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Original Article

Experimental and numericalinvestigation of a gearlessone-motor contra-rotating fan

M Heinrich1, C Friebe

2and R Schwarze

1

AbstractA gearless one-motor concept for contra-rotating fans is presented in this article. The rotors are mounted to an electricmotor using only one shaft. The coupling between both rotors is realised by utilising the conservation of angularmomentum. The contra-rotating fans has a diameter of 200mm at a design speed of 2100 min 1 for the first stageand 1200min 1 for the second stage. It has been designed and investigated through a series of experiments by theInstitute of Air Handling and Refrigeration in Dresden. The performance map and 2D particle image velocimetrymeasurements have been conducted. Numerical models for 3D quasi-steady state and transient simulations havebeen implemented and carried out by the Institute of Mechanics and Fluid Dynamics. The results show a good agreementbetween the quasi-steady, the transient simulations and the experiment. However, when close to stall, the time-resolvedsimulations show a superior performance compared with steady-state computations.

KeywordsContra-rotating fan, gearless, computational fluid dynamics, simulation, validation

Date received: 26 August 2015; accepted: 15 January 2016

IntroductionContra-rotating fans (CRF) reduce the swirl in thewake ow as well as achieve higher power densitiescompared with single-rotor fans (SRF) with dischargeguide vanes. In practice, this approach is alreadybeing used to expand the application range of axialfans due to the higher total pressure rise and increasedefficiency. This is due to the improved recovery of kinetic energy from the rst rotor. ConventionalCRF utilise one motor for each rotor or a gear forpower transmission. The compact design makes the

CRF favourable for various applications, where dueto lack of space, large industrial fans cannot be used,e.g. tunnel and mine ventilation, as well as high-loaded fans for high-speed propulsion systems.However, using two motors or one motor with agear is more complicated, expensive and requiresmore maintenance compared with SRF.

The CRF system has been around for many years.The very rst experimental investigations of opencontra-rotating propellers were already being con-ducted in the 1930s 1,2 showing the increase of efficiencycompared with conventional propellers. In recentyears, more research work focusing on CRFs hasbeen published. Advantages of CRF has been shownby Wang et al. 3 He compared a SRF, a conventionalrotor–stator stage and a CRF. The results showed

larger pressure rise and efficiency with lower rotationrate at a given power consumption. Cho et al. 4 showedhow to design rotor blades and Wang et al. 5 investi-gated different blade pitch angles to validate andimprove recommendations for shaft power matchingof CRFs. Another focus has been on the reduction of sound emission, see Shin et al. 6 and Li et al. 7 Grassoet al. 8,9 focused on the reduction of the mechanicalstresses and broadband noise scattered from the trail-ing-edge of the blades.

The inuence of the rotor spacing has a limitedinuence on performance and efficiency (see Nouri

et al.10

). Varying the axial distance based on the rstrotor chord length between 0.17 and 0.86 resulted inno signicant change. Even when increasing to 3.1,the global performance was only decreased by 5%.Similar results were found by Sharma et al. 11 andMistry and Pradeep. 12 They concluded that the best

1 Institute of Mechanics and Fluid Dynamics, Technical UniversityBergakademie Freiberg, Freiberg, Germany2 Institute of Air Handling and Refrigeration (ILK), Dresden, Germany

Corresponding author:

Martin Heinrich, Institute of Mechanics and Fluid Dynamics, TechnicalUniversity Bergakademie Freiberg, Lampadiusstrasse 4, 09596 FreibergGermany.Email: [email protected]

Proc IMechE Part A: J Power and Energy 0(0) 1–10! IMechE 2016Reprints and permissions:sagepub.co.uk/journalsPermissions.navDOI: 10.1177/0957650916633014pia.sagepub.com

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performance is found at an axial distance of approxi-mately 0.9 chord length.

The relative rotational speed of the rst and sec-ond stage has a major impact on the performance of CRF and compressors. Wilcox and Wright 13,14 inves-tigated a transonic contra-rotating compressor. As a

result, a variation of the rotational speed of one stageonly had a signicant inuence on the overall stageperformance. A higher tip speed of the second rotorresulted in an increase in total adiabatic efficiency.Nagano et al. 15 and Shirma 16 showed that a correctspeed radio could reduce stall on the rst rotor. Royet al. 17 concluded from their results that the rst rotorhas a higher effect on the mass ow capacity whereasthe second one mainly inuences the pressure rise andoperating range. Furthermore, a wider operatingrange can be achieved by having a higher secondrotor speed compared with the rst rotor. The results

of Chen et al.18

support the signicant inuence of speed ratio on off-design performance and efficiency.In this article, a new CRF concept is presented using

only one gearless motor. The concept was developedfor compact fans by the Institute of Air Handling andRefrigeration (ILK) in Dresden, Germany. Because of the innovative design, only one shaft is needed whilethe torque is equally distributed between both rotors,thus improving the ow eld inside the fan. This way,the speed ratio between the rst and second rotor isnot xed and can vary for different operating condi-tions. Moreover, the design is lighter, more compactand requires less maintenance compared with conven-tional CRF.

Gearless one-motor conceptThe new concept for a CRF utilises only one shaft forthe drive train of the two fan blades (see Figure 1).

The rst rotor is connected to the internal rotor of theelectric motor whereas the second one is connected tothe external rotor of the motor. Grooved ball bearingsare used for the mountings on both ends. This way, theelectric motor has no xed position but can rotatefreely. For this reason, a contactless power transmission

is needed to supply power to the motor. Because of theconservation of angular momentum, if the electricmotor starts spinning, the rst fan turns in one directionwhereas the second fan turns in the opposite direction.

The hub has a total length of 290 mm and an innerdiameter of 90mm. The pipe diameter is 200 mm.Three struts are used at the spinner as well as behindsecond rotor to connect the fan with the housing. Thelatter is made of acrylic glass to grant optical access forstereo particle image velocimetry (PIV) measurements.

Because of the law of conservation of angularmomentum and under the assumption of no friction,

the magnitude of the momentum M of both fans isalways equal

M 1 ¼ M 2 ð1Þ

with the power of each rotor P i dened as

P i ¼ 2 M i ni ð2Þ

and the rotational speed of each fan ni . The shaftpower is equal to the sum of the power of each rotor

P shaft ¼ P 1 þ P 2: ð3Þ

The following ratio of rotational speed and power of each rotor can be derived from these equations:

n1n2

¼ P 1P 2

: ð4Þ

Figure 1. Basic setup of the contra-rotating fan.

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nozzle. The PIV measurement system is calibrated inadvance using a multi-level target. One blade passageis resolved using 15 steps. For each step, 200 framesare taken. Dantec Dynamic Studio (version 3.20.89) isused for calculating the ow velocity from the particlemovement.

The accuracy of the measurements at design pointis as follows:

. Flow rate: (572 5.72) m 3 /h

. Pressure rise: (128.7 0.6) Pa

. Power consumption: (43.7 0.9) W

. Efficiency: (46.8 1.1) %

The PIV measurements possess the followinguncertainty:

. In-plane velocity components: 2%

.

Out-of-plane velocity components: 4%

Numerical setupThe three-dimensional, incompressible and Reynolds-averaged Navier–Stokes equations are solved usingthe open source computational uid dynamics(CFD) library OpenFOAM Version 2.3.x. 19 Theseequations are given as follows:

r u ¼ 0, ð6Þ

@u@t

þ r uuð Þ ¼ r p þ r ð þ RS Þ ð7Þwhere denotes the density, p the pressure, þ RS

the viscous and Reynolds stress (RS) tensor and u theReynolds-averaged absolute velocity. The SIMPLEalgorithm is used for the pressure-velocity coupling.The k ! shear-stress transport (SST) turbulencemodel 20 is employed to model RS . The convectiveterms in the transport equations are discretised with

second interpolation for u and rst-order upwind inter-polation for the turbulent quantities. Gradients are dis-cretised with the central differencing scheme. In case of transient sliding mesh simulations, a rst-order explicitEuler scheme is used for temporal discretisation. Amaximum Courant number of 2.0 is used to adjust

the time step size. The nal residuals for steady-statesimulations are set to 10 5 . Furthermore, the staticpressure at the inlet is monitored to judge convergence.

Steady-state simulations employ the Multiple-Reference-Frame (MRF) approach to model the rota-tion of the fans. Here, all parts of the mesh are xed inspace and additional terms are added to the momen-tum equation to account for the rotation of the rotor.This way, no blade passage or time-dependent effectscan be resolved. Transient simulations are performedusing the sliding mesh approach. In this case, themesh in the rotor is moving around the axis of rota-

tion. The connection between the rotating and xedmesh parts is achieved using mesh interfaces. Six revo-lutions with respect to the second rotor are computedfor the unsteady simulations and the results are aver-aged of three rotations.

The numerical model of the test bench is shown inFigure 3. It consists of the air inow channel, theCRF and the outlet channel. At the inlet, a mappedboundary condition is used for the velocity and tur-bulent quantities to achieve a fully developed owprole. The average inlet velocity is varied to achievedifferent ow rates. At the outlet, the relative staticpressure with respect to the ambient conditions is setto zero and the other variables are treated as a zerogradient. The uid is air with a kinematic viscosity of

¼ 1.85 10 5 m 2 /s and a density of ¼ 1.190kg/m 3 .This resembles from ambient conditions of 24 C at101500 Pa.

The domain is meshed using an unstructured tetra-hedral mesh. It is rened in the wake of the blades andaround the tips. Furthermore, an extrusion mesh hasbeen used for the inlet and outlet channel, where theinlet pipe is twice the diameter and the outlet pipe four

Figure 3. Computational domain for the numerical model.




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times the diameter. This reduces the boundary effectson the numerical solution. Five different mesh sizes,(4.02–10.30) 106 cells, are used to assess the griddependency using the MRF approach (see Figure 4)at the example of the pressure rise coefficient (seeequation (3)). The higher the cell count, the lower the

deviation compared with the experimental results.However, a mesh resolution of 5.66 106 cellsshows the best compromise between accuracy andcomputational cost. For this reason, this grid reso-lution is used for the further computations. A cutthrough the mesh is shown in Figure 5. The localrenement around the blades and the extrusionmesh in front of the fan are clearly visible. The max-imum aspect ratio of the mesh is 10, the maximumskewness 0.7, the minimum included angle 17 and theaverage yþ value on the surfaces is about 10. For thisreason, an all- yþ wall function is used to model near-

wall turbulence.

Results

ExperimentThe experimental measurements have been conductedby Czarnecki 21 at the ILK. Two different setups areinvestigated. The rst one is a standard SRF using the

rst rotor only. In this case, the second rotor isreplaced by a straight pipe. The other aspects of thesetup are the same as for the CRF. In the second case,the CRF is analysed. The ow and pressure rise coef-cients as well as the power coefficient are dened asfollows:

’ ¼_V

2

4 D3n1

, ¼ ptot2

2 D2n21

, l ¼ P F 4

8 D5n31

ð8Þ

based on the total pressure rise ptot . The experimen-

tal measurements of the fan performance are shown inFigure 6. The general shapes of the pressure rise coef-cient curves are similar and the maxima are nearly atthe same ow rate. However, the pressure rise of theCRF is considerably higher compared with the SRF.The reason for this is the second rotor, whichreduces the swirl in the wake of the fan and thereforeincreases the static pressure rise. Furthermore, themaximum efficiency of the CRF is increased byapproximately 5% and shifted to lower ow rates.Additionally, the range of higher efficiencies is widercompared with the efficiency peak of the SRF. Thisimproves the performance at off-design conditions.

Figure 7 shows the angular velocities of the rstand second rotor for the CRF. These vary by up to20% with respect to the design speed. This indicatesthat the momentum distribution between the rst andsecond stage is uneven in case of xed rotationalspeeds. It further shows the advantage of the newapproach to vary the rotational speed based on theloading of the blades. Because of this variation, thedesign speed cannot be used for the numerical

0.36

0.38

0.4

0.42

0.44

4 6 8 10 12

ψ [ - ]

Cell count x 10 6

ExperimentSimulation

Figure 4. Pressure rise coefficient at design point as afunction of the cell count using the MRF approach.MRF: Multiple-Reference-Frame.

Figure 5. Cross-sectional view of the computational mesh.

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simulations. The rotational speed, which is differentfor each operating point, has to be adjusted accordingto the experiments.

Simulations

Figure 8 compares the performance map of the CRFfor the MRF and sliding mesh simulations with theexperiment. The MRF simulations show a reasonableagreement at high and low ow rates. However, thecharacteristic maximum and drop in pressure riseare not captured. This indicates that the unstableregion of the performance curve is not modelled

correctly. One reason for this is the MRF approach,which does not account for the unsteady behaviour of the ow. Furthermore, rotor–rotor interactions aregreatly simplied.

The transient sliding mesh simulations show abetter overall accordance with the reference data

and the slopes are met very well. Furthermore, thesudden drop in pressure is predicted well. However,it also shifts slightly to higher ow rates. Even theunstable operating range at ’ ¼ 0:15 of the CRF ismodelled correctly. This shows the strong inuence of transient and rotor-passage effects, which cannot beresolved using a MRF approach.

0

0.2

0.4

0.6

0.8

1

1.2

0 0.1 0.2 0.3 0.4 0

0.1

0.2

0.3

0.4

0.5

0.6

ψ [ - ]

E f f i c i e n c y

[ - ]

φ [-]

SRF

CRF

Pressure rise coefficient

Efficiency

Figure 6. Experimental measurements of the pressure rise coefficient and efficiency for the SRF and the CRF.SRF: single-rotor fan; CRF: contra-rotating fan.

0

400

800

1200

1600

2000

2400

0 0.1 0.2 0.3 0.4

n [ m i n - 1 ]

φ [-]

Rotor 1 (2100 min -1)

Rotor 2 (1200 min -1)

Figure 7. Rotational speed of the first and second stage of the contra-rotating fan with design speed in brackets.




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A comparison of the power coefficient is shown inFigure 9. Similar to the previous results, the MRFsimulations overpredict performance at high ow-rates due to the over-estimation of static pressurerise. At the same time, the maximum performance ispredicted at higher ow rates. In contrast to that, the

sliding mesh simulation results resemble a good pre-diction of the experimental results for all ow rates.

Stereo-PIV measurements have been performed toacquire detailed velocity elds in a measurement planelocated 30 mm behind the second rotor. Figure 10shows the circumferential averaged axial ow velocityfor the simulations and the experiment at designspeed. The measured axial velocity increases linearly

up to a maximum of about 9 m/s at a radius of 93mm.Both simulation models do well in capturing the vel-ocity prole.

Turbulent structures downstream of the fan areshown in Figure 11. At design point, only a few largeturbulent structures are visible, particularly in the wake

region of the hub. This indicates a good setup betweenthe rst and second rotor and as a consequence lowswirl behind the fan. In contrast, large vortices areformed at off-design conditions. These move in a hel-ical manner due to the rotation of the ow. This indi-cates higher swirl and turbulence compared with designconditions and is one reason for the reduced efficiencyat off-design ow rates.

0.05

0.06

0.07

0.08

0.09

0.1

0.11

0.12

0.05 0.1 0.15 0.2 0.25 0.3 0.35

λ [ - ]

φ [-]

Experiment

MRF

SlidingMesh

Figure 9. Power coefficient for the experiment and the simulation. _V ¼ 373 m3(Off-design point).

0

0.1

0.2

0.3

0.4

0.5

0.6

0.05 0.1 0.15 0.2 0.25 0.3 0.35

ψ [ - ]

φ [-]

Experiment

MRF

SlidingMesh

Figure 8. Pressure rise coefficient for the experiment and the simulation.

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The velocity magnitude at a radius of 150mm isshown in Figure 12 for the MRF and the sliding meshsimulations. The former one can only correctly pre-dict the average ow eld close to design point.The velocity prole inside the blade channels is homo-geneous. Furthermore, the interaction with the struts

is weak. In contrast to that, the ow in the transientsimulation is much more complex. Transienteffects such as separation and the development of avon Karman vortex street behind the struts canbe observed. The ow is much more uniform in thewake of the second rotor. This indicates thatthe momentum exchange between the wake ow andthe jets is much stronger in the sliding mesh simula-tions. This way, velocity gradients are smoothed and ahomogeneous velocity prole is developed.

ConclusionThis article presents a gearless one-motor concept forCRF. For comparison, the single-rotor and the CRFare investigated through several experiments, whichinclude a test rig for performance measurements andstereo PIV for ow eld measurements. The relative

speed of the CRF is 3300min 1 . As expected, theCRF shows a signicantly higher pressure rise coeffi-cient and power coefficient compared with the SRF.

The rotational speeds of the rst and second rotorvary by up to 20% with respect to the design speed.This shows the uneven momentum distributionbetween both stages in case of xed rotationalspeeds. For this reason, the proposed approach withvarying speeds based on the loading of the blades maybe better for CRF.

0

2

4

6

8

10

12

45 55 65 75 85 95

Radius [mm]

A x i a

l V e

l o c

i t y [ m / s ]

Experiment

MRF

SlidingMesh

Figure 11. Circumferential averaged axial velocity of the simulation and experiment 30mm behind the second rotor at design point.

Figure 10. Lambda2 isosurface downstream the CRF to illustrate large turbulent structures at design point (top) and off-designpoint (bottom).CRF: contra-rotating fan.




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Furthermore, a numerical model is developedfor the CRF and validated using the measurementdata. The results show that steady-state simulation iscapable of resolving the ow eld of the fan and pre-dicts the performance curve as well as local velocityproles. However, transient sliding-mesh simulationsshow a signicantly better performance, in particularwithin the unstable operating range at low ow rates.

Declaration of conflicting interestsThe author(s) declared no potential conicts of interest withrespect to the research, authorship, and/or publication of this article.

FundingThe author(s) received no nancial support for the research,authorship, and/or publication of this article.

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Figure 12. Velocity magnitude at a radius ¼ 75 mm for the steady-state (left) and transient simulations (right) at design point.

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Appendix 1

Notationd Diameterh Hub-to-tip ratiok Turbulent kinetic energy

m MassM Momentumn Rotational speedP Shaft power p Pressuret Timeu Velocity_V Flow rate

y Wall distance

l Power coefficient Kinematic viscosity Density

Shear stress tensor’ Flow coefficient

Pressure rise coefficient! Specific dissipation


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