOX still exists. The research conducted by Sanghoon et al. [10] indicates that when the combustion temperature is lower than 1650 K, the formation of an area where nitrogen oxides and soot accumulate is prevented, that is, a practically zero emission of nitrogen oxides and soot is achieved. Accordingly, to resolve the trade-off relationship between unburned CH4 and NOX and concurrently reduce the emissions of NOX and unburned CH4, it is necessary to achieve a homogeneous charge compression ignition (HCCI) low-temperature combustion (LTC) mode in the diesel engine. The LTC has two distinct characteristics [11]: long ID and low combustion temperature. LTC can reduce NOX emissions, and a longer ID can promote fuel and gas mixing; consequently, more NG can be oxidized. Therefore, the LTC technology can simultaneously decrease NOX and unburned methane emissions from diesel/NG DF engines. In this regard, the most effective way to reduce combustion temperature is the introduction of a large proportion of exhaust gas recirculation (EGR). Abdelaal et al. [12] investigated the impact of EGR on diesel/NG DF combustion and emission. They concluded that compared with pure diesel, the ID period is prolonged using dual fuel. Moreover, the ID period further increases as the EGR rate increases, and the maximum pressure rise rate (MPRR) is effectively reduced; NOX emissions are reduced as the EGR rate increases under different loads. Under heavy load conditions, when the EGR rate is 20%, the NOX specific emissions of pure diesel fuel can be reduced from 22 to 5 g/kW·h; CO and HC exhibit opposite trends, but the growth rate is not large. Hosseinzadeh et al. [13] used availability analysis to compare the thermal, chemical, and radical effects of EGR gases. The simulation results indicate that when the EGR rate is low, thermal action is more important than chemical action; the different types of free radicals in the EGR also have a significant impact on enhancing the combustion and emission characteristics of the engine. Selim et al. [14,15] reported that the ID was prolonged by the use of EGR technology at a fixed injection time; this is beneficial to the improvement of mixture quality and enhancement of combustion condition. However, the decrease in oxygen concentration leads to the deterioration of combustion. When the EGR rate is less than 5%, the thermal efficiency increases with the decrease in pressure rise rate. The pressure rise rate and thermal efficiency simultaneously
in-cylinder pressure in-cylinder temperature ignition delay injection pressure indicated thermal efficiency low heat value low-temperature combustion maximum pressure rise rate mass of natural gas mass of diesel total mass of fuel natural gas nitric oxide nitrogen oxide octane number top dead center
decrease with the EGR increase particularly when the EGR rate is more than 5%. The computational fluid dynamics (CFD) software is coupled with the chemical kinetic mechanism for simulation; this is a typical technique for the in-depth exploration of emission characteristics and combustion phenomena in the engine. In the previous work of our research group, the CFD software is coupled with a simplified diesel/NG DF mechanism for calculation, and the effects of injection ratio and multiple injection timing on the combustion and emission of DF engine are investigated [1]. The simulation is performed on the basis of experimental values, and simulation data are combined with experimental data to explain the emission characteristics and combustion phenomena in the DF engine. Mattarelli et al. [16] performed a numerical simulation on a 4-cylinder 2.8-L high-speed diesel engine at three speeds by coupling the KIVA-3V software and diesel/NG DF reaction mechanism, which contains 81 species and 421 reactions. Results show that the mean effective pressure of the original engine can be reached by using a diesel/NG DF engine; the NG substitution ratio and diesel injection time corresponding to different loads are also simulated. Shu et al. [17] coupled a simplified diesel/NG DF mechanism with CONVERGE software to explore the effect of spray cone angle on the emission from and combustion in the DF engine. The cloud map of NOX, CH4, CO, and some active molecules at different spray cone angles are given by simulation results, and the combustion and emission characteristics of diesel/NG DF engine are explained by combining the mechanisms. In summary, solving the trade-off relationship of NOX-CH4 is a hot spot in the research of diesel/NG DF engines. Increasing the IP can increase thermal efficiency and reduce the emission of unburned methane, but higher IP will promote the combustion temperature in the cylinder to increase the NOX emission. The introduction of EGR is an effective means of reducing NOX emissions. However, the coupling effect of IP and EGR is not clear, so this is the focus of this study. In this study, a diesel engine equipped with a natural gas intake system is tested, and the combustion and emission from the diesel/NG Table 1 Main parameters of test engine. Compression ratio Bore × stroke Connecting rod length Number of valves Speed Displacement Start of injection
2
17.5 123 mm × 145 mm 224 mm 4 1720 r/min 10.3 L −14 CA ATDC
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consumption is introduced, and the NG mass consumed is converted into diesel mass. The conversion formula is as follows:
Table 2 Diesel and natural gas properties [20,21]. Properties
Diesel
Air-fuel ratio (kg/kg) CN ON Lower heating value (MJ/kg) Auto-ignition temperature (°C) Boiling point (°C) Rich burn limit Lean burn limit
14.3 52.5 20–30 43.5 250 180–360 7.6 1.4
Natural gas
mdual = mdiesel +
16.4 – 130 50 650 −162 13.9 5.0
Model
Wall interaction Spray breakup Turbulence model Evaporation Turbulence Combustion
Walljet1 [25] WAVE [23] Renormalized group (RNG) k–ε Dukowicz [24] k–zeta–f [26] Chemkin
mcng
(2)
The ITE is calculated by the following formula: it
=
3600 Pi mdiesel × H µdiesel + mcng × Hµcng
(3)
where Pi is the indicated power, Hµdiesel is the low heat value for diesel, Hµcng is the low heat value for NG, and mdiesel and mcng are the masses of diesel and NG, respectively. In this experiment, the effects of EGR rate and IP on the emission and combustion of a DF engine are investigated under the following conditions: speed, 1720r/min; torque, 370 N·m (approximately 22% of load); diesel ignition amount, 27 mg/cycle; NG substitution rate, 50%. Moreover, the EGR rate varies from 0% to 40%, and the IPs are 80, 120, and 160 MPa.
Table 3 Computational models. Description
Hµcng H µdiesel
3. Numerical models 3.1. Spray modeling During the breakup process, the droplets collide and evaporate at the same time as they are atomized; the droplet change on the wall face also performs a certain function [22]. The Euler-Lagrange method is employed to simulate diesel spray, and the WAVE breakup model [23] suitably simulates the nozzles because the injector is a multi-hole type. The Dukowicz evaporation model [24] assumes that the heat and mass transfer processes are completely similar. Moreover, compared with other evaporation models, it does not require iteration; this considerably reduces the calculation time. The phenomenon of droplet wall impact is described by the Walljet1 model [25]. 3.2. Turbulence modeling Combustion in the engine cylinder is classified as turbulent combustion. Turbulence performs the mass and heat transfer functions during the combustion process. In this study, the k–zeta–f turbulence model [26] is used to calculate the in-cylinder flow. The k–zeta–f model has a higher accuracy and stability compared with the k–ε model; its computational time is only 15% more than that of the k–ε two–equation turbulence model.
Fig. 1. Calculated grid of 45° sector combustion chamber model.
DF engine are mainly explored. The CFD software coupling kinetic mechanism is used for the simulation calculation; moreover, from the microscopic point of view, the DF combustion mechanism and formation mechanism of emissions are described. 2. Experimental setup and procedure
3.3. Combustion modeling
The 6-cylinder turbocharged and inter-cooled diesel engine is used in the experiment. The NG supply system is installed on the original machine, and the matching design of the electronic control unit, which can accurately control the gas intake of NG, is performed. The main parameters of the test engine are summarized in Table 1; the main properties of the two fuels are listed in Table 2. In the experiment, the external EGR is used to control the EGR rate by adjusting the EGR valve. In view of the high temperature of the imported exhaust gas, it is cooled by the inter-cooler while preheating the fresh charge. The detailed description of the experimental device and the uncertainty of the instrument can be found in Ref. [18], and that of the test equipment and measuring instrument in Ref. [19]. The EGR rate is calculated by measuring the CO2 mole fractions in the exhaust gas and intake gas. The calculation formula is as follows:
REGR =
in out
× 100%
In order to explain the combustion phenomenon in the diesel/NG DF engine from a small molecule point of view, the calculation of AVLFIRE V2014 coupled with Chemkin is presented in this paper. The mechanism used in this study is a diesel/NG simplification mechanism constructed by our research group that includes 143 species and 746 reactions [19]. Among these, diesel is represented by n-heptane, and natural gas is represented by methane. Table 3 lists the sub-models used in the calculation. 3.4. Computational domain The calculation model in this study is generated by FIRE ESE-Diesel platform. The nozzle diameter of the injector is 0.121 mm, and the number of nozzle holes is eight; the calculation process is from the intake valve closing (−156°CA ATDC) to exhaust valve opening (140°CA ATDC). Because of the symmetry of combustion chamber and the center position of injector on the cylinder head, the number of diesel injection holes is eight. In order to improve the calculation efficiency, only oneeighth of the combustion chamber model is used (see Fig. 1).
(1)
where in and out are the mole fractions of CO2 in the intake and exhaust gases, respectively. Additionally, in this study, the concept of equivalent fuel 3
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Fig. 2. Experimental data of ICP and HRR compared with simulation results.
4. Results and discussion
the figure, experimental data agreed well with simulation results. The experimental and predicted results of CO, CH4, and NOX for diesel/NG DF engine under an 80-MPa IP and 30% EGR rate are compared in Fig. 3. As shown in the figure, the simulation data agree well with
Fig. 2 presents the experimental data of the ICP and HRR compared with the simulation of diesel/NG DF engine. As can be observed from 4
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The experimental data of the variations of HRR and ICP with different EGR rates for NG/diesel DF engine at a fixed IP (80 MPa) are shown in Figs. 6; 7 presents the corresponding simulation results. As shown in Fig. 6, the peak HRR and maximum pressure decrease as the EGR rate increases. This is because of the presence of a large number of polyatomic molecules, such as H2O and CO2 in the waste gas. Their addition decreases the oxygen concentration and increases the heat capacity of the intake charge in the cylinder, inhibits the complete combustion of the fuel, and decreases the HRR peak. The results in a decrease in the in-cylinder temperature (ICT) and leads to a drop in the in-cylinder burst pressure. This effect becomes more pronounced as the EGR increases. Fig. 8 presents the calculation results of the ITC, HRR, and key components as a function of CA at an 80-MPa IP and 10% EGR rate; the injection timings and durations are also shown (green lines). In the LTC stage, n-heptane undergoes a negative temperature coefficient phenomenon and a low temperature reaction because in the LTC zone, the in-cylinder temperature is lower, the oxygen content is higher, and the probability of n-heptane contact with oxygen increases. Therefore, nheptane dehydrogenates with oxygen to form C7H15 and HO2 (see Fig. 8), that is, the reaction, C7H16 + O2 = C7H15 + HO2, occurs. Hightemperature combustion (HTC) is mainly divided into three stages: the first stage is defined as the CA from the peak of the n-heptane concentration to the peak of the HRR; the second is defined as the CA from the peak of the HRR to the complete consumption of C7H16; the third phase is defined as the CA from the complete consumption of C7H16 to the end of in-cylinder combustion. The first phase of HTC is the premixed combustion of NG and diesel. At this stage, n-heptane is rapidly consumed mainly by the reaction, C7H16 + HO2 = C7H15 − 2 + H2O2, a large amount of heat is released, and the temperature increases. When the temperature reaches 1000 K, H2O2 is decomposed into two OH radicals by the reaction, H2O2 + M = OH + OH + M, the OH concentration rises and collides with CH4, and the NG content decreases. The HCO + O2 = CO + HO2 reaction pathway produces CO; its main depletion reaction is CO + OH = CO2 + H. The temperature continues to rise, the H2O2 completely breaks, and a large amount of OH accumulates. The OH reacts with most of the natural gas; this releases a large amount of heat, increases the temperature, and promotes the oxidation of CO (CO + HO2 = CO2 + OH, CO + OH = CO2 + H). In the second stage, as the injection continues, the amount of diesel entering the cylinder continues to increase and replenish the amount of CO in the cylinder; hence, the net CO content increases, but the generation rate decreases. The third stage of HTC is the premixed combustion of NG. The exothermic heat at this stage is all produced by natural gas combustion, and the temperature gradually rises to a higher level, which considerably promotes the oxidation of CO (CO + OH = CO2 + H); moreover, CO and OH concentrations are simultaneously reduced. Hyunhook et al. [27] indicated that the excessive global equivalent ratio (ER) is the main cause of the low thermal efficiency and high emission of the diesel engine. Therefore, a higher local equivalent ratio is used to solve this problem. In other words, the stratified combustion of the mixture can achieve higher thermal efficiency while reducing emissions. The ID period is defined as the CA from the start of injection to the CA10. In DF mode, the ID is closely related to the ER in the cylinder and the activity of reactants [28,29]. Fig. 9 shows the variation of ID under different IPs and EGR rate. It can be seen from that, with the increase of IP, the atomization effect of diesel is better, and a more uniform mixture is formed, the activity of reactants increases and the ID period is shortened. With the increase of EGR rate, the ID period increased first, then decreased and finally increased. This is because under the condition of
Fig. 3. Comparison between experimental and predicted values of CO, CH4, and NOX for diesel/NG DF engine at 80-MPa IP and 30% EGR rate.
Fig. 4. Experimental ICP and HRR of DF engine at different IPs.
experimental data; however, certain errors exist. The main reasons for this are as follows: diesel is replaced by two components, whereas NG is only replaced by methane; additionally, the influence of turbulence on chemical kinetics is not fully taken into consideration in the simulation process; finally, there is an error in the measurement of experimental values [1]. The experimental data of ICP and HRR of the diesel/NG DF engine with an EGR ratio of 30% are shown in Figs. 4; 5 shows the corresponding calculation results. Fig. 4 shows that the peak HRR and maximum ICP increase with the increase in IP when the EGR rate is 30%. On the one hand, a higher IP can improve the atomization and evaporation of fuel beam; consequently, the kinetic energy of fuel beam increases and improves the space utilization rate. On the other hand, the same mass is injected at a higher IP, which is equivalent to shortening the injection duration; consequently, fuel and air can mix more fully.
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Injection pressure (MPa)
Temperature
OH
80
120
160
Fig. 5. In-cylinder temperature and OH radical distribution at different IPs.
efficiency, and increases the ITE. As the EGR rate increases, the ITE first increases and thereafter decreases. This results from the increase in the EGR rate, the prolonged mixing time of diesel and NG/air mixture, and the large premixing ratio of the mixture, which promotes complete fuel combustion and ITE increase. When the EGR rate is further increased, more H2O and CO2 molecules in the waste gas replace the oxygen molecules of the intake gas and reduce the oxygen concentration in the intake charge; the polyatomic molecules have a higher specific heat capacity, which leads to the decrease in the ICT and deterioration of combustion in the cylinder. Additionally, CA50 deviates from the top dead center (TDC) (see Fig. 10(b)), and ITE decreases. Fig. 12 shows the experimental data of MPRR with the change in the EGR ratio of diesel/NG dual-fuel engine at different IPs. As shown in the diagram, the MPRR increases with the increase in IP. This is because the higher the IP, the better the atomization effect of the fuel, the more homogeneous mixture can be produced, the premixed combustion proportion is increased, and CA50 advances (see Fig. 10(b)); consequently, the MPRR increases. As the EGR rate increases, the MPRR initially decreases, thereafter increases, and finally decreases. When the EGR rate is small (< 10%), the ignition delay period does not considerably vary, and combustion is mainly affected by oxygen concentration; accordingly, the MPRR decreases as the EGR increases. At a moderate EGR rate, the MPRR increases as the EGR rate increases. This is because as the EGR rate increases, the ID period is prolonged, the premixing ratio of mixture increases, and CA50 approaches the TDC (see Fig. 10(b)), causing the MPRR to rise. At a large EGR rate, the combustion in the cylinder is severely anoxic, which leads to the deterioration of combustion; accordingly, the MPRR decreases. The main component of nitrogen oxide in an internal combustion engine is NO [30], which is mainly derived from nitrogen in the air involved in combustion. The emissions of NO can be divided into three categories: thermal NO, prompt NO, and fuel NO. The experimental data of NOX with the change in the EGR ratio of diesel/NG dual-fuel
Fig. 6. Experimental ICP and HRR of DF engine at different EGR rates.
small EGR rate (< 10%), the oxygen content in the cylinder decreases, the activity of reactants decreases, and the ID period is prolonged. Under the condition of large EGR rate (< 20%), the combustible mixture is diluted, the equivalent ratio increases, the reaction activity increases [29], and the ID period is shortened. With the further increase of EGR rate, the effect of oxygen content in cylinder is greater than that of equivalent ratio. At this time, the oxygen content in cylinder is very little, the activity of reactant is decreased, and the ID period is prolonged. Fig. 10 presents the indicated thermal efficiency (ITE) and CA50 experimental data of diesel/natural gas DF engine with the change in EGR rate at different IPs. The figure shows that as the IP increases, fuel atomization is promoted, the local concentration area is reduced, the local ER is increased, this promotes the mixing of diesel and NG/air mixture, accelerates the combustion rate of CH4 (see Fig. 11), it is helpful to complete combustion of the fuel, improves combustion
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EGR (%)
Temperature
OH
0
10
20
30
40
Fig. 7. In-cylinder temperature and OH radical distribution at different EGR rates.
engine at different IPs are presented in Fig. 13. The diagram shows that NOX emissions increase as the IP increases. This is because as the IP rises, the atomization effect of diesel is better and the stratification is not obvious, but more combustible mixtures are formed, complete fuel combustion is promoted, and the ICT rises (see Fig. 5); consequently, NOX emissions increase. As the EGR rate increases, the amount of exhaust gas that enters the cylinder increases; this causes the oxygen concentration and in-cylinder combustion temperature to decrease; thus, NOX emission is reduced. Fig. 13 shows that when the EGR rate is small, the IP has a considerable impact on NOX emission. This is because at a small EGR rate, the increase in IP aids in the mixing of diesel and NG; this increases the combustion rate, CA50 advances (see Fig. 10(b)), cylinder temperature rises (see Fig. 5), and NOX emissions increase. When the EGR rate increases, the ignition delay period is prolonged, the fuel and air is fully mixed, and the effect of increasing IP on the mixing process is weakened. Fig. 13 further shows that the effect of IP on NOX emission is less than that of the EGR rate. Fig. 14 shows the main path of CO generation and consumption in diesel/NG dual-fuel mode at 1200 K and different EGR ratios. The path analysis is calculated by a transient closed homogeneous batch reactor of the SENKIN program [31] under homogeneous, adiabatic, and constant volume conditions. During combustion, CO is the main intermediate product of hydrocarbon fuel that is mainly formed by reactions
HCO + O2 = HO2 + CO and HCCO + O2 = OH + 2CO. If the oxygen concentration and temperature of the mixture are sufficiently high and reaction time is adequate, then CO will be oxidized to CO2, and the oxidation reactions of CO are mainly CO + HO2 = CO2 + OH and CO + OH = H + CO2. Fig. 15 presents the experimental data of CO emissions from diesel/ NG DF engines at different IPs and EGR rates. The increase in IP promotes the mixing of diesel and intake charge, and the local ER is increased, which is conducive to the formation of a homogeneous mixture. This causes the advance in CA50 (see Fig. 10(b)) and the increase in in-cylinder combustion temperature (see Fig. 5) to promote CO oxidation. Therefore, CO emissions decrease with the IP increase. When the EGR rate increases, CO emissions exhibit an upward trend. This is because with the increase in the EGR rate, the OH radical concentration decreases (see Fig. 7). This promotes the reaction, HCCO + O2 = OH + 2CO, in the positive direction and suppresses the reaction, CO + OH = H + CO2, in the same reaction direction; accordingly, CO emissions increase. On the other hand, polyatomic molecules, such as CO2 and H2O, have high heat capacities. After the addition of waste gas, the total heat capacity in the cylinder increases, and the combustion rate decreases; this results in a decrease in the ICT. Therefore, the increase in the EGR rate can inhibit the oxidation of CO and lead to the increase in CO emissions. Fig. 15 shows that CO emissions are insensitive to the IP when the
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Fig. 8. Variations in temperature, HRR, and key components in cylinder at different CAs.
emissions decrease with the increase in the IP; this occurs because of the reasons previously outlined. It is noteworthy that the reaction chain, C7H16 → C7H15-2 → PC4H9 → C2H4 → CH4, is the main pathway for methane formation from diesel pyrolysis. As the EGR rate increases, the amount of CH4 that is produced by diesel through the branch chain reaction, PC4H9 → C2H4 → CH4, increases (see Fig. 14) and results in increased methane emissions. Fig. 16 further shows that the injection pressure has a slight effect on methane emissions at low EGR rates. When the EGR rate is low, the oxygen content is relatively sufficient, which aids in the complete combustion of CH4; this decreases CH4 emissions. As the EGR rate increases, the oxygen content in the cylinder decreases, and the injection pressure rises; this aids in the oxidation reaction between the fuel and oxygen molecules; this improves the oxidation rate of methane. Therefore, the IP has a more significant influence on CH4 emissions when the EGR rate is large. Fig. 17 presents the calculation results of methane distribution in the cylinder at different IPs; the corresponding slices of turbulent kinetic energy at different IPs are shown in Fig. 18. The figure shows that when the IP increases, the turbulent kinetic energy in the cylinder and the flame propagation speed of methane increases (see Fig. 11); this contributes to the combustion of methane in the cylinder. However, the mixture near the cylinder wall and the crevice region does not ignite because of its low temperature. Although this part of the gas re-enters the cylinder during the expansion stroke, the in-cylinder temperature is already low at this instance, and the combustible mixture cannot be ignited. Consequently, unburned methane is discharged out of the cylinder along with the exhaust gas during the exhaust stroke. Thus, it is evident that the increase in injection pressure promotes
Fig. 9. Experimental ID at different IPs and EGR rates.
EGR rate is low. This is because at such rate, the oxygen content is high, the CA50 advances (see Fig. 10(b)), and the in-cylinder combustion temperature increases (see Fig. 7), which promotes the oxidation of CO; consequently, CO emission is low. When the EGR rate is higher, there are more waste gases in the cylinder, and the specific heat capacity becomes larger. The increase in the IP promotes the mixing of fuel and intake charge, reduces the formation of a rich zone, and contributes to the oxidation of CO; thus, CO emissions are reduced. Fig. 16 shows the experimental data of CH4 emissions from diesel/ NG DF engines at different EGR rates and IPs. The figure shows that CH4 8
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(a) ITE
(b) CA50
Fig. 10. Experimental ITE and CA50 at different IPs and EGR rates.
Fig. 11. Combustion rate of CH4.
Fig. 13. Experimental NOX emissions at different IPs and EGR rates.
DF engine was carried out. As the injection pressure increased, CO and CH4 emissions decreased, whereas the MPRR, ITE, and NOX emissions increased. When the EGR rate increased, CH4 and CO emissions increased, NOX emissions decreased, the ITE first increased and thereafter decreased, whereas the MPRR first decreased, thereafter increased, and finally decreased. In addition, the increase in injection pressure increased the in-cylinder turbulent kinetic energy and the flame propagation speed of methane; this aided in the re-ignition of unburned methane in the cylinder and consumed most of the unburned methane. At low EGR rates, the injection pressure had a lower impact on CO and CH4 emissions. The CO and CH4 emissions significantly decreased as the IP increased under high EGR rate conditions. It was found by the path analysis that, as the EGR rate increased, the amount of CH4 produced by diesel through the branch chain reaction, PC4H9 → C2H4 → CH4, increased; this became one of the sources of CH4 emissions. The natural gas in the cylinder wall and crevice region was not oxidized because of the low temperature in the cylinder wall. Thus, the NG in the crevice region and cylinder wall became the main source of unburned CH4 emissions. It is worth pointing out, the diesel/NG DF had a distinct three-stage heat-release characteristic in the HTC process, and the diesel/NG dual-fuel engine can obtain higher thermal efficiencies and lower emissions at a 160-MPa IP and 20% EGR ratio. Furthermore, under the condition of introducing large proportion of EGR, the combustion of DF engine is very unstable. Therefore, in order to further understand the effect of fuel injection pressure coupling EGR rate on diesel/NG DF engine, the cycle-to-cycle variation of diesel/NG DF engine should be explored in the future.
Fig. 12. Experimental MPRR at different IPs and EGR rates.
natural gas combustion in the center of the cylinder, but the remaining methane in the crevice region and cylinder wall becomes the main source of unburned methane emissions. 5. Conclusion In this study, the effects of different IPs (80, 120, and 160 MPa) and EGR ratios (0%–40%) on the emissions and combustion of a diesel/NG
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Fig. 14. Main path of CO generation and consumption in diesel/NG dual-fuel mode at different EGR rates.
Fig. 16. Experimental CH4 emissions at different injection pressures and EGR rates.
Fig. 15. Experimental CO emissions at different IPs and EGR rates.
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Crank angle
80 MPa
(°CA ATDC)
120 MPa
160 MPa
20
60
80
Fig. 17. Distribution of methane in cylinder at different IPs.
80 MPa
120 MPa
160 MPa
Turbulent Kinetic Energy
Fig. 18. Slices of equivalent ratio and turbulent kinetic energy at different IPs at TDC.
Acknowledgements
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