Performance Enhancement of the In-Line Fan Equipped with the Guiding Vane and the Tail Body

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Research Article

Performance Enhancement of the In-Line Fan Equipped with

the Guiding Vane and the Tail Body

Hung-Cheng Yen,

1,2

Wen-Syang Hsu,

1

and Sheam-Chyun Lin

3

1_{Department of Mechanical Engineering, National Chiao Tung University, 1001 Ta Hsueh Road, Hsinchu 300, Taiwan}

2_{Department of Energy Equipment System Engineering, Industrial Technology Research Institute, Hsinchu 300, Taiwan}

3_{Department of Mechanical Engineering, National Taiwan University of Science and Technology, Taipei 106, Taiwan}

Correspondence should be addressed to Sheam-Chyun Lin; sclynn@mail.ntust.edu.tw Received 10 November 2013; Accepted 14 January 2014; Published 25 February 2014 Academic Editor: Wen-Hsiang Hsieh

Copyright © 2014 Hung-Cheng Yen et al. This is an open access article distributed under the Creative Commons Attribution License, which permits unrestricted use, distribution, and reproduction in any medium, provided the original work is properly cited.

This integrated numerical and experimental study intends to enhance the performance of an in-line fan with the implement of the guiding vane and the tail body. At first the flow flied associated with the original in-line fan is simulated and analyzed within the framework of CFD code Fluent, in which the finite volume method is applied. Next, the guiding vane is constructed based on the calculated flow characteristics, and attached in the downstream of rotor to smoothen the flow pattern. An appropriate guiding vane with high-performance and low-noise features can be achieved after several design iterations. In addition, the tail body connected to the motor is introduced for further enhancing the fan performance by reducing the sizes of wake and reversed flow behind the hub. Thereafter, to manufacture the mockup for experimental verification, the modified fan with guiding vane is plotted in the CAD/CAM format for mockup fabrication via the rapid-prototype technique. Moreover, a set of relations correlating the performance and noise of this fan prototype are executed inside AMCA test chamber and semianechoic chamber, respectively. Consequently, the feasibility of design scheme and numerical system can be verified according to these experimental results. In summary, this work provides a systematic scheme for designing and analyzing the in-line fan.

1. Introduction

In-line fans, also called axial-centrifugal fans or tubular centrifugal fans, are used extensively in the duct work for their advantage of easy installation in the pipe system as the axial-flow fan. At the same time, they generate bigger static pressure than the axial-flow fans of the same rotor diameter and rotational speed, even though they are not as much as centrifugal fans. In general, their efficiencies are inferior to those of axial-flow and centrifugal fans. Thus, the demand on the performance enhancement of in-line fan exists and motivates this research.

In-line fan is, also called axial-centrifugal fans or tubular centrifugal fans, composed of a centrifugal impeller installed inside a cylindrical, barrel-shaped, or even square housing. Thus, the generated static pressure is higher than that of axil fan, while inlet and outlet flows are in the same axial direc-tion. The above features fit with the need for transporting

flow in a pipe line perfectly. However, the outflow at the

centrifugal wheel is in the radial diction after a 90∘turn at

the wheel inlet; so another 90∘turn is needed to bring it back

to the axial-flow direction. This extra 90∘turn results in the

unfavorable flow patterns and downgrades the fan efficiency significantly. To compensate these drawbacks, a guide vane is attached after the centrifugal wheel for smoothing this direction transition and improving the efficiency. Moreover, a backward-inclined centrifugal rotor is usually used for enhancing the static-pressure gain for covering the system resistance in the pipe line. Nevertheless, user’s request on the compact housing makes the designs of barrel-shape housing and the guide vane very difficult. Thus, the challenging design task induced by this complicate flow pattern becomes the topic and motivation of this research paper.

The famous textbook on fan by Eck [1] introduced the

characteristics of different fan types and estimated their efficiencies using the related fan design parameters. He

Volume 2014, Article ID 817637, 9 pages http://dx.doi.org/10.1155/2014/817637

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also set a solid foundation for turbomachines by deriving the relation of energy transfer between fluid and impeller.

Regarding the fan design scheme, Lin and Tsai [2–4] utilized

an inverse cascade method to design centrifugal and axial-flow fans. The velocity triangles are calculated with the inputs of operating condition, specified axial velocity, and head distribution. Thereafter, the appropriate airfoil shapes and chord lengths were selected to fit the input operating condi-tions. To validate this inverse design method, the mockup of designed fan was manufactured by the CNC machine to carry out the performance test. With the above experimental and numerical efforts, they established a systematic, reliable, and user-friendly design scheme for small fans.

Klujszo et al. [5] design several guide vanes for the air

cleaner. The results show that the static pressure of air drops with increasing swirling motion while increasing turning angle; furthermore, a reduction on pressure drop is observed

by setting a 60∘ turning angle to turn the flow gradually.

Thakker and Dhanasekaran [6] carried out and examined

the numerical simulation for the effect of guide vane shape. In this case, when vortex theory is applied to improve the efficiency, three-dimensional inlet and downstream guide vanes have been designed. As a result, CFD model can predict the experimental results accurately, and the efficiency of the turbine is improved about 4.5% by employed 3D guide vanes.

Wang et al. [7] analyses the effect of guide vanes, and

the results show that vortex phenomenon appears at wall of guide vanes. While the vortex phenomenon appears, it not only decreases flow rate but also losses some kinetic energy. Furthermore, the proper setting angle of guide vane would

improve these nonuniform conditions. Li and Quan [8] use

different curvature of guide vanes by applied trimmed airfoil to increase efficiency of hydraulic turbine. The result shows that negative curvature guide vane improves the pressure distribution on both pressure side and suction sides. And a severe decrease on flow rate is observed for increasing its inlet angle.

Regarding the number of guiding vanes, Liou et al. [9]

used the laser-Doppler velocimetry (LDV) to observe flow field in a side-dump combustor. In this study, they found that guide van can diminish the friction loss associated with flow separation of the no-vane case. Nevertheless, the guide-vane itself also introduces a friction loss by introducing more guide-vanes. Consequently, the one-vane case is identified to generate the minimum friction loss among all cases considered.

After carefully reviewing the previous literature, it is found that there is no comprehensive investigation on the guide vane and flow patterns on the in-line fan by means of both numerical and testing means together. Thus, this work integrates CFD, mockup fabrication, and experimental technique together to fulfill the need of this challenging task.

2. Experimental Setups of Performance Tests

To validate the numerical results, the performance curve and noise levels of an in-line fan are tested by the standard

test procedures following AMCA 210-99 [10] and CNS-8753

codes [11], respectively. In this experiment, the fan speed

is measured by a noncontact photo tachometer, which is accurate to 0.05% with a resolution of 1 rpm. The discharge static pressure is measured using a pressure transducer, and the flow rate can be calculated with the aids of differential pressure measured across multiple nozzles. Therefore, the measurement uncertainty of discharge flow rate and static pressure are mainly affected by the accuracy and calibration of the pressure transducer. Since the static pressure is very small for a typical fan, this study chooses a pressure trans-ducer designed for exceedingly low differential pressure

mea-surement applications. With full ranges below±0.56 cm, the

accuracy of this instrument was estimated to be within 0.25% full scale. Moreover, a precision pressure calibration system is used to calibrate this ultra-low-range pressure transducer

with a resolution of 0.05 cm H₂O and an accuracy of±0.5%

of reading. In summary, the uncertainty of the performance

measurement was estimated to be around±2.0%.

This experiment measures sound pressure levels using a RION NL-14 portable sound level meter (SLM) and an AND AD-3524 FFT frequency analyzer. The sound pressure creates analog electric signals in SLM microphone; these signals were then fed into the FFT analyzer to generate the noise characteristics. The noise at fan outlet is mea-sured by employing the CNS-8753 standard. Additionally, the consistency and calibration of the above devices are checked by a piston phone (94 dB at 1000 Hz) both before and after a set of measurements. The lowest cut-off frequency is 125 Hz. The semianechoic chamber offers an appropriate test

environment with18 ± 2 dBA background noise and a 2.8 m

× 2.3 m × 2.05 m valid measuring size. The absorption rate of sound pressure for wedge is greater than 99%. The reflection of sound pressure for wedge is less than 10%. The inverse square law is valid between 125∼20 kHz, and the deviation is ±1.5 dB. The transmission decay is larger than 35 dBA.

3. Numerical Scheme

To simulate the complex flow patterns inside the small

in-line fan, the commercial software Fluent [12] is utilized

to solve the fully three-dimensional incompressible Navier-Stokes equations with the standard turbulence model. Also, the modified SIMPLE method and Pressure-Implicit with Splitting of Operators (PISO) coupling method are imple-mented to speed up the efficiency of the pressure calculation. The PISO algorithm adopts additional corrections to satisfy the momentum balance more closely, especially for transient problems. The numerical calculation is executed with the moving reference frame (MRF) model to deal with the rotating fluid for an in-line fan. Then, the fan is evaluated for each operating point on the entire performance curve.

In this work, several appropriate assumptions and bound-ary conditions were made to simulate the actual flow patterns inside the fan. They are described as follows.

(1) Pressure Outlet Boundary Condition. The atmosphere is set

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Table 1: Dimensions and parameters of the original fan. Name Dimensions (mm) Inlet diameter 93.4 Outlet diameter 93 Length of fan 266.33 Width of impeller 58.48 Diameter of impeller 141.8 Number of blades 12

to calculate the maximum airflow rate for the case of free-delivery condition.

(2) Mass Flow Inlet Boundary Condition. The mass flow rate

of each operating point at outlet boundary surface is set for calculating the corresponding static pressure.

(3) Wall Boundary Condition. This numerical model sets the

no-slip boundary condition on the solid surfaces of the fan.

(4) Moving Reference Frame (MRF). The numerical

investi-gation deals with the rotating flow via the MRF in the CFD codes. The rotating wall surfaces are treated as stationary boundaries relative to rotating frame in this model.

In order to create a similar condition as test chamber, the ample inlet and outlet domains are adopted to provide a numerical measuring station and simulate the conditions of which the fan is expected to encounter. Besides, to correctly simulate the flow patterns of the in-line fan, the fine meshes are employed in the regions of rotor blades surface where acute flow variations occur. To minimize the error related to CFD model and grid number, it is necessary to establish appropriate numerical models for in-line fan and carry out the grid independence test. A fine grid is generated near the rotor region to accurately predict the physical phenomena, and to have a lower deformed geometry, tetrahedron grid is applied here.

4. Flow Analysis and Improving Alternatives

This analysis focuses on the aerodynamic performance of an in-line fan via an integrated effort of CFD and experimental technique. At first, the commercial CFD code (Fluent) is used to simulate and analyze the detailed flow pattern inside

the original fan design (see Figure1and Table1). Based on

the reversed flows observed from numerical outcome, two design alternatives are proposed and evaluated by the cor-responding simulations. In addition, the actual performance enhancement is validated by the experimental test on the CNC-fabricated mockup of the new design.

4.1. Flow Characteristics of the Original Fan Design. The

purpose of effectively distributing the computing resources on the resolution of flow-field phenomenon, a fine grid distribution is used in the rotor region, and coarse grids are utilized in the inlet and outlet zones. Moreover, the grid independence test is applied for identifying the appropriate

266.33

141.79

93.4 93

58.48

Figure 1: Geometry of the original in-line fan.

27 28 29 30 31 163 164 165 166 167 1.5 2.0 2.5 3.0 3.5 4.0 M axim um s ta tic p res su re (mmA q) M axim un v o lu me flo w ra te (cf m)

Mesh number (millions) Max. flow rate

Max. static pressure

(a) Performance for different grid numbers

Grid number (million) CPU time (hours) Iteration number Max. flow rate (cfm) Max. static pressure (mmAq) 1.6 8 6000 165.22 29.37 2.1 11 6000 164.3 29.56 2.5 15 6000 163.39 29.27 3 21 6000 163.99 28.9 3.7 30 6000 164 28.84

(b) CPU times for different grid numbers

Figure 2: Calculated results for the grid independence test.

grid system. Figure2shows the calculated maximum static

pressure and flow rate for several grid systems with different grid numbers, and each case has been run on an i5-2500K personal computer with 16 GB Ram for about 6,000 iterations;

also, the convergence criterion is set as 10−3and the residual

number drops 4 to 8 orders.

Clearly, the results of the 3 M-grid and 3.7 M-grid systems are almost identical and sufficient to yield the detailed flow patterns inside the original in-line fan. As a result, the 3.7 M-grid system is selected to perform the related CFD

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0 5 10 15 20 25 30 35 0 0.5 1 1.5 2 2.5 3 3.5 4 4.5 5 P (mmA q) Q (CMM)

(a) Performance curve

A 23.4 21.1 18.7 16.4 14.1 11.7 9.4 7.0 4.7 2.3 0.0 V elo ci ty (m s −1 )

(b) Overall flow field

Figure 3: Calculated performance curve and flow pattern of the original in-line fan.

18.6 16.8 14.9 13.0 11.2 9.3 7.5 5.6 3.7 1.9 0.0 V elo ci ty (m s −1) B

Figure 4: Detailed flow pattern around the rib of the original fan.

the performance curve and the overall flow filed attained from the numerical simulation of original case. It is found that the maximum static pressure and maximum flow are 28.84 mmAq and 4.64 CMM, respectively.

Numerical results from the numerical simulation of pressure counter and velocity field show that there is a high

reversed flow at the fan exit as point A in Figure3. The farthest

the reversed flow from position A to the exit of the fan is, the better the efficiency of the fan will be; hence modifying the tail body geometry is necessary. And the detailed flow pattern around the upper rib is illustrated around point B in

Figure4. It is obvious that holding ribs form a blockage when

flowing through the impeller, and its short length and straight shape have negative impact on the static pressure. It shows that there exists a low-pressure region due to circulation as a result of sudden stepping down of the flow field. This effect

is shown in Figure4in which the high pressure only occurs

at the blade outlet. Elongating the rib and guiding the flow spirally to the fan exit will prevent this sudden change of the fluid flow; thus the spiral guiding vanes are designed for this purpose. Thus, the fan pressure is enhanced and also that flow is moving more smoothly.

4.2. Designs of Guiding Vane and Tail Body. To increase the

static pressure of the fan, the guiding vanes are designed, while the same rib height of the original fan is maintained.

126.3 78.9 31.6 −15.8 −63.2 −110.5 −157.9 −205.3 −252.6 −300.0 P ressur e (P a)

Figure 5: Pressure distribution of the original in-line fan.

45∘

60∘

30∘

(a) Sketch for different twist angles

(b) Guiding vanes

Figure 6: Geometries of the guiding vane.

It is easy to find from Figures4and5that the exit flow from

impeller is inclined with an angle of about 45 degrees to the horizontal axis, hence different twisted angles of the guide vanes are proposed and the prime number of guide vanes is employed in order to have a better noise performance.

Figure 6 shows several geometries of guiding vanes with

various twisted angles. Note that the vane lengths for the 30-degree and 45-degree twisted angles are only 0.5 and 0.64 times of the original rib height, respectively, while the length

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Arc Elliptic Polynomial

45

42

Figure 7: Different shapes for the tail body (unit: mm).

42 30 15 45 45∘ 22∘

Figure 8: Sketches of the polynomial shape of the tail body (unit: mm).

of twisted 60-degree vane is 1.28 times to provide a longer distance for increasing its static pressure. So this twisted angle of guiding van is used for attaining the smooth flow pattern and increasing the vane length.

The tail body is modified by producing the appropriate curve shape. The arc curve, elliptic curve, and spline curve

(Figure7) for the tail body are considered here under the

constraints of a maximum 45 mm in length and installing along the center line of motor axis. Note that the spline line with an optimum initial angle is firstly determined through iteration. Then other subsequent points are defined at 22, 45,

and 90 degrees as illustrated in Figure8. Three polynomial

curves of tail body are created for three lengths (15, 30, and 45 mm) of major axis. 40 35 30 25 20 15 10 0 P (mmA q) 0 0.5 1 1.5 2 2.5 3 3.5 4 4.5 5 Q (CMM)

With guide vanes Original 5

Figure 9: Performances for the original fan and the fan with guiding vanes.

Table 2: Performances for fans with different twist angles on the guiding vane design.

Original 30 degrees 45 degrees 60 degrees

𝑃max 28.84 35.34 37.2 35.1

𝑄_max 4.64 2.82 3.49 4.18

5. Discussions on Modified Fan

5.1. Experimental and Simulation Analysis. In order to

achieve a better performance of the fan, the parameters stud-ied in the previous sections are modifstud-ied and checked with CFD simulation firstly prior to manufacturing the prototype. And the product with modified design is developed and investigated based on the flow visualization of numerical

output. Figure 9 shows the comparison of performances

between the original fan and the modified fan with guiding

vanes. From Table2, the fan performance is enhanced by 21%

on maximum pressure with a 10% reduction on flow rate by using guiding vane.

These performance changes are due to the blockage influence of guide vanes and a pronounced enhancement for the maximum static pressure. Experimental and numerical results of the new design are found in good agreement with a minor variation. Also, AMCA test chamber established in accordance with code 210-99 is used for the performance experiment in this work.

5.2. Introducing the Guiding Vanes. In order to minimize

the calculated error, the designed fan with guiding vanes

is simulated using the 4 M-grid numerical model. Table2

shows the simulation results of maximum static pressure and flow rate for fans with different twist angles. The case of twisted 60 degrees increases the maximum static pressure to 35.01 mmAq while reducing the flow rate to 4.18 CMM. Also, the designs with 30-degree and 45-degree twist angles improve maximum static pressure to 35.34 mmAq and 37.2 mmAq, respectively. But the generated flow rate

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0 1 2 3 3.5 4 4.5 3 4 6 40 35 30 25 20 15 10 0 5

Figure 10: Performance curves for fans with different twist angles on the guiding vane design.

0 0.5 1 1.5 2 2.5 3 3.5 4 4.5 Q (CMM) Simulation Experiment 40 35 30 25 20 15 10 0 P (mmA q) 5

Figure 11: Performance comparison for test and CFD results for fans with different twisted angles on the guiding vane design.

decreases significantly to 2.82 CMM and 3.49 CMM, which represent almost 39% and 25% loss of flow rate.

It is clear that the twisted 60∘vane are the proper choice

since it not only enlarges static pressure by 21.7% but also has a

better overall performance curve as shown in Figure10.

After-wards the fan performance with twisted 60-degree guiding vane is tested via the PR-fabricated mockup. The maximum static pressure and the maximum flow of the experimental results are found to be 30.57 mmAq and 4.11 CMM, as shown

in Figure11. Thus, a rough 5% deviation is observed between

CFD and test results and this difference is presumably due to the surface roughness caused by the PR (rapid-prototyping) technique.

Since the deviation is within an acceptable range, it is concluded that the numerical result can accurately predict

its performance. As illustrated in Figure12, it is clear to find

that the in-line fan with guiding vanes not only increases the pressure effectively but also eliminates the low-pressure region at inlet significantly. Also, the modified spiral vanes have smoothened the flow pattern and take away the flow

126.3 78.9 31.6 −15.8 −63.2 −110.5 −157.9 −205.3 −252.6 −300.0 P ressur e (P a)

Figure 12: Pressure distribution of the fan with guide vanes.

V elo ci ty (m s −1) 21.6 19.5 17.3 15.1 13.0 10.8 8.7 6.5 4.3 2.2 0.0

(a) Detailed flow pattern around guide vanes

V elo ci ty (m s −1) 22.9 20.6 18.3 16.0 13.7 11.4 9.2 6.9 4.6 2.3 0.0

(b) Overall flow field

Figure 13: Flow patterns of the in-line fan with the guide vanes.

Table 3: Performances of fan with different curvature on the tail body design.

Original Arc curve Elliptic curve Polynomial curve

𝑃_max 28.84 29.6 29.09 29.36

𝑄max 4.64 4.87 4.91 4.88

separation further as shown in Figure13. And the reversed

flow quoted as point B in Figure 4 has been reduced

considerably. However, there still exists a large reversed flow area at the motor base and the fan discharge. Therefore, it motivates the addition of tail body to eliminate this problem. Several important parameters of tail body are evaluated in the following subsection.

5.3. Adding the Tail Body. Different parameters on the tail

body through numerical simulation are investigated and discussed here. The comparison of maximum static pressure and maximum flow rate for simulation results of the modified

tail body fan and the original in-line fan is listed in Table3. It

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0 5 10 15 20 25 30 35 0 0.5 1 1.5 2 2.5 3 3.5 4 4.5 5 5.5 St at ic p ressur e (mmA q)

Volume flow rate (CMM) Arc

Elliptic Polynomial

Figure 14: Fan performance curves for the tail body with different curvatures.

22.98 17.23 11.49 5.74 0.00 V elo ci ty (m s −1)

(a) Arc curve

23.79 17.84 11.89 5.95 0.00 V elo ci ty (m s −1 ) (b) Elliptic curve 24.7 22.2 19.7 17.3 14.8 12.3 9.9 7.4 4.9 2.5 0.0 V elo ci ty (m s −1) (c) Polynomial curve

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0 5 10 15 20 25 30 35 40 0 0.5 1 1.5 2 2.5 3 3.5 4 4.5 5 St at ic p ressur e (mmA q)

Volume flow rate (CMM) With arc curve tail body

With elliptic curve tail body With polynomial curve tail body

Figure 16: Performance curve for fans with guiding vane and various tail bodies.

Table 4: Performance of fan with guiding vane and various tail bodies.

Guiding

vanes only Arc curve

Elliptic curve Polynomial curve 𝑃_max 35.1 36.96 37.47 35.0 𝑄_max 4.18 4.53 4.56 4.26

flow rate to 4.88 CMM with a slight increase of static pressure at 29.6 mmAq and 29.09 mmAq, respectively.

And the fan with elliptic-curve tail body provides a higher flow rate to 4.91 CMM and the similar 29.09 mmAq static

pressure as illustrated in Figure14. Even though there is no

huge difference among the fan with different shape of tail

body. Figure15shows the flow pattern for different curve of

tail body; it can be found that there are still some reversed flows existed for the cases of arc and polynomial curves. Clearly, imposing both the modified tail body and the guide vanes together on the in-line fan is instructive and necessary to explore. And the following section shows the simulation results of these in-line fans.

5.4. Adding Guide Vane and Tail Body. In order to increase

flow rate and eliminate reversed flow at exit region simulta-neously, the guide vane and the tail body are employed in the

in-line fan design to have a better outcome. Table4shows the

performance comparison for the guide vane fan with different tail bodies. As a result, the guide vane designs with arc, elliptic, and polynomial-curve tail bodies yield the maximum static pressure and flow rate as (36.96 mmAq, 4.53 CMM), (37.47 mmAq, 4.56 CMM), (35 mmAq, 4.26 CMM),

respec-tively. Also, it is clearly indicated in Figure 16 that the

performance of guide vanes with an arc curve tail body is superior to the other two curves.

Figure 17shows the performance comparison between

fan with guiding vane and fan with both vane and tail body. Apparently, the maximum static pressure and flow rate are enlarged by 6.8% and 9.1%, respectively. Moreover, its

0 5 10 15 20 25 30 35 40 0 0.5 1 1.5 2 2.5 3 3.5 4 4.5 5 St at ic p ressur e (mmA q)

Volume fow rate (CMM) With both guiding vane and tail body With guide vanes

Figure 17: Performance comparison between fan with guiding vane and fan with both guiding vane and tail body.

23.1 20.8 18.4 16.1 13.8 11.5 9.2 6.9 4.6 2.3 0.0 V elo ci ty (m s −1)

Figure 18: Flow fields of the in-line fan with guiding vanes and tail body.

maximum static pressure and flow rate are 28.8% higher and 7.7% lower than these from the fan with guide vane only.

When implementing both the spiral guide vanes and the modified tail body, the separation point is successfully delayed, while a reversed flow is still identified at exit area

as indicated in Figure18. Additionally, Figure19illustrates

the detailed flow pattern around the guide vane, which yields the smoother flow field pattern and a larger velocity by the effect of tail body. However, when both the modified tail body and the spiral guide vanes are adopted in the design alternative, the maximum static pressure is found to have an impressive 30% enhancement with a minor 1.7% reduction on its flow rate. On the other end, the comparison on the overall performance curves between the original fan and design with

both vane and tail body is shown in Figure20. Apparently,

the new fan design with guide van and tail body results in a satisfactory outcome, especially on the aspect of static pressure gain.

6. Concluding Remarks

In this study, an integrated performance analysis on an in-line fan is carried out through a combined experimental and numerical approach. The numerical visualization on the flow characteristics provides significant information to improve

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23.7 21.3 18.9 16.6 14.2 11.8 9.5 7.1 4.7 2.4 0.0 V elo ci ty (m s −1)

Figure 19: Flow pattern of the fan with guiding vane and tail body.

40 35 30 25 20 15 10 0 0 0.5 1 1.5 2 2.5 3 3.5 4 4.5 5 5 St at ic p ressur e (mmA q)

Volume ﬂow rate (CMM)

Original With both

Figure 20: Performance comparison between the original fan and fan with guiding vanes and tail body.

the inner flow patterns for fan designers. For validating CFD prediction, a fan mockup is fabricated via RP technique to serve as the investigation subject. The result indicates that experimental P-Q curve agrees to that of the numerical simulations. The deviations between CFD and test results are approximately 5% on both the maximum volumetric flow rate and the maximum static pressure.

In summary, this integrated fan performance evaluation scheme consists of CFD analysis, mockup fabrication, and performance test. Through this systematic evaluation pro-cedure, the fan designer can visualize the flow patterns and modify its design for meeting the request under various operating points. Consequently, this study establishes an integrated aerodynamic, rapid prototype, and experiment approach that can be used as an important tool for fan designers.

Conflict of Interests

The authors declare that there is no conflict of interests regarding the publication of this paper.

Acknowledgments

The authors would like to thank the financial supports under Department of Energy Equipment System Engineering, Industrial Technology Research Institute (ITRI-102AL604), and Ministry of Education (MoE) Top University Projects (102H223401 and 102M21005).

References

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