Dimethyl Ether Dual

Energy & Fuels 2009, 23, 2719–2730

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Experimental and Numerical Study of Methanol/Dimethyl Ether Dual-Fuel Compound Combustion Zheng Chen, Mingfa Yao,* Zunqing Zheng, and Quanchang Zhang State Key Laboratory of Engines, Tianjin UniVersity, Tianjin 300072, People’s Republic of China ReceiVed December 3, 2008. ReVised Manuscript ReceiVed March 1, 2009

This paper experimentally and numerically investigates the effects of injection timing on combustion of a methanol/dimethyl ether (DME) dual-fuel compound at various methanol concentrations. In this dual-fuel compound combustion approach, methanol is directly injected into the cylinder, and DME is injected at the intake port. The experimental results indicate that methanol injection timing has an obvious effect on the heat release process of the dual-fuel compound combustion, and the effect of methanol concentration (i.e., the engine load) on injection timing is also evident. Late injection should be adopted to achieve smooth combustion and low NOx emissions at high methanol concentrations. Moderate injection timing should be adopted to achieve higher indicated thermal efficiency (ITE) at low methanol concentrations. The simulation results, using computational fluid dynamics (CFD) combined with reduced chemical kinetics, show that methanol injected at -26 °CA after top dead center (ATDC) has a remarkable effect on the high-temperature reaction of DME. In the case of injection timing at -26 °CA ATDC, the high-temperature combustion region is concentrated within the combustion chamber, which results in a higher NO concentration. With injection timing at -6 °CA ATDC, by contrast, the high-temperature combustion region is dispersed in the compression clearance and near the chamber wall, which leads to a relatively low NO concentration. Increasing injection pressure is an effective way to shorten the duration of the methanol/DME dual-fuel compound combustion achieved by later injection.

1. Introduction Alternative fuels have been widely investigated all over the world in recent years. Methanol is a promising alternative fuel to diesel and gasoline. It can be produced efficiently and economically from a variety of feedstocks, including natural gas, coal, and biomass.1-3 Methanol, CH3OH (i.e., methyl alcohol), is the simplest aliphatic alcohol and the first member of the homologous series. It is a colorless liquid, completely miscible with water and organic solvents. As a fuel, methanol has many desirable combustion and emission characteristics.4-7 Methanol has a high octane number, indicating good antiknock performance; high volatility, allowing a denser fuel-air charge; and excellent lean burn properties. These properties make * To whom correspondence should be addressed: tel 86-22-27406842, ext. 8014; fax 86-22-27383362; e-mail [email protected]. (1) Joshi, S.; Lave, L.; Lester, M.; Lankey, R. A life cycle comparison of alternative transportation fuels. SAE Tech. Pap. Ser. 2000, 2000-011516. (2) Dhaliwal, B.; Yi, N.; Checkel, D. Emissions effects of alternative fuels in light-duty and heavy-duty vehicles. SAE Tech. Pap. Ser. 2000, 200001-0692. (3) Nabi, M.; Minami, M.; Ogawa, H.; Noboru, M. Ultra-low emission and high performance diesel combustion with highly oxygenated fuel. SAE Tech. Pap. Ser. 2000, 2000-01-0231. (4) Miyamoto, N.; Ogawa, H.; Arima, T.; Kenji, M. Improvement of diesel combustion and emissions with addition of various oxygenated agents to diesel fuels. SAE Tech. Pap. Ser. 1996, 962115. (5) Miyamoto, N.; Ogawa, H.; Nurun, N.; Kouichi, O.; Teruyoshi, A. Smokeless, low NOx, high thermal efficiency, and low noise diesel combustion with oxygenated agents as main fuel. SAE Tech. Pap. Ser. 1998, 980506. (6) Tsurutani, K.; Takei, Y.; Fujimoto, Y.; Junichi, M.; Mitsuhiro, K. The effects of fuel properties and oxygenates on diesel exhaust emissions. SAE Tech. Pap. Ser. 1995, 952349. (7) Richards, B. G. Methanol-fueled Caterpillar 3406 engine experience in on-highway trucks. SAE Tech. Pap. Ser. 1990, 902160.

methanol a good fuel for spark-ignition (SI) engines.8 Its high latent heat of vaporization is useful for decreasing combustion temperature and engine-out NOx. At the same time, methanol exhibits clean-burning behavior due to its high oxygenated content and lack of carbon-carbon bonds.9 DME, which is also a renewable alternative fuel, has a high cetane number (CN ) 55∼60), low autoignition temperature, and low boiling point. It can be used as an ignition improver in compression-ignition methanol engines. Green et al.10 showed in experiments that, compared to glow plug engines, combustion and emissions of compression-ignition methanol engines improved remarkably at partial loads with the introduction of DME. Kozole and James11 reported that cold startability with spark ignition was improved from 10 to -15 °C by introducing DME into the intake manifold. Galvin,12 Cipolat,13 and Karpuk et al.14 (8) Black, F. An overview of the technical implications of methanol and ethanol as highway motor vehicle fuels. SAE Tech. Pap. Ser. 1991, 912413. (9) McCallum, P. W.; Timbario, T. J.; Bechtold, R. L.; Eckland, E. E. Methanol/ethanol: alcohol fuels for highway Vehicles; CEP: Edinburgh, Scotland, 1982; pp 52-59. (10) Green, C. J.; Lionel, K.; Neal, A. C. Dimethyl ether as a methanol ignition improverssubstitution requirements and exhaust emissions impact. SAE Tech. Pap. Ser. 1990, 902155. (11) Kozole, K. H.; James, S. W. The use of dimethyl ether as a starting aid for methanol-fueled SI engines at low temperatures. SAE Tech. Pap. Ser. 1988, 881677. (12) Galvin, M. P. Aspirated ether ignition system for methanol fueled diesel engines. In Proceedings of the 8th International Symposium on Alcohol Fuels, Tokyo, 1988; pp 601-604. (13) Cipolat, D. Methanol/dimethyl ether fueling of a compression ignition engine. In Proceedings of the 9th International Symposium on Alcohol Fuels, Firenze, Italy, 1991; pp 411-415. (14) Karpuk, M. E.; John, D. W.; James, L. D.; Daniel, E. J. Dimethyl ether as an ignition enhancer for methanol-fueled diesel engines. SAE Tech. Pap. Ser. 1991, 912420.

10.1021/ef8010542 CCC: $40.75  2009 American Chemical Society Published on Web 04/01/2009

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investigated stable ignition and combustion of methanol engines by introducing DME into the intake manifold. Karpuk and Scott15 also investigated the potential of on-board DME generation to assist methanol engines in cold starting. Their results showed that methanol could exothermically dehydrate to form DME with a fluorinated γ-alumina as catalyst in the temperature range of 250-350 °C, according to the equation 2CH3OH f CH3OCH3 + H2O + heat. This also offers the possibility of achieving dual-fuel combustion with a single fuel. With increased exploitation of alternative fuels, new combustion modes and technologies (different from typical diesel and gasoline combustion) have been proposed and developed. These combustion modes include controlled autoignition (CAI),16 homogeneous charge compression ignition (HCCI),17,18 stratified charge compression ignition (SCCI),19 and low-temperature combustion (LTC).20 Although these new combustion modes can achieve high-efficiency and low-emission combustion only at partial speeds and loads, they show a clear direction for further decreasing the pollution and fuel consumption of the internal combustion engine. Engine combustion simulation has become a useful tool for studying in-cylinder physicochemical processes. To simulate chemically controlled combustion phenomena such as HCCI, chemical kinetics must be taken into account. Various types of models have been developed, including zero-dimensional models,21 multizone models,22 and multidimensional computational fluid dynamics (CFD) models coupled with detailed chemistry.23,24 It has been shown that the zero-dimensional model often predicts a very fast combustion rate once ignition occurs and is unable to capture the correct combustion phase.21 A zonal model can predict a reasonable overall combustion rate but requires empirical adjustments in formulating zones.22 The simultaneous calculation of CFD and detailed chemical kinetics can offer a precise simulation of the combustion process and a prediction of HC, CO, and NOx.25 Unfortunately, the coupling process incurs a high computational cost, rendering this method unattractive to researchers. As an alternative, many researchers have adopted reduced chemical kinetics coupled with CFD.26-29 Noel et al.26 used reduced chemical kinetics of n-heptane with (15) Karpuk, M. E.; Scott, W. C. On board dimethyl ether generation to assist methanol engine cold starting. SAE Tech. Pap. Ser. 1988, 881678. (16) Oakley, A.; Zhao, H.; Ma, T.; Ladommatos, N. Dilution effects on the controlled auto-ignition (CAI) combustion of hydrocarbon and alcohol fuels. SAE Tech. Pap. Ser. 2001, 2001-01-3606. (17) Gray, A. W.; Ryan, T. W. Homogeneous Charge Compression Ignition (HCCI) of Diesel Fuel. SAE Tech. Pap. Ser. 1997, 971676. (18) Hasegawa, R.; Yanagihara, H. HCCI combustion in DI diesel engine. SAE Tech. Pap. Ser. 2003, 2003-01-0745. (19) Aroonsrisopon, T.; Werner, P.; Sohm, V. M.; Foster, D. E.; Ibara, T.; Morikawa, T.; Iida, M.; Waldman, J. O. Expanding the HCCI operation with the charge stratification. SAE Tech. Pap. Ser. 2004, 2004-01-1756. (20) Akihama, K.; Takatori, Y.; Inagaki, K.; Sasaki, S.; Dean, A. M. Mechanism of the smokeless rich diesel combustion by reducing temperature. SAE Tech. Pap. Ser. 2001, 2001-01-0655. (21) Dec, J. A computational study of the effects of low fuel loading and EGR on heat release rates and combustion limits in HCCI engines. SAE Tech. Pap. Ser. 2002, 2002-01-1309. (22) Aceves, S. M.; Flowers, D. L.; Westbrook, C. K.; Smith, J. R.; Pitz, W.; Dibble, R.; Christensen, M.; Johansson, B. A Multi-Zone Model for Prediction of HCCI Combustion and Emissions. SAE Tech. Pap. Ser. 2000, 2000-01-0327. (23) Kong, S. C.; Reitz, R. D. Proc. Combust. Inst. 2002, 29, 663–669. (24) Kong, S. C.; Reitz, R. D. Combust. Theory Model 2003, 7, 417– 433. (25) Wang, Z.; Shuai, J. S.; Wang, J. X.; Tian, G. H.; An, X. L. Modeling of HCCI combustion: From 0D to 3D. SAE Tech. Pap. Ser. 2006, 200601-1364. (26) Noel, L.; Maroteaux, F.; Ahmed, A. Numerical study of HCCI combustion in diesel engines using reduced chemical kinetics of n-heptane with multidimensional CFD code. SAE Tech. Pap. Ser. 2004, 2004-011909.

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multidimensional CFD code to simulate HCCI combustion in diesel engines. Their study shows that the reduced mechanism behaves similarly to the detailed models for the description of the two-stage phenomenon and also for the temperature and pressure inside the cylinder. Kim and Lee27 used KIVA-3V/ reduced chemical kinetics to study the effects of injection strategy on the combustion process and the chemical reaction path in an HSDI diesel engine. Zheng and Yao28 adopted a fully coupled multidimensional CFD and reduced chemical kinetics model to investigate the effects of charge stratification on HCCI combustion and emissions. Mo et al.29 used a DME reduced chemical kinetics model coupled with CFD to predict combustion and emissions of DME HCCI. Their results indicate that the combination of reduced chemical kinetics and CFD can efficiently capture the combustion characteristics of HCCI and contribute to optimizing the HCCI combustion system. In our previous study,30,31 we proposed and investigated a new approach for burning methanol in engines, where the engine burned methanol and DME dual-fuel in HCCI mode. In this approach, both methanol and DME were injected into the intake manifold and mixed into the intake air, while DME may be converted from methanol with an alumina catalyst. A homogeneous mixture was formed during the intake and compression stroke. HCCI combustion can be obtained in dual-fuel mode, resulting in ultralow NOx emissions and high thermal efficiency. The major advantages of this combustion approach were that the ignition timing and combustion rate could be effectively controlled by adjusting the ratio of DME and methanol and the exhaust gas recirculation (EGR) rate, and that both the low and high load limits of HCCI could be extended. However, the engine could not run in full load mode. Furthermore, based on the development of fuel injection technology, a dual-fuel compound combustion regime is proposed. In this regime, DME is injected at the intake port and methanol is directly injected into the cylinder by a common rail system. By adjusting methanol injection timing, various combustion characteristics can be achieved at different operating conditions. In this paper, some preliminary results will be presented. In addition, a simulation of CFD coupled with reduced chemical kinetics will be introduced to capture the behavior of the dual-fuel compound combustion. 2. Experimental Apparatus and Methods Experiments were conducted on a modified four-cylinder, watercooled, direct injection diesel engine. Specifications of the testing engine are given in Table 1, and a schematic of the experimental setup is shown in Figure 1. The fuels were supplied by an electronically controlled fuel injection system. The DME injector was mounted close to the intake port. DME was injected near the intake TDC so that a homogeneous mixture of DME and air could be formed during the intake and compression stroke. The methanol injector was mounted on the cylinder head of the engine. Methanol fuel was injected directly into the cylinder by a common rail fuel injection system. In this study, the common rail pressure was fixed at a constant value of (27) Kim, S.; Lee, J. Combustion process analysis in an HSDI diesel engine using a reduced chemical kinetics. SAE Tech. Pap. Ser. 2004, 200401-0108. (28) Zheng, Z. L.; Yao, M. F. Energy Fuels 2007, 21 (4), 2018–2026. (29) Mo, C. L.; Zhang, Y. S.; Shi, Y.; Han, J.; Sun, H. Y. Experimental and numerical study on emission in an HCCI engine operated with neat dimethyl ether. SAE Tech. Pap. Ser. 2007, 2007-01-1888. (30) Yao, M. F.; Zheng, Z. Q.; Chen, Z.; Zhang, B. Experimental study on HCCI combustion of dimethyl ether (DME)/methanol dual fuel. SAE Tech. Pap. Ser. 2004, 2004-01-2993. (31) Yao, M. F.; Chen, Z.; Zheng, Z. Q.; Zhang, B. Fuel 2006, 85, 2046– 2056.

Methanol/DME Dual-Fuel Compound Combustion

Energy & Fuels, Vol. 23, 2009 2721 commands from the electronic fuel injection system. Dynamic injection timing is adopted in this study.

Table 1. Engine Specifications bore × stroke compression ratio combustion chamber intake valve opening intake valve closure exhaust valve opening exhaust valve closure swirl ratio methanol injector type nozzle hole number spray cone angle common rail pressure

112 × 132 mm 17.5 bowl in piston 13.5 °CA BTDC 38.5 °CA ABDC 56.5 °CA BBDC 11.5 °CA ATDC 2 multihole 7 152 °C 20 MPa

3. Modeling Strategy

20 MPa due to the properties of methanol and the capabilities of the injection system. In order to reduce the effect of variation of intake air cycling on the measurement of intake air flow rate, a large tank was added as a pressure stabilizer in the intake system. The flow meter was mounted at the inlet of the surge tank. Cylinder pressure was recorded by a high-speed data acquisition system. The core of the system was a computer equipped with an acquisition board. A pressure transducer (Kistler 6125A) was mounted on the cylinder head. The pressure signal was transferred to the computer through a charge amplifier. The exhaust gas was measured with an exhaust analyzer (Horiba MEXA-7100DEGR) that measures HC with a hydrogen flame ionization detector (FID), CO and CO2 by the nondispersive infrared (NDIR) method, and NOx with a chemiluminescent NOx analyzer (CLA). In order to reduce the number of variables being investigated, the DME percentage is defined as the percentage of energy released by DME in total energy:

RDME )

GDMEHDME × 100 GMeOHHMeOH + GDMEHDME

(1)

where GDME and GMeOH are the mass flow rates of DME and methanol, respectively, and HDME and HMeOH are the low heat values of DME and methanol, respectively. Total fuel/air equivalence ratio φtotal, DME fuel/air equivalence ratio φDME, and methanol fuel/air equivalence ratio φMeOH are calculated with the following equations:

φtotal )

GDMEAFDME + GMeOHAFMeOH Gair

(2)

GDMEAFDME Gair

(3)

GMeOHAFMeOH Gair

(4)

φDME )

φMeOH )

where Gair is the total mass flow rate of air and AFDME and AFMeOH are the stoichiometric air-fuel ratios of DME and methanol, respectively. The engine load is presented by indicated mean effective pressure (IMEP). There are two definitions for IMEP: net IMEP refers to the full cycle, while gross IMEP refers to the compression and expansion strokes only. Net IMEP is used in this paper. Test conditions were set as follows: engine speed, 1400 r/min; coolant temperature, 85 °C; oil temperature, 90 °C; and inlet charge temperature, 18-23 °C. The primary objective of this paper is to investigate the effects of methanol injection timing on the dual-fuel compound combustion process. All tests were performed at a constant speed of 1400 rpm. Six injection timings were investigated, as shown in Table 2. Injection timing is expressed in two ways: dynamic and static timings. Dynamic injection timing is determined by the nozzle needle lift, while static injection timing is determined by the

3.1. DME/Methanol Reduced Chemical Kinetics Mechanism. A DME/methanol reduced chemical kinetics mechanism consisting of 27 species and 35 reactions was used to simulate fuel chemistry.32 In addition, a NOx submodel is added, including the two primary chemical routes: the Zeldovich mechanism and the N2O-intermediate mechanism.33 Ultimately, the new DME/methanol reduced mechanism includes 30 species and 44 reactions, as shown in the Appendix. 3.2. Coupling of CFD with Reduced Chemical Kinetics. The three-dimensional CFD code FIRE34 is adopted in this study to capture in-cylinder flow and the basic combustion characteristics of the dual-fuel compound combustion. Based on fundamental conservation principles, FIRE solves the average conservation equations of total mixture mass, momentum, enthalpy, and species. Turbulence effects are considered by adopting a compressible version of the standard two-equation κ - ε turbulence model. The differential transport/conservation equations, in curvilinear nonorthogonal form, are transferred into an Eulerian-Lagrangian coordinate system in order to enable their solution on a body-fitted computational mesh with moving boundaries. The finite volume method is used to discretize the partial differential equations governing the mean fluid motion. The Euler implicit scheme is used for temporal integration to ensure unconditional numerical stability, while a hybrid central/upwind differencing scheme is used for spatial derivatives. The SIMPLE algorithm is then used to solve the resulting algebraic equations. The most significant step in the simulation was to combine CHEMKIN, developed by the Sandia National Laboratory in the United States, with the CFD code. FIRE offers a general species transport model for implementing chemical kinetic models. The general species transport model provides the necessary transport equations for gas-phase chemical species in the computational domain. A formatted ASCII representation of the chemical reaction mechanism can be converted to the binary file required by CFD initialization. This file contains required data about the elements, species, and reactions in the user’s mechanism. The corresponding physical properties (e.g., density, specific heat, viscosity, thermal conductivity, and diffusion coefficient) of each species and of the gas mixture are calculated based on parameters extracted from the CHEMKIN database. The coupling strategy of CFD with CHEMKIN has been shown in the literature.25 The effect of chemistry is considered such that, at the beginning of each FIRE time step (∆t), a single zone 0D reactor model is called for each computational cell to calculate chemical heat release. On the basis of the results of this model, FIRE calculates the source terms for the species transport equations and the enthalpy equation to obtain the pressure, temperature, and concentration. In fact, each cell in the CFD model is a well-stirred reactor. The single-zone 0D model is called for each reactor alone so that pressure, temperature, and composition are homogeneous in each cell. Furthermore, the interaction between the cells is implemented by heat and mass transfer. The above processes are carried out (32) Yao, M. F.; Huang, C.; Zheng, Z. L. Energy Fuels 2007, 21 (2), 812–821. (33) Bowman, C. T. Control of combustion-generated nitrogen oxide emissions. Proceedings of the 24th Symposium (International) on Combustion. The Combustion Institute: Pittsburgh, PA1992; pp 859-878. (34) AVL FIRE UsersGuide, version 8.4, AVL LIST GmbH, 2005.

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Figure 1. Schematic of experimental setup: (1) engine; (2) rotary encoder; (3) cylinder pressure acquisition system; (4) exhaust gas analyzer; (5) fuel pump; (6) electromotor; (7) high-pressure fuel pump; (8) common rail pressure transducer; (9) pressure transducer; (10) methanol injector; (11) common rail; (12) DME surge tube; (13) DME control valve; (14) surge tank; (15) DME injector; (16) air flow meter; (17) dynamometer; (18) electric-controlled fuel injection system. Table 2. Specifications of Methanol Injection Timings case

methanol injection timing (°CA ATDC)

methanol fuel/air equivalence ratio

DME fuel/air equivalence ratio

total fuel/air equivalence ratio

1 2 3 4 5 6

-6 -16 -26 -36 -46 -66

0.119, 0.338, 0.5 0.119, 0.5 0.119, 0.338, 0.5 0.119, 0.5 0.119, 0.338, 0.5 0.119, 0.338, 0.5

0.105, 0.113, 0.105 0.105, 0.103 0.107, 0.113, 0.102 0.107, 0.102 0.107, 0.113, 0.102 0.107, 0.113, 0.102

0.224, 0.451, 0.605 0.224, 0.603 0.226, 0.451, 0.602 0.226, 0.602 0.226, 0.451, 0.602 0.226, 0.451, 0.602

in turn according to the time (or crank angle) sequence until the end of calculation. 3.3. Turbulence Interaction. The source terms for the species transport equations in FIRE are calculated as

and wCO2, wH2O, wCO, wH2, and wN2 are mass fractions of CO2, H2O, CO, H2, and N2, respectively, calculated by single-zone model in each cell. 4. Results and Discussion

Fn+1wkn+1 - Fnwkn Vcell Sk ) ∆t

(5)

where the superscript n indicates the starting values and the superscript n + 1 indicates the results of the single zone reactor model. wk is the mass fraction of the kth species, and Vcell is the volume of the computational cell. The effects of both chemical kinetics and mixing are accounted for by introducing a delay coefficient according to the approach of Kong et al.,35,36 with the assumption that the reaction rate is mainly determined by kinetic (equilibrium assumption under perfectly mixed conditions) and turbulent time scales (eddy break-up assumption): Sk )

τkin

τkin Fn+1wkn+1 - Fnwkn Vcell + fτturb ∆t

(6)

where τkin is kinetic time scale and τturb is turbulent time scale. The delay coefficient f)

1 - e-r 0.632

(7)

where r)

wCO2 - wH2O - wCO - wH2 1 - wN2

(8)

4.1. Experimental Results. Six injection timings are investigated in this section. Except for cases 2 and 4, each injection timing case includes three methanol concentrations. As seen in Figure 2, for each methanol concentration, the DME concentration is kept constant while injection timing is varied. In the figure, the definitions of knock and partial burning are consistent with the literature.31 Figure 2 indicates that as DME percentage is kept constant at the same methanol concentration, partial burning appears at low methanol concentrations and early injection timings due to lower gas temperature in the cylinder caused by the high latent heat of vaporization of methanol. On the other hand, knock appears at higher methanol concentrations and early injection timings. These phenomena indicate that, for the early injection case, the DME/methanol concentration controlled region is so narrow that it is difficult to control the dual-fuel compound combustion process. Figures 3 and 4 show the history of the in-cylinder pressure and heat release rate at various methanol injection timings and concentrations. Figure 3 explores the case of low methanol concentration. Figure 3a shows that, with advanced injection timing, the maximum peak pressure first rises and then falls. Figure 3b shows that the heat release rate curve presents three (35) Kong, S. C.; Marriott, C. D.; Reitz, R. D.; Christensen, M. Modeling and experiments of HCCI engine combustion using detailed chemical kinetics with multidimensional CFD. SAE Tech. Pap. Ser. 2001, 2001-011026. (36) Kong, S. C.; Han, Z. Y.; Reitz, R. D. The development and application of a diesel ignition and combustion model for multidimensional engine simulations. SAE Tech. Pap. Ser. 1995, 950278.


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Figure 2. Relationship of DME/methanol percentage and injection timings.

Figure 4. (a) Effects of injection timing on (a) in-cylinder pressure and (b) rate of heat release at φMeOH ) 0.5.

Figure 3. (a) Effects of injection timing on (a) in-cylinder pressure and (b) rate of heat release at φMeOH ) 0.119.

obvious peaks at later injection timings (i.e., -6 °CA ATDC). With advanced injection timing, the third heat release peak disappears and the second peak first clearly increases and then decreases at the injection timing -66 °CA ATDC. Furthermore, the first peak decreases and the phase in which it appears is clearly delayed at the injection timing -66 °CA ATDC. Figure 4 shows that the trend at higher methanol concentrations is different from Figure 3. It can be seen in Figure 4that the maximum peak pressure and the second peak of heat release increase sharply when the injection timing is advanced. The above phenomena result from the effect of various combustion modes. By adopting this dual-fuel compound combustion approach, a homogeneous DME/air mixture is formed in the intake and compression strokes, and compressed ignition is reached due to its high cetane number as temperature and pressure in the cylinder reach DME’s ignition conditions. Its heat release process has obvious HCCI combustion characteristics and the combustion duration is very short. The early

combustion of DME supplies a thermo-atmosphere so that the following injected methanol can be ignited quickly after undergoing a short ignition delay. As the injection timing is relatively late (i.e., case 1), much methanol cannot be thoroughly mixed with air prior to ignition, so the primary combustion of methanol here is the diffusion combustion controlled by the mixing rate, in which the third peak of heat release rate appears due to the lower heat release rate and longer combustion duration. Therefore, for later injection, the combustion process comprises three stages: low-temperature reaction of DME, hightemperature reaction of DME, and diffusion combustion of methanol. In this case, the combustion mode is the HCCI combustion of DME and the low-temperature diffusion combustion of methanol. As the injection timing is advanced, more time is available for injected methanol to mix with air prior to ignition. Therefore, more methanol/DME/air mixture ignites simultaneously, so the maximum peak pressure and the second peak of heat release rate rise significantly. If the injected methanol concentration is increased to some extent in this case, it could cause sharp heat release, even knock (see curves of injection timing at -36, -46, and -66 °CA ATDC in Figure 4b). Therefore, late injection is suggested to achieve smooth combustion at high methanol concentrations for this dual-fuel compound combustion approach. When the injection timing is advanced to -66 °CA ATDC, the maximum peak pressure is lowest and the heat release rate curve shows little heat release due to partial burning at low methanol concentrations (see Figure 3b). This shows that a higher DME percentage is required to achieve normal combustion at low methanol concentrations and early injection timings. The above results also indicate that the injected methanol concentration has an evident effect on the methanol injection timing. Figure 5 shows the effects of methanol injection timing and concentration on IMEP. It can be seen that, with advanced

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Figure 5. Effects of methanol injection timing and concentration on IMEP.

Figure 6. Effects of methanol injection timing and concentration on ITE.

injection timing, IMEP first increases and then decreases at low methanol concentrations, whereas IMEP increases monotonously at higher methanol concentrations. Later injection results in later ignition and longer combustion duration due to the diffusion combustion of methanol, so lower IMEP is caused by the increased power loss and incomplete combustion. As the injection timing is advanced, the degree of incomplete combustion decreases, and the phase of ignition and heat release are more suitable, so IMEP increases accordingly. At the injection timing -66 °CA ATDC, IMEP decreases sharply at low methanol concentrations due to partial burning. Furthermore, IMEP increases obviously with increased methanol concentrations. Figure 6 show the change in indicated thermal efficiency (ITE) with varied methanol injection timings and concentrations. Indicated thermal efficiency is calculated from the pressure in the cylinder and the fuel flow rate.37 It can be seen from this figure that, with advanced injection timing, ITE first increases and then decreases at low methanol concentrations, whereas it increases monotonously at higher methanol concentrations. Furthermore, ITE increases with increasing methanol concentration. At later injection timing, the longer combustion duration of methanol results in increased incomplete combustion products, so ITE decreases. With the advance of injection timing, ITE increases due to higher combustion temperature and more suitable phase of heat release. At the injection timing -66 °CA ATDC, ITE at low methanol concentration decreases sharply due to partial burning. Therefore, appropriate injection timing needs to be adopted to achieve higher ITE at low methanol (37) Olsson, J. O.; Tunestal, P.; Ulfvik, J.; Johansson, B. The effect of cooled EGR on emissions and performance of a turbocharged HCCI engine. SAE Tech. Pap. Ser. 2003, 2003-01-0743.

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Figure 7. Effects of methanol injection timing and concentration on HC emissions.

Figure 8. Effects of methanol injection timing and concentration on CO emissions.

concentrations. With an increase in methanol concentration, IMEP increases and combustion temperature also rises, so ITE increases. Figure 7 shows the effects of methanol injection timing and concentration on hydrocarbon (HC) emissions. With the advance of injection timing, HC emissions first decrease and then increase. With increasing methanol concentration, HC emissions decrease due to the rise in mean gas temperature. Figure 8 shows the effects of methanol injection timing and concentration on carbon monoxide emissions. With the advance of injection timing, CO emissions first decrease and then increase at low methanol concentrations, whereas it follows a decreasing trend at higher methanol concentrations. It is noted that CO emissions at methanol equivalence ratio 0.5 show an increase at intermediate injection timings, which is unexpected and needs to be further explored. With increasing methanol concentration, CO emissions decrease. Figure 9 shows the effects of methanol injection timing and concentration on NOx emissions. It can be seen that, contrary to HC emissions, NOx emissions first increase and then decrease with the advance of injection timing. With increasing methanol concentration, NOx emissions of earlier injection timing increases sharply due to knock combustion, whereas NOx emissions may be decreased to an acceptable level with suitable postponement of injection timing. Therefore, for this dual-fuel compound combustion, late injection would be a good choice at high methanol concentrations. 4.2. Modeling of CFD Coupled with Reduced Chemical Kinetics Mechanism. 4.2.1. Computational Details. In order to better understand the combustion characteristics of the dualfuel compound combustion with port injection of DME and incylinder direct injection of methanol, and the effect of injection



Figure 9. Effects of methanol injection timing and concentration on NOx emissions.

Figure 10. Computational grid and cut face at TDC.

timing on its combustion process, we carried out the modeling of CFD coupled with a DME/methanol reduced chemical kinetics mechanism. Major submodels used in the computation include the walljet1 wall interaction model, the Dukowicz evaporation model, and the HuhGosman breakup model. These models and some model constants are adjusted to enable the simulated pressure trace to approach the experimental one. All other associated models and model constants are left at their defaults. Details of these models can be found in the literature.38 Due to the different sizes of the intake and exhaust valves, the seven-hole nozzle and the piston bowl in the real engine have an eccentricity from the cylinder axis. Since the intake flow and the valves are not simulated, the eccentricity is not taken into account for purposes of computational efficiency. This allows the simulation only of a 51° sector, approximately 1/7 of the geometry. The total number of grid cells at bottom dead center (BDC) is about 30 000, and the total number of nodes is about 33 000. The numerical mesh and the position of the cut face at TDC are shown in Figure 10. The computations are started at the time of intake valve closure (IVC ) -141.5 °CA ATDC) and continue until exhaust valve opening (EVO )123.5 °CA ATDC). Based on the experimental data, initial pressure and temperature at IVC are adjusted to enable the simulation pressure and heat release rates to approach the measured results due to the effects of intake port heat transfer, in-cylinder flow, fuel atomization cooling, and residual gas content.39 The adjusted initial pressure value is 0.102 MPa for the two simulated cases. The adjusted initial temperature of the simulated -6 °CA ATDC injection timing is 380 K, and that of the simulated -26 °CA ATDC injection timing is 375 K. The cylinder head, piston, and linear are set to be fixed wall, with surface temperatures estimated to be 503, 553, and 453 K, (38) AVL FIRE Spray, version 8.4, AVL LIST GmbH, 2005. (39) Sjoˇberg, M.; Dec, J. E. An investigation of the relationship between measured intake temperature, BDC temperature, and combustion phasing for premixed and DI HCCI engines. SAE Tech. Pap. Ser. 2004, 2004-011900.

Figure 11. Experimental and simulated pressure traces with injection timing at (a) -6 and (b) -26 °CA ATDC.

respectively. There is some heat transfer through these walls. The wall heat transfer model is the Han-Reitz model, which formulates a correctional temperature wall function that accounts for variations of gas density and turbulence Prandtl number in the boundary layer. Details of the model can be found in the original literature.40 4.2.2. Calculation Results and Analysis. 4.2.2.1. Model Validation. The experimental results show that late injection is a good method for controlling the dual-fuel compound combustion process and preventing NOx formation. Consequently, this simulation is focused on the late injection case. Two injection timings, -6 and -26 °CA ATDC, are simulated. The amount of methanol injected, corresponding to a methanol fuel/air equivalence ratio of φMeOH ) 0.119, is the same for these two injection timings. Figure 11 panels a and b illustrate the calculated and experimental pressure curves, respectively, at these two injection timings. The figure indicates that the calculated pressure traces are in a good agreement with the measured data in both cases. It is shown that the submodels, boundary condition, and initialization condition used in the simulation are effective, so they can be used for further investigation. Nitrogen oxides (NOx) in engines contain nitric oxide (NO) and nitrogen dioxide (NO2), in which NO is dominant. Therefore, the NOx formation submodel is focused on NO simulation. Figures 12 and 13 show calculated in-cylinder mean temperature and NO mass fraction of these two injection timings, respectively. It can be seen that in the case of injection timing (40) Han, Z. W.; Reitz, R. D. Int. J. Heat Mass Transfer 1997, 40 (3), 613–625.

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Figure 12. Simulated in-cylinder mean temperature for injection timing at -26 and -6 °CA ATDC.

Figure 13. Simulated NO concentration for injection timing at -26 and -6 °CA ATDC.

at -26 °CA ATDC, the maximum mean temperature is higher than that with the injection timing at -6 °CA ATDC. Similarly, in the case of injection timing at -26 °CA ATDC, the NO mass fraction is larger than that for the injection timing at -6 °CA ATDC. These trends, which agree with experimental results, indicate that the NO formation submodel is reasonable in the reduced mechanism. 4.2.2.2. Effect of Injection Timing on DME Combustion Process. Experimental investigation shows that pure DME HCCI has a two-stage heat release behavior, with low- and hightemperature reactions.41 We now investigate how the injection timing affects the heat release process of DME in the dual-fuel compound combustion, composed of HCCI and diffusion combustion. Figures 14 and 15 show the effects of injection timings (i.e., -6 and -26 °CA ATDC) on DME concentration history and temperature history, respectively. At the beginning of the lowtemperature reaction (i.e., -27 °CA ATDC), the ignition locations for both of these injection timings is at the bottom of the combustion chamber, where gas temperature is higher. At the end of the low-temperature reaction (i.e., -18 °CA ATDC), DME concentration and temperature distribution are almost homogeneous in the whole combustion chamber, except for the linear and crevice. This result shows that the methanol injected at -26 °CA ATDC has no effect on the low-temperature reaction of DME HCCI. During the high-temperature reaction of DME HCCI, the effect of injection timing is quite noticeable. At the initial stage (41) Zheng, Z. Q.; Shi, C. T.; Yao, M. F. Experimental study on dimethyl ether combustion process in homogeneous charge compression ignition mode. Trans. Tianjin UniV. 2004, 10 (4), 241–246.

Chen et al.

(i.e., -8 °CA ATDC), with injection timing at -26 °CA ATDC, DME is consumed rapidly in a small range of the central section of the combustion chamber, where the gas temperature is higher and reaches 1300 K. In contrast, when the injection timing is at -6 °CA ATDC, DME reacts simultaneously within the whole chamber, where the temperature discrepancy is slight. Until -5 °CA ATDC, with injection timing at -26 °CA ATDC, almost all the in-cylinder DME is burned except for the portion near the bottom of the cylinder head and in the crevice, whereas only a primary portion of the DME within the chamber is reacted when the injection timing is at -6 °CA ATDC. This phenomenon shows that the methanol injected at -26 °CA ATDC accelerates DME’s heat release rate. At the end of the hightemperature reaction (i.e., TDC), all the DME is burned except for that in the crevice. By contrast, the amount of remaining DME in the crevice is larger when the injection timing is at -6 °CA ATDC, which is then a leading source of HC emissions. From the view of temperature distribution, in the case of injection timing at -26 °CA ATDC, a local high-temperature region appears near the chamber wall and the maximum temperature reaches 3000 K. In contrast, with injection timing at -6 °CA ATDC, the gas temperature is approximately homogeneous at about 1350 K in most of the chamber here. This result shows that the methanol injected at -26 °CA ATDC has a remarkable effect on the high-temperature reaction of DME HCCI. It changes the location where the high-temperature reaction starts and also accelerates the high-temperature reaction rate. Furthermore, the local high-temperature region appears in favor of NOx formation. 4.2.2.3. Effect of Injection Timing on Methanol Combustion Process. Figure 16 shows the in-cylinder temperature history after methanol is injected at -26 and -6 °CA ATDC. At the early stage of injection, in the case of injection timing at -26 °CA ATDC, the in-cylinder mean temperature is lower than that of -6 °CA ATDC injection. At 25 °CA after the start of injection, the mean temperature of -26 °CA ATDC injection rises more quickly. At 20 °CA after TDC, in the case of injection timing at -26 °CA ATDC, the high-temperature combustion region is wider and concentrated within the combustion chamber. As the methanol is injected at -6 °CA ATDC, in contrast, the high-temperature combustion region is narrower and is dispersed in the compression clearance and near the chamber wall. Figure 17 shows the in-cylinder NO concentration history after methanol is injected at -26 and -6 °CA ATDC. It can be found that NO correlates with the in-cylinder temperature. At 15 °CA after the start of injection, a small quantity of methanol ignites and a little NO is formed in the combustion chamber in the case of injection timing at -6 °CA ATDC. At 25 °CA after the start of injection, in the case of injection timing at -26 °CA ATDC, more NO is formed due to the higher combustion temperature. At 20 °CA after TDC, in the case of injection timing at -26 °CA ATDC, NO is widely distributed in most parts of the combustion chamber and its concentration is far more than that of -6 °CA ATDC injection. Figure 18 shows the in-cylinder methanol vapor concentration history after methanol is injected at -26 and -6 °CA ATDC. It can be seen that, at the beginning stage of injection, little methanol is vaporized in the case of injection timing at -26 °CA ATDC due to the relatively low in-cylinder temperature. At 15 °CA after the start of injection, the methanol vapor concentration of -26 °CA ATDC injection is higher, whereas that of -6 °CA ATDC injection decreases due to the combustion of methanol. At 25 °CA after the start of injection, methanol


Figure 14. Effect of injection timings on DME concentration history.

Figure 15. Effect of injection timings on in-cylinder temperature history.


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Figure 16. In-cylinder temperature history after methanol is injected at -26 and -6 °CA ATDC.

Figure 17. In-cylinder NO history after methanol is injected at -26 and -6 °CA ATDC.

vapor of -26 °CA ATDC injection decreases greatly due to fast burning. At 20 °CA after TDC, in the case of injection timing at -26 °CA ATDC, methanol vapor is oxidized almost completely, whereas a little methanol remains near the chamber wall in the case of injection timing at -6 °CA ATDC, so that its combustion duration is prolonged. 4.2.2.4. Prediction Based on Validated Model. As found by the above study, later injection timing (e.g., -6 °CA ATDC) results in lower NOx emissions but causes longer combustion duration and further lowered thermal efficiency, especially at high load conditions. Therefore, we perform a simulation to explore how to shorten the combustion duration and improve thermal efficiency at later injection. In the simulation, two kinds of injection rates are investigated. One has injection pressure at 20 MPa while the other

is at 30 MPa, as shown in Figure 19. The amount of injected methanol is kept constant for these two injection rates. The simulated in-cylinder pressure and rate of heat release can be seen in Figure 20. The corresponding in-cylinder mean temperature and NO mass fraction are shown in Figure 21. These figures indicate that though boosted injection pressure at a constant injected mass results in an increase in the maximum mean temperature and NO concentration, it can increase the peak in-cylinder pressure and shorten combustion duration, leading to an improvement in thermal efficiency. Although the present experimental study found that it is currently difficult to increase injection pressure due to the low viscosity of methanol, advanced fuel injection systems may make it possible in the future.



Figure 18. In-cylinder methanol concentration history after methanol is injected at -26 and -6 °CA ATDC.

Figure 19. Normalized injection rates at different injection pressures.

Figure 21. In-cylinder mean temperature and NO concentration at different injection pressures.

indicating experimental direction. Consequently, more detailed numerical calculations and experimental investigations will be carried out to further the goal of high-efficiency and low-emissions combustion of methanol/DME in a dual-fuel engine. 5. Conclusions

Figure 20. In-cylinder pressure and heat release rate at different injection pressures.

The above experimental and fundamental study developed a better understanding of the important role that injection timing plays in dual-fuel compound combustion. Furthermore, it is shown that numerical simulation based on the validation model is an effective tool for exploring undefined rules and

Through experiments and simulations on the dual-fuel compound combustion with methanol/DME, the following conclusions can be drawn: (1) Methanol injection timing has an obvious effect on the heat release process of the dual-fuel compound combustion. The effect of methanol concentration (i.e., the engine load) on the injection timing is also evident. (2) For early injection, the combustion controlled region is narrower. Late injection needs to be adopted to achieve smooth combustion and low NOx emissions at high methanol concentrations. Moderate injection timing should be adopted to attain higher ITE at low methanol concentrations.

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Table 3. Species and Reactions of New DME/Methanol Reduced Chemical Kinetics Mechanism species: CH3OCH3, O2, CH3OCH2, HO2, CH3OCH2O2, CH2OCH2O2H, O2CH2OCH2O2H, HO2CH2OCHO, OH, OCH2OCHO, H2O, H2, CH2O, HCO2, H2O2, CH3OCH2O2H, CH3OCH2O, CH3O, CH2OH, H, HCO, O, CO, CO2, CH3, CH3OH, N2, N, NO, N2O reaction 1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20 21 22

CH3OCH3 + O2 ) CH3OCH2 + HO2 CH3OCH2 + O2 ) CH3OCH2O2 CH3OCH2O2 ) CH2OCH2O2H CH2OCH2O2H + O2 ) O2CH2OCH2O2H O2CH2OCH2O2H ) HO2CH2OCHO + OH HO2CH2OCHO ) OCH2OCHO + OH CH3OCH3 + OH ) CH3OCH2 + H2O OCH2OCHO ) CH2O + HCO2 OH + OH (+ M) )H2O2 (+ M) H2O2 + O2 ) HO2 + HO2 H2O2 + O2 ) HO2 + HO2 CH2OCH2O2H ) OH + CH2O + CH2O CH3OCH2O2H ) CH3OCH2O + OH CH3OCH3 + HO2 ) CH3OCH2 + H2O2 CH3OCH2O ) CH3O + CH2O CH2OH + M ) CH2O + H + M CH2OH + O2 ) CH2O + HO2 CH2O + OH ) HCO + H2O CH2O + O ) HCO + OH CH2O + HO2 ) HCO +H2O2 CO + OH ) CO2 + H CO + HO2 ) CO2 + OH

A 4.10 × 1013 1.00 × 1012 7.42 × 1011 9.00 × 1011 9.00 × 1011 1.01 × 1020 1.40 × 108 5.05 × 1016 1.24 × 1014 2.96 × 1012 0.00 1.25 × 1013 1.83 × 1020 1.00 × 1013 6.48 × 1012 1.85 × 1024 1.00 × 1014 3.43 × 109 4.16 × 1011 5.60 × 1012 9.42 × 103 1.51 × 1014

b 0.0 0.0 0.0 0.0 0.0 -1.5 1.6 -1.6 -0.4 0.0 0.0 0.0 -1.5 0.0 -0.1 -2.5 0.0 1.2 0.6 0.0 2.3 0.0

E 44 910 0 18 500 0 18 500 44090 -35 15 400 0 38 150 38 150 18 160 47 160 17 690 14 870 34 190 5000 -447 2762 13 600 -2351 23 650

(3) CFD coupled with a new DME/methanol reduced chemical kinetics mechanism can effectively simulate the dual-fuel compound combustion and its NOx formation tendency. (4) Methanol injected at -26 °CA ATDC has no effect on the low-temperature reaction of DME HCCI, but it accelerates the heat release rate of the high-temperature reaction remarkably and leads to the appearance of a local high-temperature region. (5) In the case of injection timing at -26 °CA ATDC, the high-temperature combustion region is concentrated within the combustion chamber, which results in a higher NO concentration. When the injection timing is at -6 °CA ATDC, the high-temperature combustion region is dispersed in the compression clearance and near the combustion chamber wall, which leads to a relatively low NO concentration. (6) Increasing injection pressure is an effective way to shorten the duration of the methanol/DME dual-fuel compound combustion achieved by later injection.

reaction 23 24 25 26 27 28 29 30 31 32 33 34 35 36 37 38 39 40 41 42 43 44

A

CH3 + HO2 ) CH3O + OH HCO2 + M ) H + CO2 + M HCO + HO2 ) CH2O +O2 HCO + O2 ) CO + HO2 H2O + M ) H + OH + M HO2 + M ) H + O2 + M CH3OCH2 ) CH2O + CH3 CH3OH+OH)CH2OH+H2O CH3OH + OH ) CH3O + H2O CH3OH + HO2 ) CH2OH + H2O2 CH3OH + O2 ) CH2OH + HO2 CH3O + O2 ) CH2O + HO2 H2O2 + OH ) H2O + HO2 N + CO2 ) NO + CO N2 + O ) NO + N N + O2 ) NO + O NO + M ) N + O + M NO + NO ) N2 + O2 N2O + M ) N2 + O + M N2O + O ) N2 + O2 N2O + O ) NO + NO N2O + N ) N2 + NO

1.50 × 1013 2.29 × 1026 1.00 × 1014 9.10 × 1012 1.84 × 1027 3.57 × 1024 1.60 × 1013 1.44 × 106 6.30 × 106 5.20 × 1013 4.05 × 1013 6.30 × 1010 1.00 × 1013 1.90 × 1011 1.80 × 1014 9.00 × 109 9.64 × 1014 3.00 × 1011 1.26 × 1012 1.00 × 1014 6.92 × 1013 1.00 × 1013

b 0.0 -3.0 0.0 0.0 -3.0 -2.7 0.0 2.0 2.0 0.0 0.0 0.0 0.0 0.0 0.0 1.0 0.0 0.0 0.0 0.0 0.0 0.0

E 0 35 070 3000 410 122 600 51 620 25500 -840 1500 19 360 44 906 2600 1800 3400 76 100 6500 14 8300 65 000 62 620 28 200 26 630 20 000

Acknowledgment. We gratefully acknowledge support from the National Natural Science Foundation of China (NSFC) through Projects 50376046 and 50676066.

Nomenclature ATDC ) after top dead center BTDC ) before top dead center BDC ) bottom dead center CFD ) computational fluid dynamics DME ) dimethyl ether EGR ) exhaust gas recirculation EVO ) exhaust valve open HC ) hydrocarbon HCCI ) homogeneous charge compression ignition IMEP ) indicated mean effective pressure IVC ) intake valve closure NOx ) nitrogen oxides TDC ) top dead center SI ) spark ignition EF8010542

Dimethyl Ether Dual

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