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High swirl-inducing piston bowls in small diesel engines for emission reduction
Applied Energy 88 (2011) 2355–2367
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High swirl-inducing piston bowls in small diesel engines for emission reduction B.V.V.S.U. Prasad, C.S. Sharma, T.N.C. Anand, R.V. Ravikrishna ⇑ Department of Mechanical Engineering, Indian Institute of Science, Bangalore 560012, India
a r t i c l e
i n f o
Article history: Received 13 May 2010 Received in revised form 29 October 2010 Accepted 22 December 2010 Available online 5 February 2011 Keywords: Diesel engine emissions Swirl Combustion Computational Fluid Dynamics (CFD) Piston bowl Injector sac volume
a b s t r a c t Detailed three-dimensional CFD simulations involving flow and combustion chemistry are used to study the effect of swirl induced by re-entrant piston bowl geometries on pollutant emissions from a single-cylinder diesel engine. The baseline engine configuration consists of a hemispherical piston bowl and an injector with finite sac volume. The first iteration involved using a torroidal, slightly re-entrant bowl geometry, and a sac-less injector. Pollutant emission measurements indicated a reduction in emissions with this modification. Simulations on both configurations were then conducted to understand the effect of the changes. The simulation results indicate that the selected piston bowl geometry could actually be reducing the in-cylinder swirl and turbulence and the emission reduction may be entirely due to the introduction of the sac-less injector. In-cylinder air motion was then studied in a number of combustion chamber geometries, and a geometry which produced the highest in-cylinder swirl and Turbulence Kinetic Energy (TKE) around the compression top dead centre (TDC) was identified. The optimal nature of this re-entrant piston bowl geometry is confirmed by detailed combustion simulations and emission predictions. Ó 2010 Elsevier Ltd. All rights reserved.
1. Introduction Increased pollution awareness and the consequent introduction of stringent emission norms throughout world are forcing engine manufacturers to continue to investigate for strategies for reducing emissions. During fuel injection in diesel engines, the air motion plays a very important role in fuel-air mixing, combustion and emission formation, especially around the compression TDC [1]. Along with air motion, spray characteristics such as injector-hole diameter, injection pressure, spray angle, and injection timing, also have a significant effect on diesel engine combustion. Several researchers have studied the in-cylinder processes, experimentally and computationally, to understand combustion in diesel engines as discussed in the following paragraphs. The in-cylinder air motion in diesel engines is generally characterized by swirl, squish and turbulence, which have a major impact on air–fuel mixing and combustion. Swirl motion of the air is usually generated due to the design of the intake port. A good intake port design will generate higher swirl and help to improve combustion [2]. When there is swirl in the in-cylinder air, the swirl–squish interaction produces a complex turbulent flow field at the end of compression. This interaction is much more intense in re-entrant combustion chamber geometries [3,4]. Further, intensification of swirl and turbulence are observed around TDC of compression. Around this time, most of the in-cylinder air is compressed into a ⇑ Corresponding author. Tel.: +91 80 22933226; fax: +91 80 23600648. E-mail address: [email protected] (R.V. Ravikrishna). 0306-2619/$ - see front matter Ó 2010 Elsevier Ltd. All rights reserved. doi:10.1016/j.apenergy.2010.12.068
smaller diameter combustion chamber. Thus, by conservation of angular momentum, as the radius of rotation reduces, the speed of rotation increases. Intensification of turbulence is due to the highly turbulent squish and reverse-squish motions of the air near TDC of compression. Because of these, usually two peaks in turbulence, one just before TDC and the other just after the TDC, are observed [5]. In re-entrant chambers, the intensification of swirl and turbulence are higher when compared to cylindrical chambers [6]. This leads to more efficient combustion which in turn causes higher NOX emission and less soot and HC emissions [7]. Better air–fuel mixing and combustion are possible with injectors having smaller hole diameter and higher injection pressures. Higher injection pressures produce smaller droplets which evaporate faster and mix rapidly with air [8,9]. However, wall impingement of the spray and vapor leads to flame quenching and high soot emissions. Thus, a longer spray path without wall impingement is desirable for better combustion and low emissions [10,11]. The relative position of the axes of the piston bowl and injector with respect to the cylinder axis also plays a role in in-cylinder mixture motion and combustion [12]. Injectors with finite sac volume at the tip are usually associated with large fuel droplets towards the end of injection which lead to higher HC and soot emissions [1]. In re-entrant chambers, retardation of spray can be used to control NOX emissions without much increase in soot and HC emissions. Thus, both NOX and soot can be simultaneously controlled with the use of proper combination of re-entrancy and injection timing [7,13,14]. There seems to be an optimal swirl level of in-cylinder air for minimum emissions [15]. Re-entrant lip shape and piston
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bowl radius are also observed to control the mixture distribution [16,17]. In recent years, to investigate many of the above mentioned aspects, computational tools are being significantly used [18–21]. Overall, optimization of combustion chamber geometry along with swirl and selected spray parameter is the key to reducing pollutant emissions and for better fuel economy [22,23]. Most studies in the literature concerning re-entrant chambers and injection characteristics have been conducted either on motored engines, or on large, fired engines. There are very few studies on the effect of combustion chamber geometry in medium and small engines. The present study focuses on the effect of introducing re-entrancy in the combustion chamber of a single cylinder 7.5 kW engine running at 1500 rpm. Preliminary engine modifications and the resulting exhaust emission data are presented. Detailed in-cylinder three-dimensional modeling of air motion and combustion is used to explain the trends, and is further used to arrive at an optimal bowl geometry. 2. Motivation The specifications of the engine studied are listed in Table 1. The selected engine is a naturally-aspirated, constant-speed engine. First column of Table 2 lists the emissions measured from the baseline engine, which are observed to be high. Based on published literature on diesel engines, preliminary modifications were implemented as shown in Fig. 1. The combustion chamber, which was hemispherical, was replaced with a torroidal and slightly reentrant configuration (referred as C1), since this design was shown to be effective from previous investigations in the literature [24– 26]. In addition, the original injector with a finite sac volume at the tip was replaced with a sac-less injector. The new injector has a smaller hole diameter of 0.26 mm (referred as I2). The injection timing was then delayed by 4° CA to reduce NOX emissions. After incorporating these changes, the modified engine was tested and the emission results from these tests are listed in Table 2. It is observed from Table 2 that there is a significant reduction in emissions for the modified engine except in case of CO. For the modified engine denoted by C1I2, HC emissions have significantly reduced. However, particulate matter (PM) and NOX need further reduction. Also, interestingly, CO levels have increased. Further engine modifications and testing were found to be costly and time consuming. Also, to proceed with further modifications of the engine, first there is a need to understand the effect of these changes on the in-cylinder processes, and hence on combustion and emission formation. This provides the motivation for the present investigation
Table 1 Engine specifications. Power output Speed Bore Stroke Compression ratio Swept volume Spray cone angle Max. injection pressure
7.5 kW 1500 rpm 102 mm 116 mm 16.3 948 CC 135° 380 bar
Table 2 Comparison of measured cycle average emissions after preliminary modifications. Pollutant
Baseline engine (g/(kW h))
C1I2 configuration (g/(kW h))
HC CO NOX PM
2.83 5.58 13.34 1.399
0.94 5.81 9.33 0.504
concerning numerical simulation of the in-cylinder processes. Three-dimensional Computational Fluid Dynamics (CFD) simulations are used to understand clearly the effect of changes made to the engine, on in-cylinder flow, combustion, and emission formation, and to arrive at an optimal engine configuration.
3. Modeling methodology AVL FIRE, a CFD software specially developed for in-cylinder simulations, is used in the present study to simulate the incylinder processes. It uses a cartesian coordinate system and a finite volume-based implicit discretization procedure. Unstructured meshing is utilized to mesh the complex intake manifold and cylinder geometry while taking valve motion into account. This software has been validated for engine simulations by Tatschl et al. [27–29], Priesching et al. [30], Wieser and Ennemose [31], and Suzzi et al. [32]. The SIMPLE algorithm is used for pressure velocity coupling in the solution of the flow field. The two equation k–e model has been implemented for estimating turbulence in terms of Turbulence Kinetic Energy (TKE). While the k–e model is a widely used model, it assumes isotropic turbulence. For this reason, even though it has given reasonable agreement with experiments, some researchers [6,33] have found it to underestimate turbulence during some parts of the cycle in swirling engine flows. However, the RNG k–e model was also found to underestimate the turbulence, similar to the k–e model [34]. Hence, overall, because of its simplicity and due to problems like convergence and computational time associated with other models, the k–e model has been chosen for the present study. More importantly, in the present case, the emphasis is on comparing various geometries qualitatively. Spray break-up has been modeled by the WAVE break-up model developed by Reitz and co-authors [35] which is one of the discrete droplet models. In this model, an infinitesimal axisymmetric surface displacement is imposed on the initially steady moving liquid surface which causes small axisymmetric fluctuations. It is assumed that drops are formed with a drop size proportional to the wavelength of fastest growing or most probable unstable surface wave. The maximum growth rate and its corresponding wavelength are related to pertinent properties of liquid and gas. The initial conditions required by the droplet model are calculated using the cylinder pressure, injection pressure data and analytical expressions used in the previous studies on diesel engines [36]. The Shell kinetic ignition model which was developed by Halstead et al. [37] is used to simulate the auto-ignition of diesel fuel. This model uses an eight step mechanism to predict the low temperature combustion process during ignition. For predicting the heat release, the Eddy-Dissipation Combustion Model proposed by Magnussen and Hjertager [38] is used, which can predict the heat release reasonably with low computational effort. This model assumes that the chemical reactions usually have time scales that are very short compared to the characteristic times of the turbulent transport processes. Thus, the rate of combustion is determined by the rate of intermixing on a molecular scale of eddies containing reactants and hot products. The EDC model is not suitable for predicting minor species, since detailed chemistry cannot be incorporated. In the present study, however, the emphasis is on the methodology of bowl optimization where a relative comparison of the combustion process is adequate. The Zeldovich mechanism which is based on equilibrium assumptions has been used for predictions of NOX [1]. Soot emission is predicted using Hiroyasu soot formation model [39] and Nagle-StricklandConstable soot oxidation model [40]. The computational grids used for suction and closed valve simulations are shown in Fig. 2a and b, respectively. The computation
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Piston bowl and injector tip before modification (Baseline)
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Piston bowl and injector tip after modification (C1)
Fig. 1. Modifications to the baseline engine.
Fig. 2. Computational grid and boundary surfaces.
mesh has around 190,000 cells at TDC, and 250,000 cells at BDC. Grid independence studies were also performed by comparing the results with a finer grid having 240,000 cells at TDC, and 450,000 cells at BDC. Numerical values were also compared at different locations at different crank angles throughout the simulation, and the results were found to be sufficiently grid independent. Hence, the coarser mesh having up to 250,000 cells was used for the simulations. Besides the grid, a section of the computational domain is shown, in which all the boundary surfaces used to apply various boundary conditions are labeled. The boundary conditions applied to the boundary surfaces shown in Fig. 2 are listed in Tables 3 and 4, respectively, for the suction and closed-valve duration. The boundary temperature is kept constant during the simulation
which causes higher heat transfer during compression and expansion strokes. Thus, the wall temperatures are increased during compression expansion strokes to get a better match between measured and simulated pressure curve during compression. The range for the wall temperatures is selected from previous experimental studies on diesel engines [41,42].
4. Air-motion study Simulation of in-cylinder air motion enables study of the effect of the combustion chamber geometry alone. In this study, swirl and TKE around TDC of compression are considered as important
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Table 3 Boundary conditions used for suction simulation. Boundary
Momentum boundary condition
Thermal boundary condition
Intake port surface Inlet
Wall Pressure = 100 kPa TKE = 1 m2/s2 Turbulence length scale = 0.02 m Wall Wall Wall with boundary velocity Wall with boundary velocity
350 K 315 K
Head Cylinder wall Piston top Valve surface
400 K 425 K 450 K 450 K
Table 4 Boundary conditions for simulating closed-valve part of cycle. Boundary
Momentum boundary condition
Thermal boundary condition
Head Cylinder wall Piston top Surface of compensation volume
Wall Wall Wall with boundary velocity Wall with boundary velocity
530 K 500 K 550 K Adiabatic wall
parameters. It must be emphasized here that accuracy of initial conditions for the in-cylinder air flow computations of compression (flow field at the end of suction) plays a very important role on predictions at TDC of compression [43]. Thus, detailed suction stroke simulations were conducted considering the complete geometrical details of the intake manifold and cylinder. In the present simulations, the suction stroke starts at 360° CA and ends at 148° CA (considering compression TDC as 0° CA). At the beginning of suction, a valve-overlap of 8° CA has not been considered since the opening of valves during that period is only 0.3 mm or less. The pressure at the inlet was specified as per the test conditions. For all other boundaries, ‘Wall’ boundary conditions with appropriate temperatures were specified. At the beginning of suction stroke, swirl ratio was assumed to be zero. As the angular momentum induced by intake flow is much higher compared to the angular momentum of residual air existing at the beginning of suction, initial swirl ratio does not influence the results significantly. During suction, the variation of average swirl number (ratio
of swirling speed of air to engine speed) in the cylinder was estimated for both the baseline and the C1 configuration. It was observed that there is hardly any difference between the swirl numbers of the two configurations. In fact, towards the end of suction stroke, the difference in swirl number of both the geometries is negligible. The same trend was observed for the TKE variation also. Thus, it can be concluded that the flow field at the end of suction is independent of the combustion chamber geometry. This is an important and useful observation since it enables using the same initial condition for any new bowl geometry during simulation of compression stroke. The value of volumetric efficiency predicted from simulations is 86%, which is in close comparison to that from experimental measurement (83%). This ensures that the mass inducted during suction stroke predicted by simulations is in close proximity of the experimental value. The closed-valve part of the cycle starts at 148° CA and ends at 148° CA with respect to TDC. During simulation of the closed-valve duration, the intake port was removed from the computational domain to reduce computational effort. To simplify the mesh generation procedure and to improve the quality of the mesh for compression stroke calculations, grid generation is achieved by revolution of a surface mesh. In this method, it is not possible to generate a mesh for non-axisymmetric parts. Thus, a small volume has been added near the piston crown as shown in Fig. 2 to keep the compression ratio unaltered. The result at the end of suction was taken as the initial condition and calculations for the closed-valve duration were continued. The variation of swirl number during the closed-valve part of the cycle is shown in Fig. 3. Near TDC, the intensification in swirl for the C1 configuration is less than that for the baseline configuration. This result is contrary to what was expected. To identify the reasons for this, the velocity distribution in the central plane of the cylinder is shown in Fig 4. Just before TDC, the maximum velocity in C1 is more than that of baseline. But, at TDC and after TDC, the maximum velocity in the C1 configuration is less than that in the baseline case. There is an increase of 13% in the surface area of the bowl largely due to the presence of the central projection. This might be leading to increased skin friction and negating the advantage obtained by introducing re-entrancy. Velocity distributions in both configurations at different times around TDC are shown in Fig 4. From the figure, it can be observed that the extent of the high velocity squish flow vortex is almost the same in both configurations. Due to the swirl–squish interaction, two counter-rotating torroidal vortices are formed in the
3.5 Baseline C1
Swirl number
3
2.5
2
1.5
1 -150
-120
-90
-60
-30
0
30
60
90
Crank angle Fig. 3. Variation of swirl number during closed-valve part of cycle.
120
150
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Baseline
C1 configuration
Baseline
C1 configuration Fig. 4. Distribution of velocity in the central plane of cylinder and velocity vectors in the central plane of cylinder at TDC.
35 Baseline C1
30
2
2
TKE (m /s )
25 20 15 10 5 0 -150
-120
-90
-60
-30
0
30
60
90
Crank angle Fig. 5. Variation of TKE during closed-valve part of cycle.
120
150
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Baseline
C1 configuration
Fig. 6. Distribution of TKE in the central plane of cylinder.
combustion chamber. The C1 configuration has an entry diameter less than that of the baseline configuration. Thus, the squish flow is expected to penetrate more towards the center of bowl. In the present study, however, the squish flow penetration for both C1 and baseline geometries is similar. In case of the baseline configuration, the larger vortex spreads up to the center of the cylinder. But, in the C1 configuration, the large central projection causes the larger vortex to move outwards, which in turn displaces the smaller vortex outwards. Thus, the central projection in case of C1 is obstructs the squish flow from entering deep into the combustion chamber. These reasons explain the lower swirl number in case of C1. In Fig. 5, the variation of cylinder-averaged TKE during compression and expansion strokes is shown. From the figure, it is observed that TKE is also consistently less in case of C1 as compared to the baseline. In previous studies on re-entrant chambers, it was found that TKE peaks twice, once corresponding to squish flow and the second time due to reverse-squish [5]. In both cases, intensification of TKE corresponding to squish flow is marginal and the second peak corresponding to the reverse-squish is almost negligible. This implies that the re-entrancy introduced in C1 is too small to take the advantage of the strong squish and reverse-squish flows. The TKE distribution in the central plane of the cylinder is shown in Fig. 6. Before TDC and at TDC, the maximum values of TKE are slightly higher in C1 as compared to the baseline configuration. The improvement is marginal and the region with the maximum value of TKE is significantly smaller in C1. This explains the reason for the lower average TKE values in C1. Observing at the TKE distribution at 20° CA and TDC, it seems that the central projection is damping turbulence in the central region of the cylinder. The TKE distribution at 20° CA shows that the sharp edge in the baseline configuration is causing higher turbulence during reverse-squish flow. Thus, from air motion analysis, it can be concluded that the combustion chamber geometry selected from literature is actually negatively impacting the flow field in the present case. 5. Combustion study Simulation of spray and combustion provides more information about in-cylinder processes such as spray–air interaction, wall impingement, mixing, combustion, and emission formation. The
Table 5 Models for spray simulation, model constants and input parameters. Parameter/model
Value/reference
Mass of fuel injected Initial droplet diameter Number of parcels injected Droplet break-up model
36 mg/stroke 0.28 mm for I1 and 0.26 mm for I2 37,000 Wave break-up model [35] (C1 = 0.61, C2 = 13, C3 = 1) Enable [44] Dukowicz [45] Walljet1 [46]
Turbulent dispersion model Droplet evaporation Wall interaction
models utilized in the present case as discussed in Section 3 are relatively simpler when compared to the complexity of actual processes. Thus, spray, ignition, combustion and pollutant formation models are provided with constants which are to be calibrated against at least one set of experimental data [29]. Spray mass flow rate and droplet size were calculated using needle lift diagram, injection pressure and cylinder pressure data. Instantaneous droplet sizes have been calculated according to previous studies [36] which have considered cavitation in nozzle flows. Also, the effect of sac (in case of baseline injector) has been simulated by injecting droplets with very low velocity at the end of injection. The droplets injected at a very low velocity remain in the immediate vicinity of the nozzle tip and thus undergo slow evaporation and combustion similar to that of fuel in the sac. In the present case, cylinder pressure and emission data obtained from engine tests and the spray penetration data obtained from experimental correlations have been used to calibrate model constants. The details of the spray parameters utilized in the simulations are listed in Table 5 (In Table 5 I1 and I2 represent two injectors with different hole diameters). The initial part of the spray penetration curve calculated from correlations [1] was used for calibrating the spray break-up model. By varying the break-up model constant C_2, (generally in the range of 5–60) breakup can be advanced or delayed, and thus the penetration can be reduced or increased. By trial and error, a value of 13 was found to give a good match for the penetration. The ignition model constant was adjusted to a value of 1.0e+6 such that the starting point of ignition in the pressure curve matches
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7E+06 Experimental 6E+06
Predicted
Pressure (Pa)
5E+06 4E+06 3E+06 2E+06 1E+06 0E+00 -100
-75
-50
-25
0
25
50
75
100
125
150
Crank Angle Fig. 7. Predicted pressure variation after model calibration in comparison with experimental data.
Table 6 Predicted emission values after model calibrations for the baseline configuration. NOX (g/(kW h))
Soot (g/(kW h))
Predicted
Experimental
Predicted
Experimental
10.11
10.06
0.52
0.52
with the experimental value. The combustion model constant whose value is in the range of 3–25, generally, was calibrated to a value of 7 by matching the pressure curve with the experimental value at 100% load. The predicted pressure curve after model calibration is compared with experimental data in Fig. 7. The prediction of pressure is in sufficiently good agreement with experimental data. The slight discrepancy observed after TDC may be due to the limitations of the model. Pre-exponential factors for NOX and soot reaction rates are adjusted for baseline case to match the predicted values against experimental values. Emission value predictions in comparison with the experimental values are shown in Table 6. Once the model constants are calibrated, the values are kept the same for all further simulations. Emission predictions of the C1 configuration with original and new injector, denoted by C1I1 and C1I2 respectively, along with the baseline configuration are listed in Table 7. Comparing the emission values of baseline and C1I1, there is a significant reduction in NOX and increase in soot. It is to be noted that the C1I1 configuration was not experimentally studied, but is numerically simulated to aid in comparison, especially to isolate the effect of bowl geometry alone. The comparison indicates that for the C1I1 configuration, combustion can be poorer than the baseline case. Figs. 8 and 9 show the in-cylinder pressure and Heat Release Rate (HRR) for the baseline, C1I1 and C1I2 configurations, respectively. It is observed that the difference in combustion and hence in emissions is mainly due to the new injector. Thus, optimization of com-
bustion chamber for higher swirl and TKE may help in improving combustion in the present case. For the C1I2 case, the predicted NOX emission 8.9 g/(kW h) matches well with the experimental value of 9.3 g/(kW h) (Table 2). Table 7 also indicates the effect of injector change comparing the C1I1 and C1I2 cases (reduction in soot and increase in NOX). This confirms the earlier observation of significant improvement in combustion due to the change of injector. The new injector, having smaller hole diameters, leads to faster evaporation and better combustion. The predicted pressure variation in Fig. 8 also indicates a higher peak pressure in case of C1I2 than for the C1I1 case. The heat release rate variation shown in Fig. 9 shows that the ignition delay reduces significantly and makes combustion to happen around TDC when in-cylinder temperatures are higher, and this leads to efficient combustion and lower soot. This explains that the reduction in emissions observed experimentally is mainly due to change of injector rather than change in the combustion chamber geometry. It is a very interesting observation which indicates that there is further scope for reduction in emissions through optimization of combustion chamber geometry. 6. Parametric optimization From the above analysis, it was concluded that there is scope to modify the combustion chamber geometry to further improve combustion and reduce emissions. The above investigations also show that a combustion chamber geometry which improves swirl and TKE intensification around TDC, will lead to more efficient combustion. Thus, the air-motion study was used for selection of combustion chamber geometry and then the optimal injection timing was selected using a combustion study. Simulations for different geometrical variations of the combustion chamber were commenced at the end of suction stroke. Initial conditions were taken from flow field solution at the end of suction for the baseline
Table 7 Predicted emission values for different combinations of geometry and injector. Bowl geometry
Injector
Baseline C1 C1
I1 I1 I2
Injection Timing (w.r.t. TDC) 12.4 12.4 15.6
NOX (g/(kW h))
Soot (g/(kW h))
Ignition delay (CA)
10.11 6.73 8.89
0.52 2.21 1.13
10.2 10.2 8.4
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7E+06
Baseline 6E+06
C1I1 C1I2
Pressure (Pa)
5E+06 4E+06 3E+06 2E+06 1E+06 0E+00 -150
-120
-90
-60
-30
0
30
60
90
120
150
Crank Angle Fig. 8. Variation of in-cylinder pressure for various configurations.
160 Baseline
140
C1I1 C1I2
HRR (J/deg)
120 100 80 60 40 20 0 -20
-10
0
10
20
30
40
50
60
70
Crank Angle Fig. 9. Variation of heat release rate for various configurations.
4.5 Baseline
Swirl number
4 3.5
C1
3
C2
2.5 C3
2 C4
1.5 C5
1 C6
0.5
C7
0
-150 -120
-90
-60
-30
0
30
60
90
120
150
Crank Angle Fig. 10. Swirl number variation during closed-valve duration for various bowl geometries.
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35 Baseline
30 C1
25
TKE (m2/s2)
C2
20 C3
15
C4
10
C5
5
C6
0 -150
C7
-120
-90
-60
-30
0
30
60
90
120
150
Crank angle Fig. 11. TKE variation during closed-valve duration for various bowl geometries.
Fig. 12. Predicted velocity and TKE distribution in the central plane of the C3 configuration.
geometry. For all variants of the combustion chamber geometry, bowl volume and compression ratio of engine were kept equal to that of baseline geometry. The variation of swirl number in the selected geometries along with the baseline and C1 is shown in Fig. 10. C3 configuration shows the highest swirl intensification. The peak swirl number for the C3 case is around 35% higher than that of the baseline value. Variation of TKE for all the selected geometries during the closed-valve duration is shown in Fig. 11. It is observed that, once again, the C3 geometry is better compared to the other geometries. The peak TKE for the C3 case is as much as 50% higher when compared to the corresponding baseline value. The distribution of velocity and TKE in the central plane of the C3 case is shown in Fig. 12. (Corresponding figures for baseline configuration are shown in Fig. 4 (velocity) and Fig. 6 (TKE).) From the figure, it is observed that in addition to the average values, the local values are also fairly higher for the C3 case when compared to
the baseline configuration. In order to confirm that the higher swirl and TKE intensifications for C3 case would lead to improved combustion, combustion simulations were performed for the C3 case and the two nearest competing piston bowl geometries C2 and C7. The results from these simulations are shown in terms of p–h curves in Fig. 13. The configuration C3 shows higher peak cylinder-averaged pressure and larger area under the p–h curve (thus higher power) as compared to C2 and C7. The corresponding emission predictions for C3, C2 and C7 are shown in Table 8. Lower soot and higher NOX emissions are observed for C3 as compared to C2 and C7. It is to be noted that although the soot emissions are slightly lower for the C2 case as compared to the C3 case, the difference is very small. Overall, comparing peak cylinder pressures, C2 is observed to be inferior compared to C3. For C7, soot emissions are far above the corresponding levels for C3. Thus, the bowl geometry corresponding to the C3 is selected as the optimum,
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Fig. 13. Pressure variation in C3I2, C2I2 and C7I2 configurations.
and is taken up for further studies pertaining to injection timing optimization. Combustion simulations were carried out for the C3 geometry with the new injector or C3I2 case, with corresponding injection timing (15.6° CA BTDC). Table 8 shows the emission predictions. As expected, there is a significant reduction in soot and increase of NOX emissions. Also, there is a reduction in ignition delay by 1.8° CA. This indicates considerable improvement in combustion. Higher swirl and TKE causes faster breakup and evaporation of
Table 8 Predicted emissions for C3, C2 and C7 configurations. Injection timing (w.r.t. TDC) 15.6 15.6 15.6
NOX (g/(kW h))
Soot (g/(kW h))
18.66 17.49 10.06
0.06 0.05 2.54
8E+06
Baseline, -12.4 CA 7E+06
C3I2, -15.6 CA C3I2, -5.6 CA
6E+06
Pressure (Pa)
C3I2 C2I2 C7I2
C3I2, -10.6 CA C3I2, -8.6 CA
5E+06 4E+06 3E+06 2E+06 1E+06 0E+00 -150
-120
-90
-60
-30
0
30
60
90
120
Crank Angle Fig. 14. Pressure variation in the C3I2 configuration for different injection timings.
150
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12.4 15.6 10.6 8.6 5.6
NOX (g/(kW h))
Soot (g/(kW h))
Ignition delay (CA)
10.11 18.66 9.44 7.36 5.35
0.52 0.06 0.04 0.08 0.52
10.2 8.4 8.2 7.9 8.4
10o ASoI
15o ASoI
20o ASoI
25o ASoI
30o ASoI ASoI – After Start of Injection
Fig. 15. Spray evolution in terms of droplet locations and diameters (in meters) and equivalence ratio distribution in central plane of cylinder at various crank angles.
liquid droplets resulting in improved air–fuel mixing and better combustion. Fig. 14 shows comparison of predicted pressure curve of C3I2 with 15.6° injection timing with pressure curve of baseline engine. Improvement in combustion causes significant increase in peak cylinder pressure which enables retardation of injection without affecting fuel economy. Various injection timings and the corresponding emission predictions for C3I2 are listed in Table 9. The NOX-soot trade off can be clearly observed from the table. It can also be seen from the table that ignition delay has reduced as the injection timing approaches TDC. At 5.6° CA which is a highly retarded case, the ignition delay has increased again. In this case, ignition occurs after TDC. A slight reduction in temperature and pressure after TDC, and flow destruction after TDC might be the reason for this. An injection timing of 8.6° CA seems to be an optimal value at which there is a reduction of 27% in NOX, and 85% in soot. From Fig. 14, it is also observed that the pressure curve is almost matching with that of the baseline case
indicating that there is no work loss with the proposed 8.6° CA injection timing. Thus, the bowl geometry denoted by C3 with an injection timing of 8.6° CA BTDC shows significant potential for further reduction of emissions from the selected engine. Spray droplet distributions and equivalence ratios are compared for baseline and C3 configurations in Fig. 15. Lower equivalence ratios are observed for the C3 configuration which implies improved air–fuel mixing compared to baseline configuration. This is a reason for the lower soot levels for the C3 configuration. Finally, the effect of a 6-hole nozzle configuration with the same fuel injection pressure was also studied. In order to keep the same amount of fuel delivered in both cases, the nozzle hole diameter was suitably reduced. This new configuration is denoted by C3I3. Simulations were conducted for this configuration at various injection timings of 12.4°, 8.6°, 5.6°, and 2.6° CA. For a timing of 5.6° CA which gives the best soot-NOX trade-off for this configuration, the predicted NOX level is 9.59 g/kW h, and the soot level is
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0.22 g/kW h. These values are higher than those corresponding to the C3I2 configuration with the 8.6° CA timing, thus indicating that the C3I3 configuration is inferior. This may be attributed to the lower penetration by the fuel droplets for the C3I3 case as compared to the C3I2 case. Thus, among all cases studied, the C3I2 configuration is observed to optimum leading to significant reduction in both NOX and soot levels compared to the baseline configuration. 7. Conclusions The present study concerns the effect of swirl induced by re-entrant piston bowl geometries on emissions in a diesel engine, and specifically focuses on a single cylinder, 7.5 kW constant-speed engine. The emission test results of two configurations of the selected engine are reported. The second configuration which has a slightly re-entrant combustion chamber and a sac-less injector was found to yield lower emissions. In order to understand the effect of re-entrancy and injector change on emissions, detailed, threedimensional CFD simulations of the in-cylinder processes were conducted. The effect of chamber geometry and injector change was studied using unfired and fired simulations. Simulation of closedvalve part of the cycle in the two configurations revealed that average swirl and turbulence levels around TDC of compression were higher for the baseline case than for the modified geometry. Increased surface area, presence of a large central projection and insufficient re-entrancy were identified as the reasons for the modified geometry yielding poor results. Combustion simulations revealed that the reduction in emissions observed during experiments is mainly due to the change in the injector rather than change in the piston bowl geometry, thus indicating scope for optimization of bowl geometry. Several piston bowl geometries with varying levels of re-entrancy and different heights of central projections were simulated. A highly re-entrant piston bowl and without a central projection was found to be the best for swirl and TKE intensification around TDC. Combustion simulations were carried out using the selected geometry and injection timings were optimized to keep NOX levels below those of the baseline case. An injection timing of 8.6° CA BTDC was found to be optimum since it led to a 27% reduction in NOX emissions and 85% reduction in soot levels as compared to the baseline configuration. Acknowledgements The authors would like to acknowledge the information and data provided by Prof. M.V. Narasimhan, Mr. J. Kalvani and the Indian Diesel Engines Manufacturers Association (IDEMA). References [1] Heywood JB. International combustion engine fundamentals. New York: McGraw-Hill Book Company; 1988. [2] Brandl F, Reverencic, Cartellieri W, Dent JC. Turbulent air flow in the combustion bowl of a DI diesel engine and its effect on engine performance. SAE Paper 790040. [3] Arcoumannis C, Bicen AF, Whitelaw JH. Squish and swirl–squish interaction in motored model engines. ASME J Fluid Mech 1993:105–12. [4] Arcoumannis C, Bicen AF, Vafidis C, Whitelaw JH. Three-dimensional flow field in four-stroke model engines. SAE Paper 841360. [5] Kondoh T, Fukumoto A, Ohsawa K, Ohkubo Y. An assessment of a multidimensional numerical method to predict the flow in internal combustion engines. SAE Paper 850500. [6] Payri F, Benajes J, Margeo X, Gil A. CFD modeling of the in-cylinder flow in direct-injection diesel engine. Comput Fluids 2004;33:995–1021. [7] Saito T, Daisho Y, Uchida N, Ikeya N. Effects of combustion chamber geometry on diesel combustion. SAE Paper 861186. [8] Di Giorgio F, Laforgia V. Investigation of drop size distribution in the spray of a five-hole, V.C.O. nozzle at high feeding pressure. SAE Paper 950087. [9] Benajes J, Pastor JV, Payri R, Plazas AH. Analysis of the influence of diesel nozzle geometry in the injection rate characteristic. J Fluids Eng, ASME 2004;126:63–71.
[10] Meintast U, Staudt M, Reichelt L, Renz U, Sommerhoff FA. Analysis of spray/ wall interaction under diesel engine conditions. SAE Paper 2000-01-0272. [11] Ikagami M. Role of flows and turbulent mixing in combustion and pollutant formation in diesel engines. In: Proceedings of international symposium COMODIA-90; 1990. p. 49–58. [12] Zhengbai L, Xinqun G. Investigation of effects of piston bowl and fuel injector offsets on combustion and emissions in DI diesel engines. SAE Paper 2002-011748. [13] De Risi A, Manieri DF, Laforgia D. A theoretical investigation on the effects of combustion chamber geometry and engine speed on soot and NOX emissions. ASME-ICE, Paper No. 99-ICE-207, 33-1; 1999. p. 51–9. [14] De Risi A, Donateo T, Laforgia D. Optimization of the combustion chamber of direct injection diesel engine. SAE Paper 2003-01-1064. [15] McCracken ME, Abraham J. Swirl–spray interactions in a diesel engine. SAE Paper 2001-01-0996. [16] Zhu Y, Zhao H, Melas DA, Ladommatos N. Computational study of the effects of the re-entrant lip shape and toroidal radii of piston bowl on a HSDI diesel engine’s performance and emissions. SAE Paper 2004-01-0118. [17] Tomoya A, Kenichi N, Takatoshi A, Hideki M, Masahiro K, Seiichi H. Development of new 2.2-l turbocharged diesel engine for the EURO-IV standards. SAE Paper 2004-01-1316. [18] Rakopoulos CD, Kosmadakis GM, Pariotis EG. Critical evaluation of current heat transfer models used in CFD in-cylinder engine simulations and establishment of a comprehensive wall-function formulation. Appl Energy 2010;87(5):1612–30. [19] Rakopoulos CD, Kosmadakis GM, Pariotis EG. Investigation of piston bowl geometry and speed effects in a motored HSDI diesel engine using a CFD against a quasi-dimensional model. Energy Convers Manage 2010;51(3): 470–84. [20] Jayashankara B, Ganesan V. Effect of fuel injection timing and intake pressure on the performance of a DI diesel engine – a parametric study using CFD. Energy Convers Manage 2010;51(10):1835–48. [21] Rakopoulos CD, Kosmadakis GM, Pariotis EG. Evaluation of a new computational fluid dynamics model for internal combustion engines using hydrogen under motoring conditions. Energy 2009;34(12):2158–66. [22] Shahrokh H, Gerhard R, Anthony C, Michael E, Herbert M. Application of CFD modeling in combustion bowl assessment of diesel engines using DoE methodology. SAE Paper 2006-01-3330. [23] Jaeman L, Kyoungdoug M. The effects of spray angle and piston bowl shape on diesel engine soot emissions using 3-D CFD simulation. SAE Paper 2005-012117. [24] Narasimhan MV. Project report ‘‘Project sponsored by IDEMA (Indian Diesel Engines Manufacturers Association) at MICO/IISc (Indian Institute of Science) to meet central pollution control board norms 2004 in genset engines 619 kW.’’; 2003. [25] Zhang L, Ueda T, Takatsuki T, Yokota K. A study of the effects of chamber geometries on flame behavior in a DI diesel engine. SAE Paper 952515. [26] Lisbona MG, Olmo L, Rindon G. Analysis of the effect of combustion bowl geometry of a DI diesel engine on efficiency and emissions. In: Proceedings of THIESEL 2002; 2002. p. 279–93. [27] Tatschl R, Pachler K, Winklhofer E. A comprehensive DI diesel combustion model for multidimensional engine simulation. In: Proceedings of the fourth international symposium COMODIA 98; 1998. p. 141–8. [28] Tatschl R, Wiesler B, Alajbegovic A, Kunsberg Sarre CV. Advanced 3D fluid dynamic simulation for diesel engines. In: Proceedings of THIESEL 2000; 2000. p. 113–21. [29] Tatschl R, Gabriel HP, Priesching P. FIRE-A generic CFD platform for DI diesel engine mixture formation and combustion simulation. International multidimensional modeling user’s group meeting at SAE Congress, Detroit; 2001. [30] Priesching P, Kunsberg Sarre CV, Gabriel HP, Tatschl R. A comprehensive CFD workflow for DI diesel engine analysis and optimization. JSAE Spring Convention, Paper No. 20015353, 2001. [31] Wieser KJ, Ennemose A. 3D-CFD diesel combustion and accurate heat transfer modeling for diesel engines. In: Proceedings of THIESEL 2002; 2002. p. 445–54. [32] Suzzi D, Berg EV, Pastor JV, Bianchi GM, Tatschl R. Simulation of primary break-up of diesel jets by a hybrid method combining volume of fluidcalculations and the classical discrete droplet modeling rate approach with a 3D computational fluid dynamics code. ILASS-Europe zone, Nottingham; 2004. [33] Beard P, Mokaddem K, Baritaud T. Measurement and modeling of the flowfield in a DI diesel engine: effect of piston bowl shape and engine speed. SAE Paper 982587; 1998. [34] Miles P, Megerle M, Hammer J, Nagel Z, Reitz RD, Sick V. Late-cycle turbulence generation in swirl-supported, direct injection diesel engines. SAE Paper 200201-0891; 2002. [35] Lui AB, Mather D, Reitz RD. Modeling the effect of drop drag and break of fuel sprays. SAE Paper 930072; 1993. [36] Kuensberg CV, Kong SC, Reitz RD. Modeling the effects of injector nozzle geometry on diesel spray. SAE Paper 1999-01-0912; 1999. [37] Halstead M, Kirsch L, Quinn C. The autoignition of hydrocarbon fuel at high temperatures and pressures – fitting of a mathematical model. Combust Flame 1977;30:45–60. [38] Magnussen BF, Hjertager BH. On mathematical modeling of turbulent combustion with special emphasis on soot formation and combustion. In: Sixteenth international symposium on combustion; 1977. p. 719– 29.
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Contents lists available at ScienceDirect
Applied Energy journal homepage: www.elsevier.com/locate/apenergy
High swirl-inducing piston bowls in small diesel engines for emission reduction B.V.V.S.U. Prasad, C.S. Sharma, T.N.C. Anand, R.V. Ravikrishna ⇑ Department of Mechanical Engineering, Indian Institute of Science, Bangalore 560012, India
a r t i c l e
i n f o
Article history: Received 13 May 2010 Received in revised form 29 October 2010 Accepted 22 December 2010 Available online 5 February 2011 Keywords: Diesel engine emissions Swirl Combustion Computational Fluid Dynamics (CFD) Piston bowl Injector sac volume
a b s t r a c t Detailed three-dimensional CFD simulations involving flow and combustion chemistry are used to study the effect of swirl induced by re-entrant piston bowl geometries on pollutant emissions from a single-cylinder diesel engine. The baseline engine configuration consists of a hemispherical piston bowl and an injector with finite sac volume. The first iteration involved using a torroidal, slightly re-entrant bowl geometry, and a sac-less injector. Pollutant emission measurements indicated a reduction in emissions with this modification. Simulations on both configurations were then conducted to understand the effect of the changes. The simulation results indicate that the selected piston bowl geometry could actually be reducing the in-cylinder swirl and turbulence and the emission reduction may be entirely due to the introduction of the sac-less injector. In-cylinder air motion was then studied in a number of combustion chamber geometries, and a geometry which produced the highest in-cylinder swirl and Turbulence Kinetic Energy (TKE) around the compression top dead centre (TDC) was identified. The optimal nature of this re-entrant piston bowl geometry is confirmed by detailed combustion simulations and emission predictions. Ó 2010 Elsevier Ltd. All rights reserved.
1. Introduction Increased pollution awareness and the consequent introduction of stringent emission norms throughout world are forcing engine manufacturers to continue to investigate for strategies for reducing emissions. During fuel injection in diesel engines, the air motion plays a very important role in fuel-air mixing, combustion and emission formation, especially around the compression TDC [1]. Along with air motion, spray characteristics such as injector-hole diameter, injection pressure, spray angle, and injection timing, also have a significant effect on diesel engine combustion. Several researchers have studied the in-cylinder processes, experimentally and computationally, to understand combustion in diesel engines as discussed in the following paragraphs. The in-cylinder air motion in diesel engines is generally characterized by swirl, squish and turbulence, which have a major impact on air–fuel mixing and combustion. Swirl motion of the air is usually generated due to the design of the intake port. A good intake port design will generate higher swirl and help to improve combustion [2]. When there is swirl in the in-cylinder air, the swirl–squish interaction produces a complex turbulent flow field at the end of compression. This interaction is much more intense in re-entrant combustion chamber geometries [3,4]. Further, intensification of swirl and turbulence are observed around TDC of compression. Around this time, most of the in-cylinder air is compressed into a ⇑ Corresponding author. Tel.: +91 80 22933226; fax: +91 80 23600648. E-mail address: [email protected] (R.V. Ravikrishna). 0306-2619/$ - see front matter Ó 2010 Elsevier Ltd. All rights reserved. doi:10.1016/j.apenergy.2010.12.068
smaller diameter combustion chamber. Thus, by conservation of angular momentum, as the radius of rotation reduces, the speed of rotation increases. Intensification of turbulence is due to the highly turbulent squish and reverse-squish motions of the air near TDC of compression. Because of these, usually two peaks in turbulence, one just before TDC and the other just after the TDC, are observed [5]. In re-entrant chambers, the intensification of swirl and turbulence are higher when compared to cylindrical chambers [6]. This leads to more efficient combustion which in turn causes higher NOX emission and less soot and HC emissions [7]. Better air–fuel mixing and combustion are possible with injectors having smaller hole diameter and higher injection pressures. Higher injection pressures produce smaller droplets which evaporate faster and mix rapidly with air [8,9]. However, wall impingement of the spray and vapor leads to flame quenching and high soot emissions. Thus, a longer spray path without wall impingement is desirable for better combustion and low emissions [10,11]. The relative position of the axes of the piston bowl and injector with respect to the cylinder axis also plays a role in in-cylinder mixture motion and combustion [12]. Injectors with finite sac volume at the tip are usually associated with large fuel droplets towards the end of injection which lead to higher HC and soot emissions [1]. In re-entrant chambers, retardation of spray can be used to control NOX emissions without much increase in soot and HC emissions. Thus, both NOX and soot can be simultaneously controlled with the use of proper combination of re-entrancy and injection timing [7,13,14]. There seems to be an optimal swirl level of in-cylinder air for minimum emissions [15]. Re-entrant lip shape and piston
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bowl radius are also observed to control the mixture distribution [16,17]. In recent years, to investigate many of the above mentioned aspects, computational tools are being significantly used [18–21]. Overall, optimization of combustion chamber geometry along with swirl and selected spray parameter is the key to reducing pollutant emissions and for better fuel economy [22,23]. Most studies in the literature concerning re-entrant chambers and injection characteristics have been conducted either on motored engines, or on large, fired engines. There are very few studies on the effect of combustion chamber geometry in medium and small engines. The present study focuses on the effect of introducing re-entrancy in the combustion chamber of a single cylinder 7.5 kW engine running at 1500 rpm. Preliminary engine modifications and the resulting exhaust emission data are presented. Detailed in-cylinder three-dimensional modeling of air motion and combustion is used to explain the trends, and is further used to arrive at an optimal bowl geometry. 2. Motivation The specifications of the engine studied are listed in Table 1. The selected engine is a naturally-aspirated, constant-speed engine. First column of Table 2 lists the emissions measured from the baseline engine, which are observed to be high. Based on published literature on diesel engines, preliminary modifications were implemented as shown in Fig. 1. The combustion chamber, which was hemispherical, was replaced with a torroidal and slightly reentrant configuration (referred as C1), since this design was shown to be effective from previous investigations in the literature [24– 26]. In addition, the original injector with a finite sac volume at the tip was replaced with a sac-less injector. The new injector has a smaller hole diameter of 0.26 mm (referred as I2). The injection timing was then delayed by 4° CA to reduce NOX emissions. After incorporating these changes, the modified engine was tested and the emission results from these tests are listed in Table 2. It is observed from Table 2 that there is a significant reduction in emissions for the modified engine except in case of CO. For the modified engine denoted by C1I2, HC emissions have significantly reduced. However, particulate matter (PM) and NOX need further reduction. Also, interestingly, CO levels have increased. Further engine modifications and testing were found to be costly and time consuming. Also, to proceed with further modifications of the engine, first there is a need to understand the effect of these changes on the in-cylinder processes, and hence on combustion and emission formation. This provides the motivation for the present investigation
Table 1 Engine specifications. Power output Speed Bore Stroke Compression ratio Swept volume Spray cone angle Max. injection pressure
7.5 kW 1500 rpm 102 mm 116 mm 16.3 948 CC 135° 380 bar
Table 2 Comparison of measured cycle average emissions after preliminary modifications. Pollutant
Baseline engine (g/(kW h))
C1I2 configuration (g/(kW h))
HC CO NOX PM
2.83 5.58 13.34 1.399
0.94 5.81 9.33 0.504
concerning numerical simulation of the in-cylinder processes. Three-dimensional Computational Fluid Dynamics (CFD) simulations are used to understand clearly the effect of changes made to the engine, on in-cylinder flow, combustion, and emission formation, and to arrive at an optimal engine configuration.
3. Modeling methodology AVL FIRE, a CFD software specially developed for in-cylinder simulations, is used in the present study to simulate the incylinder processes. It uses a cartesian coordinate system and a finite volume-based implicit discretization procedure. Unstructured meshing is utilized to mesh the complex intake manifold and cylinder geometry while taking valve motion into account. This software has been validated for engine simulations by Tatschl et al. [27–29], Priesching et al. [30], Wieser and Ennemose [31], and Suzzi et al. [32]. The SIMPLE algorithm is used for pressure velocity coupling in the solution of the flow field. The two equation k–e model has been implemented for estimating turbulence in terms of Turbulence Kinetic Energy (TKE). While the k–e model is a widely used model, it assumes isotropic turbulence. For this reason, even though it has given reasonable agreement with experiments, some researchers [6,33] have found it to underestimate turbulence during some parts of the cycle in swirling engine flows. However, the RNG k–e model was also found to underestimate the turbulence, similar to the k–e model [34]. Hence, overall, because of its simplicity and due to problems like convergence and computational time associated with other models, the k–e model has been chosen for the present study. More importantly, in the present case, the emphasis is on comparing various geometries qualitatively. Spray break-up has been modeled by the WAVE break-up model developed by Reitz and co-authors [35] which is one of the discrete droplet models. In this model, an infinitesimal axisymmetric surface displacement is imposed on the initially steady moving liquid surface which causes small axisymmetric fluctuations. It is assumed that drops are formed with a drop size proportional to the wavelength of fastest growing or most probable unstable surface wave. The maximum growth rate and its corresponding wavelength are related to pertinent properties of liquid and gas. The initial conditions required by the droplet model are calculated using the cylinder pressure, injection pressure data and analytical expressions used in the previous studies on diesel engines [36]. The Shell kinetic ignition model which was developed by Halstead et al. [37] is used to simulate the auto-ignition of diesel fuel. This model uses an eight step mechanism to predict the low temperature combustion process during ignition. For predicting the heat release, the Eddy-Dissipation Combustion Model proposed by Magnussen and Hjertager [38] is used, which can predict the heat release reasonably with low computational effort. This model assumes that the chemical reactions usually have time scales that are very short compared to the characteristic times of the turbulent transport processes. Thus, the rate of combustion is determined by the rate of intermixing on a molecular scale of eddies containing reactants and hot products. The EDC model is not suitable for predicting minor species, since detailed chemistry cannot be incorporated. In the present study, however, the emphasis is on the methodology of bowl optimization where a relative comparison of the combustion process is adequate. The Zeldovich mechanism which is based on equilibrium assumptions has been used for predictions of NOX [1]. Soot emission is predicted using Hiroyasu soot formation model [39] and Nagle-StricklandConstable soot oxidation model [40]. The computational grids used for suction and closed valve simulations are shown in Fig. 2a and b, respectively. The computation
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Piston bowl and injector tip before modification (Baseline)
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Piston bowl and injector tip after modification (C1)
Fig. 1. Modifications to the baseline engine.
Fig. 2. Computational grid and boundary surfaces.
mesh has around 190,000 cells at TDC, and 250,000 cells at BDC. Grid independence studies were also performed by comparing the results with a finer grid having 240,000 cells at TDC, and 450,000 cells at BDC. Numerical values were also compared at different locations at different crank angles throughout the simulation, and the results were found to be sufficiently grid independent. Hence, the coarser mesh having up to 250,000 cells was used for the simulations. Besides the grid, a section of the computational domain is shown, in which all the boundary surfaces used to apply various boundary conditions are labeled. The boundary conditions applied to the boundary surfaces shown in Fig. 2 are listed in Tables 3 and 4, respectively, for the suction and closed-valve duration. The boundary temperature is kept constant during the simulation
which causes higher heat transfer during compression and expansion strokes. Thus, the wall temperatures are increased during compression expansion strokes to get a better match between measured and simulated pressure curve during compression. The range for the wall temperatures is selected from previous experimental studies on diesel engines [41,42].
4. Air-motion study Simulation of in-cylinder air motion enables study of the effect of the combustion chamber geometry alone. In this study, swirl and TKE around TDC of compression are considered as important
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Table 3 Boundary conditions used for suction simulation. Boundary
Momentum boundary condition
Thermal boundary condition
Intake port surface Inlet
Wall Pressure = 100 kPa TKE = 1 m2/s2 Turbulence length scale = 0.02 m Wall Wall Wall with boundary velocity Wall with boundary velocity
350 K 315 K
Head Cylinder wall Piston top Valve surface
400 K 425 K 450 K 450 K
Table 4 Boundary conditions for simulating closed-valve part of cycle. Boundary
Momentum boundary condition
Thermal boundary condition
Head Cylinder wall Piston top Surface of compensation volume
Wall Wall Wall with boundary velocity Wall with boundary velocity
530 K 500 K 550 K Adiabatic wall
parameters. It must be emphasized here that accuracy of initial conditions for the in-cylinder air flow computations of compression (flow field at the end of suction) plays a very important role on predictions at TDC of compression [43]. Thus, detailed suction stroke simulations were conducted considering the complete geometrical details of the intake manifold and cylinder. In the present simulations, the suction stroke starts at 360° CA and ends at 148° CA (considering compression TDC as 0° CA). At the beginning of suction, a valve-overlap of 8° CA has not been considered since the opening of valves during that period is only 0.3 mm or less. The pressure at the inlet was specified as per the test conditions. For all other boundaries, ‘Wall’ boundary conditions with appropriate temperatures were specified. At the beginning of suction stroke, swirl ratio was assumed to be zero. As the angular momentum induced by intake flow is much higher compared to the angular momentum of residual air existing at the beginning of suction, initial swirl ratio does not influence the results significantly. During suction, the variation of average swirl number (ratio
of swirling speed of air to engine speed) in the cylinder was estimated for both the baseline and the C1 configuration. It was observed that there is hardly any difference between the swirl numbers of the two configurations. In fact, towards the end of suction stroke, the difference in swirl number of both the geometries is negligible. The same trend was observed for the TKE variation also. Thus, it can be concluded that the flow field at the end of suction is independent of the combustion chamber geometry. This is an important and useful observation since it enables using the same initial condition for any new bowl geometry during simulation of compression stroke. The value of volumetric efficiency predicted from simulations is 86%, which is in close comparison to that from experimental measurement (83%). This ensures that the mass inducted during suction stroke predicted by simulations is in close proximity of the experimental value. The closed-valve part of the cycle starts at 148° CA and ends at 148° CA with respect to TDC. During simulation of the closed-valve duration, the intake port was removed from the computational domain to reduce computational effort. To simplify the mesh generation procedure and to improve the quality of the mesh for compression stroke calculations, grid generation is achieved by revolution of a surface mesh. In this method, it is not possible to generate a mesh for non-axisymmetric parts. Thus, a small volume has been added near the piston crown as shown in Fig. 2 to keep the compression ratio unaltered. The result at the end of suction was taken as the initial condition and calculations for the closed-valve duration were continued. The variation of swirl number during the closed-valve part of the cycle is shown in Fig. 3. Near TDC, the intensification in swirl for the C1 configuration is less than that for the baseline configuration. This result is contrary to what was expected. To identify the reasons for this, the velocity distribution in the central plane of the cylinder is shown in Fig 4. Just before TDC, the maximum velocity in C1 is more than that of baseline. But, at TDC and after TDC, the maximum velocity in the C1 configuration is less than that in the baseline case. There is an increase of 13% in the surface area of the bowl largely due to the presence of the central projection. This might be leading to increased skin friction and negating the advantage obtained by introducing re-entrancy. Velocity distributions in both configurations at different times around TDC are shown in Fig 4. From the figure, it can be observed that the extent of the high velocity squish flow vortex is almost the same in both configurations. Due to the swirl–squish interaction, two counter-rotating torroidal vortices are formed in the
3.5 Baseline C1
Swirl number
3
2.5
2
1.5
1 -150
-120
-90
-60
-30
0
30
60
90
Crank angle Fig. 3. Variation of swirl number during closed-valve part of cycle.
120
150
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Baseline
C1 configuration
Baseline
C1 configuration Fig. 4. Distribution of velocity in the central plane of cylinder and velocity vectors in the central plane of cylinder at TDC.
35 Baseline C1
30
2
2
TKE (m /s )
25 20 15 10 5 0 -150
-120
-90
-60
-30
0
30
60
90
Crank angle Fig. 5. Variation of TKE during closed-valve part of cycle.
120
150
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Baseline
C1 configuration
Fig. 6. Distribution of TKE in the central plane of cylinder.
combustion chamber. The C1 configuration has an entry diameter less than that of the baseline configuration. Thus, the squish flow is expected to penetrate more towards the center of bowl. In the present study, however, the squish flow penetration for both C1 and baseline geometries is similar. In case of the baseline configuration, the larger vortex spreads up to the center of the cylinder. But, in the C1 configuration, the large central projection causes the larger vortex to move outwards, which in turn displaces the smaller vortex outwards. Thus, the central projection in case of C1 is obstructs the squish flow from entering deep into the combustion chamber. These reasons explain the lower swirl number in case of C1. In Fig. 5, the variation of cylinder-averaged TKE during compression and expansion strokes is shown. From the figure, it is observed that TKE is also consistently less in case of C1 as compared to the baseline. In previous studies on re-entrant chambers, it was found that TKE peaks twice, once corresponding to squish flow and the second time due to reverse-squish [5]. In both cases, intensification of TKE corresponding to squish flow is marginal and the second peak corresponding to the reverse-squish is almost negligible. This implies that the re-entrancy introduced in C1 is too small to take the advantage of the strong squish and reverse-squish flows. The TKE distribution in the central plane of the cylinder is shown in Fig. 6. Before TDC and at TDC, the maximum values of TKE are slightly higher in C1 as compared to the baseline configuration. The improvement is marginal and the region with the maximum value of TKE is significantly smaller in C1. This explains the reason for the lower average TKE values in C1. Observing at the TKE distribution at 20° CA and TDC, it seems that the central projection is damping turbulence in the central region of the cylinder. The TKE distribution at 20° CA shows that the sharp edge in the baseline configuration is causing higher turbulence during reverse-squish flow. Thus, from air motion analysis, it can be concluded that the combustion chamber geometry selected from literature is actually negatively impacting the flow field in the present case. 5. Combustion study Simulation of spray and combustion provides more information about in-cylinder processes such as spray–air interaction, wall impingement, mixing, combustion, and emission formation. The
Table 5 Models for spray simulation, model constants and input parameters. Parameter/model
Value/reference
Mass of fuel injected Initial droplet diameter Number of parcels injected Droplet break-up model
36 mg/stroke 0.28 mm for I1 and 0.26 mm for I2 37,000 Wave break-up model [35] (C1 = 0.61, C2 = 13, C3 = 1) Enable [44] Dukowicz [45] Walljet1 [46]
Turbulent dispersion model Droplet evaporation Wall interaction
models utilized in the present case as discussed in Section 3 are relatively simpler when compared to the complexity of actual processes. Thus, spray, ignition, combustion and pollutant formation models are provided with constants which are to be calibrated against at least one set of experimental data [29]. Spray mass flow rate and droplet size were calculated using needle lift diagram, injection pressure and cylinder pressure data. Instantaneous droplet sizes have been calculated according to previous studies [36] which have considered cavitation in nozzle flows. Also, the effect of sac (in case of baseline injector) has been simulated by injecting droplets with very low velocity at the end of injection. The droplets injected at a very low velocity remain in the immediate vicinity of the nozzle tip and thus undergo slow evaporation and combustion similar to that of fuel in the sac. In the present case, cylinder pressure and emission data obtained from engine tests and the spray penetration data obtained from experimental correlations have been used to calibrate model constants. The details of the spray parameters utilized in the simulations are listed in Table 5 (In Table 5 I1 and I2 represent two injectors with different hole diameters). The initial part of the spray penetration curve calculated from correlations [1] was used for calibrating the spray break-up model. By varying the break-up model constant C_2, (generally in the range of 5–60) breakup can be advanced or delayed, and thus the penetration can be reduced or increased. By trial and error, a value of 13 was found to give a good match for the penetration. The ignition model constant was adjusted to a value of 1.0e+6 such that the starting point of ignition in the pressure curve matches
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7E+06 Experimental 6E+06
Predicted
Pressure (Pa)
5E+06 4E+06 3E+06 2E+06 1E+06 0E+00 -100
-75
-50
-25
0
25
50
75
100
125
150
Crank Angle Fig. 7. Predicted pressure variation after model calibration in comparison with experimental data.
Table 6 Predicted emission values after model calibrations for the baseline configuration. NOX (g/(kW h))
Soot (g/(kW h))
Predicted
Experimental
Predicted
Experimental
10.11
10.06
0.52
0.52
with the experimental value. The combustion model constant whose value is in the range of 3–25, generally, was calibrated to a value of 7 by matching the pressure curve with the experimental value at 100% load. The predicted pressure curve after model calibration is compared with experimental data in Fig. 7. The prediction of pressure is in sufficiently good agreement with experimental data. The slight discrepancy observed after TDC may be due to the limitations of the model. Pre-exponential factors for NOX and soot reaction rates are adjusted for baseline case to match the predicted values against experimental values. Emission value predictions in comparison with the experimental values are shown in Table 6. Once the model constants are calibrated, the values are kept the same for all further simulations. Emission predictions of the C1 configuration with original and new injector, denoted by C1I1 and C1I2 respectively, along with the baseline configuration are listed in Table 7. Comparing the emission values of baseline and C1I1, there is a significant reduction in NOX and increase in soot. It is to be noted that the C1I1 configuration was not experimentally studied, but is numerically simulated to aid in comparison, especially to isolate the effect of bowl geometry alone. The comparison indicates that for the C1I1 configuration, combustion can be poorer than the baseline case. Figs. 8 and 9 show the in-cylinder pressure and Heat Release Rate (HRR) for the baseline, C1I1 and C1I2 configurations, respectively. It is observed that the difference in combustion and hence in emissions is mainly due to the new injector. Thus, optimization of com-
bustion chamber for higher swirl and TKE may help in improving combustion in the present case. For the C1I2 case, the predicted NOX emission 8.9 g/(kW h) matches well with the experimental value of 9.3 g/(kW h) (Table 2). Table 7 also indicates the effect of injector change comparing the C1I1 and C1I2 cases (reduction in soot and increase in NOX). This confirms the earlier observation of significant improvement in combustion due to the change of injector. The new injector, having smaller hole diameters, leads to faster evaporation and better combustion. The predicted pressure variation in Fig. 8 also indicates a higher peak pressure in case of C1I2 than for the C1I1 case. The heat release rate variation shown in Fig. 9 shows that the ignition delay reduces significantly and makes combustion to happen around TDC when in-cylinder temperatures are higher, and this leads to efficient combustion and lower soot. This explains that the reduction in emissions observed experimentally is mainly due to change of injector rather than change in the combustion chamber geometry. It is a very interesting observation which indicates that there is further scope for reduction in emissions through optimization of combustion chamber geometry. 6. Parametric optimization From the above analysis, it was concluded that there is scope to modify the combustion chamber geometry to further improve combustion and reduce emissions. The above investigations also show that a combustion chamber geometry which improves swirl and TKE intensification around TDC, will lead to more efficient combustion. Thus, the air-motion study was used for selection of combustion chamber geometry and then the optimal injection timing was selected using a combustion study. Simulations for different geometrical variations of the combustion chamber were commenced at the end of suction stroke. Initial conditions were taken from flow field solution at the end of suction for the baseline
Table 7 Predicted emission values for different combinations of geometry and injector. Bowl geometry
Injector
Baseline C1 C1
I1 I1 I2
Injection Timing (w.r.t. TDC) 12.4 12.4 15.6
NOX (g/(kW h))
Soot (g/(kW h))
Ignition delay (CA)
10.11 6.73 8.89
0.52 2.21 1.13
10.2 10.2 8.4
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7E+06
Baseline 6E+06
C1I1 C1I2
Pressure (Pa)
5E+06 4E+06 3E+06 2E+06 1E+06 0E+00 -150
-120
-90
-60
-30
0
30
60
90
120
150
Crank Angle Fig. 8. Variation of in-cylinder pressure for various configurations.
160 Baseline
140
C1I1 C1I2
HRR (J/deg)
120 100 80 60 40 20 0 -20
-10
0
10
20
30
40
50
60
70
Crank Angle Fig. 9. Variation of heat release rate for various configurations.
4.5 Baseline
Swirl number
4 3.5
C1
3
C2
2.5 C3
2 C4
1.5 C5
1 C6
0.5
C7
0
-150 -120
-90
-60
-30
0
30
60
90
120
150
Crank Angle Fig. 10. Swirl number variation during closed-valve duration for various bowl geometries.
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35 Baseline
30 C1
25
TKE (m2/s2)
C2
20 C3
15
C4
10
C5
5
C6
0 -150
C7
-120
-90
-60
-30
0
30
60
90
120
150
Crank angle Fig. 11. TKE variation during closed-valve duration for various bowl geometries.
Fig. 12. Predicted velocity and TKE distribution in the central plane of the C3 configuration.
geometry. For all variants of the combustion chamber geometry, bowl volume and compression ratio of engine were kept equal to that of baseline geometry. The variation of swirl number in the selected geometries along with the baseline and C1 is shown in Fig. 10. C3 configuration shows the highest swirl intensification. The peak swirl number for the C3 case is around 35% higher than that of the baseline value. Variation of TKE for all the selected geometries during the closed-valve duration is shown in Fig. 11. It is observed that, once again, the C3 geometry is better compared to the other geometries. The peak TKE for the C3 case is as much as 50% higher when compared to the corresponding baseline value. The distribution of velocity and TKE in the central plane of the C3 case is shown in Fig. 12. (Corresponding figures for baseline configuration are shown in Fig. 4 (velocity) and Fig. 6 (TKE).) From the figure, it is observed that in addition to the average values, the local values are also fairly higher for the C3 case when compared to
the baseline configuration. In order to confirm that the higher swirl and TKE intensifications for C3 case would lead to improved combustion, combustion simulations were performed for the C3 case and the two nearest competing piston bowl geometries C2 and C7. The results from these simulations are shown in terms of p–h curves in Fig. 13. The configuration C3 shows higher peak cylinder-averaged pressure and larger area under the p–h curve (thus higher power) as compared to C2 and C7. The corresponding emission predictions for C3, C2 and C7 are shown in Table 8. Lower soot and higher NOX emissions are observed for C3 as compared to C2 and C7. It is to be noted that although the soot emissions are slightly lower for the C2 case as compared to the C3 case, the difference is very small. Overall, comparing peak cylinder pressures, C2 is observed to be inferior compared to C3. For C7, soot emissions are far above the corresponding levels for C3. Thus, the bowl geometry corresponding to the C3 is selected as the optimum,
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Fig. 13. Pressure variation in C3I2, C2I2 and C7I2 configurations.
and is taken up for further studies pertaining to injection timing optimization. Combustion simulations were carried out for the C3 geometry with the new injector or C3I2 case, with corresponding injection timing (15.6° CA BTDC). Table 8 shows the emission predictions. As expected, there is a significant reduction in soot and increase of NOX emissions. Also, there is a reduction in ignition delay by 1.8° CA. This indicates considerable improvement in combustion. Higher swirl and TKE causes faster breakup and evaporation of
Table 8 Predicted emissions for C3, C2 and C7 configurations. Injection timing (w.r.t. TDC) 15.6 15.6 15.6
NOX (g/(kW h))
Soot (g/(kW h))
18.66 17.49 10.06
0.06 0.05 2.54
8E+06
Baseline, -12.4 CA 7E+06
C3I2, -15.6 CA C3I2, -5.6 CA
6E+06
Pressure (Pa)
C3I2 C2I2 C7I2
C3I2, -10.6 CA C3I2, -8.6 CA
5E+06 4E+06 3E+06 2E+06 1E+06 0E+00 -150
-120
-90
-60
-30
0
30
60
90
120
Crank Angle Fig. 14. Pressure variation in the C3I2 configuration for different injection timings.
150
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12.4 15.6 10.6 8.6 5.6
NOX (g/(kW h))
Soot (g/(kW h))
Ignition delay (CA)
10.11 18.66 9.44 7.36 5.35
0.52 0.06 0.04 0.08 0.52
10.2 8.4 8.2 7.9 8.4
10o ASoI
15o ASoI
20o ASoI
25o ASoI
30o ASoI ASoI – After Start of Injection
Fig. 15. Spray evolution in terms of droplet locations and diameters (in meters) and equivalence ratio distribution in central plane of cylinder at various crank angles.
liquid droplets resulting in improved air–fuel mixing and better combustion. Fig. 14 shows comparison of predicted pressure curve of C3I2 with 15.6° injection timing with pressure curve of baseline engine. Improvement in combustion causes significant increase in peak cylinder pressure which enables retardation of injection without affecting fuel economy. Various injection timings and the corresponding emission predictions for C3I2 are listed in Table 9. The NOX-soot trade off can be clearly observed from the table. It can also be seen from the table that ignition delay has reduced as the injection timing approaches TDC. At 5.6° CA which is a highly retarded case, the ignition delay has increased again. In this case, ignition occurs after TDC. A slight reduction in temperature and pressure after TDC, and flow destruction after TDC might be the reason for this. An injection timing of 8.6° CA seems to be an optimal value at which there is a reduction of 27% in NOX, and 85% in soot. From Fig. 14, it is also observed that the pressure curve is almost matching with that of the baseline case
indicating that there is no work loss with the proposed 8.6° CA injection timing. Thus, the bowl geometry denoted by C3 with an injection timing of 8.6° CA BTDC shows significant potential for further reduction of emissions from the selected engine. Spray droplet distributions and equivalence ratios are compared for baseline and C3 configurations in Fig. 15. Lower equivalence ratios are observed for the C3 configuration which implies improved air–fuel mixing compared to baseline configuration. This is a reason for the lower soot levels for the C3 configuration. Finally, the effect of a 6-hole nozzle configuration with the same fuel injection pressure was also studied. In order to keep the same amount of fuel delivered in both cases, the nozzle hole diameter was suitably reduced. This new configuration is denoted by C3I3. Simulations were conducted for this configuration at various injection timings of 12.4°, 8.6°, 5.6°, and 2.6° CA. For a timing of 5.6° CA which gives the best soot-NOX trade-off for this configuration, the predicted NOX level is 9.59 g/kW h, and the soot level is
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0.22 g/kW h. These values are higher than those corresponding to the C3I2 configuration with the 8.6° CA timing, thus indicating that the C3I3 configuration is inferior. This may be attributed to the lower penetration by the fuel droplets for the C3I3 case as compared to the C3I2 case. Thus, among all cases studied, the C3I2 configuration is observed to optimum leading to significant reduction in both NOX and soot levels compared to the baseline configuration. 7. Conclusions The present study concerns the effect of swirl induced by re-entrant piston bowl geometries on emissions in a diesel engine, and specifically focuses on a single cylinder, 7.5 kW constant-speed engine. The emission test results of two configurations of the selected engine are reported. The second configuration which has a slightly re-entrant combustion chamber and a sac-less injector was found to yield lower emissions. In order to understand the effect of re-entrancy and injector change on emissions, detailed, threedimensional CFD simulations of the in-cylinder processes were conducted. The effect of chamber geometry and injector change was studied using unfired and fired simulations. Simulation of closedvalve part of the cycle in the two configurations revealed that average swirl and turbulence levels around TDC of compression were higher for the baseline case than for the modified geometry. Increased surface area, presence of a large central projection and insufficient re-entrancy were identified as the reasons for the modified geometry yielding poor results. Combustion simulations revealed that the reduction in emissions observed during experiments is mainly due to the change in the injector rather than change in the piston bowl geometry, thus indicating scope for optimization of bowl geometry. Several piston bowl geometries with varying levels of re-entrancy and different heights of central projections were simulated. A highly re-entrant piston bowl and without a central projection was found to be the best for swirl and TKE intensification around TDC. Combustion simulations were carried out using the selected geometry and injection timings were optimized to keep NOX levels below those of the baseline case. An injection timing of 8.6° CA BTDC was found to be optimum since it led to a 27% reduction in NOX emissions and 85% reduction in soot levels as compared to the baseline configuration. Acknowledgements The authors would like to acknowledge the information and data provided by Prof. M.V. Narasimhan, Mr. J. Kalvani and the Indian Diesel Engines Manufacturers Association (IDEMA). References [1] Heywood JB. International combustion engine fundamentals. New York: McGraw-Hill Book Company; 1988. [2] Brandl F, Reverencic, Cartellieri W, Dent JC. Turbulent air flow in the combustion bowl of a DI diesel engine and its effect on engine performance. SAE Paper 790040. [3] Arcoumannis C, Bicen AF, Whitelaw JH. Squish and swirl–squish interaction in motored model engines. ASME J Fluid Mech 1993:105–12. [4] Arcoumannis C, Bicen AF, Vafidis C, Whitelaw JH. Three-dimensional flow field in four-stroke model engines. SAE Paper 841360. [5] Kondoh T, Fukumoto A, Ohsawa K, Ohkubo Y. An assessment of a multidimensional numerical method to predict the flow in internal combustion engines. SAE Paper 850500. [6] Payri F, Benajes J, Margeo X, Gil A. CFD modeling of the in-cylinder flow in direct-injection diesel engine. Comput Fluids 2004;33:995–1021. [7] Saito T, Daisho Y, Uchida N, Ikeya N. Effects of combustion chamber geometry on diesel combustion. SAE Paper 861186. [8] Di Giorgio F, Laforgia V. Investigation of drop size distribution in the spray of a five-hole, V.C.O. nozzle at high feeding pressure. SAE Paper 950087. [9] Benajes J, Pastor JV, Payri R, Plazas AH. Analysis of the influence of diesel nozzle geometry in the injection rate characteristic. J Fluids Eng, ASME 2004;126:63–71.
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