A Novel High-Efficiency Impulse Turbine for Use inOscillating Water Column Devices

Shahab Natanzi1, Joao Amaral Teixeira

2, George Laird

1

1Dresser-Rand Company Limited85, Papyrus Road, Peterborough, United Kingdom

SNatanzi@dresser-rand.comGLaird@dresser-rand.com

2Cranfield University, Department of Power and PropulsionCranfield, Bedfordshire, United Kingdom

j.a.amaral.teixeira@cranfield.ac.uk

Abstract — This paper presents a novel impulse turbineconfiguration for use in OWC systems which was developed aspart of a UK Department of Trade and Industry part-fundedresearch programme carried out by Dresser-Rand Company Ltdand Cranfield University. The turbine demonstrates increasedlevels of efficiency in comparison with existing impulse designsacross a wide range of incident flows. This is achieved by adesign which does not contain any moving parts besides therotor, meaning that good reliability and economic cost ofproduction should be achievable. The results presented are basedon a validated and calibrated numerical approach and describesthat a total-static efficiency of about 75% for a full scale turbinehas been achieved. The numerical results have been validatedagainst the data obtained at unique test facility located atCranfield University which is capable of simulating theaerodynamic conditions that turbines are exposed to in OWCpower plants. Moreover, a unique numerical study of the effectsof inlet flow distortion on the performance of the impulse turbinefor use in OWC power plants is also presented.

Keywords — Wave Energy, Impulse Turbine, HydroAir,

Efficiency, Output Power

I. INTRODUCTION

Oscillating Water Column (OWC) devices have been thesubject of research for over thirty years and a small number ofthese devices have been operated, mostly in prototype or pre-commercial forms.

An OWC device comprises a structure forming an airchamber on the sea surface which encloses a volume of air.The incident waves make the free surface of the water insidethe chamber rise and fall. The rising water level forces the airthrough a turbine at the top of the structure. As the water levelfalls, the air from outside the chamber is sucked back into thechamber through the turbine. It should be noted that in mostOWC configurations the turbine rotates only in one directionirrespective of the direction of the air flow.

OWC systems comprise of a chain of energy conversions,wave to pneumatic (buoy), pneumatic to mechanical (air

turbine) and mechanical to electrical (generator). Of these thefirst two are the most problematic when the overall efficiencyof the conversion is taken into account. Accepting that thewave to pneumatic conversion efficiency can be reasonablyhigh for a judiciously sized and appropriately designedcollector buoy, knowing that the mechanical to electricalconversion efficiency for electrical systems can be over 80%,it leaves the turbine as the single most critical element in theefficiency chain of an OWC system. Figure 1 shows aschematic view of an OWC power plant. In practice, thecollector chamber can be either a land based structure or atethered buoy.

Figure 1 - Oscillating Water Column (OWC) wave energy converter

Over the last thirty years, there have been a large numberof research projects addressing turbine design for use in OWCsystems. However, mainly due to the low efficiency of theOWC turbines, the reported overall wave-to-wire efficienciesare often low, and hence such plants are still not economicallyfeasible. This is because, the design of a turbine for use in thereciprocating flow conditions, experienced in theseapplications, is complicated due to the alternating direction ofthe flow. Therefore, a self-rectifying turbine that can provide aunidirectional rotation, regardless of direction of flow, is asolution for these systems. Additional difficulties arise from

Wave Propagation

Water

Air

Turbine

Generator

Airflow

OWC

the fact that the turbine has to operate over a wide range offlow conditions associated with a variety of sea states; to saynothing of the fact that waves are composed by trains oflargely irregular water fluctuations.

Over the last 30 years, many different turbine designs havebeen examined some of which are listed in Table 1.

TABLE 1Turbines for Wave Power Generation

Turbine Type Sub-type

Wells Turbine Without guide vanesWith guide vanesSelf pitch controlled bladesLeading edge slatBiplane without guide vanesBiplane with outer guide vanesBiplane with outer and innerguide vanesContra-rotating

Impulse turbines Impulse turbine with self pitchcontrolled guide vanesImpulse turbine with fixed guidevanesMcCormick counter-rotatingturbine

Dennis Auld turbineRadial turbinesCross flow turbineSavonius Turbine

In 2006 Setoguchi and Takao [3] reviewed the status of theturbines listed in Table 1 on the basis of:

- The starting characteristic- The running characteristic i.e. the efficiency versus flow-ratecharacteristic- Peak efficiency, and stall margin

The authors presented the turbine peak efficiency valuesunder steady flow conditions (Table 2).

TABLE 2PEAK EFFICIENCY OF VARIOUS AIR TURBINES FOR WAVE ENERGY [3]

TurbineType

Wellsturbine

withguidevanes

Wellsturbine

with self-pitch-

controlledblades

BiplaneWells

turbinewithguidevanes

Impulseturbine

with self-pitch-

controlledguidevanes

Impulseturbine

withfixedguidevanes

PeakTurbine

Efficiency0.492 0.496 0.534 0.564 0.390

The same authors also evaluated the conversion efficiencyof wave energy systems for various turbines under irregularflow conditions. This was carried out because the authorsbelieved that the characteristics of the turbines under steadyflow conditions do not provide sufficient information aboutthe suitable turbine for wave power conversion. This lack of

information exists because the performance of the wave powerconverter also depends on the OWC’s energy absorptionefficiency, which is closely related to the pressure differenceacross the turbine. Therefore, they presented the energyconversion efficiency of the system which is obtained bymultiplying the efficiency of the air chamber and the turbineefficiency. This is presented in Table 3.

TABLE 3PEAK CONVERSION EFFICIENCY OF WAVE ENERGY SYSTEM UNDER

IRREGULAR FLOW CONDITIONS [3]

TurbineType

Wellsturbine

withguidevanes

Wellsturbine

with self-pitch-

controlledblades

BiplaneWells

turbinewithguidevanes

Impulseturbine

with self-pitch-

controlledguidevanes

Impulseturbine

withfixedguidevanes

PeakConversionEfficiency

0.30 0.44 0.34 0.47 0.37

They concluded that the impulse type turbines have thepotential to be superior to the Wells type turbines in overallperformance under irregular flow conditions; to say nothing ofthe fact that in comparison, the starting characteristic of theimpulse turbine with self pitching guide vanes was superior;that the lower rotational speed of an impulse turbine wasdesirable to reduce noise and ease mechanical design; and thatthe impulse turbine maintain its efficiency over a wider rangeof flow rates.

Most of the OWC plants operated to date have utilisedWells turbines, except for the Port Kembla OWC which isusing a variable pitch Denniss-Auld turbine. The outputs ofthe Wells turbine OWCs have been around 400-500 kW. Interms of overall wave-to-wire efficiency under real seaconditions, their performance have been found to be relativelylow [4].

Figure 2 - Schematic of fixed guide vane impulse turbine showing [19]

The main problematic design element of an impulseturbine utilised in bidirectional flows, is the requirement tohave guide vanes both upstream and downstream of the rotorblade as shown in Figure 2. Since the function of the vanes isto direct the inlet flow onto the rotor with the correctorientation, and given that the outflow of impulse turbine rotoris broadly axial, it is inevitable that the angle of incidenceonto the downstream guide vanes is rather poor. This resultsin large areas of separated flow around the downstream guidevanes and correspondingly large losses. This is demonstratedgraphically in Figure 3 which shows a velocity vector plotaround the impulse turbine blades from a CFD calculation.Although changing the shape of the guide vanes are proved tobe effective, but similar separated flow has been observed forvarious studies with different guide vane profiles [5].

Figure 3 - Velocity vector plot around impulse turbine blades

The observed separation increase the losses through theturbine and therefore decreases the turbine efficiencysignificantly. One solution to this problem is to have variablegeometry guide vanes that are able to provide the desired flowturning when they are upstream of the rotor, and adopt a lowloss configuration when they are downstream. The experiencewith this approach in the field, however, has been that thereare significant reliability and maintainability issues associatedwith this choice [3].

II. VARYING RADIUS TURBINE (VRT)

The Dresser-Rand/Cranfield University HydroAir project,which the UK DTi co-sponsored, commenced by an extensiveliterature review undertaken by Herring [7], and numericalstudies of the baseline Wells and impulse turbines performedby Banks [19]. After that a decision was taken to focus onimpulse turbines for further development. The rationalebehind this decision was that it seemed more feasible toimprove the peak efficiency of the impulse turbine andmaintain the wide operating range (at high efficiencies) thanto broaden the efficient operating range of the Wells turbine.Although various methods exist to delay the onset of stall foraerofoils used in aircraft, none of these seemed to offer thepromise of sufficiently wide operating range to make theWells turbine an ideal solution for use in real OWC flows.Initially the key approach that seemed likely to give highefficiencies over a wide range of incident flows was toimplement the variable pitch mechanism for the Wells turbine

blades [6]. However, this would require a complex mechanismto actuate the blades as part of a rotating assembly and it wasfelt that it would be significantly difficult to make such asystem reliable in a real marine environment.

At this stage, extensive Computational Fluid Dynamic(CFD) analysis was performed on Wells turbine, mainly byBanks [19]. Comparing the results with the availableexperimental data in the literature, some confidence had beengained in the ability of CFD to be used as a tool in the designof impulse turbines. Therefore, it was decided to attempt toproduce a baseline impulse turbine geometry with equal orbetter performance than those reported in the literature overwhich improvements could then be sought.

In the course of developing the baseline impulse turbine, itbecame clear that it was likely to prove difficult to achieve asubstantial increase in efficiency if the flow separation on thedownstream guide vane could not be removed or at leastsubstantially decreased. Since there was no obvious way toremove the separation around the downstream guide vanewithout recourse to variable geometry or boundary layerblowing (neither of which is desirable in a marineenvironment due to the associated reliability concerns), it wasrealised that decreasing the flow velocity in this region wouldsubstantially reduce the pressure drop, even if the separationwas still present. This can be demonstrated by adopting theapproach of assigning loss coefficients to the guide vanes.According to the classical turbine theory, the stagnationpressure loss coefficient equation in incompressible flow canbe calculated for the guide vanes using the equations below inwhich P is the total pressure, V is the inlet flow velocity to theguide vanes, and Y represents the stagnation pressure losscoefficient.

2

0201

5.0 V

PPYGV

This equation describes that for a particular guide vaneprofile, which has a constant pressure loss coefficient (

GVY ),

there is a correlation between the pressure drop across the

component and the inlet velocity ( 2VP ). This shows thatthe total pressure drop across the guide vanes is related to thesecond power of inlet velocity. In other words, if the velocityof the flow through the guide vane row is reduced, the resultwould be a greater reduction in pressure drop which results inhaving a more efficient guide vane row. An obvious way todecrease the velocity is to increase the cross sectional area ofthe flow path for a given volumetric flow rate. This could beachieved with a moderate flaring of the annulus such that thespan of the guide vanes would be larger than that of the rotorblades. Simply increasing the height of the annulus at theguide vanes however, would lead to problems with the flowangle at the inlet to the rotor, because the axial velocity wouldincrease, in the convergent section of the annulus,disproportionately to the tangential velocity.

The solution to achieving both a velocity reduction whilekeeping control of the variation of the velocity trianglesinvolved moving the guide vanes outwards towards a largerradius by reference to their original position. Thisarrangement was patented and designated the varying radiusturbine (VRT) and a schematic view of it is shown in Figure4.

Figure 4 - Schematic of the varying radius turbine layout [19]

There were two approaches here that could be taken; firstlykeeping the height of the blades (guide vanes and rotors)constant, or calculating the effect that changing height of theblades can have on the flow angles [20]. In this section, tosimplify the explanation, the height of the blades is keptconstant, although the results presented in the next section arefor uneven blade heights.

By maintaining the height of the guide vanes, whilemoving them outward from the original annulus to a greaterdiameter, so that the guide vanes are placed at a greater radiusthan the rotor, the axial and tangential components would bothbe accelerated as the fluid proceeded towards the rotor,because angular momentum must also be conserved. Thiswould maintain a good angle of incidence when acting as anIGV. After exiting the rotor, the fluid would then bedecelerated both axially and tangentially meaning thatalthough the angle of incidence onto the downstream guidevane would remain poor, the velocities and hence pressuredrop would be much smaller.

Due to the novelty of the arrangement, there was no guidancein the literature as to what an appropriate offset and ductgeometry might be. Initially, it was calculated that to give apredicted increase in efficiency to 60 % from the 40 % of thebaseline design, the required duct offset was 3.6 blade heights.This analysis was based on the fact that in the absence of anyother effects (e.g. increased wetted area losses), the pressuredrop (and hence the efficiency) should obey the simplerelation [7]:

Pr

rP

2

where P is the pressure drop with a constant mean radius ofr and P is the pressure drop that would be expected from a

VRT arrangement with the guide vanes located at a meanradius of r .

A large number of CFD simulations carried out indicatedthat the concept was effective in increasing performance withthe initial case giving a predicted efficiency of 58.9 % at thedesign point. This suggested that the simplistic analysis ofHerring lead to reasonably accurate predictions of the benefitthat might be expected for modest guide vane offsets.

III. EXPERIMENTAL PROCEDURE

As part of the Dresser-Rand and DTi sponsored HydroAirproject, a reciprocating flow test facility was designed andinstalled at the Gas Turbine Laboratories at CranfieldUniversity for testing bidirectional turbines intended for use inOWCs. A full description of the design, construction andoperation of this test rig can be found in [7] and [8]; however,an outline is presented here.

The rig consists of a circular chamber containing a flap thatcan be controlled by means of a computer to deliver anoscillating airflow to a duct in which test turbines can belocated. A schematic of this arrangement is shown in Figure 5.The fixed baffle means that by rotating the flap backwards andforwards, the air on the turbine side of the baffle is alternatelycompressed and rarified and an oscillating flow is set upthrough the turbine located in the test section. Both thefrequency and the amplitude of the flap movements can beaccurately controlled to allow the simulation of airflowscaused by waves of different heights and periods. Althoughthe results that will be presented in this work are for constantfrequency and amplitude sinusoidal flows, the rig is alsocapable of producing time varying irregular pressurefluctuations that are more representative of the flowconditions seen in a real OWC. A CAD model of the test rig isshown in Figure 6.

Figure 5 - Schematic of the oscillating flow bidirectional turbine

Figure 6 - CAD model of the test rig

The rig and the VRT turbine were designed in tandem suchthat the turbine could be tested in a range of flows andmaintain adequate Reynolds numbers for scaling

purposes( 5102 ). Due to the low velocities of the

flow 150 ms , there were not expected to be any scaling issuesassociated with Mach number effects and so these were notconsidered [7]. In order to capture any possible frequencyeffects, Herring concluded that it is necessary to attainStrouhal number similarity in order to be able to scale theresults to full size. Strouhal number is defined as:

av

fDSt

where f is the frequency of the flow oscillations, D is a

characteristic length, in this case taken to be the turbinediameter, and

av is the mean axial flow velocity. The turbine

was designed to be a one quarter scale of a turbine that wouldbe suitable for an OWC similar to LIMPET. Full-scale wavesin locations suitable for the installation of OWC devicestypically have a period of around 10s. This, in tandem with a

design axial flow velocity of 25 1ms for the turbine, meantthat to maintain Strouhal similarity, the test rig would have tobe capable of producing oscillations of period T equal to

ScaleFullT _314.0 . The thesis of Herring shows that this is

possible with the rig that was built, but only for a relativelysmall range of flows.

TABLE 4FLOW CAPABILITIES OF THE OSCILLATING TEST RIG [7]

Parameter Value Units

Total swept volume 16.43 3mMax flow rate 5.88 13 smMax RMS flow rate 4.15 13 smMax frequency 0.14 HzMin period for full oscillation 7 s

Max frequency for small oscillations 1 Hz

The maximum flows that test rig was capable of producingare summarised in Table 4. The flow rate could be calculatedby measuring the flap movements by means of a positionencoder located on the motor shaft, or by static and totalpressure sensors positioned on the guide vane leading edges.Both static and total pressures could be measured at thesepositions and thermocouple measurements were also made fortemperature correction of the pressures.

The turbine speed and mechanical torque were measuredusing a torque transducer located on the shaft between theturbine and electrical generator and the maximum turbinespeed was limited by the rating of the electrical generatorwhich was 1500 rpm. With measurements of flow rate,pressure drop, turbine rotational speed and mechanical torque,the efficiency of a test turbine could be determined bothinstantaneously (subject to the data acquisition limits) and onaverage.

IV. NUMERICAL PROCEDURE

The ANSYS TurboGrid package was used to generate thecomputational meshes for the guide vanes and rotor blade. Allmeshes were body fitted hexahedral meshes with O-gridsaround the blades and H/J/L-type topology. The tip region ofthe rotor blade was meshed with a non-matching H-grid. Firstcell heights from the blade surfaces were set to give anestimated y+ < 2 at maximum Reynolds number. In all cases,the maximum mesh angles were > 25°. Three different meshdensities were prepared for these cases; coarse (<500,000nodes), medium (<1,000,000 nodes) and fine (>1,000,000nodes).

The flow in the stator domains were calculated in thestationary frame while the rotor blade flow was solved in therotating frame. Stage type interfaces were used to join thedomains. These interfaces are an option within CFX packagewhich allows for both pitch and frame change with all flowquantities being circumferentially averaged and appropriatelyscaled. Such interfaces incur a one time mixing loss (due tothe averaging) but are fully conservative. Rotationallyperiodic boundaries were set along the sides of all domainsand all hub, shroud, and blade surfaces were set as non-slipwalls, with the shroud casing of the rotor domain being set asa counter rotating wall. A generalised grid interface (GGI)was used to join the two sides of the non-matching H-gridused in the rotor tip gap. A total pressure was specified at theinlet to the upstream guide vane and an average static pressure(atmospheric) imposed at the outlet of the downstream guidevane. The values of mass flow were therefore calculated basedon the pressure variation across the passage.

Figure 7 shows a typical comparison between theexperimental and the numerical data, obtained for one of theearlier versions of the VRT, and describes that a sufficient

compatibility can be expected between numerical andexperimental results [19].

Figure 7 - Comparison of experimental and numerical results [19]

V. RESULTS AND DISCUSSION

After verifying the effectiveness of the VRT design,several optimisation processes were carried out on thisconcept which led to introduction of other generations of VRTover the last three years. The results presented herecorrespond to a full scale design denoted as VRT-2 and wereobtained through a calibrated and validated CFD approachestablished at Dresser-Rand Company. It ought to bementioned that a prototype scale of this turbine wasmanufactured at Dresser-Rand's Peterborough facility and wasinstalled on an off-shore OWC platform in Australia during2010. A schematic view of the full scale VRT-2 is shown inFigure 8 while a manufactured prototype is shown in Figure 9.

Figure 8 - Full scale VRT-2 configuration

Figure 9 - Manufactured VRT-2 prototype

Although the VRT-2 make use of a new guide vane profile,new 3D rotor profile, as well a new duct concept, the wholesystem is still based on the same approach outlined in theprevious section. Figure 10 shows the predicted efficiency ofthe VRT-2 with a 2 meter diameter rotor as obtained fromnumerical analysis. The present status of the design representsthe outcome of a series of optimisations studies on variousdesign parameters such as hub-tip ratio, guide vanes turningangles, guide vane twist, tip clearance effects, and axial andradial offset of the guide vanes in comparison to the rotorposition.

As can be seen from Figure 10, the present VRT-2 designcan achieve a peak efficiency of over 75 percent. This can beconsidered as a step change in efficiency of the impulseturbine with fixed guide vanes for use in OWC power plants.

0

0.25

0.5

0.75

1

0 0.5 1 1.5 2 2.5 3 3.5

Flow Coefficient

Eff

icie

ncy

Figure 10 - Full scale VRT-2 Characteristic (numerical results)

Figure 11 - Velocity vectors on downstream guide vane (a) Flow-Coefficient= 0.97 (b) Flow-Coefficient = 1.57

The new design approach was also beneficial to reduce(and in some cases to completely remove) the separation onthe downstream guide vanes. This is thought to be the mainfactor which allows the VRT-2 to achieve these peakefficiency values. Figure 11 shows the velocity vectors on thedownstream guide vanes for two flow coefficients. Figure 11(b) shows the velocity vectors on a downstream guide vane ofthe VRT-2 operating at a flow coefficient of 1.57. As can beseen, the majority of the separation on the downstream guidevane is recovered compared to what was observed in Figure11 (a) which is the same turbine operating at a flowcoefficient of 0.97. It should be highlighted that both cases(Figure 11 a and b) achieved similar efficiency values. Thiscan be explained by that the operating under higher flowcoefficients means the flow velocity is increased inside theturbine and therefore greater losses are associated with theupstream turbine elements when compared to the turbineoperating under a lower flow coefficient. This approach canbe used to reduce the noise levels from the turbine sinceboundary layer separation can be attributed to the generatednoise of the turbo-machines.

VI. ENVIRONMENTAL EFFECTS - INLET FLOW CONDITIONS

The performance of turbo-machines is dependent onvarious environmental factors, amongst which inlet flowconditions play a significant role. The importance of studyingthe inlet flow conditions is magnified in the OWC turbineinvestigation process due to the nature of the environment inwhich OWC plants operate. There are a number of publishedpapers which have investigated the effects of inlet flowdistortion on various types of turbo-machines such as gas

turbines [9], [10], [11] and compressors [12], [13]. However,due to the flow conditions in which these turbo-machinesusually operate (i.e. high flow velocities), these results are notdirectly relevant to the OWC turbine studies.

In 1987, Raghunathan et al. [14] experimentallyinvestigated the effects of inlet flow distortion and high levelsof turbulence on the performance of a Wells turbine withguide vanes, using an average Reynolds number of around2.6× 510 . The authors concluded that a large change in theturbulence level (from 0.4% to 6.7%) increases the turbinetotal-to-total efficiency by 2%. However, the stall point wasobserved to be postponed from a flow coefficient of 0.24 to0.3. They also concluded that inlet flow distortion leads to avery small performance loss in the Wells turbine.

There is no literature available on the effect of inlet flowdistortion on the performance of the impulse turbine for use inOWC power plants. Therefore, this section focuses onevaluating the performance of the designed impulse turbine(VRT-2), in presence of inlet flow distortion, in order to fillthe present gap in this knowledge. An investigation of thepredicted turbine performance in the operational environmenthas been performed. The inlet flow distortion in the presentedmodel was caused by introducing a moderate wind flow (10m/s wind velocity) to the system.

A numerical model was set up to include a complete OWCchamber, a full scale VRT-2, and a large domain set torepresent the atmosphere. The atmospheric domain surroundsthe whole system. Figure 12 displays a schematic view of thenumerical model.

Figure 12 - A whole OWC power plant model

(a)

(b)

VRT-2

Atmospheric Domain

OWCChamber

Rotor sub-domain

Oscillatingsea surface

Turbineinner duct

Turbine

outer duct

Since the inclusion of all the turbine blading in the modelinvolved some very large grid counts, ~100million nodes, andcorrespondingly large computational resources and runningtimes, a decision was taken to dispense with the detailedrepresentation of the blades. Instead it was the effects of theblades on the flow that were modeled. This involved thenumerical representation, through source terms, of the turningexerted by the blades together with the momentum losses. Theappropriate quantities were obtained from previous VRT-2stage simulations. Figure 13 shows the added turning of theflow passing through the guide vane sub-domain. This guidevane turning shown in Figure 13 was achieved by imposingdirectional loss on the desired flow direction using an extrapolar coordinate system. The pressure drop imposed on therotor domain was related to the mass flow rate by using aquadratic equation.

Figure 13 - Imposed numerical turning on guide vane sub-domain and andpressure drop on the rotor sub-domain

The simulations were performed using 32 processors out ofwhich 16 processors had 16GB memory and the rest had 8GBmemory. The solution’s time-step was set to be 0.01 secondand the simulation was run for 1500, 2000, and 3000 time-steps each with a maximum of 3 and 5 coefficient loops and aresidual target of 0.0001. It was shown that the results wereclosely comparable for all the cases and every cycle. Hence,for the purposes of clarity, the results presented below areonly for one full cycle. Figure 14 illustrates the velocitycontours on a X-Z plane. This is a meridional plane alignedlengthways. Four cases are shown corresponding to maximumexhalation and inhalation flows, with and without wind.

Figure 14 - Velocity contours during maximum exhalation and inhalation

Figure 15 shows the ΔP and mass flow rate across the turbine for one full cycle and for both cases with and withoutwind. The ΔP shown in this figure is the difference in pressure measured between the inside of the chamber and the outsideof the turbine outlet which is atmospheric pressure. Therefore,a positive ΔP corresponds to the exhalation period when the pressure inside the chamber is higher than atmosphericpressure. A negative ΔP is indicative of the inhalation cycle, when the pressure inside the chamber is lower than theatmospheric pressure and therefore the flow is being suckedinto the device. Although as expected the maximum andminimum flow rates correspond to maximum and minimumΔP respectively there is interestingly a small lag between the two peaks.

Moreover, comparing Figures 15 (a) and (b), the resultsshow that in the presence of wind, the mass flow rateincreases during inhalation and decreases during exhalationwhen compared to the case without wind. This can beexplained by the fact that during the inhalation, wind willfreely flow into the turbine which results in having around 7%more flow during inhalation; while during exhalation the windblocks the turbine’s outlet resulting in a reduction of some12% less mass flow. This is expected to correspond to a slightincrease in efficiency during inhalation and about a 3%decrease in efficiency during exhalation. The results alsoshowed that the inlet flow distortion causes a certain amountof non-uniformity of the flow entering the rotor blades duringinhalation period. This can on its own cause further disruptionto the performance of the turbine.

-6000

-4000

-2000

0

2000

4000

6000

0 1 2 3 4 5 6 7 8 9 10 11 12 13 14ΔP

(P

a)

-300

-200

-100

0

100

200

300

Mass

flo

wra

te(k

g/s

)

-6000

-4000

-2000

0

2000

4000

6000

0 1 2 3 4 5 6 7 8 9 10 11 12 13 14

Time (s)

ΔP

(P

a)

-300

-200

-100

0

100

200

300

Mass

flo

wra

te(k

g/s

)

ΔP (Pa) Mass flow rate (kg/s)

Figure 15 - ΔP and mass flow rate across the turbine in a full cycle (a) with wind (b) without wind

There are various other data that can be extracted from thiskind of simulations and hysteresis data can be one of them. Asregards to wave energy devices, it is usually presumed that the

10m/s wind

10m/s windNo wind

No wind

With 10 m/s wind

Without wind

hysteresis effect is minor due to a very low air-flowfrequency. The hysteresis effect may have various causes suchas asymmetry in the boundary layer on the blade surfaces.There are a number of studies on this subject mainlyperformed by Raghunathan and Setoguchi et al [15], [16],[17], [18]. However, all of them have studied the hysteresiseffects in a small scale model and mostly on the turbinesolely, not as a whole system as performed here.

-6000

-5000

-4000

-3000

-2000

-1000

0

1000

2000

3000

4000

5000

6000

-250 -225 -200 -175 -150 -125 -100 -75 -50 -25 0 25 50 75 100 125 150 175 200 225 250

Mass flow rate (kg/s)

ΔP

(P

a)

Without wind With wind

Figure 16 - Flow rate against pressure drop across the turbine

Figure 16 shows the hysteresis loop on the OWC system inthe two cases, with and without wind effects. In other words,it shows the differential pressure across the turbine against themass flow rate passing through the turbine. This indicates thatthe relationship between the mass flow rate and the pressuredifferential is not unique, but depends on whether the air isforced out of the turbine (exhalation) or sucked in (inhalation).It also explains that the magnitudes of hysteresis for bothcases (with and without wind) are very similar.

VII. CONCLUSIONS

The results presented in this paper are a selection from thedata acquired over a 5 year research program. From these anumber of conclusions based on this research program can bedrawn:

• A new turbine design (VRT-2) was generated whichintroduces a considerable step up in efficiency of theimpulse turbine with fixed guide vanes for use inoscillating water column power plants. Various turbinedesign elements were investigated and optimised. Thisincludes hub-tip ratio, guide vanes twist, tip clearance,blade profiles, and axial and radial offset of the duct. Asa result, a peak efficiency of over 75% was observed forthis design.

• The negative effect of the flow separation on the outletguide vane was minimised and in some cases wascompletely removed, using the VRT approach as well asoptimised elements such as new guide vane profiles andduct shape.

• A novel transient OWC power plant model was created.Based on this model, it was observed that the turbinemass flow and efficiency is likely to be reduced when itis exposed to windy conditions. Also, the transientsimulations helped to increase the understanding of thehysteresis effects on the OWC systems.

ACKNOWLEDGEMENT

The authors would like to acknowledge the financialassistance provided by Dresser-Rand Company Ltd and theUK Department of Trade and Industry (now BERR). Theworks of Professor Christopher Freeman on the design of theturbine, Dr. Kevin Banks in performing various optimisationanalyses, and Dr. Steven Herring in carrying out the physicalexperiments reported in this paper are also gratefullyacknowledged.

REFERENCES

[1] R. Shaw. Wave Energy a Design Challenge. Ellis Horwood Ltd, 1982.

[2] A. Sarmento, F. Neumann, and V. La Regina. The need and availablemeans for international cooperation in marine renewable technology. InWorld Renewable Energy Congress, 2005.

[3] T. Setoguchi, and M. Takao,. Current status of Self rectifying air turbinesfor wave energy conversion. Energy Conversion and Management, vol. 47,no. 15-16, pp. 2382-2396.

[4] I. Webb, C. Seaman, and G. Jackson. Oscillating water column waveenergy converter evaluation report. Technical report, Carbon Trust, 2005.

[5] T. Setoguchi, S. Santhakumar, H. Maeda, M. Takao, and K. Kaneko. Areview of impulse turbines for wave energy conversion. Renewable Energy,23:261–292, 2001.

[6] S. Herring. OWC Performance Matrix and Design Review (Unpublishedreport). 2005.

[7] S. Herring. Design and evaluation of turbines for use in OWC powerplants. PhD thesis, School of Engineering, Cranfield University, 2007.

[8] S. Herring and G. Laird. A new test facility for evaluating turbines for usein OWC power plants. In Proceedings of the 7th European Wave and TidalEnergy Conference, Porto, Portugal, 2007.

[9] M. Davis, and A. Hale, (2007), "A parametric study on the effects of inletswirl on compression system performance and operability using numericalsimulations", Vol. 1, pp. 1.

[10] M. Davis, A. Hale, and D. Beale, (2002), "An argument for enhancementof the current inlet distortion ground test practice for aircraft gas turbineengines", Journal of Turbomachinery, vol. 124, no. 2, pp. 235-241.

[11] Anon (1987), "Engine response to distorted inflow conditions (Paperpresented at the Propulsion and Energetics 68TH Annual SpecialistsMeeting)".

[12] V. G. Filipenco, S. Deniz, J. M. Johnston, E. M. Greitzer, and N. A.Cumpsty, (2000), "Effects of inlet flow field conditions on the performance ofcentrifugal compressor diffusers: Part 1-discrete-passage diffuser", Journal ofTurbomachinery, vol. 122, no. 1, pp. 1-10.

[13] P. Bry, P. Laval, and G. Billet, (1985), "Distorted Flow Field inCompressor Inlet Channels.", Journal of Engineering for Gas Turbines andPower, vol. 107, no. 3, pp. 782-791.

[14] S. Raghunathan, T. Setoguchi, and K. Kaneko, (1987), "Wells air turbinesubjected to inlet flow distortion and high levels of turbulence.", InternationalJournal of Heat and Fluid Flow, vol. 8, no. 2, pp. 165-167.

Exhalation

Inhalation

[15] T. Setoguchi, Y. Kinoue, T. H. Kim, K. Kaneko, and M. Inoue, (2003),"Hysteretic characteristics of Wells turbine for wave power conversion",Renewable Energy, vol. 28, no. 13, pp. 2113-2127.

[16] S. Raghunathan, (1995), "The wells air turbine for wave energyconversion", Progress in Aerospace Sciences, vol. 31, no. 4, pp. 335-386.

[17] M. Mamun, Y. Kinoue, T. Setoguchi, T. H. Kim, K. Kaneko, and M.Inoue, (2004), "Hysteretic flow characteristics of biplane Wells turbine",Ocean Engineering, vol. 31, no. 11-12, pp. 1423-1435.

[18] Y. Kinoue, T. H. Kim, T. Setoguchi, M. Mohammad, K. Kaneko, and M.Inoue, (2004), "Hysteretic characteristics of monoplane and biplane Wellsturbine for wave power conversion", Energy Conversion and Management,vol. 45, no. 9-10, pp. 1617-1629.

[19] K. Banks. Optimisation of bidirectional impulse turbines for wave powergeneration. PhD thesis, School of Engineering, Cranfield University, 2009.

[20] S. Natanzi. Design and investigation of turbines for use in oscillatingwater column power plants. PhD thesis, School of Engineering, CranfieldUniversity, 2010.