a methodology for in cylinder flow field evaluation in a low stroke

8/8/2019 A Methodology for In Cylinder Flow Field Evaluation in a Low Stroke

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SAE TECHNICALPAPER SERIES 2002-01-1119

A Methodology for In-Cylinder Flow Field

Evaluation in a Low Stroke-to-Bore SI Engine

G. Cantore, S. Fontanesi and E. MattarelliUniversity of Modena and Reggio E.

G. M. BianchiUniversity of Bologna

Reprinted From: Modeling of SI Engines and Multi-Dimensional Engine Modeling(SP1702)

SAE 2002 World CongressDetroit, Michigan

March 4-7, 2002


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2002-01-1119

A Methodology for In-Cylinder Flow Field Evaluation in a LowStroke-to-Bore SI Engine

G. Cantore, S. Fontanesi and E. MattarelliUniversity of Modena and Reggio E.

G. M. BianchiUniversity of Bologna

Copyright 2002 Society of Automotive Engineers, Inc.

ABSTRACT

This paper presents a methodology for the 3DCFD simulation of the intake and compressionprocesses of four stroke internal combustion engines.The main feature of this approach is to provide veryaccurate initial conditions by means of a cost-effectiveinitialization step. Calculations are applied to a lowstroke-to-bore SI engine, operated at full load andmaximum engine speed. It is demonstrated that initialconditions for this kind of engines have an importantinfluence on flow field development, particularly interms of mean velocities close to the firing TDC.

Simulation results are used to discuss the

choice of a set of parameters for the flow fieldcharacterization of low stroke-to-bore engines, as wellas to provide an insight into the flow patterns duringthe overlapping period.

INTRODUCTION

The level of turbulence within the cylinder, aswell as the mean flow field, plays a fundamental role inthe combustion process of S.I. engines. Therefore, anyaccurate simulation of combustion should be precededby the analysis of the in-cylinder flow during the intakeand compression strokes. Furthermore, the analysis of

the cold flow can be very useful to address theoptimization of the intake duct and valve assembly.

The computational domain for themultidimensional simulation of the intake andcompression stroke is generally restricted to onecylinder, and to a portion of the intake and exhaustpipes attached to the same cylinder. In order to keepinto account the influence of the whole system on thesingle cylinder, time-varying conditions must beapplied to the boundaries. These conditions aregenerally provided by 1-D simulations, previouslycarried out on the whole engine. Sometimes, the

multidimensional and the 1D code are coupled, andthey run parallel, exchanging information through theinterfacing boundaries. Unfortunately, this kind ofsimulation is very demanding from a computationalpoint of view, since more than one complete engine

cycle must be calculated by the multidimensional codein order to reach a satisfactory cycle-by-cycle stability[1]. Furthermore, complex phenomena such as

injection, mixing and combustion should be adequatelymodelled.

Another critical issue in the CFD simulation ofthe intake and compression strokes is the definition ofinitial conditions. In order to include the overlappingperiod, the calculation should start at Intake ValveOpening (IVO), or just before. At this time of theengine cycle, the flow field in the computationaldomain is far from trivial, particularly in a tuned intakemanifold. Even if the intake valve is closed, the freshmixture does not stay still, since the duct is run throughby strong compression and expansion waves. In thiscase, a 1-D simulation carried out on the whole enginecan not help very much. The 1D model of the intakemanifold is roughly simplified, particularly at the valveport, and all the 3-D details are completely out ofreach. The same could be repeated for the exhaustduct.

In order to give more accurate initialconditions, one possibility could be to perform a 3DCFD simulation of one cylinder for a few completeengine cycles. The boundary conditions are providedby a 1D simulation of the whole engine, while, for the

sake of simplicity, combustion is not modelled. Besidesome conceptual objections, this methodology is verydemanding, since the simulation must be performed onthe whole domain and for more than one completeengine cycle, in order to reach a satisfactory cycle-by-cycle stability.

The methodology proposed in this paper forthe intake and compression strokes analysis is aimedat strongly increasing the cost-effectiveness of suchkind of simulations, by providing accurate boundaryand initial conditions and limiting the computationaldemand at the same time. It is important to remark that

this methodology, although particularly suitable forhigh performance engines, where ramming effects playa fundamental role, can be applied to every kind ofengine.


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The idea underlying this methodology is tosplit the domain into two separate regions, the firstrepresenting the portions of the intake and exhaustducts considered, the latter being the cylinder. Thesimulation of the flow through a duct is relativelysimple and fast, while it is much more demanding tomanage a cylinder, the volume of which changeswidely during the calculation. Furthermore, duringintake and exhaust strokes, pressure and velocity

conditions are more uniform within the cylinder than inthe attached pipes. Then, it is convenient to calculatethe flow field in the intake and the exhaust duct byseparating them from the cylinder. Forcing time-varying boundary conditions, (calculated by means ofa previous 1-D engine simulation) at each end of thepipe one can get a very accurate flow field within thesepipes at any time of the cycle. The results obtained willthen be mapped into the complete computationaldomain for the intake and compression strokes finalsimulation.

The four valve pentroof combustion chamber

is the most widespread configuration for S.I.homogeneous charge engines. Such a configuration isused for both fuel efficient, low speed automotiveengines and high speed racing engines. The formerhave stroke-to-bore ratio larger or equal to one, whilefor the latter such a ratio can be lower than 0.5. Whilefor low speed engines the in-cylinder flow field has

been extensively studied [28], for racing engines thenumerical and experimental approach is still not wellestablished. One of the most critical topics is the flowfield characterization at the end of the intake process.It is well known that, for low specific power engines,the mean velocity field at BDC is organized according

to clear patterns, which can be described in terms oftumble number [9]. As the stroke-to-bore ratiodecreases, these patterns become more and morecomplex. For a stroke-to-bore ratio close to 0.5, themacro-vortex generated by the intake flow directedtoward the exhaust side is no more representative forthe flow field, since at least one other vortex, havingcomparable dimension and intensity but rotating in the

opposite way, takes place within the cylinder [1012].As a consequence, such kind of engines have very lowtumble numbers, sometimes negative [7]. In spite ofthis, turbulence intensity is usually high enough toprovide good combustion velocity.

From all the above considerations, thetraditional tumble ratio has no meaning for engineshaving ultra-low bore-to-stroke ratios. Therefore, afurther purpose of this paper is to address thedefinition of more representative parameters for suchkind of engines.

Another difference between low and high bore-to-stroke ratio engines is the intensity of the meanvelocity field in the last part of compression stroke, atmaximum engine speed. Obviously, the shorter is

stroke, the higher is top revving speed; as aconsequence, when combustion is started by the sparkplug, also mean velocities within the cylinder arehigher. Furthermore, at high engine speed, ignitionadvance is usually larger than at low speed. Sincemacro-vortex intensity is decreasing when

approaching top dead center, the larger the sparkadvance, the higher the mean velocities will be. Thiscondition is not ideal for combustion, since the flamekernel is stretched out and can be carried away fromthe spark plug by the mean velocities. As a result,combustion does not start in the center of thechamber, and a portion of mixture can remainunburned. Therefore, when performing CFD simulationof low stroke-to-bore ratio engines, the residual mean

flow during the compression stroke should be carefullycontrolled, particularly below the spark plug [1319].

In SI four stroke racing engines, high poweroutput is obtained by optimizing gas exchangeprocesses at top engine speed. The overlappingperiod is particularly critical: the exhaust systemshould be tuned in order to produce a suction, whilethe pressure trace upstream of the intake valve shouldpresent a deep timed with the maximum suction withinthe cylinder (generally occurring close to ExhaustValve Closure, EVC). The tuning of the intake andexhaust system is generally carried out with the help of

1D engine cycle simulation codes. These codesrequire experimental inputs to model the flow throughthe valves, in and out of cylinder. Particularly, theinstantaneous effective area of the valves must beprovided. Such values are determined by means ofsimple experiments at the steady flow bench.Unfortunately, these experiments are not able toaccount for the shrouding effect generated by thepiston crown, very close to the valves during theoverlapping period. The 3D CFD simulation of theintake stroke can be used as a powerful source ofinformation for assessing the influence of the pistonshrouding on the valve flow.

In this paper, the methodology mentionedabove has been applied to a low stroke-to-bore ratioengine. Some results from an experimentally validated1-D model of this engine have been used to providethe boundary conditions and to assess the quality ofthe initialisation results. Finally, intake andcompression strokes results have been discussed andcompared to those obtained with a simplifiedapproach.

FLOW FIELD INITIALISATION

All the calculations presented in this paperhave been performed by means of the CFD codeVECTIS by Ricardo Software ltd, Burr Ridge, IL.

The computational domain for the initialisationis made up of two completely disconnected regions,visible in figure 1. Symmetry has been invoked, andonly one half of the system has been modelled. Theintake side is on the left, the exhaust on the right. Atthe intake pipe inlet, as well as at the exhaust pipeoutlet, the flow is expected to be almost one-dimensional.

To account for the influence of the valvemotion, moving boundaries, following the valveactuation profiles, have been applied to the valvecurtain areas. The presence of these moving domainsimplies that the initial mesh is increasingly distorted, up


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to a critical point. Here, the computational domainmust be refreshed by switching to a new undistortedmesh. The operation is repeated up to the finalposition. For the considered domain, the completevalve motion has been covered with a set of 59undistorted meshes. The VECTIS pre-processorallows the user to semi-automatically create such a setof meshes. Unfortunately, the mesh refreshing timeduring the simulation is not known a-priori. Therefore,

a user subroutine has been implemented by theauthors in order to predict critical distortion. Thesubroutine is very helpful in reducing the number ofmeshes and the pre-processing time.

In order to correctly represent the wallinfluence in the valve curtain region, at least four cellshave been placed along the valve axis direction, evenat the minimum lift. Therefore, a number of cellsranging from 50000 to 120000 has been adopted,according to the valve position. Nevertheless, toprevent an excessive mesh distortion at very low lifts,a minimum lift of 0.5 mm has been imposed. Valve

timing has been adjusted in order to keep the integralof the valve curtain area equal to the actual value.

Boundary conditions, derived from a previous1-D engine simulation have been applied in order totake into account the effects of the whole system. Thissimulation has been carried out by using the WAVEcode, by Ricardo Software.

As far as the intake duct is concerned, a time-varying mass flow rate has been imposed at the inletsection, while the 1-D in-cylinder pressure has beenenforced as a static pressure boundary condition for

the intake valve curtain area. It is worthwhile to remarkthat simulation results can change when a differentkind of boundary conditions is applied (for examplewith pressure traces forced at both ends of the pipe).The set-up used in the paper is the best compromise,found by the authors, between numerical stability andaccuracy.

For the exhaust duct, a different approach hasbeen used in order to limit the numerical and physicaldifficulties associated to the sonic flow at the ExhaustValve Opening (EVO). The WAVE in-cylinder pressurehas been imposed on the valve curtain as a totalpressure, while a static pressure is enforced at theexhaust duct outlet. This strategy is only partiallysuccessful, but it should be considered that theinfluence of blow-down on the intake process is notvery significant.

To reach satisfactory cyclic stability, fourcomplete engine cycles have been simulated, andresults in terms of local velocity and pressure traceshave been compared. It has been found that threecomplete cycles should be performed before startingthe final simulation of the intake and compression

strokes. However, no relevant differences can benoticed between cycle 2 and 3. It is worthwhile toremark that, due to the relatively limited meshdimension, a complete cycle simulation runs in 16hours on a four-processor IBM RS6000 SP3.

Figure 1: the computational domain for theinitialization

A comparison between multidimensional and1-D results has been performed at different locationsthroughout the domain, in order to assess the qualityof the flow field initialization. Figure 2 shows thepressure traces on a cross section in the middle of theintake duct. The 3D results correspond to a celllocated in the middle of such a cross section. Asexpected, the curves agree fairly well in terms of bothwaves intensity and pulse/deep timing. The differencesexisting between the two trends are due to localeffects, represented in a lumped fashion by the 1-Dcode. The agreement between multidimensional and1-D simulation is still satisfactory when considering theintake valve upstream section, figure 3, and theexhaust valve downstream section, figure 4. A furtherconfirm appears in figure 5, where the mass flow ratesthrough the intake valves are plotted.

Finally, the flow field at the intake pipe inletand at the exhaust duct outlet has been analyzed.While the flow is very uniform at the intake inlet,velocity gradients have been observed at the exhaustoutlet. However, in this region pressure differences arenegligible.

C. A. DEG.

ABS.PRESSURE

1-D 3-D

Figure 2: Comparison between 1-D and 3-D CFDresults: pressure trace in the mid of the intake duct.


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C. A. DEG.

ABS

.PRESSURE

1-D 3-D

Figure 3: Comparison between 1-D and 3-D CFDresults: pressure trace upstream of the intake port.

C. A. DEG.

ABS.PRESSURE

1-D 3-D

Figure 4: Comparison between 1-D and 3-D CFDresults: pressure trace downstream the exhaust valve.

C. A. DEG.

MASSFLOW

RATE

1-D 3-D

Figure 5: Comparison between 1-D and 3-D CFDresults: mass flow rates through the intake valves.

INTAKE AND COMPRESSION STROKES CFD

SIMULATION

For the intake and compression strokesimulation, the combustion chamber and the cylinder

volume have been added to the spatial domain. Themoving solid walls are now the valves and the piston

crown. As for the initialization, symmetry has beeninvoked, and only one half of the domain has beenmodeled. Due to the change in the computationaldomain, a new set of 35 undistorted grids has beendefined. The angular position for mesh refreshing hasbeen automatically calculated by an authorsimplemented subroutine, so as to prevent cell criticaldistortion. The mesh dimensions are ranging from aminimum of 60000 to a maximum of 325000 cells.

Since the map of the velocity field within thecylinder is not available from the previous calculations,the simulation starts at 40 c.a. degrees before IVO,from a totally still condition within the cylinder. In thisway, before the intake valve starts to open, the in-cylinder flow can adjust itself, fitting itself to the flow inthe exhaust duct.

As was done for the initialization, theinstantaneous mass flow rate has been assigned atthe intake duct inlet, while a static pressure trace is

forced at the exhaust duct outlet. A variable time stephas been adopted for keeping the maximum Courantnumber in the optimum range, i.e. between 0.1 and 1,throughout the entire simulation, and the convergencecriterion has been set to 1.0e-05.

The simulation runs in 32 hours on a four-processor IBM RS6000 SP3. It should be pointed outthat a simulation carried out on the same spatialdomain and computational platform, with just 2complete initialization cycles, would require about 135hours.

In order to assess the physical consistency ofthe simulation, a further comparison with the wellestablished 1-D engine model has been carried out. Infigure 6, the pressure trace upstream of the intakevalve is considered, finding a very good agreement.Figure 7 presents the instantaneous mass flow ratethrough the intake and the exhaust valve. Thedifferences which can be observed for the intake massflow rate in the first part of the intake process are dueto the shrouding effect of the piston, which can beaccounted for only in the multidimensional simulation.Such an effect will be discussed in the following. As afinal control, the total amount of air delivered by the

cylinder during the intake stroke has been calculated.The value predicted by the 3D CFD analysis is 3%lower than the value found by the 1D approach,confirming the physical soundness of the simulation.

Finally, the proposed methodology has beencompared to a simplified approach, which will bereferred to in the following text as the pseudo-strokeone. According to this approach, the exhaust duct isnot considered and calculations start at TDC. Time-varying boundary conditions, given by the same 1Dengine simulation, are applied at the intake duct inlet.On the same computational platform the simplified

calculation takes about 16 hours.


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C. A. DEG.

ABS.PRESSURE

1-D 3-D

Figure 6: Comparison between 1-D and 3-D CFDresults, during the intake process: pressure traceupstream of the intake port.

C. A. DEG.

MASSFLOW

RATE

1-D Int. 3-D Int. 1-D Exh. 3-D Exh.

Figure 7: Comparison between 1-D and 3-D CFD

results, during the intake process: mass flow ratestrough the intake and the exhaust valve.

RESULTS

Computational data have been processed inorder to get an insight into the flow through the intakeassembly. In 1-D engine simulation codes, dischargecoefficients are generally used to evaluate the fluid-dynamic efficiency of valves and port, as well as tomodel the flow in and out of cylinder. Such coefficientsare experimentally determined at the steady flowbench, according to the following expression:

=

0

1s00ref

d

p

paA

mC

(1)

where:m is the mass flow rate; 000 ,, pa are the

total quantities (density, speed of sound and pressure,

respectively) at the upstream section; 1sref p,A are the

geometric area and the static pressure at thedownstream section. Finally:

=

+

1

0

1

2

0

1

0

1

1

2

p

p

p

p

p

p sss(2)

The definitions above have been used tocalculate the instantaneous intake dischargecoefficients, on the basis of 3D-CFD results. Only the

flow from the intake duct toward the cylinder has beenconsidered. The upstream quantities (pressure,density, speed of sound) are instantaneous valuesaveraged over a cross section of the intake manifold,at about 50 mm from the valve seat. The mass flowrate too is picked up in this section. As far as thedownstream pressure is concerned, a value averagedon the valve curtain was assumed.

It should be pointed out that a comparisonbetween instantaneous and steady dischargecoefficients can be carried out only under thehypothesis of quasi-steady flow. Unfortunately, this

hypothesis is not verified when the intake valve isclosing and strong ramming effects are taking placeupstream the valve. Therefore, the results presented inthe following include only the first part of the intakeprocess, i.e. with the valve lift increasing up to themaximum height. As far as 1D engine simulationcodes are concerned, transients effects are generallyaccounted for by adding a term representing the flowinertia to the equation for steady flow.

Calculated instantaneous dischargecoefficients and their steady experimental counterpartsare compared in figure 8. In the plot, the reference

area for discharge coefficient is the inner area of thevalve seat, while lift is normalized against valve seatdiameter. It is remarkable that, exactly when the valveis closest to the piston crown, i.e. in the middle of theopening period, the discharge coefficient falls down.Figure 8 confirms that, in low stroke-to-bore engines,valve permeability is strongly influenced by the pistoncrown. Finally, when the piston is far enough from thevalve, transient discharge coefficients are close to thesteady ones.

0

0.2

0.4

0.6

0.8

1

0 0.1 0.2 0.3 0.4 0.5

LIFT / VALVE SEAT DIAMETER

Cd

transient steady

Figure 8: Comparison between steady intake valvedischarge coefficients (measured at the flow bench),and the transient discharge coefficients, calculatedduring the CFD simulation of the intake stroke.


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The influence of piston shrouding can beaccounted for in 1-D simulations, by entering correctedvalues of valve effective area. Such a correction canbe operated by replacing the values of effective areameasured at the steady flow bench with the valuescalculated during the intake stroke simulation. Figure 9shows this correction, whose entity in terms of meaneffective area is about 16%. Running the 1-Dsimulation with the corrected intake areas, the

difference with the previous simulation results is muchmore limited, as visible in figures 10 and 11. Figures10 and 11 show pressure and mass flow rate traces atthe intake valve. The intake valve shrouding seems toproduce a slight shifting of the pressure trace, and afall in the first part of the flow rate plot. The differencein terms of trapped mass is 4%.

CRANK ANGLE

EFFECTIVEAREA

Corrected Baseline

Figure 9: Correction of intake effective area in order toaccount for the shrouding effect of the piston. Input for

1-D engine simulations.

Now, an important question to be addressedconcerns the influence of modified boundary and initialconditions on the 3D CFD simulation of the intakeprocess. How much will the difference in the 1Dresults, shown in figures 10 and 11, change the flowrate predicted by the 3D code ?

The answer has been found by repeating allthe 3D calculations with the new boundary conditions,and comparing new and old results. Considering themass flow rate traces through the intake valve,negligible differences have been observed. Therefore,small variations in the boundary conditions have verylittle effect on the intake process simulation results.

The opinion of the authors is that, in 3Dsimulations of pulsating flows within complexcomponents, such as the valve assembly, point wiseflow patterns are governed more by local geometricaldetails, than by boundary conditions. It should beobserved that the relevant variations in boundaryconditions occur only for a limited crank angle interval.Furthermore, the inlet boundary is located relatively far

away from the valve assembly. Therefore, smalldifferences between boundary conditions are dumpedby the whole gas-dynamic system.

CRANK ANGLE

PRESSURE

Baseline Corrected

Figure 10: Influence of the effective area correction onthe pressure trace at the intake valve. Results of 1-Dengine simulations.

CRANK ANGLE

MASSFLOWR

ATE

Baseline Corrected

Figure 11: Influence of the effective area correction onthe mass flow rate through the intake valve. Results of1-D engine simulations.

The definition of a set of parameters representing thein-cylinder flow field at BDC is a particularly difficulttask, when ultra-low stroke-to-bore engines areconsidered. Figure 12 gives an idea about thecomplexity of the flow patterns and can be used as abasis for discussion. Pictures A and B show thevelocity field on two parallel planes, the symmetryplane and the one passing through the valve axis. Alsopictures C and D present velocity on two parallelplanes, orthogonal to the symmetry plane and passingthrough the cylinder axis (C) and through the intakevalve axis (D).

The in-cylinder velocity field is moulded by the jet through the intake valve curtain, and by theinteraction between such a jet and the combustionchamber walls and the piston crown. It is quite clear,even only considering the bore dimension againststroke, that there is not the faintest chance of findingone organized flow structure extending all over thevolume. It seems indeed reasonable to split thecylinder volume in two parts, the first including theintake valves, and the second containing the exhaustones.

In the region under the intake valves, velocity vectorsare mainly oriented parallel to the symmetry plane, seefigures 12B and 12C, and two main eddies can berecognized. The first one is generated by the intakeflow directed toward the exhaust side (referred to inthe


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Figure 12: Velocity field at BDC for a low bore-to-stroke engine at top revving speed. Plots on the symmetry plane (A),on a section parallel to the symmetry plane and containing the valve axis (B), on the plane orthogonal to the symmetryplane including the cylinder axis (C) and on the plane orthogonal to the symmetry plane, containing the valve axis (D).

following as direct tumble), the second one is acounter-rotating vortex created by the flux toward thecylinder liner (referred in the following as reversetumble). Looking at figure 12B, under the intake valvethe former seems to be much smaller and weaker thanthe latter. The situation is completely reversed when

analyzing the velocity field on the symmetry plane,figure 12A.

Under the exhaust valves, see figure 12B, thevalve jet creates two clockwise loops, the former in the

region between intake and exhaust valves, the latter inthe corner far from the intake valve. In the same place,but on the symmetry plane, the eddy rotational versusis reversed.

In the central region of the chamber and under

the exhaust valves, an organized motion takes placeon planes orthogonal to the symmetry one. This eddy,clearly visible in figure 12C, will be referred to in thefollowing as cross tumble.


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After the examination of the velocity fieldpresented in figure 12, which is quite similar to othersobserved by the authors, the following conclusions canbe drawn.

The cylinder should be split into two regions, theformer under the intake and the latter under theexhaust valves, each one governed by one or apair of vortexes. These main eddies, the length

scale of which is comparable to the stroke, shouldbe able to survive during the compression stroke,thus acting as a source of turbulence whencombustion starts.

In the intake region, the reverse tumble isrecognized as the main vortex

In the exhaust region, direct tumble and crosstumble play together the most relevant role.

For sake of simplicity, the cylinder is split by aplane orthogonal to the symmetry plane and including

the cylinder axis. In this way, since the intake valvesare always larger than the exhaust ones, the region ofthe chamber between intake and exhaust is allotted toexhaust.

The intensity of each vortex can be quantified bythe ratio of the equivalent solid body angular velocity tothe engine rotational speed. Further details about theequations used for these calculations are shown inAppendix A.

In figures 13-15, the normalized values of directtumble, reverse tumble and cross tumble, are plottedagainst crank angle during the intake and compressionstrokes. All the values of tumble ratio are divided byone reference positive number, which can not bedisclosed in the paper. The results of the completesimulation are compared to the ones obtained from thepseudo-stroke calculations. It can be observed that thetraces obtained from the two methodologies are quitedifferent during the intake stroke, especially for directtumble, while such differences diminish as the pistonapproaches TDC. At BDC, when simulation results areusually analyzed, the mean flow field predicted by thecomplete simulation presents patterns which cannot be

found in the pseudo-stroke analysis. Thus, theimportance of a proper gas-dynamic initialization isclearly demonstrated.

When considering absolute values of the tumbleratios, it is interesting to observe that a relativemaximum is always reached in the second half of theintake stroke, when the mass flow rate entering thecylinder reaches a peak. Another maximum isgenerally found in the middle of the compressionstroke, for the spinning-up effect. Only for direct tumblein the pseudo-intake analysis such an effect does notoccur.

The comparison presented in figure 13 (directtumble) is particularly interesting. The value of the firstpeak, predicted by the pseudo-stroke is twice the one

estimated in the complete simulation and is retardedtowards BDC. Non-negligible differences persist alsothroughout the compression stroke: in the completesimulation, tumble ratio becomes negative and has apeak, while in the pseudo-stroke it remains positiveand decreases to a near-zero value.

As far as reverse tumble is concerned (figure 14),complete and pseudo-stroke simulations yield quite

different results during the intake stroke. However, inboth cases the parameter is negative for almost thewhole stroke. Therefore, also under the intake valvesthe flow rotational versus is concordant with the directtumble eddy. Such a result can be explainedconsidering that during the intake stroke the maincomponent of the flow through the intake valve curtainis the one oriented toward the exhaust side. This jet isabruptly deflected downward by the exhaust valve dishand the spark plug. Then, a large eddy takes positionjust under the intake valves. In the second part of theintake stroke, a better balance is reached between theflow rate directed toward the exhaust side and the one

oriented toward the cylinder liner. As the latterincreases, the previous eddy under the intake valvedecreases and/or is pushed toward the exhaust side.At BDC, the value of reverse tumble is alreadypositive, and continues to increase at a fast rate whenintake valve is closing.

It is important to remark that at BDC, in thecomplete simulation, direct and reverse tumble haveabout the same value, while the cross tumble numberis lower. Then, if tumble were defined in the traditionalway, i.e. all over the cylinder and for the velocitycomponents parallel to the symmetry plane , its value

would be close to zero. Conversely, in the pseudo-stroke the direct tumble is dominant, and the traditionaldefinition would yield a completely different number.

In figure 15 the differences between the twomethodologies in terms of predicted cross tumble arevisible only in the first half of the intake stroke.

-0.3

0

0.3

0.6

0.9

1.2

-300 -240 -180 -120 -60 0

C. A. DEG. (after firing TDC)

TR/TR*

Complete Pseudo

Figure 13: Comparison between the methodologyproposed in the paper (referred to as complete), and asimplified approach (referred to as pseudo) in terms of

normalized direct tumble.


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-0.6

-0.3

0

0.3

0.6

0.9

-300 -240 -180 -120 -60 0


TR/TR*

Complete Pseudo

Figure 14: Comparison between the methodologyproposed in the paper (referred as complete), and asimplified approach (referred as pseudo) in terms ofnormalized reverse tumble.

-0.6

-0.3

0

0.3

0.6

0.9

-300 -240 -180 -120 -60 0


TR/TR*

Complete Pseudo

Figure 15: Comparison between the methodology

proposed in the paper (referred as complete), and asimplified approach (referred as pseudo) in terms ofnormalized cross tumble.

The final evaluation of the attitude of the meanflow field to promote turbulence production in the lastpart of compression is assessed by the ratio ofturbulence intensity to mean piston speed. In figure 16such a ratio is plotted for both the complete simulationand the pseudo-stroke. The two plots are similar, butthe pseudo-stroke values are slightly higher, with adifference of 5-10%, from 100 up to 30 degrees beforeTDC. In this range, it can be observed that turbulence

intensity decreases very slowly. This behaviorsuggests that the main vortexes are indeed able tosustain turbulence, thus combustion velocity. Thisresult is qualitatively confirmed by the experimentalvalues of spark advance, which, for the analyzedengine, are almost constant in the range of highengine speed.

Finally, figure 17 presents the values of meanvelocity, averaged on the gasket plane and normalizedagainst mean piston speed. This parameter is plottedagainst crank angle. Results of the complete and thepseudo-stroke simulations are compared. In this case,

the two methodologies does not yield the same output.For the complete simulation, the gasket velocitydecreases at a constant rate up to 40 deg. beforeTDC. Here, the occurrence of squish sustains thevelocity field, and a relative maximum can be observed

a few degrees later. Such a relative maximum is notpresent in the pseudo-stroke plot, which does notseem to be influenced by squish. This is anotherevidence of the differences on the flow field producedby the two methodologies. Such a difference canbecome important for the combustion evolution, sincethe high values of normalized average velocity suggestthat the mean flow field is strong enough to carry awaythe flame kernel from the middle of the combustion

chamber. The difference between the approachproposed in the paper and the pseudo-intake isconfirmed also by the point-wise analysis of thevelocity field at 60 deg. BTDC. According to the formerapproach the velocity magnitude under the spark plugis 34% lower than the one provided by the pseudo-stroke simulation.

0.4

0.5

0.6

0.70.8

0.9

1

-100 -80 -60 -40 -20 0


U'/MPS

Complete Ps eudo

Figure 16: Comparison between the methodologyproposed in the paper, and a simplified approach(intake pseudo-stroke) in terms of normalized turbulent

velocity.

0

0.5

1

1.5

2

-100 -80 -60 -40 -20 0


U_gsk/MPS

Complete Pseudo

Figure 17: Comparison between the methodologyproposed in the paper, and a simplified approach(pseudo-stroke) in terms of normalized mean velocity,averaged on the gasket plane.

CONCLUSION

The spatial domain for the CFD analysis of the

intake and compression processes in internalcombustion engines is usually made up of one cylinderand a portion of the attached piping. While boundaryconditions can be easily provided by a 1D simulationcarried out on the whole engine, it is much moredifficult to create an accurate map of the initial flow


12/13

field. The methodology proposed in this paper allowsone to calculate an accurate flow field within the intakeand exhaust ducts attached to the cylinder. This isdone by forcing at the ends of the manifolds time-varying boundary conditions, calculated by means of aprevious 1-D engine simulation. The flow field is usedas an initial condition for the simulation of the intakeand compression strokes. Such a simulation starts 40degrees before IVO, in order to allow the in-cylinder

flow field to adjust itself, fitting the exhaust flow.

The proposed methodology has been appliedto a low stroke-to-bore engine, operated at maximumspeed and full load. The physical soundness of the 3DCFD analysis has been assessed by comparing theresults with the ones obtained by using anexperimentally validated 1D engine model. Theagreement is satisfactory. Particularly, the difference interms of predicted intake delivered mass is 3%.

Comparison has been made between thesteady intake discharge coefficients, measured at the

traditional flow bench, and the instantaneouscoefficients, calculated on the base of the 3D CFDresults, during the valve opening period. The influenceof piston shrouding on intake permeability is verystrong, and it should be included in any accurate 1Dengine simulation. However, the resulting variations inthe boundary conditions for the 3D CFD analysis donot have a relevant influence in terms of flow ratethrough the valves.

A set of three parameters (direct tumble,reverse tumble and cross tumble) is proposed for thecharacterization of the flow field in low bore-to-stroke

engines. These parameters correspond to the strengthof three large eddies, which seems to be effective topromote turbulence production in the last part of thecompression stroke.

Finally, results obtained using the proposedmethodology have been compared to those yielded bya simplified approach (pseudo-intake). Relevantdifferences have been observed in terms of predictedmean flow field during both intake and compressionstrokes. These differences are expected to conditionthe development of combustion.

ACKNOWLEDGMENTS

The authors wish to acknowledge RicardoSoftware, Burr Ridge, IL, for the use of the VECTIScode, granted to the University of Modena and ReggioE.

The authors also wish to acknowledge EnricoNeodo and Davide Balestrazzi for the excellent workdone during their degree theses, on which this paper isbased.

APPENDIX A

The in-cylinder flow field is often characterizedby means of various non-dimensional parameters,such as the swirl ratio and the tumble ratio. In thepresent paper, Direct Tumble ratio TRd, Reverse

Tumble ratio TRr and Cross Tumble ratio TRc aredefined as the ratio of an equivalent solid body angularvelocity to the engine rotational speed.

Figure 18: Tumble parameters definition

Figure 18 highlights the definition of the threetumble ratios, as well as the domains over which theyare computed. Particularly, TRd is the ratio of theangular momentum about an axis perpendicular to thesymmetry plane passing through the center of mass ofthe exhaust side region of the cylinder, to that given bysolid body rotation of the region charge mass revvingat crank shaft speed about the correspondent center ofrotation (supposed coincident with the center of mass).

If ncells is the number of cells in the

considered region, e is the crank shaft rotational

speed, i are the cell densities, iii zyx ,, are the

Cartesian coordinates of the generic cell centroid,

ccc zyx ,, are the Cartesian coordinates of the center

of mass, and iii wvu ,, are the velocity components of

the generic cell i , the numerical implementation of theintegration over the considered region becomes:

( ) ( )[ ]

( ) ( )[ ]

=

=

+

=

ncells

i

CiCiie

ncells

i

CiiCiii

d

zzyy

zzvyywTR

1

22

1

(A1)

TRr is the ratio of the angular momentumabout an axis perpendicular to the symmetry planepassing through the center of mass of the intake sideregion of the cylinder, to that given by solid bodyrotation of the region charge mass revving at crankshaft speed about the correspondent center of rotation(supposed to be coincident with the center of mass).


13/13

( ) ( )[ ]

( ) ( )[ ]

=

=

+

=ncells

i

CiCiie

ncells

i

CiiCiii

r

zzyy

yywzzv

TR

1

22

1

(A2)

TRc is the ratio of the angular momentumabout an axis parallel to the symmetry plane passing

through the center of mass of the exhaust side regionof the cylinder, to that given by solid body rotation ofthe region charge mass revving at crank shaft speedabout the correspondent center of rotation (supposedcoincident with the center of mass).

( ) ( )[ ]

( ) ( )[ ]

=

=

+

=ncells

i

CiCiie

ncells

i

CiiCiii

c

zzxx

xxwzzu

TR

1

22

1

(A3)

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a methodology for in cylinder flow field evaluation in a low stroke

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