experimental investigation of the influence of fouling on compressor cascade characteristics and...

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Experimental investigation of the influence of fouling oncompressor cascade characteristics and implicationsfor gas turbine engine performanceD Fouflias1∗, A Gannan1, K Ramsden1, P Pilidis1, D Mba1, J Teixeira1, U Igie1, and P Lambart2

1Gas Turbine Engineering Group, Department of Power and Propulsion, School of Engineering, Cranfield University,Cranfield, Bedfordshire, UK

2R-MC Power Recovery Ltd, Stamford, Lincolnshire, UK

The manuscript was received on 26 February 2010 and was accepted after revision for publication on 14 May 2010.

DOI: 10.1243/09576509JPE992

Abstract: This article describes the findings of a study which examined the influence of foulingon the behaviour of a cascade and by making use of these results the performance implicationsfor gas turbine engines of exposure to airborne foulants.

A suction-type compressor cascade tunnel with a plenum chamber was employed for inves-tigating fouling blade effects. The tests showed that such a testing arrangement allows theextraction of pressure and corrected velocity distribution data downstream of the blades thatis comparable with what can be obtained from blow-type cascade tunnels.

This study presents experimental results for smooth clean cascade blades and for uniformlyfouled blades. For all the cases considered, mid-span-corrected velocity distributions and pres-sure losses taken one chord downstream of the blades were investigated in order to identify theeffects of fouling on the blades. The result of fouling on exit flow angle was investigated as well.

In the present study, cascade clean and fouled cases were used to predict real engine perfor-mance. Results are obtained in terms of stage polytropic efficiency, thermal efficiency, usefulpower, and compressor efficiency deterioration. Roughening the cascade blades uniformly withparticles of 254 μm size, the compressor efficiency dropped by 7.7 percentage points.

Keywords: compressor cascade, polytropic efficiency, thermal efficiency, useful power, compres-sor efficiency

1 INTRODUCTION

The performance of the compressor of an industrialgas turbine can suffer significantly from fouling dueto the ingestion of particles like sand and dust. Invery hostile environments when particles mix with oilvapour, the outcome is a substantial loss in power out-put and cycle efficiency due to compressor fouling.In order to recover this performance loss and subjectto manufacturer’s firing temperature limitations, theengine fuel flow could be increased. This, however,would reduce the turbine blades’ creep life and result

∗Corresponding author: Gas Turbine Engineering Group, Depart-

ment of Power and Propulsion, School of Engineering, Cranfield

University, Cranfield, Bedfordshire MK43 0AL, UK.

email: [email protected]

in large increases in gas turbine engine operating costs.Therefore, investigating the effects of blade roughnessgives valuable knowledge about the mechanisms offouling and how this can be reduced.

Surface roughness effects on compressor bladesstrongly influence the performance of gas turbineengines. Gbadebo et al. [1], applied surface rough-ness on stator blading of a single-stage low-speedaxial compressor. With surface flow visualization andexit loss measurements it was shown that the three-dimensional (3D) separation at the hub was increasedby the presence of roughness. The separation identi-fied by Gbadebo caused a significant reduction in thestage total pressure rise. In addition, the experimen-tal work illustrated that applying surface roughnessbetween the leading edge and the location of thepeak suction caused a significant reduction in stageperformance, whereas applying surface roughness

JPE992 Proc. IMechE Vol. 224 Part A: J. Power and Energy

1008 D Fouflias, A Gannan, K Ramsden, P Pilidis, D Mba, J Teixeira, U Igie, and P Lambart

downstream of the suction peak had a virtuallynegligible effect. Applying roughness on the blades thestage total pressure rise was reduced, but at higher flowcoefficients the effect of roughness was insignificant.

Nikuradse [2] investigated the flow in pipes thatwere roughened with sand grains and divided the flowinto three regions. In the first region, the roughnessof the surface is completely covered by the laminarsublayer, and it was noted that surface roughness hasno influence on the flow. In the second region, someroughness elements protruded through the sublayercausing an increase of the pressure drop coefficient.In the third region, the roughness elements protrudecompletely through the sublayer and the pressuredrop coefficient does not increase further.

Schaffler [3] investigated experimentally the effectof Reynolds number on four multi-stage axial com-pressors and reported three different regimes of oper-ation characterized by the condition of the bladeboundary layer, the laminar separation, the turbu-lent attached flow with hydraulically smooth bladesurface, and the turbulent attached flow with hydrauli-cally rough blade surface. Schaffler defined a lowercritical Reynolds number below which laminar sep-aration takes place. Above this critical Reynolds num-ber, the efficiency increases as the Reynolds numberincreases.This increase stops when the Reynolds num-ber reaches the upper critical Reynolds number, whichis 5 × 105, and the blade surface behaviour is hydrauli-cally rough. Also, it is stated that the lower criticalReynolds number for standard production blades isequal to or below the value of 105. Below this value,significant blade flow separation occurs and the bladecan stall.

Bammert and Milsch [4] ran experiments with com-pressor cascades consisting of roughened NACA 65series blade sections in different geometrical vari-ations and illustrated that loss coefficients rise byincreasing the roughness grade defined as the ratio ofsand grain size over the chord length. Also, they illus-trated that increasing the surface roughness resultedin a decrease of the flow turning angle.

Bammert and Woelk [5] investigated the influ-ence of blade surface roughness on the aerodynamicbehaviour and characteristic of a three-stage axialcompressor model by using emery grain uniformlyroughened blades. For a relative roughness ks/c =4.51 × 10−3, the authors reported a reduction in staticpressure ratio of 30 per cent and a reduction in over-all efficiency of 13 per cent. The equivalent sand grainroughness ks represents the size of sand grains whichgive the same skin friction coefficients in internalpassages as the roughness being evaluated [6].

Bons [7], having done a review of surface rough-ness and effects in gas turbines, states that the firstevidence of roughness influence takes place whenthe roughness reaches an admissible level ks,adm =100c/Rec and this represents a design guideline. The

author also states that for a turbulent boundary layer,exceeding ks,adm a modest increase in blade profilelosses may occur; however, for a laminar bound-ary layer, the profile loss could increase by up toa factor of 2. On the other hand, if the laminarboundary layer was already prone to separation, atlower Reynolds numbers roughness-induced transi-tion could minimize separation and reduce profilelosses.

Tarabrin et al. [8], analysing computational results,reported a 4.5 per cent reduction in mass flow, 4 percent reduction in pressure ratio, and 2 per cent reduc-tion in compressor efficiency, as successive stages ofa gas turbine foul for the first six stages. Tarabrinet al. [9] reported that the gas turbine unit sensitiv-ity to axial compressor fouling decreases by increasingthe turbine entry temperature (TET) and keepingthe compressor pressure ratio constant. Keeping theTET constant and increasing the compressor pres-sure ratio, the gas turbine unit sensitivity to foulingincreases. They also stated that a 1 per cent decreasein axial compressor efficiency of a single shaft gas tur-bine due to fouling can cause a reduction in usefulpower (UW) output of 2.82 per cent, keeping the TETand pressure ratio constant.

Kurz and Brun [10] developed a model for a two-shaft gas turbine engine with power turbine in order toinvestigate engine performance degradation. By using2.1 per cent loss in compressor efficiency, 5 per centreduction in airflow, 5 per cent reduction in pressureratio, and a 0.5 per cent reduction in gas generatorturbine efficiency, the authors reported an 8.6 percent reduction in power and an efficiency drop of3.5 per cent.

Zaba [11] investigated the effect of fouling on theperformance of an industrial gas turbine. He statedthat if all the compressor stages are equally anduniformly fouled, the percentage reduction in thecompressor volume flowrate is approximately equalto the percentage reduction in the compressor effi-ciency. The author also reports that fouling of the firststages has greater impact on the compressor volumeflowrate than fouling of the rear stages and the per-centage change in the volume flowrate is greater thanthe percentage change in the compressor efficiency.For the case he investigated experimentally, he foundthat for heavily fouled first stages the change in volu-metric flowrate is approximately equal to 2.5 times thepercentage change in compressor efficiency.

Meher-Homji and Bromley [12] ran performancesimulations for a 39.6 MW industrial gas turbine witha firing temperature of 1377.6 K. Imposing a 6 per centdeterioration in mass flow and 5 per cent deteriora-tion in the compressor efficiency, the authors foundthat the power output drops by 5.5 MW, hence by14.3 per cent.

Howell [13] developed a method to relate cascadetests to the performance of a compressor stage. This is

Proc. IMechE Vol. 224 Part A: J. Power and Energy JPE992

Experimental investigation of the influence of fouling on compressor cascade characteristics and implications 1009

used here to predict the performance deterioration ofa compressor in an engine and the associated changesin the whole engine performance, from changes incascade performance due to fouling.

2 EXPERIMENTAL FACILITY

The compressor cascade wind tunnel employed forthis investigation is a suction wind tunnel, with across-sectional area of 0.043 m2, and is designed foran air mass flowrate of approximately 5 kg/s, whichcan be varied via a throttling valve installed at the exit.This corresponds to an inlet Mach number of 0.3 anda Reynolds number of 3.8 × 105 formed with the inletvelocity and the chord length. This value is above thelower critical Reynolds number of about 105 [3], belowwhich laminar separation occurs.

The degree of turbulence and the mean velocityof the inlet air flow were measured with a hot wireanemometer as 2.25 per cent and 52.6 m/s, respec-tively. The test section of the cascade comprises nine2D blades of NACA 65 thickened profile with circulararc camber line. The blades are set at zero incidence;the blade span is 180 mm, the chord length 60 mm, andthe pitch-to-chord ratio s/c 0.8. The cascade bladesrepresent a stator blade mean section of a gas turbinecompressor with 50 per cent reaction.

The blades have a span-to-chord ratio (aspect ratio)of 3. The selection of the particular number of bladesand aspect ratio was made in order to minimize theinfluence of the cascade wall boundary layer effectson the test blade section.

In a compressor cascade tunnel, the increase instatic pressure across the blades causes wall boundarylayer thickening and contraction of the flow. Owingto the contraction of the flow the meridional veloc-ity increases. This contrasts with the diffusing actionof the compressor cascade itself. As a consequence,the sidewall boundary layer effectively reduces thestatic pressure increase that theoretically would beachieved. To reduce this effect, Dixon [14] and manyothers suggest that at least seven blades are needed forcompressor cascade testing and that each blade has aminimum aspect ratio of 3.

Figure 1 illustrates the cascade tunnel arrange-ment. A plenum chamber was installed behind thecascade test section. The ideal cascade experimentwould discharge the airflow from the blading intoan infinite space (e.g. the atmosphere). In the cur-rent project, however, the facility employed a suctionpump (fan). Accordingly, the blading was designed todischarge into the largest practical volume feasibleto ensure a uniform exit static pressure. In addi-tion, it is equally important that the effect of theplenum chamber shape itself and its exit ductingon the fan does not distort the cascade dischargeflow.

intake

cascade test section plenum chamber

centrifugal fan cascade tunnel exit cone

settling chamber

Fig. 1 Actual cascade tunnel left side view

3 ROUGHNESS AND MEASURING PARAMETERS

The three middle cascade blade profiles were rough-ened by covering the surfaces of the blades withvery thin double-sided sticky tape with a thicknessof 0.09 mm. Application of carborundum allowed theblade surface roughness distribution to be nearly uni-form. The carborundum grain sizes were defined withemery grade numbers which represent different sievesizes via which the grains pass through. The grit num-bers involved were 220, 180, 120, and 60 meaning grainsizes of 63, 76, 102, and 254 μm, which correspondto K /c values of 0.0010, 0.0013, 0.0017, and 0.0042,respectively. The roughness height K was assumed tobe equal to the average size (mean diameter) of thecarborundum grains.

Traverse pressure measurements one chord up-stream of the blades’ leading edges (see Fig. 2) weretaken with an L-shaped pitot static tube. The samepitot static probe was used to evaluate the mass flowby traversing along a cross-sectional plane perpendic-ular to the flow in the area of the settling chamber. Theplane was located at a distance of 50 per cent of thesettling chamber width upstream of the first cascadeblade, marked blade 1 in Fig. 2.

data collection line 1

perspex walls

blade 1 data collection line 2

SSPS

top

bottom

0 mm

120

-40

Fig. 2 Cascade test section



A three-hole cobra probe was used for downstreamtraverses at a distance of one chord from the bladetrailing edge. At this distance, most of the mixingwould have taken place and pitchwise flow angle vari-ation will be only small, according to Gostelow [15].The probe was nulled for the measurements and theprobe traversing took place at mid-span. The widthof the probe head is 2.4 mm, the height is 0.8 mm, andthe total length is 500 mm. The three holes of the probeare located in the same plane parallel to the probe axis,and their diameter is 0.5 mm. The head of the probe istrapezoidal with a characteristic wedge angle δ = 45◦.The pressure readings extracted from this yaw probewere combined in order to provide particular coeffi-cient values that were related to the probe calibrationchart so as to work out the flow exit angles of theblades. Velocity within the flow field was calculatedfrom the measurements of total and static pressure.

The blade profile total pressure loss coefficient

ω = P1 − P2

P1 − p1(1)

was obtained via analysis of the pressure data obtainedfrom the up and downstream traverses.

Flow fluctuations gave rise to a variation in mea-sured total pressure at any point in the flow of notmore than 100 Pa. This represents a variation of plusor minus 0.5 per cent. The total temperature was mea-sured in the test cell and not in the flow. In fact, the flowtotal temperature was assumed to be the test cell ambi-ent temperature. All data for pressure and temperaturewere non-dimensionalized.

4 EXPERIMENTAL RESULTS – SMOOTH BLADES

The velocity distributions one chord downstream ofthe cascade were investigated at mid-span for clean(smooth) blades. A statistical average of all the resultstaken at different ambient conditions is presented inFig. 3. It can be seen that the passage velocity increasesfrom 81 to 85 m/s towards the upper end of the cas-cade. This variation occurs because at the cascade exit

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pitchwise distance (mm)

Out

flow

vel

ocity

(m

/s)

Fig. 3 Statistical average velocity distribution at 60 mmtraverse

0.940

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p2/P

amb

Fig. 4 Normalized rear static pressure distribution at60 mm traverse

there is a pitchwise gradient in static pressure (seeFig. 4) of about 0.2 per cent. This causes the flow todeviate towards the lower static pressure region andprogressively increases the values of exit flow angleα2 at the exit of the cascade blades. This pressuregradient is due to the fact that the settling chamberis smaller than what would be required to allow thecascade to discharge to uniform static pressure. Thedimensions of the settling chamber were dictated bysize limitations of the test location.

When employing suction-driven cascade experi-ments, exit conditions in terms of static pressure arealways greatly influenced by the downstream dis-charge conditions. Given that the pressure gradientidentified above is of modest proportions this wasjudged not to constitute an insuperable obstacle tothe obtaining of useful data with the present rigconfiguration.

The measured wakes were typically around 10 mmwide with the velocity defect falling from 83 to 73.5 m/sin the wake.

Figure 5 illustrates a statistical average of pressureloss coefficient values calculated at different ambientconditions. The wake loss coefficient of the middleblade reaches a maximum value of 0.16 and the leftand right blade loss coefficients both reach a valueof around 0.18. The loss coefficient corresponding to

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Tot

al p

ress

ure

loss

coe

ffici

ent w

Fig. 5 Statistical average loss distribution at 60 mmtraverse



272829303132333435363738394041424344

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Exi

t flo

w a

ngle

(de

gree

s)

Fig. 6 Statistical average exit flow angle distribution at60 mm traverse

the three passages investigated is almost constant at avalue of 0.04.

It is important to note that all the cascade experi-ments have been undertaken at a nominal inlet Machnumber of 0.3. In fact, the range of Mach numbervaries from Mach 0.297 to Mach 0.31. Examples oftypical variation of losses with Mach number over arange of incidences reported [16, 17] show that forMach number less than 0.5 losses are independentof Mach number. For the test cases considered here,experimental examination of losses at Mach numbersless than 0.3 has not been undertaken. These valuesare somewhat lower than Mach 0.5, which is charac-teristically found in the mid-span of the first stage ofindustrial compressors. However, the fan capacity didnot allow the experiments to reach that number.

A statistical representation of the distribution ofblade exit flow angle, α2, is illustrated in Fig. 6. It can beobserved that the middle blade wake exit flow angle isapproximately 36◦ whereas in the passage the exit flowfalls to 34◦. In addition, there is a small (1◦) decrease inthe maximum value of the wake exit flow angle, fromone side of the cascade to the other.

5 EXPERIMENTAL RESULTS – ROUGHENEDBLADES

Roughness was applied uniformly on the three middlecascade blades. The particles used were carborundumof 63, 76, 102, and 254 μm average size stuck with verythin double-sided sticky tape on the blades (see Fig. 7).

Roughness is usually described by the arithmeticaverage value Ra, defined as the arithmetic mean ofthe absolute departures of the roughness profile fromthe mean line. This is the line fitted through the profilewhere the areas of the profile above and below this lineare equal [3].

However, this arithmetic mean value Ra is not suf-ficient in defining the hydrodynamic characteristic of

Clean blades Roughened blades

Blade4

Blade5

Blade6

Fig. 7 Illustration of roughened blades

the blade surface because a number of small rough-ness particles are fully submerged in the laminarsublayer and do not protrude or are not felt by theturbulent flow. Nevertheless, they must be taken intoaccount in terms of the arithmetic mean value. Schaf-fler [3] used an effective roughness height k whichdescribes the peaks rather than the average and isdefined as the difference between the arithmetic aver-ages of the ten highest peaks and the ten deepestgrooves existing per millimetre length. The measuringlength used by Schaffler was 5 mm.

Schaffler [3], after measuring the surface rough-ness of several blades, found a correlation betweenthe effective roughness height k and the arithmeticaverage height Ra (see equation (2))

k = 8.9 Ra (2)

For this project the parameter Rz was used as well,which is the average height difference between thefive highest peaks and the five lowest valleys withina sampling length of 0.8 mm [18].

A surface texture-measuring device (Taylor HobsonSutronic 25) was used involving a stylus following thesurface of the blades in the streamwise and spanwisedirections and taking roughness measurements at sta-tions of 25 per cent, 50 per cent, and 75 per cent of thechord and 25 per cent, 50 per cent, and 75 per centof the span. All the span and streamwise measure-ments were averaged, and the results are presentedin Tables 1 and 2 for upper and lower blade surfaces,respectively. The number of sampling lengths assessed

Table 1 Upper surface roughness

Upper surface spanwise Upper surface streamwiseroughness (μm) roughness (μm)

Particlesize (μm) Ra Rz k Ra Rz k

0 0.42 2.56 3.76 0.6 4.67 5.3463 6.73 37.22 59.93 7.2 40.11 64.4176 10.02 51.78 89.20 10.11 51.33 90.00102 14.44 73.44 128.55 14.62 73.67 130.14254 22.71 114.00 202.13 21.47 107.44 191.05



Table 2 Lower surface roughness

Lower surface spanwise Lower surface streamwiseroughness (μm) roughness (μm)

Particlesize (μm) Ra Rz k Ra Rz k

0 0.37 2.11 3.36 0.47 3.22 4.1563 7.4 41.33 65.86 7.38 41.11 65.6676 10.4 53.22 92.56 10.11 51.11 89.99102 14.53 74.22 129.35 14.33 72.33 127.57254 24.04 122.78 214.00 22.44 109.22 199.76

from the measuring device was 5, and a mean value forevery roughness parameter measured was calculatedwithin an assessment length l defined from the sum ofthe sampling lengths. This provides a better statisticalestimate of the parameter’s measured value.

Bons [7] collected a wide variety of proposed corre-lations which have been employed to convert mea-surable surface roughness parameters (Ra, Rz , orRq) to equivalent sand grain roughness ks (includingequation (2)). Bons noted that many of the correla-tions vary by up to a factor of 5. This shows that nosingle correlation appears to capture all of the rele-vant physics for both engineered and service-relatedroughness.

The effect of fouling on the pitchwise velocity dis-tribution at one chord downstream of the cascadeblades was investigated for clean and fouled blades.The corrected velocity, defined as the ratio of exit flowvelocity to the square root of the ambient tempera-ture (V2/

√Tamb), in the middle of the passage remains

almost constant at a value of 4.9 (Fig. 8). However, asthe fouling level increases this ratio undergoes signif-icant changes. Increasing the particle size to 63, 76,102, and 254 μm the velocity ratio in the wake dropsfrom 4.39 (clean blade) to 4.21, 4.14, 4.10, and 3.91,respectively. These values correspond to percentagefalls of 10.4 per cent, 14 per cent, 15.5 per cent, 16.3 percent, and 20.2 per cent from the passage value of 4.9.This indicates that the wake-corrected velocity behindthe roughened blades drops continuously though notlinearly.

In addition, the wake widens as the fouling levelincreases. However, the increase is not gradual. For

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Tam

b (m

/s /

K0.

5)

0 microns 63 microns 76 microns 102 microns 254 microns

PS SS

Fig. 8 Roughness wakes at 60 mm traverse

clean blades the wake is approximately 10 mm thick.However, for fouling levels of 63, 76, and 102 μm thewake thickness remains almost constant at 24 mm.Applying particles of 254 μm on the blades, the wakethickness increases considerably to about 39 mm.

As seen in Fig. 8 all the wakes move towards theright with respect to the 0 μm wake as the foulinglevel is increased. This happens due to the higherboundary layer growth of the suction surface of theblades compared to the pressure surface boundarylayer. The effect of the pitchwise static pressure incre-ment towards the upper end of the cascade (see Fig. 4)behind the blades is assumed not to be responsible forthis wake shift towards the upper end of the cascade.This is because the flow Mach number there is verylow. For such cases, probable incidence changes donot cause significant changes in the blade drag coeffi-cient, which is related to the thickness of the boundarylayer affecting the blade wake.

The pressure losses corresponding to the all threemiddle passages behind the blades were investigated.Figure 9 illustrates the pressure loss distribution onechord downstream of the blades towards the cascadeexit streamtube. Taking into account the middle pas-sage, as the fouling level applied on the blade surfaceincreases from 0 μm (smooth blades) to 63, 76, 102,and 254 μm the total pressure loss coefficient increasesfrom 0.16 to 0.25, 0.27, 0.285, and 0.35, respectively.This indicates that the wake total pressure loss coeffi-cient behind the roughened blades rises continuouslybut not linearly. The passage loss varies smoothlyaround the value of 0.04 and it seems that the foulingdoes not affect this area significantly.

It was noted that as the fouling level increases from0 to 76 μm, the middle passage exit flow angle (onechord downstream) corresponding to the nulling pointin the passage between blades 4 and 5 increases from34◦ to 37◦. Increasing the fouling level further to 102and 254 μm, the passage exit flow angle at the samepoint gets values of 38◦ and 39.5◦ (see Fig. 10). Hence,the passage exit flow angle seems to increase in alower proportion as the roughness increases more

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coe

ffici

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0 microns 63 microns 76 microns 102 microns 254 microns

Fig. 9 Roughness cases loss distribution at 60 mmtraverse



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particle size (microns)

Exi

t flo

w a

ngle

(de

gree

s)

Fig. 10 Blade 4–5 passage exit flow angle versus particlesize

than 102 μm. Up to the 102 μm the increase in exitflow angle is 4◦, and from 102 to 254 μm the increaseis reduced to 1.5◦ only. This can be attributed to theincrease of the passage blockage which progressivelyreduces the deviation and therefore the exit flow angle.Though the trend is unmistakable, the exact angularvariation is subject to the measurement accuracy esti-mated at ±1◦, as represented through the error bars inFig. 10.

6 BLADE THICKNESS EFFECT

The effect of blade thickness due to roughness has notbeen taken into account in the current study. Gbadeboet al. [1], in order to separate the effect of thicknessfrom the effect of roughness, performed tests by cov-ering the leading edge/peak suction region with thincardboard strips of similar thickness to that of theemery paper used for applying roughness. Comparingcontours of stage pressure rise coefficient for thick-ened blades with these of smooth and roughenedblades, they found that the thickness has negligi-ble contribution to wake thickening. According toSaravanamuttoo et al. [16], test results for subsoniccompressor blade sections shows that, at low Machnumbers, the losses for zero incidence are very lowcompared to those at higher incidence and higherMach number. The current cascade runs with an inletMach number of 0.3 and at a nominal incidence of0◦. The corresponding losses are expected to be lowand depend only secondarily on blade thickness. How-ever, at high Mach number and incidence far awayfrom nominal, as expected, the losses increase dramat-ically. In this case, blade thickening due to roughnesssignificantly increases the pressure losses.

The effect of adding thickness to the middle cas-cade blades by attaching the double-sided sticky tapewas investigated experimentally. Figure 11 shows thataddition of the double-sided sticky tape has almost no

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coe

ffici

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smooth blades - statistical average data blades with double sided sticky tape

Fig. 11 Effect of adding thickness on the total pressureloss coefficient

effect on total pressure loss coefficient, especially forthe two side cascade blades.

7 THEORETICAL BACKGROUND

Howell [13] developed a fluid dynamic theory whosebasis is knowledge of the 2D flow past cascade bladesand of the appropriate correction factors neededto provide the mean stage conditions in an actualcompressor.

In Howell’s work, the blade profile drag and liftcoefficients are obtained as follows

CDp = sc

(�p0

(1/2)ρV 21

)cos3 αm

cos2 α1(3)

CL = 2sc

cos αm(tan α1 − tan α2) (4)

where

tan αm = 12(tan α1 + tan α2) (5)

In order to proceed towards a representation of areal compressor stage, it was considered that the stagereaction R is 50 per cent.Then knowing the blade pitch,the blade height, and the blade lift coefficient values,the overall drag coefficient can be calculated

CD = CDp + CDa + CDs (6)

The drag coefficient, related to wall annulus fric-tion losses, is CDa, and the drag coefficient CDs, whichdescribes the secondary losses, particularly those thatgive rise to trailing edge vortices, can be calculated asfollows, employing Howell’s method

CDa = 0.02 s/H (7)

CDs = 0.018C 2L (8)

Annulus wall boundary layers give rise to spanwisedistributions of axial velocity, which tend to be more



peaky as the air passes through the compressor stages.Because of this, the axial velocity at the blade mid-span is higher than the design value and less work willbe generated.

Theoretically an increase in the work will take placeat the blade ends, but due to stalling at these areas noincrease in work occurs and the result is a reductionof the work done for the whole blade below the designvalue. The work done factor λ, which is the ratio ofactual rotor blade whirl velocity to the ideal changeof whirl velocity [19], applies a correction for all theselosses, and the mean value for multi-stage compressorused by Howell [13] was 0.86.

Howell and Bonham [20], in order to plot thestage characteristics curve using cascade testingdata, employed the stage temperature rise coeffi-cient cp�Ts/0.5U 2 (to use Howell’s notation), stageefficiency ηs (or polytropic efficiency ηp), and stagepressure rise coefficient �ps/0.5ρU 2. These quanti-ties had to be calculated for different incidences asfollows [21]

cp�Ts

(1/2)U 2= 2λ

(Va

U

)(tan α1 − tan α2) (9)

ηs = 1 − 2sin 2αm

CD

CL(10)

�ps

(1/2)ρU 2= ηs

(cp�Ts

(1/2)U 2

)(11)

The cascade tunnel results were analysed accordingto Howell’s theory [13] in order to relate the cascadewith an actual stage. The blade profile loss coefficientwas calculated and then the overall drag coefficientwas estimated. Calculating the lift coefficient and themean cascade flow angle αm, the polytropic efficiencyηp related to an actual stage was calculated. The actualstage polytropic efficiency calculated was then com-pared with the polytropic efficiency of an industrialgas turbine (165 MW power output). Running cascadeexperiments for different levels of fouling, differentvalues of actual stage polytropic efficiency were calcu-lated and for each case the percentage deteriorationin polytropic efficiency was calculated compared tosmooth blade cases. The fall in polytropic efficiencyobtained from the experiments was used as an inputto the gas turbine zero-dimensional performancesimulation tool, Turbomatch [22, 23], developed atCranfield University, for the examination of the per-formance of a large industrial engine. In this study, itis assumed that the percentage deterioration in poly-tropic efficiency of this engine equals the percentagedeterioration of the polytropic efficiency derived fromthe experimental cascade expressed as stage results.An assumption is made that the mid-radius reactionis 50 per cent. It is worth noting that isolated cascadetests in other than 50 per cent reaction stages can onlybe applied to one of the two blade rows of the stageat appropriate stagger and camber angles. The other

blade row in the stage of the compressor would needseparate cascade test data. Only in the case of a 50 percent reaction stage can the results of a single cascadetest be applied to both blade rows.

For the case of smooth blades, the values of stageflow coefficient Va/U and stage loading coefficient�H /U 2 were 0.524 and 0.247, respectively.

Another parameter used as an input to the Turbo-match simulation was the percentage deteriorationin the cascade passage non-dimensional mass flow�(W1T 0.5

1 /P1)p obtained from the experimental resultsdue to fouling. This was assumed to be equal to thepercentage reduction in non-dimensional mass flowof the industrial gas turbine engine examined.

8 EXPERIMENTAL PERFORMANCE SIMULATIONRESULTS

The polytropic efficiency was calculated for smoothand fouled blades of the current project compressorcascade tunnel, Fig. 12 showing that for fouling levelsof 63, 76, 102, and 254 μm, the percentage deteriora-tion in polytropic efficiency is 2.2 per cent, 3.9 per cent,4.9 per cent, and 7.7 per cent, respectively.

The conversion of the loss data from the cascadetests into a typical change in polytropic efficiencyof the real compressor stage takes into account thechange in outlet angle as given in Fig. 10.

Table 3 shows the percentage reduction of the cas-cade passage non-dimensional mass flow with respectto the fouling particle size applied on the blades.

0

1

2

3

4

5

6

7

8

0 20 40 60 80 100 120 140 160 180 200 220 240 260

Fouling particle diameter (microns)

Δηp

%

Fig. 12 Experimental percentage deterioration in poly-tropic efficiency due to fouling

Table 3 Cascade passage percentage reductionin non-dimensional mass flow

Particle size (μm) �(W1T 0.51 /P1)p%

0 0.00063 0.476 0.5102 0.7254 1.7



0.270

0.280

0.290

0.300

0.310

0.320

0.330

0.340

0.350

0.360

0.370

1100 1200 1300 1400 1500

TET (K)

η th

0 microns 63 microns 76 microns102 microns 254 microns

Fig. 13 Thermal efficiency versus TET

13.5

14.0

14.5

15.0

15.5

0 20 40 60 80 100 120 140 160 180 200 220 240 260

Fouling particle diameter (microns)

PR

TET=1100 K TET=1200 K TET=1300 KTET=1400 K TET=1500 K

Fig. 14 Pressure ratio reduction due to fouling

For this study, it was assumed that the compressoris fouled uniformly all the way through. Turbomatchperformance simulation runs were carried out whilevarying the TET between 1100 and 1500 K, combinedwith the effect of fouling. The thermal efficiency of theengine was examined first. The thermal efficiency ηth

increases as the TET increases (Fig. 13), and this canbe attributed to the parallel increase of the compres-sor pressure ratio (Fig. 14). Increasing the particle sizefrom 0 to 254 μm for constant TET levels the pressureratio was found to decrease by 1.3 per cent. However,keeping the fouling level constant and increasing theTET from 1100 to 1500 K the pressure ratio increasedby 8.3 per cent.

As the fouling increases, the pressure ratio drops andthe thermal efficiency decreases (i.e. referring to valuesat constant TET in Figs 14 and 13, respectively). Con-sidering the case of smooth blades (0 μm), the TETincreases from 1100 to 1500 K and the thermal effi-ciency increases by 15.5 per cent. Taking into accountthe case of the highest fouling of 254 μm for therange of the same TET increase, the thermal efficiency

increases by 25.5 per cent. Therefore as the fouling par-ticle size increases, the percentage thermal efficiencygain in the same level of TET increase gets higher.

As a result of the diverging constant pressure linesin the temperature–entropy diagram (Fig. 15), theUW progressively increases with increasing TET. Asa result of compressor fouling, the compressor effi-ciency reduces and the compressor work increases.As TET increases, the effect of increasing compressorwork with increasing fouling decreases.

Finally, the outcome is that the engine performancemeasured by the thermal efficiency is less sensitive tocomponent inefficiency as TET increases. This fact isillustrated by the reducing range of thermal efficiencychange with increasing TET shown in Fig. 13. In con-clusion, the engine performance deterioration due tofouling is highest at low TET.

From the Turbomatch results obtained it was shownthat as the fouling level increases the UW of the gas tur-bine decreases, but increasing the TET this drawbackcould be handled (see Fig. 16). Taking into account the

Fig. 15 Temperature versus entropy [24]

60708090

100110120130140150160170180190200

1100

TET (K)

UW

(M

Wat

ts)

0 microns 63 microns 76 microns102 microns 254 microns

1200 1300 1400 1500

Fig. 16 UW versus TET



case of 0 μm (smooth blades), incorporating a foulinglevel of 254 μm on the blades at the same TET of 1100 Kthe UW drops by 19.2 per cent and only by 9.8 per centfor the level of 1500 K. In order for the engine to recoverthe original UW at 0 μm after suffering from foulinglevel of 254 μm at the TET of 1100 K, it has to increaseits TET by almost 50 K. For the case of 102 μm pass-ing to 254 μm, the TET should increase less in order torecover the original useful clean engine output.

Also, for constant values of fouling level, the UW wasincreased linearly with respect to the TET increase. Thedecrease in UW as the fouling level (particle size) onthe blades increases is caused due to a reduction in theengine mass flow capacity.

Using predicted data obtained with the Turbomatchcode, Fig. 17 shows that for the industrial engine whosedesign point is at a TET of 1378 K, the rate of reductionof compressor efficiency with roughness particle sizeincrease is nearly independent of TET. Referring to theTET value of 1100 K, as the particle size increases from0 to 254 μm the compressor efficiency falls as statedearlier. Also, increasing the TET from 1100 to 1500 Kthe compressor efficiency ηc keeps falling by almost1.2 per cent for the smooth (0 μm) and all the foulingcases examined. This drop in efficiency is due to thefact that the engine non-dimensional mass flowratedecreases as the TET increases. This happens becausefor a compressor constant speed running line, as theTET increases the pressure ratio increases (movingfrom point A to point B), as Fig. 18 illustrates. How-ever, the non-dimensional mass flowrate of the enginefalls following the trend of the efficiency line passingfrom the design point shown in Fig. 19.

Keeping the TET constant and increasing the foulinglevel gradually, the performance drawbacks in terms ofpercentage compressor efficiency, thermal efficiency,and UW deterioration are illustrated in Table 4. Fromthe experimental cascade results correlated with thereal uniformly roughened fouled engine via Howell’stheory, it was found that by increasing the roughness

Fig. 17 Compressor efficiency versus TET

0

5

10

15

20

25

30

35

40

0.060 0.080 0.100 0.120 0.140 0.160 0.180 0.200

WeTe0.5/Pe

PR

N / Ndp =0.56 N / Ndp=0.6 N / Ndp=0.67N / Ndp=0.79 N / Ndp=0.9 N / Ndp=1point A, TET=1100 K point B, TET=1500 K

AB

Fig. 18 Industrial engine pressure ratio versus non-dimensional mass flowrate

0.700

0.750

0.800

0.850

0.900

0.950

1.000

1.050

1.100

0.060 0.080 0.100 0.120 0.140 0.160 0.180 0.200

WeTe0.5/Pe

η c

N / Ndp =0.56 N / Ndp=0.6 N / Ndp=0.67N / Ndp=0.79 N / Ndp=0.9 N / Ndp=1point A, TET=1100 K design point point B, TET=1500 K

AB

Fig. 19 Industrial engine compressor efficiency versusnon-dimensional mass flowrate

Table 4 Percentage efficiency and UW reductions due tofouling of industrial gas turbine

0 μm 63 μm 76 μm 102 μm 254 μm

�ηc (%) 0 2.2 3.9 4.9 7.7UW (1100 K) (MW) 88.2 84.2 80.4 78.3 71.3�UW (1100 K) (%) 0.0 4.6 8.9 11.3 19.2UW (1500 K) (MW) 196.5 191.5 187.4 185.0 177.3�UW (1500 K) (%) 0.0 2.5 4.6 5.8 9.8ηth (1100 K) 0.316 0.307 0.297 0.292 0.274�ηth (1100 K) (%) 0.0 2.9 5.9 7.6 13.2ηth (1500 K) 0.365 0.360 0.355 0.352 0.344�ηth (1500 K) (%) 0.0 1.4 2.7 3.4 5.6

up to a level of 254 μm the drawbacks in compressorefficiency can be as high as 7.7 per cent. Taking intoaccount these deterioration percentages in terms ofcompressor efficiency, increasing the fouling towardsthe level of 254 μm at 1100 K TET, the percentagedeterioration in the UW (�UW (1100 K) per cent)produced by the engine was found to be 19.2 percent. This percentage deterioration was eliminatedalmost by half, as the TET was increased to 1500 K.For the same TET increase and fouling particle sizeof 254 μm, the percentage deterioration in thermal



efficiency was eliminated by 57.6 per cent. For lowerlevels of fouling, and increasing the TET in the samerange, similar trends in terms of UW and thermal effi-ciency were found. Therefore, it can be stated thatone performance deterioration inhibitor when foul-ing presents, is the parameter TET which must beincreased properly.

9 CONCLUSIONS

The analysis of the cascade experiments showed thatas a result of increasing the fouling level, the width andthe depth of the wakes increased significantly. Moresignificantly it was observed in all fouling cases thatthe wakes shifted towards the blade suction side dueto the increased boundary layer growth on the suc-tion surfaces of the blades. Increasing the roughnessto 254 μm caused the total pressure loss coefficient(measured one chord downstream of the blades) todouble (i.e. from 0.16 to 0.35). Also, by increasing thelevel of fouling, the wake-corrected velocity defect andthe total pressure loss coefficient rise measured down-stream of the blades were observed to vary graduallythough not linearly.

From the experimental results, it can be observedthat in the middle of the blade passage, the increasein the exit flow angle when the fouling level was raisedfrom 0 to 102 μm was reduced by almost half whenfouling was further increased from 102 to 254 μm.This is due to the resulted added exit flow blockagerelated to the widening of the wakes as the roughnessincreases.

This article presents a method of establishing a rela-tionship between the compressor cascade tested anda real compressor stage. Howell’s theory of correlatingcascade data with compressor stages was employed.The performance simulation tool Turbomatch wasused to examine a real engine subject to the foul-ing comparable to the four different fouling casesexamined experimentally. The results showed that byincreasing the fouling level, the deterioration in bothpolytropic and overall efficiencies of the compressorincreases continuously. By increasing the fouling levelto 254 μm, the percentage deterioration in compres-sor efficiency reached a value of 7.7 per cent whencompared to the smooth blades.

The compressor efficiency falls when the foulinglevel on the blades increases due to the pressureratio degradation caused. However, for the compres-sor examined, by increasing the TET from 1100 to1500 K the efficiency falls by 1.2 per cent, no matterhow severely fouled the compressor may be. Again,in order for the engine to recover the loss in UW dueto the increasing compressor work caused by foul-ing, the TET must increase. Increasing the TET by 36.4per cent (from 1100 to 1500 K) and keeping the foul-ing level constant, the engine percentage performance

deterioration in terms of UW and thermal efficiencywas reduced by almost 50 per cent.

As the TET increases, the thermal efficiency of theengine increases as well, as a result of the increase incompressor pressure ratio. Also, for the same rangeof TET increase, the higher the level of fouling on theblades, the higher the gain in thermal efficiency, com-pared to clean blades. Keeping the TET constant andincreasing the fouling level, the pressure ratio dropsand the thermal efficiency decreases.

The UW increases linearly with the TET, and for con-stant TET, by increasing the fouling level, the powerreduces due to the reduction in the mass flow capac-ity of the engine. At lower TETs, the decrease inUW due to fouling is higher than that correspond-ing to higher TETs as a result of the extra marginin UW incorporated, as the diverging lines of thetemperature–entropy diagram show.

ACKNOWLEDGEMENTS

The authors thank Paul Lambart, Russell Gordon,Andy Lewis, and Jonathon O’Donnell from the R-MCPower Recovery Limited for their technical advice andsupport.

© Authors 2010

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APPENDIX

Notation

c blade chordcp specific heat at constant pressureCD drag coefficientCL lift coefficientH blade height, enthalpy

k roughness parameterks equivalent sand grain roughnessks,adm admissible sand roughnessK roughness height (particle size)l assessment lengthN engine rotational speedp static pressureP total pressurePamb ambient pressureR stage reactionRa arithmetic average roughnessRec Reynolds number based on true chord and

inlet conditionsRq root mean square roughnessRz ISO ten-point height roughness

parameters blade pitchT total temperatureTamb ambient temperatureU blade speedV velocityW mass flowrate

α air flow angleβ blade metal angleδ yaw probe wedge angle� change in valueη efficiencyλ work done factorρ density

Subscripts

0 stator outlet1 rotor inlet, cascade inlet2 rotor outlet, cascade outlet3 stator inlet4 stator outleta axial, annulusc compressordp design pointe engine compressor inletm meanp profile, peak, passage, and polytropics secondary, stageth thermal

Abbreviations

CW compressor powerPR pressure ratioPS pressure surfaceSS suction surfaceTW turbine power


experimental investigation of the influence of fouling on compressor cascade characteristics and...

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compressor blades

compressor fouling

surface roughness effects

applied surface roughness

effect of roughness

effects of blade roughness

roughness elements

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