THE EFFECT OF FORWARD SWEEPIN A TRANSONIC COMPRESSOR ROTOR

Harald PASSRUCKER Martin ENGBERMTU Aero Engines, Dachauer Straße 665, 80995 München / Germany

Harald.Passrucker@muc.mtu.de Martin.Engber@muc.mtu.de

Stephan KABLITZ Dietmar K. HENNECKEDarmstadt University of Technology, Gas Turbines and Flight Propulsion,

Darmstadt / Germanykablitz@gfa.tu-darmstadt.de hennecke@gfa.tu-darmstadt.de

ABSTRACTThis paper presents design and testing of a transonic compressor rotor with

forward sweep. The rotor was used to investigate the influence of forward sweep onperformance and stability of a single stage transonic compressor compared with abaseline design with radially stacked blade sections. The comparison was donenumerically with the 3D Navier-Stokes code TRACE_S and experimentally in theDarmstadt Transonic Compressor test rig. It was found that the new rotor with forwardsweep has an increased efficiency and also a much better stall margin (much more in therig test than predicted by the 3D Navier-Stokes calculation). Particularly close to stallthe forward sweep diverts the flow towards the blade tip region which helps to stabilizethis region. For that reason it is possible to throttle the forward swept rotor muchfurther as the radially stacked rotor although the forward-swept rotor does alreadysuffer from separated flow in the hub.

INTRODUCTIONA major part of the losses in transonic compressor rotors is created near the blade tip.

Shock losses and the interaction of the shock with other flow phenomena, like tip clearanceflow or boundary layers, also contribute to these losses. The tendency of the shock to causeboundary layer separation can account for an amount of loss which is significantly higher thanthe actual shock loss. Therefore, sweep has been considered as a method to reduce shockstrength and to improve efficiency and surge margin. Denton et al. [1] analyzed by CFD thatthe effect of sweep and lean on transonic fan efficiency and pressure ratio is remarkablysmall, but have a significant influence on the stall point of the fan. Ulrich [2] numericallyinvestigated the influence of sweep and lean on a transonic rotor with the same result. Thereare mainly 3 physical effects how sweep does influence the flow in a blade row.Basic effects of swept blades on compressor flow

a) Influence on the blade loading (figure 1): The pressure gradient perpendicular to aplane end wall must be zero, since there can be no acceleration perpendicular to the wall. Inthe case of figure 1 (aft sweep), the blade loading near the lower wall must be reduced nearthe leading edge where the loading rapidly falls to zero (no blade) as one movesperpendicularly away from the wall. Conversely the loading on the lower wall will tend to beincreased near the trailing edge since there can be little pressure difference between it and themore highly loaded region above it. The opposite effect occurs near the upper wall. Generallythe loading in the tip region is reduced in the front area with the forward-sweep which results

in a leading edge which is more tolreant to changes in incidence. Furthermore the tip leakageis reduced in this area (lower loading).

b) Influence on the shock position (figure 2): In the spanwise direction, the shock cannotintersect the outer casing obliquely. It must either turn normal to the casing or possiblybifurcate in a shock/boundary-layer intersection. This requirement on the spanwise shockshape near the casing is an inviscid phenomenon. In the absence of an endwall, the shockshapes for the forward- and aft-swept rotors would be bent forward or beackward in similarfashion. In the presence of the endwall, however, the shock must turn normal to the casing,moving upstream for an aft-swept rotor and downstream for the forward-swept rotor.Generally, a shock position which is further downstream in the tip region, leads to a betterstall margin because the rotor can be throttled further until the bow shock detaches from theleading edge.

c) Influence on the accumulation of low momentum fluid near the endwall (figure 3): Ina conventional rotor, fluid particles inside the blade boundary layer move radially outwarddue to the imbalance between the centrifugal forces and the pressure gradient. Theaccumulation of low momentum fluid near the endwall is considered to be a major cause ofincreased aerodynamic loss and reduced operation range. In the case of a forward-swept rotor,two mechanisms lessen the accumulation of low momentum fluid near the endwall. First, theradially migrating boundary layer flow cannot reach the endwall region due to the forwardsweep of the blade. Second, the region of high pressure on the suction surface after the peakin the pressure distribution is located further upstream at the tip region than at the hub region.Therefore, radial migration of the low momentum fluid is suppressed and accumulation oflow momentum fluid near the tip is reduced.

Figure 1: Effect of sweep on blade loading(Denton [3])

Figure 2: Endwall effect on shock structurenear the casing (Hah et al. [4])

Figure 3: Secondary flow in a forward swept rotor (Yamaguchi et al. [5])History of different swept Rotors tested at TU-Darmstadt

The Darmstadt Transonic Compressor test rig was brought into operation in close co-operation with MTU Aero Engines Munich in 1993. A series of 3 rotors was used toinvestigate the influence of sweep and lean on performance and stability of a single stagetransonic compressor. The baseline Rotor No.1 (Rotor 1) with radially stacked blade sections

was designed by Schulze et al. [6]. To investigate the influence of blade sweep, especially inthe tip region, a Rotor No.2 (Rotor 2) was designed with considerable aft-sweep. It wasinvestigated in 2000 (Blaha et al. [7]). Rotor No.3 (Rotor 3) features forward sweep and wastested in 2001.

TEST FACILITY AND MEASUREMENT EQUIPMENTThis section briefly describes the test rig (figure 4, 5) and conventional instrumentation

used to determine speedlines and efficiency.

Figure 4: Sketch of the test facility Figure 5: Cross section of the test compressorInlet total pressure and temperature are taken in the settling chamber in front of a bellmouth.At the inlet, wall static pressure is measured to determine the mass flow by using a calibratednozzle. Pressure losses in the inlet duct are taken into account by an experimentallydetermined loss coefficient. The downstream flow conditions are taken from fixed totalpressure and total temperature probe rakes located on the bearing support struts behind thestator (figure 6), while the stator is traversed circumferentially. Shaft speed, power and torqueare measured by a Torquemeter measuring device between the 800kW DC-drive and thecompressor.

Figure 6: Total pressure and total temperature probe rakes

Measurements of speedlines were performed by recently upgraded Pitot type totalpressure and total temperature rakes mounted on the five struts downstream of the stator.Complementing the eleven pitot type total pressure probes two static pressure taps are locatedat the same axial position to gather static pressure information at hub and casing. Bytraversing the stator upstream of the rakes in increments of 5% stator pitch, the stator exit

plane is resolved with 20 positions pitchwise and 13 probe locations from hub to tip, yielding260 single pressure values. For determination of the total pressure ratio the data is at firstaveraged circumferentially, using an arithmetic average (since measuring the wholetemperature distribution required for the massflow weighed average would take way too muchtime). For averaging in the radial direction, the pressures are weighted according to localmassflow using the measured radial distribution of total temperature between two statorwakes. Isentropic efficiency is calculated by comparing compressor work input to the flowtaken from pressure measurements with work input at the shaft which is measured with theTorquemeter device of the test facility. Total temperature measurements at stator exit givegood general information about radial distributions of efficiency but quantitatively preciseaveraged results are too difficult to obtain due to the rather long duration of temperaturemeasurements and become even less reliable at part speed conditions. The data aquisitiontakes 4 minutes for each operating point. For a 95% confidence level (U95) at 100% speedand peak efficiency operating point this yields:• mass flow rate +/- 1.1%• pressure ratio +/- 0.5 %• isentropic efficiency +/- 1.4 %

GEOMETRIC DESIGN OF ROTOR 3Figure 7 shows both rotors. Rotor 1 was designed conventionally with blade profiles

stacked radially along their centers of gravity. Rotor 3’s design features are lower bladenumber, pronounced forward-sweep and higher blade chord length in the tip region to reach abetter stability and efficiency.

Rotor 1 (radially stacked) Rotor 3 (forward-swept)

Figure 7

The blade number of Rotor 3 was reduced compared to Rotor 1 from 16 blades to 14blades. For this reason the solidity of Rotor 3 is generally lower than of Rotor 1 (figure 9).The blade chord length of Rotor 3 was increased in the tip region (figure 9) which increasesthe solidity there and improves stability and efficiency. The inlet and exit metal angles (figure10) are more or less the same for both rotors to deliver the same work. Figure 11 shows thestagger line of Rotor 3 profiles (center of gravity) in axial direction. The light-grey curvesillustrate the displacement to introduce the forward sweep. A small lean (grey curve) wasrequired in the tip region to balance the mechanical stress in this area. The black curve is the

actual stacking line, the sum of both displacements. The sweep displacement is counted indirection of the blade chord, the lean displacement perpendicular to the chord. The sweepstacking line with the backward displacement in the middle was designed in a manner tolower the stress at the leading edge. This is important for FOD (foreign object damage) cases.

0.0 0.2 0.4 0.6 0.8 1.0

relativ height X/H [-]

chor

d le

ngth

L [m

]

ROTOR 1

ROTOR 3

0.005

0.0 0.2 0.4 0.6 0.8 1.0

relativ height X/H [-]

L/T

[-]

ROTOR 1

ROTOR 3

0.1

Figure 8: Chord length Figure 9: Solidity

0.0 0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9 1.0

relativ height X/H [-]

ROTOR 1

ROTOR 3

exit angle

inlet angle

blade chord angle5

Figure 10: Blade metal angles

0.100

0.110

0.120

0.130

0.140

0.150

0.160

0.170

0.180

0.190

0.200

axial displacement

rad

ius

[m]

sweepleanboth

Figure 11: Profiles stagger line Rotor 3

COMPUTATIONAL GRID AND BOUNDARY CONDITIONSThe simulation was performed with the steady 3D Navier-Stokes-Solver TRACE_S

(details are found in Fritsch et al. [8]) on a relative fine H/O-grid (figure 12) with a H-grid inthe rotor tip gap (0.9mm) (Rotor H(105×33×65), O(129×9×65), Hgap(129×5×9); StatorH(105×33×65), O(129×9×65)) for both rotors. The k-ε model was uses for turbulencemodelling with wall function. Design speed is 20,000 rpm. At the inlet total temperature, totalpressure and flow angles are forced. The strong gradient of the total pressure at the casingboundary layer was accounted for by an appropriate boundary condition for that region.Information between rotor and stator domain is transferred with a mixing plane interface. Atthe stator outlet the average pressure with radial equilibrium was set.

Figure 12: Computational grid (Rotor 3)

3D-ANALYSIS AND TEST RESULTSThere is good overall agreement between calculation and measurement (see figure 13,

15). Surge margin of Rotor 3 is fortunately much better in the measurement. The reason isthat the numerical stability is determined by the stator tip region of where separation occurs.In the experiment this is no trigger for rotating stall. Therefore in reality there is separation inthe stator but no rotating stall at this operation point and the compressor can be throtteledfurther. This also shown in the compressor maps (figure 13, 14). The speed line of the rotoralone is still rising and delivers more pressure rise, whereas the stage’s speed line turnshorizontally with no further pressure rise.

With Rotor 3 it was possible to increase the peak efficiency (Navier-Stokes 1.5%,measurement 1.5%) with the above described features (reduced blade number, increasedblade chord length in tip region, forward sweep). The blade number reduction reduces theblockage and therefore a displacement of the choke margin to higher massflow with a wideroperating range. Furthermore there is a much higher stability at the throttled condition.Unfortunately it is not possible to allocate the efficiency profit and the higher stability to eachdesign change, but numerical experiments show that the efficiency profit can be accounted toequal parts to increased blade chord length, forward sweep and slight profile adaptions. Thehigher stability is attributed mainly to the forward sweep.

Figure 16 shows the measured compressor map for both rotors. Rotor 3 was measuredfrom 100% down to 30% speed, Rotor 1 only from 100% to 80% speed. Rotor 3 has nearly aconstant peak efficiency from 80% to 100% speed, which covers the most operating points inan engine application. Rotor 1 has its peak efficiency at 80% speed and is about 0.4% betterthan Rotor 3. The difference to 100% speed in efficiency is 2.1%. The stability at all speedlines of Rotor 1 is significantly lower than Rotor 3. The last stable point at 90 % speed ofRotor 3 has been taken in the experiment, although there is a strong “negative” pressure rise.Rotor 3 obviously suffers from strong separation in the hub region and produces a lot oflosses. But in the tip region the rotor is still stable and rotating stall can not yet be detected.

Figure 13: Stage compressor map

ROTOR (Navier-Stokes)

1.30

1.35

1.40

1.45

1.50

1.55

1.60

1.65

1.70

1.75

1.80

1.85

14 14.5 15 15.5 16 16.5 17

massflow [kg/sec]

tota

l pre

ssu

re r

atio

[-]

isen

tro

p e

ffic

ien

cy[%

]

ROTOR 1

ROTOR 3

100%

∆η=2.5

Figure 14: Rotor compressor map

STAGE (measurement)

1.30

1.35

1.40

1.45

1.50

1.55

1.60

1.65

1.70

1.75

1.80

1.85

12.5 13 13.5 14 14.5 15 15.5 16 16.5 17

massflow [kg/sec]

tota

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re r

atio

[-]

isen

tro

p e

ffic

ien

cy[%

]

ROTOR 1

ROTOR 3

100%

∆η=2.5

Figure 15: Stage compressor map

STAGE (measurement)

1.00

1.10

1.20

1.30

1.40

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2.00

2 7 12 17

massflow [kg/sec]

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atio

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cy[%

]

ROTOR 1

ROTOR 3

100%

90%

80%

50%

40%

65%

30%

∆η=5

Figure 16: Stage compressor map

STAGE (Navier-Stokes)

1.30

1.35

1.40

1.45

1.50

1.55

1.60

1.65

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14 14.5 15 15.5 16 16.5 17

massflow [kg/sec]

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cy[%

]

ROTOR 1

ROTOR 3

100%

∆η=2.5

Figure 17 illustrates radial distributions of stage total pressure ratio and isentropicefficiency near choke (m=16.7), peak efficiency (m=16.4), near stall (m=12.6) and one pointinbetween (m=15.2). Remarkable is the near stall radial distribution, where the rotor stilldeliverrs high pressure rise in the tip region which indicates high stability. In the mid there isa little penetration in pressure rise. The other curves seems to be consistent with a continuouspressure rise at different throttle conditions. The highest efficiency is as expected in the midregion. Towards the tip there is a strong decrease of efficiency and pressure rise as a result ofthe high mach number with the resulting shock losses. Radial distributions of stage totalpressure ratio and isentropic efficiency of Rotor 1 are displayed in figure 18 near choke(m=16.4), peak efficiency (m=16.2) and near stall (m=15.0). Here the character of the radialpressure rise distribution is not changing at throttled conditions.

0.0 0.2 0.4 0.6 0.8 1.0relative height X/H [-]

tota

l pre

ssu

re r

atio

[-]

m=16.7, PIT=1.31m=16.4, PIT=1.48m=15.2, PIT=1.54m=12.6, PIT=1.55

100%

0.1

0.0 0.2 0.4 0.6 0.8 1.0

relative height X/H [-]

isen

tro

pic

eff

icie

ncy

[%

]

m=16.7, PIT=1.31m=16.4, PIT=1.48m=15.2, PIT=1.54m=12.6, PIT=1.55

100%

10

Figure 17: Rotor 3 (measurement)

0.0 0.2 0.4 0.6 0.8 1.0relative height X/H [-]

tota

l pre

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re r

atio

[-]

m=16.4, PIT=1.37m=16.2, PIT=1.45m=15.0, PIT=1.53

100%

0.1

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relative height X/H [-]

tota

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re r

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[-]

m=16.4, PIT=1.37m=16.2, PIT=1.45

m=15.0, PIT=1.53

100%

0.1

Figure 18: Rotor 1 (measurement)

Figure 19 clearly shows the difference in leading edge and trailing edge contour of Rotor3 with forward sweep and the baseline Rotor 1. The streamlines of Rotor 3 are generallydiverted towards the tip region as a result of the forward sweep compared to Rotor 1. Thisfeature is more prominent towards the stall condition.

peak efficiency near stall

- - - - Rotor 1 —— Rotor 3Figure 19: Meridian stream lines (Navier-Stokes)

Figure 20 illustrates the isentropic mach number distribution of both rotors’ bladesections at peak efficiency operation conditions. The biggest difference can be detected in thetip region, where the most losses are located. The loading near the leadind edge of Rotor 3 isreduced which causes less tip leakage. The mach number upstream of the shock is also lowerwhich results in lower shock losses. The shock position is much further downstream on theprofile which is improving stability. In the hub and mid sections the loading of Rotor 3 isincreased as a result of the reduced blade number. In the mid section the shock system is splitin two which also helps to reduce losses at a given pressure rise.

hub mid tip

- - - - Rotor 1 —— Rotor 3Figure 20: Isentropic mach number _ peak efficiency (Navier-Stokes)

CONCLUSIONSThe present study indicates that the numerical prediction of global values like massflow,

pressure rise and efficiency with TRACE_S is very close to the experiment and allows thedesigner to optimize the blades with a 3D Navier-Stokes solver. The calculation as well as themeasurement show an increased efficiency and also a much increased stall margin (muchmore in the rig test than predicted by the 3D Navier-Stokes calculation) for the forward sweptrotor compared to a radially stacked rotor. The increased efficiency results not only from theforward sweep but also from the increased chord length in the tip region with simultaneousreduction of the blade number. Particularly close to stall the forward sweep diverts the flowinto the tip region which improves stability in this region. Even if separation occurs in the hubregion, the forward swept rotor can be operated in a stable conidtition without developingrotating stall.

ACKNOWLEDGEMENTSThe work presented here was supported by the German Ministry of Economics Affairs.

We are also grateful to the management of MTU Aero Engines for the permission to publishthe result.

REFERENCES[1] Denton, J.D., Xu l. (2002), The Effects of Lean and Sweep on Transonic Fan

Performance, ASME Paper 2002-GT-30327, Amsterdam - The Netherlands

[2] Ulrich, M. (1999), Einfluß von 3D-Gestaltungselementen bei der Beschaufelungs-auslegung auf Wirkungsgrad und Stabilitätsgrenze einer Hochdruckverdichterstufe,Diplomarbeit, MTU Aero Engines GmbH – Fachhochschule Konstanz

[3] Denton, J.D. (1999), The Exploitation of 3D Flow in Turbomachinery Design, VKILecture Series 1999-02 – Turbomachinery Blade Design Systems, Belgium

[4] Hah, C., Puterbaugh, S.L., Wadia A.R. (1998), Control of Shock Structure and SecondaryFlow Field inside Transonic Compressor Rotors through aerodynamic Sweep, ASMEPaper 98-GT-561, Stockholm - Sweden

[5] Yamaguchi, N., Tominaga, T., Hattori, S., Mitsubishi, T. (1991), Secondary-LossReduction by Forward-Skewing of Axial Compressor Rotor Blading, Proceedings of 1991Yokohama International Gas Turbine Congress, Vol. 2, pp. 61-68

[6] Schulze, G., Hennecke, D.K., Sieber J., Wörhl B. (1994), Der neue Verdichterprüfstandan der TH Darmstadt, VDI Berichte Nr. 1109, Germany

[7] Blaha, C., Kablitz S., Hennecke, D.K., Schmidt-Eisenlohr, U., Pirker, K., Haselhoff, S.(2000), Numerical Investigation of the Flow in an Aft-Swept Transonic CompressorRotor, ASME Paper 2000-GT-0490, Munich - Germany

[8] Fritsch, G., Möhres, W. (1997), Multistage Simulations for Turbomachinery Design onParallel Architectures, presented at the Parallel Computational Fluid Dynamics Conf.