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Numerical study of engine parameters on combustion and performance characteristics in an n-heptane fueled HCCI engine
Accepted Manuscript Research Paper Numerical study of engine parameters on combustion and performance characteristics in an n-heptane fueled HCCI engine M.M. Hasan, M.M. Rahman, K. Kadirgama, D. Ramasamy PII: DOI: Reference:
S1359-4311(16)32030-0 https://doi.org/10.1016/j.applthermaleng.2017.09.121 ATE 11178
To appear in:
Applied Thermal Engineering
Received Date: Revised Date: Accepted Date:
29 September 2016 11 April 2017 23 September 2017
Please cite this article as: M.M. Hasan, M.M. Rahman, K. Kadirgama, D. Ramasamy, Numerical study of engine parameters on combustion and performance characteristics in an n-heptane fueled HCCI engine, Applied Thermal Engineering (2017), doi: https://doi.org/10.1016/j.applthermaleng.2017.09.121
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Numerical study of engine parameters on combustion and performance characteristics in an n-heptane fueled HCCI engine M.M. Hasan1, M.M. Rahman1,2, K. Kadirgama1,2 and D. Ramasamy1,2 1
Automotive Engineering Research Group Faculty of Mechanical Engineering, Universiti Malaysia Pahang 26600 Pekan, Pahang, Malaysia Email: [email protected] Phone: +60149670211 2 Automotive Engineering Centre, Universiti Malaysia Pahang 26600 Pekan, Pahang, Malaysia Abstract Homogeneous charge compression ignition (HCCI) is an alternative combustion concept which offers significant benefits in terms of its high efficiency. However, the operational range using HCCI combustion in terms of speed and load is restricted due to the absence of the direct control of the onset of ignition and the heat release rate. In this work, a zero-dimensional single-zone numerical simulation with reduced fuel chemistry was developed and used to investigate the effect of various engine parameters on combustion and performance characteristics in an HCCI engine fueled with n-heptane. The simulations show good agreement while comparing the results with the published experimental results and capture important combustion phase trends as engine parameters are varied with a minimum percentage of error which is less than 6%. The combustion phase was advanced and the combustion duration was shortened with the increase of intake air temperature and the decrease of the engine speed. The maximum load was successfully increased with increasing the intake air pressure. The highest load in this work was 11.27 bar in IMEPg at the condition of 200 kPa in intake air pressure and 333 K in intake air temperature. Keywords: HCCI; single zone; n-heptane; engine parameters; combustion; performance. 1. Introduction With increasing concern about fuel economy and emissions, the internal combustion engine industry is constantly looking for better alternatives to spark ignition (SI) and compression ignition (CI) engines. Homogeneous charge compression ignition (HCCI) engine is one of the alternatives under extensive research in recent years [1, 2]. The name demonstrates its two essential characteristics, homogeneous mixture, and compression ignition. The HCCI concept promises several advantages. In short, it is more efficient than the SI engine and cleaner than the CI engine. Compared with the SI engine, the higher compression ratio can be used and leaner fuel-air mixture can be applied on an HCCI engine [3, 4]. At the same time, compared with the CI engine, the cylinder mixture is more evenly distributed in an HCCI engine, where fuel rich pocket is not possible. Without the soot-inducing fuel-rich pockets and NOX-inducing lean regions, the overall result is that the HCCI engine can achieve higher fuel efficiency with lower NOX and soot emissions [5].
Nomenclature a Ac Aw B Cd dQ dV hc hj i j l L m N p p0 pT Qh R Rc Ru s
T0 Tw U V Vc Vd p0
crank radius effective valve open area wall area bore discharge co-efficient heat release dependent on the variation of crank angle d variation of cylinder volume heat transfer coefficient enthalpy of flows entering or leaving the system mass fraction of species number of flows in or out of the system connecting rod length stroke length total mass in the system engine speed in-cylinder pressure upstream stagnation pressure downstream static pressure heat transfer ratio of connecting rod length compression ratio universal gas constant distance between crank axis and piston pin axis upstream stagnation temperature wall temperature internal energy instantaneous cylinder volume clearance volume displacement volume upstream stagnation pressure
W W net Y iin m in Sp
W mv Vd W
work net work done inlet mass fraction mass flow rate instantaneous piston speed mean molecular weight of the mixture displacement volume work
Greek symbols ratio of specific heats ?? indicated torque θ crank angle Ø equivalence ratio ω rotational speed mass reaction rate of the species i i Abbreviations AFR air-fuel ratio CFD computational fluid dynamics CI compression ignition EVO exhaust valve open HCCI homogeneous charge compression ignition IVC inlet valve close LTR low-temperature reaction MCS main combustion stage NTC negative temperature coefficient SI spark ignition SOC start of combustion TDC top dead center
However, differing from both CI and SI engines, HCCI engines lack a physical event which controls the ignition timing. It is the main challenge which ultimately influences the power and efficiency [6]. Unlike SI engines and CI engines, HCCI engines have no direct mechanism to control the SOC. Its SOC is fully dependent on auto-ignition. According to Yao, Zheng [7], this auto-ignition is affected by several factors like fuel auto-ignition chemistry and thermodynamic properties, combustion duration, wall temperatures, intake temperature and
pressure, compression ratio, amount of EGR, engine speed, engine temperature, convective heat transfer to the engine, and other engine parameters. Higher intake temperatures and intake pressures cause earlier combustion timing due to the faster chemical kinetics [8]. For equivalence ratio sensitive fuels, higher equivalence ratios (up to stoichiometric) have a tendency to propel ignition timing due to the improvement of charge reactivity. The fuels without equivalence ratio sensitivity demonstrate an inverse pattern, where higher equivalence ratios have a tendency to bring about more delayed combustion timing. The lower specific heat ratio from higher equivalence ratios causes the prerequisite of more compression heating to achieve ignition temperatures. Nonetheless, in real operation, for both sorts of fuels the utilization of higher equivalence ratios likewise increases the in-cylinder wall and residuals temperatures, along these lines bringing on advanced combustion timing [9]. For fuels with little to no equivalence ratio sensitivity, there are contending impacts where lower specific heat ratios have a tendency to delay combustion while higher in-cylinder wall and residual temperatures have a tendency to propel combustion. To understand the effect of engine parameters on combustion and performance characteristics of HCCI engine, a fruitful study is needed which can be done by developing a more accurate simulation model based on zero-dimension and single zone rather than experiments or multidimensional modeling. Experimental work is costly and requires more time. On the other hand, modeling is a way of exploring engine behavior in ways that might otherwise impossible in a true experimental approach. In addition, a multi-dimensional model requires substantial computational time as compared to zero-dimensional and quasi-dimensional models, which are a simplified version of a multi-dimensional model [10]. Thus, the use of simplified model becomes increasingly important because of its advantages in computational time and resources. A zero-dimensional simulation would be an interim solution until the cost and time of running a multi-dimensional model is comparable with the current cost of the zerodimensional model. The use of chemical kinetics mechanisms also helps in investigating the combustion behavior of an HCCI engine. The purpose of this study is to establish a numerical model based on single zone and zero-dimension for HCCI combustion over a wide range of operating conditions using n-heptane. The HCCI engines were modeled to investigate the effects of a number of engine parameters such as air-fuel ratio, intake air temperature, intake air pressure, compression ratio and engine speed on the combustion phasing of the low-temperature reaction (LTR) and main combustion stage (MCS). A reduced fuel chemistry for n-heptane was developed and compared to the combustion data obtained in [11]. The numerical simulation was able to capture the combustion phasing trends observed during the experiments with changes in several engine parameters. Based on the numerical results, it is believed that the model used in this study is suitable for investigating fuel chemistry effects on HCCI combustion. 2. Numerical Methodology 2.1 Numerical Model In order to design an engine for various fuels, accurate models are needed which are able to model both combustion and performance. In this study, a zero-dimensional single zone model for a single cylinder, four stroke HCCI engine was established using MATLAB. The proposed method takes into account the detailed thermodynamic aspects of the engine. The model integrates the ordinary differential equation system corresponding to the chemical and thermal evolution of a closed homogeneous system under an imposed volume history reproducing the
engine cycle. This single-zone in time model treats the cylindrical combustion chamber as a uniform reactor with uniform temperature, pressure, and composition throughout. The reactor volume changes based on the slider-crank relations which determine the motion of the piston in the engine cylinder. The simulation handles only the closed part of the cycle; intake and exhaust processes are not considered. This kind of model aims to simulate the auto-ignition process in the core of the air-fuel mixture and enables the use of detailed and reduced chemical kinetic models with which we can investigate the chemical reactions contributing to cylinder inside pressure and temperature. This study only considers the results obtained from combustion between inlet valve close (IVC) and exhaust valve open (EVO). The emissions produced after EVO are not of interest in this study. The effect of wall heat losses is neglected and the accumulated gas is assumed to be an ideal gas. Toward the start of the simulation, the engine parameters and beginning working condition were characterized, which depended on the experimental data. At that point, the mixture composition in the combustion chamber was introduced. It was accepted that the initial composition in the combustion chamber comprises a mixture of air and fuel. An ordinary valve profile was resolved to take into account the equation from [12], which was utilized to represent the valve movement furthermore to get the inlet mass flow rate. The entire simulation process is shown in Figure 1. The main part of the simulation is the place the main combustion happened, which was after IVC and before EVO. At this stage, the piston was in upward movement after IVC, compacting the gas mixture. The chemical kinetics plays an imperative part here as it will decide the start of combustion. Once the piston passes the top dead center (TDC) mark, it will be in descending movement once more, expanding the mixture and the simulation stops at EVO. The simulation solves the energy and species equations for the whole procedure. 2.2 Chemical Kinetic Mechanism In order to simulate the combustion in an HCCI engine, it is obvious to solve a series of chemical reactions. It can be done by developing a chemical reaction mechanism for a particular fuel. In this study, a reduced n-heptane mechanism was used which was developed by Seiser et al. [13]. The mechanism consists of 159 species and 770 elementary reactions. The mechanism was then utilized in the Cantera chemical kinetics software package [14] to get the chemical properties and reactions mechanisms. Cantera is an open source software package for chemical kinetics mechanism and is widely used [15]. The chemical kinetics files are available in Chemkin format. Cantera can convert this file to use it in MATLAB. The utilization of a chemical reaction mechanism empowers the investigation of the chemical species interaction, where the interaction is impacted by the temperature, pressure and species mass fraction. Besides, the chemical reaction mechanism would have the capacity to give a superior comprehension in the combustion study.
. Input 1. IVC (ATDC) 2. Pintake (Pa) 3. Tintake [K] 4. AFR 5. Bore (m) 6. Stroke (m) 7. Compression ratio 8. Connecting rod length (m) 9. RPM 10. Swirl ratio 11. Wall temperature [K]
θCA = θIVC Volume = f(θCA)
Start of combustion = f(T, φ) Burn duration (∆θ) = f(θSOC)
Reactant properties (CANTERA)
Motored temperature and pressure (Polytropic compression)
θCA = θSOC
Combustion phase (T, P)
Product properties (CANTERA)
θCA = θSOC + ∆θ T, P (Polytropic expansion) θCA = θEVO End of model T and P history during IVC→EVO
Figure 1: An algorithm flow chart for zero-dimensional single-zone model simulation. 2.3 Governing Equations One approach in engine modelling utilizes the geometry of the reciprocating engine to calculate the engine volume over the complete combustion cycle. The piston work in the energy equation
can be determined by the cylinder volume. The instantaneous cylinder volume at any crank angle location is calculated by below equation.
1 V V c 1 R c R 1 cos R 2 sin 2 2
(1)
Where V is for instantaneous volume, V c is the clearance volume, R is the ratio of connecting rod length to crank radius and Rc is the compression ratio. With the rate of change of volume
1 1 cos dV V c Rc sin 2 2 d R sin 2
(2)
The equation for compressible flow through a flow restriction is generally used to calculate the mass flow rate to the combustion chamber [16]. The equation uses the effect of the real gas flow and discharge coefficient C d which can be acquired from the experiments. A pressure difference is necessary between combustion chamber and ports to flow the mass in or out of the chamber. The equation is divided into two cases: choked and subsonic flows. For choked flow, the following conditions are followed:
pT 2 p0 1
1
1
2 2 1 C d Ac p0 m in RT 0 1
(3)
While for subsonic flow,
pT 2 1 p0 1 1
C d Ac p 0 pT 2 pT m in 1 R T 0 p0 1 p0
1
1
2
(4)
where pT , p0 , T 0 and ?? indicate downstream static pressure, upstream stagnation pressure, upstream stagnation temperature and ratio of specific heats, respectively. When the flow of gas mixture toward the inside of the cylinder chamber, p 0 acts as the intake port pressure and pT acts as the cylinder pressure. In contrast, when the flow of gas mixture toward the outside of the cylinder chamber, p 0 acts as the cylinder pressure and pT acts as the exhaust port pressure.
The mass in the system which enters through intake valve and leaves the system through exhaust valve obeys below equation [17]
dm m j dt j
(5)
where j indicates the total number of flows enter the system where m j is considered as positive or exit from the system and ?? indicates the total mass of air/fuel mixture present inside the system. Conservation of species is employed to look for the evolution of species within the combustion chamber which is the result of chemical reactions. The rate of change of the mass fraction of species ?? is given by below equation [17]
d Y i Y i Y iin i dt V
(6)
Where Y iin is the inlet mass fraction and i is the mass reaction rate of the species ??. The temperature change in the single-zone model is given by below equation. dT 1 pv R i dm 1 d Q h pdV d hi Y i C B m j h j dt C A i R dt dt m dt m dt j
(7)
Where Q h is the heat transfer and h j is the enthalpy of flows entering or leaving the system as well as C A and C B are defined as v (8) CAcp T (9) C B h v Once the temperature is obtained, the in-cylinder pressure is determined by using the ideal gas law p
Ru T
(10)
W mv
Where Ru is the universal gas constant and W mv is the mean molecular mass of the mixture. The convective heat transfer rate of HCCI engines can be described by Newton’s law of cooling [18], as
d Qh hc Aw T T w dt
(11)
Where hc is the heat transfer coefficient, Aw is the wall area and T w is the wall temperature. The wall area is the sum of the piston, cylinder head and cylinder wall area, which is
Aw
B2 B2
4 BL V c R 1 cos R2 sin 2 4 B 2 4
(12)
In this study, the heat transfer coefficient is modelled using the Woschni correlation. This correlation was calibrated in a gasoline engine which was run in HCCI mode. The Woschni heat transfer coefficient uses bore, ??, as the characteristic length, mean piston speed, S p as the characteristic velocity, ?? as instantaneous in-cylinder temperature and ?? as the instantaneous incylinder pressure which can be expressed as below equation where SI units are used for all variables (?? in m, ?? in K, V in m3 and S p in m/s) except for ?? which is in bars.
h c 129.8 B 0.2 p
0.8
T 0.55 2.228S p
0.8
(13)
With the help of cylinder pressure and volume as well as convective heat transfer rate of HCCI engines, heat release rate can be calculated which is expressed by below equation. dQ dV 1 dp d Q h p V d 1 d 1 d dt
(14)
dQ is the heat release dependent on the variation of crank angle d and is the ratio of specific heat values. Indicated mean effective pressure was calculated by below equation.
imep
W net Vd
(15)
Where V d is the cylinder swept volume and W net is the net work done which can be calculated by the following equation.
W net pdV
(16)
Where p is the cylinder pressure and dV is the variation of cylinder volume. Indicated torque and indicated power were calculated by below equations.
W net 4
(17)
2N 60
(18)
ip
Where is the indicated torque and N is the engine speed in rpm. 2.4 Operating Conditions In the present study, the simulation models were validated by the experimental results that were conducted on a four stroke single cylinder diesel engine which was modified for HCCI operation at a fixed engine speed of 900 rpm under the full engine load condition for n-heptane fuel [11]. The detail specifications of the engines are shown in Table 1. The schematic diagram of the engine setup is shown in Figure 2. Table 1: Engine model specifications [11]. Engine parameters Type Bore (mm) Stroke (mm) Displacement (L) No. of cylinders Compression ratio Connecting rod length (mm) Combustion chamber Intake Valve Close (°CA) Exhaust Valve Open (°CA) Fuel system
Value 4 stroke, single cylinder 82.55 114.3 0.6117 1 4.6-16 254 Pancake shape -144 140 Air-assist port fuel injection
The engine was in standard ASTM D2699 setting. Then it was modified for HCCI combustion and to control various engine parameters including air-fuel ratio (AFR), intake air temperature and pressure and exhaust gas recirculation. Necessary hardware and software were assembled for the modification. An air-assist port fuel injection was installed in the intake manifold to provide privilege to the liquid fuels for perfect atomization. In order to reduce the effect of pressure pulsation on intake and exhaust gasses, surge tanks were installed both in intake and exhaust systems. Data acquisition system was installed to control all the hardware and software of the engine system. As the engine specification is favorable for low speed, the data acquisition is also based on low-speed. It helps to control engine speed and load as well as maintain a stable speed and load for HCCI combustion. It also controls various engine parameters like as engine coolant and lubricating oil temperatures, intake air pressure and temperature, exhaust back pressure, fuel injection timing, and the quantity of fuel injected. In addition, for the more proper controlling system an eddy-current dynamometer was used to control the engine speed and loading. Besides, AC electric drive motor was coupled to a diesel engine for engine motoring. The engine was first run in diesel mode and then switched to HCCI once the condition was ready. Upon the completion of model validations, simulations were run to investigate the effects of various engine parameters on HCCI combustion and performance. Details of the simulation matrix are provided in Table 2. The reason of selecting this range of different engine parameters is that the value lower than this range cannot achieve the test fuels’ auto-ignition temperature which ultimately is the reason for failing to achieve HCCI combustion or creates misfiring. Similarly,
the value higher than this range creates higher in-cylinder pressure and temperature which ultimately causes knocking.
Figure 2: Schematic diagram of the HCCI engine set up, reproduced from [11]. Table 2: Simulation matrix. Simulation no. 1 2 3 4 5
Air-fuel ratio 35-55 50 50 50 50
Intake air temperature (K) 333 313-393 333 333 333
Intake air pressure (kPa) 100 100 100-200 100 100
Compression ratio 10 10 10 10-16 10
Speed (rpm) 600 900 900 900 600-1200
3. Results and Discussion 3.1 Model Validations To further validate the applicability of the developed reaction mechanism for engine simulations, numerical simulations were performed and compared against the HCCI engine experiments. As shown in Figure 3, the numerical simulation was able to capture elements of HCCI combustion of n-heptane, particularly the LTR. Figure 3(a) plots the in-cylinder pressure and Figure 3(b)
plots the heat release rate obtained from experiments done by [11] and our simulations, at a fixed engine speed of 900 rpm under constant intake temperature of 333 K and AFR of 50 conditions. As can be seen, the peak cylinder pressure is adequately reproduced, indicating that the important reaction pathways are very well represented. In comparison, the MCS predicted by the numerical simulation is advanced compared to the experimental result as shown in Figure 3(b). A spike in the predicted heat release rate was observed during the MCS. This spike was caused by the rapid oxidation of all CO accumulated to this point in the simulation to CO2. The conversion of CO to CO2 is predicted to happen rapidly once the required temperature is reached because only a few reactions are involved. Rapid oxidation of CO to CO2 has not been observed. The assumption of uniform mixture temperature and mixture composition contributes to this overly-rapid heat release rate. (a) In-cylinder Pressure 60
225 Experiment (Guo et al., 2010) The Present Study
200
50
Experiment (Guo et al., 2010) The Present Study
Heat Release Rate (J\0CA)
175
40
Pressure (bar)
(b) Heat Release Rate
30
20
10
150 125 100 75 50 25 0
0 -100 -75
-50
-25
0
25
Crank Angle
50
75
100
-25 -50 -40 -30 -20 -10
0
10 20 30 40 50
Crank Angle
Figure 3: Comparison between zero-dimensional single zone model with experimental results from [11]. AFR=50, Tin=333 K, Pin=100 kPa, CR=10, N=900 rpm. 3.2 Influence of Air-Fuel Ratio Figure 4 represents the effect of AFR on HCCI combustion and performance for a constant intake air temperature and pressure of 333K and 100 kPa conditions respectively. AFR was varied from 35 to 55. It can be seen from the Figures that stable HCCI combustion was achieved for a wide range of AFR and it is also proved that HCCI combustion is possible to happen in a very lean mixture. As shown in Figure 4(a) and 4 (b), in-cylinder pressure and heat release rate were decreased with the increase of AFR. The combustion was also delayed with the increase of AFR. It is also seen that the phasing of LTR was advanced and phasing of MCS was delayed with increasing AFR. The reason behind these phenomena is that the meeting of hydrocarbon and oxygen molecules during combustion worsens at leaner mixtures resulting in delayed autoignition. The LTR heat release profiles were found to be quite similar, although they were
advanced as AFR increased. However, increasing the AFR significantly decreases the heat release rate of the MCS. The decreasingly delayed phasing of the MCS contributes to the decreased heat release rate [19]. To investigate the influence of AFR on the performance of HCCI engines, five typical AFR were selected for the simulation. As represented in Figure 4(d), 4(e) and 4(f), IMEP, IT and IP decrease with increasing AFR. The reason behind these phenomena is that less amount of air and fuel are derived into the engine when higher AFR is used. It reduces the rate of combustion which ultimately provides lower IMEP, IT and IP [20]. One of the major limitations of HCCI combustion is that the unavailability of a highly diluted mixture to reduce the speed of the chemical reactions which may create excessive pressure. It ultimately damages the engine. By using a higher amount of AFR, this problem can be solved. Under the specific operating condition, the maximum IMEP, IT, and IP occur when AFR is 35. (a) In-cylinder Pressure
(b) Heat Release Rate
AFR=35 AFR=40 AFR=45 AFR=50 AFR=55
50
110
AFR=35 AFR=40 AFR=45 AFR=50 AFR=55
300
Heat Release Rate (J\0CA)
60
Pressure (bar)
(c) Cumulative Heat Release
350
40
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20
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100 90
Cumulative Heat Release (%)
70
200
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0
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-30
60 50 40 30 20
-20
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40
(e) Indicated Torque
(d) Indicated Mean Effective Pressure
35
3.5
6
30
3.0
5
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2
1
Indicated Power (kW)
7
Indicated Torque (N-m)
4.0
25
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0 50
-10
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0 45
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(e) Indicated Power
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8
35
AFR=35 AFR=40 AFR=45 AFR=50 AFR=55
10
Crank Angle
Crank Angle
Indicated Mean Effective Pressure (bar)
70
0
0 0 -40
80
0.0 35
Air-Fuel Ratio
40
45
50
Air-Fuel Ratio
55
35
40
45
50
55
Air-Fuel Ratio
Figure 4: Influence of air-fuel ratio on combustion and performance characteristics in HCCI engine. Tin=333 K, Pin=100 kPa, CR=10, N=900 rpm. 3.3 Influence of Intake Air Temperature Intake air temperature is the most critical and widely used engine operation parameter to control the phasing of HCCI combustion [21]. Figure 5 illustrates the effect of intake temperature on HCCI combustion and performance for a constant AFR of 50. Intake temperature of the air was
varied from 313 K to 393 K. Stable HCCI combustion was obtained for a wide range of temperatures. As shown in Figure 5(a) the auto-ignition timing was advanced with the increase of intake air temperature. The phasing of both LTR and MCS were advanced. Furthermore, it can also be clearly seen from Figure 5(b) that higher pressure rise rate creates rapid heat release rate. Due to this phenomenon knocking takes place which deteriorates the HCCI combustion. For this reason, cylinder pressure decreases with increasing intake air temperature. The LTR heat release profiles were found to be quite similar, although they were advanced as temperature increased. However, increasing the intake temperature significantly enhances the heat release rate of the MCS. The increasingly advanced phasing of the MCS contributes to the increased heat release rate [22].
(a) In-cylinder Pressure
(b) Heat Release Rate
313 K 333 K 353 K 373 K 393 K
110
313 K 333 K 353 K 373 K 393 K
200
Heat Release Rate (J\0CA)
50
40
Pressure (bar)
(c) Cumulative Heat Release
250
30
20
100 90
Cumulative Heat Release (%)
60
150
100
50
10
0
0
-50
80 70 60 50 40 30 20
313 K 333 K 353 K 373 K 393 K
10 0
-40
-30
-20
-10
0
10
20
30
40
-10 -40
-30
Crank Angle
-10
0
10
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(e) Indicated Torque
25
2.5
2
1
Indicated Power (kW)
5
Indicated Torque (N-m)
3.0
3
20
15
10
5
0 353
373
393
Intake Air Temperature (K)
0
10
20
30
40
2.0
1.5
1.0
0.5
0 333
-10
(f) Indicated Power
30
4
-20
Crank Angle
6
313
-30
Crank Angle
(d) Indicated Mean Effective Pressure
Indicated Mean Effective Pressure (bar)
-20
0.0 313
333
353
373
393
Intake Air Temperature (K)
313
333
353
373
393
Intake Air Temperature (K)
Figure 5: Influence of intake air temperature on combustion and performance characteristics in HCCI engine. AFR=50, Pin=100 kPa, CR=10, N=900 rpm. To investigate the influence of intake air temperature on the performance of HCCI engines, five typical temperatures were selected for the simulation. As is shown in Figure 5(d), 5(e) and 5(f), IMEP, IT and IP decrease with increasing intake temperature. This is because there are key chemical reactions in the low-temperature phase. These reactions depend on the system temperature and AFR. C7H15O2=C7H14OOH is the most important reaction in low-temperature
phase. C7H14OOH retards the process of low-temperature oxidation, and it plays the most important role in the occurrence of the negative temperature coefficient (NTC) phase [23, 24]. Under the specific operating condition, the maximum IMEP, IT, and IP occur when intake temperature is 313 K. Obviously, the intake temperature is very critical, i.e., a small change of intake temperature leads to large variations in performance of HCCI. 3.4 Influence of Intake Air Pressure Intake air pressure boosting has a significant effect on the combustion and performance of HCCI engines as well as is a common way to improve power output. In this study, the effect of boosting on HCCI combustion and performance was examined at a constant AFR and intake air temperature condition. As shown in Figure 6(a), boosting the intake pressure tends to significantly increase the peak cylinder pressure during the compression stroke. The reason behind this phenomenon is that with increasing intake air pressure the collision frequency among molecules increases. For this reason, combustion reaction velocity of HCCI combustion also increases [25]. Since the quantity of fuel injected is increased as intake pressure increases to maintain a constant AFR, the LTR stage is advanced and intensified. This leads to a shorter NTC delay period and significantly advances phasing of the MCS, as shown in Figure 6(b). The reason for this is the combustion reaction velocity which increases with the increasing boost pressure [26, 27]. To investigate the influence of intake air pressure on the performance of HCCI engines, five typical pressures ranging from 100 kPa to 200kPa were selected for the simulation. As is shown in Figure 6(d), 6(e) and 6(f), IMEP, IT and IP increases with increasing intake air pressure. The maximum IMEP, IT, and IP were obtained at 200 kPa. This is because higher intake air pressure brings a higher amount of air into the cylinder and to maintain constant AFR higher amount of fuel is also injected into the cylinder. For this reason, a higher amount of IMEP, IT and IP are achieved at higher intake air pressure. On the other hand, the engine can operate stably at a leaner AFR with an increasing inlet pressure, which leads to lower values of IMEP, IT, and IP. However, if proper tuning is carried out to control combustion phasing using different intake air temperature, EGR or compression ratio, more benefits can be achieved from boosting intake air pressure. The results indicate that an increase in the boost pressure causes the need of leaner mixture, and requires more advanced injection timing to achieve the maximum engine torque [28, 29].
450
90
100 kPa 125 kPa 150 kPa 175 kPa 200 kPa
400 350
Heat Release Rate (J\0CA)
80 70
Pressure (bar)
110
60 50 40 30 20
100 90
Cumulataive Heat Release (%)
100 kPa 125 kPa 150 kPa 175 kPa 200 kPa
100
300 250 200 150 100 50
80 70 60 50 40 30
10
0
0
0
-50
-10
-30
-20
-10
0
10
20
30
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-30
Crank Angle
-20
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100 kPa 125 kPa 150 kPa 175 kPa 200 kPa
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(e) Indicated Torque
(d) Indicated Mean Effective Pressure 12.5
(f) Indicated Power
60
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7.5
5.0
Indicated Power (kW)
10.0
Indicated Torque (N-m)
Indicated Mean Effective Pressure (bar)
(c) Cumulative Heat Release
(b) Heat Release Rate
(a) In-cylinder Pressure 110
40
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4
3
2
2.5 10
0.0
1
0 100
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Intake Air Pressure (kPa)
0 100
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Intake Air Pressure (kPa)
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Figure 6: Influence of intake air pressure on combustion and performance characteristics in HCCI engine. AFR=50, Tin=333 K, CR=10, N=900 rpm. 3.5 Influence of Compression Ratio Compression ratio is another important engine parameter which also has a significant effect on combustion and performance characteristics of HCCI engines [30]. Figure 7 shows that increasing the compression ratio advances the combustion process and increases the peak cylinder pressures. This is primarily due to the effect of increased compression temperatures and pressures as the compression ratio increases, which enhances n-heptane oxidation [31]. In the case of the higher compression ratios, the temperature around TDC is relatively high. This is conductive to initiation of the chemical reactions. As a result, a large amount of energy is released which ultimately creates an increased pressure and temperature [32]. As shown in Figure 7(b), the phasing of both LTR and MCS phases were advanced with increasing compression ratio. To investigate the influence of compression ratio on the performance of HCCI engines, five typical compression ratios were selected for the simulation. As is shown in Figure 7(d), 7(e) and 7(f), IMEP, IT and IP increases with increasing compression ratio. This is because there are key chemical reactions which occur during the compression stroke. These reactions depend on the in-cylinder pressure and temperature. Due to increasing compression ratio oxidation of n-
heptane is enhanced which ultimately help to increase IMEP, IT and IP [24, 33]. Under the specific operating condition, the maximum IMEP, IT, and IP occur when the compression ratio is 16. It is obvious that the compression ratio is very critical, i.e., a small change of compression ratio leads to large variations in performance of HCCI.
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Figure 7: Influence of Compression Ratio on combustion and performance characteristics in HCCI engine. AFR=50, Tin=333 K, Pin=100 kPa, N=900 rpm. 3.6 Influence of Speed Engine speed is an important parameter which has a significant effect on combustion and performance of HCCI engines. Figure 8 illustrates the effect of engine speed on HCCI combustion and performance for a constant intake temperature of 333 K. Engine speed was varied from 600 rpm to 1200 rpm. From the Figure 8(a) it is clearly seen that in-cylinder pressure increases with increasing engine speed up to 900 rpm and then decreases with further increasing engine speed. The most important reason is heat loss through cylinder wall during the compression stroke. With the engine speed increasing, the accumulated heat loss during the compression stroke reduced significantly because of the shorter cyclic period, which increases the in-cylinder temperature at the end of compression stroke. Therefore, the chamber could
achieve a higher maximum pressure at higher engine speed due to less heat loss through the cylinder wall [34]. In addition, peak pressure location is retarded with the increase of engine speed. This is due to a significantly retarded MCS, as shown in Figure 8(b). The heat release rate during the MCS is also reduced with retarded combustion due to the increasing combustion chamber volume after top dead center. This leads to lower combustion chamber temperatures and a corresponding decrease in the oxidation reaction rates [35]. However, the effect of engine speed on the LTR phase tends to be relatively weak, reflecting its strong dependence on temperature history [23].
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Figure 8: Influence of speed on combustion and performance characteristics in HCCI engine. AFR=50, Tin=333 K, Pin=100 kPa, CR=10. To observe the influence of engine speed on the performance of HCCI engines, five typical speeds were selected for the simulation. As is shown in Figure 8(d), 8(e) and 8(f), IMEP, IT and IP increases with increasing engine speed up to 900 rpm and then decreases with further increasing engine speed. Maximum IMEP, IT, and IP were obtained as 5.69 bar, 27.72 Nm and 2.61 kW. At higher engine speed the performance was less. This is because of that, with the increase of engine speed, the combustion duration is shortened with respect to time and affects the performance of HCCI engines [36]. Moreover, volumetric efficiency depends on engine
speed. Because of the changing volumetric efficiency and increased friction forces at higher engine speed, the IMEP, IT and IP were less[37]. 4. Conclusions A zero-dimensional single zone model was established to investigate HCCI combustion and performance characteristics over a wide range of operating conditions. A reduced n-heptane chemical reaction mechanism was used to solve the chemical reactions during combustion. The validated numerical simulation was able to capture trends in combustion phasing variation with critical engine parameters, particularly the LTR. From the simulation, it was seen that HCCI combustion was achieved easily for very lean mixture. Decreasing the AFR advances the phasing of the MCS and IMEP, IT and IP decreases with AFR. The phasing of both the LTR and MCS were advanced with increasing intake air temperature and compression ratio. However, opposite results were found in terms of performance while increasing intake air temperature and compression ratio. That means, performance decreases with increasing intake air temperature and increases with increasing compression ratio. In addition, it was obtained from the simulation that the effects of turbocharging on the phasing of LTR are relatively weak. Moreover, increasing the engine speed peak pressure location is retarded due to retarded MCS. IMEP, IT, and IP were increased up to a certain limit then decreased with increasing speed. This is because of with the increase in engine speed; the combustion duration is shortened with respect to time. Acknowledgements The authors would like to thank Universiti Malaysia Pahang for financial support under project no. RDU1503101 and providing laboratory facilities. References [1] Mohanamurugan S, Sendilvelan S. Emission and combustion characteristics of different fuel In A HCCI engine. International Journal of Automotive and Mechanical Engineering. 2011;3:279-92. [2] Hairuddin A, Wandel A, Yusaf T. An introduction to a homogeneous charge compression ignition engine. Journal of Mechanical Engineering and Sciences. 2014;7:1042-52. [3] Hasan MM, Rahman MM, Kadirgama K. A review on homogeneous charge compression ignition engine performance using biodiesel-diesel blend as a fuel. International Journal of Automotive & Mechanical Engineering. 2015;11:2199-211. [4] Mohanamurugan S, Sendilvelan. Emission and combustion characteristics of different fuel in HCCI engine. International Journal of Automotive and Mechanical Engineering. 2011;3:279-92. [5] Polovina D, McKenna D, Wheeler J, Sterniak J, Miersch-Wiemers O, Mond A, et al. Steadystate combustion development of a downsized multi-cylinder engine with range extended HCCI/SACI capability. SAE Technical Paper; 2013. [6] Izadi Najafabadi M, Abdul Aziz N. Homogeneous charge compression ignition combustion: challenges and proposed solutions. Journal of combustion. 2013;2013. [7] Yao M, Zheng Z, Liu H. Progress and recent trends in homogeneous charge compression ignition (HCCI) engines. Progress in Energy and Combustion Science. 2009;35:398-437.
[8] Saxena S, Bedoya ID. Fundamental phenomena affecting low temperature combustion and HCCI engines, high load limits and strategies for extending these limits. Progress in Energy and Combustion Science. 2013;39:457-88. [9] Mangus M, Kiani F, Mattson J, Depcik C, Peltier E, Stagg-Williams S. Comparison of neat biodiesels and ULSD in an optimized single-cylinder diesel engine with electronically-controlled fuel injection. Energy & Fuels. 2014;28:3849-62. [10] Komninos N, Rakopoulos C. Modeling HCCI combustion of biofuels: A review. Renewable and Sustainable Energy Reviews. 2012;16:1588-610. [11] Guo H, Neill WS, Chippior W, Li H, Taylor JD. An experimental and modeling study of HCCI combustion using n-heptane. Journal of Engineering for Gas Turbines and Power. 2010;132:022801. [12] Assanis D, Polishak M. Valve event optimization in a spark-ignition engine. Journal of Engineering for Gas Turbines and Power. 1990;112:341-7. [13] Seiser R, Pitsch H, Seshadri K, Pitz W, Gurran H. Extinction and autoignition of n-heptane in counterflow configuration. Proceedings of the Combustion Institute. 2000;28:2029-37. [14] Goodwin D. An open-source, extensible software suite for CVD process simulation. Chemical Vapor Deposition XVI and EUROCVD. 2003;14:2003-08. [15] Andrae J. A kinetic modeling study of self-ignition of low alkylbenzenes at engine-relevant conditions. Fuel Processing Technology. 2011;92:2030-40. [16] Heywood JB. Internal combustion engine fundamentals: Mcgraw-hill New York; 1988. [17] Kuo KK. Principles of combustion. 1986. [18] Stiesch G. Modeling engine spray and combustion processes: Springer Science & Business Media; 2013. [19] Polat S. An experimental study on combustion, engine performance and exhaust emissions in a HCCI engine fuelled with diethyl ether–ethanol fuel blends. Fuel Processing Technology. 2016;143:140-50. [20] Maurya RK, Agarwal AK. Experimental investigation on the effect of intake air temperature and air–fuel ratio on cycle-to-cycle variations of HCCI combustion and performance parameters. Applied Energy. 2011;88:1153-63. [21] Maurya RK, Agarwal AK. Experimental study of combustion and emission characteristics of ethanol fuelled port injected homogeneous charge compression ignition (HCCI) combustion engine. Applied Energy. 2011;88:1169-80. [22] Uyumaz A. An experimental investigation into combustion and performance characteristics of an HCCI gasoline engine fueled with n-heptane, isopropanol and n-butanol fuel blends at different inlet air temperatures. Energy Conversion and Management. 2015;98:199-207. [23] Zhang CH, Xue L, Wang J. Experimental study of the influence of λ and intake temperature on combustion characteristics in an HCCI engine fueled with n-heptane. Journal of the Energy Institute. 2014;87:175-82. [24] Peng Z, Zhao H, Ma T, Ladommatos N. Characteristics of homogeneous charge compression ignition (HCCI) combustion and emissions of n-heptane. Combustion science and technology. 2005;177:2113-50. [25] Liu H, Yao M, Zhang B, Zheng Z. Effects of inlet pressure and octane numbers on combustion and emissions of a homogeneous charge compression ignition (HCCI) engine. Energy & Fuels. 2008;22:2207-15. [26] Christensen M, Johansson B. Supercharged homogeneous charge compression ignition (HCCI) with exhaust gas recirculation and pilot fuel. SAE Technical Paper; 2000.
[27] Olsson J-O, Tunestål P, Johansson B. Boosting for high load HCCI. SAE Technical Paper; 2004. [28] Hosseini V, Checkel MD. Intake pressure effects on HCCI combustion in a CFR engine. Proceedings of the combustion institute, canadian section, spring technical meeting2007. [29] Olsson J-O, Tunestål P, Haraldsson G, Johansson B. A turbocharged dual-fuel HCCI engine. SAE Special Publications. 2001;2001. [30] Olsson J-O, Tunestål P, Johansson B, Fiveland S, Agama R, Willi M, et al. Compression ratio influence on maximum load of a natural gas fueled HCCI engine. SAE Technical Paper; 2002. [31] Li H, Guo H, Neill WS, Chippior W, Taylor JD. An experimental and modeling study of HCCl combustion using n-heptane. Proceedings of the ICEF2006. 2006:1-11. [32] Iida M, Hayashi M, Foster D, Martin J. Characteristics of homogeneous charge compression ignition (HCCI) engine operation for variations in compression ratio, speed, and intake temperature while using n-butane as a fuel. Journal of engineering for gas turbines and power. 2003;125:472-8. [33] Liu M-B, He B-Q, Zhao H. Effect of air dilution and effective compression ratio on the combustion characteristics of a HCCI (homogeneous charge compression ignition) engine fuelled with n-butanol. Energy. 2015;85:296-303. [34] Zhang C, Wu H. Combustion characteristics and performance of a methanol fueled homogenous charge compression ignition (HCCI) engine. Journal of the Energy Institute. 2015. [35] García MT, Aguilar FJJ-E, Lencero TS. Experimental study of the performances of a modified diesel engine operating in homogeneous charge compression ignition (HCCI) combustion mode versus the original diesel combustion mode. Energy. 2009;34:159-71. [36] Canakci M. An experimental study for the effects of boost pressure on the performance and exhaust emissions of a DI-HCCI gasoline engine. Fuel. 2008;87:1503-14. [37] Nishi M, Kanehara M, Iida N. Assessment for innovative combustion on HCCI engine by controlling EGR ratio and engine speed. Applied Thermal Engineering. 2016;99:42-60. List of Figures Figure 1: An algorithm flow chart for zero-dimensional single-zone model simulation. Figure 2: Schematic diagram of the HCCI engine set up, reproduced from [11]. Figure 3: Comparison between zero-dimensional single zone model with experimental results from [11]. AFR=50, Tin=333 K, Pin=100 kPa, CR=10, N=900 rpm. Figure 4: Influence of air-fuel ratio on combustion and performance characteristics in HCCI engine. Tin=333 K, Pin=100 kPa, CR=10, N=900 rpm. Figure 5: Influence of intake air temperature on combustion and performance characteristics in HCCI engine. AFR=50, Pin=100 kPa, CR=10, N=900 rpm. Figure 6: Influence of intake air pressure on combustion and performance characteristics in HCCI engine. AFR=50, Tin=333 K, CR=10, N=900 rpm.
Figure 7: Influence of Compression Ratio on combustion and performance characteristics in HCCI engine. AFR=50, Tin=333 K, Pin=100 kPa, N=900 rpm. Figure 8: Influence of speed on combustion and performance characteristics in HCCI engine. AFR=50, Tin=333 K, Pin=100 kPa, CR=10.
List of Tables Table 1: Engine model specifications [11]. Table 2: Simulation matrix.
HIGHLIGHTS OF MANUSCRIPT
HCCI combustion was achieved easily for very lean mixture. The phasing of combustion was advanced with increasing intake air temperature. Turbocharging effect on the phasing of low temperature reaction is relatively weak. Better performance was found with increasing compression ratio. Peak pressure location was retarded with increasing engine speed.
S1359-4311(16)32030-0 https://doi.org/10.1016/j.applthermaleng.2017.09.121 ATE 11178
To appear in:
Applied Thermal Engineering
Received Date: Revised Date: Accepted Date:
29 September 2016 11 April 2017 23 September 2017
Please cite this article as: M.M. Hasan, M.M. Rahman, K. Kadirgama, D. Ramasamy, Numerical study of engine parameters on combustion and performance characteristics in an n-heptane fueled HCCI engine, Applied Thermal Engineering (2017), doi: https://doi.org/10.1016/j.applthermaleng.2017.09.121
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Numerical study of engine parameters on combustion and performance characteristics in an n-heptane fueled HCCI engine M.M. Hasan1, M.M. Rahman1,2, K. Kadirgama1,2 and D. Ramasamy1,2 1
Automotive Engineering Research Group Faculty of Mechanical Engineering, Universiti Malaysia Pahang 26600 Pekan, Pahang, Malaysia Email: [email protected] Phone: +60149670211 2 Automotive Engineering Centre, Universiti Malaysia Pahang 26600 Pekan, Pahang, Malaysia Abstract Homogeneous charge compression ignition (HCCI) is an alternative combustion concept which offers significant benefits in terms of its high efficiency. However, the operational range using HCCI combustion in terms of speed and load is restricted due to the absence of the direct control of the onset of ignition and the heat release rate. In this work, a zero-dimensional single-zone numerical simulation with reduced fuel chemistry was developed and used to investigate the effect of various engine parameters on combustion and performance characteristics in an HCCI engine fueled with n-heptane. The simulations show good agreement while comparing the results with the published experimental results and capture important combustion phase trends as engine parameters are varied with a minimum percentage of error which is less than 6%. The combustion phase was advanced and the combustion duration was shortened with the increase of intake air temperature and the decrease of the engine speed. The maximum load was successfully increased with increasing the intake air pressure. The highest load in this work was 11.27 bar in IMEPg at the condition of 200 kPa in intake air pressure and 333 K in intake air temperature. Keywords: HCCI; single zone; n-heptane; engine parameters; combustion; performance. 1. Introduction With increasing concern about fuel economy and emissions, the internal combustion engine industry is constantly looking for better alternatives to spark ignition (SI) and compression ignition (CI) engines. Homogeneous charge compression ignition (HCCI) engine is one of the alternatives under extensive research in recent years [1, 2]. The name demonstrates its two essential characteristics, homogeneous mixture, and compression ignition. The HCCI concept promises several advantages. In short, it is more efficient than the SI engine and cleaner than the CI engine. Compared with the SI engine, the higher compression ratio can be used and leaner fuel-air mixture can be applied on an HCCI engine [3, 4]. At the same time, compared with the CI engine, the cylinder mixture is more evenly distributed in an HCCI engine, where fuel rich pocket is not possible. Without the soot-inducing fuel-rich pockets and NOX-inducing lean regions, the overall result is that the HCCI engine can achieve higher fuel efficiency with lower NOX and soot emissions [5].
Nomenclature a Ac Aw B Cd dQ dV hc hj i j l L m N p p0 pT Qh R Rc Ru s
T0 Tw U V Vc Vd p0
crank radius effective valve open area wall area bore discharge co-efficient heat release dependent on the variation of crank angle d variation of cylinder volume heat transfer coefficient enthalpy of flows entering or leaving the system mass fraction of species number of flows in or out of the system connecting rod length stroke length total mass in the system engine speed in-cylinder pressure upstream stagnation pressure downstream static pressure heat transfer ratio of connecting rod length compression ratio universal gas constant distance between crank axis and piston pin axis upstream stagnation temperature wall temperature internal energy instantaneous cylinder volume clearance volume displacement volume upstream stagnation pressure
W W net Y iin m in Sp
W mv Vd W
work net work done inlet mass fraction mass flow rate instantaneous piston speed mean molecular weight of the mixture displacement volume work
Greek symbols ratio of specific heats ?? indicated torque θ crank angle Ø equivalence ratio ω rotational speed mass reaction rate of the species i i Abbreviations AFR air-fuel ratio CFD computational fluid dynamics CI compression ignition EVO exhaust valve open HCCI homogeneous charge compression ignition IVC inlet valve close LTR low-temperature reaction MCS main combustion stage NTC negative temperature coefficient SI spark ignition SOC start of combustion TDC top dead center
However, differing from both CI and SI engines, HCCI engines lack a physical event which controls the ignition timing. It is the main challenge which ultimately influences the power and efficiency [6]. Unlike SI engines and CI engines, HCCI engines have no direct mechanism to control the SOC. Its SOC is fully dependent on auto-ignition. According to Yao, Zheng [7], this auto-ignition is affected by several factors like fuel auto-ignition chemistry and thermodynamic properties, combustion duration, wall temperatures, intake temperature and
pressure, compression ratio, amount of EGR, engine speed, engine temperature, convective heat transfer to the engine, and other engine parameters. Higher intake temperatures and intake pressures cause earlier combustion timing due to the faster chemical kinetics [8]. For equivalence ratio sensitive fuels, higher equivalence ratios (up to stoichiometric) have a tendency to propel ignition timing due to the improvement of charge reactivity. The fuels without equivalence ratio sensitivity demonstrate an inverse pattern, where higher equivalence ratios have a tendency to bring about more delayed combustion timing. The lower specific heat ratio from higher equivalence ratios causes the prerequisite of more compression heating to achieve ignition temperatures. Nonetheless, in real operation, for both sorts of fuels the utilization of higher equivalence ratios likewise increases the in-cylinder wall and residuals temperatures, along these lines bringing on advanced combustion timing [9]. For fuels with little to no equivalence ratio sensitivity, there are contending impacts where lower specific heat ratios have a tendency to delay combustion while higher in-cylinder wall and residual temperatures have a tendency to propel combustion. To understand the effect of engine parameters on combustion and performance characteristics of HCCI engine, a fruitful study is needed which can be done by developing a more accurate simulation model based on zero-dimension and single zone rather than experiments or multidimensional modeling. Experimental work is costly and requires more time. On the other hand, modeling is a way of exploring engine behavior in ways that might otherwise impossible in a true experimental approach. In addition, a multi-dimensional model requires substantial computational time as compared to zero-dimensional and quasi-dimensional models, which are a simplified version of a multi-dimensional model [10]. Thus, the use of simplified model becomes increasingly important because of its advantages in computational time and resources. A zero-dimensional simulation would be an interim solution until the cost and time of running a multi-dimensional model is comparable with the current cost of the zerodimensional model. The use of chemical kinetics mechanisms also helps in investigating the combustion behavior of an HCCI engine. The purpose of this study is to establish a numerical model based on single zone and zero-dimension for HCCI combustion over a wide range of operating conditions using n-heptane. The HCCI engines were modeled to investigate the effects of a number of engine parameters such as air-fuel ratio, intake air temperature, intake air pressure, compression ratio and engine speed on the combustion phasing of the low-temperature reaction (LTR) and main combustion stage (MCS). A reduced fuel chemistry for n-heptane was developed and compared to the combustion data obtained in [11]. The numerical simulation was able to capture the combustion phasing trends observed during the experiments with changes in several engine parameters. Based on the numerical results, it is believed that the model used in this study is suitable for investigating fuel chemistry effects on HCCI combustion. 2. Numerical Methodology 2.1 Numerical Model In order to design an engine for various fuels, accurate models are needed which are able to model both combustion and performance. In this study, a zero-dimensional single zone model for a single cylinder, four stroke HCCI engine was established using MATLAB. The proposed method takes into account the detailed thermodynamic aspects of the engine. The model integrates the ordinary differential equation system corresponding to the chemical and thermal evolution of a closed homogeneous system under an imposed volume history reproducing the
engine cycle. This single-zone in time model treats the cylindrical combustion chamber as a uniform reactor with uniform temperature, pressure, and composition throughout. The reactor volume changes based on the slider-crank relations which determine the motion of the piston in the engine cylinder. The simulation handles only the closed part of the cycle; intake and exhaust processes are not considered. This kind of model aims to simulate the auto-ignition process in the core of the air-fuel mixture and enables the use of detailed and reduced chemical kinetic models with which we can investigate the chemical reactions contributing to cylinder inside pressure and temperature. This study only considers the results obtained from combustion between inlet valve close (IVC) and exhaust valve open (EVO). The emissions produced after EVO are not of interest in this study. The effect of wall heat losses is neglected and the accumulated gas is assumed to be an ideal gas. Toward the start of the simulation, the engine parameters and beginning working condition were characterized, which depended on the experimental data. At that point, the mixture composition in the combustion chamber was introduced. It was accepted that the initial composition in the combustion chamber comprises a mixture of air and fuel. An ordinary valve profile was resolved to take into account the equation from [12], which was utilized to represent the valve movement furthermore to get the inlet mass flow rate. The entire simulation process is shown in Figure 1. The main part of the simulation is the place the main combustion happened, which was after IVC and before EVO. At this stage, the piston was in upward movement after IVC, compacting the gas mixture. The chemical kinetics plays an imperative part here as it will decide the start of combustion. Once the piston passes the top dead center (TDC) mark, it will be in descending movement once more, expanding the mixture and the simulation stops at EVO. The simulation solves the energy and species equations for the whole procedure. 2.2 Chemical Kinetic Mechanism In order to simulate the combustion in an HCCI engine, it is obvious to solve a series of chemical reactions. It can be done by developing a chemical reaction mechanism for a particular fuel. In this study, a reduced n-heptane mechanism was used which was developed by Seiser et al. [13]. The mechanism consists of 159 species and 770 elementary reactions. The mechanism was then utilized in the Cantera chemical kinetics software package [14] to get the chemical properties and reactions mechanisms. Cantera is an open source software package for chemical kinetics mechanism and is widely used [15]. The chemical kinetics files are available in Chemkin format. Cantera can convert this file to use it in MATLAB. The utilization of a chemical reaction mechanism empowers the investigation of the chemical species interaction, where the interaction is impacted by the temperature, pressure and species mass fraction. Besides, the chemical reaction mechanism would have the capacity to give a superior comprehension in the combustion study.
. Input 1. IVC (ATDC) 2. Pintake (Pa) 3. Tintake [K] 4. AFR 5. Bore (m) 6. Stroke (m) 7. Compression ratio 8. Connecting rod length (m) 9. RPM 10. Swirl ratio 11. Wall temperature [K]
θCA = θIVC Volume = f(θCA)
Start of combustion = f(T, φ) Burn duration (∆θ) = f(θSOC)
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θCA = θSOC
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Product properties (CANTERA)
θCA = θSOC + ∆θ T, P (Polytropic expansion) θCA = θEVO End of model T and P history during IVC→EVO
Figure 1: An algorithm flow chart for zero-dimensional single-zone model simulation. 2.3 Governing Equations One approach in engine modelling utilizes the geometry of the reciprocating engine to calculate the engine volume over the complete combustion cycle. The piston work in the energy equation
can be determined by the cylinder volume. The instantaneous cylinder volume at any crank angle location is calculated by below equation.
1 V V c 1 R c R 1 cos R 2 sin 2 2
(1)
Where V is for instantaneous volume, V c is the clearance volume, R is the ratio of connecting rod length to crank radius and Rc is the compression ratio. With the rate of change of volume
1 1 cos dV V c Rc sin 2 2 d R sin 2
(2)
The equation for compressible flow through a flow restriction is generally used to calculate the mass flow rate to the combustion chamber [16]. The equation uses the effect of the real gas flow and discharge coefficient C d which can be acquired from the experiments. A pressure difference is necessary between combustion chamber and ports to flow the mass in or out of the chamber. The equation is divided into two cases: choked and subsonic flows. For choked flow, the following conditions are followed:
pT 2 p0 1
1
1
2 2 1 C d Ac p0 m in RT 0 1
(3)
While for subsonic flow,
pT 2 1 p0 1 1
C d Ac p 0 pT 2 pT m in 1 R T 0 p0 1 p0
1
1
2
(4)
where pT , p0 , T 0 and ?? indicate downstream static pressure, upstream stagnation pressure, upstream stagnation temperature and ratio of specific heats, respectively. When the flow of gas mixture toward the inside of the cylinder chamber, p 0 acts as the intake port pressure and pT acts as the cylinder pressure. In contrast, when the flow of gas mixture toward the outside of the cylinder chamber, p 0 acts as the cylinder pressure and pT acts as the exhaust port pressure.
The mass in the system which enters through intake valve and leaves the system through exhaust valve obeys below equation [17]
dm m j dt j
(5)
where j indicates the total number of flows enter the system where m j is considered as positive or exit from the system and ?? indicates the total mass of air/fuel mixture present inside the system. Conservation of species is employed to look for the evolution of species within the combustion chamber which is the result of chemical reactions. The rate of change of the mass fraction of species ?? is given by below equation [17]
d Y i Y i Y iin i dt V
(6)
Where Y iin is the inlet mass fraction and i is the mass reaction rate of the species ??. The temperature change in the single-zone model is given by below equation. dT 1 pv R i dm 1 d Q h pdV d hi Y i C B m j h j dt C A i R dt dt m dt m dt j
(7)
Where Q h is the heat transfer and h j is the enthalpy of flows entering or leaving the system as well as C A and C B are defined as v (8) CAcp T (9) C B h v Once the temperature is obtained, the in-cylinder pressure is determined by using the ideal gas law p
Ru T
(10)
W mv
Where Ru is the universal gas constant and W mv is the mean molecular mass of the mixture. The convective heat transfer rate of HCCI engines can be described by Newton’s law of cooling [18], as
d Qh hc Aw T T w dt
(11)
Where hc is the heat transfer coefficient, Aw is the wall area and T w is the wall temperature. The wall area is the sum of the piston, cylinder head and cylinder wall area, which is
Aw
B2 B2
4 BL V c R 1 cos R2 sin 2 4 B 2 4
(12)
In this study, the heat transfer coefficient is modelled using the Woschni correlation. This correlation was calibrated in a gasoline engine which was run in HCCI mode. The Woschni heat transfer coefficient uses bore, ??, as the characteristic length, mean piston speed, S p as the characteristic velocity, ?? as instantaneous in-cylinder temperature and ?? as the instantaneous incylinder pressure which can be expressed as below equation where SI units are used for all variables (?? in m, ?? in K, V in m3 and S p in m/s) except for ?? which is in bars.
h c 129.8 B 0.2 p
0.8
T 0.55 2.228S p
0.8
(13)
With the help of cylinder pressure and volume as well as convective heat transfer rate of HCCI engines, heat release rate can be calculated which is expressed by below equation. dQ dV 1 dp d Q h p V d 1 d 1 d dt
(14)
dQ is the heat release dependent on the variation of crank angle d and is the ratio of specific heat values. Indicated mean effective pressure was calculated by below equation.
imep
W net Vd
(15)
Where V d is the cylinder swept volume and W net is the net work done which can be calculated by the following equation.
W net pdV
(16)
Where p is the cylinder pressure and dV is the variation of cylinder volume. Indicated torque and indicated power were calculated by below equations.
W net 4
(17)
2N 60
(18)
ip
Where is the indicated torque and N is the engine speed in rpm. 2.4 Operating Conditions In the present study, the simulation models were validated by the experimental results that were conducted on a four stroke single cylinder diesel engine which was modified for HCCI operation at a fixed engine speed of 900 rpm under the full engine load condition for n-heptane fuel [11]. The detail specifications of the engines are shown in Table 1. The schematic diagram of the engine setup is shown in Figure 2. Table 1: Engine model specifications [11]. Engine parameters Type Bore (mm) Stroke (mm) Displacement (L) No. of cylinders Compression ratio Connecting rod length (mm) Combustion chamber Intake Valve Close (°CA) Exhaust Valve Open (°CA) Fuel system
Value 4 stroke, single cylinder 82.55 114.3 0.6117 1 4.6-16 254 Pancake shape -144 140 Air-assist port fuel injection
The engine was in standard ASTM D2699 setting. Then it was modified for HCCI combustion and to control various engine parameters including air-fuel ratio (AFR), intake air temperature and pressure and exhaust gas recirculation. Necessary hardware and software were assembled for the modification. An air-assist port fuel injection was installed in the intake manifold to provide privilege to the liquid fuels for perfect atomization. In order to reduce the effect of pressure pulsation on intake and exhaust gasses, surge tanks were installed both in intake and exhaust systems. Data acquisition system was installed to control all the hardware and software of the engine system. As the engine specification is favorable for low speed, the data acquisition is also based on low-speed. It helps to control engine speed and load as well as maintain a stable speed and load for HCCI combustion. It also controls various engine parameters like as engine coolant and lubricating oil temperatures, intake air pressure and temperature, exhaust back pressure, fuel injection timing, and the quantity of fuel injected. In addition, for the more proper controlling system an eddy-current dynamometer was used to control the engine speed and loading. Besides, AC electric drive motor was coupled to a diesel engine for engine motoring. The engine was first run in diesel mode and then switched to HCCI once the condition was ready. Upon the completion of model validations, simulations were run to investigate the effects of various engine parameters on HCCI combustion and performance. Details of the simulation matrix are provided in Table 2. The reason of selecting this range of different engine parameters is that the value lower than this range cannot achieve the test fuels’ auto-ignition temperature which ultimately is the reason for failing to achieve HCCI combustion or creates misfiring. Similarly,
the value higher than this range creates higher in-cylinder pressure and temperature which ultimately causes knocking.
Figure 2: Schematic diagram of the HCCI engine set up, reproduced from [11]. Table 2: Simulation matrix. Simulation no. 1 2 3 4 5
Air-fuel ratio 35-55 50 50 50 50
Intake air temperature (K) 333 313-393 333 333 333
Intake air pressure (kPa) 100 100 100-200 100 100
Compression ratio 10 10 10 10-16 10
Speed (rpm) 600 900 900 900 600-1200
3. Results and Discussion 3.1 Model Validations To further validate the applicability of the developed reaction mechanism for engine simulations, numerical simulations were performed and compared against the HCCI engine experiments. As shown in Figure 3, the numerical simulation was able to capture elements of HCCI combustion of n-heptane, particularly the LTR. Figure 3(a) plots the in-cylinder pressure and Figure 3(b)
plots the heat release rate obtained from experiments done by [11] and our simulations, at a fixed engine speed of 900 rpm under constant intake temperature of 333 K and AFR of 50 conditions. As can be seen, the peak cylinder pressure is adequately reproduced, indicating that the important reaction pathways are very well represented. In comparison, the MCS predicted by the numerical simulation is advanced compared to the experimental result as shown in Figure 3(b). A spike in the predicted heat release rate was observed during the MCS. This spike was caused by the rapid oxidation of all CO accumulated to this point in the simulation to CO2. The conversion of CO to CO2 is predicted to happen rapidly once the required temperature is reached because only a few reactions are involved. Rapid oxidation of CO to CO2 has not been observed. The assumption of uniform mixture temperature and mixture composition contributes to this overly-rapid heat release rate. (a) In-cylinder Pressure 60
225 Experiment (Guo et al., 2010) The Present Study
200
50
Experiment (Guo et al., 2010) The Present Study
Heat Release Rate (J\0CA)
175
40
Pressure (bar)
(b) Heat Release Rate
30
20
10
150 125 100 75 50 25 0
0 -100 -75
-50
-25
0
25
Crank Angle
50
75
100
-25 -50 -40 -30 -20 -10
0
10 20 30 40 50
Crank Angle
Figure 3: Comparison between zero-dimensional single zone model with experimental results from [11]. AFR=50, Tin=333 K, Pin=100 kPa, CR=10, N=900 rpm. 3.2 Influence of Air-Fuel Ratio Figure 4 represents the effect of AFR on HCCI combustion and performance for a constant intake air temperature and pressure of 333K and 100 kPa conditions respectively. AFR was varied from 35 to 55. It can be seen from the Figures that stable HCCI combustion was achieved for a wide range of AFR and it is also proved that HCCI combustion is possible to happen in a very lean mixture. As shown in Figure 4(a) and 4 (b), in-cylinder pressure and heat release rate were decreased with the increase of AFR. The combustion was also delayed with the increase of AFR. It is also seen that the phasing of LTR was advanced and phasing of MCS was delayed with increasing AFR. The reason behind these phenomena is that the meeting of hydrocarbon and oxygen molecules during combustion worsens at leaner mixtures resulting in delayed autoignition. The LTR heat release profiles were found to be quite similar, although they were
advanced as AFR increased. However, increasing the AFR significantly decreases the heat release rate of the MCS. The decreasingly delayed phasing of the MCS contributes to the decreased heat release rate [19]. To investigate the influence of AFR on the performance of HCCI engines, five typical AFR were selected for the simulation. As represented in Figure 4(d), 4(e) and 4(f), IMEP, IT and IP decrease with increasing AFR. The reason behind these phenomena is that less amount of air and fuel are derived into the engine when higher AFR is used. It reduces the rate of combustion which ultimately provides lower IMEP, IT and IP [20]. One of the major limitations of HCCI combustion is that the unavailability of a highly diluted mixture to reduce the speed of the chemical reactions which may create excessive pressure. It ultimately damages the engine. By using a higher amount of AFR, this problem can be solved. Under the specific operating condition, the maximum IMEP, IT, and IP occur when AFR is 35. (a) In-cylinder Pressure
(b) Heat Release Rate
AFR=35 AFR=40 AFR=45 AFR=50 AFR=55
50
110
AFR=35 AFR=40 AFR=45 AFR=50 AFR=55
300
Heat Release Rate (J\0CA)
60
Pressure (bar)
(c) Cumulative Heat Release
350
40
30
20
250
100 90
Cumulative Heat Release (%)
70
200
150
100
50
10
-30
-20
-10
0
10
20
30
40
-40
-30
60 50 40 30 20
-20
-10
0
10
20
30
-10 -40
40
(e) Indicated Torque
(d) Indicated Mean Effective Pressure
35
3.5
6
30
3.0
5
4
3
2
1
Indicated Power (kW)
7
Indicated Torque (N-m)
4.0
25
20
15
10
5
0 50
-10
55
0
10
20
30
40
2.5
2.0
1.5
1.0
0.5
0 45
-20
(e) Indicated Power
40
40
-30
Crank Angle
8
35
AFR=35 AFR=40 AFR=45 AFR=50 AFR=55
10
Crank Angle
Crank Angle
Indicated Mean Effective Pressure (bar)
70
0
0 0 -40
80
0.0 35
Air-Fuel Ratio
40
45
50
Air-Fuel Ratio
55
35
40
45
50
55
Air-Fuel Ratio
Figure 4: Influence of air-fuel ratio on combustion and performance characteristics in HCCI engine. Tin=333 K, Pin=100 kPa, CR=10, N=900 rpm. 3.3 Influence of Intake Air Temperature Intake air temperature is the most critical and widely used engine operation parameter to control the phasing of HCCI combustion [21]. Figure 5 illustrates the effect of intake temperature on HCCI combustion and performance for a constant AFR of 50. Intake temperature of the air was
varied from 313 K to 393 K. Stable HCCI combustion was obtained for a wide range of temperatures. As shown in Figure 5(a) the auto-ignition timing was advanced with the increase of intake air temperature. The phasing of both LTR and MCS were advanced. Furthermore, it can also be clearly seen from Figure 5(b) that higher pressure rise rate creates rapid heat release rate. Due to this phenomenon knocking takes place which deteriorates the HCCI combustion. For this reason, cylinder pressure decreases with increasing intake air temperature. The LTR heat release profiles were found to be quite similar, although they were advanced as temperature increased. However, increasing the intake temperature significantly enhances the heat release rate of the MCS. The increasingly advanced phasing of the MCS contributes to the increased heat release rate [22].
(a) In-cylinder Pressure
(b) Heat Release Rate
313 K 333 K 353 K 373 K 393 K
110
313 K 333 K 353 K 373 K 393 K
200
Heat Release Rate (J\0CA)
50
40
Pressure (bar)
(c) Cumulative Heat Release
250
30
20
100 90
Cumulative Heat Release (%)
60
150
100
50
10
0
0
-50
80 70 60 50 40 30 20
313 K 333 K 353 K 373 K 393 K
10 0
-40
-30
-20
-10
0
10
20
30
40
-10 -40
-30
Crank Angle
-10
0
10
20
30
40
-40
(e) Indicated Torque
25
2.5
2
1
Indicated Power (kW)
5
Indicated Torque (N-m)
3.0
3
20
15
10
5
0 353
373
393
Intake Air Temperature (K)
0
10
20
30
40
2.0
1.5
1.0
0.5
0 333
-10
(f) Indicated Power
30
4
-20
Crank Angle
6
313
-30
Crank Angle
(d) Indicated Mean Effective Pressure
Indicated Mean Effective Pressure (bar)
-20
0.0 313
333
353
373
393
Intake Air Temperature (K)
313
333
353
373
393
Intake Air Temperature (K)
Figure 5: Influence of intake air temperature on combustion and performance characteristics in HCCI engine. AFR=50, Pin=100 kPa, CR=10, N=900 rpm. To investigate the influence of intake air temperature on the performance of HCCI engines, five typical temperatures were selected for the simulation. As is shown in Figure 5(d), 5(e) and 5(f), IMEP, IT and IP decrease with increasing intake temperature. This is because there are key chemical reactions in the low-temperature phase. These reactions depend on the system temperature and AFR. C7H15O2=C7H14OOH is the most important reaction in low-temperature
phase. C7H14OOH retards the process of low-temperature oxidation, and it plays the most important role in the occurrence of the negative temperature coefficient (NTC) phase [23, 24]. Under the specific operating condition, the maximum IMEP, IT, and IP occur when intake temperature is 313 K. Obviously, the intake temperature is very critical, i.e., a small change of intake temperature leads to large variations in performance of HCCI. 3.4 Influence of Intake Air Pressure Intake air pressure boosting has a significant effect on the combustion and performance of HCCI engines as well as is a common way to improve power output. In this study, the effect of boosting on HCCI combustion and performance was examined at a constant AFR and intake air temperature condition. As shown in Figure 6(a), boosting the intake pressure tends to significantly increase the peak cylinder pressure during the compression stroke. The reason behind this phenomenon is that with increasing intake air pressure the collision frequency among molecules increases. For this reason, combustion reaction velocity of HCCI combustion also increases [25]. Since the quantity of fuel injected is increased as intake pressure increases to maintain a constant AFR, the LTR stage is advanced and intensified. This leads to a shorter NTC delay period and significantly advances phasing of the MCS, as shown in Figure 6(b). The reason for this is the combustion reaction velocity which increases with the increasing boost pressure [26, 27]. To investigate the influence of intake air pressure on the performance of HCCI engines, five typical pressures ranging from 100 kPa to 200kPa were selected for the simulation. As is shown in Figure 6(d), 6(e) and 6(f), IMEP, IT and IP increases with increasing intake air pressure. The maximum IMEP, IT, and IP were obtained at 200 kPa. This is because higher intake air pressure brings a higher amount of air into the cylinder and to maintain constant AFR higher amount of fuel is also injected into the cylinder. For this reason, a higher amount of IMEP, IT and IP are achieved at higher intake air pressure. On the other hand, the engine can operate stably at a leaner AFR with an increasing inlet pressure, which leads to lower values of IMEP, IT, and IP. However, if proper tuning is carried out to control combustion phasing using different intake air temperature, EGR or compression ratio, more benefits can be achieved from boosting intake air pressure. The results indicate that an increase in the boost pressure causes the need of leaner mixture, and requires more advanced injection timing to achieve the maximum engine torque [28, 29].
450
90
100 kPa 125 kPa 150 kPa 175 kPa 200 kPa
400 350
Heat Release Rate (J\0CA)
80 70
Pressure (bar)
110
60 50 40 30 20
100 90
Cumulataive Heat Release (%)
100 kPa 125 kPa 150 kPa 175 kPa 200 kPa
100
300 250 200 150 100 50
80 70 60 50 40 30
10
0
0
0
-50
-10
-30
-20
-10
0
10
20
30
40
-40
-30
Crank Angle
-20
-10
0
10
20
30
100 kPa 125 kPa 150 kPa 175 kPa 200 kPa
20
10
-40
40
-40
-30
Crank Angle
-20
-10
0
10
20
30
40
Crank Angle
(e) Indicated Torque
(d) Indicated Mean Effective Pressure 12.5
(f) Indicated Power
60
6
50
5
7.5
5.0
Indicated Power (kW)
10.0
Indicated Torque (N-m)
Indicated Mean Effective Pressure (bar)
(c) Cumulative Heat Release
(b) Heat Release Rate
(a) In-cylinder Pressure 110
40
30
20
4
3
2
2.5 10
0.0
1
0 100
125
150
175
200
Intake Air Pressure (kPa)
0 100
125
150
175
Intake Air Pressure (kPa)
200
100
125
150
175
200
Intake Air Pressure (kPa)
Figure 6: Influence of intake air pressure on combustion and performance characteristics in HCCI engine. AFR=50, Tin=333 K, CR=10, N=900 rpm. 3.5 Influence of Compression Ratio Compression ratio is another important engine parameter which also has a significant effect on combustion and performance characteristics of HCCI engines [30]. Figure 7 shows that increasing the compression ratio advances the combustion process and increases the peak cylinder pressures. This is primarily due to the effect of increased compression temperatures and pressures as the compression ratio increases, which enhances n-heptane oxidation [31]. In the case of the higher compression ratios, the temperature around TDC is relatively high. This is conductive to initiation of the chemical reactions. As a result, a large amount of energy is released which ultimately creates an increased pressure and temperature [32]. As shown in Figure 7(b), the phasing of both LTR and MCS phases were advanced with increasing compression ratio. To investigate the influence of compression ratio on the performance of HCCI engines, five typical compression ratios were selected for the simulation. As is shown in Figure 7(d), 7(e) and 7(f), IMEP, IT and IP increases with increasing compression ratio. This is because there are key chemical reactions which occur during the compression stroke. These reactions depend on the in-cylinder pressure and temperature. Due to increasing compression ratio oxidation of n-
heptane is enhanced which ultimately help to increase IMEP, IT and IP [24, 33]. Under the specific operating condition, the maximum IMEP, IT, and IP occur when the compression ratio is 16. It is obvious that the compression ratio is very critical, i.e., a small change of compression ratio leads to large variations in performance of HCCI.
(a) In-c ylinder P res s ure
(c ) C umulative Heat R eleas e
(b) Heat R eleas e R ate
100
350
CR=10 CR=11.5 CR=13 CR=14.5 CR=16
90 80
110
CR=10 CR=11.5 CR=13 CR=14.5 CR=16
300
100 90
P re s s u re (b a r)
0
H e a t R e le a s e R a te (J \C A )
70 60 50 40 30 20
C u m u la tiv e H e a t R e le a s e (% )
250
200
150
100
50
80 70 60 50 40 30 20 CR=10 CR=11.5 CR=13 CR=14.5 CR=16
10 0
10
0
0
-50 -40
-30
-20
-10
0
10
20
30
40
-10 -40
-30
C ran k A n gle
-20
-10
0
10
20
30
40
-40
-30
C ran k A n gle
-10
0
10
20
30
40
C ran k A n gle
(e) Indic ated T orque
(d) In dic ated Mean E ffec tiv e P res s u re
-20
(f) Indic ated P ow er
6
30
5
25
4
3
2
1
In d ic a te d P o w e r (k W )
2.5
In d ic a te d T o rq u e (N-m )
In d ic a te d Me a n E ffe c tiv e P re s s u re (b a r)
3.0
20
15
10
0 10
12
14
C om pres s ion R atio
16
1.5
1.0
0.5
5
0
2.0
0.0 10
12
14
C om pres s ion R atio
16
10
12
14
16
C om pres s ion R atio
Figure 7: Influence of Compression Ratio on combustion and performance characteristics in HCCI engine. AFR=50, Tin=333 K, Pin=100 kPa, N=900 rpm. 3.6 Influence of Speed Engine speed is an important parameter which has a significant effect on combustion and performance of HCCI engines. Figure 8 illustrates the effect of engine speed on HCCI combustion and performance for a constant intake temperature of 333 K. Engine speed was varied from 600 rpm to 1200 rpm. From the Figure 8(a) it is clearly seen that in-cylinder pressure increases with increasing engine speed up to 900 rpm and then decreases with further increasing engine speed. The most important reason is heat loss through cylinder wall during the compression stroke. With the engine speed increasing, the accumulated heat loss during the compression stroke reduced significantly because of the shorter cyclic period, which increases the in-cylinder temperature at the end of compression stroke. Therefore, the chamber could
achieve a higher maximum pressure at higher engine speed due to less heat loss through the cylinder wall [34]. In addition, peak pressure location is retarded with the increase of engine speed. This is due to a significantly retarded MCS, as shown in Figure 8(b). The heat release rate during the MCS is also reduced with retarded combustion due to the increasing combustion chamber volume after top dead center. This leads to lower combustion chamber temperatures and a corresponding decrease in the oxidation reaction rates [35]. However, the effect of engine speed on the LTR phase tends to be relatively weak, reflecting its strong dependence on temperature history [23].
(a) In-cylinder Pressure 300
600 rpm 750 rpm 900 rpm 1050 rpm 1200 rpm
110 600 rpm 750 rpm 900 rpm 1050 rpm 1200 rpm
250
40
30
20
100 90 80
200
Cumulative Heat Release
Heat Release Rate (J\0CA)
50
Pressure (bar)
(C) Cumulative Heat Release
(b) Heat Release Rate
60
150
100
50
10
70 60 50 40 30 20
600 rpm 750 rpm 900 rpm 1050 rpm 1200 rpm
10
0
0 0
-50 -40
-30
-20
-10
0
10
20
30
40
-10 -40
-30
Crank Angle
-10
0
10
20
30
40
-40
(e) Indicated Torque
25
2.5
2
1
Indicated Power (kW)
5
Indicated Torque (N-m)
3.0
3
20
15
10
5
0 900
1050
Speed (rpm)
1200
0
10
20
30
40
2.0
1.5
1.0
0.5
0 750
-10
(f) Indicated Power
30
4
-20
Crank Angle
6
600
-30
Crank Angle
(d) Indicated Mean Effective Pressure
Indicated Mean Effective Pressure (bar)
-20
0.0 600
750
900
1050
Speed (rpm)
1200
600
750
900
1050
1200
Speed (rpm)
Figure 8: Influence of speed on combustion and performance characteristics in HCCI engine. AFR=50, Tin=333 K, Pin=100 kPa, CR=10. To observe the influence of engine speed on the performance of HCCI engines, five typical speeds were selected for the simulation. As is shown in Figure 8(d), 8(e) and 8(f), IMEP, IT and IP increases with increasing engine speed up to 900 rpm and then decreases with further increasing engine speed. Maximum IMEP, IT, and IP were obtained as 5.69 bar, 27.72 Nm and 2.61 kW. At higher engine speed the performance was less. This is because of that, with the increase of engine speed, the combustion duration is shortened with respect to time and affects the performance of HCCI engines [36]. Moreover, volumetric efficiency depends on engine
speed. Because of the changing volumetric efficiency and increased friction forces at higher engine speed, the IMEP, IT and IP were less[37]. 4. Conclusions A zero-dimensional single zone model was established to investigate HCCI combustion and performance characteristics over a wide range of operating conditions. A reduced n-heptane chemical reaction mechanism was used to solve the chemical reactions during combustion. The validated numerical simulation was able to capture trends in combustion phasing variation with critical engine parameters, particularly the LTR. From the simulation, it was seen that HCCI combustion was achieved easily for very lean mixture. Decreasing the AFR advances the phasing of the MCS and IMEP, IT and IP decreases with AFR. The phasing of both the LTR and MCS were advanced with increasing intake air temperature and compression ratio. However, opposite results were found in terms of performance while increasing intake air temperature and compression ratio. That means, performance decreases with increasing intake air temperature and increases with increasing compression ratio. In addition, it was obtained from the simulation that the effects of turbocharging on the phasing of LTR are relatively weak. Moreover, increasing the engine speed peak pressure location is retarded due to retarded MCS. IMEP, IT, and IP were increased up to a certain limit then decreased with increasing speed. This is because of with the increase in engine speed; the combustion duration is shortened with respect to time. Acknowledgements The authors would like to thank Universiti Malaysia Pahang for financial support under project no. RDU1503101 and providing laboratory facilities. References [1] Mohanamurugan S, Sendilvelan S. Emission and combustion characteristics of different fuel In A HCCI engine. International Journal of Automotive and Mechanical Engineering. 2011;3:279-92. [2] Hairuddin A, Wandel A, Yusaf T. An introduction to a homogeneous charge compression ignition engine. Journal of Mechanical Engineering and Sciences. 2014;7:1042-52. [3] Hasan MM, Rahman MM, Kadirgama K. A review on homogeneous charge compression ignition engine performance using biodiesel-diesel blend as a fuel. International Journal of Automotive & Mechanical Engineering. 2015;11:2199-211. [4] Mohanamurugan S, Sendilvelan. Emission and combustion characteristics of different fuel in HCCI engine. International Journal of Automotive and Mechanical Engineering. 2011;3:279-92. [5] Polovina D, McKenna D, Wheeler J, Sterniak J, Miersch-Wiemers O, Mond A, et al. Steadystate combustion development of a downsized multi-cylinder engine with range extended HCCI/SACI capability. SAE Technical Paper; 2013. [6] Izadi Najafabadi M, Abdul Aziz N. Homogeneous charge compression ignition combustion: challenges and proposed solutions. Journal of combustion. 2013;2013. [7] Yao M, Zheng Z, Liu H. Progress and recent trends in homogeneous charge compression ignition (HCCI) engines. Progress in Energy and Combustion Science. 2009;35:398-437.
[8] Saxena S, Bedoya ID. Fundamental phenomena affecting low temperature combustion and HCCI engines, high load limits and strategies for extending these limits. Progress in Energy and Combustion Science. 2013;39:457-88. [9] Mangus M, Kiani F, Mattson J, Depcik C, Peltier E, Stagg-Williams S. Comparison of neat biodiesels and ULSD in an optimized single-cylinder diesel engine with electronically-controlled fuel injection. Energy & Fuels. 2014;28:3849-62. [10] Komninos N, Rakopoulos C. Modeling HCCI combustion of biofuels: A review. Renewable and Sustainable Energy Reviews. 2012;16:1588-610. [11] Guo H, Neill WS, Chippior W, Li H, Taylor JD. An experimental and modeling study of HCCI combustion using n-heptane. Journal of Engineering for Gas Turbines and Power. 2010;132:022801. [12] Assanis D, Polishak M. Valve event optimization in a spark-ignition engine. Journal of Engineering for Gas Turbines and Power. 1990;112:341-7. [13] Seiser R, Pitsch H, Seshadri K, Pitz W, Gurran H. Extinction and autoignition of n-heptane in counterflow configuration. Proceedings of the Combustion Institute. 2000;28:2029-37. [14] Goodwin D. An open-source, extensible software suite for CVD process simulation. Chemical Vapor Deposition XVI and EUROCVD. 2003;14:2003-08. [15] Andrae J. A kinetic modeling study of self-ignition of low alkylbenzenes at engine-relevant conditions. Fuel Processing Technology. 2011;92:2030-40. [16] Heywood JB. Internal combustion engine fundamentals: Mcgraw-hill New York; 1988. [17] Kuo KK. Principles of combustion. 1986. [18] Stiesch G. Modeling engine spray and combustion processes: Springer Science & Business Media; 2013. [19] Polat S. An experimental study on combustion, engine performance and exhaust emissions in a HCCI engine fuelled with diethyl ether–ethanol fuel blends. Fuel Processing Technology. 2016;143:140-50. [20] Maurya RK, Agarwal AK. Experimental investigation on the effect of intake air temperature and air–fuel ratio on cycle-to-cycle variations of HCCI combustion and performance parameters. Applied Energy. 2011;88:1153-63. [21] Maurya RK, Agarwal AK. Experimental study of combustion and emission characteristics of ethanol fuelled port injected homogeneous charge compression ignition (HCCI) combustion engine. Applied Energy. 2011;88:1169-80. [22] Uyumaz A. An experimental investigation into combustion and performance characteristics of an HCCI gasoline engine fueled with n-heptane, isopropanol and n-butanol fuel blends at different inlet air temperatures. Energy Conversion and Management. 2015;98:199-207. [23] Zhang CH, Xue L, Wang J. Experimental study of the influence of λ and intake temperature on combustion characteristics in an HCCI engine fueled with n-heptane. Journal of the Energy Institute. 2014;87:175-82. [24] Peng Z, Zhao H, Ma T, Ladommatos N. Characteristics of homogeneous charge compression ignition (HCCI) combustion and emissions of n-heptane. Combustion science and technology. 2005;177:2113-50. [25] Liu H, Yao M, Zhang B, Zheng Z. Effects of inlet pressure and octane numbers on combustion and emissions of a homogeneous charge compression ignition (HCCI) engine. Energy & Fuels. 2008;22:2207-15. [26] Christensen M, Johansson B. Supercharged homogeneous charge compression ignition (HCCI) with exhaust gas recirculation and pilot fuel. SAE Technical Paper; 2000.
[27] Olsson J-O, Tunestål P, Johansson B. Boosting for high load HCCI. SAE Technical Paper; 2004. [28] Hosseini V, Checkel MD. Intake pressure effects on HCCI combustion in a CFR engine. Proceedings of the combustion institute, canadian section, spring technical meeting2007. [29] Olsson J-O, Tunestål P, Haraldsson G, Johansson B. A turbocharged dual-fuel HCCI engine. SAE Special Publications. 2001;2001. [30] Olsson J-O, Tunestål P, Johansson B, Fiveland S, Agama R, Willi M, et al. Compression ratio influence on maximum load of a natural gas fueled HCCI engine. SAE Technical Paper; 2002. [31] Li H, Guo H, Neill WS, Chippior W, Taylor JD. An experimental and modeling study of HCCl combustion using n-heptane. Proceedings of the ICEF2006. 2006:1-11. [32] Iida M, Hayashi M, Foster D, Martin J. Characteristics of homogeneous charge compression ignition (HCCI) engine operation for variations in compression ratio, speed, and intake temperature while using n-butane as a fuel. Journal of engineering for gas turbines and power. 2003;125:472-8. [33] Liu M-B, He B-Q, Zhao H. Effect of air dilution and effective compression ratio on the combustion characteristics of a HCCI (homogeneous charge compression ignition) engine fuelled with n-butanol. Energy. 2015;85:296-303. [34] Zhang C, Wu H. Combustion characteristics and performance of a methanol fueled homogenous charge compression ignition (HCCI) engine. Journal of the Energy Institute. 2015. [35] García MT, Aguilar FJJ-E, Lencero TS. Experimental study of the performances of a modified diesel engine operating in homogeneous charge compression ignition (HCCI) combustion mode versus the original diesel combustion mode. Energy. 2009;34:159-71. [36] Canakci M. An experimental study for the effects of boost pressure on the performance and exhaust emissions of a DI-HCCI gasoline engine. Fuel. 2008;87:1503-14. [37] Nishi M, Kanehara M, Iida N. Assessment for innovative combustion on HCCI engine by controlling EGR ratio and engine speed. Applied Thermal Engineering. 2016;99:42-60. List of Figures Figure 1: An algorithm flow chart for zero-dimensional single-zone model simulation. Figure 2: Schematic diagram of the HCCI engine set up, reproduced from [11]. Figure 3: Comparison between zero-dimensional single zone model with experimental results from [11]. AFR=50, Tin=333 K, Pin=100 kPa, CR=10, N=900 rpm. Figure 4: Influence of air-fuel ratio on combustion and performance characteristics in HCCI engine. Tin=333 K, Pin=100 kPa, CR=10, N=900 rpm. Figure 5: Influence of intake air temperature on combustion and performance characteristics in HCCI engine. AFR=50, Pin=100 kPa, CR=10, N=900 rpm. Figure 6: Influence of intake air pressure on combustion and performance characteristics in HCCI engine. AFR=50, Tin=333 K, CR=10, N=900 rpm.
Figure 7: Influence of Compression Ratio on combustion and performance characteristics in HCCI engine. AFR=50, Tin=333 K, Pin=100 kPa, N=900 rpm. Figure 8: Influence of speed on combustion and performance characteristics in HCCI engine. AFR=50, Tin=333 K, Pin=100 kPa, CR=10.
List of Tables Table 1: Engine model specifications [11]. Table 2: Simulation matrix.
HIGHLIGHTS OF MANUSCRIPT
HCCI combustion was achieved easily for very lean mixture. The phasing of combustion was advanced with increasing intake air temperature. Turbocharging effect on the phasing of low temperature reaction is relatively weak. Better performance was found with increasing compression ratio. Peak pressure location was retarded with increasing engine speed.