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Research Paper

Effect of a cross-flow opening on the performance of

a centrifugal fan in a combine harvester: Computational and

experimental study

Mekonnen Gebreslasie Gebrehiwot a,b,*, Josse De Baerdemaeker a, Martine Baelmans b

aDivision of Mechatronics, Biostatistics and Sensors (MeBioS), Department of Biosystems, K.U.Leuven, Kasteelpark Arenberg 30,

B-3001 Heverlee, BelgiumbDepartment of Mechanical Engineering, K.U.Leuven, Celestijnenlaan 300A, Heverlee 3001, Belgium

a r t i c l e i n f o

Article history:

Received 26 January 2009

Received in revised form

28 October 2009

Accepted 11 November 2009

Published online 11 December 2009

In modern harvesting machines, one of the critical factors to fulfil the current demand of

capacity and output under a wide range of field and crop conditions is the capacity of the

cleaning fan. In order to obtain an effective cleaning action, thefan has to generate a forceful

and even air flow over and through the complete width of sieves. This paper presents the

effectof a cross-flowopening on thedistribution of flow along thewidth of a forward curved,

wide centrifugal fan withtwo parallel outlets.Computational Fluid Dynamics (CFD)is utilized

to study theeffect of theaddition of a cross-flowopening on theperformance of thefan using

three fans of similar geometries but different in their cross-flow opening. Velocity profiles at

the outlets of oneof thefans are measured using X-wire hot-wire anemometers and they areused to validate the CFD simulations. Loads on the fan are created by using perforated plates

whose resistancecurves have been determined in a wind tunnel. It is found outthat addition

of thecross-flow opening throughout thewhole widthof thefan plays a crucial role in having

uniform air flow along the width of the outlets of the fan. Comparisons between the simu-

lations and measurements generally show good agreement.

2009 IAgrE. Published by Elsevier Ltd. All rights reserved.

1. Introduction

The process on a typical agricultural combine can be summa-

rized as shown in Fig. 1. The crop header apparatus is used to

reap grain from the crop and feeds the threshing apparatus,

which separatesgrain fromstraw and chaff, collectively known

as material other than grain (MOG). The grain, chaff and small

bits of straw fall through the openings in theconcave under the

rotors in thethreshing apparatusand thestrawis carried bythe

straw walker towards the rear end of the combine. While

a substantial portion of the grain is separated from the MOG in

the threshing cylinders, the threshed crop material comprises

both grain kernels and discardable material, such as chaff and

straw particles, and hence further cleaning is required. This

important step is performed in the cleaning section.

The combine cleaning section includes a grain pan located

below the threshing cylinders, a fan and oscillating cleaning

sieves. The grain pan, which has a corrugated surface, is

installed below the threshing mechanism to receive the

mixture of grain, chaff and pieces of straw that have passed

through the openings in the concave under the rotors, and

guide it to the oscillating sieves in a uniform and even manner.

A stream of air from the fan is used to remove light materials

from the mixture, to assist in positioning particles over sieve

* Corresponding author. Department of Mechanical Engineering, K.U.Leuven, Celestijnenlaan 300A, Heverlee 3001, Belgium.E-mail address: [email protected] (M.G. Gebrehiwot).

A v a i l a b l e a t w w w . s c i e n c e d i r e c t . c o m

j o u r n a l h o m e p a g e : w w w . e l s e v i e r . c om / l o c a t e / i s s n / 1 5 37 5 1 1 0

b i o s y s t e m s e n g i n e e r i n g 1 0 5 ( 2 0 1 0 ) 2 4 7 2 5 6

1537-5110/$ see front matter 2009 IAgrE. Published by Elsevier Ltd. All rights reserved.

doi:10.1016/j.biosystemseng.2009.11.003
mailto:[email protected]://www.elsevier.com/locate/issn/15375110http://www.elsevier.com/locate/issn/15375110mailto:[email protected]

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openings, and to move particlesalongthe surface if they do not

pass through the openings. In order to obtain an effective

cleaning action, the fan has to generate an even air flow with

proper speed over and through the complete width of sieves.

The proper air speed can be determined from aerodynamic

properties of agricultural materials which are terminal velocity

and drag coefficient (Kutzbach and Quick, 1999; Khoshtaghaza

and Mehdizadeh, 2006).

With the increasing power and output demands of the

modern grain combine, the cleaning section capacity has

become a limiting factor. The most readily achieved method of

increasing the cleaning capacity is by increasing the width of

the combine and the sieves to spread the crop material across

a wider area in a thinner veil. Increasing the width of the

cleaning sieves, so as to increase cleaning section capacity,also

involves havingto modifythe air flow across the increased size

of the cleaning sieves. The inherently uneven air distribution of

cleaning fans is accentuated with an increase in the width of

the cleaning fans.

Two kinds of fans commonly used for this function are

centrifugal and cross-flow fans (Fig.2). Cross-flow fansproduce

even wind distribution along the width of the sieves but are

criticized for their inability to generate enough static pressure

and the necessarily stable and steep pressurevolume charac-

teristics. The major concern with the wide centrifugal cleaning

fans (also called paddle fans), however, is that the required

wind distribution over the width of the sieves can hardly be

realized.As thewidthof thefan increases, theuniformity ofthe

flow at the outlets deteriorates. In wide cleaning shoes

(1.72.0 m), the radial fan drawing air from each side inlet,

delivers little air in the sides and in the middle part. Therefore,

34 blowers are arranged side by side with gaps in between

them for suction, as patented by Claas and Tophinke (1981).

This arrangement has its own limitations as there is not

enough wind on parts of the sieves due to the gaps between

fans. When the space between the fans is made smaller to

reduce the unevenness of wind distribution, the suction is

hampered.

Several researchers have studied air flow in combine

cleaning shoes. According to Streicher et al. (1986), researchers

Kerber and Lucas (1969) discussed how shoe design changes

improved the uniformity of the flow. Kerber and Lucas (1969)

showed how appropriate air flow levels could be achieved for

unloaded conditions but did not discuss velocities under load.

Streicher et al. (1986) also mentioned that Hengen (1963), in his

laboratory measurements of shoe air flow, showedlow airflow

levelsin thecentre of thesievewhichthey attributed to theexit

velocity profile of the paddle-wheel type fans used. Under

loadedconditions,velocities were lower in the front and higher

in therear comparedto theunloadedcase.Streicher etal. (1986)

used thermistors to measure air velocities at multiple locations

in the combine cleaning shoe during harvesting of wheat at

MOG flow rates up to 10 kg s1. They found that in general,

velocities in the chaffer decrease as total material flow rate

(grain and MOG) increases. However at intermediate total

material flowrates, velocities at therear increase slightly.They

also concludedthatthe velocities at thesidesof themid section

of the chaffer appear to be the most responsive to changes in

total material flow rate through the combine.

Researchers have suggested geometric changes to the

paddle fan to obtain an efficient cleaning action. Peters (1995)

presented a cleaning fan with two outlets, with the first outlet

Nomenclature

Cf resistance coefficient

f non-dimensional flow rate

v flow velocity (ms1)

j non-dimensional pressure

Dp pressure drop (Pa)

m fluid viscosity (kg m1 s1)

Q flow rate (m3 s1)

a permeability of the medium (m2)

N fan speed (rad s1)

C2 pressure-jump coefficient (m1)

r density (kg m3)

Dm thickness of the medium (m)

D diameter (m)

Fig. 1 Schematic diagram showing the main components of the combine harvester. (1) Thresher, (2) elevator, (3) threshing

cylinders, (4) straw walker, (5) grain pan, (6) fan, (7) cleaning sieves, (8) clean seed auger and (9) tailings auger.


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having two ducts. The primary duct directs an air blast

through the grain and chaff falling from the threshing system

to the step pan, the secondary duct directs a secondary airblast through the grain and chaff from the step pan to

a conditioning pan and thesecond outlet directs an air blast to

the chaffer. The inventor claims to provide a cleaning fan that

better facilitates the separation of chaff from grain in

a cleaning system of a combine. The inventor however did not

deal with the unequal distribution of wind across the width of

the sieves. Jonckheere (1997) patented a centrifugal fan,

modified to have a cross-flow inlet opening throughout its

width as in a cross-flow fan in addition to the well known

traditional side inlets, installed in a generally volute-shaped

fan housing with two outlets. His housing has a main outlet

duct directed to a chaffer sieve and a lower grain sieve and an

additional outlet duct directed to a pre-cleaning sieve and anassociated grain pan. He claimed to provide an economical

and effective solution for grain losses without affecting the air

flow through the inlets.

Recently, Craessaerts et al. (2008) considered the interaction

between the settings of the cleaning section (e.g., fan speed,

lower sieve opening and upper sieve opening) and the MOG

content in the grain bin to assess the performance of the

cleaning shoe. They used a non-linear genetic polynomial

regression technique to rank a pool of potential sensors as

possible regression variables fora prediction model of the MOG

content in the grain bin. They showedthat the cleaning section

settings, like lower and upper sieve openings, had a minor

effect on the MOG content inthe grain bin in comparisonto thefan speed and the loadings by chaff, straw and grain on the

upper sieve.

Computational and experimental methods may be

employed to assess the performance of fans and other turbo-

machinery. The use of Computational Fluid Dynamics (CFD)for

turbomachinery flows has significantly increased in the past

years. Flow analysis techniques using Unsteady Reynolds-

Averaged Navier-Stokes (URANS) approach have lead to

remarkable progress in several engineering applications.

Recently, more attentionhas been paid to thestudy ofunsteady

phenomenain turbomachines.Zhangetal. (1996) computedthe

three-dimensional viscous flow in a blade passage of a back-

swept centrifugal impeller at the design point using the

standard k3 model. Studies by Dilin et al. (1998) and Thakur

et al. (2002) for a centrifugal blower, Muggli et al. (2002) for

a mixed flow pump, Yedidiah (2008) for design of a centrifugalpump and Miner (2000) for axial and mixed flow pumps also

demonstrated the accuracy of CFD for turbomachinery

performance prediction. Furthermore, when combined with

measurements, CFD provides a complementary tool for simu-

lation, design, optimization andanalysis of theflow field inside

a turbomachine.

For experimental validation, hot-wire anemometry (HWA)

is a widespread measurement technique in the study of

turbomachinery flow. Arias-Garcia et al. (2001) have performed

time-averaged HWA velocity measurements of both steady

and pulsating flow in a Close Coupled Catalyst manifold. Ker-

gourlay et al. (2006) used HWA to measure the velocity

components in the near field, downstream of axial fans.From the literature review, it can be seen that a lot of

research has been done with respect to the efficiency of the

cleaning shoe. It has been shown that the required wind

distribution over the width of the sieves can hardly be realized

using the ordinarytwin-inlet centrifugalfan. As a consequence

of the total capacity increase of the combine, the width of the

fan has been increased and this has accentuated the inher-

ently uneven air distribution on the sieves. However there is

limited research that focused on the design modification of the

fan, to improve the performance of the cleaning shoe.

To this end, this paper investigates the effect of an addi-

tional inlet opening on the performance of the cleaning fan.

CFD is combined with experimental measurements using hot-wire anemometers to study the influence of a cross-flow

opening on the performance and flow distribution of a forward

curved wide centrifugal fan with two parallel outlets. The

objective is to have a paddle fan that is capable of delivering

uniform distribution of air flow over the width of the cleaning

section.

2. Materials and methods

Three forward curved fans shown in Fig. 3 are considered. All

fans are of the same dimension. Fan-I is an ordinary forward

curved centrifugal fan with two axial inlets, while fan-II and

Fig. 2 (a) Centrifugal fan and (b) cross-flow fan.

b i o s y s t e m s e n g i n e e r i n g 1 0 5 ( 2 0 1 0 ) 2 4 7 2 5 6 249

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fan-III have a cross-flow inlet in addition to the axial inlet

opening in fan-I. The cross-flow inlet in fan-II runs over the

whole width of the fan while in fan-III, it is limited to 2/3rd of

the width, covering 1/3rd of it at each side. All fans have the

same impeller of six forward curved blades at a uniform pitchwith an external to internal diameter ratio D2/D1 of 1.64 and

width to diameter ratio w/D2 of 4.7.

The standard description of a fans performance is given in

terms of thepressure rise (Dp) asa functionof the flow rate(Q).

These variables are best considered in dimensionless form.

The length and velocity scales commonly used are the fan

diameter (D) and speed (N ), respectively. The non-dimen-

sional pressure coefficient (j) and flow coefficient (f) are

defined as follows,

j Dp

r N2D2(1)

f Q

ND3(2)

whereDp is the load (pa), Qis flow rate (m3 s1), N is the angular

speed (s1), r is density (kg m3) and D is the diameter (m).

2.1. Resistance of perforated plates

Perforated plates of effective opening ranging from 35% to

100% (completely open) are used to represent the loads at the

outlets of the fan. The resistance coefficients of the perforated

plates used were determined in a wind tunnel. The experi-

mental setup of the wind tunnel is shown in Fig. 4. The setup

consists of a fan for creating flow, a standard orifice plate tomeasure the mass flow rate in the tunnel and a test section

where the perforated plate is placed. All measurements were

taken using two Halstrup differential pressure sensors with

accuracy of1% full scale. One sensor measured the pressure

drop across the test section and the other sensor measured

the pressure drop across the orifice plate, which is used to

determine the mass flow rate through the tunnel. The

measurements were conducted at nine different flow rates.

The pressure drops in the empty (without perforated plates)

wind tunnel were also measured and these measurements

were used to correct the raw data of the pressure drops over

the perforated plates. Measurements of resistance of each

plate were repeated using two and three plates in the test

section to ensure repeatability. When multiple plates were

considered, they were placed sufficiently far from each other

to prevent flow interactions. For measurements with two

plates, the intermediate distance was1.4 m, whereas for three

plates it was 0.7 m.

2.2. Computational method (CFD)

2.2.1. Governing equations and numerical scheme

In order to assess the effect of the cross-flow opening on the

performance of the centrifugal fan, the commercial CFD soft-

ware FLUENT is utilized. The transient, three-dimensional,

viscous, incompressible URANS equations are solved. More-

over, the Re-Normalization Group (RNG) k3 is used as a turbu-

lence model. The pressure correction is realized with the

SIMPLE algorithm (Chernobrovkin and Lakshminarayana, 1999;

Fluent, 2006). The calculation is performed unsteady, becauseof

the highlytransient flow in the blade channels (Seo et al., 2003).The sliding mesh technique is used to obtain the final

unsteady results (Fluent, 2006). In this case, the grids change

their relative position during the calculations according to the

angular velocity of the impeller. Using the sliding mesh

model, the rotor domain was defined as a moving zone with

rotational speed of 900 rpm. The time step, Dt, of the unsteady

calculations was set to 3.7 $ 104 s, chosen considering the

rotational speed of the impeller in such a way that one

complete impeller revolution is performed after each 180 time

steps. This value was chosen to minimize the computational

Fig. 3 Geometrical configuration of the fans under consideration.

Fig. 4 Wind tunnel setup to measure air flow resistances

of perforated plates.


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time while retaining good accuracy. It is small enough to get

the necessary time resolution and to capture the phenomena

associated with the blade passage and their interactions with

the volute casing wall.

The perforated plates at the end of the outlet ducts of the

fan are modelled as a porous jump boundary condition. The

porous jump model is a one-dimensional approximation of

a porous medium. The thin porous medium has a finitethickness over which the pressure change is defined as

a combination of Darcys Law and an additional inertial loss

term (Fluent, 2006):

Dp

m

av C2

12rv2

Dm (3)

where m is the fluid viscosity, a is the permeability of the

medium, C2 is the pressure-jump coefficient, v is the velocity

normal to the porous face, and Dm is the thickness of the

medium. Appropriate values for a and C2 can be calculated

from the known pressure-drop/velocity curves of the perfo-

rated plates determined in Section 2.1.Due to the complexity of the geometry,it was impossible to

converge from an initialized flow field to the final solution

using the numerical schemes described above. Instead,

a gradual solution procedure was executed to ensure stable

convergence. The CFD simulation process began with a steady

flow calculation using the Multiple Reference Frames (MRF)

technique, low under-relaxation factors (URF), which are

parameters that can be adjusted to increase stability at the

cost of slower convergence, and first-order discretization

schemes to generate the initial condition. Then, the desired

elements of the model were gradually added until the final

solution was achieved with the desired model and numerical

precision (Gebrehiwot et al., 2006).

2.2.2. Computational domain, grid generation and boundary

conditions

The geometrical models and their meshes are generated using

GAMBIT, the pre-processer of FLUENT. The whole fan process

including its rotor-stator interaction is modelled. Since the fan

is symmetrical, half of the domain is simulated to conserve

computational resources. The computational domain of fan-II

together with the boundary conditions is shown in Fig. 5. The

grid is divided into two zones, the moving zone and the

stationary zone. The moving zone, the domain in contact with

the rotating blades, is modelled as a sliding mesh zone so that

an unsteady solution using the Sliding mesh techniqueprovided by FLUENT is possible. Interface zones are used

between the two zones for data exchange between the rotor

and stator. The pressure-jump boundary condition at the

outlets is explained in Section 2.2.1. In order to resolve reliable

turbulence phenomena near the walls, finer elements are

used in the boundary layer near all walls such that each wall-

adjacent cells centre is located within the log-law region This

is expressed by 30< y< 300 where y is a mesh-dependent

dimensionless distance that quantifies to what degree the

wall layer is resolved (Ferziger and Peric, 2002). No-slip

condition and standard wall functions are used at the walls

and impeller blades. At theinlet a gauge pressure of 101325 Pa

has been applied.

2.3. Experiments

The experimental measurements were conducted in a test

setup of the fan, constructed to simulate the situation within

the cleaning section of the combine harvester. A standard hot-

wire system from Dantec Dynamics, with software (Stream-

Ware 2.8) designed for calibrating the hot-wire probes and

recording measurement data, was used to make wind velocity

measurements at the outletsof the fan. The fan is driven by an

electric motor and the speed is regulated by a frequency regu-

lator. At one end of the fan shaft, an external trigger (rotation

counter) is mounted in order for the velocity measurements to

be phase-locked with the rotation of the rotor. All measure-

ments were taken at 900 rpm.Fig. 6 shows schematic presentation of the fan, hot-wire

anemometer probes and the geometry of perforated plates

used. Perforated plates of different open area ratio that corre-

spond to different pressure resistance, whose air flow resis-

tance coefficients are measured in Section 2.1, were placed at

the outlets of the fan as loads. The probe used (Fig. 7(b)) is

a Dantec 55P61 X-wire type, in which two platinum-plated

tungsten wires of 5 mm diameter are inclined at 45 to the air

flow and at 90 to each other. This type of probe gives two

independent measurements, which can be used for measuring

flow velocities in two dimensions (Brunn, 1995).

The transient velocity measurements are taken inside the

outlet ducts before the perforated plates at points along 4equidistant vertical lines in one half of the width of the fan.

Measurementsare taken at 4 6 pointsfor theupperoutlet and

4 10 points for the lower outlet. An automated Dantec

Dynamics traverse system is used to move the hot-wire sensor

to each measurement point. The measurement points are

sufficiently far from the perforated plates to avoid the effect of

the perforated plates on the stream lines of the flow.

3. Results and discussion

The measured pressure drops over each perforated plate are

shown in Fig. 7 as a function of the free-stream velocity. The

Fig. 5 Computational domain of fan-II, grid system and

definition of boundary conditions.


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pressure drop curves for each perforated plate agree with the

well known second order pressure drop characteristics. The

pressure drop increases with the increasing free-stream

velocity of air and the relationship between them is quadratic.

The total resistance coefficients for each plate are then

calculated as the ratio of the measured pressure drop and the

dynamic pressure based on the free-stream velocity as,

Cf 2Dprv2 (4)

where Cf total resistance coefficient; Dp pressure drop

across one perforated plate and v free-stream velocity.

3.1. Grid sensitivity

Grid sensitivity analysis has been done on fan-II using coarse,

medium and fine meshes at a load ofj 0.59, which corre-

sponds to a perforated plate of effective opening of 32%. The

number of cells is shown in Table 1. The non-dimensional

flow coefficient, f, has been monitored at the lower outlet for

each case, and comparison between the results for one revo-

lution of the impeller is shown in Fig. 8. The computationapproaches convergence as computational cell numbers

increase. When the mesh is increased to fine, the influence of

cell size becomes very weak and the computation accuracy is

acceptable. Thus, the fine computational grid is adopted forall

fans.

Fig.9 comparesthe calculated andmeasured time history of

the non-dimensional, area-averaged wind velocity through

both outlets at no load for one cycle of the impeller of fan-II at

900 rpm. The wind velocity, both calculated and measured, at

the outlets of the fan is a transient wave with the number of

pulses corresponding to the number of blades of the impeller.

Moreover, from the same figure it can be seen that at no load,

the computed results match the measurement results to within3%.At higherloads, corresponding to effectiveopeningof 32%

(j 0.59), however, the error is larger but within 10%. In

Fig. 6 Schematic figure of the measurements setup: (a) fan-II together with the perforated plates at the outlets and the hot-

wire probe with its axis aligned to the flow direction (b) enlarged figure of the X-wire probe (Brunn, 1995) and (c) perforated

plate with staggered holes.

Fig. 7 Resistance curves of perforated plates. , Empty

section, D61% open, > 51% open, C 45% open, B 32%

open.

Table 1 Number of cells of three different meshes usedfor grid sensitivity analysis

Coarse mesh Medium mesh Fine mesh

Rotating zone 114 019 150 025 228 038

Stationary zone 147 659 190 295 306 765

Total number

of cells

261 678 340 320 534 803


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general, this range of error is not uncommon in turboma-

chinery simulations (Chernobrovkin and Lakshminarayana,

1999). Possible causes of such discrepancy between experi-

mental and simulateddata maybe the automaticallygenerated

computational grids and other discretization errors on the

simulation side and errors that may be incurred during cali-

bration of the hot-wire probe and during measurements. It

should be stressed that CFD can never replace measurements

Fig. 8 Grid analysis: comparison of the flow coefficient at the lower outlet for the coarse, medium and fine meshes.

C Coarse mesh, : medium mesh, - fine mesh.

Fig. 9 Comparison of results of the time history of wind velocity through both outlets of fan-II from CFD simulations and

hot-wire anemometer measurements: (a) upper outlet (j=0), (b) lower outlet (j=0), (c) upper outlet (j=0.59) and (d) lower

outlet (j=0.59), - - - - - - Measurement, dCFD.


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completely but the amount of experimentation and the overall

cost of the research canbe significantlyreduced.Therefore, the

CFD simulations are used as a supplement to the experimental

measurements. The CFD simulations give distribution of air

flow velocity and pressure in the entire domain which is diffi-

cult to get from experiments because of the time and cost

involved. Consequently, for the purpose of evaluating alterna-

tive designs before experimentaltesting takes place,an error ofthis magnitude is acceptable.

The velocity vector plots of the cross-flow opening of fan-II

and fan-III at no load are shown in Fig. 10. The cross-flow

opening of fan-II (Fig. 10(a)) is actually not fully an inlet, as air

is coming out of some parts of it. There is a relatively large

amount of air coming out of the frontal middle area across the

width. In the other parts of the opening, air is entering the fan

through this opening at a relatively smaller velocity. Thus,

even when running at no load this opening is not completely

an inlet. Close monitoring of the flow in this opening shows

that a small net amount of air enters the fan. Fig. 10(b) shows

that in fan-III, closing the mid section of the cross-flow

opening through which flow is observed to come out in fan-II,did not improve thesituation as largeamount of air comesout

through the frontal area of the cross-flow opening.

Velocity contours of the outlets of all fans at representative

loads of j 0 and j 0.59 are shown in Fig. 11. From the

contours at both high and low loads, it can be seen that in fan-I

there isverylowflowcoming out through the endsof theupper

outlet, the flow is concentrated to the middle span of the fan.

This creates a problem duringcleaning becausepoor airoutput

at the ends means that those crop materials disposed towards

theoppositesidesofthesievesdonotbenefitfromthesameair

Fig. 10 Velocity vector plots of the cross-flow opening at

no load: (a) fan-II and (b) fan-III.

Fig. 11 Representative contour plots of total air speed of air at the outlets of the fans with the dashed lines representing the

symmetry planes, (a) unloaded (j[0) and (b) loaded (j[0.59).


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flow regime as those crop materials disposed towards the

lateral centre of the sieve. Thus the effective cleaning area of

the combine is significantly reduced by end air effects at

opposite ends of the fan. This problem becomes even worse

when the combine is operated on hills or a field with uneven

terrain during which the combine is tilted such that most of

the crop material tends to be accumulated on one side of the

cleaning sieve.

Additionof thecross-flow opening as in fan-IIhas improved

theflow distributionin the upper outlet. But in fan-III (reducingthe area of the cross-flow opening to 2/3 of that of fan-II),

worsens the distribution of air in the upper outlet. The perfor-

mance curves of the three fans drawn from the CFD results are

displayed in Fig. 12 together with the characteristics of fan-II

from the hot-wire anemometer measurements.

From theCFD results (see fan-I andfan-II), it canbe observed

that availability of the cross-flow opening increases the flow

rate created by thefan at low loads.At higherloads however, all

fans generate similar flow rates.

Looking at the slopes of the characteristic curves for the

three fans, it can be seen that addition of the cross-flow

opening reduces the amount of flow delivered by the fan at

higher loads however its presence throughout the whole widthof the fan improves the uniformity of the flow through the

outlets of the fan.

4. Conclusions

Three forward curved wide centrifugal based fans are studied

numerically and experimentally for the effect of addition of

a cross-flow inlet on the uniformity of low at the outlets and

the total performance. CFD is used to study the effect of

addition of a cross-flow opening on the performance of the

centrifugal fan. Experiments are performed using X-wire hot-

wire anemometer to measure the velocity at the outlets of the

fan to validate the simulations. Comparison between the

results demonstrated that the three-dimensional URANS CFD

simulations can predict the performance of the cleaning fan

with a reasonable error of less that 10%. Availability of the

cross-flow opening increases the flow rate created by the fan

at lowloads. At higher loads however, allfans generate similar

flow rates. Moreover, the addition of the cross-flow opening

across the whole width plays an important role for axial

distribution of air in the outlets. End air effects at opposite

ends of the fan are reduced thereby increasing the effectivecleaning area of the combine. This effect is especially impor-

tant when the combine is operated on hills or a field with

uneven terrain during which the combine is tilted such that

most of the crop material tends to be accumulatedon one side

of the cleaning sieve.

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