Numerical Modeling and Experimental Study of Combustion and Soot

Energy & Fuels 2007, 21, 1483-1492

1483

Numerical Modeling and Experimental Study of Combustion and Soot Formation in a Direct Injection Diesel Engine T. L. Chan*,† and X. B. Cheng†,‡ Research Centre for Combustion and Pollution Control, Department of Mechanical Engineering, The Hong Kong Polytechnic UniVersity, Hung Hom, Kowloon, Hong Kong, and School of Energy and Power Engineering, Huazhong UniVersity of Science & Technology, Wuhan 430074, China ReceiVed October 18, 2006. ReVised Manuscript ReceiVed February 23, 2007

In the present study, an improved multidimensional numerical code has been used to simulate the combustion and emission formation processes of direct injection diesel engine. Soot formation and oxidation processes have been modeled according to a hybrid particle turbulent transport controlled rate and soot oxidation rate expression. A reasonable agreement of the measured and computed data of in-cylinder pressure, heat release rate, NO, and soot emissions for different engine operation conditions has been made. The effects of fuel injection pressure and timing on the diesel engine combustion and emissions formation have been further computed based on an improved multidimensional combustion and soot model. The effects of different testing conditions on the gaseous and particulate emissions have been discussed. Predicted trends of soot and NO formation have also been presented together with the corresponding measured data. It has been demonstrated that the developed multidimensional engine combustion and emission formation model has provided good insight for new designs of different engine parameters and in-cylinder engine events.

1. Introduction The major problems associated with diesel engines are the high levels of nitrogen oxides (NOX) and particulate emissions. Because of their damage to our environment and their adverse effects on human health, there is worldwide concern over the exhaust of gaseous and particulate emissions from motor vehicles. Engine designers are facing increasingly stringent diesel vehicle emission standards on particulates and NOX. Several improved techniques are being explored in order to meet the challenging stringent vehicle emission standards such as the fuel injection pressure, the rate of fuel injection, the retardation of fuel injection timing, the inlet manifold pressure, the combustion chamber design, and the aftertreatment device, etc. Among those engine exhaust emissions, it is difficult to decrease the NOX and soot emissions simultaneously. It leads to more NOX formation typically if a better mixing and faster combustion process has been proposed. In this effort, the application of multidimensional models for a diesel engine appears promising in the engine design field. Soot formation is a persistent problem in fossil fuel combustion processes. The detailed kinetic mechanisms of soot formation in the pyrolysis, the combustion of hydrocarbons, and the soot particle dynamics have been studied extensively.1-4 Some * Corresponding author. Tel.: (852) 2766 6656. Fax: (852) 2365 4703. E-mail: [email protected]. † The Hong Kong Polytechnic University. ‡ Huazhong University of Science & Technology. (1) Wang, H.; Frenklach, M. A. Detailed Kinetic Modeling Study of Aromatics Formation in Laminar Premixed Acetylene and Ethylene Flames. Combust. Flame 1997, 10, 173-221. (2) Mauss, F.; Schafer, T.; Bockhorn, H. Inception and Growth of Soot Particles in Dependence on the Surrounding Gas-phase. Combust. Flame 1994, 99, 697-705. (3) Pitsch, H.; Barths, H.; Peters, N. Three-dimensional Modeling of NOx and Soot Formation in DI-Diesel Engine using Detailed Chemistry Based on the InteractiVe Flamelet Approach; SAE Technical Paper 962057, SAE: Warrendale, PA, 1996.

numerical models have been developed to represent the physical and chemical processes of the soot formation in a diesel engine, which have improved the complexity and capability of soot models.5-8 A better understanding of the in-cylinder phenomena in a diesel engine is important for the development of NOX and particulate matter (PM) reduction strategies. The in-cylinder phenomena are complicated because they include the liquid fuel atomization, vaporization, ignition, and combustion processes accompanied with their emission formation. Both in-cylinder computational fluid dynamics (CFD) simulations and advanced experimental diagnostics are being actively pursued in order to analyze the in-cylinder phenomena. Multidimensional models resolve the flow field spatially and temporally and include the submodels for their physical processes.9-11 Such models have been proven to be useful in predicting the in-cylinder events. Computer models such as the (4) Gopalakrishnan, V.; Abraham, J. Computed NO and Soot Distribution in Turbulent Transient Jets under Diesel Conditions. Combust. Sci. Technol. 2004, 176, 603-641. (5) Surovikin, V. F. Analytical Description of the Processes of Nucleusformation and Growth of Particles of Carbon Black in the Thermal Decomposition of Aromatic Hydrocarbons in the Gas Phase. Solid Fuel Chem. 1976, 10, 92-101. (6) Dan, T.; Takagishi, S.; Senda, J.; Fujimoto, M. Effect of Ambient Gas Properties for Characteristics of Non-reacting Diesel Fuel Spray; SAE Technical Paper 970352, SAE: Warrendale, PA, 1997. (7) Sung, N.; Lee, S.; Kim, H.; Kim, B. A Numerical Study on Soot Formation and Oxidation for a Direct Injection Diesel Engine. Proc. Inst. Mech. Eng. Part D: J. Automobile Eng. 2003, 217, 403-413. (8) Golovitchev, V. I.; Tao, F.; Chomiak, J. Numerical EValuation of Soot Formation Control at Diesel-like Conditions by Reducing Fuel Injection Timing; SAE Technical Paper 1999-01-3552, SAE: Warrendale, PA, 1999. (9) Belardini, P.; Bertoli, C.; Ciajolo, A.; D’Anna, A.; Del Giacomo, N. Three Dimensional Calculations of DI Diesel Engine Combustion and Comparison with In Cylinder Sampling ValVe Data; SAE Technical Paper 922225, SAE: Warrendale, PA, 1992. (10) Garo, A.; Said, R.; Borghi, R. Model of Soot Formation: Coupling of Turbulence and Soot Chemistry. In Soot Formation in CombustionMechanisms and Models; Bockhorn, H., Ed.; Springer-Verlag: Berlin and Heidelberg, 1995; pp 527-550.

10.1021/ef0605201 CCC: $37.00 © 2007 American Chemical Society Published on Web 04/10/2007

1484 Energy & Fuels, Vol. 21, No. 3, 2007

KIVA code are available to provide a better understanding of the engine performance and emission formation. They are especially useful when coupling with the experiments to identify the combustion and emission trends and provide direction to engine designers and, then, improve engine performance and efficiency. Hence, the aim of the present work is intended to investigate the nature of emission formation from a single cylinder of diesel engine for different operating conditions using a hybrid soot particle turbulent transport controlled rate and soot oxidation rate model. The effect of different fuel injection parameters such as the injection pressure and start of injection (SOI) on the characteristics of emissions from a diesel engine is presented. The effects of different engine operating conditions on the gaseous and particulate emissions are also discussed. On the basis of the detailed in-cylinder phenomena, insight into the governing processes provided by the simulation results, the various aspects of the combustion, and emission formation are also discussed. 2. Soot Model It is now well-established that the soot production process occurs in two main phases: soot formation and soot oxidation.12 These processes depend on the fuel composition, cylinder gas pressure and temperature, and local fuel and oxygen concentration. Soot formation is intrinsic to the diffusion of the diesel combustion process. Considerable progress has been made in the last few decades toward the advanced understanding of the physical mechanisms of soot formation and burn out in combustion systems. So far, the proposed soot models are classified into three categories, namely, empirical models, semiempirical models, and detailed chemistry models. Although the detailed mechanisms of the soot formation and oxidation process in diesel engines are not so clear, several theories have been proposed to explain these processes. 2.1. Previous Soot Models. Khan et al.13 presented a correlation equation of soot formation based on high-pressure diesel engine data. As a result, they further assumed that the rate of soot formation was controlled entirely by the formation of soot particles (i.e., the soot inception rate). Parameters in this model were local values of temperature, unburned hydrocarbon concentration, and equivalence ratio. Their proposed model did not take into account the complete soot oxidation process but is just an empirical approach. Tesner et al.14 presented a kinetic scheme on the soot/carbon formation. It was based on a chain-type process involving the radical nuclei from which the soot particle will grow later. It is a semiempirical model, which incorporated some aspects of the physics and chemistry of the phenomenon, as opposed to a correlation of experimental data. Magnussen and Hjertager15 presented a model for soot formation from turbulent combustion. They used Tesner’s soot formation model14 in which the soot oxidation was determined by the amount of oxygen and the dissipation rate of turbulent eddies. (11) Flagan, R. C.; Seinfeld, J. H. Fundamentals of Air Pollution Engineering; Prentice Hall Inc.: New York, 1988. (12) Heywood, J. B. Internal Combustion Engine Fundamentals; McGrawHill Book Company: New York, 1988. (13) Khan, I. M.; Greeves, G.; Wang, C. H. T. Factors Affecting Smoke and Gaseous Emissions from Direct Injection Diesel Engine and a Method of Calculation; SAE Technical Paper 730169, SAE: Warrendale, PA, 1973. (14) Tesner, P. A.; Snegirio, T. D.; Knorre, V. G. Kinetics of Dispersed Carbon Formation. Combust. Flame 1971, 17, 253-260. (15) Magnussen, B. F.; Hjertager, B. H. On Mathematical Modeling of Turbulent Combustion with Special Emphasis on Soot Formation and Combustion. Proceedings of the Sixteenth Symposium (International) on Combustion; The Combustion Institute: Pittsburg, PA, 1976; pp 719-729.

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Surovikin5 also improved Tesner’s soot formation model by adding an intermediate step to describe the growth of the radical nuclei and their conversion. It was a detailed chemistry kinetic model including detailed polycyclic aromatic hydrocarbons (PAHs) kinetics and soot particle dynamics. The oxidation process was included in their formulation. The steps included the following: (i) the formation of radical nuclei; (ii) the growth of the nuclei to a critical diameter and the transformation to incipient particles with a physical surface; (iii) the growth of the incipient particles into carbon particles. Their rates of oxidation are expressed as

dn ) n0 + (f - gn)n - g0Nn dt

(1)

where n is the concentration of free radical nuclei; n0 is the spontaneous production rate of a radical; f is the branching reaction coefficient; g is the termination reaction coefficient; g0 is the rate of loss of nuclei due to the collision with soot particles; and N is the soot particles concentration. Surovikin’s model5 and the Nagle and Strickland-Constable (NSC) oxidation model7,8 are still used widely in phenomenological and numerical models of combustion systems including diesel engines. Their models have been implemented in the standard KIVA-3V code for engine simulation.16 In spite of the uncertainties of the soot formation and oxidation mechanisms, there are some important processes which have been adopted by many researchers. Hiroyasu et al.17 developed a simple soot model by which the soot mechanism was later applied for modeling the soot formation in a diesel engine by Belardini et al.9 This model predicts the production rate of net soot mass, M˙ s, by a single-step reaction between the formation rate of soot mass M˙ sf and the oxidation rate of soot mass M˙ so according to the following:

M˙ s ) M˙ sf - M˙ so

(2)

The Arrhenius soot mass formation and oxidation rates can be obtained from

M˙ sf ) KfMfv

(3)

M˙ so ) KoMs

(4)

where Kf is the formation rate coefficient; Ko is the oxidation rate coefficient; Mfv is the fuel vapor mass; and Ms is the formed soot mass. The soot formation and oxidation rate coefficients, Kf and Ko, can be expressed as

Kf ) AfP0.5 exp(-Ef/RT)

(5)

Ko ) XO2AoP1.8 exp(-Eo/RT)

(6)

where Af and Ao are the pre-exponential factors; Ef and Eo are the activation energies; XO2 is the oxygen molar fraction; R is the specific gas constant; and T is the gas temperature. The activation energies and Arrhenius pre-exponential constants were modified from Belardini et al.9 to fit the data of the (16) Amsden, A. A. KIVA-3V: A Block-structured KIVA Program for Engines with Vertical or Canted ValVes; Los Alamos National Laboratory Report LA-13313-MS, LANL: Los Alamos, NM, 1997. (17) Hiroyasu, H.; Kodota, T. Models for Combustion and Formation of Nitric Oxide and Soot in Direct Injection Diesel Engines; SAE Technical Paper 760129, SAE: Warrendale, PA, 1976.

Combustion and Soot in a DI Diesel Engine

Energy & Fuels, Vol. 21, No. 3, 2007 1485

engine baseline case. This model, however, was found to give relatively low peak in-cylinder soot concentrations. A more realistic prediction was obtained by using the Nagle and Strickland-Constable (NSC) oxidation model18 in place of eq 4. In this model, the carbon oxidation occurs by two mechanisms whose rates depend on the surface chemistry involving more reactive A sites and less reactive B sites and the conversion of A sites to B sites. The chemical reactions are the following:

A + O2 f A + 2CO

(7)

B + O2 f A + 2CO

(8)

AfB

(9)

On the basis of eqs 4 and 6, the soot mass oxidation rate M˙ so is then replaced by the NSC oxidation rate:

M˙ so )

6mc M M˙ Fsds s t

(10)

where M˙ t is net reaction rate; mc is the carbon molecular weight; Fs is the soot density; and ds is generally an averaged soot particle diameter. 2.2. Hybrid Soot Particle Turbulent Transport Controlled Rate and Oxidation Rate Model. Magnussen and Hjertager15 deduced from the experiments on free diffusion flames that the effect of the turbulent mixing rate has a predominant role in the production of soot particles in a diesel engine. The soot was formed and contained in the turbulent eddies within the flame, and the burn-up of soot was related to the dissipation of the turbulence. In the direct injection (DI) diesel engine with strong air swirl, the mixing rate depends on the fuel injection and air swirl processes. The mixing process is controlled by the slower phenomenon, which is the air swirl. By comparing the typical time scales of soot formation and oxidation with those of turbulent flows in a diesel engine, Garo et al.10 concluded that the oxidation of soot particles is the process that would be largely influenced by the turbulent motions whereas the nucleation and coalescent coagulation would not. The formation of chains will be influenced to a certain extent. The Kolmogorov time scales are clearly smaller than the smallest eddies of the turbulence and are likely to influence the coagulation. The time scales of soot formation processes are critical for an incomplete oxidation. Soot is a solid-phase particle and each stage of soot formation includes complicated physical and chemical processes. The soot formation rate and the soot properties vary with its particle size. The particle size is quite small ranging from 1 to 10 nm in diameter which is of the same order as the gas mean free path. Hence, the soot can be considered as acting like the gas-phase species. After the soot particles are produced from a combustion process, the soot particles are then convected by the medium of gas flow. It is assumed that no slip occurs in between the soot and the fluid. Thus, the spatial distribution of soot is a function of the gas flow. The soot concentration transport equation is solved together with all other flow and spray equations as follows: (18) Kong, S. C.; Han, Z.; Reitz, R. D. The DeVelopment and Application of a Diesel Ignition and Combustion Model for Multidimensional Engine Simulations; SAE Techncial Paper 950278, SAE: Warrendale, PA, 1995.

( )

∂Ms ∂Ms 1 ∂ ∂ Ds + M˙ s + (ViMs) ) ∂t ∂xi Fs ∂xi ∂xi

(11)

where Vi is the velocity in the direction of space coordinates i ) 1, 2, 3; Fs is the soot density; and Ds is the effective soot diffusivity which can be determined from11

Ds )

kBTCc 3πµds

(12)

where kB is the Boltzman constant; µ is the gas viscosity; and T is the gas temperature. When the soot particle diameter approaches the same order as the mean free path λ of the suspending fluid, the resisting force offered by the fluid is smaller than that predicted by the Stokes law. To account for noncontinuum effects so that ds becomes smaller and smaller, a slip correction factor Cc is introduced into eq 12.11 If ds . λ, then Cc ≈ 1 and, in the free molecule regime, Ds varies as ds-1. On the other hand, if ds , λ, then Cc ≈ 1 + (1.657)(2λ/ds) and, in the free molecule regime, Ds varies as ds-2. In the realistic DI diesel engine combustion process, the oxidation reaction takes place on the surfaces of soot particles as soon as the soot particles formation occurs. The soot oxidation rate is proportional to the local concentration of soot and oxidant. The soot oxidation rate will be zero when one of those soot particles and oxidants is zero. In turbulent combustion, the soot particles are contained within the turbulent eddies. These soot particles are burnt up swiftly with the dissipation of these eddies in the soot oxidation zone. Since the chemical reactions usually have time scales which are short when comparing with those of the turbulent transport processes, the soot oxidation rate is determined by the rate of intermixing on a molecular scale of turbulent eddies (or the rate of dissipation of turbulent eddies). A hybrid soot oxidation rate model can then be expressed as

(

Mox MsSs , M˙ so ) C min Ms, κ Ss MsSs + MfuSf

)

(13)

where C is constant and taken as 4;15 κ is turbulent energy; is the dissipation rate of turbulent eddies; /κ is the eddy turnover time that represents the turbulent time scale; Mox is the mass fraction of oxidant; Mfu is the mass fraction of fuel; Ss is the chemical equivalence ratio of soot-oxygen; and Sf is the chemical equivalence ratio of fuel-oxygen. In the present study, Hiroyasu’s soot formation and oxidation model17 was used as the “original soot model”. The Arrhenius soot mass formation and oxidation rate and rate coefficients are given in eqs 3-6, respectively. But, the aforementioned hybrid soot oxidation rate model was implemented in the present improved KIVA-3V code in place of Hiroyasu’s soot oxidation model as the “present hybrid soot model” in order to compare with the original soot model. 3. Computational Models In the present study, the multidimensional combustion and emissions characteristics of a DI diesel engine have been simulated using an improved KIVA-3V code based on the modified turbulence, fuel spray, ignition and combustion, and spray models.18 This code solves the unsteady, compressible, and turbulent-reacting flows on finite-volume grids. With the addition and modification of many built-in submodels, it is now being widely applied and validated for engine combustion simulations. These improved submodels have been adequately


Chan and Cheng

described in the literature18-20 and hence are only briefly described here. 3.1. Turbulence Model. The renormalization group (RNG) k- turbulence model has been used for simulating the variabledensity turbulent flow within the engine combustion chamber. The modified RNG k- model21 is similar to the standard k- model except that an extra term is added into the dissipation equation of which the change of the mean strain rate and compressibility of the flow are accounted for. The RNG k- turbulence model predicts the shorter spray penetration and increases the mixing processes over the standard k- turbulence model. 3.2. Spray Model. The wave breakup model (also referred to as the Kelvin-Helmholtz (KH) breakup model) has been applied to simulate the primary breakup of fuel spray. This model assumes that the aerodynamic instabilities are responsible for the liquid fuel breakup within the dense core region. The wave or KH breakup model has been combined with the socalled Rayleigh-Taylor (RT) breakup model in order to estimate the disintegration of the blobs into the secondary fuel droplets. The RT breakup model describes the development of instabilities on a liquid-gas interface subject to the strong normal acceleration pointed toward the gas phase. The changing rate of fuel drop size is related to the frequency and wavelength of the fastest growing surface wave. The exponential growth of the perturbation of wavelength Λ with the maximum growth rate Ω dominates the whole breakup process. The fastest growing perturbation ultimately leads to the breakup phenomena.22 0.7 Λ 9.02(1 + 0.45xZ)(1 + 0.4T ) ) ro (1 + 0.865We 1.67)0.6

(14)

g

Ω)

0.34 + 0.38Weg1.5 (1 + Z)(1 + 1.4T ) 0.6

x

σ Flro3

(15)

where Weg ) (FgVrel2ro)/σ and Wel ) (FlVrel2ro)/σ are the Weber numbers for gas and liquid, respectively; Fl is the density for liquid; Fg is the density for gas; ro is the radius of the droplet; Vrel is the magnitude of the relative velocity; Z ) Wel0.5/Rel is the Ohnesorge number; Rel is the liquid Reynolds number; and T ) ZWeg0.5 is the Taylor number. 3.3. Ignition and Turbulent Combustion Model. The reaction mechanism used for simulating the hydrocarbon autoignition was the low temperature flame chemistry of the Shell model. This model is well suited for the simulation of autoignition phenomena in a diesel engine. The reaction scheme consists of a multistep chemical kinetics and chain propagation mechanism containing two routes, namely, the formation of branching agents and additional chain terminating reactions. The pre-exponential constant in the rate-limiting step, commonly referred to as Af04 in KIVA-3V code, is typically the important adjusted parameter. (19) Rutland, C. J.; Ayoub, N.; Han, Z.; Hampson, G.; Kong, S. C.; Mather, D.; Montgomery, D.; Musculus, M.; Patterson, M.; Pierpont, D.; Ricart, L.; Stephenson, P.; Reitz, R. D. Diesel Engine Model DeVelopment and Experiments; SAE Technical Paper 951200, SAE: Warrendale, PA, 1995. (20) Tennison, P. J.; Georjon, T. L.; Farrell, P. V.; Reitz, R. D. An Experimental and Numerical Study from a Common Rail Injection System for Use in an HSDI Diesel Engine; SAE Technical Paper 980810, SAE: Warrendale, PA, 1998. (21) Han, Z.; Reitz, R. D. Turbulence Modeling of Internal Combustion Engines using RNG k- Models. J. Combust. Sci. Technol. 1995, 106, 267295. (22) Reitz, R. D. Modelling Atomization Process in High-pressure Vaporizing Sprays. J. Atomization Spray Technol. 1987, 3, 309-337.

The laminar and turbulent characteristic time model was used to simulate the diesel engine combustion processes.21 For the high temperature flame chemistry combustion model based on the characteristic-time model, the changing rate of partial density of species i due to the conversion from one chemical species to another can be determined by

dYi Yi - Y/i )dt τc

(16)

where Yi is the mass fraction of species, i; Y/i is the instantaneous local thermodynamic equilibrium value of Yi; and τc is the characteristic time for the achievement of such equilibrium condition. The characteristic time τc can be formulated as

τc ) τl + fτt

(17)

where f is the delay coefficient for the controlling role of turbulent effects; τl is the laminar time scale, which is derived from the correlated one-step reaction rate from a single-droplet autoignition experiment; and τt is the turnover time of a turbulent eddy. The value of τc is assumed to be the same for these seven species considered, namely, fuel, O2, N2, CO, CO2, H2, and H2O, in order to predict each temperature accurately at the thermodynamic equilibrium. Among these seven species, six reactive species are accounted for in order to solve the instantaneous local thermodynamic equilibrium values Y/i . By using this combustion model, the chemical source term in the species continuity equation and the chemical heat release in the energy equation are computed. Since the total chemical time scale includes the turbulent time scale, the effects of turbulence on mean reaction rate are then accounted for. 3.4. NO Formation Model. In the present study, the extended Zeldovich thermal mechanism was used for modeling of nitric oxide (NO) emission from the diesel engine. Other sources of NO formation are neglected. This model consists of the following mechanisms: k1

} NO + N N2 + O {\ k -1

k2

N + O2 {\ } NO + O k -2

k3

N + OH {\ } NO + H k -3

(18)

(19)

(20)

Since the characteristic time of thermal NO formation is much longer than the characteristic physical and chemical time scales in the reaction front, it is possible to treat the thermal NO formation decoupled from the main combustion process. The rate coefficients for eqs 18-20 can be obtained from our recent study of Gui et al.23 The concentrations of O, OH, O2, H, and N can be approximated by a thermodynamic equilibrium assumption. 3.5. Heat Release Rate Model. The heat release rate dQ/dφ is computed by the following formula derived from the first law of thermodynamics:15 (23) Gui, B. C.; Chan, T. L.; Leung, C. W.; Xiao, J.; Wang, H. W.; Zhao, L. B. Modelling Study on the Combustion and Emissions Characteristics of a Light-duty DI Diesel Engine Fueled with Dimethyl Ether (DME) using a Detailed Chemical Kinetics Mechanism; SAE Technical Paper 200401-1839, SAE: Warrendale, PA, 2004.


(

)


(

)

m + 1 φ - φVB m -6.908 φ - φVB dQ e ) 6.908mfH dφ ∆φ ∆φ ∆φ

m+1

(21)

where Q is the heat release; φ is the crank angle; mf is the mass of fuel; H is the heating value; ∆φ is the duration of combustion, ∆φ ) φVE - φVB; φVB is the start of crank angle for combustion; φVE is the end of crank angle for combustion; and m is a constant and taken as 2.9.16 The heat transfer to the exposed cylinder wall Qw is computed by

dQW ) RAW(T - TW) dφ

Table 1. Engine Specifications type bore stroke con-rod length displacement rate power speed intake valve closure compression ratio type of injector swirl ratio

single-cylinder, DI 100 mm 115 mm 190 mm 903 cm3 9.5 kW 1600 rpm -131° ATDC 17.4 4 holes × 0.30 mm 1.3

Table 2. Engine Operating Conditions

(22)

where R is the heat transfer coefficient computed from Woschni’s heat transfer coefficient;12 Aw is the exposed cylinder area; T is the global temperature; and Tw is the cylinder wall temperature. The ignition delay is the crank angle interval from the start of the nozzle lift curve to the start of the rapid pressure rise. 4. Experimental and Computational Conditions Our recent numerical study on the combustion and emissions characteristics of a single cylinder of light-duty DI diesel engine fueled with dimethy ether (DME) was performed.23 This engine was well-characterized in our previous experimental and computational studies including their exhaust emissions measurements and in-cylinder pressure measurements. The tested engine specifications and operation conditions are listed in Tables 1 and 2, respectively. Because of the symmetry of the engine piston and becuase the fuel injector is located at the center of the cylinder and its four orifices are equally dispersed, one-fourth of the cylinder domain (a 90° sector) was then simulated. The geometry corresponds closely to that of the tested engine. The adaptive grid employed has about 14 539 cells at top-dead-center (TDC) and 15 420 cells at bottomdead-center (BDC). The tested engine was fully instrumented for temperature, pressure, airflow rate, and exhaust emissions measurements including in-cylinder pressure and fuel injection pressure. The exhaust gaseous samples such as the concentrations of NO/NOX, CO, CO2, HC, and particle size and mass concentrations were measured continuously using the California Analytical Instruments (CAI) model 400 HCLD for NO/NOX, CAI model 300 NDIR for CO/ CO2, CAI model 300 HFID analyzers for HC, TSI Model 3934 scanning mobility particle sizer (SMPS), TSI model 3310A aerodynamic particle sizer (APS), and TEOM model 1105, respectively.24,25 To ensure the reliability and repeatability of the measured data, each engine test condition was first allowed to run at the required constant engine speed/load for at least 15 min until the steady-state values had been reached as listed in Table 2. Detailed experimental procedures can be found in our recent work.24,25

5. Results and Discussions

engine speed (rpm)

1600 (baseline)

1600

1600

start of injection (BTDC) injector open pressure (MPa) fuel injected (mg/cycle) injection duration (CA) high engine load low engine load

25° 20 41.5 20° 75% 25%

20° 20 41.5 20° 75% 25%

12° 20 41.5 16° 75% 25%

in Figure 1. Comparing with the measured and computed cylinder pressure data, the larger difference of computed pressure data for using the present hybrid soot model and the original soot model is about 7.6% and 11.3%, respectively. Over the range of engine conditions considered, the computed pressure data show a good agreement with the measured data. This indicates that the computational models used are of sufficiently high accuracy based on the baseline conditions of the experimental setup, as shown in Figure 1. Figure 2 shows the net concentration of the soot formation and oxidation during the combustion process. However, the predicted soot concentration using the present hybrid soot model and original soot model is significantly different. The predicted soot concentration using the present hybrid soot model reaches its maximum value at about 25° crank angle after top-dead-center (CA ATDC); it then decreases rapidly. Comparing with the measured soot and NO

Figure 1. In-cylinder pressures for baseline engine case.

5.1. In-Cylinder Combustion Analysis. Comparison between the measured and computed in-cylinder pressure, and soot and NO concentration data for the engine baseline case are shown in Figures 1-3. The computed in-cylinder pressure curve using the present hybrid soot model shows a better agreement with the measured one than the original soot model, as shown (24) Wong, C. P.; Chan, T. L.; Leung, C. W. Characterisation of Diesel Exhaust Particle Number and Size Distributions using Mini-dilution Tunnel and Ejector-diluter Measurement Techniques. Atmos. EnViron. 2003, 37, 4435-4446. (25) Leung, D. Y. C.; Luo, Y.; Chan, T. L. Optimization of Exhaust Emissions of a Diesel Engine Fuelled with Biodiesel. Energy Fuels 2006, 20, 1015-1023.

Figure 2. In-cylinder soot concentrations for baseline engine case.


Figure 3. In-cylinder NO concentrations for baseline engine case.

Figure 4. Fuel vapor mass fraction field and droplet distribution at 10° BTDC for baseline engine case.

data under the steady-state condition, the present hybrid soot model demonstrates a better agreement, as shown in Figures 2 and 3. Figure 4 shows the cross-sectional “S-S” which is obtained by rotating 45° on the X-Y plane and passing through the spray axis for the visualization results of isosurfaces. It also shows that the impingement point of the fuel spray tip locates near to the upper edge of the piston bowl. A typical strong gas swirl motion within the cylinder demonstrates an important role in combustion process and formation, as shown in Figure 5. This strong gas swirl motion in the cylinder speeds up the fuel-air mix process and leads to the burning out of most fuels during the diffusion combustion period. The combustion process and emission formation for the engine baseline case are shown in Figures 6-10. The computed ignition time occurs at about 2° CA before top-dead-center (BTDC). The premixed-combustion is dominant at the early stage of the combustion process, which forms the first peak of its heat release rate. The premixed-combustion stage is very short for only about 5-8° CA. The small droplets of spray fuel leading to the jet edge are effectively mixed with the surrounding air before the ignition process takes place. The ignition process takes place at the spray front where the equivalent ratio and temperature in the cylinder are higher than other positions. The maximum heat release rate and in-cylinder pressure occur at this process stage. It also shows that the injected fuel reaches the edge of the piston bowl after the injection process has taken place. Due to the shape of the bowl, the injected fuel is turned toward to its cylinder head and eventually induces a counterclockwise vortex. A typical velocity vector plot can be found in Figure 5. The vortex pushes the fuel into the surface of the engine piston, and then, the concentration of fuel vapor increases when approaching the chamber wall. As a result, the flame exists near the surface of the engine piston. This phenomenon can be observed from the contribution of fuel vapor concentration and

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temperature fields in Figures 6 and 7. The fuel vapor is accumulated and forms extremely rich fuel areas at the near wall regions of the chamber. Figure 6 also shows that the ignition is expected to start at the near wall region and upper edge of the piston bowl rather than at the central spray region as the general view suggests. Partial unburnt fuel vapor is around at the rim of the bowl, but most of the fuel vapor diffuses on the top region of the engine piston due to the effect of the strong swirl gas motion in the engine cylinder. The combustion processes occur simultaneously on the top of the engine piston and inside the piston bowl. The nucleation of soot starts to form at the region where the ignition occurs. Shortly after ignition has taken place, a large number of soot particles are formed throughout the entire fuel stream during the premixed combustion phase. After 4° CA ATDC, a fully developed diffusion flame is formed. Soot formation mainly occurs during the diffusion combustion process. Since the soot emission is controlled by the dynamic characteristics of the fuel-air mixture flame, the soot formation processes depend on the fuel composition, gas pressure and temperature, and local fuel and oxygen concentration inside the engine. The air-fuel ratio obviously affects the soot formation. From Figures 2 and 11, the temporal evolution of the soot concentration first increases and then reaches the maximum of soot mass concentration when the global formation and oxidation rates are equal. Although the soot oxidation increases with increasing of the combustion temperature, the soot formation is also increased gradually with increasing of the numbers of fuel molecules. The high soot mass concentration is achieved at the upstream of the fuel jet. For the present hybrid soot model, the high soot mass concentration region is moved to the low soot concentration domain due to its flow characteristics. Eventually most of the soot particles are obtained in these domains. The high soot mass concentration field appears at the rich fuel mixture ratio rather than at the lean fuel mixture ratio. It is caused by the surface growth of soot particles promoted by their collisions with the acetylene molecules and the deposition of the gas-phase hydrocarbons on the hot surfaces of the soot particles. On the other hand, the rate of fuel supply decreases quickly and the ratio of fuel-air decreases due to the lower injection rate toward the end of injection, which results in a decrease in soot concentration. Figure 8 shows that the peak local soot concentrations locate at different regions where the fuel vapor contours in their rich mixtures condition. The formation fields of soot always appear near the wall of the piston bowl because the mixing process appears to be slower. Figure 9 shows that the high concentration of NO is also formed at the region of high temperature during the diffusion burn. A strong convection of oxygen is formed in this region. Figure 10 shows the predicted histories of OH mass fraction. The maximum OH mass fraction occurs at about 0° CA TDC, which is within the diffusion combustion portion of the engine cycle. Figure 11 shows the prediction of soot formation, soot oxidation, and net soot production, respectively. Soot oxidation is a heterogeneous process in which the oxidation reaction takes place at the surfaces of soot particles. Soot particles can be oxidized by the O atoms, OH radicals, and O2. Considering the OH radical as the primary oxidizing agent, Neoh et al.26 studied the consumption of soot mass based on the fundamental kinetic theory. In the flame environment, about 10% of collisions with (26) Neoh, K. G.; Howard, J. B.; Sarofim, A. F. Soot Oxidation in Flames. In Particulate Carbon Formation during Combustion; Siegla, D. C., Smith, G. W., Eds.; Plenum: New York, 1981; pp 261-282.



Figure 5. Typical velocity vector at 4 °CA ATDC for baseline engine case.

Figure 6. Isosurfaces of the equivalence ratio after TDC histories for baseline engine case.

Figure 7. Isosurfaces of temperature after TDC histories for baseline engine case.

OH radicals are effective into the gasification of a carbon atom. It is suggested that OH radicals may be very important in soot oxidation for the diffusion flame, even in the presence of significant concentrations of molecular oxygen. This is because the OH diffuses much faster than O2 and it attacks the soot precursors significantly. 5.2. Fuel Injection Timing. The computational results are highly sensitive to different fuel injection timing using the same spray, ignition, and, combustion models. Figures 12 and 13 show the heat release rate and in-cylinder temperature for -25, -20, and -12 SOI timings, respectively. The results of the retardation

of fuel injection decrease significantly for ignition delay but increase its combustion duration. It is noted that the average and peak heat release rate is lower when the retardation of fuel injection timing has taken place. Once the ignition is started, the combustion temperature rises rapidly to its peak for higher than 1500 K in these three cases. The highest average combustion temperature is for the -25 SOI case. It also produces the highest NO emissions among these three cases. It is interesting to note that all the fuel/air mixture is probably consumed at about 20° CA ATDC. The peak combustion temperature for the -25 SOI case begins to drop due to the in-cylinder


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Figure 8. Isosurfaces of soot mass fraction after TDC histories for baseline engine case.

Figure 9. Isosurfaces of NO mass fraction after TDC histories for baseline engine case.

Figure 10. Isosurfaces of OH mass fraction after TDC histories for baseline engine case.

expansion. Although the retardation of its fuel injection timing leads to a slight reduction of thermal efficiency, and slight increase of its exhaust gas temperature and fuel consumption rate, it has the benefit of reducing the rate of NO formation significantly. In general, the formation of NO and soot have strong relationships with the air-fuel ratio and combustion temperature. The soot concentration shows significant increase as the fuel

injection timing is retarded, as shown in Figure 14. The NO concentration, however, shows the reverse trend. It decreases as the fuel injection timing is retarded, as shown in Figure 15. The long ignition delay provides more time to form a flammable mixture for the premixed combustion phase. Therefore, a large number of nucleation species is provided. In the present study, the longest ignition delay occurs in the -12 SOI case but provides the lower in-cylinder temperature. This lower temper-


Figure 11. Soot mass formation and oxidation for baseline engine case at 75% engine load.

Figure 12. Heat release rate for different fuel injection timings.

Figure 13. Average temperature of engine cylinder for different fuel injection timings.

Figure 14. In-cylinder net soot concentrations for different fuel injection timings.

ature provides better conditions for the nucleation and, also, slows the oxidation of the soot particles. However, it should be noted that the peak concentration of soot appears in the regions where there is a small quantity of NO formation. It is due to the fact that the chemical dynamic reaction mechanism for the formation of the soot and NO is different under the same


Figure 15. In-cylinder NO concentrations for different fuel injection timings.

Figure 16. Fuel burnt concentrations for different injection timings.

temperature and pressure condition. The maximum soot formation rate does not appear in the regions where the combustion temperature is the highest. For the -25° SOI case, the computed highest combustion temperature reaches 2386 K for 16° CA ATDC, while the peak soot mass fraction is 2.48 × 10-3. However, the calculated maximum soot mass fraction is 4.25 × 10-3 and appears at 11.5° CA ATDC where the combustion temperature is 2125 K. The computed results show that when the flame temperature reaches above 2400 K, the soot concentration in the combustion diffusion flame is negligible. The rate of soot oxidation achieved at the highest value when the flame temperature is between 2350 and 2450 K. High temperature and lean fuel are the most important factors to form the soot emissions. Significant amounts of NO begin to appear in the in-cylinder region at about 1° CA BTDC for the baseline engine conditions, as shown in Figure 15. Comparing with the in-cylinder temperature, it can be observed that the highest NO concentration level is found in the same regions where the highest temperature exists. The peak NO formation rate occurs at about 16° CA ATDC, and the temperature reaches 2386 K here. Higher NO concentration levels may also be found in excessively rich fuel regions where they are transported after being formed much earlier in the engine combustion cycle. It turns out that the rates of the NO formation and depletion processes are highly dependent on the temperature and fuel concentration distribution. If the parameters of engine speed, load, and amount of injected fuel per cycle are kept constant in the present computation, the significantly higher fuel vaporization and burnt rate for the earlier fuel injection timing cases can then be observed, as shown in Figure 16. The fuel burnt mass curve increases significantly with advancing of the SOI timing. The large amount of diffusion flame will burn on the squish region. It can be found that the formation field of NO is primarily close to the wall of the piston bowl and just around the fuel injector


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Figure 19. Soot and NO tradeoff curve for different fuel injection timings at 25% engine load.

Figure 17. Measured and simulated normal NO and soot formations for different injection timings.

5.4. Effect of Engine Load. The proposed hybrid soot model was used to simulate different engine operating conditions. Figure 18 shows the soot-NO tradeoff results for the high engine load case for different fuel injection timings. The computed emission results for the low engine load are also shown in Figure 19. It shows that the soot and NO curve do not exhibit at the same tradeoff trend for the high engine load case. In fact, for the low engine load condition, the fuel ignition delay becomes longer when the fuel injection timing is further retarded. The longer ignition delay due to the retarded fuel injection timings results in a higher NO and a lower soot mass concentrations. The present hybrid soot model predictions somewhat capture the above phenomena at late fuel injection timings. 6. Conclusions

Figure 18. Soot and NO tradeoff for different injection timings at 75% engine load.

tip. When the combustion process is taking place, the NO peak concentration moves away from the fuel injector tip and moves to the squish region. 5.3. Soot-NO Tradeoff. Figure 17 shows the predicted trend in the normalized NO and soot concentrations for the three fuel injection timing cases. Figure 18 shows the soot and NO tradeoff for different fuel injection timings at 75% engine load. They exhibit the lowest NO concentration levels at the -12 SOI case and the lowest soot level at the -25 SOI case. Excessive retardation of fuel injection timing results in a longer ignition delay and increases the duration of premixing combustion. NO formation is reduced at the -20 SOI case as shown in Figure 19 with a slight decease in the soot formation. NO can be reduced significantly with a corresponding penalty of increasing the soot for the retardation of the SOI. Hence, it can be determined that the optimal fuel injection timing is the -20 SOI case for the tested diesel engine. However, for the retardation of fuel injection timing, the NO levels decrease slightly. These predicted results are consistent with the experimental observations. It can be concluded from the results shown in Figures 17 and 18 that the fuel injection timing of -20 SOI is a better scheme than the other cases in the present study.

The application of the improved CFD code for the simulation of combustion and emission formation in a high-speed diesel engine has been presented and discussed. The measured and computed data for the in-cylinder pressure, soot, and NO emissions agree reasonably well. The soot formation and oxidation process used in the present study is modeled according to a hybrid particles turbulent transport controlled rate and soot oxidation rate expression. The effects of soot concentration transport on the engine have also been investigated. The improved soot emission models have been used to evaluate the engine performance for different operating conditions such as the fuel injection timing and engine load. The results show that the soot concentration increases significantly with retarding of the fuel injection timing; on the contrary, the NO concentration decreases slightly. The optimal fuel injection timing is found to be -20 SOI in the present study. The results clearly show that the processes of soot and NO formation are predominantly governed by the local fuel-air ratio, in-cylinder gas pressure and temperature, and the engine operating conditions. It has also been demonstrated that the developed multidimensional engine combustion and emission formation model has provided good insight for new designs of different engine parameters and incylinder engine events. Acknowledgment. This work was supported by the grants from the Research Grants Council of the Hong Kong Special Administrative Region, China (RGC Project No. PolyU 5265/04E) and the Central Research Grants of The Hong Kong Polytechnic University (Project Nos. B-Q853 and A-PA2V). EF0605201

Numerical Modeling and Experimental Study of Combustion and Soot

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