Modeling Diesel Engine Flow Processes For Performance and Emissions

Transcription

1 Modelin Diesel Enine Flow Processes For Performance and Emissions B.Venkateshwar Rao, K.Madhu Murthy G.Amba Prasad Rao * Abstract- The paper deals with the simulation of typical direct injection diesel enine processes followin phenomenoloical modelin considerin heat transfer, heat losses (both convection and radiation), and variable specific heats. A double-weibe function is used to model the heat release. It was found that early injection timin leads to hiher levels of pressure and temperature in the cylinder. Fuel injection timin and enine speed, compression ratio, inlet chare pressure and temperature are observed to be pertinent parameters affectin diesel enine performance. Effect of exhaust as recirculation on the formation and emission of oxides of nitroen and soot density are also studied. Hih speed diesel fuel C 1.8 H 18.7 is considered for calculations. Numerical experiments are performed by writin a computer code in C ++ and heat release (both pre-mixed and diffusion phases), in-cylinder pressure and temperature histories and emissions are predicted while incorporatin different enine parameters. All trends obtained are in accordance with the wellestablished literature. It is observed that fuel injection timin and fraction of exhaust as recirculation as critical factors in affectin enine performance as well as emissions. Keywords Diesel Enine, Phenomenoloical Model, Wiebe Function, EGR, Performance, Exhaust Emissions D I.INTRODUCTION IRECT injection diesel enines exhibit better performance as far as fuel economy is concerned compared to asoline enines. Of late, due to strinent emission norms researchers and leadin manufacturers are aimin for the development of clean diesel enines. In this reard, computer simulations are found to be prominent tools for arrivin at the optimum desins and to make the diesel technoloy be more competitive. Thouh there are various models such thermodynamics and fluid dynamics based models available, phenomenoloical models are attractive in the liht of less computational complexities involved. These models can be made more attractive by imposin all possible practical conditions the diesel enine experiences and to predict performance near to actual cycle simulations. Typical direct injection diesel enine combustion process comprises of four phases viz; inition delay, pre-mixed, diffusion and late burnin. Abu-Nada et al. [1-3] carried out enine simulations takin into account the effect of heat transfer, friction, and B.Venkateshwar Rao is Graduate student at the National Institute of Technoloy, Waranal, India. K. Madhu Murthy is Professor in the Dept. of Mechanical Enineerin at National Institute of Technoloy, Waranal, India. G. Amba Prasad Rao is Associate Professor in the Dept. of Mechanical Enineerin at National Institute of Technoloy, Waranal, India ( ambaprasadrao@mail.com). temperature dependent specific heats on the overall enine performance. Miyamoto et al. [4] the model was oriinally developed for spark inition (SI) enines; they claimed that it could be extended and modified to simulate compression inition (CI) enines as well. This results in a sinificant shift in the rate of heat release model from the simple Weibe function commonly used for SI enines. A double peak heat release model becomes more representative CI enines [4].Arrele et al. [6]studied the influence of injection parameters and runnin conditions on heat release in a Diesel enine. Galindo et al. [7]used four different Weibe functions to account for pilot injection, premixed, diffusion and late combustion in the heat release model. Chemla et al. [8] used a zero-dimensional rate of heat release model for the simulation of direct injection diesel enine. Aithal [9] studied effect of EGR fraction on diesel enine performance considerin heat loss and temperature dependent properties of the workin fluid. The objective of the present work is to analyze the performance of a CI enine usin a phenomenoloicalthermodynamics based model considerin double-weibe function for diesel combustion process. The developed model predicts in-cylinder temperatures and pressures as functions of the crank anle, variation of premixed, diffusion, heat release pattern by varyin the enine operatin and desin parameters. The novelty of present model lies in considerin effects of heat losses (both convection and radiation) and temperaturedependent specific heats in addition to inition delay and EGR fraction. These features can be very useful in providin more realistic estimations for the performance and emission parameters in comparison to the existin models for CI enines. II.PHENOMENOLOGICAL MODELING The overnin equation for calculatin variation of incylinder pressure with respect to crank anle is: dp k 1 dqin dqloss d V d d In Eq. (1), the rate of heat loss dqloss d 1 ha( )( T Tw ) dq loss d P dv P dk k V d k 1 d is expressed as: (2) The convective heat transfer coefficient is iven by the Woschni model as [1]: (1) 185

2 h D P T w (3) 2NS The mean piston speed is related as U p 6 (4) Based on thewoschni s formula calculate the convective mode of heat losses, assume it as 7% of total losses, based on that calculate 3% of heat losses, assume this 3% of heat losses is due to radiation. On the other hand, the rate of the heat input dq in d (heat release) can be modeled usin a dual Weibe function [5, 7 and 8]: mp 1 mp md 1 md dqin Q p Qd a m p exp a a m d exp a d p p p d d d Where p and d refers to premixed and diffusion phases of combustion. The parameters θ p and θ d represent the duration of the premixed and diffusion combustion phases. Also, Q p and Q d represent the intated enery release for premixed and diffusion phases respectively. The constants a, m p, m d are selected to match experimental data. For the current study, these values are selected as 6.9, 4, and 1.5 respectively [5]. It is assumed that the total heat input to the cylinder by combustion for one cycle is: Qin m f LHV (6) Equations for inition delay and fraction of fuel burned in premixed combustion are taken from Heywood [1]. Eq. (1) is discretized usin a first order finite difference method to solve for the pressure at each crank anle (θ). Once the pressure is calculated, the temperature of the ases in the cylinder can be calculated usin the equation of state as: T P( ) V ( ) mr (5) (7) The instantaneous cylinder volume, area, and displacement are taken from literature [1]. The initial NO formation rate written as d[ NO] 61 1/ 2 dt T 16 69,9 exp [ O2 ] T 1/ 2 e [ N ] mol/cm 3.s (8) The formation rate of soot particles is expressed in the Arrhenius form dm dt soot A m Initial temperature f fuel p.5 E f exp RT T[] mata m ( ma m P[ ] V[] ( m m ) R T[] a T ) 2 e (9) (1) (11) The followin assumptions were made in modelin the compression and power strokes: i. All thermodynamic variables were assumed to be spatially uniform throuhout the enine at any instant of time (zero-dimensional model) and variables varied temporally durin the enine cycle. ii. Compression and expansion were assumed to take place in a series of quasi-equilibrium processes. iii. Compression stroke bean at and expansion stroke ended at 36.Fuel injection bean at i and ended at f. iv. For i i, fuel was injected but not burned due to the inition delay. Fuel combustion took place for i. v. For i, only a prescribed fraction of the instantaneous fuel mass available in the cylinder was burned. vi. EGR fraction was defined as the ratio of the mass of exhaust as to the mass of fresh air m m a at BDC. vii. Effect of soot formation on the reduction of the flame temperature (due to radiation) was considered. viii. EGR was assumed to consist of CO, H O, O N (effect of residual ases 2 2 2, nelected). 2 The enery equation expressin the relationship between pressure and crank anle was solved usin Euler s method (First order finite difference expression), to obtain the work output and cycle efficiency. The specifications of enine and other pertinent parameters are iven Table 1. III. RESULTS AND DISCUSSION Numerical experiments are performed, makin use of the equations and imposin relevant conditions for the chosen diesel enine confiuration, for predictin enine performance and emissions. Enine performance modelin To bein with, heat release calculations are done for different crank anles; the pattern is shown in Fi.1, takin into consideration the heat transfer in the enine and variable specific heats of the air, usin a dual Weibe function. This result is treated as base line data; it can be observed that there are three dominant combustion staes, namely premixed, diffusion and late combustion phases. The predicted combustion phases are in ood areement with the Lyn and Ways model [1]. 186

3 hihest for late injection timin since heat release continues even after piston crosses the TDC. Also, the early injection timin develops hiher pressures and temperatures leadin to hiher power output compared to delayed injection timin. The effect is illustrated in Fis. 5 and6. Fi.1. Heat release pattern for the diesel enine at N=25 rpm, Φ=.6, and FIB 8 btdc The predicted heat release pattern is utilized to derive the in-cylinder pressure history to obtain enine performance. Incylinder pressure variation with crank anle is obtained as shown in Fi.2, it can be observed that a deviation of combustion curve from the motorin pressures and peak pressure reachin about 9 bar occurrin very near to TDC. Fi. 4 Variation of premixed combustion heat profiles for different injection timins. Fi.3. 2 In-cylinder pressure for the diesel enine at N=25 rpm, Φ=.6, and FIB 8 btdc. Enine heat transfer from hot ases to walls comprises both convection, radiation and affects its performance, and emissions. The predominant portion of heat is lost throuh convection and radiation losses are sinificant in diesel enine due to soot formation accountin to about 25 to 35 per cent of the total heat transfer. For a iven mass of fuel within the cylinder, hiher heat transfer to the combustion chamber walls will lower averae combustion as temperature and simultaneous reduction in pressure decreases the work per cycle transferred to the piston. Fi. 3represents rate of heat loss versus crank anle for the Diesel enine runnin at 25 rpm and Φ=.6. Fi. 3 Diesel enine heat loss history for N=25 rpm and Φ=.6 and FIB 8 btdc. By advancin the fuel injection timin, inition delay will increase which causes increase in the fraction of fuel burned in pre-mixed combustion phase and thus increases premixed combustion stae as illustrated in Fi.4. In early injection timin, most of the heat release takin place before piston reaches the TDC and for the case of late injection timin heat release still continues even after piston crosses the TDC. However, an interestin observation is that 2 dees after TDC, the as temperature and pressure in the cylinder is still Fi.5. In-cylinder pressures for different injection timins. Fi. 6.In-cylinder temperatures for different injection timins. With increase in enine speed, the absolute time in milliseconds ets reduced. The actual time for accumulation of fuel will be reduced with lesser amount of fuel injected. This effect reduces the inition delay in milliseconds, causin decrease in the fraction of fuel burned in pre-mixed combustion. However, the duration of pre-mixed combustion increases with increase in speed. As the pre-mixed heat release decreases with speed, the combustion phasin shifts to diffusion with relatively hiher values. The speed has sinificant effect on heat loss as the speed decreases, the absolute time taken for heat exchane increases resultin in more heat loss to the cylinder walls. These effects are well predicted throuh the present numerical experiments. With the consequent reduction in heat losses, eventually the enery is utilized in developin more work. This can be observed with increase in in-cylinder pressures for the iven conditions. Also, increase in enine speed decreases time for exchane of heat with cylinder walls resultin in reduced heat losses and finally yieldin hiher temperatures within the cylinder. The ultimate effect of speed is that it increases enine efficiency. Enine emissions modelin: Thouh direct injection diesel enines are known to be fuel efficient compared asoline enines, challenes still exist with 187

4 reards to emission of hiher levels of NOx and PM. Thus researchers and leadin manufactures are resortin to limit these harmful emissions. Numerical experiments are performed to observe the emission formation phenomenon. Dilution of the intake air with cooled re-circulated exhaust as limits the production hih temperatures with subsequent reduction of NO x due to a lowerin of the adiabatic flame temperature, reduction in oxyen content of the intake mixture and reduction of heat input inside the combustion chamber. The popular technique to accomplish this task is exhaust as recirculation (EGR). EGR also reduces the mixture-averaed ratio of specific heats (k) of the chare, pressure, temperature leadin to a reduction in the thermodynamic cycle efficiency. Also, with increase in EGR fraction heat release in both premixed and diffusion combustion phase reduces, as shown in Fis.7 and 8 respectively. Utilizin Zeldovich mechanism related equations for NOx formation with in-cylinder temperature history the appropriate equations are solved for predictin NOx emissions. Apart from, EGR retardin of fuel injection beinnin is also known one of the techniques for NOx reduction methods. The effect of EGR alone is plotted as Fi.9. Thus, it can be noticed that as the EGR fraction increases, there appears to substantial reduction in NOx formation and emission and also with retardin timin there is reduction in NOxemission[1]. Fi. 7. Variation of pre-mixed combustion with EGR fraction Fi. 8. Variation of diffusion combustion with EGR fraction Fi. 9. Variation of NO x emissions with EGR fraction In enines, the formation mechanisms for NOx and soot (or PM) are opposite to each other. Hih levels of EGR can also lead to hih levels of PM due to a reduction in the oxyen levels and poor combustion quality. Additionally, EGR leads to hiher specific fuel consumption and enine noise while adversely affectin the lubricatin oil quality and enine durability. Also, early injection timin leads to hiher levels of pressure and temperature which causes drop in the net soot density. Effect of EGR is shown in Fi.1. From Fis.9 and 1, it can be observed clearly that the factors which decrease the NOx would increase soot and vice-versa. Therefore, a trade-off should be obtained without emittin either harmful pollutants or in deterioration of enine ross performance [1]. Fi.1. Variation of soot emissions with EGR fraction IV. CONCLUSIONS In the present work, DI diesel enine flow processes for performance and emissions are modeled usin dual Wiebe function for heat release analysis. Based on the work, the followin conclusions are drawn: Advancin of fuel injection timin, inition delay increaseswhich causes increase of premixed combustion with a minor effect on diffusion combustion. Early injection timin leads to hiher levels of pressure and temperature inside the combustion chamber. Early injection timin leads to the formation of hiher levels of NO x emission and lower levels of soot emission. Increase in enine speed, inition delay increases in crank anle dees, but decreases in milliseconds causin drop in premixed combustion however, with increased diffusion phase. Thermal efficiency, pressure and temperature will increase by increasin the enine speed. By increasin the fraction of EGR supply to the enine, heat input, pressure, temperature, adiabatic flame temperature, and efficiency will decrease. Increase in EGR fraction reduces the formation of NO x emissions and increase the formation of soot emission. ACKNOWLEDGEMENTS The authors wholeheartedly thank the authorities of NIT Waranal for their cooperation and permittin to publish the present work. REFERENCES [1] E. Abu-Nada, I. Al-Hinti, B. Akash, Al-Sarkhi, Thermodynamic analysis of spark inition enine usin a as mixture model for the workin fluid, International Journal of Enery Research 31 (27) [2] E. Abu-Nada, Sakhrieh,I. Al-Hinti, A. Al-Ghandoor, B. Akash Computational thermodynamic analysis of compression inition enine International Communications in Heat and Mass Transfer 37 (21) [3] E. Abu-Nada, I. Al-Hinti, A. Al-Sarkhi, B. Akash, Effect of piston friction on the performance of SI enine: a new thermodynamic approach, ASME Journal of Enineerin for Gas Turbines and Power 13 (2) (28) [4] J I Ghojel Review of the development and applications of the Wiebe function: a tribute to the contribution of Ivan Wiebe to enine research Int. J. Enine Res. Vol. 11,

5 [5] N. Miyamoto, T. Chikahisa, T. Murayama, R. Sawyer, Description and analysis of diesel enine rate of combustion and performance usin Weibe's functions, SAE paper 8517, [6] J. Arrèle, J.M. Garcia, J.J. Lopez, C. Fenollosa, Development of a zerodimensional Diesel combustion model. Part 1: analysis of the quasisteady diffusion combustion phase, Applied Thermal Enineerin 23 (23) [7] J. Galindo, J.M. Lujan, J.R. Serano, L. Hernandez, Combustion simulation of turbocharer HSDI Diesel enines durin transient operation usin neural networks, Applied Thermal Enineerin 25 (25) [8] F.G. Chemla, G.H. Pirker, A. Wimmer, Zero-dimensional ROHR simulation for DI diesel enines a eneric approach, Enery Conversion and Manaement 48 (27) [9] AithalSM.Impact of EGR fraction on diesel enine performance considerinheat loss and temperature dependent properties of the workin fluid.int J Enery Res 28; 33: [1] Heywood JB. Internal Combustion Enine Fundamentals, New York: McGraw Hill; TABLE 1. SPECIFICATIONS OF ENGINE Fuel Cetane number 45 Lower heatin value (kj/k) Molecular weiht (k/kmol) Stoichiometric air fuel 14.36:1 ratio Compression ratio 18:1 Bore X Stroke (m X m).15 X.125 Connectin rod lenth.1 (m) Swept volume (m 3 ) Enine speed rane (rpm) 1 5 Equivalence ratio,φ.6 Injection timin 12 to 8 Duration of combustion 7 Wall temperature (K) 4 189