International Journal of Automotive Technology , Vol. 13, No. 6, pp. 861−872 (2012)DOI 10.1007/s12239−012−0087−3
Copyright ©2012KSAE/067−02pISSN 1229−9138/eISSN 1976-3832
861
OPTIMIZA TION OF INTAKE PORT DESIGN FOR SI ENGINE
Y . L. QI 1), L. C. DONG 2), H. LIU 3), P . V . PUZINAUSKAS 4) and K. C. MIDKIFF 4)*
1)
Caterpillar Inc., 100 NE Adams Street, Peoria 61629, IL, USA
2)
School of Chemistry & Chemical Engineering, Chongqing University, Chongqing 400044, China
3)
Zhejiang Yinlun Machinery Co., Ltd., Hi-tech Industrial Development Zone, Taizhou, Zhejiang 317200, China
4)
Department of Mechanical Engineering, The University of Alabama, Tuscaloosa 35401, AL, USA
(Received 13 September 2010; Revid 21 May 2011; Accepted 19 May 2012)
ABSTRACT −It is well known that in-cylinder flow is very important factor for the performance of SI engine. An appropriate in-cylinder flow pattern can enhance the turbulence intensity at spark time, therefore increasing the stability of combustion,reducing emission and improving fuel economy. In this paper, the effect of intake port design on in-cylinder flow is studied.It is found a vortex existed at the upper side of intake port of a production SI engine ud in the study, during the intake stroke,which will reduce both tumble ratio and volumetric efficiency. A minor modification on intake port is made to eliminate the vortex and increa tumble ratio while keeping volumetric efficiency at the same level. It is demonstrated that the increa in tumble in the new design results in a 20 per cent increa in the fuel vaporization. In this study, both KIV A and STAR-CD are ud to simulate the engine cold flow, as well as ICEM CFD and es-ice ud as pre-processor respectively due to the
complexity of engine geometry. Simulation results from KIV A and STAR-CD are compared and analyzed.
KEY WORDS :SI intake port design, Flow recirculation, High tumble, Fuel vaporization, CFD, KIV A, STAR-CD, ICEM CFD, es-ice
1. INTRODUCTION
Early rearch on spark-ignition engines in pasnger cars demonstrated that combustion is slow when the fuel/air mixture is introduced into the engine cylinder in a quiescent manner without large-scale fluid motion and turbulence generation. Thus, modern internal combustion engines are designed to promote turbulent combustion in the engine cylinder through enhanced in-cylinder flow fields. The details of the in-cylinder flow fields have important effects on the progress and efficiency of combustion. Carefully designed in-cylinder flow fields can enhance combustion,yielding shorter burn times, reduced emissions, and improved fuel economy. Therefore, engine manufacturers have focud on the design of specially shaped intake ports to produce large scale fluid motions, such as swirl and tumble, in the engine cylinder. Furthermore, more stringent emissions legislation and consumer demands have led to incread efforts in recent years to predict and measure in-cylinder f
low fields and their effects on combustion. The in-cylinder flow field can be significantly affected by minor changes in intake port geometry yielding profound effects on engine performance. Traditional (empirical) flow improvements by trial and error of core boxes in a flow bench are costly and time consuming. Even sophisticated flow measurement techniques such as PIV are difficult to employ for tuning high-performance engines, becau of the unsteady nature of the flow inside of engines. In recent year, an almost parallel development in CFD (computational fluid dynamics) techniques has resulted in the development of three dimensional engine models that are capable of predicting in cylinder flow and combustion with a fair degree of accuracy. This development has drastically reduced the engine design cycle time. Moreover, CFD analysis results allow for visualization of the flow field along any desired region, which can be very uful in identifying problem regions.
2. LITERATURE REVIEW
Simulation of steady flow rig by CFD (Guy et al ., 2006;H ong and Tarng, 2001; Mistreanu et al ., 2006a, 2006b,2006c, 2006d, 2006e) has been widely applied in academic and industrial rearch. A more complex and realistic CFD application on engine simulation is to simulate the engine flow and combustion. In the simulation, dynamic mesh is necessary becau valves and piston will keep moving in gas exchange process. Fyhr and Dahlberg (2004) employed a new CFD approach where t
he complete gas exchange and combustion systems are modeled in CFD. Their study focud on the gas exchange system where boundary conditions are placed as far as possible from the esntial components. The effect of combustion on the gas exchange
*Corresponding author . e-mail: Cmidkiff@eng.ua.edu
862Y. L. QI et al.
is accounted for by a simplified source term approach.Both a 5-cylinder diel and a gasoline N/A (natural aspirated) engine have been studied. The diel system includes a rotating turbine and the EGR (exhaust gas recirculation) circuit and the gasoline N/A system is complete from air intake to tailpipe. It had been shown that it is possible to model the complete gas exchange system in 3D CFD. The method can be ud as a powerful tool for improving and understanding almost any phenomenon in the engine that is governed by coupled wave dynamics and 3-D effects, e.g., initial and boundary conditions for single component 3D simulations, EGR studies, and cylinder to cylinder deviations in residuals or volumetric efficiency. Sadakane et al . (2005) performed intake-port design optimization to obtain both high volumetric efficiency and high combustion speed. In conventional intake-port designs,higher tumble ratios usually generate lower coefficients of flow. To obtain both hi
gh tumble ratio and high coefficient of flow, the intake-port design was optimized by CFD analys and flow visualization. In the ca of the developed intake-port, strong airflow can be generated to both the intake side and exhaust side, and a high tumble ratio and high coefficient of flow are obtained.
The purpo of current study is to increa the hybrid engines tolerance to EGR by improving intake port design.Hybrid vehicle with EGR is a good choice for future car due to its high efficiency and low NO x emission. However,high EGR input will dilute the fuel/air ratio, causing some rious performance problems, such as incomplete combus-tion, torque fluctuation, engine misfire. The drawbacks will greatly lo attraction of EGR technique to automobile manufacturers and consumers. An efficient way to overcome the drawbacks is to intensify tumble, which will increa turbulence intensity at ignition timing. The enhancement of turbulence intensity will speed turbulence flame velocity and improve combustion quality, therefore increasing engine tolerance to EGR. In this study, intake port will be optimized gradually to get both higher tumble ratio and acceptable volumetric efficiency. Both KIV A and STAR-CD will be ud to simulate the whole engine process.
3. GOVERNING EQUATIONS
The continuity equation for species m is given as follows:
(1)
Where ρm is the mass density of species m , ρ is the total mass density, u is the fluid velocity vector. D is diffusion coefficient for Fick’s diffusion law. ∆ is the Kronecker delta function. ρm c is the source term due to chemical reaction. ρs is the source term for spray evaporation. By summing Equation (1) over all species, it is easy to obtain the total fluid density equation.
The moment equation for fluid mixture is: (2)
Where P is the fluid pressure. The dimensionless quantity a is ud in conjunction with PGS (Pressure Gradient Scaling)Method. This is a method for enhancing computational efficiency in low Mach number flow, where the pressure is nearly uniform. A tkesw indicates whether the flow is laminar or turbulent, σ is the viscous stress tensor, which is given with µ and λ being the first and cond coefficients of viscosity, respectively. F s is the rate of momentum gain per unit volume due to spray. The specific body force g keeps constant.
The internal energy equation:
(3)
Where I is the specific internal energy. J is the heat flux vector, which is the sum of contributions due to heat conduction and enthalpy diffusion. ρε is the dissipation rate of turbulent kinetic energy. The final two terms, Q c and Q s are sources due to chemical heat relea and spray interaction.
The flow process in the engine cylinder are turbulent.In turbulent flows, the rates of transfer and mixing are veral times greater than rate due to laminar flow.Therefore, it is critical to correctly model the turbulence in order to obtain good predictions that describe the detailed physics and chemical process in the cylinder. Three important length scales reprent different aspects of the turbulent flow behavior. They are integral scale,Kolmogorov scale, and microscale. The integral scale is a measure of largest scale structure of flow field, which is limited in size by system boundaries. The Kolmogorov scale indicates the size of smallest eddies, where the turbulence energy dissipates due to molecular viscosity.The microscale is defined by relating the fluctuating strain rate of turbulent flow field to turbulent intensity. The three length scale can be computed as
(4)(5)(6)
All three length scales vary through the four strokes of IC engine. At the end of the compression stroke, the integral length scales are of the order of clearance height,microlength scales are of the order 1 mm and Kolmogorov length scale are of the order 10-2 mm.
The turbulence quantities (κ and ε) that appear in conrvation equation are given by their own transport equation:
∂ρm ∂t --------∇ρm u ()⋅+∇ρD ∇ρm ρ
-----⎝⎠⎛⎞ρ·m c
ρ·s δm ,l
++⋅=∂ρu ∂t ---------∇ρuu ()⋅+1a
2----∇P A tkesw ∇23--ρk ⎝⎠⎛⎞–∇σF s ρg ++⋅+–=∂ρI ()∂t
------------∇ρuI ()⋅+P ∇u ∇J A tkesw ρεQ ·c Q
·s +++⋅–⋅–=l I C k 32
⁄ε------=l K v 3ε---⎝⎠
⎛⎞14
⁄=l M u ′
∂u ∂x
⁄-------------≈
OPTIMIZATION OF INTAKE PORT DESIGN FOR SI ENGINE 863
(7)(8)
A tumble flow is usually formed late in the intake stroke
and survive through compression stroke until the piston reaches near TDC. Tumble is qualified using tumble ratios.The tumble ratio, TR x and TR y , are defined as following.
(9)(10)
Where ωx and ωy are the angular velocities of a solid-body rotating flow at y-z plane and x-z plane.
4. ENGINE SPECIFICATION AND BOUNDARY CONDITION
The geometry of the engine is provided from the manufacturer in a NASTRAN format shown as Figure 1; it is a faceted surface mesh named as baline intake port configuration.
Figure 2 shows intake valve and exhaust valve lift profiles, which is also provided by the manufacturer and using during the computation. It is found that there is overlap between intake valve lift curve and exhaust valve lift curve that will effect the mesh procedure in KIV A
(explained below).
Specifications of the engine and the computational conditions are listed in Table 1.
5. MESH PROCEDURE
Two CFD software packages, KIV A and STAR-CD, are employed to simulate intake and compression process in this work. ICEM CFD is ud as pre-processor for KIV A and es-ice for STAR-CD. The detailed meshing process for each pre-processor will be discusd in the following:
∂ρk ()∂t -------------∇ρuk ()⋅+23--ρk ∇u ⋅()–σ:∇u ∇µPr k ------∇k ⎝⎠⎛⎞ρεW s
+–⋅+⋅+=∂ρε()∂t
-------------∇ρu ε()⋅+23--c 1c 3–23--c µc ηk ε--∇u ⋅+⎝⎠⎛⎞ρε∇u ⋅()–∇u ⋅+= ∇+µPr e ------∇ε⎝⎠⎛⎞εk --c 1c η–()σ:u c 2ρεC s W
s
+–∇[+⋅TR x ωx
ωc -----ωx 2πN ----------==TR y ωy
ωc -----ωy 2πN
----------=
=Figure 1. Engine geometry with baline intake port configuration.
Figure 2. Intake valve and exhaust valve lift profiles.
Table 1. Specification of SI engine and boundary condition.Model Chrysler cirrus 1996Bore ×Stroke 87.5×101mm Squish 6.56mm Engine speed 2000RPM Connect rod length 151mm Compression ratio 10:1Inlet air pressure 1,000 kPa Inlet air temperature 293 K Piston face temperature 493 K Cylinder wall temperature 493 K Cylinder head temperature
493 K
Figure 3. Standard procedure of KIV A-prep mesh for KIV A code.
Figure 4. Modified procedure of ICEM CFD mesh for KIV A code.
864Y. L. QI et al. The original structure of KIV A is introduced in Figure 3.
A more advanced meshing software package ICEM CFD
was ud in current study due to the complexity of the
engine geometry, as well as ENSIGH T as an advanced
post-processor. The modified structure of KIV A is shown
in Figure 4. This meshing/blocking procedure is shown as
Figure 5. After completing above procedure, a hexahedron
and structure meshed engine is obtained shown as Figure 6.
The interior mesh of engine is shown by cutting-plane as
Figure 7.
Dynamic mesh is involved in this simulation due to
piston and valves’ movement. V alve movement can be
classified as vertical valve and canted valve according to
the moving path of valve. It is not only a matter of
generating the initial grid but also dynamically valve
snapping and rezoning of the canted valves in this study.
Moreover, due to the intake/exhaust valves overlap at their
lowest position, the grids are generated with the valves in
their clod (TDC) positions, and provide a sufficient
number of planes near the top of the cylinder to allow for
valve snapping. Considering combustion chamber resolution
at the time of combustion when the piston at its TDC, at
least a bare minimum of three planes of cells in a pentroof
combustion chamber between the valves and the piston are
maintained. Subroutine REZPENT in KIV A-3V is ud to
adjust vertex position in the dynamic mesh for pentroof
ca. This subroutine was revid according to engine
geometry, mesh and valve lift profile. In current ca,
sixteen areas are created in cylinder and are treated
differently to avoid inverted cell appearing in the engine
cycle.
In STAR-CD, after read in PROAM, the engine
combustion dome, piston crown, cylinder wall, ports and
valves are assigned to different cell types in PROSTAR.
After gathering all the desired cells and remove unud
cells and vertices from the model, a databa file of the
engine surface mesh is written in PROSA TR for mesh
generation in es-ice where the ur generates a template
that approximates the desired engine geometry. The
template contains all connectivity information for the final
grid and is specified by a limited number of parameters that
can be altered to generate in a variety of configurations.
Then the mesh is generated by either trimming the template
(using the same AutoMesh methodology incorporated in
pro-STAR) or mapping the surface of the template to the
surface of the problem geometry.
6. ASSESSMENT OF BASELINE INTAKE PORT
CONFIGURATION
In this ction, a cold flow study using the KIV A3VREL2
CFD code is carried out on the baline intake port
configuration. The simulation results are illustrated as
velocity vector diagrams. Figure 8 shows the cutting plane
that is t at y = 0mm and y = 20mm, respectively.
Figure 9 shows the velocity vector plot at 60o CA in two Figure 5. Schematic of topology for blocking.
Figure 6. Hexahedron and structure mesh.
Figure 7. Interior mesh.
OPTIMIZATION OF INTAKE PORT DESIGN FOR SI ENGINE 865
vertical cutting planes. At the same scales, the flow patterns show similarities in the two different locations.The flow jet tends to be stronger as the flow moves further outwards from the cylinder center. V ector plot at y = 20mm reveals that there are two tumble jets inside the cylinder, one is the jet with a large vortex guided by the dome chamber; the other is a counter rotating jet with little smaller scale.
Figure 10 (a) exhibits a tumble vortex at y = 0 mm plane when the intake process going to 90o CA. Note that the maximum velocity is located at y = 20 mm around the inlet valve curtain in Figure 10 (b), becau the flow jet from one intake valve was countervailed by the one form another valve at y = 0 mm plane. When the piston arrives at BDC,the tumble vertices tend to be weak at any cutting planes,as shown in Figure 11. Further obrvation reveals, in Figure 12, flow paration was detected at the valve stem step region just after 90o CA, although there is still a tumble vortex. After 90o CA, the flow paration is growing,which can be en in Figure 11 (b). Figure 13 indicates that backflow occurs once the piston pass BDC.
A flow paration occurs when the boundary layer encounters a sufficiently large adver pressure gradient.The fluid flow becomes detached from the surface of the object, and instead takes the forms of eddies and vortices.In aerodynamics, flow paration can often result in incread drag, par
ticularly pressure drag that is caud by the pressure differential between the front and rear surfaces of the object as it travels through the fluid. For this reason much effort and rearch has gone into the design of aero-or hydrodynamic surfaces that keep the local flow attached for as long as possible.
7. OPTIMIZATION ON FLOW SEPARATION
After completing the above flow field analysis, the baline
intake port geometry must be altered to eliminate the flow recirculation region. Due to the flow paration that occurs at the valve stem step position, the step is removed and the intake duct is smoothed around the valve stem. Two simultaneous computer simulations are carried out to predict the in-cylinder flow field of the same engine with the preliminary modified intake port under the same
Figure 8. Cutting planes.
Figure 9. V elocity vectors at 60o CA.
Figure 10. V elocity vectors at 90o CA.
