Computational flow modeling for optimizing industrial fan perfomance characteristics

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European Conference on Computational Fluid Dynamics ECCOMAS CFD 2006 P. Wesseling, E. Oñate, J. Périaux (Eds) TU Delft, The Netherlands, 2006

COMPUTATIONAL FLOW MODELING FOR OPTIMIZING

INDUSTRIAL FAN PERFORMANCE CHARACTERISTICS

W.K. Ng* and M. Damodaran**

Division of Thermal and Fluids Engineering School of Mechanical and Aerospace Engineering

Nanyang Technological University, 50 Nanyang Avenue, Singapore 639798 *Graduate Student **Associate Professor

e-mail: ngwa0003@ntu.edu.sg e-mail: mdamodaran@ntu.edu.sg

Key words: Industrial Centrifugal Ventilation Fan, Sliding Mesh Model

Abstract. The paper describes the computational flow analysis inside industrial ventilation fans to facilitate computational design to achieve optimal fan performance characteristics eventually. Finite volume computation using the notion of sliding meshes with rotational zonal grids is used to solve the unsteady Navier-Stokes equations to simulate internal flow inside both a single inlet and double inlet centrifugal fan for operational conditions specified by an industry vendor. Fan performance characteristics are extracted and verified with experimental data provided by the vendor and also shows the impact of various turbulence models on the predictions. The impact of two fan blade section profiles, namely a blunt edge section and an airfoil section on the performance are also assessed.

1 INTRODUCTION

Much work has been done by various researchers in the application of Computational Fluid Dynamics (CFD) to turbomachinery flows. The state of the art of this practice is shown

in Lakshiminarayana1, Elder et al.2 and in Horlock and Denton3. Many parameters such as

efficiency, pressure rise, flow coefficient, work input, rotor blade tip speeds, and geometrical ratios characterise the operational performance of industrial ventilation fans. While many of these parameters could be obtained by routine wind tunnel tests for specific fan geometries, the routine use of CFD technology in industrial applications enables the possibility of complementing the experimental methods of obtaining the fan performance characteristics of ventilation fans. Flows in these fans are complex, viscous, unsteady and three-dimensional and CFD enables a better understanding of the flow phenomena which would otherwise be impossible or expensive to gather experimentally. In this work, fan performance characteristics is obtained from CFD simulation of a centrifugal industrial fan unit

manufactured by Kruger4 in order to gain insight into the complex flowfields and also to set a

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W.K. Ng and M. Damodaran

rotation on the gauge static pressure distribution at outlet and surface pressure distribution on blades and the impeller unit. While most often the unsteady solution that is sought in a sliding mesh model is time-periodic with the unsteady solution repeating itself over a period related to the moving domains, transient analysis of the flowfield as a result of a sudden reduction in rotational speed on fan performance characteristics can also be assessed.

2 NUMERICAL SIMULATION OF FAN FLOWFIELDS

The flow inside the industrial fan configuration is computed by a finite volume discretisation of the Navier Stokes equations, the general form of which can be written as follows:

V D A S

V A V

QdV U Q d dV

t (1)

where Q ( , , )V E T,D (0, ,V q)T,S (0, g, 0)Tand pI where is the

density of the fluid which can be assumed a constant for incompressible flow, V is the fluid

velocity of the control volume, E is the total energy, D is the non-convective terms, S is a

source term, p is the pressure, is the shear stress tensor, q is the heat fluxes from

conduction and other heat sources and U is the velocity of the control volume which is zero

for the inertial frame of reference. Since the operational speeds of the industrial fan are in the low speed regimes, incompressible flow field assumption is made for these simulations. These equations are discretised into an algebraic form over all the cells in the computational mesh representing the domain of calculation. These sets of algebraic equations are then solved

using the Fluent5 flow solver.

3 MESH GENERATION

The configuration of the single inlet and double inlet industrial fan are as shown in Figures 1(a) and 1(b) respectively.

(a) (b)

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The rotating part which is the impeller has plates on which the blades are arranged. The impeller is then installed in a casing (volute). Flow enters through the eye of the impeller and leaves through the rotating blade passages out to the casing towards the outlet.

The geometrical fan configurations shown in Figure 1 are provided by the industries in

Parasolid6 format and imported into Gambit7 software for the purpose of meshing. The geometry of the blade and its sectional profile influence the fan performance characteristics. The blades are arranged on a impeller plate. On each plate, there are 11 blades arranged equidistantly. The double inlet fan has 22 blades on both sides of the impeller plate. With such a blade arrangement, it is possible to use structured mesh for grid generation according

to Roth & Peikert8. The impeller plate is divided into 11 sections and meshed with

quadrilateral face mesh as shown in Figure 2(a). Using the Cooper scheme, the quadrilateral volume mesh for one section is created by projecting the quadrilateral face mesh upwards. This volume mesh for one section is copied and rotated about the X-axis to create structured volume meshes for the impeller as shown in Figure 2(b).

Blade 11 Blade 2 Blade 3 Blade 7 Blade 6 Blade 5 Blade 4 Blade 9 Blade 10 Blade 1 Blade 8

As the sliding mesh model is used for numerical flow simulation, interfaces need to be created between different fluid zones such that one fluid zone remains stationary, while the other fluid zone is rotating with the specified rotation velocity. For the rotating reference frame Eqn (1) is written in terms of the absolute velocities i.e.

V D A S

V A V

QdV U Q d dV

t (2)

where U r is the relative velocity, is the angular velocity of the rotating frame/zone.

Figure 3 shows the interface location between the blue and grey fluid zones. The purple

(a) Division of blade arrangement into 11 sections with quadrilateral face mesh for one section

(b) Completed volume mesh for blade arrangement plate

Figure 2. Steps in creating volume mesh for blade arrangement plate which includes (a) division of blade arrangement plate into 11 sections with quadrilateral face mesh for one section and (b) completed volume

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coloured fluid zone and the light blue coloured fluid zones for the blade arrangement are rotating about X-axis with the rotation velocity. The grey coloured fluid zone is set to be

stationary. According to Casey9, it is advisable that the computational domain covers the

region upstream and downstream of the blade stages. Figure 3 shows that the grid structure in the upstream and downstream regions of blade stages i.e., the grey coloured areas. Structured volume meshes are created for the single and double inlet fans. As far as possible mesh skewness has been kept low for improved accuracy. The quality of the volume meshes is checked by doing grid dependence studies and other mesh quality tests.

Interface between moving fluid zone and stationary fluid zone

moving fluid zone stationary fluid zone

4 INITIAL AND BOUNDARY CONDITIONS

The volume flow rate extracted from the fan has been fixed by using the gauge static pressure at the pressure inlet boundary. The stagnation pressure at the outlet is an unknown. Since the inlet is exposed to atmospheric air, the inlet has gauge total pressure 0 Pa. Using Bernoulli equation, as shown in equation (3); the gauge static pressure at inlet for a certain volume flow rate was calculated and inputted at the pressure inlet boundary i.e.

2

T

P P

where P is the gauge static pressure at inlet in Pa, PT is the gauge total pressure at inlet in Pa,

is the density of air at 1.22kg/m3, is the velocity at the inlet obtained by dividing the

volume flow rate by the total inlet area. The problem has to be set up with a rotating reference frame for the moving fluid zones first before activating the sliding mesh model. In the rotating reference frame (moving reference frame), the rotational velocity is specified from the operational conditions of the fan. The blade surfaces and rotating impeller walls are set as moving walls but the rotating velocity is zero relative to adjacent fluid zone. This is because the adjacent fluid zone surrounding the impeller blades are already sliding with the prescribed rotational velocity. The moving fluid zones which are attached to the blade and impeller surfaces are set to the specified rotational velocity about the axis of rotation of the impeller. The impeller will rotate clockwise with respect to the outlet in the direction, which expels air towards the outlet. After the solution has converged for the moving reference frame, the sliding mesh model is started to achieve the steady state solution for the ventilation fan. The

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unsteady solver is activated and a suitable time step is selected such that the cyclic behaviour of the ventilation fan is captured. The unsteady solution will exhibit a cyclic behaviour after the initial start up phase has passed. It takes several revolutions of the rotor before the solution reaches steady state convergence. A suitable time step for the unsteady computation using sliding mesh is computed as

Nn

t 60 where N is the angular speed of rotation in rpm and n is

the number of time steps for the fan to complete one revolution. For the case of the fan operating at 2500 rpm, the time step sizes for unsteady state iteration and mesh motion are

both of the _O(10 4)_{seconds. The computation is done using the segregated solver. Since the}

realizable k model produces non-physical turbulent viscosities in situations when the

computational domain contains both rotating and stationary fluid zones which are sliding

meshes, it is not advisable to use realizable k model for sliding mesh model and hence the

standard k model has been selected for this study.

5 RESULTS AND DISCUSSIONS

5.1 Computed Steady State Flow Fields in Single Inlet Fan

It is useful to plot the pressure contour at the outlet as the gauge total pressure will be used later to compute the total efficiency of the single inlet fan. The total efficiency of the fan is defined as BHP Q P_T t where Q is discharge in m 3

/s, BHP is brake horsepower in watts and

PT is gauge total pressure in Pa. Figure 4 shows that there is uneven gauge total pressure

distribution at the outlet. This is because the geometry of single inlet fan is not symmetrical about Y-axis. The flow simulation result gives a gauge total pressure of 2140.107 Pa at the outlet and it is very accurate when compared with the experimental result which gives 2200Pa.

The presence of high computed vorticity at the blade tip region is shown in Figure 5. This implies that the blade tips could be redesigned to minimize flow separation and improve

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the performance of the fan. Rotating stall could be observed when the computed flow fields in the blade passages are animated into a video format. Rotating stall occurs because of the lack of air to fill in between the blade passages. As a result, the flow is shared unequally such that some blade passages have large flow rates while other blade passages record lower flow rates. The stall cells always rotate in the direction of the spinning rotor. Stall will run in the direction in which the angle of incidence is increased. For the rotating blades, it means that the stall cells moves backward relative to the rotating blades. This means that when the rotating blades are rotating clockwise as shown in Figure 6, the stall cells are rotating in an anti-clockwise direction relative to the blades. The rotation of the stall cells is in the same sense as the rotor but at a greatly reduced speed. These observations proved that the sliding mesh model serves as a good representation of flow in this complete fan configuration.

Blade 1 Blade 2 Blade 3 Blade 4 Blade 5 Blade 6 Blade 7 Blade 8 Blade 9 Blade 10 Blade 11

The computed streamlines of the flow leaving immediately aft of the impeller blades are shown in Figures 7(a) and 7(b) respectively. Flow separation with presence of shed vortices can be seen.

. Unsymmetrical streamlines due to effects of rotation Rotation direction Detaching flows at blades with presence of vortices

Figure 5. Contour plot of vorticity magnitude at blade arrangement (3 D view) for 0.95m3/s, 2500rpm

Figure 6. Contour plot of static pressure for blade passages (Top view) for 0.95m3/s, 2500rpm

Figure 7. Comparison between streamline aft blade arrangement in (a) top view and (b) zoom in 3 dimensional view for 0.95m3/s, 2500rpm

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Figure 8 shows the computed streamlines emanating from the blade unit to the outlet along the extension tube at the outlet

Computed skin friction lines are plotted from the computed flow variables which insight gained from plot of skin friction lines throws light on the flow separation from the blade surfaces. Figures 9(a) and 9(b) show the plot of computed skin friction lines on the pressure and the suction side of a blade respectively. It has been observed that different blades have different patterns of skin friction lines due to presence of rotating stall, which creates an uneven fluid distribution.

Figure 9. Plot of skin friction line on (a) pressure side and (b) suction side of Blade 11 for 0.95m3/s, 2500rpm

Figure 10 compares the impact of using different turbulence models on the variation of gauge total pressure at the outlet against volume flow rate for single inlet fan operating at 2500rpm. Figure 11 compares the plot of computed total efficiency against volume flow rate with the experimental results measured by the vendor of the fans. Both the experimental results and flow simulation results showed the maximum total efficiency when the volume flow rate is

1.9m3/s.

Figure 8. Streamline aft blade for 0.95m3/s, 2500rpm

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W.K. Ng and M. Damodaran 0 500 1000 1500 2000 2500 0 0.5 1 1.5 2 2.5 3 3.5 4

Volum e Flow Rate (m 3/s)

G a u g e T o ta l P re s s u re a t o u tl e t (P a ) Standard K Epsilon Experimental Result Standard K Omega RSM K Epilson Realizable

0.00E+00 1.00E+01 2.00E+01 3.00E+01 4.00E+01 5.00E+01 6.00E+01 7.00E+01 8.00E+01 0 1 2 3 4

Volume Flow Rate (m3/s)

T o ta l E ff ic ie n c y Standard K Epsilon Experimental Result Standard K Omega RSM K Epilson Realizable

The total efficiency of the single inlet fan at 1.9m3/s for flow simulation result is 73.8 %

which is close to the experimental result of 76.7 %. Figure 12 shows the plot of computed gauge total pressure at outlet against volume flow rate for single inlet fan running at different rpm. The surge line connects the surge points corresponding to different rpm. Effective fan operation should not exceed the surge line but should be near to the surge line to avail the maximum efficiency.

Figure 10. Plot of gauge total pressure at outlet against volume flow rate for single inlet fan running at 2500rpm for different turbulence models

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W.K. Ng and M. Damodaran -500 0 500 1000 1500 2000 2500 3000 3500 4000 4500 0 1 2 3 4 5 6

G a u g e T o ta l P re s s u re a t o u tl e t (P a ) 1200rpm 1800rpm 2500rpm 2900rpm 3600rpm

Surge line

.

5.2 Computed Steady State Flow Fields in Double Inlet Fan

Figure 13 shows the computed gauge total pressure contours at the outlet for the

extension tube for a flow rate of 1.2m3/s and fan rotating at 2200rpm. Since the double inlet

fan has its top and bottom blade arrangements aligned midway to each other, there is

asymmetry in the pressure distribution at the outlet.

Figure 12. Plot of gauge total pressure at outlet against volume flow rate for single inlet fan running at different rpm with Standard K Epsilon Turbulence Model

Figure 13. Contour plot of gauge total pressure at outlet for 1.2m3/s, 2200rpm

Average gauge static pressure: 1737.845 Pa

Figure 14. Contour plot of gauge static pressure and blade arrangement on the impeller (3 D view) for

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The surface pressure distribution of the impeller blades as shown in Figure 14 also shows the presence of rotating stall cells. Computed surface pressure contours on the blade pressure and suction sides are shown in Figures 15(a) and (b) respectively.

A B C E D F G L J H I K

Hub line EF of Pressure side Midline CD of Pressure side Tip line AB of Pressure side Midline IJ of Suction side Hub line KL of Suction side Tip line GH of Suction side

Pressure side has higher gauge static pressure than suction side

Figure 15. Contour plot of gauge static pressure at (a) pressure side and (b) suction side of blade for 1.2m3/s, 2200rpm

(a) Pressure side (b) Suction side

Figure 16. Spanwise plot of static pressure along line on (a) pressure side, (b) suction side and (c) both sides of blade surfaces for 1.2m3/s, 2200rpm

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Figure 16(a) and (b) shows the surface pressure distribution at various spanwise stations near the hub, mid-span and tip stations shown in Figure 15(a). It can be seen that gauge pressure distribution does not vary significantly across the span of the blade. The spanwise plot of gauge static pressure along the blade surfaces at the three spanwise stations as shown in Figure 16(c) shows that the pressure side of blade has a higher gauge static pressure than the suction side of the blade. In addition, the gauge static pressure increases from C to D and I to J such that the blade edge farther away from the centre of rotation has a higher pressure. For any fan, the minimum pressure occurs at the centre of rotation of the fan wheel and the maximum pressure occurs just at the discharge side of the wheel. This means the rotating plate has a low-pressure region while the outlet has a high-pressure region.

A

A B

B

Figure 17(a) and (b) show the computed vorticity contours on the blade surfaces. The plot of the variation of the magnitude of vorticity along the lines A-A and B-B in the spanwise direction indicated in Figure 17(a)-(b) is shown in Figure 18.

Blade tips observed to have higher vorticity magnitude

A A

B B

(a) Pressure side of blade (b) Suction side of blade Figure 17. Contour plot of vorticity magnitude at (a) pressure side and (b)

suction side of blade for 1.2m3_{/s, 2200rpm}

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Higher vorticity regions are seen near the blade tip regions embedded into the impeller as in Figure 19. Figure 20 shows a plot of computed streamlines emanating from the blades and rotating impeller section and these are coloured according to the magnitude of the vorticity. As swirling flow takes in these regions, this implies that the blade tip shape also influences fan performance. According to Figure 21 and Figure 22, flow separation occurs at high

vorticity magnitude regions.

Figure 21 shows the plot of computed gauge total pressure at outlet against volume flow rate for double inlet fan running at 2200rpm.

0 200 400 600 800 1000 1200 1400 1600 1800 2000 0 1 2 3 4 5 6 7

G a u g e T o ta l P re s s u re a t O u tl e t (P a )

Flow Simulation Result Experimental Result

Figure 22 compares the plot of computed total efficiency against volume flow rate with the experimental results measured by the vendor of the fans. Both the experimental results and flow simulation results showed the maximum total efficiency when the volume

flow rate is 3.33.m3/s. The total efficiency of the double inlet fan at 3.33m3/s for flow

simulation result is 73.9 % which is close to the experimental result of 76.1 %.

Figure 21. Plot of gauge total pressure at outlet against volume flow rate for double inlet fan running at 2200rpm Figure 19. Contour plot of vorticity magnitude on

blades and rotating impeller plate for 1.2m3/s, 2200rpm

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W.K. Ng and M. Damodaran 0.00E+00 1.00E+01 2.00E+01 3.00E+01 4.00E+01 5.00E+01 6.00E+01 7.00E+01 8.00E+01 0 2 4 6 8

T o ta l E ff ic ie n c y

Flow Simulation Result Experimental Result

In order to compare the performance of a single and a double inlet centrifugal fan, both fans are run at 2200 rpm. Figures 23 and 24 compares respectively the computed and experimental variations of gauge total pressure and efficiency with volume flow rate.

0 200 400 600 800 1000 1200 1400 1600 1800 2000 0 1 2 3 4 5 6 7

G a u g e T o ta l P re s s u re a t O u tl e t (P a

) Double inlet fan flow simulation result_{Double inlet fan experimental result}

Single inlet fan flow simulation result Single inlet fan experimental result

0.00E+00 1.00E+01 2.00E+01 3.00E+01 4.00E+01 5.00E+01 6.00E+01 7.00E+01 8.00E+01 0 2 4 6 8

Double Inlet fan flow simulation result Double inlet fan experimental result Single inlet fan flow simulation result Single inlet fan experimental result

It can be seen that the double inlet fan generates a higher gauge total pressure than the single inlet fan for the same rpm and volume flow rate. Since the double inlet fan has two impellers instead of one impeller, it will generate a higher static pressure rise. The single inlet fan is not suitable for high volume flow rate as it can be seen from Figure 23 that the gauge total pressure at the outlet decreases drastically at higher volume flow rate as the single inlet fan has a smaller outlet area when compared with the double inlet fan. By having the same volume flow rate, the flow velocity through the outlet of the single inlet fan has to be higher than that of the double inlet fan, which will cause a drastic drop in gauge total pressure at the outlet. Figure 24 shows that the single inlet fan has a higher total efficiency than the double inlet fan despite giving out a lower gauge total pressure at the outlet. This is because the

Figure 22. Plot of total efficiency for double inlet fan running at 2200rpm

Figure 23. Plot of gauge total pressure at outlet against volume flow rate for single inlet fan and double inlet fan

running at 2200rpm

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single inlet fan has a much lower power requirement than the double inlet fan. To rotate the impellers at the same rpm, the double inlet fan requires more power as it is turning two impellers instead of one. The higher gauge total pressure at the outlet for the double inlet fan is not enough to compensate for a much higher power requirement. The single inlet fan has a much lower total efficiency at higher volume flow rate due to a sudden reduction of gauge total pressure at the outlet. Hence the single inlet fan has a higher total efficiency than the double inlet fan at low volume flow rates but a lower total efficiency at high volume flow rates. The double inlet fan gives a higher gauge total pressure at the outlet but has a lower total efficiency due to a higher power requirement. Hence, for industrial purposes, the double inlet fan is desirable when a higher gauge total pressure is desired. The single inlet fan is desirable at low volume flow rates as it gives a higher total efficiency.

5.3 Computed Flow Fields in Double Inlet Fan Using Airfoil Blade Sections

Bleier10 observed that centrifugal fans which use blades made up of airfoil sections

have the highest efficiencies of all centrifugal fans. Hence, it is possible to replace the blunt blade sections used in the single inlet and double inlet fans discussed in scetions 5.1 and 5.2 with airfoil blades as shown in Figure 25 so that improved efficiencies could be obtained. A computation was done for the same double inlet fan using an airfoil profile for the blade as specified by the vendor. The computed pressure contours at the outlet of the fan using blades with airfoil sections is shown in Figure 26.

The double inlet centrifugal fan using blunt edge blades outlined in section 5.2 has a gauge total pressure of 1737.845 Pa which is higher than the gauge total pressure at the outlet for the fan using airfoil section blades which is 1647.416 Pa for the same volume flow rate and rpm. The power required for rotation for the airfoil blade fan is lower than the blunt bladed fan. The total efficiency of the airfoil blade fan is 51.1 % which is slightly lower than the total efficiency of blunt bladed fan at 52.1 %. This discrepancy is due to the flow simulation result for airfoil bladed fan giving a lower gauge total pressure at the outlet when compared with experimental result. Figure 27 shows that for the fan using airfoil section blades, the flow

Figure 25. Plot of total efficiency for double inlet fan running at 2200rpm

Figure 26. Contour plot of gauge total pressure at outlet for 1.2m3/s, 2200rpm

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between the blade passages is shared more equally such that there is an almost equal distribution of gauge static pressure in each blade passages.

Blade 1 Blade 2 Blade 3 Blade 4 Blade 5 Blade 6 Blade 7 Blade 8 Blade 9 Blade 10 Blade 11 Detaching flows at blades with presence of vortices

Figure 28 shows the presence of flow separation and vortices between blade passages for the airfoil blade fan. Figures 29 and 30 show that the blunt edged bladed fan has a higher vorticity magnitude than that from the airfoil bladed fan.

Dimensionless plot of computed gauge total pressure vs. volume flow rate for airfoil bladed fan is compared with experimental data provided by the vendor. Figure 31 shows that the gauge total pressure at the outlet increases with decreasing volume flow rate until the surge point is reached for the airfoil blade fan, while the blunt bladed fan has a higher gauge total pressure than the airfoil bladed fan for the corresponding volume flow rate. Figure 32 compares the computed total efficiency vs. volume flow rate curves for the airfoil bladed fan and double inlet fan. Differences between flow simulation result and experimental results can be observed although the trends are very similar. It appears by comparing the experimental results that the total efficiency of the airfoil bladed fan is higher than that of the blunt bladed fan due to a lower power requirement. Comparing flow simulation results, the airfoil bladed

Figure 27. Contour plot of static pressure for blade passages (Top view) for 1.2m3/s,

2200rpm for ADA 450

Figure 28. Streamline aft blade arrangement in zoom in 3 dimensional view for 1.2m3/s,

2200rpm for airfoil blade fan

Figure 29. Contour plot of vorticity magnitude for double inlet fan blade for 1.2m3/s, 2200rpm

Average vorticity magnitude: 1864.045 1/s

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fan has a higher total efficiency than the blunt bladed fan at high volume flow rates. However, the blunt bladed fan can reach a higher maximum efficiency than the airfoil bladed fan. This is also observed by comparing the experimental results between the airfoil bladed fan and blunt bladed fan.

0 200 400 600 800 1000 1200 1400 1600 1800 2000 0 1 2 3 4 5 6 7

G a u g e T o ta l P re s s u re a t O u tl e t (P a )

Airfoil Blade fan Flow Simulation Result Double Inlet fan Flow Simulation Result Airfoil Blade fan Experimental Result Double Inlet fan Experimental Result

0.00E+00 1.00E+01 2.00E+01 3.00E+01 4.00E+01 5.00E+01 6.00E+01 7.00E+01 8.00E+01 0 1 2 3 4 5 6 7

Airfoil Blade Fan Flow Simulation Result Double Inlet Fan Flow Simulation Result Airfoil Blade fan Experimental Result Double Inlet fan Experimental Result

6 CONCLUSIONS

Industrial fan performance characteristics have been investigated using CFD on a sliding mesh model. Generally, the computed results appear to be reliable and similar to the experimental fan performance characteristics provided by the vendor. However, there are some differences in values of gauge total pressure at outlet and total efficiency due to the difference in setup and measurement between wind tunnel results and numerical simulated results. The airfoil bladed fan has a higher gauge total pressure at the outlet with less power required for rotation of the fan. Since airfoil blades have a lower level of vorticity than the blunt bladed fan, there are fewer vortices in the flow field around an airfoil bladed fan. The blunt bladed fan has a slightly higher gauge total pressure than the airfoil bladed fan for the corresponding volume flow rate. However, the total efficiency of the airfoil bladed fan is higher than that of blunt bladed fan due to a lower power requirement. Blunt bladed fan can achieve a higher maximum total efficiency than the airfoil bladed fan. The experimental result is obtained by running the ventilation fan at a certain rpm. The Fan Law is then applied to calculate the corresponding gauge static pressure at the outlet. The gauge dynamic pressure is calculated by using empirical formulae. This gauge dynamic pressure is then added to the gauge static pressure to obtain the gauge total pressure. Hence, the experimental result is based on the assumption of the applicability of the Fan Law and the formula for the calculation of gauge dynamic pressure. Moreover, the experimental result is obtained in a wind tunnel test, which is different from the flow simulation setup. This could explain why there is a difference between experimental and flow simulation result and this matter needs to

Figure 31. Plot of gauge total pressure at outlet against volume flow rate for airfoil blade fan and double inlet fan

running at 2200rpm

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be addressed in the future. Further work on the numerical simulation of the airfoil bladed fan is being done to address the discrepancies with the experimental data by refining the grids. The blade shape design optimisation and fan performance optimisation program will be developed on top of this simulation program using the notions of reduced order modelling techniques for design.

ACKNOWLEDGEMENT

This work is partially funded by Kruger Ventilation Industries, Singapore, which also provided inputs on their fan design geometries and experimental data for this work.

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http://www.krugerfan.com/centri_bdb.php

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