1. International Conference on Computer Aided Engineering (CAE-2013)
Department of Mechanical Engineering, IIT Madras, India
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NOISE PREDICTION OF ENGINE COOLING FAN WITH SHARK TEETH AND FIN ARRANGEMENT
E. Sandeepkumar1 , K. R. Anandakumaran Nair2, N. S. Ramanathan3
1 Design Engineer, Advanced Engineering, Lucas TVS Ltd., Padi, Chennai, Tamil Nadu, India, 600 050
2 Executive Director (Development), Lucas TVS Ltd., Padi, Chennai, Tamil Nadu, India, 600 050
3 Head - Advanced Engineering, Lucas TVS Ltd., Padi, Chennai, Tamil Nadu, India, 600 050
 sandeepkumar.engg@lucastvs.co.in
ABSTRACT
Automobile cooling system fans have predominantly been axial fans, due to their low cost, thickness, and ease of mounting. However, since flow resistances in front and back of the fan are very strong, air tends to flow diagonally at the fan outlet side. The fan does not operate as an axial flow fan. The purpose of this study is to reduce the noise of an axial flow fan by understanding the aerodynamic noise source distribution around a rotating fan blade by measuring the noise signal around the blade. Noise distribution over the fan blade is measured by microphone. It is found that the noise source is located near the rotating fan blade, especially around leading and trailing edges. The separation and reattachment of flow are observed near the leading edge. The tip vortices and vortex shedding are found near the trailing edge. So the generation of noise is mainly by the formation of vortex structure around the blade. The performance of a conventional axial flow fan and other fan, which have shark teeth like arrangement at the trailing edge of the blade and fins at the tip of the bade are compared. Then the effects of teeth and fin arrangement are investigated in detail. The model of a conventional engine cooling fan and a fan with teeth and fin arrangement is created using 3D product development software. Flow analysis is carried out using commercial Computational Fluid Dynamics analysis software. As a result, the fan with teeth and fin arrangement showed a considerable reduction in noise than the conventional fan at the operating speed.
Keywords: Axial Fan, Shark Teeth, Fin, turbulence, and trailing edge.
1. INTRODUCTION
Aerodynamic noise generated by axial fans is a major source of annoyance. Noise is undesirable or unwanted. Noise has become an important subject in the design, installation and operation of fans. The main causes of aerodynamically generated noise are:
 At the entry and exit of the fan, i.e., suction and exhaust noise,
 Rotation of blades through air or gas,
 Passage of blades through wakes,
 Turbulence of air,
 Shedding of vortices from blades, and
 Separation, stalling and surging.
The aerodynamic loads applied to the blades are highly unsteady. This loading forms a source system creating acoustic waves propagating in the surrounding medium to the observer. Several studies have shown that a significant share of the noise is directly linked to these forces [1, 2].
Previous work by Kameier, F. et. al. [3] investigating the effects of varying tip clearances on noise and performance showed the existence of a broadband noise source for large tip–casing clearances. This source appeared in the rotor wall pressure spectrum at about half the blade passing frequency (BPF) and radiated a fluctuating tonal component into the far field, the so-called tip clearance noise (TCN).
The uneven blade spacing of a rotor as a mean to reduce the tonal noise annoyance has been rigorously approached by Andrea Cattanei et. al. [4] resulting in an optimization procedure where the geometry generation and the related constraints have been kept distinct from the search of the optimal geometry. A general method of optimization based on the evaluation of the rotor interference function has been developed which may be applied to rotors of arbitrary blade number.
Efforts have been taken to reduce the tonal noise generated by axial fan. Changing the blade spacing resulted in performance reduction and changing the blade profile involves developing complete new design of axial fan. Hence a study was made to understand the shark teeth and shark fin kind of arrangement in the existing axial fan which is of 370 mm in diameter.
One fan with shark teeth arrangement at the leading edge was done in the 3D model and simulated, but there was a drastic reduction in the performance of the fan. Then the shark teeth arrangement was given to the trailing edge. This gave satisfactory results in simulation. Hence a sample was made and tested for its flow performance and noise. Another fan was taken and shark fin arrangement was given to the suction side of the blade. After obtaining satisfactory results in

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simulation, a sample was made and tested for its flow performance and noise.
Taking these two aspects of blade modifications into account the work was carried out to perform the 3D simulation for performance prediction and the experimental test results were obtained and correlated.
2. AIR FLOW TEST/SIMULATION
At the end of design process designed fan has to be evaluated assure its performance, efficiency, structural strength and the noise. Making a prototype for the evaluation of design increases the product cost and it is time consuming. Hence the evaluation of the design through computational simulation is desirable. Once simulation results are satisfactory prototype can be made to confirm the performance. Simulation results alone cannot be taken as the final result because the simulation tools have their own advantages and disadvantages. In some cases simulation may predict closer to the test results whereas in some cases it may have large deviations. So the correlation between test and the simulation results are inevitable for the design validation. In this work existing fan which has a 370 mm outer diameter with 8 blades is first tested and simulated to establish a correlation. Then it is compared with a 370 mm fan having shark teeth and another 370 mm fan having shark fin arrangement. In this study first the experiment set up is explained, then the simulation method is explained, then correlation between the test and simulation performance curve is established.
2.1. Experiment Setup
The Experimental setup is called air flow performance tester which has a blower connected to the 4.8m test chamber at one end and at the other fan model which is to be tested mounted. The airflow from the blower into the test chamber is controlled by opening and partial closing of a number nozzles which can be controlled by the knobs in control panel. The air flow rate and the pressure inside chamber can be monitored at the monitor. An image of the chamber is given in Fig. 1
Fig. 1: Experimental setup
2.1.1 Air Flow Testing
An existing axial fan with 370mm outer diameter is considered to establish correlation between test performance and simulation performance. The fan is designed to operate at 3400 rpm and 160 Pa pressure drop to produce a volumetric flow rate of 3287 cu.m/hr.
Fan is mounted in the test chamber at appropriate location. Then the blower at the other end is started to deliver air at the required flow rate. The air delivered by the blower can leave chamber only through the fan. Motor is powered to run at the corresponding design speed. The static pressure inside the chamber can be maintained at particular value by adjusting the flow rate from the blower. Rotating fan can only suck the air from the chamber and deliver it to the atmosphere. If we maintain the static pressure in the chamber at a particular value then the flow rate delivered by the blower is equal to flow rate delivered by the fan. The test set up has instruments in the chamber to measure static pressure inside the chamber and the flow rate delivered by blower.
To obtain the free delivery point of the fan that is flow rate at 0 Pa pressure, the static pressure inside chamber is maintained at 0 pa pressure and the flow rate delivered by the blower is measured. After obtaining the free delivery point, flow rate from blower is decreased systematically to obtain the performance curve at different pressure drops. The efficiency of the fan is obtained by ratio output to the input power. Then the performance curve of fan is plotted by considering the volumetric flow rate in the x axis and pressure & efficiency in the y axis.
2.1.2 Test Model Calculation
A model calculation to compute the test efficiency is shown below.
Input power = Voltage × Current
Output power = Operating Pressure × Test Flow Overall Efficiency = Output power/ Input power
Considering 160 Pa operating point (Existing Fan):
Input power = 13.5 × 41.18
= 555.93 W
Output power = 160 × (3287/3600)
= 146.39 W
Overall Efficiency = 26.33%
2.2. CFD Simulation Of The Fan
Generally CFD simulation is carried out in the fan to predict its performance and efficiency. Commercial 3D product development software is used for modeling.

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ANSYS 'ICEM CFD’ is used for finite volume meshing. And ANSYS CFX is used for the simulations. In the test setup chamber is inlet of fan and the outlet is atmosphere. In simulation both the inlet and the outlet chamber should be modeled as a fluid domain. The Blower is replaced by the mass flow rate or the pressure boundary condition at inlet.
2.2.1 3D Model For Simulation
To perform a structural or thermal analysis the component model alone is sufficient. But to perform a CFD simulation the component model alone is not sufficient. Static and Rotating fluid domains need to be modeled for the fan component analysis. So the rotating domain is first modelled by subtracting the fan from a disc having dimensions slightly greater than the fan. After various trials cylinder dimensions were concluded to have an increase in diameter of 10 mm and increase in thickness of 2 mm respectively on either side of the fan. Inlet and outlet also should be big enough to facilitate the simulation. After trying out various inlet and outlet domain shape, it was concluded that the inlet domain should have same dimensions like the test set up. The outlet domain also modelled with the same dimension as inlet. A shark teeth fan model for simulation is shown in Fig. 2. Dimensions of the inlet, outlet and rotating domain is shown in Fig. 3.
Fig. 2: Shark Teeth arrangement at trailing edge
Fig. 3: Dimensions of the inlet and outlet domain
2.2.2 3D Model Meshing
In CFD simulation meshing is a critical part which needs to be carried out in such a way that the number of nodes is less and the quality of the element is adequate. As the number of nodes increases the simulation time also increases. By grid independent study the number of nodes is fixed to be 1.3 million. Quality is maintained above 0.15. Meshing is broadly classified as two types, one is free meshing and the other is mapped meshing. In free meshing tetrahedral elements are used whereas in mapped meshing hexahedral elements are used. Even though mapped meshing reduces the node count significantly, it is time consuming in complicated geometry. Since our fan has fillets and curved surfaces it is laborious to do the mapped meshing so free meshing is preferred. Surface mesh in the existing fan is shown in Fig. 4. Section view of the rotating domain surface mesh is shown in Fig. 5. Surface mesh in the shark fin fan is shown in Fig. 6. Surface mesh in the fluid domain is shown in Fig. 7.
Fig. 4: Surface mesh in the existing fan
Fig. 5: Section view of rotating domain mesh
Fig. 6: Surface mesh in the fan with shark fins

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Fig. 7: Surface mesh of the domain
2.2.3 Simulation Procedure
The Free delivery point flow rate is obtained by defining 0 Pa static pressure boundary condition at inlet and outlet. Torque is measured at the fan surfaces. From the torque obtained, input power can be calculated. Then the fan efficiency is calculated by ratio of output to the input. Overall efficiency is obtained by multiplying the motor efficiency to the fan efficiency.
After getting the free delivery point the other points in the performance curve can be obtained by decreasing the flow rate value or increase the pressure rise value in the inlet boundary condition. The defined boundary conditions are shown in Fig. 8.
Fig. 8: Imposed boundary conditions
2.2.4 Post Processing
Convergence of the simulation is decided by the residual value of the mass continuity and the three momentums. The residual value is usually set as 1e-4 but in some cases where the rotating region are used, the residual values never comes less than 1e-3. Since there are large fluctuations in the residuals, convergence of the simulation is obtained by creating a monitor point. Monitor point is nothing but monitoring the desired output. If the desired value remains same in the successive iterations then the solution is converged. The simulation can be stopped and the results can be obtained.
In this study, pressure and mass flow rate are the desired values. In addition torque is also measured on the surface of the fan to find the power consumption and efficiency. The Fig. 9 below shows velocity vector plot.
Fig. 9: Velocity vector plot showing the direction of flow
2.3. Test/Simulation Correlation
Results obtained from the test and the simulation was compared and the performance curve was plotted. Flow and efficiency results were within 10% difference at the operating point of 160 Pa. At higher flow rate simulation under predicts the test performance whereas in lower flow rate simulation over predicts the test performance. The Comparison table and the performance curve for:
 Existing 370 mm fan (Table 1, Fig. 10)
 Shark teeth (Table 2, Fig. 11); and
 Shark fin fan (Table 3, Fig. 12) is given below.
Table 1: Existing fan test/simulation performance Existing 370 mm Fan Test vs. Simulation Correlation S.No Pressure Speed Flow Efficiency Flow Efficiency (Pa) (rpm) (cu.m/hr) (%) (cu.m/hr) % 1 0 3613 4288 0 4390.1 0 2 80 3575 3848 16.47 4082.1 13.4 3 160 3486 3287 26.33 3409.5 23.9 4 240 3423 2724 30.86 2819.9 29.4 5 320 3360 2241 32.46 2090.5 28.1 Test Results Simulation Results
Fig. 10: Existing fan test/simulation performance curve

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Table 2: Shark teeth fan test/simulation performance Existing 370 mm Fan vs. Shark Teeth Fan Simulation Correlation S.No Pressure Speed Flow Efficiency Flow Efficiency (Pa) (rpm) (cu.m/hr) (%) (cu.m/hr) % 1 0 3613 4390.1 0 4463.7 0 2 80 3575 4082.1 13.4 4250.4 13.5 3 160 3486 3409.5 23.9 3582.8 23.6 4 240 3423 2819 29.4 2996 29.4 5 320 3360 2090.5 28.1 2213.8 29 Existing 370 mm Fan Shark Teeth Fan
Fig. 11: Shark teeth fan test/simulation performance curve
Table 3: Shark fin fan test/simulation performance Existing 370 mm Fan vs. Shark Fin Fan simulation correlation S.No Pressure Speed Flow Efficiency Flow Efficiency (Pa) (rpm) (cu.m/hr) (%) (cu.m/hr) % 1 0 3613 4390.1 0 4483.6 0 2 80 3575 4082.1 13.4 4217.9 12.1 3 160 3486 3409.5 23.9 3700.4 21.8 4 240 3423 2819 29.4 3219.1 28.3 5 320 3360 2090.5 28.1 2245.6 26.6 Existing 370 mm Fan Shark Fin Fan
Fig. 12: Shark fin fan test/simulation performance curve
2.3.1 Test Model Calculation
A model calculation to compute the simulation efficiency is shown below:
Input power = (2πNT/60)
*T is obtained on the fan surfaces from simulation
Output power = Pressure × Simulation Flow
Fan Efficiency = Output power/ Input power
η (Overall) = η (Fan Efficiency) * η (Motor)
Considering 160 Pa operating point (Existing Fan):
Motor Efficiency is 70% (Determined from testing)
Torque = 1.2164 Nm
Input power = ((2 × 3.14 × 3486 1.5764)/60)
= 443.84 W
Output power = 160 × (3409.50/3600)
= 151.53 W
Fan Efficiency = 34.14%
Overall Efficiency = 34.14 × 70
= 23.9%
3. NOISE TEST IN ANECHOIC ROOM
The aero acoustic loads applied in fan blades are highly unsteady. These loads form a base in generating acoustic waves that propagate to the observer in surrounding.
The main sources of generation are:
 Turbulent in-flow
 Self-noise (turbulent or laminar boundary layers, boundary layer separation)
 Trailing edge noise
 Secondary flows at the blade borders [5]; and
 Noise due to the rotor stator interaction [6, 7]
Among the several mechanisms of noise generation, nature of inflow is considered as major. But the dominant source cannot be distinguished.
Fig. 13: Semi anechoic room for noise test
The axial fan is placed in the centre of the anechoic room shown in Fig. 13. Microphone is placed in front of the fan at a distance of 1 m. Noise is measured in dB at different voltages with free flow condition i.e. 0 Pa. The room is considered to be semi anechoic because the floor is not with sound absorbing material.

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4. RESULT AND DISCUSSION
The flow simulations were done to predict the performance and efficiency of the shark fin and shark teeth fan. Since there was no drastic difference in performance and efficiency for the two fans, a prototype was made and tested. Both the fans showed good correlation with simulation and test at the operating point. Noise test was done experimentally further to check the effect of Shark teeth and shark fin arrangement on the fans. Noise test results of Existing 370 mm fan (Table 4), Shark teeth fan (Table 5) and Shark fin fan (Table 6) is given below.
Table 4: Existing fan noise test
Voltage (V)
Existing Fan Noise Test (dB)
13.5
79.4
13
79.12
12
77.13
Table 5: Shark teeth fan noise test
Voltage (V)
Shark Teeth Noise Test (dB)
13.5
76.25
13
75.8
12
74.8
Table 6: Shark fin fan noise test
Voltage (V)
Shark Fin Noise Test (dB)
13.5
76.4
13
75.2
12
73.5
The acoustic measurements were satisfactory. The results proved that, the tip vortices and vortex shedding can be minimized at the trailing edge by the shark teeth arrangement and the turbulent inflow and self-noise can be minimized by the shark fin arrangement.
The shark teeth arrangement showed a maximum of 3.32 dB decrease in noise at 13 V and a minimum of 2.33 dB at 12 V.
The shark fin arrangement showed a maximum of 3.92 dB decrease in noise at 13 V and a minimum of 3 dB at 13.5 V.
5. CONCLUSION
Even a moderate increase in performance may result in a significant increase of the noise level. The designer works with the blade geometry changing sweep and lean angles (i.e. changing the blade profiles stacking) or blade number and spacing to reduce the tonal noise.
These solutions often have an impact on performance or require a complete blade redesign. Hence adding a Shark teeth and Shark fin kind of arrangement to the existing fan to study the performance and noise was taken up in this work.
As a mean to reduce the tonal noise, the existing bade was optimised with Shark teeth and fin arrangement. Based on the optimisation, two axial fans were realised and tested for its air flow performance in the air flow chamber and tested for noise in the semi anechoic room. Even though there was a small reduction in fan performance, there was a significant reduction in noise. The two types of arrangements are in consideration to be implemented for future design of axial fans.
6. ACKNOWLEDGEMENT
Simulation part of the work was supported by “The thermal turbo machines laboratory, Department of mechanical engineering, IIT Madras”. The authors express their appreciation to Dr. B.V.S. Prasad of IIT Madras. We gratefully acknowledge all his support.
7. REFERENCES
[1] Baad PK. "Effects of acoustic loading on axial flow fan noise generation," Noise Control Engineering 1977; 8:5–15.
[2] Wen-Shiang Chiu, Lauchle GC, Thompson DE. "Subsonic axial flow fan noise and unsteady rotor force,‖ Journal of Acoustic Society of America 1989; 85(2):641–7.
[3] Kameier, F. andNeise,W. "Experimental study of tip clearance losses and noise in axial turbo machinery and their reduction," Trans. ASME J. Turbomach., 1997, 119, 460–471.
[4] Andrea Cattanei, Riccardo Ghio, Alessandro Bongioviet, "Reduction of the tonal noise annoyance of axial flow fans by means of optimal blade spacing," ScienceDirect, Applied Acoustics 68 (2007) 1323— 1345.
[5] Fukano T, Kodama Y, Takamatsu Y. "The effect of the tip clearance on the noise of low pressure axial and mixed-flow fans," Journal of Sound and Vibration 1986; 105:299–308.
[6] Morfey CL. "Sound generation in subsonic turbomachinery," Trans ASME 1970:450–8.
[7] Chen MZ, Wu XH, "Vortex simulation of rotor/stator interaction in turbomachinery," Trans ASME 1999;121:358–64.