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May 21, 2015 - In this work, an improved KIVA-3V code was coupled with the CHEMKIN solver to model combustion and emission of a DMDF engine. First ...
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Numerical Study of Combustion and Emission Characteristics of a Diesel/Methanol Dual Fuel (DMDF) Engine Ruzhen Zang and Chunde Yao* State Key Laboratory of Engines, Tianjin University, Tianjin 300072, People’s Republic of China ABSTRACT: In comparison to the extensive experimental studies, numerical studies of a diesel/methanol dual fuel (DMDF) engine are very limited. In this work, an improved KIVA-3V code was coupled with the CHEMKIN solver to model combustion and emission of a DMDF engine. First, a chemical reaction mechanism including 65 reactions and 43 species was developed and validated with premixed flame species profiles. Second, new engine experiments at varied loads in both diesel and DMDF modes were performed to further validate the present model. Results show that the model predicts the engine combustion and emission in two modes well. At last, parameter studies on diesel injection timings and exhaust gas recirculation (EGR) rates were performed. Results show that the ignition delay time in DMDF mode is longer than that in diesel mode. Methanol supplied through the intake port significantly reduces soot emissions and slightly increases NOx emissions at different injection timings. In two modes, soot emissions are both formed mainly in the bowl bottom, bowl lip near the combustion chamber center, and squish area. With the increasing EGR rate, NOx emissions are reduced significantly and soot emissions increase significantly in two modes. Combustion characteristics in DMDF mode are more sensitive to EGR than in diesel mode.

1. INTRODUCTION Because of the petroleum energy crisis and environmental concerns, the solutions using alternative fuels for internal combustion engines have been suggested by many investigators.1−3 Methanol is one of these alternative fuels. Now natural gas is one of the most widely used alternative fuels in the world; however, methanol has greater potential than natural gas.4 This is because methanol has the lower cost and easier storage. In addition, methanol also has abundant sources, such as natural gas, coal, and biomass.5 Thus, methanol applied in internal combustion engines is attracting more and more attention from researchers. Now, there are mainly three applications of methanol in diesel engines, such as diesel/methanol blends,6,7 direct injection of methanol in the chamber,8 and port injection.9−12 However, because of the non-miscibility and difficulty to apply in the engine, direct mixture and direct injection have small application potential. The so-called port injection is that methanol is injected into the intake port to produce the homogeneous mixture, and then diesel is injected into the cylinder near the top dead center and initiates combustion. In comparison to the other two methods, the port injection method can use the larger proportion of methanol. The diesel engine applying port injection of methanol is called the diesel/ methanol dual fuel (DMDF) engine. However, by reference to extensive experimental studies on DMDF engines, there are very limited publications on threedimensional computational fluid dynamics (CFD) simulation. Ni et al.13 investigated the effects of substitution ratios of methanol on soot and NOx (x = 1 and 2) emissions. Their model is based on the global mechanism and uses diesel fuel as the soot precursor. Their results show that, when the methanol ratio increases, soot emissions decrease significantly, while NOx emissions increase. Using a modified multi-dimensional CFD model, Li et al.14 optimized the combustion of a DMDF engine and performed parametric studies on the engine combustion © 2015 American Chemical Society

and emissions. However, their model is only validated with limited experimental data, such as a pure fuel engine. In this work, first, we descripted in detail the numerical model used and performed new engine experiments; second, we validated the present model with premixed flame species profiles and engine data in both diesel and DMDF modes at varied loads in a medium-duty diesel engine; and last, we performed parameter studies on diesel injection timings and exhaust gas recirculation (EGR) rates in a DMDF engine.

2. NUMERICAL MODEL In KIVA-3V,15 the renormalization group (RNG) k−ε model16 was used to model turbulence. The Kelvin−Helmholtz (KH)− Rayleigh−Taylor (RT) model17 was selected to model the droplet breakup process. The spray collision was modeled according to Nordin.18 The wall heat transfer was modeled according to Han and Reitz. 19 The chemistry solver CHEMKIN-II20 was coupled with KIVA-3V code for chemistry calculations by calling DVODE.21 The turbulence combustion model was a characteristic time combustion model and improved by Xu et al.22 considering equilibrium entropy instead of the equilibrium concentration of species. In addition, n-tetradecane (n-C14H30) was selected to represent the physical properties of diesel, and n-heptane was used to represent the chemical properties of diesel, because the ignition behavior of n-heptane is similar to diesel and the molecular structure of n-heptane is relatively simple. A reduced n-heptane/methanol mechanism23 was used to model the combustion chemistry of DMDF. Moreover, to model engine emissions, a NOx mechanism was added from GRI3.024 and an acetylene (C2H2) mechanism was added from the detailed chemical Received: March 27, 2015 Revised: May 4, 2015 Published: May 21, 2015 3963

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The calculation of the rate of the oxidation reaction Ṙ total is described by Patterson et al.28

reaction mechanism of n-heptane/toluene/methanol.25 The reason for considering acetylene oxidation pathways was that acetylene was selected as the soot precursor in the soot model. The final mechanism included 43 species and 65 reactions. In this model, we improved the soot formation model developed by Hiroyasu and Nishida26 and calculated the change rates of soot net mass and soot formation by eqs 1 and 2, respectively dMsoot dM form dMoxid = − dt dt dt

(1)

dM form = A f mC2H2P 0.5 exp( −Ef /RT ) dt

(2)

3. EXPERIMENTAL SECTION New engine experiments were performed to validate the present model. A four-cylinder, turbocharged, intercooling medium-duty diesel engine was used in this study. Specifications of the engine are shown in Table 1. The schematic diagram of the experimental setup is shown in Figure 1. The methanol quantity injected in the intake port was controlled by an electronic control unit. The engine torque and speed were controlled by a hydraulic dynamometer. In this test, two fuels used in experiments were commercial diesel fuel (sulfur content is less than 350 ppm) and methanol (purity is 99.9%). The cylinder pressure was traced using a pressure sensor (Kistler 6125C). The recorded cylinder pressure data were the average value of 100 continuous cycles. After that, heat release rate (HRR) was calculated according to the collected cylinder pressure. In addition, NOx emissions were measured with a chemiluminescence detector (CLD) analyzer. Soot emissions were measured with a filter smoke meter (AVL 415S). The computational mesh is shown in Figure 2. Because the diesel injector had seven nozzle holes, a 360° mesh was used in this study.

where Af = 4000, mC2H2 is the mass of acetylene, P is the incylinder pressure in bar, and Ef = 12 500 cal/mol. The soot oxidation model was from the Nagle and StricklandConstable27 model. The soot oxidation rate was calculated by eq 3 dMoxid 6MWc = MsootṘ total dt ρsoot Dsoot

(3)

where MWc and ρsoot are the carbon molecular weight and soot density (2 g/cm3) respectively, and Dsoot and Msoot are the soot diameter (2.5 × 10−6 cm) and soot mass, respectively.

Figure 2. Computational mesh of YC4D140 engine simulations.

Table 1. Engine Specifications parameter

specification

The calculations were from intake valve closure (IVC) to exhaust valve open (EVO).

number of cylinders engine aspiration bore × stroke (mm) displacement (cm3) compression ratio IVC (deg CA ATDC) EVO (deg CA ATDC) nozzle (number × bore diameter) (mm) injection pressure (MPa) maximum power (kW at rpm)

4 stroke, 4 inline turbocharged 108 × 115 4214 17:1 −130.3 112.2 7 × 0.16 28 103 at 1600

4. RESULTS AND DISCUSSION 4.1. Model Validations. The above developed model needed to be validated extensively. In this section, first, the model was validated with the premixed flame species profiles; and second, the model was validated with in-cylinder pressures, HRR, and emissions of an actual DMDF engine at varied loads. 4.1.1. Premixed Flame Validation. The laminar premixed flame species profiles29 were used to validate the present reaction mechanism. CHEMKIN-PRO30 package was used to

Figure 1. Schematic of the YC4D140 engine setup. 3964

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the predictive accuracy of the skeletal mechanism for acetylene is very important for soot modeling in the engine because acetylene is a key species in the soot formation. Figure 4 shows the comparisons of the concentration of acetylene between calculations and measurements. It is seen that the concentration of acetylene is also well-predicted. 4.1.2. Direct-Injection Diesel Validations for Pressure, HRR, and Emissions. The species profile validations in the premixed

conduct the simulation. The mole fraction of species was calculated by a premixed laminar burner-stabilized flame model. The temperature profiles in calculations were from the experiments.29 Two groups of validations including two different fuels, i.e., pure n-heptane (flame A), and methanol/ n-heptane dual fuel (flame B), were calculated. The experimental conditions are shown in Table 2 (X is the mole Table 2. Operating Conditions for Validations29 flame

Xn‑heptane

Xmethanol

X O2

A B

0.0481 0.0407

0.0000 0.0407

0.5288 0.5103

X

Ar

0.4231 0.4083

C/O

mass flux

0.318 0.307

0.002115 0.001960

fraction of species). Figure 3 shows the comparisons of mole fractions of species between calculations and measurements.

Figure 4. Comparisons of the concentration of acetylene between experiments29 and simulations in two different flames.

Table 3. Engine Operating Conditions for Validations for Cylinder Pressure and HRR

Figure 3. Comparison of the concentration profiles of major reaction products between experiment29 and simulation in two different premixed flames. The solid lines are simulations, and the symbols are experiments.

It is seen that the experimental results are well-predicted. The inconsistency of the species mole fraction exists in the zone that is the nearest to the burner. This is caused by both the experiment uncertainty (the preheat zone temperature is overestimated because of the heating effect of the nozzle, which results in the faster oxidation rate at this zone) and the modeling uncertainty (the temperature is estimated at this zone because the thermocouple cannot be placed in it). In addition,

simulation case

case 1

case 2

case 3

case 4

case 5

combustion mode engine speed (revolutions/min) load (%) SR (%) SOI (deg CA ATDC)i diesel mass (mg/cycle) initial pressure (bar) initial temperature (K)

diesel 1660

diesel 1660

diesel 1660

DMDF 1660

DMDF 1660

25 0 −11.4 23.2 1.57 372

50 0 −11.4 38.4 1.76 378

75 0 −11.4 55.8 2.0 385

75 30 −11.4 39.1 1.96 379

75 40 −11.4 35.8 1.94 378

i

CA represents crank angle, and ATDC represents after top dead center. 3965

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Figure 5. Comparisons of in-cylinder pressure and HRR.

flame only showed the good predictive ability of the skeletal mechanism. However, when coupled with the improved KIVA code, these validations were not sufficient for engine simulation. Thus, new engine experiments were performed to further confirm the predictive ability of the present model. Table 3 shows the engine operating conditions for validations for cylinder pressure and HRR, which include three different loads (i.e., 25, 50, and 75%) and substitution

ratios of methanol (SR) (i.e., 0, 30, and 40%). SR is the percentage reduction of diesel mass consumption in DMDF mode compared to diesel mode at the same engine speed and torque. SR is calculated using the following formula: SR = 3966

qd,d − qm,d qd,d

× 100% (4) DOI: 10.1021/acs.energyfuels.5b00644 Energy Fuels 2015, 29, 3963−3971

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Energy & Fuels where qd,d and qm,d are mass consumption rates of diesel in diesel and DMDF modes, respectively.

Figure 5 shows the comparisons between calculations and measurements in two modes. The pressure maximum deviation between experiments and simulations is 0.5 MPa. The relative error of simulation is 4%, which is acceptable for threedimensional (3D) CFD engine simulation. Panels a−c of Figure 5 show that the in-cylinder pressures and HRRs are wellpredicted at varied loads in diesel mode. Panels d and e of Figure 5 show that there are obvious differences in HRR between simulation and experiment in DMDF mode. It is seen that the peak HRR of premixed combustion is overestimated and the peak HRR of diffusion combustion is underestimated in simulation results. Li et al.31 attributed the discrepancy between the experiment and simulation (HRR calculation) in DMDF to the different calculated methods of HRR. The calculated HRR is the difference between reaction released heat and the wall transfer heat. The HRR in experiments is calculated from the experimental in-cylinder pressure data. However, as shown in panels a−c of Figure 5, the accurate predictions of HRR of diesel mode prove that their explanation is not reasonable. In addition, with the increasing of SR, a higher premixed HRR peak is expected because of the longer ignition delay time. Simulation results predict this expectation well. Thus, we conclude that the discrepancy between the experiment and simulation (HRR calculation) in the case of DMDF mode is due to the inappropriate definition of the heat capacity ratio in the derivation of HRR in a combustion analysis system (AVL IndiCom) in experiments. The experimental HRR is calculated on the basis of the cylinder pressure by eq 5

Figure 6. Comparisons of NOx and soot emissions.

dQ g dφ

=

dQ w γ dV 1 dP P + V + γ − 1 dφ γ − 1 dφ dφ

(5)

where dQg/dφ is HRR, γ is the heat capacity ratio, V is the instantaneous volume, P is in-cylinder pressure, and dQw/dφ is the heat loss rate. As we know, γ of mixture gas is determined by the temperature and species in cylinder. The empirical value of γ is suitable for experimental HRR calculation in diesel mode. However, when in DMDF mode, the value needs to be adjusted considering different species in cylinder. Thus, the inappropriate value of γ in DMDF mode results in the inconsistency in HRR between calculations and measurements.

Figure 7. Comparisons of in-cylinder pressure histories between two modes at different injection timings. The solid lines are diesel mode, and the dashed lines are DMDF mode.

Figure 8. Reaction path analysis of n-heptane/methanol dual fuel (T0 = 750 K, and P0 = 40 atm).23 The dotted line is the consumption path of radicals; the solid line is the conversion paths of other species; and the dashed line is the production path of radicals. The percentages are the distribution ratio of OH and HO2 radicals. 3967

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Energy & Fuels Further work needs to be performed to find the suitable value of γ to correct the error of experimental HRR calculation in DMDF mode. In addition, the simulated pressure curve is a little lower than the experimental results. The reason is that, in experiments, methanol is injected in the intake port but not in the intake manifold, which results in the uneven methanol/air distribution. Figure 6 shows the comparisons of NOx and soot emissions between calculations and measurements at five different SRs of methanol (0, 15, 20, 30, and 40%). The changing trend in soot emissions is predicted well. Simulation results show that soot emissions are formed more in diesel mode. With the increasing SR of methanol, soot emissions decrease rapidly. In addition, at SR 0.4, the simulation results show that the NOx and soot emissions for the simulation go down simultaneously. There are several reasons about why the NOx and soot emissions for the simulation go down at SR 0.4. For the reduction of NOx emissions, NOx emissions are determined by the oxygen concentration, the combustion temperature, and the combustion duration. First, the cylinder gas temperature is reduced at compression progress because of high latent heat of vaporization of methanol. Second, because of the premixed

methanol, the combustion duration is shortened and, hence, the duration of high temperature is also shortened. For the reduction of soot emissions, first, methanol injected from the intake port in DMDF mode reduces the diesel concentration in the cylinder; second, the longer ignition delay of methanol premixed mode leads to more premixed diesel fuel vaporization, which reduces the local diesel-rich combustion zones; and last, there is no carbon−carbon bond in methanol, which makes methanol oxidation difficult to produce the soot inception, such as acetylene or aromatic hydrocarbon. Above all, the current model can predict the species profiles in premixed flames, in-cylinder pressure, HRR, and emissions in the DMDF engine well. Thus, this model can be used to conduct the parameter study, as shown in the next section. 4.2. Parameter Studies. In this section, the effects of fuel injection timings and EGR rates on the combustion and emission characteristics of the DMDF engine were investigated. 4.2.1. Effect of Injection Timings. The comparisons of the cylinder pressure between diesel and DMDF modes at different injection timings are shown in Figure 7. The middle SR (20%) was selected to investigate the effects of injection timings because SR values range from 0 to 40% in DMDF engine experiments. It is seen that the compression curve of DMDF is lower than that in diesel mode, which is due to the higher specific heat capacity of premixed mixture gas in DMDF mode. The simulation results also confirm that the ignition delay time of DMDF is longer than that in diesel mode in all cases, which has been found by many experimental studies. The reasons are likely the sum of the lower oxygen concentration, the higher autoignition temperature of methanol, the higher specific heat capacity in the intake charge, and the chemical interaction between diesel and methanol. Xu et al.23 analyzed in detail the reaction path of fuel blends, as shown in Figure 8. They point out that the added methanol converts the reactive OH radicals generated from n-heptane low-temperature oxidation into H2O2, which is stable below 1000 K. The longer ignition delay time that resulted from methanol addition is helpful to generate more premixed diesel vapor, which could generate higher pressure in the cylinder, as shown in Figure 7. It shows that the peak pressures of DMDF are higher than those in diesel mode in all cases. When the injection timing of diesel is

Figure 9. Comparisons of NOx and soot emissions between two modes at different injection timings.

Figure 10. Temperature distributions of different injection timings at 0°, 6°, and 20° CA ATDC in DMDF mode (SOI means start of injection). 3968

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Figure 11. Temperature distributions of two modes at 6° and 20° CA ATDC (injection timing is −14° CA ATDC).

Figure 12. Equivalence ratio distributions of two modes at CA5 (injection timing is −14° CA ATDC, and CA5 is the time of 5% burn point).

−5.5° CA ATDC, the difference of the maximum pressure between DMDF and diesel modes is not obvious. This is because the main heat release appears in a bigger volume. When the injection timing is advanced, the peak pressures of two modes both increase quickly, which is due to more premixed diesel fuel vapor in the cylinder. In addition, it should be noted that there are no significant differences between the sensibilities of combustion characteristics of two modes to injection timing. Figure 9 shows the relationship between diesel injection timings and emissions of an engine. It is seen that, when diesel is injected in advance, NOx emission increases and soot emission decreases in both modes. The reason is shown in Figure 10. It is seen that there is a larger high-temperature area when the injection timing is advanced, which accelerates NOx formation and soot oxidation. In addition, when diesel is injected in advance, the more premixed diesel also reduces the local diesel-rich combustion zones, which inhibits the soot formation. In comparison to diesel mode, NOx emission in DMDF increases slightly and soot emission in DMDF decreases significantly at different injection timings, as shown in Figure 9. The reason for the increase of NOx emission is that there is a larger high-temperature area in DMDF mode, as shown in Figure 11. The longer ignition delay time in DMDF produces more premixed fuel, which results in the larger hightemperature area. The reasons for the reduction of soot emissions are mainly that methanol added in DMDF mode reduces the diesel amount and the longer ignition delay reduces the local diesel-rich combustion zones, as shown in Figure 12. In addition, the higher combustion temperature is helpful to accelerate soot oxidation.

Figure 13. Soot density distributions of two modes at 12°, 20°, and 40° CA ATDC (injection timing is −14° CA ATDC).

Figure 13 shows the position of soot formation in the cylinder in DMDF and diesel modes. It is seen that soot emissions are formed mainly at the bowl bottom, bowl lip near the combustion chamber center, and squish area, where there exist better cooling. The difference between two modes is only in the amount of soot formation. 4.2.2. Effect of EGR Rates. In the traditional diesel engine, EGR is helpful to reduce NOx emissions but results in higher particulate matter (PM), total hydrocarbon (THC), and CO emissions. To overcome the trade-off relationship between PM and NOx emissions, low-temperature combustion (LTC) has been proposed. Dual fuel with EGR dilution is one strategy for achieving LTC. However, dual fuel combustion is difficult to expand the operation range because of its premixed nature, and hence, EGR is required to lower combustion reactivity and 3969

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NOx emissions. However, there has been not experimental studies in a DMDF engine with EGR dilution. Therefore, a numerical study about effects of EGR on the combustion and emission characteristics of a DMDF engine was performed at 1660 revolutions/min engine speed. When the EGR rate changes, the mass of fuel is kept constant. Figure 14 shows that EGR retards the ignition of diesel in all cases. The reason is obvious; EGR increases the specific heat capacity of premixed mixture gas. The change of specific heat capacity can be seen from Figure14, in which the compression curve with EGR dilution is lower compared to the compression curve without EGR dilution. In addition, combustion characteristics (i.e., ignition time and peak pressure) are more sensitive to EGR in DMDF mode than in diesel mode. The reason is shown below. As we know, the ignition delay time of DMDF could be expressed simply by eq 6 as an Arrhenius-type expression τ = x1P x2φdiesel x3(1 + φmethanol x4)(1 + fEGR )e x5Ea / RT

Figure 14. Comparisons of cylinder pressure histories between two modes at different EGR rates. The solid lines are diesel mode, and the dashed lines are DMDF mode.

(6)

where φdiesel is the equivalence ratio of diesel, φmethanol is the equivalence ratio of methanol, P is the in-cylinder pressure, and x1, x2, x3, x4, and x5 are the adjustment parameters (x1 > 0; x2 < 0; x3 < 0; and x5 > 0). We assumed that the chemical effect of methanol on the ignition of diesel is reflected in the parameter x4 (x4 > 0 in DMDF mode, and x4 = 0 in diesel mode); φdiesel(DMDF) < φdiesel(diesel) . In addition, because of the higher specific heat capacity of mixture gas in DMDF mode, P(DMDF) < P(diesel) and T(DMDF) < T(diesel). On the basis of the above reasons, we could conclude that ⎛ dτ ⎞ ⎛ dτ ⎞ ⎜⎜ ⎟⎟ ⎟⎟ < ⎜⎜ ⎝ dfEGR ⎠diesel ⎝ dfEGR ⎠DMDF

(7)

Thus, combustion characteristics are more sensitive to EGR in DMDF mode. Figure 15 shows the comparison of emission characteristics between two modes at different EGR rates. It is seen that, when the EGR rate increases, NOx emissions decrease significantly and soot emissions increase in two modes. The reasons are shown in Figure 16. It shows that, when the EGR rate increases, the more exhaust gas in the

Figure 15. Comparisons of NOx and soot emissions between two modes at different EGR rates.

Figure 16. Temperature distributions of different EGR rates at 0°, 6°, and 20° CA ATDC. 3970

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(4) Olah, G. A.; Goeppert, A.; Prakash, G. K. S. Beyond Oil and Gas: The Methanol Economy; John Wiley & Sons: Weinheim, Germany, 2011. (5) Chmielniak, T.; Sciazko, M. Appl. Energy 2003, 74 (3−4), 393− 403. (6) Huang, Z. H.; Lu, H. B.; Jiang, D. M.; Zeng, K.; Liu, B.; Zhang, J. Q. Proc. Inst. Mech. Eng., Part D 2004, 218 (D4), 435−447. (7) Sayin, C.; Ozsezen, A. N.; Canakci, M. Fuel 2010, 89 (7), 1407− 1414. (8) Fang, X. Z.; Liu, X. J.; Jin, W. H.; Yan, S. F. Trans. CSICE 2003, 21 (6), 411. (9) Liu, J.; Yao, A.; Yao, C. Fuel 2014, 134, 107−113. (10) Wei, L. J.; Yao, C. D.; Wang, Q. G.; Pan, W.; Han, G. P. Fuel 2015, 140, 156−163. (11) Udayakumar, R.; Sundaram, S.; Sivakumar, K. SAE Tech. Pap. Ser. 2004, DOI: 10.4271/2004-01-0096. (12) Yao, C.; Cheung, C.; Cheng, C.; Wang, Y. Energy Fuels 2007, 21, 686−691. (13) Ni, P.; Wang, X.; Wang, Z.; Mao, G.; Wei, S. Numerical modeling of soot and NOx emissions of a diesel/methanol dual fuel engine. Proceedings of the 2011 International Conference on Computer Distributed Control and Intelligent Environmental Monitoring (CDCIEM); Changsha, Hunan, China, Feb 19−20, 2011; pp 483−486. (14) Li, Y.; Jia, M.; Chang, Y.; Liu, Y.; Xie, M.; Wang, T.; Zhou, L. Energy 2014, 65, 319−332. (15) Amsden, A. A. KIVA-3V: A Block-Structured KIVA Program for Engines with Vertical or Canted Valves; Los Alamos National Laboratory: Los Alamos, NM, 1997; LA-13313-MS. (16) Kong, S.; Han, Z. Y.; Reitz, R. D. SAE Tech. Pap. Ser. 1995, DOI: 10.4271/950278. (17) Ricart, L. M.; Reltz, R. D.; Dec, J. E. J. Eng. Gas Turbines Power 2000, 122 (4), 588−595. (18) Nordin, N. Complex chemistry modeling of diesel spray combustion. Ph.D. Thesis, Chalmers University of Technology, Göteborg, Sweden, 2001. (19) Han, Z. Y.; Xu, Z.; Trigui, N. Int. J. Eng. Res. 2000, 1 (1), 127− 146. (20) Kee, R. J.; Rupley, F. M.; Miller, J. A. CHEMKIN-II: A FORTRAN Chemical Kinetics Package for the Analyses of Gas Phase Chemical Kinetics; Sandia National Laboratories: Livermore, CA, 1989; Sandia Report SAND 89-8009. (21) Brown, P. N.; Byrne, G. D.; Hindmarsh, A. C. SIAM J. Sci. Stat. Comput. 1989, 10 (5), 1038−1051. (22) Xu, G. L.; Yao, C. D.; Rutland, C. J. Int. J. Engine Res. 2014, DOI: 10.1177/1468087413516119. (23) Xu, H. J.; Yao, C. D.; Xu, G. L. Fuel 2012, 93, 625−631. (24) Smith, G. P.; Bowman, T.; Frenklach, M. GRI-Mech 3.0, 1999; http://www.me.Berkeley.edu/gri-mech. (25) Xu, H. J.; Yao, C. D.; Xu, G. L.; Wang, Z. D.; Jin, H. F. Combust. Flame 2013, 160 (8), 1333−1344. (26) Hiroyasu, H.; Nishida, K. SAE Tech. Pap. Ser. 1989, DOI: 10.4271/890269. (27) Nagle, J.; Strickland-Constable, R. F. Proceedings of the Fifth Carbon Conference; Pergamon Press: Oxford, U.K., 1962; Vol. 1, pp 154−164. (28) Patterson, M. A.; Kong, S. C.; Hampson, G. J.; Reitz, R. D. SAE Tech. Pap. Ser. 1994, DOI: 10.4271/940523. (29) Xu, H. J. Study on the chemical kinetics of diesel/methanol duel fuel combustion. Ph.D. Thesis, Tianjin University, Tianjin, China, 2012. (30) Chemkin-Pro, Release 15083, Reaction Design: San Diego, 2009. (31) Li, Y.; Jia, M.; Liu, Y.; Xie, M. Appl. Energy 2013, 106, 184−197.

cylinder makes the in-cylinder temperature drop significantly, which is helpful to reduce NOx formation. In addition, the reduction of the O2 concentration induced by EGR increases the equivalent ratio of fuel, which results in more soot emissions.

5. CONCLUSION A skeletal chemical reaction mechanism was developed to model the combustion and emission of the DMDF engine. Premixed flame species profiles were used to validate this mechanism. Results show that this mechanism agrees well with the major reaction product concentration in both pure n-heptane fuel and n-heptane/methanol dual fuel. Acetylene was particularly concerned, because it was used as the soot precursor in this work. In addition, new engine experiments at varied loads in both diesel and DMDF modes were conducted to further validate the predictive ability of the present model. Results indicate that KIVA-3V code coupled with this mechanism can predict well with the in-cylinder pressures, HRR, and emissions in two modes. The parameter study results confirm that the ignition delay time of DMDF is longer than that in diesel mode. In DMDF mode, methanol injected in the intake port markedly reduces the soot emission and slightly increases NOx emissions. Soot emission is formed mainly at the bowl bottom, bowl lip near the combustion chamber center, and squish area in both modes. The difference between two modes is only in the quantity of soot formation. Under DMDF mode, both supplying methanol and advancing injection timing could increase the peak cylinder pressure. There are no significant differences between the sensibilities of combustion characteristics of two modes to injection timing. EGR retards the ignition in both diesel and DMDF modes. With the increasing EGR rate, NOx emissions in two modes are reduced significantly and soot emissions increase significantly. In DMDF mode, combustion characteristics (i.e., ignition time and peak pressure) are more sensitive to EGR than that in diesel mode. In the next work, more parameter studies (e.g., multiple injections, combustion chamber shape, and compression ratio) and optimizations of a DMDF engine need to be performed to deeply understand the DMDF combustion.

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AUTHOR INFORMATION

Corresponding Author

*Telephone: +86-22-2740-6649. Fax: +86-22-2738-3362. E-mail: [email protected]. Notes

The authors declare no competing financial interest.

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ACKNOWLEDGMENTS The authors thank the National Natural Science Foundation of China (51336005) and the National High-Technology Research and Development Program (863 Program) (2012AA111719) of the Ministry of Science and Technology of China.

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REFERENCES

(1) Papagiannakis, R. G.; Rakopoulos, C. D.; Hountalas, D. T.; Rakopoulos, D. C. Fuel 2010, 89 (7), 1397−1406. (2) Qi, D. H.; Chen, H.; Geng, L. M.; Bian, Y. ZH.; Ren, X. CH. Appl. Energy 2010, 87 (5), 1679−1686. (3) Balki, M. K.; Sayin, C.; Canakci, M. Fuel 2014, 115, 901−906. 3971

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