Flowin Turbo Machines

5/20/2018 Flowin Turbo Machines

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Over more than five decades, the VKI hasspecialized in the area of compressors,pumps and turbines, covering theoretical,experimental and numerical teaching andresearch activities. The VKI is recognizedas one of the centres of excellence in the

area of turbomachinery-related activities andhas attracted world-renowned specialists inthe field. The principal research activitiesare related to design methods, low- andhigh-speed compressor and turbine flowsand heat transfer. Significant contributionswere made in the areas of inverse designand multi-disciplinary optimization meth-ods, unsteady flows and blade row inter-actions, aero-thermal investigations in HPand LP turbomachinery components, microgas turbines, blade-cooling applicationsand instrumentation development.

FLOW IN

TURBOMACHINES

2006, von Karman Institute for Fluid Dynamics, Rhode-St-Gense, Belgium


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Flow In Turbomachines

DETERMINATION OF THE EFFICIENCY

OF A COOLED HP TURBINE IN

A COMPRESSION TUBE FACILITY

The experimental verification of the efficiency of anyturbine stage is a crucial step in the development ofany new turbine. The isentropic efficiency is definedas the ratio between thepower effectively extractedfrom the fluid to the powerobtained from an isentropicexpansion, which requirestesting the turbine stage un-der adiabatic conditions.

Efficiency is usually meas-ured in continuously run-ning facilities that allowsstabilizing the turbineregime and performing de-tailed area traverses with ac-curate instrumentation. Insuch type of facilities, twoindependent techniques canbe used to measure theisentropic efficiency: thethermodynamic and the me-chanical method.

Since the late 70s, the use of short-duration facilitieshas enabled testing at the actual engine levels ofReynolds number, Mach number, gas-to-wall and gas-to-coolant temperature ratios. The determination ofthe efficiency in such facilities is challenging becauseall the needed quantities should be measured during therunning time, i.e., usually in less than 1s. Temperaturefluctuations are usually quite large, implying tempera-ture inaccuracies and thus the mechanical method ispreferred, based on accurate torque measurements.

The current investigation demonstrated the potentialof the VKI compression tube facility to determine the

efficiency of a cooled turbine stage using the me-chanical method. The formulation proposed heretakes into account the different thermodynamic prop-erties of the coolant and leakage flows. It also usesmass-flow averaging in order to obtain a value that isrepresentative of the whole stage.

The measurement of the power relies on the meas-urement of the acceleration during the run time.This requires also the determination of the inertia ofthe rotating assembly, of the rotational speed historyand the mechanical losses. The very fine precision

achieved on the rotational speed results in a high ac-curacy on the acceleration and hence the power.

Inlet total pressure and temperature as well as stageexit total pressure were determined with accurateprobes (Figure 1). Each quantity is the result of aver-aging several heads placed at different circumferen-tial locations around the test section. The knowledgeof the overall mass flow combined with that of the pitch-wise and radial distribution of total temperature and

flow angle allows thecalculation of mass-averaged values usingquantities determinedat each test and di-mensionless profilesresulting from several

tests.

Results are presentedin Figure 2 indicatingan efficiency of0.9088 at designpoint. Figure 2 alsoshows that the evolu-tion of the efficiencyas a function of therotational speed isin line with the 1D pre-diction method. A lev-

el of repeatability between 0.27 and 0.51% is achieved,depending on the test conditions. The uncertaintyanalysis provides a more conservative value of 0.68%for the random error. The largest contributors are themass flow, the stage exit and inlet total pressure.Regarding the systematic error, an uncertainty level of1.2% is achieved mostly due to the contributions of themass flow, the inlet total pressure and the inertia.

Figure 2: Measured levels of efficiency,comparison with non-isentropic simple

radial equilibrium (NISRE) and turbine map

Figure 1: -a,f,g- combined probes with a Kiel head anda thermocouple used as reference probes upstream anddownstream of the stage and inside the cavities, -b,c-rake of Kiel heads, -d- rake of thermocouples, -e- fivehole probe

-a-

-b-

-c- -d- -e- -f-

-g-



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Flow in Turbomachines

TIME-AVERAGED AND TIME-RESOLVED

MEASUREMENTS ON A FILM-COOLED

ROTOR BLADE IN A TRANSONICHIGH PRESSURE TURBINE

The high-pressure turbine is located down-stream of the combustion chamber;hence, the blades experience a harshenvironment. Turbine blades mustbe made of super-alloys and beactively cooled. Rotor blades ofmodern high-pressure tur-bines are equipped withseveral span-wise rows ofcooling holes. This researchfocuses on the effect of rotorfilm cooling on the perform-ance of a highly loaded tran-sonic turbine stage. The threedifferent film-cooling config-urations investigated are dis-played in Figure 1. The presentinvestigation was carried out forrotor coolant flow rates of 0%,0.5% and 0.8% of the mainstreammass flow.

Tests were performed respecting the engine temper-ature ratios:Tgas/Twall and Tcoolant/ Twall. To reduce the

cost of the testing the three different blade-cooling

geometries are tested on a unique rotor disk dividedinto three sectors (Figure 2). Measurements includedthe rotor mid-span static pressure (Figure 1c) and the

heat transfer at 15%, 50% and 85% span.

The increase of the coolant rate causes an increase ofthe stator downstream static pressure, i.e. an increasein the stage degree of reaction. As a consequence, the

stator exit velocity triangle changes and therotor static pressure field is modified ac-

cordingly. In the presence ofcoolant, the heat flux down-

stream of the coolant holes isslightly reduced. Figure 3

presents the effect of the ro-tor cooling on the time-re-solved static pressure on

the rotor crown at 50%span. The large andabrupt pressure increaseat gauge 3, stator 0.8 is

due to the vane trailingedge shock that sweeps

the rotor blade suctionside from gauge 5 towards

the leading edge. At gauge 3the amplitude of the fluctua-

tion amounts to ~20% of P01. In

spite of the small increase in cool-ing rate from 0.5% to 0.8%, the shock

passage tends to occur sooner andsooner. This means that the rotorcoolant ejection affects the vane trail-ing edge shock position and intensi-ty. However this effect is also partlyattributed to the ejection of the rotorhub disk leakage flow.

Figure 1: Cooling configurations: a) Two rows of16 cylindrical holes on the pressure side; b) Tworows of 16 fan-shaped holes on the pressure side;c) Single row of 16 fan-shaped holes on the suc-tion side

Figure 2: View from upstream ofthe rotor disk with the threedifferent sectors of cooled blades

Figure 3: Pressure fluctuations as a function ofvane passing events at mid-span for the differentcooling rates

41 2006, von Karman Institute for Fluid Dynamics, Rhode-St-Gense, Belgium


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The optimization is performed on the stator bladeMach number distribution. The variation of the staticpressure across a pitch at the exit of the stator results

from the effect of the right running shock (RRS infigure 1) and of the reflection of the left running shock(RLRS). Because the combination of these two contri-butions was already well balanced in the initial design,only a modest reduction is achieved with the newdesign.

More promising results were obtained on the rotor.The original rotor profile was redesigned for two stag-ger angles while keeping the inlet and exit blade met-

al angles unchanged (Figure 1). For each geometry, theunsteady force is quantified using an unsteady quasi3D Euler code. The effect of the vane trailing edgeshock on the rotor depends on the vane exit Machnumber (shock strength), on the incidence of the shockon the rotor surface, on the local Mach number distri-bution and on how the shock is reflected on the rotorsurface. The comparison between the original stageand the new configurations shows that the best resultsare obtained with the low stagger, or aft-loaded, air-foil. For this blade profile a decrease of ~50% of thetime-resolved force modulus is achieved (Figure 2).

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Figure 2: Unsteady blade force andmodulus for the three rotor designs

ROTOR - STATOR INTERACTION

IN A TRANSONIC HP TURBINE,

MECHANISMS AND REDESIGN

HP turbine stages with choked vanes present largepitch-wise variations of static pressure, due to thepresence of the vane trailing edge shocks. The vanetrailing edge shock sweeping the crown of the rotorcauses very large fluctuations of unsteady force in therotor. This high-intensity and high-frequency forcingfunction may lead to high-cycle-fatigue problems thateventually would result in unexpected blade failure.Moreover it has been proved that the azimuthal non-uniformities present at the rotor exit are due to a large

extent to the vane exit static pressure signature.

The aim of this research was to investigate how thestator and rotor blade could be redesigned to reducethe rotor pulsating force. For this purpose, a numberof design and analysis tools are used. The new designstarts from the existing BRITE transonic stage that iscurrently tested in the VKI CT-3 facility.

The first step is the redesign of the stator blade pro-file to reduce the vane exit pitch-wise static pres-suregradient. Several approaches were applied and themost effective results were obtained with an opti-mization procedure using an Artificial Neural Network.

Figure 1: Illustrationof vane trailing edgeshock system; threerotor designs testedfor rotor/stator inter-action


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CLOCKING EFFECTS IN A ONE-AND-A-HALF

STAGE TRANSONIC TURBINE

The understanding of pressure and heat transfer fluc-tuations associated in a multi-row environment is fun-damental to improve both the aerodynamic perform-ance and mechanical integrity of future generations ofaero-engines. The current research focuses on the ex-perimental characterization of the time-averaged andtime-resolved flow field on the second stator of a oneand a half stage high-pressure turbine, at variousclocking positions. Clockingis the relative pitch-wise posi-tion of the first stator with re-

spect to the second stator. Anumber of previous investiga-tions have highlightedpotential performance im-provements when the secondstator is aligned with the wakeavenues of the first stator.

Tests have been carried out inthe VKI compression tube fa-cility CT-3 at engine represen-tative conditions. The test pro-gram includes four differentclocking positions, as dis-played in Figure 1. Pneumaticprobes located upstream anddownstream of the secondstator provide the time-averaged component of thepressure field. For the second stator airfoil, both time-averaged and time-resolved pressure fields are meas-ured at 15, 50 and 85% span with 54 fast responsepressure and heat transfer gauges.

Regarding time-averaged results, significant differ-ences on the front suction side are observed at 15%span (Figure 2). Overall, clocking 1 shows a smoother

acceleration on the front suction side; i.e., lower aero-dynamic losses should be obtained for this relative po-sition. The time-resolved data show fluctuations ofstatic pressure up to 33% of the inlet total pressure.The pressure fluctuations are attributed to the passageof pressure gradients linked to the traversing of theupstream rotor. Figure 2 indicates how the rotor trail-ing edge shock pattern is being modified as the rotortraverses the vane passage. The pattern of these fluc-tuations changes noticeably as a function of clocking,especially at 15% of the span (Figure 3).

The time-resolved pres-sure distribution was in-

tegrated along the air-foil surface to determinethe unsteady forces ap-plied on the second sta-tor. The largest steadyforces at 15% span areobtained for clocking 2.When compared withclocking 0, the axialcomponent of the forceincreased by 3.4%and the azimuthalcomponent by 5.3%.Concerning the time-re-solved forces, the bladesection experiencesfluctuations up to 29%

of the mean force. Clocking the vanes could attenuatethe unsteady forces and moments by 17% and 28% re-spectively. This should lead to an optimisation from amechanical integrity point of view.

Figure 1: Turbine airfoil geometryand clocking positions

Figure 2: Time-averaged staticpressure distribution at 15% span Figure 3: Phase-

locked averageof the pressurefluctuations alongthe second statorat mid-span forclocking 0


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DETERMINATION OF THE MASS FLOW

IN A COMPRESSION TUBE FACILITY

Compression tube facilities produce a blow-down ofhot gas on a cold model in order to simulate bothaerodynamic and heat transfer effects. The VKI com-pression tube facility CT-3 is displayed in Figure 1.Constant conditions are usually maintained duringabout 0.5s. Owing to this short testing time, the ac-cu-rate determination of the overall mass flow in the testsection is a challenging task.

In this facility standard metering systems such as aVenturi or orifice plates are difficult to install. Indeed,they require very long straight tubes upstream anddownstream of the test section. A novel technique hasbeen developed consisting in the prediction of the lo-cation of the piston, the total pressure and the totaltemperature downstream of the piston. Therefore themass of air in the upstream tube is known at any time.To have a coherent estimate of all the quantities, themodel is fitted to measured values. Balances of massflow and energy are performed as a function of timein the different volumes of the rig, i.e. the upstreamtube, the settling chamber and the downstream dump

tank. Turbine coolant and leakage flows are also beingconsidered.

To illustrate how accurately the model fits the rig op-eration, the pressure histories in the tube (Ptube) and

upstream of the stage (P01) are reported in Figure 2 (left).

The measured and predicted total temperature coin-cides only sporadically (Figure 2, right). The thermo-couple is implemented on the vertical end plate of thetube because it must be away from the path of the pis-ton. In this zone, the circulation of the flow is not op-timum. The thermocouple is capable of measuring cor-rectly the tube temperature only from time to time,

indicating that the predicted temperature history is re-liable. The location of the piston is checked only in twopoints of the trajectory using a photodiode.

In Figure 2, an extreme case is shown where theinlet total pressure is not constant. If the down-stream sonic throat is not well matched to theincoming flow rate, the mass flow conservationcannot be applied strictly between the differentplanes owing to storage or release of air in theupstream and downstream volumes. In orderto take into account possible differences be-tween the tube exit and the stage inlet, which

are separated by a settling chamber, theturbine is assimilated to a sonic orifice.The mass flow can thus be computed ac-cording to:

P01 and T01 are measured but S and cD should be de-

termined. For this purpose, once the model is fitted toan ensemble of measured P0tube and P01 histories, a

test is simulated in which the total pressure during therun time is perfectly constant. In this case, the massflow conservation between the tube exit and the stageinlet can be applied and the determination of Sand cDis straightforward.

The uncertainty on the mass flow was split into twocontributions. While the piston speed and the tubepressure measurement are accounted for as randomerrors, the reservoir dimension and the temperaturemeasurement (~4K) are considered as systematic er-rors. While the random error is acceptable (0.58%), thesystematic error remains quite large (0.84%). Thelargest contributions are linked to the measurementof the tube temperature and to the piston speeddetermination.


Figure 1: Sketch of the compression tube facility

Figure 2: Left: Measured andpredicted pressure in the tube

and turbine inlet; Right: Temperatures

m c P

c TSD

p

=

+

+( )

( )01

01

1

2 1

1

1

2



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LOW-PRESSURE TURBINE AERODYNAMIC

PERFORMANCE

The new generation of civil aircraft engines is charac-terized by large fans and high by-passratios. As the propulsion effi-ciency is directly linked tothis component, the require-ments imposed for its de-sign are extremely se-vere. Consequently, thedesign of the low-pres-sure turbine, which isdriving the fan and the

low-pressure compres-sor, has to be performedvery carefully.

The current aerody-namic performanceof LP turbines is alreadyhigh but significant savings are still achievablethrough a reduction of the weight of this componentby e.g. using less airfoils per blade row. A reducednumber of blades implies a higher loading. A frontloaded blade design philosophy can address this re-quirement even if strong adverse pressure gradientsenhance the risk of separation along the rear suction

side portion, considering the low Reynolds numberconditions existing in this part of the machine. The ex-tension of the profile losses, and therefore the per-formance of the turbine, will be directly related to theoccurrence of separation and/or transition. It has beenshown that the unsteady/periodic effect of upstreamwakes and their interaction with the boundary layermay lead to a reduction of the profile losses.

Research on the effect of periodically wake-inducedtransition to allow high blade loading with acceptableprofile loss levels is conducted in the LaboratoryJacques Chauvin [AJ24, MP131, MP132]. This workis performed in the Light Piston Isentropic

Compression Tube facility CT-2, allowing a correct sim-ulation of Mach and Reynolds numbers as encoun-tered in a real engine. A stationary linear cascademodel is subjected to wakes shed from upstream barsmounted on a fast rotating disk (Figure 1). This wakegenerator provides the advantage of being able todrive the bars at the velocity required to simulate thecorrect inlet velocity triangle.

The cascade was made of very high lift blades. Theprofile losses and the exit flow angle were determinedfor an exit Mach number equal to 0.8. The experimen-tal parameters were the freestream exit Reynolds

number (1.9 x 105 6.8 x 105) and Strouhal number(0 0.9). The positive impact of the wakes can be ap-preciated in Figure 2, especially at low Reynolds num-

ber where the losses are reduced by 46.7% when in-creasing the Strouhal number.

The boundary layer status and characteristics were de-termined, quantitatively, by means of a heat fluxmeasurement technique. These measurements werecarried out by means of platinum thin film gauges, de-posited on a ceramic substrate. They operate in a tran-sient mode, based on the semi-infinite slab principle.

Because of their very small thickness, they providequantitative information in a broad frequency spec-trum and do not perturb the flow by, e.g. promotinglaminar to turbulent transition. They allow the charac-terization of the boundary layer status by differentiat-ing the laminar from the turbulent state. In this way, itbecomes possible to observe the temporal evolutionof the transition and to compute critical parameterssuch as the intermittency factor. A space-time diagramof the phase locked averaged suction side heat trans-fer coefficient is presented in figure 3. The successivewakes of the bars are identified in the leading edge re-gion (small s/L values) along the time axis. This resultclearly shows the periodic suppression of the separa-tion bubble (characterized by very low heat transfercoefficient values) in the wake path of the upstreambars and its complex redistribution under these

unsteady conditions.


Figure 1: Wake generator

Figure 2:Profile losscoefficient as afunction of Sr

Figure 3:Phase lockedheat transfercoefficient



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EXPERIMENTAL STUDY OF A VERY HIGH

LIFT, LOW PRESSURE TURBINE BLADE

One of the important research objectives in gas tur-bine engines is to decrease the production and main-tenance costs by reducing the total engine weightwhile maintaining high levels of efficiency. In this per-spective, the low pressure turbine, accounting forabout 1/3 of the total engine weight, is of prime inter-est. The solution consisting in a reduced number ofblades per row results in a new design philosophycalled the high lift design.

A front loaded, very high lift, low pressure turbine

blade has been designed at the VKI, for both com-pressible and incompressible flow regimes. Thegeometries were adapted in order to present the samepressure coefficient distribution. The experimentalstudy for the incompressible flow regime, carried outin the low speed C-1 cascade wind tunnel of the labo-ratory Jacques Chauvin, led to the characterizationof the cascades overall aerodynamic performance(losses and mean outlet flow angle at midspan) andof the secondary flows. This investigation was per-formed by varying inlet incidence and freestreamReynolds number over a wide range. Conventionalpneumatic pressure probes were used to quantify lo-cal total and static pressures as well as the pitch- and

yaw flow angles. A single hot wire probe provided theturbulence characteristics, namely intensity and lengthscales. A typical pitch and spanwise loss distributionis pre-sented in Figure 1 for the nominal flow condi-tions; the main secondary flows are indicated as well.This investigation clearly showed, among other con-clusions, how the profile losses measured at midspandecreased when increasing Reynolds number orwhen decreasing inlet flow incidence. The mean out-let flow angle, measured from the axial direction, de-creased when decreasing the Reynolds number orwhen increasing the inlet incidence.

The characterization of the compressible profile wasperformed in the Isentropic Compression Tube facili-ty, including the influence of periodic incoming wakes.

The high lift design philosophy implies, among oth-er features, the existence of a rather important adversepressure gradient along the rear part of the suctionside. This eventually leads to the development of aseparation bubble whose typical characteristics (sep-aration, maximum displacement and reattachment)have been quantified, analyzed and modelled; one ofthese models [AJ42] is compared in Figure 2 to exist-ing models extracted from the literature (Roberts,Mayle or Hatman and Wang). The main difference withthe models of Roberts and Mayle is the description ofthe transition process inside the free shear layer of theseparated flow. Roberts and Mayle consider that the

complete transition to turbulent flow is reached beforethe reattachment respectively as a punctual or an ex-tended process. The present results agree better withthe bubble physics as described by Hatman and Wang.The transition is initiated in the free shear layer andends before or after the reattachment depending onthe bubble type: long or short. The model developedfrom the present results is valid for all bubble typesand can be used in numerical predictions as a criteri-on for the separated flow mode of transition.

PERFORMANCE ANALYSIS OF COOLINGPASSAGES

Nowadays, the need for high power and high effi-ciency gas turbine engines leads to a continuous in-crease of the turbine inlet temperature (TIT). The lat-ter is usually far above the melting temperature of themost advanced vane and blade materials. A safe op-eration of the engine depends therefore on the effi-ciency of the cooling system or systems and a com-plete understanding of the convection mechanisms



Figure 1: Downstream losses and secondary flows

Figure 2: LocalReynolds number atthe maximum bubbledisplacement as afunction of the localReynolds number atseparation


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resulting from the cooling techniques is mandatory inorder to predict reliable metal temperatures and con-sequently component lifetimes. The present contribu-

tion therefore focuses on the analysis of the aero-ther-mal performance of a stationary rib-roughenedcooling channel flow, as encountered inside high pres-sure turbine vanes.

The analysis is carried out on the aerodynamic (ve-locity field) and convective heat transfer data recent-ly acquired in a large scale, stationary cooling chan-nel test section. The selected channel and ribgeometries constitute a quite detailed test case understudy at the VKI for the last few years [AJ67,MP6,MP30,MP35]. Several, sometimes redundant, meas-urement techniques were used for this purpose in or-der to provide high quality data. Both the aerodynamic

and heat transfer data are acquired in between twoconsecutive ribs. The use of a Digital Particle ImageVelocimetry (DPIV) technique allows the acquisition ofhighly detailed flow field measurements over differ-ent and mutually perpendicular planes in the fullinterrib space. A quasi 3D view of the flow field istherefore available and its interpretation provides anexhaustive description of the mean flow topology.The heat transfer data were gathered with the help ofa Liquid Crystal Thermometry technique. It was ap-plied on all the channel walls, rib surfaces included,and provided in this way the complete heat transfercoefficient distribution along the investigated channelregion.

The combined analysis of the aerodynamic resultswith the wall heat transfer data allows pointing out therole played by the mean flow features and by the ve-locity fluctuations in the heat transfer process. An earlyanalysis of the present data put in evidence a strongcorrelation existing between high levels of heat trans-fer and velocity fluctuations normal to the wall.This is shown in Figure 1, on the ribbed wall, along thechannel symmetry line and also along a second line,closer to the channel lateral wall. This correlation didhowever not seem to be unique in all flow regions.As a matter of fact, mainly on top of the rib, all the ve-

locity component fluctuations were of the same order,as reported in Figure 2. The analysis of the PIV data,also by means of a Coherent Structure Eduction tech-nique, pointed out the existence of a stationaryrecirculation region located on top of the obstacle,which could induce different flow mechanisms re-sponsible for the heat transfer process.

This analysis shows that it might be possible to de-termine which are the flow field features directly re-sponsible for the heat transfer process developmentand augmentation. In particular, the most promisingidea seems to be the one which looks at the wall

events generated by the turbulent structures evolvingin the near wall regions. These events make the heattransfer process more effective but also much more

complicated to study and to understand. In order toget additional information on the flow behavior, somenumerical simulations using advanced turbulencemodeling and Large Eddy Simulation are under way.The analysis of the numerical results and a deeper in-terpretation of the available data will lead to the finaldefinition of the flow mechanisms responsible for theheat transfer field development.

AERODYNAMIC PERFORMANCE

INVESTIGATION OF FIXED RIB-ROUGHENED

COOLING PASSAGES

In modern gas turbine engines, the continuous in-crease of power for an expected lifetime has resultedin a continuous increase of cycle pressure ratio andturbine inlet temperature. The latter implies that

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Figure 1: Comparison between heat transfer andvelocity fluctuations normal to wall between tworibs

Figure 2: Comparison between velocityfluctuations on top of the rib


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advanced materials and cooling techniques must beadopted for a safe operation of the HP gas turbineblades. Apart from the progress made in the metallur-

gical domain, a continuous cooling of the blades of theturbine first stage allows operating at temperatureswhich are far above the materials melting point, with-out affecting the component integrity and geometry. Acomplete and correct understanding of the convectionmechanisms associated to the applied cooling tech-niques is therefore of major importance. The presentexperimental study therefore deals with the detailedaerodynamic investigation of the turbulent flow insiderib-roughened turbine blade cooling channels of vari-ous geometries. Inside these passages, the forced con-vection cooling process is significantly enhanced bythe presence of ribs (turbulence promoters) installedon one or more walls of the channels. The effects of the

rib size and orientation on the flow behaviour, andtherefore on the heat transfer and pressure distribu-tion along the channel walls, are closely related to thesafe operation of HP gas turbine blades.

The present research effort is devoted to the experi-mental analysis of the flow inside two cooling chan-nels of different geometries on the basis of detailedaerodynamic measurements performed by DigitalParticle Image Velocimetry (DPIV). The principal aimsare to provide new information about the behaviourof such a complicated flow, useful for its understand-ing, to complement the wall heat transfer coefficientdistributions already available and to create a wide andreliable data base for numerical code validation [AJ67].

The experiments are carried out on scaled-up modelsof turbine nozzle blade cooling channels, working ingeometrical and flow similarity conditions. In bothconfigurations, ribs are installed on one wall. The firstgeometry is characterized by a square section, the ribshave an angle of attack with respect to the "mean" flowdirection equal to 90 and a high blockage ratio of30%. The second channel presents a rectangular sec-tion with the ribs inclined at 45 to the main axis of thechannel.

In both channels, a global three dimensional view ofthe flow was attempted by measuring the whole flowfield in between two consecutive ribs over different

and mutually perpendicular planes. Where previousinformation was available, comparisons were madewith, e.g., results from alternative measurement tech-niques, showing most of the time a remarkable agree-ment and providing complementary information.Moreover, the data analysis provides a lot of informa-tion regarding particular flow structures in terms oftheir location and size, as shown in Figure 1 for thesquare section channel. For some of these structures, amodel describing their evolution was proposed [MP6].

The combined interpretation ofthe PIV results and of the avail-able heat transfer measure-ments on the channel walls,highlighted the existing linksbetween the aerodynamic andthermal behaviour of the flow[MP30, MP35]. A typical exam-ple, obtained in the second

cooling channel, is reported in Figure 2. The meanflow path measured by DPIV as close as possible tothe ribbed wall and the Nu number distribution(coloured contour plot) measured by Liquid CrystalThermometry on the ribbed wall are fully coherent withrespect to separation and reattachment area location.

COMPARISON OF TURBINE TIP LEAKAGE

FLOW FOR FLAT TIP AND SQUEALER TIP

GEOMETRIES AT HIGH-SPEED CONDITIONS

In modern gas turbine engines, the recent trend of in-creased combustor outlet temperature to achieve

higher thermal efficiency and higher power outputposes a challenge to the aerodynamics, heat transfer,and material capabilities of the first stages of the high-pressure turbine. Tip leakage flows are a source ofaerodynamic inefficiency and high thermal loadingnear the tip. It is quite difficult to cool the blade tipthrough conventional film-cooling techniques. Notonly is the leakage flow path three-dimensional in na-ture, but strong secondary flows also cause very hotmainstream flow to enter the tip clearance region. Theblade tip is one of the most frequently inspected andrepaired parts of the turbine.



Figure 1: 3D flowtopology upstreamof the rib

Figure 2: Comparisonbetween wall heat trans-fer and streamlines


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Within the EC-fundedprogram AITEB, thetip leakage flow char-

acteristics for flat andsquealer tip geome-tries are studied in thevon Karman InstitutesIsentropic Light PistonCompression Tube fa-cility, CT-2, at differentReynolds and Machnumber conditions fora fixed value of the tipgap in a non-rotating,linear cascade arran-gement. This repre-sents one of the first

high-speed tip flowdata sets for the flat tip

and squealer tip geometries. Static pressure meas-urements are made at the inlet and exit of the cascade,and the inlet endwall boundary layer profiles are ob-tained to provide proper boundary conditions for CFDcalculations. Static pressure measurements are madeon the blade tip, blade surface, and on the correspon-ding endwall. Aerodynamic losses are computed us-ing total pressures measured downstream of the cas-cade by a traverse of a 3-hole pressure probe. Finally,tip heat transfer measurements are performed alongboth a flat and a squealer tip configuration [MP133,MP174].

Oil flow visualizations provide valuable information ofthe flow on the tip and near-tip surfaces. While themain flow through the tip gap of the flat tip blade trav-els from pressure to suction side, there is also a re-gion just downstream of the leading edge regionwhere the flow along the tip actually travels in thestreamwise direction. For the squealer tip blade, flowvisualization results are in line with a recirculating flowwithin the squealer tip cavity where the flow on thecavity floor is moving from suction side to pressureside.

Static pressure measurements made on the endwallover the blade tip show that the squealer tip providesa significant decrease in velocity through the tip gapwith respect to the flat tip blade. The flat tip blade

shows a region of highvelocity extendingdownstream of theleading edge region.For the flat tip, an in-crease in Reynoldsnumber causes a largeincrease in tip velocitylevels, but the squealer

tip is relatively insensitive to changes in Reynoldsnumber. The only region of the endwall where thesquealer tip does not result in lower overall velocities

compared to the flat tip is near the leading edge. Flowvisualization results show evidence of flow impinge-ment existing near the leading edge of the squealercavity.

Twelve heat transfer gauges (Figure 2) were distrib-uted over the tip area, along the suction side and pres-sure side in the front part of the tip area and along thecamberline on the rear part of the tip area. The tech-nique applied for this purpose is based on the use ofthin films and a transient approach. During the blow-down, the gaugesrecord the local sur-face temperature

evolution as a func-tion of time. This in-formation is used asthe main boundarycondition for an un-steady conductioncalculation within thesubstrate supportingthe heat transfergauges.

This calculation finally provides the wall heat flux andsubsequently the heat transfer coefficient and Nusseltnumber distributions. A typical example of normalizedNusselt number distribution is presented in Figure 3.It demonstrates the effect of freestream Reynoldsnumber on the augmentation of heat flux. It also com-pares the present larger scale stationary cascade re-sults with the results obtained on the high speed tur-bine in rotation. Both configurations were run atsimilar working conditions and present very compa-rable conclusions.

EXPERIMENTAL STUDY

OF THE TRANSITIONAL FLOW

IN AXIAL TURBOMACHINES

The actual evolution in gas turbine engines is to de-crease the production and maintenance costs througha reduction of the total engine weight. One of the so-lutions consists in reducing the number of bladeskeeping constant the total amount of work per bladerow. A high lift, high load design philosophy musttherefore be applied. As a consequence, the suctionside boundary layer undergoes severe adverse pres-sure gradients along its rear part. Considering the low

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-a- -b-

Figure 1: Flow visualization onsquealer (a) and flat (b) tipblades

Figure 2: Typical heat flux gaugeimplementation


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MULTIDISCIPLINARY OPTIMIZATION OF

TURBOMACHINERY COMPONENTS

The basic principle of the VKI design and optimizationmethod is to mimic the traditional design procedurein which the designer makes a first optimization bymeans of a geometry generation and an approximatebut fast analysis method, followed by a more accuratebut computational computationally expensive verifi-cation by a Navier-Stokes solver.

This blade design algorithm, of which a flowchart ispresented in Figure 1, starts from the user-definedaerodynamic and mechanical requirements as there

are: inlet and outlet flow angles, the outlet static pres-sure, the Reynolds number, the maximum/minimumblade cross-sectional area and moment of inertia (Iminand Imax), maximum allowable lean and rake angles, etc.

The Artificial Neural Network(ANN) provides the ap-proximate between the geometry , theboundary conditions , and the performance . TheANN Learning defines this relation using the informa-tion stored in the database. The optimizer is based on agenetic algorithm using the trained ANN to evaluate theperformance of the new blade geometries.

The resulting geometry, which is optimum accordingto the ANN predictions, is then verified by means ofthe accurate Navier-Stokes solver TRAF3D. The geom-etry and performance are added as a new sample tothe database and a new ANN learning is started. Asthe database grows after each Navier-Stokes compu-tation, the approximate relation is expected to becomemore accurate. In this way the system is self-learning.The approximate optimization cycle is repeated untilthe performance check confirms that the ANN opti-mum is also the real optimum geometry.


Figure 2: Non dimension-al downstream losses

Figure 3: Smooth bladevelocity distribution

P f G B= ( , )

B

GP


Reynolds number environment prevailing in low pres-sure turbines, this could eventually lead to a heavy

separation which will se-

riously hinder the bladeaerodynamic perform-ance. A careful controland understanding oflaminar to turbulent tran-sition are therefore of ma-jor importance.

In order to tackle thisproblem, the continuoushigh speed facility S-1 ofthe von Karman Institutewas equipped with a new

test section for aerodynamic performance determination

of low pressure turbine blades mounted in a linear cas-cade environment (Figure 1). This continuous, cold-flow,high speed cascade tunnel operates at Reynolds andMach numbers similar to those encountered in the lowpressure section of a modern gas turbine (Re= 50 300 x 105 , M= 0.5 1.0). Static and total pressure andfreestream turbulence characteristics are measured up-stream of the cascade, time-averaged and time-resolvedstatic pressure and semi-quantitative skin friction meas-urements are performed along the airfoil and perform-ance measurements are conducted downstream of theblade row by pneumatic or fast response probes and op-tical techniques. Provision is also made for periodic up-stream wake/boundary layer interaction. A rotating disk,equipped with cylindrical bars, generates wakes up-stream of the cascade, with a correct simulation of the up-stream velocity triangle. The test section dimensions are225 x 500mm.

The well-known T106 blade profile with an increasedpitch-to-chord ratio was studied in this new setup un-der steady conditions, at low Reynolds numbers. Theoriginal profile was also equipped with a local rough-ness element in an attempt to control the transitiononset along the suction surface. The evolution of thedownstream losses as a function of Reynolds numberis presented in non-dimensional form in Figure 2 with-

out (smooth) and with (rough) artificial roughness.Along the smooth profile, the laminar suction sideboundary layer definitely separates at low Reynoldsnumber, leading to high losses and poor performance.The importance of the separation depends on theReynolds number, as seen in Figure 3, presenting theblade Mach number distribution for three different exitReynolds number values. The extension of the sepa-ration, and therefore the losses, clearly reduce whenincreasing the Reynolds number. The introduction ofartificial roughness resulted in an earlier induced tran-sition, avoiding separation, and therefore leading to abetter performance for the Reynolds number range ofinterest.

Figure 1: New cascade modelin the S-1 facility


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The cost of an opti-mization mainly de-pends on the accuracy

of the ANN because ithas a direct impact onthe required numberof Navier-Stokes cal-culations. The DOE(Design of Experi-ment) technique isused to create themost representativedatabase with a mini-mum number of sam-ples. Investiga-tions tofind the optimal pa-rameter setting of the

GA aim for an acceler-ated convergence by reducing the number of functionevaluations of the optimizer [MP177].

A new topic of re-search is multi-pointoptimization, wherethe optimum geome-try is a function of theperformance at designand off-design condi-tions. This has beenachieved without anincrease of the design

time by a paralleliza-tion of the calcula-tions, using as manyprocessors as the op-erating points [MP217,MP178].

The system is presently being extended to MultiDisciplinary Optimization. It is an extension of the ex-isting method (Figure 1) combining the mechanical andaerodynamic optimization (Figure 2) [IB18, MP250]. TheGA searching for the optimum geometry gets its inputfrom a Finite Element Stress Analysis (FEA) as well asfrom the Navier-Stokes flow analysis.

The main advantages of such an approach are:

The method has been used to optimize a 2D micro-gasturbine impeller geometry by searching for the op-timum combination of blade height and trailing edgeangle. It aims to maximize the efficiency while re-specting the stress limitations (Figure 3).

MICRO GAS TURBINES

Micro gas turbines have experienced a growing interestduring the last decade. Their large energy density(Whr/kg) makes them an attractive replacement formuch heavier batteries for the propulsion of small air-planes (UAV) or portable power units [MP207,AJ64].Designing micro gas turbines by a simple scaling oflarge high performance gasturbines will not providegood results because the scaling conditions arenot satisfied:


Figure 1: Bladedesign algorithm

Figure 2: Multidisciplinary optimization flow chart

Figure 3: 2D rotor stress dis-tribution at reference bladeheight and thickness

- The existence of only one master geometryi.e. the one defined by the geometrical parame-

ters used in the GA optimizer. It is input to thefully automatic grid generators of both analysisprograms and eliminates possible approxima-tions and errors when transmitting the geo-me-try from one discipline to another- The existence of a combined objective functionaccounting for all disciplines. This allows a moredirect convergence to the optimum geometrywithout iterations between the aerodynamicallyoptimum geometry and the mechanically accept-able one.- The possibility to do parallel calculations. TheFEA and NS analyses can be made in parallel if

each discipline is independent; i.e., if stress cal-culations do not need the pressure distributionon the vanes or flow calculations are not influ-enced by geometry deformations.

- The small dimensions result in a large heattransfer between the hot turbine and cold com-pressor and the assumption of adiabatic flow is

no longer valid [MP109,MP168]. A conjugateheat transfer model (CHT) is needed to evaluatethe heat flux and its impact on the compressorand turbine performance.- Micro gas turbines operate at very low Reynoldsnumbers (10000) and machining techniques maynot allow very smooth surfaces. The increase ofrotational speed results in larger Coriolis forces.The large heat fluxes may induce buoyancy flows.Hence the flow structure may be very differentfrom the traditional one and it is doubtful that theturbulence models are still valid at the micro gasturbine operating conditions.



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Both topics are the subject of intense research inrecent years at VKI.

Conjugate heat transfer method for turbomachines

Most CFD predictions restrict the computations to thefluid domain assuming adiabatic walls or constanttemperature of the solid boundaries. The large heattransfer in micro gas turbines however requires aConjugate Heat Transfer (CHT) calculation on a domaincovering both the fluid and solid parts of the micro gasturbine.

The VKI method couples a non-adiabatic Navier-Stokes (NS) solver for the flow in the fluid domain witha Finite Element Analysis (FEA) for the heat conduc-tion in the solid. Continuity of temperature and heatflux at the common boundaries is obtained by aniterative adjustment of the boundary conditions.

The method starts with an initial guess of the temper-ature distribution Twat the solid boundary of the flow

solver. Substituting the heat flux, predicted by theCFD, into the following relation qw=h(Tw -Tfl)and im-

posing a value for the heat transfer coefficient h, pro-vides a relation that can be used as a boundary con-dition for the solid conduction computation. Theresulting temperature Twat the common boundary is

then imposed as a new wall boundary condition to thefluid solver. This loop is iterated until convergence. Itcan be shown that the value of h influences the con-vergence rate but has no effect on the result once the

method has converged. [MP220,MP225,MP248]

The main advantages of this approach are:- The possibility to use standard NS and FEA solversthat have been extensively verified and for which thelimitations and capabilities are well known.- The possibility to reuse the solid grid to computestresses due to centrifugal forces and even thermalstresses, using the calculated temperature field.- The CHT methods using only one code for solid andfluid calculations suffer from a slow convergence inthe solid part, because the time constants in solids areone order of magnitude higher than in fluids.

- The possibility to mix simpler axisymmetric andmore expensive full 3D calculations, as is demon-strated in the next example.

The method is used to calculate the heat transfer inmicro gas turbines with 10, 20 and 40mm impeller di-ameters. It aims to reveal the different contributions,to quantify the impact of the heat transfer on per-formance and to define possible ways to reduce it.The computational domains are schematically shownin Figure 2. Two separate 3D Navier-Stokes computa-tions are made respectively for the compressor andturbine. Two 3D Finite Element heat transfer Analysesare required: one for the solid compressor and one forthe turbine. A 2D axisymmetric model is used for thestator heat transfer analysis and for the leakage flowbetween the rotor and stator.

The temperature distribution in the solid of a 20mmsteel impeller is shown in Figure 3. In spite of the largetemperature difference between the compressor andturbine impeller there is only a fairly small amount ofheat transfer in the 10mm long but small diametershaft. The analysis shows that the heat transfer by theleakage flow is of equal importance. The largest heattransfer is through the diffuser and can be reduced byinstalling a thermal barrier or by changing the mate-rial. The loss in efficiency due to the heat transfer isevaluated at a few percent only.


Figure 1:Cross sectionof a regenera-tive micro gasturbine

Figure 2: Numericalmodel

Figure 3: Temperature and heat transfer

distribution in a micro gas turbine rotorand stator (steel)



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Time resolved PIV in a rotating channel

An accurate prediction

of the micro gas turbineperformance requires acorrect estimation ofthe impact of Coriolisforces and heat transferon low Reynolds num-ber flows over roughsurfaces. The RC-1 facil-ity has been speciallybuilt to study these ef-fects and to evaluatetheir impact on turbu-lence models. It allowshigh resolution, time re-

solved Particle ImageVelocimetry in rotating

diverging channels. The aim is to build a database forvalidation/modification of existing turbulence modelsof the CFD codes.

The facility consists of a 0.7m long divergent Plexiglaschannel, mounted on a rotating disk. It has a 6 totalaperture angle, a 0.079m inlet hydraulic diameter andrepresents a large scale model of a micro gas-turbineimpeller passage. The volume flow is measured witha Venturi connected to the rotating channel and ad-justed by means of an upstream fan / throttle valvecombination. It respects the main scaling parametersof the radial impeller passage: i.e. Reynolds numberbetween 3 x 103 and 3 x 104, rotational number (Ro)

between 0.1 and 0.55 and Buoyancy numbers up to0.73. A protective cage, connected to an aspirationsystem, collects the seeding particles at the channel

exit.

Relative velocity fields in the rotating channel aremeasured by means of Particle Image Velocimetry us-ing a high speed camera and a continuous laser, bothrotating with the channel. The CMOS camera canrecord up to 7.8kHz with a 640 x 480 pixel resolutionand has a 2Gb internal memory to store the pictures.The air-cooled laser diode provides up to 25W laserlight at 806nm wave length. It is coupled by an opticalfibre cable to the line generator. This compact systemallows a direct and hence more accurate measure-ment of the relative velocity as well as time-accuratemeasurements of the flow variations [MP236,MP229].

Measurements are taken in a plane parallel to thechannel bottom wall, halfway between inlet and out-let. It corresponds to the blade plane in the micro gas-turbine blade. Mean and instantaneous flow fields areobtained in the stationary channel and at different ro-tational speeds.The increasing influence of rotation on the boundarylayer mean flow field can be noticed in Figure 5.Boundary layer thickness increases on the suction sideand decreases on the pressure side. The flow is moreturbulent and less likely to separate on the pressureside than on the suction side.Figure 6 shows how at Ro = 0.33 the vorticity concen-trates closer to the pressure side wall but spreads overa large part of the domain near the suction side.

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Figure 4: Close up

of the rotating channel

Figure 5: Velocity profiles in nonrotating channel (a); on suctionside (b) and pressure side (c) atRo= 0.33

a ) b ) c )

a ) b ) c )

Figure 6: Instantaneous velocityvectors and vorticity in stationarychannel (a), on suction side (b) andpressure side (c) at Ro= 0.33



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54

DESIGN OF A HIGHLY LOADED AXIAL

COMPRESSOR STAGE

FOR THE INVESTIGATIONOF THE INFLUENCE OF CASING

TREATMENTS ON THE STABILITY MARGIN

Over the past 30 years, continuous research has beencarried out concerning the possible benefits of casingtreatments for improving axial compressor perform-ance at off-design conditions. This problem has gainedrecently new interest with the trend to increased bladeand stage loading to reduce costs and engine weight.

In the frame of the European FP6 project VITAL, it isproposed to investigate new engine designs, withhigher bypass ratios but with reduced weight in orderto gain on propulsive efficiency and hence save on fuelburn. It is also proposed to investigate new engine ar-chitectures, like geared turbofans or contra-rotatingturbofan engines, which would enable substantialnoise reductions by lowering the fan speed. All theseobjectives result in higher stage loadings and requireaerodynamic designs which go beyond classicalknow-how.

Based on specifications provided by the engine man-ufacturers, the VKI was in charge of performing the full

3D aerodynamic design of a highly loaded lowpressure compressor (booster) representative of acontrafan architecture (Figure 1).

After calculation of the similarity parameters of non-dimensional massflow, rotational speed and Rey-nolds number to the test rig scale and conditions, themeridional design has shown the necessity to use anIGV row in order to meet the relative flow angles tothe rotor and hence the turning and diffusion for theprescribed work distribution.

The blade profiles have been designed as ControlledDiffusion Blades by the use of an inverse method de-veloped at VKI starting from initial subsonic NACA 65

profiles.

A final design and off-design 3D Navier-Stokes calcu-lation of the geometry permitted the validation of thedesign against specifications and computed the oper-ating map of the stage (Figure 2).

In order to experimentally assess the performance aswell as the stability margin of this particular stage, theVKI has designed a new test section for its high speedaxial compressor test rig R-4, which has been totallyreconditioned after 20 years of inactivity (Figure 3).The stall inception mechanism as well as the tip leak-age vortex structure and its evolution while ap-proaching stall will be investigated in detail by fast re-sponse pressure probes, hot wire measurements andunsteady static pressure measurements in the rotorcasing wall.

It is proposed to investigate two types of casing treat-ments and compare their effect with that of a smoothcasing. Based on preliminary literature research it issuggested to adopt as a first configuration a grooved

casing with either circumferential grooves or obliquegrooves, the latter aligned with the blade chord. Bothtypes have shown large improvements in the operat-ing range with little orno loss in efficiency.The exact geometry ofthe grooves, theirspacing, width anddepth, has still to befixed. The second con-figuration is to be de-termined.



Figure 1: A 3D view of the axial compressorstage

Figure 2: Performance map and predicted stallmargin from 3D Navier-Stokes calculations

Flowin Turbo Machines

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