Rotor Passage Automotive Torque Converter High Frequency …downloads.hindawi.com/journals/ijrm/2001/424583.pdf · 2019-08-01 · Measurement Part I: FlowField at the Design Condition

International Journal of Rotating Machinery2001, Vol. 7, No. 4, pp. 253-269Reprints available directly from the publisherPhotocopying permitted by license only

(C) 2001 OPA (Overseas Publishers Association) N.V.Published by license under

the Gordon and Breach Science Publishers imprint,member of the Taylor & Francis Group,

Flow Field in the Turbine Rotor Passagein an Automotive Torque Converter Based on the High

Frequency Response Rotating Five-hole ProbeMeasurement

Part I: Flow Field at the Design Condition (Speed Ratio 0.6)

Y. F. LIU, B. LAKSHMINARAYANA* and J. BURNINGHAM

Center for Gas Turbines and Power, The Pennsylvania State University, 153 Hammond Building, University Park, PA 16802

(Received 11 May 2000; In finalform 23 May 2000)

The relative flow field in an automotive torque converter turbine was measured at threelocations inside the passage (turbine 1/4 chord, mid-chord, and 4/4 chord) using a high-frequency response rotating five-hole-probe. "Jet-Wake" flow structure was found inthe turbine passage. Possible flow separation region was observed at the core/suctionside at the turbine 1/4 chord and near the suction side at the turbine mid-chord. Themass averaged stagnation pressure drop is almost evenly distributed along the turbineflow path at the design condition (SR 0.6). The pressure drop due to centrifugal andCoriolis forces is found to be appreciable. The rotary stagnation pressure distributionindicates that there are higher losses at the first half of the turbine passage than at thesecond half. The major reasons for these higher losses and inefficiency are possible flowseparation and a mismatch between the pump exit and the turbine inlet flow field. Thefuel economy of a torque converter can be improved through redesign of the core regionand by properly matching the pump and the turbine. The Part of the paper deals withthe design speed ratio (SR:0.6), and Part II deals with the off-design condition(SR: 0.065) and the effects of speed ratio.

Keywords: Torque converter; Turbine; Flow measurement; Rotating frame of reference; Five-holeprobe

INTRODUCTION

A major effort has been made in recent years to re-duce the automobile fuel consumption and air

pollution. One of the important components thatinfluence the fuel economy of the automatictransmission car is the torque converter. Sinceseveral million torque converters are built every

*Corresponding author. Tel." 814 865-5551, Fax: 814 865-7092, e-mail: [email protected]

253

254 Y. F. LIU et al.

year, even a small improvement in efficiency wouldresult in significant fuel savings. The objective ofthis research was to improve the torque converterperformance through improved knowledge of theflow field.A typical torque converter consists of a pump,

a turbine, and a stator, and it employs oil asthe working fluid. The pump is connected to theengine crankshaft and converts energy from theautomobile engine to fluid power. In contrast,the turbine extracts this energy from the fluid tothe transmission shaft. The stator acts as a reac-tion member, which is placed between the turbineexit and the pump inlet. Its function is to redirect.the turbine exit flow back into the pump..Some of the earlier investigations on torque

converter are briefly described below. Browarzik(1994) used hot film to examine the unsteady flowfield at the inlet and exit of a pump and turbine ofa Fottinger type torque converter. Marathe et al.(1997) measured the turbine exit flow using a highfrequency response five-hole probe. Later, Dong(1998) measured the pump and turbine exit flowusing an improved miniature five-hole probe. Byand Lakshminarayana (1995) used pressure taps tomeasure the static pressure in the pump andturbine. Brun and Flack (1997a, b) used a laservelocimeter to investigate flow in the pump andturbine blade passages. A review of the investiga-tions carried out by two groups: the University ofVirginia and the Pennsylvania State University,and an assessment of the fluid dynamics of anautomotive torque converter is given by VonBackstrom and Lakshminarayana (1996).None of the prior investigators, however, report

detailed pressure and velocity field in the turbinerotor of a torque converter. The thrust of thispaper is to bridge this gap. The novelty of thisinvestigation is in the use of a rotating five-holeprobe and the measurement of the flow field in therelative system.Compared with the pump and stator flow, the

turbine flow has many complex features. (1) Theturbine passage has a large turning angle (124);and due to this increased flow turning the

curvature effect on the flow is large. (2) Theturbine has a lower rotational speed than thepump, and the effects of the centrifugal andCoriolis force on the flow field in the turbine aresmaller than that in the pump. Furthermore, therotational effects vary with the speed ratio. At thenear-stall condition, the turbine speed is almostzero, and the rotating effects are negligible. (3) Therelative total velocity at the turbine inlet and theexit is relatively high due to large blade turningangle.The pressure and velocity fields at three axial

locations of the turbine (turbine 1/4 chord, mid-chord, and 4/4 chord) are presented and inter-

preted in this paper. The investigation was carriedout in a 245mm torque converter using a

miniature five-hole probe with high frequencyresponse. For the sake of brevity, only the dataat the design condition (SR =0.6) are reported inthis paper. The flow field at the off-designcondition and effects of speed ratio are given inPart II of the paper.

FACILITY AND ROTATING PROBEMEASUREMENT TECHNIQUE

This experimental investigation was accomplishedusing the Torque Converter Research Facility atthe Pennsylvania State University. The detaileddescription of this system can be found in Dong(1998). The pump has 32 blades, and the turbinehas 36 blades. The inlet blade angle of turbine is61.4 and the outlet blade angle is -62.5 Theflow field in the turbine passage was measuredusing a high-frequency response five-hole probe atfive different speed ratios, 0.8, 0.6, 0.4, 0.2 and0.065. The accuracy of this rotating probe is thesame as the probe used in the stationary frame ofreference (Marathe, 1997); +0.15m/s in totalvelocity, -+-1.5 in flow angles, and + 700 Pa inpressure.

In order to simulate all the operating conditionsof the torque converter, tests were conducted atfive speed ratios: 0.8, 0.6, 0.4, 0.2 and 0.065,

TURBINE ROTOR PASSAGE 255

TABLE Test conditions and performance parameters

Speed ratio 0.8 0.6 0.4 0.2 0.065

Pump rpm 1100 1060 1090 1070 800Turbine rpm 880 696 436 214 52Torque Ratio 0.976 1.2590 1.507 1.771 1.922Efficiency % 78.08 75.54 60.28 35.41 12.49

ranging from the peak efficiency condition(SR=0.8) to the near stall condition (SR=0.065). The pump and turbine speeds for thesespeed ratios are listed in Table I. The performancedata, torque ratio, and efficiency, measured by theload cell are also listed in the same table. The oil

temperature inside the torque converter is con-trolled at 60 + 1.Three cross sections are chosen for the study of

turbine passage flow; turbine 1/4 chord, mid-chord, and 4/4 chord. The measurement planes areshown in Figure 1. The turbine 1/4 chord locationis mainly designed to assess the turbine inlet flowlosses, the extent of flow separation, the pumpblade wake decay, and so on. The mid-chordlocation is very important in understanding thedominant flow structure in the turbine passage todetermine the major loss mechanism and flow

3/4 Ch%rd /. -4 Chor//

2/4 Cho

,,4

"11/!11/FIGURE Cross sections of the troque converter used forrotating probe measurements.

transport mechanism. The turbine 3/4 chord is notsuitable for measurement because of the largeblockage effect of the probe. At the turbine 4/4chord, the stem of the probe is located outside theturbine passage, and the probe can travel radicallystraight, but not at the same angle as the turbineblade. Thus the probe blockage effect is largelyreduced. The turbine 4/4 chord flow field data iscritical in determining the size of the core flowseparation, the turbine passage periodic flowdecay, and the flow mixing mechanism down-stream of the turbine blade trailing edge.The rotating probe measurement system de-

signed, developed, and used in this research isdescribed in Dong (1998). It consists of (1) a probetraversing mechanism to measure the flowfield atvarious chordwise and tangential locations, (2) asignal conditioner (amplifiers and battery) mount-ed on the rotating probe casing, (3) electrical con-duits to transmit the data to slip-ring unit, and(4) a slip-ring unit. The probe design and dataacquisition and processing systems in the station-ary system are identical to those described earlier.One of the most difficult problems encountered

was in the design of the probe traverse system. Theprobe must be traversed both tangentially andradially relative to the blade passage. The probe isdesigned to be mounted on the pump shell at somespecified location (percentage of the blade chordlength, as shown in Fig. 1). The radial traverse ofthe probe is accomplished by a sliding sleeve,which has a series of index holes on its outerdiameter. The sleeve can be translated (shell-to-core) in the pump shell and be fixed in incrementalpositions by a set screw, which engages the holeson the sleeve. The guide holes for the probe sleeveare manufactured at the specified locations on thepump shell and are aligned with the local bladeangle. To carry out the tangential traverse, theblade system was designed to be rotated relative tothe pump shell. The contour of shell is manufac-tured to match perfectly with the contour of theblade tip to keep the tip leakage flow negligible.The angular position of blade relative to the probeis determined by a group of index holes, which are


accurately drilled by a CNC (Computer NumericalControl) machine. Once the probe tangentialposition is fixed, the shell and blade parts are heldtogether by eight bolts. The pump cover isredesigned into two pieces; an annular ring and aflat plate cover. The electrical wires used totransfer the signal to the slip-ring unit areconnected through two small holes in the pumpcover from the pump shell to the pump shaft. Alltest parts mentioned above were manufactured byCNC machining, and the pump blade part wasmade by five-axis CNC milling.The error in the rotating probe measurement is

strongly dependent on the noise of the electronicsystem. The Kulite pressure transducers, whichare used in the high-frequency response probe,require a very stable voltage DC power supply.The output signals of these transducers need tobe amplified by low noise amplifiers. Theseamplifiers are powered by a DC power supply.For the rotating frame measurement, the signalhas to be transferred from the rotating frame tothe stationary frame, which is then processedthrough a data acquisition system (DAS-50 and486 PC). The contribution to the electronic noisecomes from the pressure transducer, the amplifier,the power supply, and the data acquisitionsystem.A slip-ring unit was used to transfer the

electrical signal from the rotating frame to thestationary frame. This slip-ring unit is a custom-designed, high quality, brush-type unit. It has 37channels, and the brushes and rings are made ofgold alloy. At the shaft speed range from 0 to1500rpm, the noise level of this slip-ring unit isabout 5 to 7mV for a 3 to 5V level signal. Thisrepresents a signal noise ratio of 0.10% to 0.25%.For the flow measurement, this noise level isacceptable after amplification, where the DCsignal is about 3 to 5 V and the AC signal is about20 to 50 mV. However, this noise is not acceptablebefore amplification, because it is even higher thanthe flow signal and will be amplified by theamplifier gain (250 in this system). Therefore, theamplifier and the transducer power supply have to

be installed in the rotating frame before transmit-ting through the slip-ring unit.A schematic of the electronic system of the

rotating five-hole probe is shown in Figure 2. Twogroups consisting of a rechargeable battery and a

voltage regulator are used as the transducer powersupply. The DC power supplies (+ 12 V, 12 V)for the five amplifiers are transmitted through theslip-ring unit by five parallel channels each. Twovoltage regulators are used in the rotating frame toreduce the noise of these power supplies due to theslip-ring unit. All the above electronic units andfive signal amplifiers are custom-designed minia-ture parts. They are mounted on the pump shell,rotating with the pump passage. The output signalfrom each channel of the five-hole probe istransmitted to the stationary frame through fourparallel channels of the slip ring unit to reduce thenoise due to the slip-ring unit. Seven slip-ringchannels were parallel connected to the commonground. The measured noise of this system is low,total about 5 to 6 mV, which is the same level asthat measured in the stationary five-hole probesystem. Since the probe is in the rotating frame,the transducers are not located at the same radius

Amplifier

Amplifier 2 ." Amplifier 4

Amplifier 5

VoltageRegulator

Signal

Power Voltage [ Battery

Regulator Battery

_R_ot_a_ti_n_g _R_e_fe c_e_F__m_e

PowerSupply

,gUni, Psi;V

DAS-50

486-PC

Stationary Reference Frame

FIGURE 2 Schematic of electrical system for rotating probemeasurements.


as the tip of the probe, and all the measuredpressures were corrected for the centrifugal forceeffect. The accuracy of the centrifugal forcecorrection depends on the spatial error in measur-ing the location of the probe tip and the transducersensors.

DATA CAPTURE AND PROCESSING

Steady Flow Field Data Processing

The velocities and pressures presented in this paperare normalized in the following way

WnormW

Vref ’D np ,/ ntVref’ - V1 (1)

P- PhubPrefPnorm-- Pre------ pV2e (2)

where D is the diameter of the torque converter,which is 245 mm, and Phub is the static pressuremeasured at the stator hub.The absolute stagnation pressure is defined as:

(Po)a P +- pV2o (3)

where (Po)a is the absolute stagnation pressure, Pis the static pressure, and Vo is the absolute totalvelocity.

In a rotating system, the rotary stagnationpressure is defined as:

(Po)r P + pW2o

P* (Po)r -PU2 (4)

where U is the local blade speed, P is the staticpressure, and Wo is the relative total velocity.

For incompressible, inviscid flow, the value ofP* is constant along a streamline. For the viscousflow, the change of this parameter between twopoints on the same streamline represents theviscous loss.

Unsteady Flow Field Data Processing

The unsteady flow field data processing procedureused here is as follows:

(1) Calculate the instantaneous pressure of the fiveholes (P1, P2, P3, P4, Ps) from the instanta-neous voltages using the pressure transducercalibration curves.

(2) Calculate the instantaneous flow parameters((Po)r, P, (Wo)t, Wz, fl) from the pressure usingthe five-hole probe calibration.

(3) Decompose the instantaneous flow parametersinto time mean average component, periodiccomponent, and unsolved component usingthe phase lock averaging technique.

(4) Calculate the RMS (Root Mean Square) valueand the associated unsteadiness quantity.

For the flow parameters in the turbine rotor

passage, the instantaneous value Gi, is decomposedas follows:

+ a + a’ (5)

where is the time-averaged value, is theperiodic value, and G’ is the unresolved flowcomponent. The periodic unsteady componentresults from the relative motion of the rotor withrespect to the stator/rotor, while the unresolvedcomponent is any flow field fluctuation that is notcorrelated with the rotor speed, such as turbulenceand vortex shedding.

Since the data acquisition system is clocked bythe shaft encoder, every instantaneous pressuremeasurement was recorded during each revolu-tion. After all the instantaneous pressure measure-ments were acquired for the particular surveypoint, the instantaneous flow parameters werecalculated from the pressure. Using the phase-lockaveraging technique, the flow parameters werethen ensemble averaged at each survey point ofone revolution, according to"- _1)_n’2" G (6)

///i=1


where G is the ensemble-averaged pressure, and nis the total number of measurement in thatparticular survey point. The unresolved compo-nent for each flow parameter can then becalculated as

G’ Gi- -- (7)

RESULT AND DISCUSSIONS

The frequency spectrums at speed ratio 0.6 at threetypical measurement planes: turbine 1/4 chord,mid-chord and 4/4 chord are shown in Figure 3.Only the spectrums of the pressure at the center

The corresponding ensemble root mean squareof the random fluctuation is calculated as

The level of unresolved unsteadiness in flowparameters is determined by this RMS valuenormalized by the corresponding property.The time-averaged component G is obtained by

averaging all the ensemble-averaged flow param-eters in each survey point as follows"

NIP.

G --Z-j (9)j=l

where NR is the number of points per revolution.The time-averaged value is the time average of allmeasurements at a fixed point in space. Theperiodic component is then obtained from

(--G (10)

The RMS (Root Mean Square) of the periodicfluctuation can be calculated from the periodiccomponent of flow parameters as follows:

The periodic unsteadiness is calculated from theRMS (Root Mean Square) value of the periodicfluctuation normalized by the appropriate prop-erty. For the pressure and velocity quantity, thenormalizing quantities are the total pressure dropthrough the turbine rotor passage and the massaveraged through flow velocity respectively. Forthe flow angle property, the normalizing quantityis the flow turning angle through the turbine rotor.

32*

./., 32(fa-.f,,)-- T-

frequency (KHz)

a. Turbine 1/4 Chord

frequency (KHz)

b. Turbine Mid-Chord

-f;

O01frequency (KHz)

c. Turbine 4/4 Chord

FIGURE 3 Frequency spectrums of rotating five-hole probeat center of turbine passage (SR =0.6).


hole (No. 1) are shown. The pump and the turbineshaft frequencies dominate the turbine passage atmid-chord and 4/4 chord. This is mainly caused bythe varying gravitational forces sensed by theprobe during one revolution. The analysis of thedata indicates that this is not due to vibration.The other dominant frequency is the pump bladepassing frequency (32*(fps-fts)) percieved in theturbine rotating frame. This represents the up-stream effect of pump. The effect of the upstreampump diminishes very rapidly as the flow pro-gresses from the turbine 1/4 chord to the turbinetrailing edge location. The absence of the statorpassing frequency indicates that the influence ofthe downstream stator is negligible.

Static Pressure (P) Field

The static pressure contours at the design condi-tion (SR=0.6) are shown in Figure 4. At theturbine 1/4 chord, the static pressure hasthe highest value near the shell region, and thepressure gradient is mainly in the radial direction.This indicates that the blade loading is small at thislocation, which is consistent with blade staticpressure measurements carried out by By andLakshminarayana (1995). While at the turbinemid-chord, both the radial and the tangentialpressure gradients dominate. At the turbine 4/4chord, again the pressure gradient is in the radialdirection, but the gradient is much smaller thanthe gradient at the turbine 1/4 chord.

In centrifugal turbomachines, both the effects ofrotation and curvature are important. The curva-ture terms include both the radius of curvature dueto the sloping path or curvature of meridionalstreamline and curvature due to swirling androtating flows. The equilibrium for an invisid flowin the meridional plane is described by thefollowing equation:

lOP U2m Vu2 cos (12)p Oz Rz r

where P is the static pressure, Vm is the meridionalvelocity, Rz is the radius of meridional curvature, r

SHELL

CORE


Level P10 0.92109 0.8798

0.83850.7973

6 0.75615 0.71484 0.6736

0.63240.59120.5499

SS

SHELL

CORE


SS

CORE

SHELL

Level P0 0.9560

0,9012

70.84630.7914

6 0.73655 0.68174 0.6268

0.57190.51700.4622

Level P10 0.39639 0,36418 0,33197 0.29976 0.26755 0,23534 0.2031

0.17090,1387

0.1065


FIGURE 4 Contour plots of normalized static pressure P(SR=0.6, Eq. 2).

is the radius of swirling flow, and e is the anglebetween turbine axis and the line perpendicular tothe meridional line.


From this equilibrium equation, e is greaterthan 90 at the turbine 1/4 chord. This indicatesthat the effects of centrifugal force due to swirlingflow and the meridional curvature results in theobserved radial static pressure gradient. At theturbine mid-chord, the contribution of the cen-trifugal force due to the swirling flow to the radialpressure gradient is insignificant, because thedifference in radius at the core and shell is verysmall. Thus, the radial pressure gradient at theturbine mid-chord is not as significant as that atthe turbine 1/4 chord. Finally, at the turbine 4/4chord, the effects of centrifugal force due to

swirling and meridional curvature cancel eachother, and the radial pressure gradient is even

smaller than that at the turbine mid-chord.

Absolute Stagnation Pressure (Po)a Field

The contour plots of the absolute stagnationpressure are shown in Figure 5. At this speedratio, the absolute stagnation pressure contour atthe turbine 1/4 chord location (Fig. 5a) show thatthe pressure has the highest value near the shell,and the pressure gradient is mainly in the radialdirection, from the shell to the core. On the otherhand, at the turbine mid-chord, both radial andtangential gradient dominate. At the turbine 4/4chord, the radial pressure gradient dominates.

Furthermore, from these three figures (Figs.5a-c), it can also be seen that the absolutestagnation pressure drop is evenly distributedfrom the turbine 1/4 chord to the turbine 4/4chord. This is different from the static pressuredrop, because the flow turning plays a veryimportant role in absolute stagnation pressuredrop.The mass-averaged absolute stagnation pressure

drop from the turbine inlet to the turbine exit isshown in Figure 6a. It is clear that the absolutestagnation pressure drop is almost evenly distrib-uted along the turbine flow path. This even

distribution should be expected, as this is thedesign condition. The incidence angle is relativesmall and the pump/turbine matches relatively well

SS

SHELLLevel (Po)=10 2.9546

2.78162.6087

7 2.43576 2.2627

2.08981.9168

3 1.74381.57091.3979

CORE


SHELL

I-II"

Level (Po)a10 1.75579 1.66438 1.5730

1.48176 1.39045 1.29904 1.20773 1.11642 1.0251

0.9337


SS

SHELL Levello o.6oo

!9 0.60338 0.55677 0.5100

0.46330.4167

4 0.37003 0.32332 0.2767

0.2300

CORE


FIGURE 5 Contour plots of normalized absolute stagnationpressure (Po)a (SR 0.6, Eq. 3).

at this operating condition. There is a pressure lossfrom the turbine trailing edge to the exit measure-ment location.


3.5

3Inlet

’ne 114 Chord

.5

ieMid-Chord

Jr

I0.5 _1_ rd__

iTurinelExitl’0 0.25 0.5 0.75

Non-Dimensional Meridlonal Length

3.5

2.5

1.5

(b)

-CORE Turbine Ixlt,-0 0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9

hlH

FIGURE 6 Absolute stagnation pressure distribution(SR=0.6). (a) Mass averaged absolute stagnation pressure(Poa)m drop (SR--0.6); (b) Radial distribution of blade-to-blade averaged absolute stagnation pressure (Poa)b (SR =0.6).

The radial distribution of blade-to-blade massaveraged absolute stagnation pressure from theturbine inlet to the turbine exit (SR --0.6) is shownin Figure 6b. It is clear that the turbine receives amuch higher total pressure at the shell than at the

core. This is caused by the nearly separated flownear the core at the pump exit.

In addition, there is approximately 50% greaterpressure drop near the shell as compared to thecore. Also there is a greater drop associated with thefirst half of the turbine passage duct, particularlynear the shell. This can be attributed to the inletflow field, with high pressure and velocity fluid nearthe shell. It can also be seen that the total pressure inthe second half of the turbine is much more uniformfrom core to shell than is the first half. The possiblecause is the intense mixing caused by the secondaryflow, which tends to provide more uniform pressuredistribution at the exit.

Furthermore, there is some pressure drop fromthe turbine 4/4 chord to the turbine exit, thisrepresents the pressure loss due to the trailing-edgewake decay, dissipation of the secondary kineticenergy as well as turbulence kinetic energy. Allthese factors could collectively result in pressureloss downstream of the turbine trailing edge. Thepressure loss is high near the shell and low near thecore region, this is due to high velocity and highpressure near the shell at the turbine trailing edgelocation and this gives high mixing losses.

Relative Total Velocity (Wo)t Field

The contour plots of the relative total velocity areshown in Figure 7. At the turbine 1/4 chord, therelative total velocity is high near the shell and lownear the core/suction corner. The flow is nearlyuniform in the blade-to-blade direction. A lowvelocity region is observed near the core. This ismainly caused by the non-uniform flow at the exit

of the pump, where low velocity is observed nearthe core (Dong, 1998). At the turbine mid-chord,high velocity fluid is located near the pressure side/shell corner, and low velocity fluid is concentrated

is very low and could be separated. From thispoint of view, the low velocity region near thecore/suction corner at the turbine 1/4 chord mayalso be a separated area. The secondary flowtransports this flow from the core region at the


SHELL

SS

CORE


SHELL

SS

CORE


Level (Wo)10 0.72449 0.64948 0,5744

7 0.49946 0,4244

5 0.34944 0.27443 O.19942 O.1244

0,0494

Level (Wo)10 0,5015

9 0.45138 0.40127 0.35106 0.30095 0,25074 0.20063 O.15042 O.1003

0.0501

ss

SHELL Level (Wo)10 0,70239 0,68398 0.66547 0.64696 0.62855 0,61 O04 0.59163 0,57312 0.5546

0.5362

CORE


FIGURE 7 Contour plots of normalized relative Totalvelocity (W0)t (SR 0.06, Eq. 1).

turbine 1/4 chord to the suction side at the turbinemid-chord. At the turbine 4/4 chord, the velocity isrelatively uniform, the low velocity region ob-served earlier has been eliminated.

It is clear by comparing the data in Figure 4, 5and 7 that the static pressure drop and relativevelocity change from the turbine 1/4 chord to themid-chord is not significant, while the stagnationpressure drop is substantial (approximately 50%drop in normalized total pressure). This shows themajor influence of radius change and the rotationeffect (Coriolis and centrifugal forces). The flowturning effect is augmented substantially by theserotation effects.The radial distributions of blade-to-blade rela-

tive total velocity from the pump exit to theturbine exit are shown in Figure 8. It is clear thatthe change in relative total velocity from the pumpexit to the turbine exit is relatively small. The dropin absolute stagnation pressure is caused predo-minantly by the centrifugal and the Coriolis forcesand not by the acceleration of the relative flow.The flow is highly non-uniform in the radialdirection at the turbine 1/4 chord and turbinemid-chord, which becomes nearly uniform at theturbine 4/4 chord and the turbine exit. This may beattributed to possible flow separation near thecore/suction corner at the turbine 1/4 chord andthe turbine mid-chord.The most noticeable feature is that the turbine

inlet velocity is significantly higher than those atthe other four planes. It is important to recall thatthe turbine inlet measurements were taken in thegap between the pump and turbine and not at theleading edge of the turbine. The higher relativetotal velocity can be explained by consideringvelocity triangles. The total velocity still contains a

large tangential component created by the pump,which has a higher rotational speed. Thus, there ishigh angular momentum at the turbine inlet. Atthe turbine exit, the turbine also imparts a highrelative tangential velocity. Thus, the relative totalvelocity would be high at the exit as well.However, the turbine rotates much more slowlythan the pump, and the velocities are not as highas the inlet velocities.


0,9

0.8

0.7

0,6

0.4

0,2

0.1

0,9

0,7

0,6

0,2 /

o6 o,5 o’. ’o,5’(W=)

a. Pump Exit b. Turbine 1/4 Chord

0.9

0.8

0.7

0,6

o,s

0.4

0.2

0.1

0.6

0,2

0.1

%’" o," ’o’.’ "o.

c. Turbine Mid-Chord

0.9

0,8

0,7

0,6

0.5

0,4

0,2

0,]

o ’o,5’ ’o’ o.5

(W=),

d. Turbine 4/4 Chord

e. Turbine Exit

FIGURE 8 Radial distribution of blade-to-blade averagedrelative total velocity (Wot)b (SR =0.6, Eq. 1).

SHELL

CORE


SHELL

ii

CORE


SHELL

SS

Level W10 0.54239 0.48798 0.43367 0.37936 0.32495 0.27064 0.21623 0.16192 O.1076

0.0532

1098765432

Level W0.49230.44300.39380.34460.29540.24610.19690.14770.09850.0492

Level W10 0.42129 0.41098 0.40077 0.39056 0.38025 0.37004 0.35973 0.34952 0.3393

0.3290

CORE


FIGURE 9 Contour plots of though flow velocity (Wz)(SR--0.6, Eq. 1).

Through Flow Velocity (Wz) Field

The contour plots of through flow velocity areshown in Figure 9. The through flow velocity is thevelocity component perpendicular to the measure-ment plane.

The overall trend in the through flow velocitydistribution is very similar to that of the totalvelocity field (Fig. 7). A "jet-wake" flow structureis observed at the turbine 1/4 chord and theturbine mid-chord. At the turbine 1/4 chord, massflow is concentrated near the shell region, and a


possible flow separation area is observed near thecore/suction side corner. At the turbine mid-chord,a different "jet-wake" flow pattern exists. Highvelocity fluid is observed near the pressure side/shell corner, and low velocity fluid (wake) region islocated near the middle of suction side. However,at the turbine 4/4 chord (turbine trailing edgelocation), the through flow is very uniform and noflow separation area exists.One reason for the low velocity region at the

turbine 1/4 chord (Figs. 7 and 9) is the pump exitflow and the high incidence angle at the turbineleading edge. Due to the "jet-wake" flow structureinside the pump passage, the velocity is high nearthe shell and very low near the core region at thepump exit. In addition, the incidence angle at theturbine leading edge is about 15 This probablycauses flow separation at the core/suction sidecorner at the turbine 1/4 chord. As the flowprogresses from the turbine 1/4 chord to theturbine mid-chord, the meridional curvature andcamber curvature effect begins to dominate. Thesecondary flow tends to redistribute the flow field.This may cause the low velocity region to movefrom the core to the suction side as the flowprogresses. As the flow continues from the turbinemid-chord to the turbine trailing edge, themeridional curvature effect continues to influencethe flow field. The flow near the core regionaccelerates due to the convex curvature, and theflow near the shell region decelerates. As a result ofthis, flow reattaches before the turbine trailingedge near the core. This is also augmented by thesecondary flow.

Relative Flow Yaw Angle (/t) Field

The contour plots of the yaw angle measured bythe rotating five-hole probe are shown in Figure 10.At the turbine 1/4 chord, the blade angle is 48.12,and at the turbine mid-chord it is 0. The largeoverturning at the turbine 1/4 chord can beattributed to large positive incidence and mis-match between the pump exit flow and turbineinlet flow. The core flow region has very low

(:)t

a. Turbine Inlet

33.5662

b. Turbine 1/4 Chord


(J)tgo

E


FIGURE 10 Contour plots of relative flow Yaw Angle (/30(SR- 0.6).

through flow and total relative velocities andoperates at substantial off-design condition withflow angle ranging from 12 to 20 This off-designcondition is located near the suction side at theturbine mid-chord. Most of the core flow has


relative flow angle from --2 to 8 The flow isunderturned in most regions, even though thepressure drop is substantial (Fig. 5); confirmingthe earlier conclusion regarding the dominanteffect of centrifugal and Coriolis forces.The blade angle at the turbine 4/4 chord is

-58.47, and the relative flow yaw angle contourplot at this location shows that the flow is wellaligned with the turbine blade. The maximumdeviation angle is about 11 near the core/suctioncorner and the middle of suction side. The presenceof large deviation angle at these two areas can beattributed to the residual effect of possible flowseparation, which plays a very significant role atthe turbine 1/4 chord and turbine mid-chord..

Rotary Stagnation Pressure (P*) Field

The contour plots of rotary stagnation pressureare shown in Figure 11. The difference in rotarystagnation pressure along flow path representslosses. The contour plots shown in Figure 11indicate that these values are low near the core andhigh near the shell. This is partially attributed topump exit flow and not the losses within theturbine passage. On the other hand, the differencein these values between 1/4 chord and mid-chordrepresents flow losses inside the turbine passage.The flow losses are highest near the blade suctionside and quite significant near the shell region.The rotary stagnation pressure decreases near

the shell region and increases near the core regionfrom the turbine mid-chord to the turbine 4/4chord. The increase in rotary stagnation pressurenear the core region is due to the flow reattach-ment before the turbine trailing edge. Because ofthe flow reattachment, the high mixing losses nearthe separation region are substantially reduced.

Overall, it can be concluded from the rotarystagnation pressure contour plots that there arehigher losses at the first half of the turbine passagethan the second half, and possible flow separationis the major sources of loss.The radial distributions of blade-to-blade aver-

aged relative total pressure (Por)b from the pump

1,0

a. Turbine Inlet

" /"1

b. Turbine 1/4 Chord

2


I


FIGURE 11 Contour plots of rotary stagnation pressure (P*)(SR=0.6, Eq. 4).

exit to the turbine exit are shown in Figure 12. It isclear that the relative total pressure decreases asthe flow progresses from inlet to exit. Examiningthe radial distribution of relative total velocity


-" Tucbine 1/4 ChordTufoine Mid-Chord

< Turbine 4/4 Chord1.2 _-- ,

0.8

0.6

0.4

0.20 0.25 0.5 0.75

FIGURE 12 Radial distribution of blade-to-blade averagedrelative total pressure (Por)b (SR’-0.6).

(Fig. 8), it is concluded that the drop in relativestagnation pressure is mainly caused by the drop instatic pressure, not by the change in the relatiyetotal velocity.

In addition, the relative stagnation pressure islow near the core region at all measurementplanes. This feature may be attributed to severalfactors. First, this low stagnation pressure may becaused by the "jet-wake" flow structure at thepump exit, which causes low pressure rise nearthe core region inside the pump passage. Second,the low stagnation pressure near the core isattributed to the low mass flow near the coreregion, which is typical in mixed flow turbo-machines. Finally, there exists flow separation nearthe core region. This causes higher losses, resultingin low stagnation pressure near the core region.

Blade-to-blade Averaged Rotary StagnationPressure (e*)b

The radial distribution of blade-to-blade averagedrotary stagnation pressure (P*)b for speed ratio 0.6is shown in Figure 13. This quantity represents the

0.7

0.6

0.5

0.4

0.3

0.2

0.1

0

-0.10

!*- Turbine 1/4 Chord, Turbine Mid-Chord:.< Turbine 4/4 Chord!’=__, ;riiNiEit

-iIlillllllllllllllllllll,,l i,,,,i,,,

0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9

h/H

FIGURE 13 Blade-to-blade averaged rotary stagnation pres-sure (P*)b (SR =0.6).

difference between local stagnation pressure andthe stagnation pressure calculated from the in-viscid Euler equation. This quantity should remainnearly constant in an inviscid flow, and thedifference in this quantity between any twostations represents the hydrodynamic loss. Sincea decrease in this quantity represents a pressureloss, it is clear that there are greater lossesoccurring in the first half of the turbine passage.This is particularly true in the end-wall regions.The flow separation is most likely the source ofthese losses. Though the drop in rotary stagnationpressure is much smaller in the second half of thepassage duct, it is still greatest in the mid-span andshell regions. However, because the flow reattachesbefore the turbine trailing edge, and because thesecond half of the passage duct does not experi-ence the flow disturbances associated with theinlet, the losses are smaller in the second half of theturbine.

Blade-to-blade Periodic Unsteadiness in Relative

Pressure, Velocity, and Flow Angle

The radial distributions of blade-to-blade aver-aged periodic unsteadiness at the turbine 1/4


chord, 2/4 chord, and turbine 4/4 chord are shownin Figures 14, 15 and 16 respectively. The periodicunsteadiness in Po, P, Wo, near the core region (hiH=0-30%) at the turbine 1/4 chord is quite high(7-10%) and low near the shell region (about2.5%). This can be attributed to the flow separa-tion near the core/suction side corner at theturbine 1/4 chord.. At this area, the fluctuationsin all the flow parameters (pressure, velocity, andflow angles) are quite high. The unsteady level inthe free stream region (h/H= 30-100%) is rela-tively low and uniform. There exists a regionshowing high fluctuations at the turbine 2/4 chord,which is located near the mid-span (h/H=55-80%). The unsteady level is about 2-5 times ofthose observed in the free stream region. This canalso be explained by the possible flow separationand the associated high fluctuations near thesuction side at the turbine 2/4 chord. On the otherhand, the radial distribution of the periodicunsteadiness at the turbine 4/4 chord is relativelyuniform from the core (2.5%) to the shell (about1.5%).

00 0.1CORE

0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9h/H SHELL

FIGURE 14 Radial distribution of blade-to-blade averagedperiodic unsteadiness in flow field (SR=0.6, Turbine 1/4Chord).

O0 0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9CORE h/H SHELL

FIGURE 15 Radial distribution of blade-to-blade averagedperiodic unsteadiness in flow field (SR=0.6, Turbine 2/4Chord).

The higher unsteadiness at the turbine 1/4 chordand turbine 2/4 chord is due to the followingreasons. First, the measurement planes at theturbine 1/4 chord and turbine 2/4 chord are muchcloser to the pump trailing edge than the turbine

3

2.75

2.5

o 2.25

1.75

1.5

1.25.2

0.75

0.5

0.25

ooCORE

0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9h/H SHELL

FIGURE 16 Radial distribution of blade-to-blade averagedperiodic unsteadiness in flow field (SR-0.6, Turbine 4/4Chord).


4/4 chord. Hence, the pump blade wake has notdecayed completely at these two planes, and hassignificant effect on the unsteady flow field.Secondly, there exist large pressure gradients asshown in Figure 4. The pressure gradient is fromthe shell to the core and from the pressure side tothe suction side for the turbine 1/4 chord and 2/4chord respectively. These pressure gradients drivethe secondary flow from the high pressure regionto the low pressure area, this results in theaccumulation of low momentum fluid at thesuction/core corner at the turbine 1/4 chord andnear the suction side at the turbine 2/4 chord. Atthe turbine 4/4 chord, however, the static pressureis very uniform. There is no clear pressure gradientand no low momentum fluid region. This canexplain why the unsteadiness near the turbinetrailing edge is low and uniform.The basic feature of the periodic unsteadiness in

relative pressure, velocity, and flow angle, is quitesimilar at these three measurement planes (Figs.14, 15 and 16). That is, the high periodicunsteadiness region is limited to the core/suctioncorner at the turbine !/4 chord, and the suctionside region at the turbine 2/4 chord. On the otherhand, the unsteady level is relatively low anduniform at the turbine 4/4 chord. This agrees wellwith the through flow velocity distribution shownin Figure 9. The magnitude of the unsteady level inthe free stream regions is only about 20-30% ofthat observed in the flow separation region. Thisindicates that the low pressure, low velocity regionmay be separated, and these may be the regionsthat contribute to the high losses and inefficiency.

CONCLUSION

From the data presented at the design condition(SR=0.6), the following conclusions can bedrawn:

(1) Upstream pump exit flow field has a dominanteffect on the flow field in the entrance region ofthe turbine passage; however, these effects

(2)

()

(4)

(5)

(6)

(7)

(8)

diminish very rapidly as the flow progressesfrom the turbine 1/4 chord to the turbinetrailing edge.The spectrum analysis reveals the presence offlow unsteadiness at the pump blade passingfrequency near the turbine 1/4 chord, withinsignificant trace at the turbine mid-chord.The spectrum at the turbine exit shows neitherthe pump nor the downstream stator influenceon the turbine exit flow field.The static pressure gradient is mainly in theradial direction at the turbine 1/4 and 4/4chord. While at the turbine mid-chord, boththe radial and the tangential pressure gradientdominant. This static pressure drop is causedby both the flow turning and the effects ofrotation and curvature.The mass averaged stagnation pressure drop isalmost evenly distributed along the turbineflow path. In addition, there is approximately50% greater pressure drop near the shell ascompared to the core. This can be attributed tothe inlet flow field which has high pressure andvelocity near the shell.Examining the static pressure, relative velocity,and absolute stagnation pressure field, it isconcluded that the drop in absolute stagnationpressure is caused predominantly by thecentrifugal and the Coriolis forces and not bythe acceleration of the relative flow. The flowturning effect is augmented substantially bythe rotation effects.The rotary stagnation pressure distributionindicates that there are a lot of losses occurringin the first half of the turbine passage. The flowseparation and the mismatch between thepump exit and turbine inlet flow field are themost likely sources of these high losses.The main feature of the unsteady flow in theshell region is that the flow fluctuation isrelatively low and uniform. The unsteady levelsare much lower than that observed in the cornerflow separation, or flow mixing areas.The core region is the most inefficient regiondue to possible flow separation. Pump/turbine


mismatch is another source of inefficiency ofturbine. The efficiency and fuel economy canbe improved through a redesign of the coreregion and by improved matching of the pumpexit flow and the turbine inlet flow field.

Acknowledgements

This project was sponsored by the PowertrainDivision of the General Motors Corporation. Theauthors wish to express their gratitude to Mr.Donald G. Maddock of GM for his assistance,comments, and encouragement. Assistance byY. Dong is gratefully acknowledged.

NOMENCLATURE

AD

Hh/H

npntPfhubenormPref(Po)a(Poa)b

(Poa)m

(Po)rp*

P.SSRS.S

areadiameter of torque converterpump shaft frequency (np/60)turbine shaft frequency (nt/60)flow parameterprobe radial height measured from core,.mmblade height, mmrelative radial position (=0 at core,at shell)pump rotating speedturbine rotating speedstatic pressurestatic Pressure at hubnormalized Pressure (Eq. (2))reference pressure (Eq. (2))absolute stagnation pressure (Eq. (3))blade-to-blade averaged absolute stagna-tion pressuremass averaged absolute stagnationpressurerelative total pressure (Eq. (4))rotary stagnation pressure (Eq. (4))pressure sidespeed ratio, nt/npsuction side

U

VrefWnorm(Wo)(Wot)b

Wz()P

blade speedabsolute total velocityreference velocity (Eq. (1))normalized velocity (Eq. (1))relative total velocityblade-to-blade averaged relativevelocitythrough flow velocityrelative flow angleoil density

total

Subscript

b, muns

blade-to-blade averaged, mass averagedperiodic unsteadiness

Superscript

Ensemble AveragedTime AveragedPeriodicUnresolved

ReferencesBrowarzik, V. (1994) "Experimental Investigation of Rotor/Rotor Interaction in a Hydrodynamic Torque ConverterUsing Hot-Film Anemometry", ASME Paper No. 94-GT-246, presented at the International Gas Turbine andAeroengine Congress and Exposition, The Hague, Nether-lands, June 13 16.

Brun, K. and Flack, R. D. (1997a) "Laser VelocimeterMeasurement in the Turbine of an Automotive TorqueConverter: Part I-Average Measurements", ASME Journalof Turbornachinery, 119, 646-654.

Brun, K. and Flack, R. D. (1997b) "Laser Ve!ocimeterMeasurement in the Turbine of an Automotive TorqueConverter: Part II-Unsteady Measurements", ASME Jour-nal of Turbomachinery, 119, 655-662.

By, R. R. and Lakshminarayana, B. (1995) "Measurement andAnalysis of Static Pressure Field in a Torque ConverterTurbine", ASMEJournal ofFluids Engineering, 117, 473 -478.

Dong, Y. (1998) "An Experimental Investigation on FluidDynamics of an Automotive Torque Converter", Ph.DThesis, Department of Aerospace Engineering, The Pennsyl-vania State University.

Marathe, B. V., Lakshminarayana, B. and Maddock, D. G.(1997) "Experimental Investigation of Steady and UnsteadyFlow Field Downstream of Automotive Torque ConverterTurbine and Inside the Stator, Part !: Flow at the Exit ofTurbine, Part 2: Unsteady Pressure on the Stator BladeSurface", ASMEJournalofTurbomachinery, 119(3), 624-645.

Von Backstrom, T. W. and Lakshminarayana, 13. (1996)"Perspective: Fluid Dynamics and Performance of Auto-motive Torque Converters: An Assessment", ASME Journalof Fluids Engineering, 118, 665-676.

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Rotor Passage Automotive Torque Converter High Frequency …downloads.hindawi.com/journals/ijrm/2001/424583.pdf · 2019-08-01 · Measurement Part I: FlowField at the Design Condition