Some Critical Technical Issues on the Steady Flow Testing of Cylinder Heads

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SAE TECHNICALPAPER SERIES 2001-01-1308

Some Critical Technical Issues on the

Steady Flow Testing of Cylinder Heads

Hongming XuJaguar Cars

SAE 2001 World CongressDetroit, Michigan

March 5-8, 2001


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2001-01-1308

SOME CRITICAL TECHNICAL ISSUES ON

THE STEADY FLOW TESTING OF CYLINDER HEADS

Hongming Xu

Jaguar Cars

Copyright 2001 Society of Automotive Engineers, Inc.

ABSTRACT

There are considerable diversities in the techniques usedfor the steady flow testing of engine cylinder heads, and

this paper presents and discusses the important issuesinvolved in the flow bench experiment. The work aims toprovide information necessary for setting up or upgradingthe experimental system of cylinder head testing. Thedefinitions of discharge/flow coefficients and swirl/tumbleratios are compared and examined, followed by theprinciples of selecting the test conditions such aspressure drop and flow rate. Techniques for measuringthe angular flow momentum in cylinders are discussedand the link between the steady flow parameters and theengine combustion performance is highlighted. Someconclusions and recommendations are drawn from thediscussion.

1. INTRODUCTION

The steady flow testing of cylinder heads is a widelyadopted procedure in the development of engines [1, 2]and it is used to assist and assess the design of theengine ports and the combustion chamber concerningthe engine flow capacity and the in-cylinder flow patternof the charge motion, which are critical to the enginecombustion performance. Although considerable effortshave been made by research workers to explore themost effective methodology for steady flow tests, thereare considerable diversities in the definitions of thetechnical terms and techniques used in the presentexperiment [1, 3], and the configurations of the flowbench vary considerably with users.

The absence of a standard methodology has obviouslyraised difficulties in the interpretation of available dataand prevented comparisons between the intake flowscharacterized by different engine groups [4]. Forexample, a swirl ratio of 3 quantified by one group couldmean a different value to another, and the ambiguity withthe tumble ratios is even greater. In the enginedevelopment process, there is often a need to make

reference to bench mark designs or data published in theliterature. It is important, therefore, to understand theoriginal definitions of the terms and the effect of theexperimental techniques on the result. The experimenta

techniques for the steady flow bench test and theiimplications have been discussed by the early work of [35, 6], and more recently by [1, 4, 7] but a criticacomprehensive review of the important issues is noavailable.

This paper presents and discusses the importantechnical issues involved in the steady flow bench testThe purpose of the work is to provide the informationnecessary for setting up or upgrading the experimentasystem, either by selecting a commercially availablebench product as often practised by industry, or bydesigning a flexible piping system which is usually a low

cost solution for university research groups. The criticainformation is also expected to be useful to engineersworking on the development of engine cylinder headsparticularly those involved in the steady flow testsBecause there are so many techniques which have beenproposed and used in the practice, it is sensible tochoose only a few representative ones for review. TheRicardo and AVL techniques are discussed here sincethey are probably the most widely used in automotiveindustry.

The text below is divided into six sections. Following thisintroduction, the definitions of discharge and flowcoefficients are presented. The third section is devotedto the methodology of defining the swirl and tumbleratios, and the fourth section discusses the selection oftest conditions such as pressure drop, flow rate and alsothe issues concerning blowing and suction systemsThen, the experimental techniques for measuring theangular flow momentum flux is reviewed in the fifthsection, and the link between the steady flow bench tesand engine combustion performance is discussed in thesixth section. Finally, the main conclusions andrecommendations drawn from the discussion aresummarized.


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2. DISCHARGE COEFFICIENT AND FLOW

COEFFICIENT

A basic requirement of the port and valve assemblydesign is to enable the engine to have high volumetricefficiency for achieving high torque and power. Duringthe early stage of engine development, the performanceof the port/valve assembly in terms of the air flowcapacity is usually assessed under steady flow testconditions, using the ratio of the measured mass flowrate to the theoretically calculated flow rate through areference flow area in the port/valve assembly. This flowrate ratio is commonly called discharge coefficient orflow coefficient, depending on which flow area is usedas the reference, Table 1.1. The 'discharge coefficient',Cd, refers to the flow rate ratio referenced to the gap

between the valve lips and the valve seats, and thereseveral ways of defining this gap area [9]. The 'flowcoefficient', Cf, refers to that corresponding to the flow

area in the port, either the minimum flow area (portthroat) or the valve inner seat area, in which the valvestem blocking effect can be included or neglected. Theyare actually alternative ways of expressing the same

data, but it will be shown below that there are somedifferences in terms of their emphasis of presentation.The calculations of Cdand Cfare simple, once the actual

flow rate is measured and the reference area isdetermined.

The choice of the reference flow area should not bearbitrary and the name of the 'flow coefficient' should notbe confused with the 'discharge coefficient', although ioften is. For conciseness, this paper uses the term 'staticflow coefficient' when addressing the two at the sametime. In particular, the definition of discharge coefficientseems to have indeed been a convention or arbitrarychoice for different engine groups. Therefore, whenevaluating discharge coefficients, one must define ocheck the reference area carefully. This is an unfortunate

situation which has often prevented the comparisonbetween published data. The author recommends usingboth the discharge coefficient and the flow coefficient

which are based on the valve inner curtain area (DvLvand the valve inner seat area (Dv2/4), respectively, fothe following reasons:

1. The 'discharge coefficient' (Cd), which decreases

with valve lift, reflects the flow restriction produced by thevalve and seat lips at low valve lifts which then determinethe flow orifice area. In fact, the geometry of valve andseat lips are critical to the flow at lower valve lifts [8, 9]

Using the valve inner curtain area (DvLv) makes thecomparison between data irrespective of the valve seatangle and it also allows a linear relationship between thetheoretical flow and the valve lift for easy interpretation

Table 1.1 Parameters for evaluating flow breathing capacity

Common

name

Reference Equation Feature

DischargeCoefficient

Valve gap area, whichhas several definitions[1]

Cd =&m

A Vv o (2.1a)

Vo =2p

(2.1b)

Value decreases with lift,identify flow restriction by

valve and seat lips

Flow Coefficient

Valve inner seat area Cf =&m

A Vp o (2.2a)

Vo =2p

(2.2b)

Value increases with lift,identify flow restriction byport geometry

Ricardo MeanflowCoefficient

Averaged over thetime between IVO andIVC Cf=

C df

1

2

2 1

(2.3)

Overall port efficiency

AVLMeanCoefficient

Averaged over thetime between TDC andBDC

( )C

C

C Cdf

m f

=

1 1

0

3

2

1

2

(2.4)

Overall port efficiency

Gulp Factor(Mach Index)

Local sonic speedZ=

B

D

S

n a Cv

s

f

2

2 1 1(2.5)

Overall restriction of theports at rated enginespeed


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1. The 'flow coefficient'(Cf), which increases with valve

lift, reflects the restriction by the port geometry, when thegap area between the valve and seat lips becomescomparable to or beyond the port throat area. Since alarge proportion of charge enters the cylinder at highervalve lifts, the influence of the 'flow coefficient' to theengine breathing capacity is more important. Using C fonly, however, will not reflect clearly the difference in theflow capacity at smaller valve lifts simply because thescale of C

f at low valve lift is too small. This is

demonstrated in Fig 1 which presents the static flowcoefficients of two cylinder heads. Compared withcylinder head one, cylinder head two has a poorer designof the valve and seat geometry which led to flowseparation with a low discharge coefficient at low valvelifts. This difference can be often omitted if only C f is

used.

Other flow parameters such as Mean Flow Coefficientand Gulp Factor have been proposed to providenormalized parameters for assessment. Ricardo andAVL have different definitions for the mean flowcoefficient [10], and the main differences are that

Ricardo consider the intake process to start at IVO(intake valve open) and end at IVC (intake valve close)and the total intake flow quantity is dependent of camprofile and valve open duration (see Table 1). AVLassume that the intake process takes place onlybetween TDC and BDC for 180 crank angle degrees andthe instantaneous flow velocity in the valve gaps isproportional to the instantaneous piston speed. AVLdefined a 'standard lift curve' for the flow parameterintegration. Therefore, a simple conversion between theRicardo and AVL flow parameters is not possible.

The gulp factor, also called the inlet Mach index, is the

ratio of the mean effective flow velocity in the port throatduring intake process to the local sonic speed. Itcorresponds closely to the mean Mach number in thevalve throat. Taylor et al correlated volumetric

efficiencies (v) measured on a range of engine andintake valve designs with the gulp factor and found that

0 1 2 3 4 5 6 7 8 9 100.0

0.2

0.4

0.6

0.8

1.0

C f 1

C f 2

C d 2

C d 1

Cd

,Cf

valve lift (mm)

Figure 1 Typical profiles of discharge coefficientand flow coefficient [8].

the vdecreases rapidly for Z 0.5 [9]. This limit shouldbe re-examined for the modern four valve cylinder headdue to the difference in the ratio of surface to volume othe inlet port. The basic requirement is to ensure that theflow will not become choked at high engine speeds.

3. SWIRL RATIO AND TUMBLE RATIO

BACKGROUND

Evaluating swirl or tumble strength in engine cylinders isfar more difficult than assessing the engine breathingcapacity. The main issues concerned are firstly thetechniques for measuring the flow, as will be discussedin the fourth section, and secondly, the method tocalculate the data, which involves the definitions of SwirRatio' and 'Tumble Ratio. The process to define thetumble ratio is actually 'borrowed' from the conventionaway of defining the swirl ratio. This section discusses thedefinitions, in which the term 'swirling flow' is used torefer to both the swirl and tumble motion.

There are a number of different ways of defining swirand tumble ratios, and the two methods widely used inautomotive industry are the Ricardo and AVL systems. Itis necessary to be aware of the difference between thetechniques used by Ricardo and AVL in in-cylinder flowstudies.

DEFINITIONS

The swirling flow is usually characterized by the momentof angular momentum about a chosen axis. The angularmomentum flux in the engine cylinder, G, is a function ofcrank angle during the induction process. In the steady

flow test, it is a function of valve lift for a given flow rateor pressure drop. The linear ratio of this angulamomentum flux to the fictitious engine speedcorresponding to the test condition is called the 'Rig SwirNumber' and 'Stationary Swirl Number' in the Ricardoand AVL systems, respectively. For the fictitious enginespeed, Ricardo adopted the isentropic ideal velocityacross the port and AVL used the axial flow velocityequal to the mean piston speed, which led to theequations for calculating the swirl parameters in Table 2.

Rewriting the equations in Table 3.1, it follows that

NN

SR

SA

=SB

CDn f2

vv

where, n-number of inlet valves, D- inner diameter oinlet valve seat.

Using geometry of a 4.0 liter V8 engine, we have

N

N

SR

SA

= 0.92 Cf


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Table 3.1 Ricardo and AVL swirl parameters

System Name of the swirl

parameter

Definition Equation

Ricardo Rig swirl numbertangential swirl velocity

ideal velocity

NSR =8

0&mBV

G (3.1)

AVL Stationary swirlnumber

speedenginefictitious

rotationvortexequivalent

NSA=

22

S

Q G (3.2)

where, &m - mass flow rate, B - cylinder bore, Vo - ideal velocity head, S - engine stroke, - charge density, Q -volumetric flow rate, G - flow momentum torque.

This indicates that the Ricardo rig swirl number is smallerthan the AVL stationary swirl number, approximately by afactor of the corresponding flow coefficient at each valvelift. Note that the value of Cf varies with the valve lift.

The most often used swirl parameter is the engine swirlratio, which has been broadly defined as

Rs=Charge vortex rotation speed ( )

Engine speed ( N / 30)

s (3.3)

where N - crankshaft revolution speed, rpm.

Under steady flow test conditions, the vortex rotationspeed is commonly calculated assuming that the chargemotion is a solid body rotating flow which, at the end ofinduction process, has momentum equal to the sum ofthe angular momentum introduced during the wholeinduction process. The swirl ratio is calculated by

integrating the rig or stationary swirl numbers as afunction of crank angle during the induction process andthen divided by the fictitious engine speed. Omitting thederivation of the equations, we have

Ricardo Swirl Ratio

RSR=BS

n D

N d

dv v

2

SRC

C

f

f

2

1

2

1 2

(3.4)

AVL Swirl Ratio

RSA

=1

0

2

NC

CGdSA

m

( )

(3.5)

where 1, 2 - inlet valve open and close position crankangle respectively; C(), Cm - instantaneous and meanpiston speed respectively.

The differences in the definitions of the Ricardo and AVLswirl ratios are partly due to the different assumptionsmade by the two groups on the engine induction andpartly due to their own way of integrating the moment ofthe flow momentum flux. Ricardo defined the flow asoccurring between IVO and IVC and therefore the swirratio is affected by the valve lift profile. AVL assumed

that the suction flow occurs only between TDC and BDCand the measurement corresponding to the 'standardvalve lift' profile mentioned earlier is also used in thecalculation of their swirl ratio. A simple conversion fromthe Ricardo swirl ratio to the AVL swirl ratio is again notpossible, since the cam profile must be consideredUnlike AVL, Ricardo defined a smaller rig swirl ratio at agiven valve lift but the definition of the engine swirl ratiohas covered a longer integration period, and it isnormalized by the engine geometry data including thebore, stroke and intake valve diameter. Note that theAVL swirl ratio is related to the engine stroke (see Table2).

For a tumble motion, the flow angular momentum flux ismeasured as for measuring the swirl, using thetechniques described in the next section. The Ricardoand AVL tumble ratios are defined in the same way asfor the swirl ratios in Eq (3.4) and (3.5).

4. EXPERIMENTAL TECHNIQUES

The discussion of the experimental techniques in asteady flow test mainly concerns the measurement of theangular momentum flux. The characterization of flowstructures is less sensitive to the flow conditions such as

pressure drop and mass flow rate than the flow capacitytest. It is known that the swirl flow angular momentumincreases linearly within the range of pressure drop from220 to 2400 mm water gauge [1]. The techniques fomeasuring the swirl ratio are relatively simple and theyhave been 'transferred' into the measurement of tumbleSpecial attention must be given so that thecharacterization of the tumble flow has as muchrelevance as possible to the combustion performance inthe engine, since the tumble motion in a steady flow isnot a well defined flow pattern. This section thus startswith tumble measurement.


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4.1 TUMBLE ADAPTOR

The measurement techniques used for characterizingtumble are not standardized. The commonly usedmethods for measuring tumble is to use a tumbleadaptor, such as those used by Imperial College andRicardo [7, 10], which converts the tumble motion in thecylinder into a swirling flow in the extended perpendicularpipe. AVL prefer to use in-cylinder velocitymeasurements from LDV and FEV prefer to use an in-cylinder paddle wheel or ring. The advantage of using atumble adaptor is that the in-cylinder flow is reflected bythe top surface of a dummy piston, simulating the flowpattern in the engine cylinder. However, the detailedgeometry of the tumble adapter, which significantlyaffects the measurement, varies considerably with users.The commonly used configurations are the "T-pipe" andthe "L-type", Fig 2, and the positions where the flowmomentum measurement is taken also vary with users.Ricardo recommend that the distance between thecenter of the cross pipe and the firing surface of cylinderhead be one half of the cylinder bore plus 20mm(probably for the flange thickness) and the swirl torque

meter be situated at about 0.5 m downstream of the 'T'junction [10].

The measurement of tumble using a flow adptor has itslimitations. Firstly, the conversion of tumble into swirlingflow is subject to flow velocity head losses within thetumble adaptor and in the conversion pipe. Secondly,when the swirl impulse meter is used, the conversion ofthe flow angular momentum to a mechanical torque issubject to errors, although of a smaller magnitudecompared to a paddle wheel anemometer [1, 3]. For LDVmeasurements in the extended conversion pipe, theintegration of the velocity profiles is subject to theassumption that the axial flow is uniform, which wasfound not to be true [4, 11]. Of course, the secondproblem also exists for the measurement of swirl.

A A

IN

B

BB - B perspex plat e

EX

EXIN

IN

swirl torque meter

perspex

A - A

EX

perspex

(a) (b)

Figure 2 Configurations of tumble adaptor(a) the "T" type; (b) the "L" type [4]

Overall, due to the existing uncertainties in quantifyingthe tumble strength under steady flow conditions with atumble adaptor, it can be argued that the interpretation othe tumble ratio measured in flow rigs is not as straighforward as for the swirl ratio. Nevertheless, the simplicityof the tumble adaptor for characterizing cylinder headsduring engine development stage under steady flowconditions is such that it outweighs the above limitationsThere are a number of evidences which have proven thetumble adaptor a useful tool for designing advanced

spark-ignition engines [4]. The issue regarding thecorrelation of the flow measurement to the combustionperformance will be discussed in section 6.

4.2 PADDLE WHEEL AND SWIRL TORQUE METER

The most often used instruments for measuring swirl arethe paddle wheel and the flow torque meter, Fig 3a and3b. The paddle wheel has a longer history and theconfiguration varies significantly. The onlystandardization appears to be in aligning the paddlewheel axis with the cylinder axis [1]. It has also beenproposed to use a paddle ring to measure the tumble

motion in the engine cylinder, for example FEV [12]There are, however, many reports showing that thepaddle wheel significantly underestimates the swirl ratioby up to 60% due to disturbance of the flow, friction ofthe wheel bearing and slip between the vanes and theflow [1]. The swirl torque meter, on the other hand, isnow widely adopted. It features a flow straighteningelement, on which the flow angular momentum about thestraightener axis is turned into the restraining torqueNote that the measurement of swirl is sensitive to theposition of the torque meter. Ricardo and AVLrecommend that the distance of the meter from thecylinder head firing surface should be 1.75 times of thecylinder bore, as a compromise between allowing theswirl to develop and limiting the swirl decay. ImperiaCollege recommend that the swirling flow be measuredat 3 times of the cylinder bore downstream to allow theswirl to settle so that the flow center moves closer to thecylinder center [7, 13].

Although the intake flow patterns in engines arecommonly characterized by the ratios of swirl, tumbleand cross tumble, which are the normalized angularmomentum moments about the three perpendicular axesin the cylinder, the steady flow test usually measures onlyone flow component. This is acceptable for engineswhich are designed to have only one dominant flow

pattern. If a complex flow pattern such as an inclinedswirl is involved, e.g., when valve disablement isemployed in a multi-valve combustion chambermeasurements have to be taken on more than onecomponent. Tipplemann, who originally proposed theconventional swirl meter, has developed a newmeasuring system with a spherical honeycomb whichallows measurement of all the three components of theintake flow in engine cylinders simultaneously [14, 15]. Aschematic diagram of the device is shown in Fig 3c. Noinformation or review about the practical application ofthe so called '3-D flow meter' has been published.


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(a) (b) (c)

(d)

Figure 3 Devices for measuring flow momentum flux

(a) paddle wheel(b) swirl torque meter

(c) 3-D momentum meter(d) rig swirl number variation with valve lift [4]

The advantage of the 3-D flow meter is obviously that itsaves the amount of work required in the multi-component measurements. For sure, the measurementof swirl with this meter will be similar to its predecessor,the one component swirl meter. The main concern is thatthe pivot of the spherical element of the swirl meterseems to have to be located in a position below theequivalent BDC piston boundary due to its structure.There is no simulated deflection by a dummy piston. Thisimplies that it can measure the flow momentum flux but

with the moment about a point different from the in-cylinder flow vortex center conventionally defined. It isimportant to emphasize that the moment of momentumof a flow depends on its integration center. Obviously, weexpect that the in-cylinder bulk flow vortex center isaround the cylinder geometry center in an ideal case,although the real flow can be much more complex.Further investigations are required to study the 3-D flowmeter's calibration and correlation with respect to othermore widely used techniques. It would be a very usefulmeasurement device if the relationship between the '3-D'measurement and the flow momentum data interested tothe engine development can be found.

4.3 OPTICAL METHODS

Commonly used optical techniques for measuring theswirl include laser Doppler velocimetry (LDV), particleimage velocimetry (PIV), particle tracking velocimetry(PTV) and other visualization methods. Figure 4 shows aLDV set-up for testing the Jaguar V8 cylinder head andthe measured rid swirl number was comparable with thausing a swirl torque meter, Fig 3d [4]. The advantagesand disadvantages of these techniques are weldocumented in literature. It must be noted that afterobtaining sufficient measurements of the flow field, themethods of processing the data affect significantly theestimate of swirl and tumble ratios. Indeed, theevaluation of the moment of momentum flux for a flow inwhich the swirl and axial velocities only vary with radiusis unambiguous [1], but the results of the integration ofthe momentum flux for an off-centered swirling flowdepend on the choice of the integration center. Both thegeometry center and the flow center have been used inthe published literature. Using LDV measurements, AVLrecommend that the center of integration be at adistance equal to half the bore below the cylinder head

surface [16].

Integration about the geometry center is one option, buone should bear in mind that the calculation does notnecessarily capture the essence of the flow patternbecause a perfectly centered flow is rare in enginesJackson et al at Ricardo [17] found that the flowmomentum integration about the center-point (either mid-cylinder or eddy center) can produce misleadinginterpretation of the flow characteristics and they proposeto use a tumble ratio based on vorticity obtained fromtheir PIV measurement. Obviously, one should find asuitable technique for a particular engine design toachieve the best correlation, and further investigations onthis issue are required.

Figure 4 A LDV set-up for testing the Jaguar V8cylinder head


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4.4 OTHER CYLINDER HEAD TESTS

The conventional steady flow test is used to assess theintake port design in terms of the static flow coefficientsand swirl and tumble ratios. Generally, a static flowcoefficient of a port/valve assembly refers to the intakeflow from the port into the cylinder through the inlet valve.As input data for modeling, static flow coefficients for theback-flow are also useful. The back-flow coefficientscan be tested with the flow reversed in the port. Thesteady flow rig can also be used to examine the staticflow coefficient of the exhaust port and the cylinder headwith manifolds.

While the steady flow bench test has become astandard technique in the development of cylinderheads, there are some other experimental systemswhich fall between the steady flow rig and the operatingengine. They have demonstrated good abilities inrevealing the fluids mechanics during the intake processof engines and here are two interesting examples. Thefirst is that employed by Imperial College [18], Fig 5a,which was a pulse flow arrangement with the overhead

camshaft driven by an electric motor. The multi-cylinderengine cylinder head was coupled to a custom-builtadapter with a plenum tank connected to a suction fan.The main advantages of this system are: (1) the dynamiceffect of the flow system can be studied, including that ofthe multi-runner manifold; (2) optical access can beeasily provided for optical diagnostics; (3) the runningcost is relatively low compared against its testingcapability. The pulsating flow has also been used toexamine the exhaust port design [19].

The second important experimental system is the wateranalogue rig used by Ricardo [17], Ford [20], GM [21],and VW [22]. The set-up of the rig at Ford SRL is shownin Fig. 5b. It requires thermal, geometrical, dynamic andkinetic similarities to simulate the gas flow, and the

Reynolds number (VpB/) and Strouhal number (B/tVp) inthe experiment must be the same as in the engineoperation. The water analogue slows down the flowprocess and significantly reduces the requirement of themeasuring time scales, allowing the use of ParticleTracking Velocimetry (PTV) or Particle ImageVelocimetry (PIV). The piston moves only during theintake and the exhaust processes and there are nocompression or expansion processes because of theincompressibility of water. Ford and Ricardo have shownthat the dynamic water rig resulted in a much better

correlation between the flow parameters and thecombustion performance than the steady state flowbench. It should be noted that Ford and Ricardo useddifferent data interpretation techniques after theyobtained the velocity data in the cylinder. A little more willbe said about this issue following the next section whichdiscusses the steady flow test condition.

5. PRESSURE, PRESSURE DROP AND FLOW RATE

It has been widely adopted that the steady flow test is

(a)

(b)

Figure 5 Transient flow bench for cylinder headtesting

(a) air flow [18](b) water analogue [20]

carried out under conditions of a constant pressure dropof 254 or 508 mm water gauge (WG). Larger pressuredrops up to 1727 mm (68 inches) WG have been usedby Ford [23]. It may also be interesting to be aware that aconstant flow rate is sometimes preferred as a steadyflow condition and advantages have been claimed fosuch an arrangement. The pressure drops can be

created by either blowing or suction. Some enginegroups including Ricardo prefer a blowing systemwhereas others including AVL use a suction system. Tohighlight the principles for selecting the test conditionsthis section discusses the effect of the pressurepressure drops and flow rate on the steady flow testingresults.

5.1 REYNOLDS NUMBER

It is known that the Reynolds number is one of the keyparameters in designing experiments in fluids since non-


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2.0x104 4.0x104 6.0x104 8.0x104 1.0x105 1.2x105 1.4x105 1.6x105 1.8x105 2.0x1050.0

0.1

0.2

0.3

0.4

0.5

0.6

0.7

1mm

2mm5mm

7mm

9mm

11 mm

Flowcoe

fficient(Cf)

Reynolds number

Figure 6 Flow coefficients as a function of Reynoldsnumber at valve lip

dimensional properties of the fluids are similar as long asthe Reynolds number is the same [24]. A requirement ofthe steady flow test condition is that the flow is fullyturbulent so that the loss of velocity head is independentof the pressure drop or flow rate. Experiments haveshown that the flow head loss in a pipe flow is due to

1) friction loss defined by the Darcy-Weisbachequation

hf = fL

D g

V2

2(5.1)

2) head losses caused by bends, elbows andvalves

hc = KV2

2g(5.2)

where fis a function of Reynolds number and the relativeroughness of the pipe surface; K is determined by thegeometry of the flow passage. Nikurade [24] proved thatthe friction loss of a turbulent flow remains constantwhen the Reynolds number in the range of 20,000 -

200,000 for a relative roughness of the pipe wall of 1/30 -1/250. This is in consistence with the work of Annandand Roe [25] which shows that the discharge coefficientremains fairly constant for high Reynolds numbers. Dataavailable within Jaguar showed a slide of the flowcoefficient with the Reynolds number and an explanationis to be found, Fig. 6.

Three Reynolds numbers can be defined in the steadyflow test for cylinder heads:

For flow in the port

Rep =

V dp p =

42

Q

n dd

p

p

=4 &m

n dv p (5.3)

For flow in the valve gap

Rev =

LVv =

Qn D L

Lv v

v

=4 &m

n Dv v (5.4)

For flow in the cylinder

Rec =

BVc =

Q

n BB2

=4 &m

n Bv (5.5)

Ricardo [10] reported that the non-dimensional propertiesof the engine port flow become substantially independenof the pressure drop when the 'Port Reynolds Numberexceed 60,000 at low valve lifts and 90,000 at high valvelifts. For a valve inner diameter of around 30 mm, theiestimated minimum pressure drop required to achievethis range of Reynolds numbers is about 10 inch watergauge, which has been widely used [2]. On the otherhand, according to Eq (5.3) to (5.5), the Reynolds

number is a function of the mass flow rate ( &m ) only for a

given flow configuration. It thus seems to be that thecylinder head can be tested with a constant mass flowrate for a chosen Reynolds number larger than athreshold. To make a choice, a comparison is neededbetween the constant pressure drop and the constantflow rate methods.

5.2 PRESSURE DROP AND FLOW RATE

Experiments have been carried out by Vafidis et al [5, 6to compare the steady flow discharge coefficientsmeasured with different constant pressure drops and a

constant flow rate, respectively. The comparison wasalso made to the dynamic coefficient obtained in anoperating engine. It was found that at lower valve liftsL/D


11/16

constant pressure drop, and thus the test result was lessdependent on the test condition.

The earlier work at Imperial College, Figure 7, shows thatthe influence of the pressure drop across the valve onthe discharge coefficient (Cd) was more significant at

lower valve lifts and smaller pressure drops (p) with aclear tendency that Cdincreased with p [6]. It was foundthat for pressure drops larger than 250 mm H2O, thestatic flow coefficient became independent of thepressure drop, except for the smaller valve lifts of L/D

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Some Critical Technical Issues on the Steady Flow Testing of Cylinder Heads