optimal design of the intake system in 4-S I.C.E

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SAE TECHNICAL

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International Congress & Exposition

Detroit, Michigan

February 28-March 3, 1994

The Engineer ing Society

For Advanc ing Mob i l i ty

Land Sea Air and Space

I N T E R N A T I O N A L

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940210

Optimal Design of the Intake

System in 4-S I.C.E.

Rafael Royo, Jos Corbern, and Antonio PrezUniversidad Politchnica de Valencia

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940210

Optimal Design of the Intake

System in 4-S I.C.E.

Rafael Royo, Jos Corbern, and Antonio PrezUniversidad Politchnica de Valencia

ABSTRACT

This article sets global rules for the design of thedifferent elements of the intake system for a wide range of4-s. i.c.e, examining the different phenomena related withthe gas exchange process. With this objective, a broad

analysis of a set of very different automotive engines wascarried out, using a full wave action model developed bythis same research group and widely checked bycomparison with experimental results. The design criteriapresented in the paper have been extensively andsuccessfully applied in different development projectscarried out for several automotive engine companies.

INTRODUCTION

The design of the intake system commonly implies acomplex process of synthesis which is mainly based upon

experimentation on the engine test bench. A great part ofthe bibliography concerning this subject [1], [2], [3], [4], [5],considered only specific elements of the intake system orparticular engines, so it seemed necessary to undertake awider analysis, with the final objective of trying to set upsome design criteria with the necessary universality. Inthis way, the article tries to exceed the limits of the designof specific elements of the intake system or individualengines, and instead sets global rules for the design of thewhole intake system for a broad range of 4-s. i.c.e.

METHODOLOGY

The study has been carried out using a calculationcode to model tie operation of the engine, a method whichallows an exhaustive analysis of the different factors thataffect the gas exchange process in the engine. Thispermits the easy modification of the different operatingand design parameters, from which it is possible todiscern the influence of different factors normally

juxtapositioned. In this way, the comprehension of thmost important phenomena that affects the gas exchangprocess is provided, including the influence of pressurwaves in the intake manifold upon the charge of thecylinder.

The code used for modelling the operation of the

engine is based on a full wave action model [6], onedimensional, unsteady and non-homentropic, solved bmeans of the well known Lax & Wendroff method [7Special modifications have been introduced to improvthe conservation properties of the scheme when there arpipes with high cross section variation rates [8]. Thedevelopment of this code was carried out by the sameresearch group [9], and a high accuracy in the results habeen achieved, especially in all those related to the gaexchange process.

This code has been extensively used in severadevelopment projects carried out for different automotiv

companies. The results of these studies have beesuccessfully tested on the engine bench. In figures 1 an2,

Fig. 1 Comparison of measured and calculated

volumetric efficiency.





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comparisons between measured and calculated curves ofvolumetric efficiency vs. engine speed at full load are

shown for two spark ignition engines of 1600 cm3 and

1800 cm3 of displacement. The larger one is a 4-valves

per cylinder engine.

As can be seen, a very close resemblance in theresults obtained with the use of the code is provided in theoverall operating range for both engines.

In trying to achieve the necessary universality for theconclusions of this article, a wide range of engines withvery different characteristics concerning size, geometry,operating conditions and specific application have beenanalyzed. In this way, the range of the study extends from

a multivalve spark ignition engine with 1180 cm3 ofdisplacement and 6500 rpm of maximum speed up to a

turbocharged diesel engine with 12000 cm3 of

displacement and 2000 rpm of maximum speed.The appendix summarizes the most important

characteristics of the different engines which have beenanalyzed, encompassing the following:

* spark and compression ignition,* natural aspirated and turbocharged,* different number of cylinders (four and six),* fast and slow types,* wide range of automotive application.

These engines have been extensively analyzed alongdifferent development projects carried out by the researchgroup.

The wide parametric study of the different describedengines has been carried out considering a variety ofrepresentative operating conditions (speeds and loads)and a complex combination of the different parameterswhich defines the geometry of the intake system. Thefrequency parameters which will be defined in the nextsection have been used to characterize the geometry ofthe different parts of the intake system.

The use of the calculation code has made this largeand complex study possible. The analysis of the results ofthe work, which consisted of calculating hundreds ofdifferent intake geometries for the six engines, has finallyprovided a set of rules for determining the design of theintake system.

INFLUENCE OF PRESSURE WAVES IN THE

INTAKE MANIFOLD UPON THE VOLUMETRIC

EFFICIENCY. FREQUENCY PARAMETER.

To determine the criteria for the design of the intakesystem, first an approach to the complex phenomenawhich affects the gas exchange process in the engine isnecessary. From these, the transmission of the pressurewaves along, the intake manifold is one of the mostinfluential factors upon the cylinder charge process [1][2], [10].

Considering the real engine operation during the intakestroke, the vacuum associated with the increase ofcylinder volume caused by the displacement of the piston,

produces the transmission of a rarefaction pressure wavetowards the intake pipes. This pulse travels along theintake system. towards the end of the pipe that is openedto the atmosphere where a reflection takes placeproducing a wave of the opposite sign.

In this way, a positive pressure pulse, with lessamplitude than the original one, returns through thedifferent elements to the cylinder, reaching the intakevalve with a certain delay time in relation to the originapressure wave. During this delay time, the engine rotatesa crank angle (delay angle, see figure 3), whose adequatecharacterization has a substantial importance in the study

of the charge process.This delay angle [1] could be adequately represented

by means of the parameter referred to as FrequencyParameter (here abbreviated as FP parameter) [10]which has been shown to be very useful in the study of theintake system. This parameter is defined as the relationbetween the natural frequency of the intake system andthe frequency associated with the rotation of the engine.

Assuming a simplified geometry, such as a singlecylinder engine with an intake manifold consisting only ofa straight and constant section pipe running directly to theatmosphere, the defined frequency parameter has an

explicit and useful meaning:(1)

as the expression of delay angle is

(2)

so in this case the frequency parameter is explicitlyrelated with the described delay angle:

Fig. 2 Comparison of measured and calculated

volumetric efficiency.

FPNATURAL FREQUENCY

ENGINE FREQUENCY---------------------------------------------------------------------

30 a

L a-------------= =

delay12 n L

a------------------=





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(3)

FP=4: delay angle = 90 degrees

For example, from (3) (as it is shown in figure 3), afrequency parameter of 4 implies a delay angle of 90degrees, which means that the reflected pressure wave

returns upstream to the intake valve around the BDC,producing a substantial increase in the mass flow rate intothe cylinder which provides an extra charge in thecylinder. In this case, the main part of the supplied mass isretained, since the immediate closing of the valveminimizes the effect of the subsequent back flow.

Figure 4 shows the calculated mass flow rate duringthe intake period through the inlet valve for the engine ofthe study referred to as SI1, calculated with a simplifiedsingle cylinder geometry, and running at maximum enginespeed with intake lengths leading to FP = 4, FP = 2 andFP = 10.

In figure 5, instantaneous pressure in the pipe just

upstream the valve is represented for the same enginewith the above shown frequency parameters.The value of FP, determined by the intake pipe length

and the engine speed, has a important influence on thevolumetric efficiency. For example, FP = 2 implies a delayangle of 180 which means that the reflected pressurepulse returns to the valve when it is practically closed. Asit can be seen in figure 5 for the same simplified geometry,in these conditions during the main part of the intakeperiod, a pulse of rarefaction acts upstream from thevalve, from which the volumetric efficiency finallybecomes quite poor (see figure 4). Frequency parameterwith values greater than 6 imply that the reflected

pressure pulse arrives to the valve too early (as it can beseen in figure 4 for the same engine with FP = 10). So, asubstantial proportion of the increment of charge causedby the effect of the pressure

Fig. 3 Delay angle between original and reflected

pressure pulses for FP=4.

Fig. 4 Mass flow rate through the valve during the

intake period for the SI1 engine at full load

and maximum engine speed for several

frequency parameters. (Negative mass flow

rate means flow coming into the cylinder).

FP360

delay

-----------------=

Fig. 5 Intake valve pressure during the inlet period

for the SI1 engine at full load and maximum

engine speed for several frequency

parameters.

Fig. 6 Volumetric efficiency vs. frequency

parameter at full load and maximum speed

for several studied engines.





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pulse is lost, since the back flow has a great magnitude inthese conditions (see figure 4).

It is important to point out that there is a certain optimalrange of values in the frequency parameter [3,4,5] forwhich the volumetric efficiency is substantially increasedby the effect of the pressure pulses acting upstream fromthe intake valve.

In figure 6, the calculated volumetric efficiency atmaximum engine speed and full load for the several

engines considered in this study has been represented vs.the frequency parameter of the intake pipe. The describedsimplified conditions of a single cylinder engine with astraight intake pipe have also been considered for thecalculations.

An improvement in volumetric efficiency is alwaysobtained if the reflected pressure pulses return before thevalve closes, which is accomplished when FP are greaterthan 3. This increase is progressively reduced with theincrease of FP, which implies that the length of the pipes isalso gradually decreasing. For these different engines, anoptimal range between 3 and 4.5 has been observed.Intense losses in volumetric efficiency are detected if the

reflected pressure pulses return after the valve closes,when FP are lower than 3.

The differences in the level or volumetric efficiencybetween these engines are caused by the magnitude ofthe available effective cross-section of the intake valve inrelation to the size of the engines and the operatingconditions.

The cause of fluctuations observed in the differentcurves, which takes place within the optimal range of thefrequency parameter, comes from the effect of residualpulses of previous operating cycles. They are combinedwith the actual pulses and depending upon the delayangle, amplify or dampen the amplitude of pressurewaves finally acting upstream from the valve. Thesefluctuations are not detected for multicylinder engines,since residual pulses are completely dampened by theirmovement through the whole intake system.

In actual conditions for multi-cylinder engines, thepressure waves are combined with pulses coming fromthe other cylinders and reflected in different points of themanifold. So, the resultant pressure upon the intake valveduring the inlet process has a very complex shape, fromwhich it is nearly impossible to find out the contribution ofthe different original pulses.

But in a similar way to that shown for the simplified

single cylinder geometries, the frequency parameter canalso be defined for each part of the intake manifold inmulticylinder geometries.

So, as it is shown in figure 7 for a four cylinders engine,it could be considered a first frequency parameter (FP1)similar to those described for the single cylinder engine,associated with the geometry of the runners, a secondfrequency parameter (FP2) identified with the geometryup till the filter, and finally, even a third frequency

parameter (FP3) related with the whole intake system.The expressions for these frequency parameters are

more complex [10], and are not as physicallycomprehendible as FP1.

The different frequency parameters are related to thereflection and transmission of the pressure waves in thedifferent parts of the manifold. Each reflection implies animportant dampening of the pressure pulses, so the firstreflected waves have the most influence on the finapressure acting upstream from the valve and then uponthe volumetric efficiency.

So, it could be concluded that, to improve the chargeof the cylinder, it is essential that:

* Reflected pressure waves arrive to the valve just

before it closes, obtaining of frequency parametervalues in the range of [3-4.5] for one part of theintake system (tuning) [1], [5].

* These pressure waves have the highest amplitude(ram effect) [1].

So the maximum volumetric efficiency will be obtainedwith optimal values for FP1 (which is obtained with thelength of the runners), and the level will be much greater ifthe cross-section of these pipes is small, since in this waythe reflected pressure waves have the highest amplitude.

MAIN DESIGN CRITERIA OF THE INTAKE

SYSTEM

This section discusses, from the analysis of the resultsof the study and according to our experience indevelopment projects, the methodology for improving thecharge of the cylinder with the design of the intakesystem.

From the different parameters that affect the volumetricefficiency, first, the influence of the valve-port geometrywill be analyzed.

Fig. 7 Frequency parameters and transmission of

pressure pulses in a multicylinder geometry.





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INTAKE PORT-VALVE GEOMETRY- At full chargeconditions, the most important pressure losses in thewhole intake system are produced in the flow through thevalve-port geometry.

These losses (with a critical influence on the volumetricefficiency at high engine speeds, and thus on maximumengine power) are characterized by the correspondingeffective cross-section of the intake port-valve, which ismeasured experimentally in the steady flow rig.

The required effective cross-section valve that

achieves the optimal performance for a specific enginedepends basically on its size and the operating speedrange. For high speed conditions and with an adequatemanifold design, the larger the magnitude of the intakevalve effective cross-section, the larger the volumetricefficiency.

This is the reason for the recent tendency to usemultivalve geometry engine with which, if the manifolddesign is adequate, high volumetric efficiency could beachieved at maximum speeds, from which the power willreach very high values.

But this good performance at maximum speeds hasthe disadvantage of producing poor volumetric efficiency

for medium and low speeds, since the greater effectivecross-section from the multivalve geometry causes alsoan increment in the back flow for lower speeds.

For older intake design conceptions, with valves thathave smaller effective cross-sections and lower maximumengine speeds, more compact manifolds were commonlyused. From the conclusions of the study, this tendency isjustified: if enough effective cross-section is not available,the potential improvement of volumetric efficiency fromthe optimal design of the intake manifold will be muchlower, so from the economic and manufacturing points ofview, it could be better to use smaller and more compactmanifolds.

From the study carried out, an estimation of themagnitude of the required effective cross-section of intakevalves has been made up, as a characteristic dependingmainly upon the engine size and its operating speed.

For fast engines, the values for maximum effectivecross-section must be at least about 12% of the geometricsection of the piston head. For the appropriate use ofpressure pulses in the manifold during the final part of theintake period, the available effective cross-section aroundthe bottom dead centre must still be quite high, -approximately 75% of maximum effective cross-sectionvalve.

Lower values are required for slow engines -approximately about 10% of the geometric section of the

piston head. This value could be reduced for largeturbocharged engines, such as those ones commonlyused for trucks.

INTAKE MANIFOLD DESIGN.- The most important factorfor the optimization of the cylinder charge is the design ofthe intake manifold, from which the engine performancesat full charge are completely dependent upon.

The optimization of the design of the intake manifoldhas the objective of getting the adequate use out of the

pressure pulses, trying to achieve the maximum cylindercharge for a selected operating range of the engine.

So the design of the intake manifold depends basicallyupon the selected optimization speed and the availableeffective cross-section valves.

These design factors are closely related: a greateffective area could produce loss of volumetric efficiencyat low velocities if the design of the manifold isinadequate, since the back flow at the end of the intakeperiod can have a significant influence.

From the different parameters of the design of theintake manifold, first the cross-section of the runners wilbe analyzed.

Cross section of the runners - As global criterium, the useof small cross-section pipes in the runners are alwaysadvantageous. An approximate magnitude of around 10-20% greater than the maximum effective area of intakevalve is suitable. So, a high amplitude of pressure pulsesis provided, and so, if the lengths of the runners areadequate to benefit from these effects, the final cylindercharge could be substantially increased.

In the opposite way, if the lengths of the runners and

the engine speed imply value of frequency parameterlower than 3, the selection of pipes of small cross-sectioncause an increase in the losses of volumetric efficiency.This effect is due to the enlargement in the amplitude ofrarefaction pulses, which in these conditions actpredominantly upstream from the valve during the intakeprocess, when very low values of frequency parameterare provided.

With the use of runners that have tapered sectionsfrom plenum to cylinder-head, intermediate performancesbetween those achieved with runners that have a constantsection is obtained. Its use is suitable when enoughvolumetric efficiency is necessary for a wide speed rangeThe improvement which can be obtained with its use islower than the magnitude which is achieved with constantsection runners, but it is extended to a wider speedinterval.

Manifold Design criteria.- Depending upon the operatingspeed range and the geometric requeriments, two criteriafor the design of the intake manifold are proposed in thispaper. These criteria provide the optimal use of pressurepulses in the intake system to improve the volumetricefficiency for a determined speed range.

The first design criterium is based on using theadequate values of the parameter frequency for therunners, FP1, and is the most convenient design method

since it allows the achievement of the highest volumetricefficiency.

If it is not possible to obtain adequate values for thefrequency parameter of the runners, because thenecessary lengths of these pipes are too large, thesecond design criterium could be applied, which is basedon the optimizing of the second frequency parameter,FP2, associated with the geometry of the intake systemup till the





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filter.

First design criterium - As was previously said, the use ofthis criterium provides the maximum levels of volumetricefficiency, since the amplitude of the pressure pulsesreflected in the end of the runners is the largest, but itsutilization is restricted by the required pipe lengths, whichcould be too long if the selected speed range foroptimization is too low.

This first design criterium is widely applied for fastengines of high performance, for which the achievementof high volumetric efficiency is an objective that takes

priority over other possible requirements.

When the appropriate design of the runners ispossible, and if the effective intake valve cross-section issufficient, the optimal frequency parameters FP1 willcorrespond to values from 3 to 4.5.

If the effective cross-section valve is not big enough inrelation to the selected optimization speed range, theadequate values of the frequency parameter FP1 areusually higher than 5, which implies the possibility of usingmore compact manifolds, but in this case the level ofvolumetric efficiency finally provided will be substantiallylower.

Referring to the section of the runners, as was saidbefore, an approximate magnitude of around 10-20 %greater than the maximum effective cross-section ofintake valve is suitable.

In this case it is more adequate to use tapered sectionpipes, since it provides more flexible performances andgood volumetric efficiency in wide operating ranges, whichis particulary interesting for automotive engines with broadspeed intervals. With stronger section tapering,improvements are obtained for a wider operating range,but the maximum levels will be lower.

This first design criterium also requires the use of aplenum with enough volume to provide an adequatereflection of the pressure pulses and so the sufficientamplitude of waves acting upstream from the intake valve.The minimum value of volume which is adequate to use isabout the same as the engine displacement.

In this case, the frequency parameter associated withthe whole intake system FP2 has a minor influence on the

volumetric efficiency, even less if the optimization isintended for operating conditions of medium enginespeed.

For the designs optimized for high engine speeds, theinfluence could be more important, and in theseconditions it is more convenient to get extreme values forthe frequency parameter FP2 -very close to FP1 or lowerthan 2, which implies very short or long pipes (L2)connecting the plenum and the filter.

Regarding the section of this intermediate pipe (S2 inthe figure 8), for this design criterium it is preferable thatthe effect of pressure pulses in this part of the systemshould be minimal, because they usually produce anegative influence. In this way, to reduce the amplitude ofpressure pulses in the intermediate pipe, it is convenientto use a section around 2.5 to 3 times greater than therunners.

In these conditions, the influence of the final part of theintake system, including the filter and the inlet pipe (L3), ingeneral is also not very significant.

Second design criterium - When the application of the firstcriterium is not possible, because it implies the utilizationof runners that are too long (since the speed range foroptimization is too low), the only other way is theoptimization of the frequency parameter FP2, associatedwith the geometry of the group 4-1 or 3-1 (for six cylindersengines). From the results of the study, a value close to3.5 for FP2 is the optimal parameter.

This method of design allows the adequate utilizationof pressure pulses in this part of the intake system, but itis necessary to point out that the levels of volumetricefficiency will be substantially reduced, because theamplitude of the reflected waves will be significantly lower.The only advantage is that it allows the use of runners

with as short a length as necessary to conform with theusual requirements of manufacturing.

Referring to the section of the runners, in this case it isalso convenient to use small sections (with a magnitude ofabout 10-20 % larger than the maximum effectivecross-section of the valve), but constant along the wholepipe. In this way, the amplitude of all pressure pulses ishigher, so the final effect upon the volumetric efficiencywill also be more pronounced.

For the second design criterium, it is more adequate touse manifold designs without a plenum, in order to avoidthe dampening of the waves travelling through it. Insteadit is better to use a simple but well designed junction

between the runners and the intermediate pipe, with theobjective of maximizing the amplitude of the pressurepulses which goes through this junction.

For the intermediate pipe, it is convenient to usesections (S2) that are not too big -approximately twice thesection of the runners, since it is necessary for thepressure pulses to have enough amplitude.

For slow engines with four cylinders, the final part ofthe intake system (filter and inlet pipe L3) has animportant

Fig. 8 Nomenclature which is used for the different

parts of the intake system.





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influence on the performance of the engine. It isnecessary to take into account the effect of this part of thesystem on the design process. It is preferable to avoidinlet pipes that are too short, which could haveundesirable effects at very low speeds. The section of thispipe (S3) is not too important, and it would be adequate touse slightly bigger dimensions than those of theintermediate pipe (S2).

For slow engines of six cylinders with normal intakesystems of the type 6-2-1, the final part of the geometry

does not have an important influence on the performanceof the engine, which means that the effect of interferencebetween both 3-1 groups of cylinders is not importantonce the lengths of the 3-1 part are long enough.

INTAKE VALVE TIMING - It has been observed that themodification of the intake valve timing makes it possible tohave a slight adjustment in the design of the intakesystem. In this way, some small modifications in theperformance curves of the engine are feasible for the finalfitting of the engine.

From the results of the study, and according to ourdesign experience, a global tendency has been detected:

delaying the intake opening point improves the volumetricefficiency for high speeds, and reduces it for the range oflower engine velocities. The opposite tendency isobserved from the advancing of the valve opening.

As it is shown in figure 9 for the engine referenced asSI3, calculated with a manifold designed for theimprovement of the volumetric efficiency at mediumspeed, the before detected tendencies could be describedas a rotation of the curve of volumetric efficiency aroundone engine speed, since progressively more influence isobserved when moving away from that velocity.

It is interesting to point out that, from the results of thedifferent analyzed designs and engines, the rotation

centre is usually placed very close to the speed selectedfor the optimization of engine performance with the designof the intake manifold.

So, it can be concluded that there is a significantrelation

between the influence of the intake valve timing and thedesign of the manifold. From the analysis of the calculatedmass flow rate through the intake valves, it is observedthat, for the optimal intake design, little variations in thevalve timing have opposite effects on the mass flow rateduring the intake process which are reciprocallycompensated. For this reason there is no variation involumetric efficiency for the design engine speed whenthe timing of the intake valve is slightly modificated.

BRIEF COMMENTS ABOUT THE DESIGN OF THEEXHAUST SYSTEM. - Although the main objective of thisstudy is the design of the intake system, since it is themost influence factor on the volumetric efficiency, it seemsinteresting to comment briefly upon the most importantfeatures referring to the design of the exhaust system.

First it is necessary to avoid, whenever it is possible,the interference of the exhaust process betweencylinders, since the pressure pulses produced by theopening of exhaust valve in one cylinder could act uponthe other one when it is beginning its exhaust processThis interference could seriously affect the possibility ofan adequate evacuation of burnt gases from the cylinder

To avoid this, it is necessary to use the appropriatearrangement of the exhaust manifold, usually 4-2-1 and toconnect the cylinders in the proper order to avoid pulseinterference.

It is also important to use exhaust manifolds with pipesthat have a large enough section, because this parameterhas a critical influence on final pressure losses and in thisway also on the exhaust back pressure, a parameterwhich significantly affects the maximum power of theengine.

RESULTS OF THE APPLICATION OF THE

ESTABLISHED CRITERIA FOR THE DESIGN OF

THE INTAKE SYSTEM IN THE STUDIED ENGINES

Some of the results obtained from the utilization of thedescribed design criteria will be described in this sectionThey are applied to the same engines used initially in thestudy, with the objective being to achieve the maximumvolumetric efficiency for several characteristic ranges oftheir operating intervals.

The results of the volumetric efficiency obtained witheach predesign are compared with the averaged value forthe whole speed range corresponding to the engine intheoretical operation without manifolds. This level issimilar for all the considered engines, around a value of0.8, excepting for the CI3, turbocharged engine, for whichthe volumetric efficiency is higher, 0.9, because of thescavenging produced by the high intake pressure.

For these calculations, pressure losses have only beenconsidered in the head cylinder. In the manifold, straightpipes have been assumed, and other pressure losses arenot taken into account. This is the reason for the highcalculated levels of volumetric efficiency.

For the first four engines, SIs and CI1, of fast type, the

Fig. 9 Influence of little variations of the intake

timing on the curve of volumetric efficiency.





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first design criterium has been used: the length of therunners has been calculated to get adequate values of

FP1, around 4.5, and with a tapered section from plenumto cylinder-head. For the plenum, a volume equal to theengine displacement has been selected.

With the SI1 engine, as it can be seen in figure 10, thedesign options proposed achieve a very high level ofvolumetric efficiency for the selected ranges ofoptimization -medium speeds for geometry 1, andmaximum speed for geometry 2. With this second design,high values of maximum power can be obtained. Theseelevated values of volumetric efficiency are possiblebecause this engine has large effective cross-section ofthe intake valves, according to its multivalve geometry.

The results obtained for SI3, as is shown in figure 11are quite a lot worse, mainly at high speeds, and theimproved range around the selected optimization engine

velocities are

quite narrow. A similar tendency is observed for CI1. Theproblems for both engines are caused by the poormagnitude of the available effective cross-section of theintake valves in relation with the high maximum enginespeeds. So, as can be seen in figure 11, with thishandicap it is very difficult to improve the charge at highengine speeds -only a limited optimization is achieved atmedium speed with both geometries proposed.

For both of the slow engines which have been studiedCI2 and CI3, the second design criterium has beenapplied to maximize the volumetric efficiency at mediumand low speeds, optimizing the value of the secondfrequency parameter FP2, corresponding with thegeometry of the intake system up till the filter. In this case,the lengths of the runners can be as short as is required,but the determining parameter is the length of theintermediate pipe connecting the plenum and filter (L2)With this

Fig. 10 Calculated volumetric efficiency with the

proposed intake system geometries for the

SI1 engine.



SI3 engine.


proposed intake system geometries for theC12 engine.



CI3 engine.





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methodology of design, as is shown in figures 12 and 13,the levels of volumetric efficiency obtained are very high,mainly at low speeds (results of geometry 1, figure 12),but the range of improvement is narrower and a sharploss of volumetric efficiency is observed at high speeds.

For the CI2, geometry 1 (as it is shown in figure 12) isselected to improve the charge at low speeds, andgeometry 2 is chosen for medium speeds.

As can be seen in the same figure 12, the manifold

optimized for low engine speeds gives very poorvolumetric efficiency for higher velocities.The proposed intake system for CI3, the turbocharged

engine, produces a strong maximum of volumetricefficiency at medium speeds, which is very appropriate forthis type of engine, since it allows an instantaneouscharge of the cylinder in this speed range, thus avoidingthe inertia of the turbocharger.

CONCLUSIONS

* The most important factor for the optimization of thevolumetric efficiency is the design of the intakemanifold, with the objective being to adequately usethe effect of the pressure waves in the pipes.

* The improvement in volumetric efficiency which canbe obtained in a speed range with the design of theintake manifold, are directly depending on theavailable effective cross-section of valves.Increasing the effective area, the improvementcould also be extended to a wider speed range.

* From the results of the study of very differentautomotive engines, two criteria for the design ofthe intake manifold have been established, which

imply the optimal use of the pressure waves in therunners or in the whole geometry of the manifold.

* With the first design criterium, which optimizes thedesign of the runners, the maximum level ofvolumetric efficiency is achieved for a selectedspeed range.

* The application of this first criterium is limited by therequired lengths of the runners, which can be toolong if the selected engine speed range is very low.

* In the above cases, the second design criteriumcan be used, providing an optimal frequency

parameter FP2 for the manifold, which mainlyimplies the use of an adequate length for theintermediate pipe connecting the plenum and thefilter.

* Both design criteria are based on getting values offrequency parameters FP1 or FP2 close to 4, whichmeans that the reflected pressure wave returns tothe intake valve around the bottom dead centre.

* In this paper some other instructions for the designof the rest of the intake system are also detailedsuch as the section of the pipes and the dimensionsof other elements.

REFERENCES

1. Winterbone, D.E. A comparison of synthesis andanalysis models for wave action manifolds. 1989,

IMechE C372/037.

2. Margary, R. Nino, E. Vafidis, C. The effect of IntakeDuct Length on the In-Cylinder Air Motion in aMotored Diesel Engine. SAE paper 900057, 1990.

3. Heinz Duelli, W. Doctoral Thesis: Berechnungen undversuche zur optimierung von ansaugsystemen frmehr-zylindermotoren mit einzelzylinder-einspritzung, 1984.

4. Morel, T. and others; Characterization of ManifoldDynamics in the Chrysler 2.2 S.I. Engine by

Measurements and Simulation. SAE paper 9006791990.

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Computational fluid mechanics and heat transferHemisphere Publishing Corporation, 1984.

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SYMBOLS

a Sound speed.FP Parameter of frequency.L Length of the pipe.n Engine speed (rpm).

APPENDIX. MAIN CHARACTERISTICS OF THE

ENGINES ANALYZED IN THE STUDY

ENGINE:TYPE

SI1(A.N.)

SI2(A.N.)

SI3(A.N.)

CI1(A.N.)

CI2(A.N.)

CI3(T.C.)

CYLINDERS 4 4 4 4 4 6

PISTONBORE (mm)

82 88 80 87 100 130

STROKE(mm)

83.5 82 80 96 127 150

VALVETIMING

IVO 2 6 10 -3 3 -7

IVC 43 41 30 25 29 27

EVO 39 48 50 42 41 48

EVC 2 2 -10 -10 2 -4

COMPRESIONRATIO

10 9.2 10.2 21 16 15.5

MAXIMUMSPEED(Rpm)

6500 5500 5750 4300 2800 2000




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