Analysis of the Start of Combustion of a Diesel Fuel in a HCCI Process

Jan 23, 2008 - The CHEMKIN software, in conjunction with a one-zone model, has been used to compute the autoignition delay time for different engine ...
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Energy & Fuels 2008, 22, 987–995

987

Analysis of the Start of Combustion of a Diesel Fuel in a HCCI Process through an Integral Chemical Kinetic Model and Experimentation Miguel Torres Garcia,* Ricardo Chacartegui Ramirez, Francisco Jimenez-Espadafor Aguilar, and Tomas Sanchez Lencero Escuela Superior de Ingenieros de SeVilla, AVda. Camino de los Descubrimientos, s/n. 41092 SeVilla, Spain ReceiVed September 10, 2007. ReVised Manuscript ReceiVed NoVember 16, 2007

In this article, theoretical and experimental results for the chemical autoignition delay time of the homogeneous combustion compression ignition (HCCI) combustion mode with diesel fuel are presented. These results allow a further analysis of the chemical kinetic mechanisms that control the fuel oxidation in HCCI internal combustion engines. The CHEMKIN software, in conjunction with a one-zone model, has been used to compute the autoignition delay time for different engine performance parameters. The numerical results have been validated with experimental results obtained in a diesel engine adapted to HCCI combustion. Finally, the numerical model has been used to analyze different engine parameters such as fuel composition, intake temperature, exhaust gas recirculation (EGR), fuel air/ratio, and engine speed.

1. Introduction The autoignition combustion of homogeneous air-diesel fuel mixtures (homogeneous combustion compression ignition, HCCI) is a technology with a great potential in NOx, soot, and smoke emission reduction.1 It is based on the self-ignition of a homogeneous air-diesel fuel mixture without an external ignition source. The absence of a direct ignition mechanism involves the fact that ignition is controlled by the fuel oxidation kinetics. One of the main drawbacks of the combustion process is the control of the ignition point; it limits the applicability of this technology to different combustion systems. In diesel engines, there is a delay time between the start of injection and the start of combustion. The start of injection is given by the rising of the injector needle (it can be measured with an appropriate sensor). The start of combustion, or ignition point, is controlled by physical processes, such as the atomization of the fuel, fuel evaporation, and fuel-air mix, and by the chemical kinetic reactions among the fuel, air, and products of combustion. The ignition point has very important effects on the engine performance and on the polluting emissions. An ignition point situated around 4-8° of crank angle before top dead center (BTDC) minimizes the specific fuel consumption (SFC). If the ignition point is ahead with regard to the optimum, the pressure before top dead center (TDC) rises, penalizing the specific work and therefore the SFC. This also causes high pressure values before TDC, with high temperatures, increasing the NOx emissions. The temperature and pressure evolution of the air-fuel mixture are the main factors that determine the start of combustion. The presence of combustion products within the combustion volume is another determinant factor in the start of * To whom correspondence should be addressed. Phone: 0034954486111. Fax: 0034954487243. E-mail: [email protected]. (1) Thring, R. H. Homogeneous Charge Compression Ignition (HCCI) Engines. SAE Paper No. 892068.

the combustion. The exhaust gas recirculation (EGR) is an effective technique to reduce the NOx emissions, and it is widely used in different combustion systems, like the combustion engines or boilers, to achieve the environmental regulations. In the case of the HCCI, the EGR helps to stabilize and center the combustion process. The presence of exhaust gases contributes to the reduction of the oxygen concentration in a significant way; it increases the combustion duration, and it retards the start of the ignition. The EGR can be carried to the exhaust temperature using additional cooling to eliminate its direct heating potential. The direct simulation of the oxidation reactions of commercial diesel fuel is unavailable due to its composition, formed by thousands of chemical species, and the uncertainty of some mechanisms of oxidation. Thus, the commercial fuel is usually modeled departing from a replacement fuel (primary reference fuel, PRF) of well-known composition and kinetic mechanisms. Within the numerical kinetic models, there are on one hand detailed models with thousands of chemical reactions and hundred of species2 and on the other hand reduced models, which come from the detailed ones but with minor chemical information. In the 1980s, the kinetic reduced models based on hydrocarbon chain breaking acquired great importance due to the computational effort that the detailed models require, mainly when they are used with multizone models. The advances in the computer sciences have increased the use of detailed models. The HCCI simulation of air-diesel fuel mixtures is carried out with kinetic chemical models and commercial solvers. Two commercial solvers widely extended are the Hydrodynamics, Chemistry and Transport (HCT)3 and the CHEMKIN.4 Many (2) Zheng, J.; Yang, W.; Miller, D. L.; Cernansky, N. P. A Skeletal Chemical Kinetic Model for the HCCI Combustion Process. SAE Paper No. 2002-01-0423, 2002. (3) Lund, C. M. A General Computer Program for Calculling TimeDependent Phenomena Involving Onen-Dimensional Hydrodynamics, Transport, and Detailed Chemical Kinetics. LLNL Report, UCRL-52504, Aug 1987.

10.1021/ef700541z CCC: $40.75  2008 American Chemical Society Published on Web 01/23/2008

988 Energy & Fuels, Vol. 22, No. 2, 2008

investigation groups have used these programs for the simulation of homogeneous combustion in different combustion systems, mainly in engines, using one-zone models5 and multizone models6 to simulate the combustion process. The conclusions of these groups show that to obtain an accurate prediction of the pressure evolution in a HCCI it is indispensable to use multizone models or phenomenological one-zone models; however, for the estimation of some features as the start of the combustion, it is possible to obtain accurate results with onezone models. The high dependence of the ignition with the temperature evolution in the combustion chamber shows that many simulation models integrate the temperature distribution as a fundamental feature. At the present time, some multizone models combined with computational fluid dynamics (CFD) discretize the combustion volume in different zones, with a homogeneous distribution of temperatures in each one,7 establishing the connection of each one of the zones with the kinetic model. Other multizone models carry out a discretization of the combustion volume in different zones by means of statistic laws where each one has a mass fraction and a temperature8 connecting each one of them among them and with the kinetic model. An alternative to the multizone analysis of the HCCI combustion is to use a one-zone model with high kinetic sensibility, where it is assumed that the whole combustion volume is as a homogeneous mixture reactor with flat temperature. The one-zone models present good acuity for the prediction of the start of the combustion, but they have limitations in the simulation of the pressure evolution and in the predicted emissions of NOx, HC, and CO.9,10 This article presents a numerical integral model for the prediction of the ignition point in HCCI processes with mixtures of air-diesel fuel and combustion products. It uses a detailed kinetic model for a replacement fuel in one-zone model, developed with the solver CHEMKIN and validated with experimental results on an adapted to HCCI combustion diesel engine. 2. Experimental Study The experimental part of this work was based on the tests done on the DEUTZ FL1 906 engine. The original diesel engine was modified to adapt it to HCCI combustion, and to control the injection point with different operative conditions of engine speed, air/fuel ratio, and intake temperature. The main characteristics of the engine are shown in Table 1, and the experimental installation is represented in Figure 1. The modified engine systems were the following: (4) Lee, D.; Goto, S. Chemical Kinetic Study of a Cetane Number Enhancing Additive for an LGP DI Diesel Engine. SAE Paper No. 200001-0193, 2000. (5) Christensen, M.; Johanson, B.; Amneus, P.; Mauss, F. Supercharged Homogeneous Charge Compression Ignition. SAE Paper No. 980787, 1998. (6) Aceves, S. M.; Flowers, D. L.; Westbrook, C. K.; Smith, J. R.; Pitz, W. J.; Dibble, R. HCCI Combustion and Emissions. SAE Paper No. 200001-0327, 2000. (7) Aceves, S. M.; Martinez-Frias, J.; Flowers, D. L.; Smith, J. R.; Dibble, R. W.; Wright, J. F.; Hessel, R. P. A Decoupled Model of detailed Fluid Mechanics Followed by detailed chemical Kinetics for Prediction of Iso-Octane HCCI Combustion. SAE Paper No. 2001-01-3612, 2001. (8) Pöttker, S.; Eckert, P.; Merker, G. P. Homogeneous Charge Compression Ignition with Synthetic Fuels. MTZ Motortechnische Zeitschrift, Dec 2005. ISSN 0024-8525 10814. (9) Kimura, S.; Aoki, O.; Ogawa, H; Muranaka, S.; Enomoto, Y. New Combustión Concept for Ultra-Clean and High-Efficiency Small DI Diesel Engines. SAE Paper No. 1999-01-3681, 1999. (10) Kimura, S.; Aoki, O.; Kitahara, Y.; Aiyoshizawa, E. Ultra-Clean Combustión Technology Combining a Low-Temperature and Premixed Combustión Concept for Meeting Future Emission Standards. SAE Paper No. 2001-01-0200, 2001.

Torres Garcia et al. Table 1. Deutz FL1 906 Engine Characteristics type cylinder volume bore stroke compression ratio maximum power maximum torque fuel lubricating injection pump refrigeration a

DI monocylinder 4 stroke 708 cm3 95 mm 100 mm 19:1 15CV (11 kW) to 3000 rpma 45 Nm to 2100 rpm diesel gear oil pump mechanic air

rpm ) revolutions per minute.

Figure 1. Experimental installation scheme.

1° injection system: The injection pump was extracted out of the engine block, and its design was modified to control the injection point. The gear of the crankshaft coupling was substituted by a transmission belt. 2° exhaust gas recirculation (EGR) system: An EGR system is designed including an exhaust gas refrigeration system. With this system, the EGR bypassed mass fraction and the temperature of the recirculation gases can be controlled. 3° measurement equipment: Different sensors were located in the cylinder to monitor the pressure in the combustion chamber, injection pressure, needle rising, engine speed, temperature, and fuel mass flow and in the intake system to monitor the pressure evolution, temperature, and air mass flow. The start of the combustion is highly dependent on the intake temperature. To carry out a full analysis of the intake temperature effect under different operative conditions, a heating system for the intake air was designed (Figure 1). Also, a heat exchanger cooler in the EGR gases was installed to control the recirculated gas temperature and thus to control the temperature of the cylinder intake mixture. The installation incorporates stagnation boxes to eliminate the strong pulsating waves of the monocylinder engine.

3. Fuel Definition The development of a numeric model with high sensibility in the fuel oxidation kinetics needs the complete definition of the reactions that govern the process and the fuel composition. Due to the complex composition of the commercial fuels formed by thousands of hydrocarbons and with not completely characterized reactions for intermediate species, primary reference fuels (PRFs) are usually used to simulate the fuel behavior. With this simpler oxidation model, using PRF of well-known composition and oxidation reactions, can be obtained a similar behavior in determined characteristics as the cetane index, this has been validated with commercial fuels.11,12 (11) Richter, M.; et al. The Influencia of charge Inhomogeneity on the HCCI Combustion Process. SAE Technical Paper No. 2000-01-2868, 2000.

Analysis of the Start of Combustion of a Diesel Fuel

Energy & Fuels, Vol. 22, No. 2, 2008 989

Table 2. Substitution Fuel Characteristics

cetane index octane number boiling temperature F (kg/m3) (A/F)e

isooctane

n-heptane

100 118 700 15.13

0 98.4 686 15.18

5% iso-octane/ 95% n-heptane 56 101 687 15.16

diesel fuel 56 170 842 14.5

In this work, the blend of two hydrocarbons has been used as the PRF. These hydrocarbons are 2,2,4-trimethyl pentane (isooctane) with an octane number (ON) of 100 and n-heptane with an ON of 0. Their main features are shown in Table 2. A blend of 95% n-heptane and 5% iso-octane is used for the simulation of the commercial diesel fuel, with an equivalent cetane index of 56, which corresponds with the cetane index of the commercial diesel fuel used in the experiments. 4. Model for the Prediction of the Start of the Combustion The first numerical studies to analyze the effect of the EGR in HCCI engines was based on CFD codes combined with simplified kinetic models. However, in HCCI engines, the combustion is dominated by the chemical kinetics of the reactions, with a minor influence of transport phenomena or the injection characteristics. A good accuracy is achieved using a substitution fuel of equal cetane index13 and detailed kinetic oxidation models.14 In this work, the kinetic model used to simulate the combustion of the PFR has 1064 species and 4391 reactions. This detailed model allows carrying out a sensitivity analysis of the EGR effect on the ignition point. The great dependence of the start of combustion on the chemical kinetics, and the poor information used by the reduced models, shows that the reduced kinetic models that only provide global autoignition features should be disregarded. The detailed model used in this work is able to predict the cold flame phenomenon, which consists of a short period of combustion of low intensity previous to the start of the main combustion, due to the high reactivity that possesses the n-heptane.15,16 The in-cylinder temperature and the mass within the cylinder at the moment of the intake valve closing has been estimated from the results obtained from a one-dimensional CFD code used to simulate the evolution in the intake conduct. This estimated temperature is an input for the kinetic model used to predict the start of the combustion (Figure 2). Experimental tests with the same engine operative conditions as those of the numerical simulation were done to determine the start of the combustion, registering the pressure evolution and evaluating the heat released rate (HRR) applying the first principle of thermodynamics. Both results were compared by setting the match among the numerical results and the experimental results (Figure 2). (12) Christensen, M.; et al. Supercharged Homogeneous Charge Compression Ignition. SAE Paper No. 980787, 1998. (13) Hashimoto, K.; et al. Evaluation of Ignition Quality of LPG with Cetane Number Improver. SAE Technical Paper No. 2002-01-0870, 2002. (14) Curran, H. J.; Pitz, W. J.; Westbrook, C. K.; Callahan, C. V.; Dryer, F. L. Oxidation of Automotive Primary Reference Fuels at Elevated Pressures. Proc. Combust. Inst. 1998, 27, 379–387. Lawrence Livermore National Laboratory, Livermore, CA, UCRL-JC-133410. (15) Eng, J. A.; et al. The effect of POx on the Auto-ignition Chemistry of n-heptane and iso-octane in a HCCI engine. SAE Technical Paper No. 2002-01-2861, 2002. (16) Olsson, J.-O.; et al. Compression Ratio Influence on Maximum Load of Natural Gas Fueled HCCI Engine. SAE Technical Paper No. 2002-010111, 2002.

Figure 2. Exhaust pressure evolution (top figure), intake pressure evolution (middle figure), and exhaust and intake mass flow (bottom figure).

4.1. In-Cylinder Temperature and Mass at the Intake Valve Close. A one-dimensional CFD model applying eqs 1-3 in each one of the volumes in which was discretized the intake of the engine was used to get a precise estimation of the mass and the temperature within the cylinder, taking into account the evolution of the pressure waves at the intake and its effect on the intake and exhaust mass flows with the crankshaft angle and the engine speed (Figure 2). The temperature evolution and the oxidation reactions will be directly related to the initial conditions and the mass within the cylinder at the beginning of the compression stroke. ∂ Fu dS ∂F + (Fu) + )0 ∂t ∂x S dx -S

[

(1)

Fu2 ∂ ∂ ∂p dx - ξ πD dx ) (FuS dx) + (Fu2S) dx ∂x 2 ∂t ∂x

(

p u2 ∂ (FS dx) U + + ∂t F 2

)] [ ( +

(2)

)]

∂ p u2 (FuS) U + + dx ∂x F 2 q˙FS dx ) 0

(3) In Figure 2 are shown the results obtained from the CFD simulation, exhaust pressure evolution (top figure), intake pressure evolution (middle figure), and intake and exhaust mass flows (bottom figure) for two engine speeds. From these results, the effect of the pressure wave propagation at the exhaust and the intake systems and the different evolution of intake and exhaust mass flows as a function of the engine speed, and thus the mass within the cylinder at the beginning of the compression stroke, can be clearly appreciated. 4.2. Mathematical Model of the Start of the Combustion. Previous works modeling oxidation reactions have been carried out by Hu,17 with the creation of a reduced model of the oxidation of PRF to low, medium, and high temperature, and the semidetailed model of Soyhan.18 In this work, a zero-dimensional kinetic model has been used to carry out the analysis. The combustion chamber is considered as a perfect mixture reactor with variable volume, with even pressure distribution, temperature, and concentration of the chemical species. Heat losses have been calculated using Woschni’s correlation.19 (17) Hu, H.; et al. Autoignition of Adiabatically Compressed Combustible Gas Mixtures. SAE Paper No. 872110, 1987. (18) Soyhan, H. S.; et al. Automatic Reduction of Detailed Chemical Reaction Mechanisms for Auto-ignition under SI Engine Conditions. SAE Technical Paper No. 2000-01-1895, 2000. (19) Woschni, G. Universally Applicable Equation for the Instantaneous Heat Transfer Coefficient in the Internal Combustion Engine. SAE Technical Paper No. 670931, 1967.

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The model has been developed to evaluate the start of the combustion, and the hypotheses of the model will make it so that the model will be valid to evaluate this feature. Just before the start of combustion in a HCCI engine, these hypotheses are suited to the real conditions, where a homogeneous mixture is compressed, and the start of combustion takes place in the central zone of the cylinder. However, after the start of the combustion, the model will be invalid due to the fact that the temperature will not be uniform and the model will give a combustion duration shorter than the real one, as the experimental values show. Also, it will be overvaluing the maximum pressure reached inside the cylinder. Therefore, taking into account the fact that the kinetic model has been developed to predict the start of the combustion, the results here exposed are not affected by the derivative restrictions of the adopted hypotheses. The equations and simplifications that govern the mathematical model integrated in the software are described in the next paragraphs. Thermal State Equation. This model considers that the thermodynamic (P, T, V) variables only depend on the time, or what is the same thing, on the crank angle. Thus, the model does not include diffusive mass or heat models, with the mixture being perfectly homogeneous in the whole volume. In this model, the autoignition of the mixture happens in the whole volume in a simultaneous way and therefore it predicts an extremely fast combustion. To determine the pressure in a chamber and its evolution, the equation of the ideal gases will be used: 1064

∑ n (t)·R·T(t) i)1

V(t)

dV dQ -P + dt dt

∑ m˙ h ) dE dt i i

i

where dQ/dt (W) is the net thermal energy flow, including heat losses and the fuel contribution, P dV/dt is the power (W), mi is the mass flow of species i (kg/s), hi is the enthalpy (J/kg), and E (J) is the total energy of the system U + mgz + mc2/2, where of these terms only U is taken into account due to the homogeneous characteristics of the system of gases considered. Released Heat during the Process of Combustion. For any reaction “i”, the released or absorbed heat will come given by ∆Hi(T) (with a positive sign if the reaction is exothermic and a negative sign if it is endothermic). The released heat for reaction i is ˙ ) a((∆H )K [A]a[B]b Q i i i Extended for all of the reactions that take place in the model,

i

p(t) )

In this work, the values of the coefficients have been obtained from the high pressure kinetic detailed model for paraffin of Curran,20–23 in which the oxidation of any long hydrocarbon chain takes place in a series of steps that can be classified according to the elementary reactions that intervene in each one of them. First Principle of Thermodynamics. The energy balance was applied to the variable volume of the chamber to predict the temperature evolution in each time step. The first principle of thermodynamics applied to control the volume formed by the combustion chamber is the following:24

1064

)

∑ C (t)·R·T(t)

dQf 3491 ) a((∆Hi)Ki[A]a[B]b dt i)1

∑

i

i)1

1064

∑ n (t) ) total mole number

Evaluating the heat losses,24 the expression for the first principle of thermodynamic takes the form

i

1

1064

∑ C (t) ) species concentration (mol/L), i

d (C (T) T dt V

given by the kinetic equations

1

Volume of the Chamber. The cylinder volume in which the gas is confined varies with the time. No crevices were considered, and deposition of the fuel was also not considered. The variance of the volume of the chamber with the rotated crankshaft angle (or time) will come given by the geometry of the engine. The relationship will be the following one: V(θ) ) Vs -

( Vs -2 Vc )(1 + λ - cos θ - (λ - sen θ) 2

2

0,5

)

where θ ) 0° at bottom dead center (BDC)

(5)

EVolution of the Species, Kinetic Equations. The model is composed of 1064 species and 3491 kinetic reactions. The expression for the evolution of the concentration of component i (Ci) is the following: d[Ci] ) dt

η

φ

∑ K (T)∏ [C

φη] -

η

1

1

R

β

∑ K (T)∏ R

1

1

[Ci] dV [CβR] V dt

η is the number of reactions where Ci is a product; each one of these reactions has Φ reactives. R is the number of reactions where Ci is a reactive; each one of these reactions has β reactives, and among them would be Ci. The stoichiometric coefficients have not been included here to simplify the expression. In this expression, the first term corresponds to the Ci generation rate, the second term to the Ci consumption rate, and the last term to the variance of the volume.

1064

∑ i)1

3491

[Ci(t)V]) )

∑ ((∆H (T))V (t) - q r

r)1

r

w-P

dV dt

4.3. Definition of the Start of the Combustion. The start of the combustion is defined from the combustion delay time. The definition of the ignition has been taken from “HCCI Engines. Key research and development, 1994”,25 where it is assumed that the autoignition takes place when the concentration of an intermediate species (chemical radical) reaches a critical value. The ignition delay is defined from the injection point until the HO2 radical maximum concentration is reached. In a HCCI combustion process, a long ignition delay allows a more homogeneous mixture before its autoignition, avoiding the associated problems of non-homogeneous mixture combustion. In Figure 3, the delay time is schematized for the injection (20) Curran, H. J.; Gaffuri, P.; Pitz, W. J.; Westbrook, C. K. A Comprehensive Modeling Study of n-Heptane Oxidation. Combust. Flame 1998, 114, 149–177. (21) Curran, H. J.; Gaffuri, P.; Pitz, W. J.; Westbrook, C. K. A Comprehensive Modeling Study of iso-Octane Oxidation. Combust. Flame 2002, 129, 253–280. (22) Curran, H. J.; et al. Oxidation of Automotive Primary Reference Fuels al Elevated Pressures. Lawrence Livermore National Laboratory, EE.UU, SAE Technical Paper, 1999. (23) Curran, H. J.; Gaffuri, P.; Pitz, W. J.; Westbrook, C. K. A Comprehensive Modeling Study of iso-Octane Oxidation. Combust. Flame 2002, 129, 253–280. (24) Heywood, J. B. Internal Combustion Engine Fundamentals; McGraw-Hill Book Company: Singapore, 1998. (25) Assanis, D. N. Homogeneous Charge Compression Ignition (HCCI) Engines. Key Research and Development Issues. ISBN 0-7680-1123-X.

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Energy & Fuels, Vol. 22, No. 2, 2008 991

Figure 3. Start of combustion definition.

Figure 6. Start of combustion versus injection point with a relative fuel/air ratio of 0.3 with different engine speeds and different intake temperatures.

Figure 4. Detail of the chemical species evolution.

Figure 7. Start of combustion versus injection point with a relative fuel/air ratio of 0.8 with different engine speeds and different intake temperatures.

Figure 5. In-cylinder pressure evolution and cumulative total energy.

in the TDC. This is the method used by the kinetic detailed models for the estimation of the start of the combustion and adopted in this work for the analysis of the numerical results. In Figure 4, a detail of the start of the combustion is shown, where the evolution of the molar fraction of the most characteristic species in the process of combustion (H2O, CO2, O2, and radical HO2-) of an iso-octane and n-heptane mixture is represented as a function of the crank angle. The maximum of the concentration of radical HO2- corresponds with the start of the combustion. To define the start of the combustion from the experimental results, the definition given by Shimazaki has been adopted.26 In the experimental results, there is not kinetic information; therefore, the start of combustion is defined from the heat release rate (HRR) obtained from the registered in-cylinder pressure by means of a pressure transducer (Figure 5) combined with a

zero-dimensional model of the combustion chamber. Figure 5 shows the registered pressure data and the corresponding HRR for this pressure evolution. The HRR integration provides the cumulative energy (Figure 5); it is accepted that the combustion has begun when the total released energy is 5% of the total. 4.4. Validation of the Kinetic-Chemical Integral Model without EGR. In order to validate the kinetic model, systematic experimental tests were carried out in the diesel engine FL1 906, defined in Table 1. A wide range of injection angles were tested, from 10° BTDC up to 220° BTDC, with different intake temperatures, different engine speeds, and different air/fuel ratios. In the numerical simulations, a substitution blend for the diesel fuel was used (95% n-heptane and 5% iso-octane28). In Figures 6 and 7 are represented the ignition point versus the injection point, where experimental results (points) are confronted with numeric results provided by the kinetic-chemical model (continuous line). These results show a good concordance among experimental and numerical results. From these figures, it can be observed that the ignition crank angle is not dependent (26) Shimazaki, N.; Akagawa, H.; Tsujimura, K. An experimental study of premixed lean diesel combustion. SAE Technical Paper No. 1999-010181, 1999. (27) Tanaka, S.; Ayala, F.; Keck, J. C. A reduced chemical kinetic model for HCCI combustion of primary reference fuels in a rapid compression machine. Combust. Flame 2003, 113, 467–481. (28) Olsson, J.-O.; et al. A Turbocharged Dual Fuel HCCI Engine. SAE Technical Paper No. 2001-01-1896, 2001.

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Table 3. Intake Gas Properties Evaluated at the Start of Combustion Temperature

rpm

EGR (% mass)

engine torque (Nm)

1500 1500 1500 1500 1500

2 17 26 36 45

10 12 15 16 18

γ

Cp (J/(kg · K))

intake total mass (×10-5 kg)

1.31 1.26 1.24 1.18 1.12

1171.42 1200.99 1239.36 1272.47 1295.91

63.24 62.71 62.89 61.35 61.67

on the injection point for very early injections, far from the TDC. From these results, we can affirm that an advance of the ignition takes place with a retardation of the injection point, due to the fact that ignition is controlled by the reaction kinetics. The development of the ignition needs the evolution of some chemical radical (p.e HO) which are only susceptible of being formed under certain pressure and temperature conditions of the mixture. An early injection of the fuel introduces it in an insufficiently pressurized and insufficiently heated ambient, as to give place to the chemical reactions that form the precursor radicals of the combustion, so there is a range of injection points in which the ignition point is not affected by the injection point. This result is shown by several investigators.27 To begin the chemical processes that give place to the autoignition, a minimum temperature must be reached. When this temperature is reached, there is a vertiginous increase of olefine and hydroperoxide radical concentrations, which are the cause of the H2O2 radical formation with the increasing temperature and later on of the hydroxide radicals that unchain the ignition. If the injection is carried out very early, there are unfavorable conditions for the formation of these radicals, and kinetic processes are slower and therefore the ignition is delayed. However, there is a small range of the injection angle where the previous trend changes. For injection angles closer to the TDC, the effect of the injection point in the start of the combustion, and the delay angle have a similar behavior to the conventional diesel injection. The characteristic ranges of the effect of the injection point remain stable for all of the studied cases. The zone that presents weak dependence of the start of the combustion regarding the start of the injection corresponds to the interval of the injection point between the intake valve close to a 320° crank angle (with the TDC being 360°). The interval of the injection point with similar behavior to a conventional diesel corresponds to values of the injection point from 320 up to 350° of the crank angle. Thus, these results indicate that, except for other effects like the mixture stratification or the fuel condensation in the walls of the cylinder, the start of the injection should be carried out among 310-320° of the crank angle. The fuel injected before finds a unfavorable environment for the development of the ignition, with a high deposition of fuel on the wall of the cylinder taking place, penalizing the emissions of unburned hydrocarbons and the specific fuel consumption (SFC). If it is injected in the second zone in which the combustion is clearly dependent on the injection point and it conditions the engine indicator diagram strongly, the conditions correspond to the conventional diesel, raising the NOx and soot emissions drastically. In the previous figures, it is shown how the start of the combustion is advanced for the same injection points with the increase of the fuel/air ratio, fundamentally due to the induced effect that the fuel/air ratio has on the temperature of the intake and the great dependence of the start of the combustion with the temperature. Likewise, for higher fuel/air ratios, the curves are flatter compared with the curves obtained for lower fuel/air ratios. This corroborates the previous conclusion; the engine reaches higher

Figure 8. Effect of EGR in compression curves. Engine speed 1500 rpm. Mass fuel constant.

temperatures with higher fuel/air ratios and gives lower temperature variations with the crank angle in the first zone. 4.5. Validation of the Kinetic-Chemical Integral Model with EGR. The addition of combustion gases to the engine intake causes an ignition point delay associated with four characteristics of different nature: the effect of heating, the dilution effect, the capacitive effect, and the kinetic effect. In the first place, it should be kept in mind that blending combustion gases with the air in the collector increases the cylinder intake temperature due to the effect of direct heating of the combustion gases to higher temperature than the ambient air. The high cetane number of the diesel fuels produces an advance of the ignition point that penalizes the mechanical energy of the engine. To eliminate this effect in the experimental test (Figure 1) and in the analytic simulations, the recirculated combustion gases have been cooled down to the ambient temperature. Besides the heating effect due to the mixture of EGR and intake air is the induced heating to the mix in the intake process. The mixture in the intake to the cylinder suffers a progressive heating through the walls of the collectors and the walls of the cylinder. This heating will be affected by the recirculation of combustion gases due to the modification of the temperatures of the combustion process and the related temperature of the engine block. The addition of combustion gases inside the cylinder reduces the quantity of oxygen for a fixed fuel debit, and therefore, it elevates the fuel/air ratio. The reduction of the air/fuel ratio due to the EGR is named the dilution effect. The total heat capacity of the intake charge grows with the recirculation of the exhaust gases. This increase is due to the bigger heat capacity of the carbon dioxide (CO2) and of the steam (H2O) than the corresponding value of the air at the start of the HCCI combustion mode. The heat capacity increment when EGR is used is named the capacitive effect. This effect makes the temperature during the compression stroke of the mixture smaller, and therefore the pressure reached, as EGR increases. This fact is experimentally corroborated when raising the proportion of EGR in the combustion chamber. The addition of exhaust gases produces changes in the thermodynamic properties of the mixture of the intake gases. EGR is formed basically by CO2, H2O, O2, N2, and other products of combustion in smaller quantities, and it is the great proportion of CO2 and H2O that produces the change of the constant pressure heat capacity (Cp) of the mixture. At the same time, the heat capacity ratio (γ) of the mixture is smaller, and therefore, in the compression stroke, the pressure reached is lower as much as the EGR proportion is higher (see Table 3). In Figure 8, experimental results with constant engine speed and different EGR ratios are shown. In this figure, it is observed how the compression curve moves progressively down as the addition of EGR raises, at the same time that the ignition point is delayed. Finally the products of combustion with EGR can participate in different chemical reactions altering the reaction kinetics, what can modify the autoignition and the later combustion. This effect is named kinetic effect of the addition of exhaust gases. To validate the chemical kinetics model with the addition of

Analysis of the Start of Combustion of a Diesel Fuel

Figure 9. Chemical kinetic model validation with EGR. m × Cp versus ignition point. Constant consumption lines with different EGR ratios.

Figure 10. Delay time evolution with fuel composition.

exhaust gases, the product of the total intake mass gases by mean heat capacity of the mixture of gases has been used. Figure 9 shows the validation of the kinetic autoignition model with EGR, with an engine speed of 1500 rpm; in the abscissa axis is represented the product of the intake mass by the mean specific heat of the mixture of gases, evaluated to the temperature of the start of the combustion provided by the model and in the ordinates is represented the start of the combustion. Each one of the lines represents points with equal fuel consumption, with the EGR proportion varying along the curve. The curves represented correspond to the range from 0.98 to 2.3 kg/cycle for 0% of EGR; the mean heat capacity is increased with the EGR addition. Also, it is shown how the ignition point is delayed as the EGR proportion rises. An almost lineal behavior is appraised among the m × Cp increment and the delay angle. The black line indicates the zone where air is not available for the combustion. 5. Sensibility Analysis of the Model: Start of the Combustion for a Mixture of Iso-octane and n-Heptane in a HCCI Engine The numerical model has the purpose of providing an accurate estimation of the start of the combustion. In this section, the computational results obtained in the engine DEUZT-DITER FL1 906 adapted for HCCI combustion are analyzed and the effects in the start of the combustion of different engine performance features are analyzed. 5.1. Effect of the Substitution Fuel Composition on the Ignition Point. In previous analyses, the substitution fuel has been composed in a proportion of 95% n-heptane and 5% isooctane. An interesting study is the one that analyzes the sensibility of the start of the combustion with the proportion of each one of these hydrocarbons. In Figure 10, the effect of the composition of the fuel on the ignition retardation is shown. The analysis has been done to an engine speed of 1800 rpm

Energy & Fuels, Vol. 22, No. 2, 2008 993

Figure 11. Ignition point as a function of intake temperature.

Figure 12. Start combustion angle as a function of injection point for different relative fuel/air ratios up to a 1500 rpm engine speed.

and a relative fuel/air ratio of 0.3, with the injection point located 45° ABDC and an initial temperature of 340 K. In this figure, the composition varies from 100% n-heptane to 100% isooctane. The evolution of the autoignition point shown in Figure 10 with respect to the octane number shows the expected tendency with the n-heptane proportion in the mixture; when this proportion raises, the mixture trend to the self-ignition raises, and therefore the delay time diminishes. These results show the possibilities of the simulations done with these paraffins; with them, the modeling of most of the commercial gasolines and diesel fuels can be carried out. A comprehensive number of investigators have worked on numeric simulations of mixtures of replacement hydrocarbons constituted by mixtures of n-heptane and iso-octane which show up as most suitable to simulate the features of the autoignition of the commercial fuels without more than to modify the proportion of each one of the two hydrocarbons. 5.2. Effect of Initial Temperature and Fuel/Air Ratio in the Start of Combustion. Due to the start of the combustion being highly dependent on the fuel oxidation reaction chemical kinetics, the ignition point is directly conditioned by the mixture intake temperature. Most of the techniques used to control the HCCI combustion take into account this factor to center the combustion process. The combined use of high compression ratios and early injection require a precise control of the initial temperature of the mass retained in the cylinder. Figure 11 shows the delay time evolution for different initial temperatures. This simulation corresponds to 2400 rpm, a fuel consumption of 3 L/h of a mixture of 5% iso-octane and 95% n-heptane, a fuel air equivalence ratio of 0.6, and an injection point of 45° BTDC. It is observed that the increase of the temperature causes an advance of the ignition, diminishing the delay time, due to the formation of the precursor radicals of the combustion. In Figure 11 is observed the high dependence

994 Energy & Fuels, Vol. 22, No. 2, 2008

Torres Garcia et al.

Figure 13. Delay time as a function of the engine speed injecting at the BDC (time) and ignition point advance versus TDC (crack angle).

of the ignition point with the initial temperature; an increase of 40 °C advances the ignition point 8° of the crank angle. An increase of the fuel/air equivalence ratio causes an increase of the initial temperature when the intake valve closes. In Figure 12 are shown the results of the simulation for different fuel/air equivalence ratios and different injection points, with an engine speed of 1500 rpm. It is observed that, for all of the injection points, the fuel/air equivalence ratio has a great influence. An increase of the fuel/air equivalence ratio causes an advance of the combustion due to the effect on the initial temperature. A highly negative gradient is observed with a fuel/air equivalence ratio of 0.3 due to the temperature decrease in the engine, and it can be appreciated how this gradient becomes softer for higher fuel/air equivalence ratios. The effect of the fuel/air ratio becomes less outstanding for injection points next to the TDC, for injection points up to 340°, with a combustion mode similar to the conventional diesel combustion. From these results, it is derived that the control of the temperature before ignition will provide an efficient method to center the combustion and therefore for the good development of the HCCI combustion. 5.3. Effect of Engine Speed in the Start of Combustion. In this epigraph is analyzed the sensibility of the start of the combustion with the engine speed (from 1000 to 3000 rpm). Figure 13 shows the results of simulations with a fuel consumption of 3 L/h, a relative fuel/air ratio of 0.5, an initial temperature of 350 K, and the injection point in BDC. The results show that when the engine speed is increased, the delay time diminishes. Although the increase of the engine speed diminishes the relative heat losses, the effect of the increase of the speed prevails on the decrease of the delay time; this fact is increased for the little influence of turbulence in the start of the combustion on this process. Therefore, the start of the combustion angle comes closer to the TDC, raising the delay angle. 6. Conclusions The one-zone integral model for the estimation of the start of the combustion in a homogeneous mixture of air-diesel fuel presents a great agreement with the experimental results, being completely valid for the prediction of the start of the combustion. Both experimental and numerical results show that the point of the start of the combustion regarding the injection point presents two clearly differentiated zones. There is one zone

where the ignition is invariant with regard to the injection point; this effect is more accused as much as higher is the engine load and then the fuel/air ratio. Another is a zone where the ignition point is clearly affected by the injection point, with a retardation of the start of the combustion taking place when the injection of the fuel is retarded (it is closer to the TDC). Each one of these zones is clearly differentiated with an inflection point located around 40° BTDC, practically invariant to all of the operation conditions placed in this engine. Experimentally and with the simulations, the great dependence of the start of the combustion angle with the EGR has been proven, showing a quasi-lineal behavior. The analysis along with the computational model has shown the great influence of the intake temperature and the fuel/air equivalence ratio on the start of the combustion. Nomenclature m ˙ ) mass flow (kg/s) ˙ , q˙ ) heat flux (W) Q ∆H ) released heat Cp ) mean constant pressure specific heat capacity (A/F)e ) stoichiometric air/fuel ratio (F/A)R ) relative fuel/air ratio S ) area (m2) A, B ) chemical species a, b ) stoichiometric coeficient c ) gas speed (m/s) C ) molar concentration (mol/L) CO ) carbon monoxide CO2 ) carbon dioxide Cp ) constant pressure specific heat capacity (J/(kg · K)). CV ) constant volume specific heat capacity (J/(kg · K)) D ) pipe diameter (m) E ) total energy (J) g ) gravity acceleration (m/s2) K(T) ) kinetics constant m ) mass (kg) n ) mole number P, p ) pressure Q, q ) heat R ) universal gas constant T ) temperature (K) t ) time (s) U ) internal energy (J) u ) speed (m/s) V ) reaction speed (mol/s)

Analysis of the Start of Combustion of a Diesel Fuel V ) volume (m3) W ) power (W) x ) length (m) z ) height (m) Greek Letters ξ ) viscous coefficient γ ) Cp/CV ) specific heats ratio F ) density (kg/m3) θ ) crankshaft angle (rad) λ ) ratio between the crankshaft radius and the connecting rod length

Energy & Fuels, Vol. 22, No. 2, 2008 995 Subscripts i ) species r ) reaction f ) fuel T ) total c ) minimum volume of the combustion chamber (piston at the TDC) s ) maximum volume of the combustion chamber (piston at the TDC) w ) extracted to refrigeration h ) combustion chamber and liner wall EF700541Z