Journal of Thermal Science Vol.20, No.6 (2011)

495502

DOI: 10.1007/s11630-011-0501-6

Article ID: 1003-2169(2011)06-0495-08

Unsteady Flow Condition of Contra-Rotating Small-Sized Axial Fan T. Shigemitsu1, J. Fukutomi1, Y. Okabe2, K. Iuchi2 and H. Shimizu2 1. Institute of Technology and Science, The University of Tokushima, 2-1 Minamijyosanjima-cho, Tokushima 770-8506, JAPAN 2. Graduate School of Advanced Technology and Science, The University of Tokushima, 2-1 Minamijyosanjima-cho, Tokushima 770-8506, JAPAN © Science Press and Institute of Engineering Thermophysics, CAS and Springer-Verlag Berlin Heidelberg 2010

Small-sized axial fans are used as air cooler for electric equipments. But there is a strong demand for higher power of fans according to the increase of quantity of heat from electric devices. Therefore, higher rotational speed design is conducted, although, it causes the deterioration of efficiency and the increase of noise. Then, the adoption of contra-rotating rotors for the small-sized axial fan is proposed for the improvement of performance. In the case of contra-rotating rotors, it is necessary to design the rotor considering the unsteady flow condition of each front and rear rotor. In the present paper, the fan performance of the contra-rotating small-sized axial fan with 100mm diameter at a designed and a partial flow rates is shown, and the unsteady flow conditions at the inlet and the outlet of each front and rear rotor are clarified with unsteady numerical results. Furthermore, the relation between the performance and the unsteady flow condition of the contra-rotating small-sized axial fan is discussed and the methods to improve the performance are considered.

Keywords: Small-sized axial fan, Contra-rotating rotor, Performance, Unsteady flow condition

Introduction Because of spread of cloud computing, establishment of ubiquitous networking society and the increase in the rate of electric parts in machines, power consumption in data centers, IT devices and machines have been increasing significantly. In the view of the issue of global warming and energy savings, there is a strong demand for the reduction of power consumption in above facilities and equipments [1]. Electrical power used for the cooling of the IT devices for data centers is huge as the same as that used for the IT devices themselves in data centers. Small-sized axial fans are used as air coolers for electric equipments i.e. laptop, desk top computers and servers. There is a strong demand for higher power of

fans according to the increase of quantity of heat from electric devices. However, the increase of the power with the increase of the fan diameter is restricted because of the limitation of the space. Therefore, higher rotational speed design is conducted, although it causes the deterioration of the efficiency and the increase of noise. On the other hand, lower rotational speed design [2] and advantages on the performance of the contra-rotating fans and pumps are verified by experimental results [3, 4]. Then, the adoption of contra-rotating rotors for small-sized fans is proposed for the improvement of the performance. In the case of contra-rotating rotors, the axial space becomes larger than conventional small-sized axial fans. However, it is an adequate choice to apply the contra-rotating rotors for small sized-fans because the

Received: May 2011 T. Shigemitsu: Associate Professor, Dr. Eng. This research is supported by Japan Science and Technology Agency and University of Tokushima and Komiya research aid. www.springerlink.com

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Nomenclature Dh

diameter at the hub (mm)

Dt L Nf Nr Pt P

diameter at the tip (mm) shaft power (W) rotational speed of front rotor (min-1) rotational speed of rear rotor (min-1) total pressure (Pa) fan static pressure (Pa)

axial space can be ensured in electrical devices as compared to that of the radial space. In the case of contra-rotating rotors, it is necessary to design the rotor considering the unsteady flow condition [5]. Furthermore, it is important to clarify the influence of the wake from the front rotor to the rear rotor on the performance and pressure interaction between front and rear rotors [6]. On the other hand, the conventional design method and the theory for the turbomachinery should be modified for small-sized axial fans because small-sized axial fans applied to electrical devices belong to an extremely small size field in turbomachinery [7]. Therefore, there is a strong demand to establish the design method for smallsized axial fans. The performance and the circumferential averaged flow conditions at the designed flow rate and the partial flow rate were clarified by the previous research [8, 9]. In the present paper, the performance curves of the contra-rotating small-sized axial fan with 100mm diameter are compared with the unsteady numerical analysis results to verify the validity of the unsteady numerical analysis results. After that, the unsteady flow conditions at a designed flow rate and a partial flow rate are clarified by unsteady numerical analysis results. Furthermore, the relations between the performance and flow conditions were discussed by the unsteady numerical analysis results, and methods to improve the performance of contra-rotating small-sized axial fan are considered.

Experimental Procedure and Numerical Analysis Conditions Experimental apparatus and methods The rotor and the primary dimensions of a contrarotating axial fan (RRtype) are shown in Fig.1 and Table 1 respectively. The hub tip ratio was Dh/Dt=45 mm/98 mm, the designed flow rate was Qd=0.016 m3/s and fan static pressure at the design point was PdRR=14.7 Pa for RRtype with the same fan static pressure of each front and rear rotor. The rotational speed of front and rear rotors of RRtype was Nf=Nr=1780 min1. In this research, an aerofoil blade was used because there was a report

fan static pressure at the designed flow rate (Pa) Q flow rate (m3/s) Qd designed flow rate (m3/s) Greek letter total pressure efficiency (%)  rotation angles of front rotor f rotation angles of rear rotor r

Pd

that implied an advantage of the aerofoil blade for the small-sized axial fan [10], however a circular-arc blade was generally used for small-sized axial fans. RRtype rotor used in the experiment was changed a little from the rotor for the numerical analysis in the blade thickness, because the thickness of the rotor for the numerical analysis was thin and lack of strength for using it for the

Fig.1 Contra-rotating small sized axial fan (RRtype) Table 1 Primary dimensions of RRtype Diameter[mm]

Front Rotor (RRtype-exp.)

Rear Rotor (RRtype-exp.)

Hub

Mid

Tip

45

72

98

Blade Number

4

Blade Profile

NACA 4409

Solidity

1.196

0.496

0.29

Stagger Angle

44.67°

61.09°

68.15°

Blade Number

5

Blade Profile

NACA 4412

Solidity

0.91

0.447

0.288

Stagger Angle

56.73°

64.54°

69.60°

Blade Number Front Rotor (RRtype-cal.)

Rear Rotor (RRtype-cal.)

4

Blade Profile

NACA 4406

Solidity

1.245

0.508

0.308

Stagger Angle

42.40°

60.11°

67.21°

Blade Number

5

Blade Profile

NACA 4409

Solidity

0.91

0.447

0.288

Stagger Angle

55.92°

63.91°

68.87°

T. Shigemitsu et al.

Unsteady Flow Condition of Contra-Rotating Small-Sized Axial Fan

experiment was worried. In this paper, the unsteady flow conditions at the designed flow rate Qd=0.016 m3/s and the partial flow rate Q=0.0096 m3/s were focused. Fig. 2 shows the schematic diagram of the experimental apparatus for RRtype. The experimental apparatus was designed based on the Japanese Industrial Standard and the air blown in the test section passed the rotor, chamber, measurement duct and booster fan and was blown out into the ambient atmosphere. The fan static pressure (P) was measured by the pressure difference between static holes in the downstream of the rotor installed at the chamber and ambient air. Furthermore, the rotational speed was controlled by the servo motor and the flow rates were measured by the orifice meter set at the measurement duct. The pressure curves from the cutoff flow rate to the large flow rate were investigated in the experiment with the constant rotational speed Nf=Nr=1780 min1 for RRtype.

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of the rotor for RRtype. At the inlet boundary, the uniform velocity was given and the constant pressure was given at the outlet boundary condition. The coupling between the front and the rear rotors was accomplished by the sliding mesh. The standard wall function and k-turbulence model were used. The unsteady numerical flow analysis was conducted at the designed flow rate Qd=0.016m3/s and at the partial flow rate Qd=0.0096m3/s. The time step number per one rotor rotation was 200 and the time step was t=1.6854×10-4 s. The data of one rotor rotation were obtained after 6 rotor rotations in unsteady numerical analysis.

Fig.3 Numerical analysis grids

Experimental and Numerical Results

Fig.2 Experimental apparatus

Numerical analysis conditions The commercial software ANSYS-Fluent was used to investigate the flow condition which couldn’t be measured by the experiment. In the numerical analysis, the numerical model which was almost the same with the experimental apparatus was used and three dimensional unsteady numerical analysis was conducted. The numerical grids used for the numerical analysis are shown in Fig.3. The numerical domains comprised the inlet, rotor, chamber and outlet duct regions. The numerical grid numbers were 218,039 for the inlet region, 667,135 for the chamber region and 39,875 for the outlet duct region. The numerical grid numbers for the rotor region were 3,613,381 for RRtype. The numbers of nodes along the front and the rear rotor blades at the hub were 150 nodes and 80 nodes. The numbers of nodes of blade-to-blade of the front and the rear rotors were 35 nodes and 30 nodes. The numbers of nodes from hub to tip of the front and the rear rotors were 110 nodes and 100 nodes. The tip clearance was kept 1mm as the same with the experimental apparatus in the numerical analysis and the number of nodes from the blade tip to the casing was 7 nodes. The numerical grids over 150,000 were ensured at the tip clearance. The y+ was 15 near the hub

Performance curve of RRtype Figure 4(a) shows fan static pressure curves of RR type obtained by the experiment. In Fig.4(a), the fan static pressure obtained by the unsteady numerical analysis is also given to compare with the experimental results. The total pressure efficiency of RRtype obtained by the unsteady numerical analysis at the designed flow rate Qd=0.016 m3/s and the partial flow rate Q=0.0096 m3/s are shown in Fig.4(b). Moreover, the total pressure efficiency of each front and rear rotor is also shown in Fig.4(b). In the unsteady numerical analysis, the static pressure was obtained at the same position in the experiment and the 200 static pressure data of one rotor rotation were averaged. The total pressure efficiency was evaluated by the following equation. PQ  t (1) L Pt is total pressure, Q is the flow rate and L is the shaft power. The total pressure of each front and rear rotor was evaluated by the averaged data of one rotor rotation at the corresponding sectional areas. Further, the sectional area averaged total pressure of each front and rear rotor was obtained by the total pressure difference between inlet of the numerical domain and 6 mm downstream of the trailing edge at the hub of the front rotor for the front rotor, and 6 mm downstream of the trailing edge at the hub of the front rotor and the static pressure measurement position in the chamber for the rear rotor. It could be found in Fig.4(a) that the fan static pressure increased

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Fig.4(a) that the unsteady numerical analysis results of RRtype showed the quantitative accordance with the experimental results, and the unsteady numerical analysis results represented the appropriate data. Then, in order to clarify the unsteady flow condition and the influence of it on the performance, we investigated the unsteady flow condition from the numerical analysis results. Unsteady flow condition of RRtype

Fig.4

Performance of RRtype

according to the decrease of the flow rates and the pressure curves showed the stable negative curve from the experimental results. On the other hand, unsteady numerical results represented the qualitative tendency of the experimental results that the fan static pressure increased at the partial flow rate, although the results obtained by the unsteady numerical analysis were slightly higher than that of the experimental results. The total pressure efficiency of RRtype was =41.2% at the designed flow rate Qd=0.016 m3/s. Moreover, the total pressure efficiency of each front and rear rotor was f=41.7% and r=40.8% at the designed flow rate Qd=0.016 m3/s. It could be found that the total pressure efficiency of the rear rotor was slightly lower than that of the front rotor. On the other hand, the total pressure efficiency decreased (=30.5%) at the partial flow rate Qd=0.0096 m3/s. Furthermore, the total pressure efficiency of the rear rotor at the partial flow rate Qd=0.0096 m3/s became lower (r=29.2%) than that of the front rotor (f=32.0%). It was confirmed from

The meridional velocity vectors of each front and rear rotor at the designed flow rate Qd=0.016 m3/s are shown in Figs.5 and 6. The rotational direction of the front rotor is the back side of the paper and that of the rear rotor is the front side of the paper. f and r represent rotation angles of each front and rear rotor leading edge from the meridional plane. f=0 deg and r=0 deg correspond to the circumferential position that the leading edge of each front and rear rotor is in accord with the meridional plane. The relative circumferential position of each front and rear rotor is that the rear rotor rotation angle is r=0deg when the front rotor rotation angle is f=18deg. Focused on the flow condition of the front rotor before the front rotor traveling to the meridional plane as shown in Fig.5(a), velocity was low near the boundary layer of the casing wall and at the inlet due to the influence of the sudden contraction of the flow passage at the inlet (not bellmouth). Moreover, leakage flow from the front rotor tip occurred when the front rotor was passing the meridional plane in Fig.5(b). After that, the leakage flow from the front rotor tip decayed when the front rotor passed the meridional plane in Figs.5(c),(d). It was found that the vortex caused by the leakage flow repeated the change with the front rotor rotation from these results. On the other hand, the rear rotor showed the similar tendency with that of the front rotor that the vortex caused by the leakage flow repeated the change with the rear rotor rotation. Fig. 7 shows the meridional velocity vectors at the designed flow rate Qd=0.016m3/s from the view point that the whole front and rear rotors can be seen. It was found that there was a low velocity region near the tip of the rotor caused by the leakage flow and the boundary layer near the casing wall. This low velocity region accounted for 1/9 of the flow passage in radial direction and it was clarified that the influence of the leakage flow from the tip clearance and the frictional loss of the casing was large for the contra-rotating small-sized fan. Furthermore, the separation at the inlet corner of the casing and the hub of the front rotor were observed in Fig.7. Therefore, it was important to increase the radius of the corner at the inlet to suppress the separation at the corner. The meridional velocity vectors of each front and rear rotor at the partial flow rate Q=0.0096 m3/s are shown in Figs.8 and 9. The rotational directions of front rotor and

T. Shigemitsu et al.

Unsteady Flow Condition of Contra-Rotating Small-Sized Axial Fan

499

(a) f=0, r=18.0 [deg]

(a) f=18.0, r=0 [deg]

(b) f=25.2, r=7.2 [deg]

(b) f=36.0, r=18.0 [deg]

(c) f=50.4, r=32.4[deg]

(c) f=54.0, r=36.0 [deg]

(d) f=75.6, r=57.6[deg]

(d) f=72.0, r=54.0 [deg]

Fig.5 Meridional velocity vectors of front rotor, Qd=0.016 [m3/s]

Fig.6 Meridional velocity vectors of rear rotor, Qd=0.016 [m3/s]

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(a)f=0, r=18.0 [deg]

(a) f=0, r=-18.0 [deg]

(b) f=25.2, r=7.2 [deg]

(b) f=25.2, r=7.2 [deg]

Fig.7 Meridional velocity vectors, Qd=0.016 [m3/s]

rear rotor are the same as shown in Figs.5 and 6. Furthermore, f, r and the relative circumferential position of each front and rear rotor are the same in Figs.5 and 6. It was found that there were the vortexes related to the leakage flow in both front and rear rotors as shown in Figs.8 and 9, however, the vortex region at the partial flow rate Q=0.0096 m3/s was larger than that at the designed flow rate Qd=0.016 m3/s in Figs.5 and 6. Figure 10 shows the meridional velocity vectors at the partial flow rate Q=0.0096 m3/s from the view point that whole front and rear rotors can be seen. It was confirmed that the vortex region at partial flow rate Q=0.0096 was larger than that of the design flow rate Qd=0.016 m3/s. This would be caused by the increase of the negative pressure gradient in axial direction at the partial flow rate Q=0.0096 m3/s. As a result, the leakage flow rate from the tip clearance increased and the boundary layer was developed. Then, the back flow occurred near the casing. The back flow region occupied about 30% of the flow passage in radial direction. The velocity between the front and the rear rotors near the casing decreased due to the vortex and the boundary layer near the casing at the partial flow rate Q=0.0096m3/s. The decrease of the velocity between front and rear rotors near the casing would influence the flow condition

(c) f=50.4, r=32.4[deg]

(d)f=75.6, r=57.6[deg]

Fig.8

Meridional velocity vectors of front rotor, Q=0.0096 [m3/s]

T. Shigemitsu et al.

Unsteady Flow Condition of Contra-Rotating Small-Sized Axial Fan

(a) f=18.0, r=0 [deg]

(b) f=36.0, r=18.0 [deg]

501

(a) f=0, r=18.0 [deg]

(b) f=25.2, r=7.2 [deg]

Fig.10 Meridional velocity vectors, Q=0.0096 [m3/s]

of the rear rotor and cause the decrease of the rear rotor efficiency as shown in Fig.4(b). Therefore, it was important to keep the uniform flow condition at the inlet of the rear rotor by the modification of the front rotor design and the casing.

Concluding Remarks

(c) f=54.0, r=36.0 [deg]

The unsteady flow conditions of the contra-rotating small-sized axial fan especially focused on the meridional velocity were investigated by the unsteady numerical analysis. The performance of unsteady numerical results showed the good agreement with that of the experimental results. The meridional velocity vectors varied with the rotor rotation, and vortexes near the casing were influenced significantly by the leakage flow. The rear rotor performance was deteriorated by the non-uniform unsteady flow condition from the front rotor. Therefore, it was important to keep the uniform flow condition at the inlet of the rear rotor by the modification of the front rotor design and the casing.

Acknowledgements (d) f=72.0, r=54.0 [deg]

Fig.9

Meridional velocity vectors of rear rotor, Q=0.0096 [m3/s]

The authors wish to show our special thanks to the supports by the Komiya research aid, the project research

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aid from The University of Tokushima and Japan Science and Technology Agency.

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JSME/KSME Fluids Engineering Conference, Nagoya, pp.14531548. Sanders, A, J., Papalia, J and Fleeter, S., (2002), MultiBlade Row Interactions in a Transonic Axial Compressor: Part I- Stator Particle Image Velocimetry (PIV) Investigation, ASME Journal of Turbomachinery, Vol.124, pp. 1018. Shigemitsu, T., Fukutomi, J and Okabe, Y., (2010), Performance and Flow Condition of Small-sized Axial Fan and Adoption of Contra-Rotating Rotors, Journal of Thermal Science, Vol.19, No.1, pp.16. Shigemitsu, T., Fukutomi, J., Okabe, Y and Iuchi, K., (2010), Performance and Flow Conditions of Contra- rotating Small-sized Axial Fan, Proceedings of ASME 2010 3rd Joint US-European Fluids Engineering Summer Meeting, Montreal, FEDSM-ICNMM2010-30418. Shigemitsu, T., Fukutomi, J., Okabe, Y and Iuchi, K., (2010)., Performance and Flow Condition of Contrarotating Small-sized Axial Fan at Partial Flow Rate, International Journal of Fluid Machinery and Systems, Vol.3, No.4, pp.271278. Ito, T., Minorikawa, G,. Nagamatsu, A and Suzuki, S., (2006)., Experimental Research for Performance and Noise of Small Axial Flow Fan(Influence of Parameter of Blade), Transactions of the JSME (in Japanese), Vol. 72, No. 715, pp.670677.