Direkt püskürtmeli bir dizel motorunda hava ve damlacık hareketlerinin modellenmesi

Abstract

Pistonlu motorlarda, yanma odası şeklinin yanma verimine etki ettiği bilinmektedir. Bu çalışmada, ele alınan bir direkt enjeksiyonlu dizel motorunun yanma odası içindeki hava hareketleri ve bu hareketle rin yakıt damlacıklarının yanma odası içindeki dağılı mına etkileri incelenmiştir. Yanma odasındaki hava hareketi, akışın iki boyut lu, eksenel simetrik ve adyabatik olduğu kabulüyle, sıkıştırmanın başladığı andan başlanarak, püskürtmenin belli bir safhasına kadar incelenebilmiştir. Türbülansm model lenmesi için k-e modeli kulla nılmıştır. Türbülansın akış alanı içindeki etkisi efektif viskozite ile sağlanmıştır. Püskürtme olayı sırasında, damlacıkların zamana bağlı bir akım alanı içindeki hareketleri incelen miş ve damlacık dağılımı, damlacıklara ait korunum denklemlerinin, hareketi takiben çözülmesiyle elde edilmiştir. Silindir içindeki akış alanının Çözülmesinden ve damlacıkların hareketlerinin incelenmesinden elde edi len sonuçların literatürde mevcut deneysel ve teorik sonuçlarla yeterli bir uyum içinde olduğu gözlenmiş tir.

The turbulent flow field of a direct-injection diesel engine strongly infuluences its combustion characteristics and is affected by the design of the combustion chamber. It is important then to under stand the mechanism by which these engine variables affect engine turbulance. In this study, in-cylinder air motion of a direct injection diesel engine and its effects on the motion and consentration of the droplets, which are injected into the combustion chamber during injection time, are studied. 1. In-Cylinder Air Motion Model: The governing equations to be solved for the in- cylinder flow are described on the assumptions that instantaneous fluid density is spatially uniform over the flow field but depends on time. The effective viscosity is introduced to represent the effect of the turbulance. Computation is made step-by-step in terms of crank angle increment starting from the beginning of the compression until the end of the calculation. It is also assumed that the working fluid is an ideal fluid and no combustion takes place. The change in air density due to compression is calculated with the approximation of a polytropic change. The polytropic index is invariably set at 1.4. Computations has been performed relying on a semi- implicit method. The first stage for obtaining the flow field after a time increment 6t, is to guess the velocity components u,v by assuming the pressure field in the present time. At the same time k and e are determined explicitly. VI Since the pressure gradients 6P/6x, 6P/6y are not explicitly expressed, in the second stage, the velocity components are estimated to obtain pressure corrections which can be solved using Newton- Raphson method by applying line-by-line double-sweeping method. The model employed here, is for a two-dimensional axisymmetric flow field. Turbulance is described a two equation model for its kinetic energy k and its dissipation rate e. It is assumed that the turbulance field is not affected by means of the motions of the droplets. When the air is being sucked into the cylinder during the intaking time, there will occur a free vortex because of the source-sink interactions of the sucking and exhaust valves. This vortex is called as swirl in the literature. It is repoted that the effects of such a swirl makes the flow field change completelly. In this study, the effects of such a swirl is not taken under consideration. In other words, all calculations have been performed in no-swirl case. As the piston goes up, its velocity changes with crank angle a. Because of this reason, the piston velocity has been calculated respect to the crank angle at every step of the calculation. The rotatio nal speed of the engine is chosen as its full load speed of 1200 rpm. The momentum equations of fluid is solved by using a lagrangien method, which is based on the assumption that the grids move with fluid together. In other words, the velocity of fluid is equal to the velocity of grid. To represent the boundary conditions on the wall it is assumed that the velocity of fluid perpendicular to the wall is zero, but its parallel component is considered to be equal to the parallel velocity of the nearest grid point. The parallel velocity of fluid on the center axis is taken to be equal to the parallel velocity at the nearest grid point, while the horizontal compo nent of fluid is zero. The engine used for the computation is a four- stroke, 102 mm bore, 106 mm stroke diesel engine having a 4 length ratio of the connecting rod to the crankarm and 10 compression ratio. The combustion chamber is cylindrical, having a bowl diameter to cylinder diameter of 0.5 and a ratio of depth to VII the bowl diameter of 0.52 and a top clerance of 5 mm. The injection nozzle is of a hole type with one 0.58 mm hole and an injection angle to the cylinder head of 43^. The fuel is assumed to be diesel fuel having a specific gravity of 0.850 gr/cm3. The following conditions are selected as the baseline operating conditions. The engine speed is 1200 rpm, the equivalance ratio $>t> =0.5, the tempera ture in the beginning of the compression (BDC) 340 °K, the injection timing -53. 2^, the injection duration 18-. Spray Dynamics: Since diesel combustion is strogly controlled by the details of the spray, an understanding of high- pressure sprays is central to the goal of optimizing the direct-injection diesel engines. Some computations of an evaporating spray have been performed in the past ten years [1,2], all of which neglect the break up process. These computations have included the simplified models : the evaporation rate is assumed to be the same as that for an isolated drop in a stagnant ambitient fuel vapor atmosphere and the slip between the droplets and the surrounding gas is either neclected or modelled by the drag law for steady flow past â nonevaporating sphere [3]. Cliffe and Lever [4] made a compherensive numerical study of the dynamics and energetics of an individual evaporating droplet, and new improved cor relations for drag and evaporation were suggested on the basis of their numerical results. In this study, a spray model that contains droplets having different numbers and the diameters in group, is employed. To calculate the number of droplets injected into the combustion chamber per unit time in the injection timing, Rosin-Rammler distrubition function is used. The droplets in the spray have been represented by five different diameter groups, each of which includes different number of droplets. It is assumed that the droplets come into the flow field at the distance from the nozzle of 1 cm, where the spray is supposed to form completely. It is also assumed that the droplets don't break VIII up or coalasce. So the total number of the droplets at during the injection timing is conserved. The droplets are moved on the five different streamline, which are selected arbitrarly at the place where the spray is formed. The only force acting on the droplets as they are moving through the flow field, is supposed to be aerodynamic drag force. So the momentum equations for each group of droplets are based on that assumption. Moreover, the force acting on the fluid due to the evaporation process of the droplets neglected, because of its very small quantity respect to other source terms. The evaporation coefficient (Cb) of the droplets is taken to be constant having the value of 0.79 mnrVs in the flow field. The surface temparature of the droplet injected into the flow field is supposed to be constant and equal to the evoparation temperature of the fuel. So the droplets start evoparating as soon as they come into the flow field and the heat energy that they transfer from their surrounding air is equal to the heat of vaporization of fuel. The momentum equations of the droplets are solved by using Taylor's series expansion with small time increments being considered only first order terms. To determine the places of the the droplets and the velocity of fluid there, a lagrangian interpola tion method is employed. At the final section of this study, the results obtained by both flow field and droplets model have been considered with experimental data and numerical calculations in the literature. Consequently the following conclusions have been reached: 1) The turbulance kinetic energy is decreasing in the bowl from the beginnig until the end of calcula tion. 2) It is observed that the temperature and pressure distribution over the flow field are nearly homogenous until the time that the stroke becomes 30 mm. After then, the velocity components of fluid started increasing more rapidly and as the result, the temperature and pressure differences at grid points have been increased, as the stroke is decreasing. 3) The velocity of the droplet ise first IX decreasing until the time that it is equal to the velocity of fluid. After then the droplets start to move with fluid. 4} The droplets have collected at the center axis after having followed a parabolical orbit. As the result, the theoretically predicted results of this simulation showed sufficent agreement with experimentally repoted data.