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Effect of inlet straighteners on centrifugal fan performance
Energy Conversion and Management 47 (2006) 3307–3318 www.elsevier.com/locate/enconman
Effect of inlet straighteners on centrifugal fan performance N.N. Bayomi a
a,*
, A. Abdel Hafiz a, A.M. Osman
b
Faculty of Engineering, Mataria, Helwan University, 11718 Masaken, El-Helmia, Cairo, Egypt b Faculty of Engineering, Shoubra, Zagazig University, Cairo, Egypt Received 31 July 2005; accepted 30 January 2006 Available online 3 April 2006
Abstract The use of straighteners in the inlet duct of centrifugal fans is suggested for eliminating any inlet distortion. An experimental investigation was performed to study the effect of inlet straighteners on the performance characteristics of centrifugal fans. Two types of straighteners were used, circular tubes and zigzag cross section, with different lengths. Circular tubes with different diameters have been investigated. The study was conducted on three types of fans, namely radial, backward with exit blade angles 60° and 75° and forward with 105° and 120°. The results confirm that the inlet straighteners exhibit different effects on the fan performance for the different blade angles. Accordingly, the results indicate the selection of long circular tube straighteners with large diameter for radial blades, long zigzag type for backward 60° blade angle and short zigzag type for backward 75° blade angle. Generally, good improvements in efficiency are observed for radial and backward blades on account of a slight drop in static head. In addition, an increase in the flow margin up to 12% and a decrease in the noise level from 3 to 5 dB are indicated compared to the free inlet condition. On the contrary, unfavorable influences are exerted on the forward fan performance. Ó 2006 Published by Elsevier Ltd. Keywords: Centrifugal fan; Straighteners; Noise; Distortion
1. Introduction In the traditional market of centrifugal fans for industrial, commercial and utility applications, strong emphasis has long been placed on the initial cost of these fans. Considerations of ease of manufacturing and installation and maintenance of the equipment in the field have tempered any improvements in performance. The growing breadth of fan applications causes variations in inlet duct configuration due to spatial restrictions. Flow non-uniformity is frequently generated at the impeller inlet, and consequently, deterioration of fan performance is expected. Generally, this is known as the inlet distortion. Deviations from a steady uniform distribution of the flow properties can include variations in swirl, velocity, turbulence, total and static pressures, velocity, temperature, flow angle and fluid density. Non-uniform inlet profiles are created in industrial fans or in ventilation systems using a 90° bend directly upstream of the inlet due to mechanical *
Corresponding author. Tel.: +20 2 4838988; fax: +20 2 2735437. E-mail addresses: [email protected], [email protected] (N.N. Bayomi).
0196-8904/$ - see front matter Ó 2006 Published by Elsevier Ltd. doi:10.1016/j.enconman.2006.01.003
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Nomenclature D D FM H L Psh R ro SM V b1 b2 DHst c g P
inlet duct diameter straighteners tube diameter flow margin straightener zigzag height straightener length shaft power radius outer radius of inlet duct surge margin volume flow rate inlet blade angle exit blade angle static head difference specific weight efficiency = cVDHst/Psh static pressure ratio
Subscripts Max maximum Op operating point Sp surge point
limitations, which dictate the radial entry to the machine. Generally, the air separates at the top surface of the bend and generates secondary flow within the cross sections of the inlet. The swirl generated by the secondary flow and the separation results in a distortion of the flow field at the fan entry. Ariga et al. [1] divided the inlet distortion for compressors into two dominant forms, radial and circumferential distortions. The former one is subdivided into tip and hub distortion. Hub distortion occurs when an axisymmetric obstacle is used at the center portion of the inlet fan such as a tachometer pick up and hub cover. Tip distortion happens when axisymmetric boundary layers of an inlet duct exist or axisymmetric obstacles such as an orifice plate are used. Circumferential distortion happens from non-axisymmetric obstacles such as struts or a bending duct. Although the non-uniformity of inlet flow appears frequently in centrifugal fans, only a few data about inlet distortions are found in the literature, and most of them are for centrifugal compressors. Field measurements of compressor performance indicated that both efficiency and pressure rise were several percentage points lower than the expected performance [1,2]. The onset of stall was influenced or magnified by severely distorted inflows [3–5]. Similarly, for centrifugal fans, Wright et al. [6] showed significant degradation in efficiency and pressure rise, as much as 10–15%, resulting from moderately to severely distorted inflow patterns. The existence of inlet distortion is considered to cause partial flow separation at the entrance of the fan compared to non-distorted conditions. Moreover, the flow range becomes narrower due to the fact that the beginning of the instability of the flow, such as rotating stall and surge, in centrifugal fans is affected by seriously distorted inflows. Consequently, it is necessary that the distorted flow be rectified before it enters the impeller. This can be done by different ways based on a mechanism in which secondary vortices are counteracted by vortices generated in the opposite sense of the secondary flow by additional vortex generators. Inlet guide vanes were employed by Madhavan and Wright [7,8], Chen et al. [9], Montazerin et al. [10], Kassens and Rautenberg [11] and Coppinger and Swain [12]. Unfortunately, additional inlet vortices occur in fans with inlet vane control, causing unstable flow at the entrance of the impeller, which further complicates the situation. This instability causes unfavorable effects on the stall point and increases noise and vibration levels, which can lead to fatigue cracks in inlet ducts as well as in the rotor [9]. Also, Jack [13] found that centrifugal fans that operate at
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inefficient lower volumes are subject to rotating stall or surge, which wastes power and generates excessive low frequency noise. Bhope and Padole [14] investigated the noise level and fluid flow in a centrifugal fan impeller. The present paper suggests the fitting of annular straighteners at the entrance of the impeller in order to rectify the non-uniformity of the flow and to eliminate the vortices that are generated by the existence of inlet distortion. The main objective of this work is to assess the use of these straighteners on the fan performance. For this purpose, two different types of straighteners, circular and corrugated (zigzag), are considered with different sizes. The investigation is conducted on five different impellers with different exit blade angles. Comparisons with free inlet fans are performed. Measurements of static head, shaft power and noise levels at different loads are conducted for the different cases. The analysis of these measurements gives some information concerning the operating range and surge margin for these types of fans. 2. Test facilities and instrumentation Experimental investigation were conducted in the Turbomachine Laboratory of the Mataria Faculty. The test rig consists of a low pressure commercial centrifugal fan of the radial type, a test inlet duct and a delivery duct. The fan wheel is comprised of 16 straight blades of 3 mm thickness with constant blade width of 60 mm welded to a back plate and a shroud. The impeller inner and outer diameters are 215 mm and 394 mm, respectively. The scroll casing is of constant rectangular width. The fan is driven by an electric motor of shaft power 3 hp at constant speed of 2800 rpm. The test inlet duct is 160 mm in diameter and 300 mm long. The exit circular duct of 100 mm diameter is connected to the rectangular outlet of the fan through a conical connection and fitted at the end with a throttle valve. Fig. 1 illustrates the test rig layout equipped with the measuring devices. In this investigation, the suggested straighteners for overcoming any tip or circumferential distortion are located in the inlet duct at a distance of 30 mm from the impeller entrance. Two types of straighteners are designed, both with constant annular cross section of inner and exit diameters 45 mm and 160 mm, respectively. One type consists of annular bundles of plastic or PVC tubes with different diameters, 2.5, 4 and 15 mm. The other type (the zigzag type) is manufactured with the same process used for catalytic converters for car exhaust. Hardened paper foil is corrugated and wound together with non-corrugated foil, making a triangular cross section of height 10 mm. The various layers of corrugated and non-corrugated foils are glued to each other, making the annular shape. The length of the straighteners, considered as a parameter, has been
Circular tube
Static taps
Centrifugal fan
Prandtl probe
Spherical valve
Flow Static taps
Pitot tube Straight blade
Computer Sound level meter
Micromanometer
Multi-channel Static and/or total pressure Switch
Wattmeter
(a) Test rig. Impeller
Straightener
Impeller
Air inlet
(b) Setting location of straighteners. Bellmouth
Fig. 1. Test rig and measuring devices layout.
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L
D
h=10 mm
d=2.5 mm
d=4 mm
d=15 mm
2640 tube
1030 tube
74 tube
Zigzag type
Circular type
Fig. 2. A schematic drawing of the different shape of straighteners.
Total suction head (m water)
varied from 225 mm to 180 mm, making a length to duct diameter ratio, L/D, of 1.4 and 1.125. Schematic drawings show the different shapes of the straighteners in Fig. 2. The average static pressures at the inlet and exit of the fan are measured through four taps equally distributed circumferentially, Fig. 1. The flow velocity distribution across the delivery duct diameter was measured using a standard cylindrical Prandtl probe with inner diameter 2 mm mounted on a traverse mechanism with accuracy ±0.1%. The probe is located at ten diameters from the delivery duct inlet to ensure uniformity of the flow. The flow through the fan is controlled by a spherical regulator valve located at the end of the delivery pipe. In order to check the flow uniformity at the fan inlet downstream of the straighteners, the total pressure distribution is measured by a shielded Pitot tube using a traversing mechanism with accuracy ±0.1%. All the pressures were measured through a multi-channel switch by a digital micro-manometer, model Yokogawa 2655, with resolution of 0.1 Pa and updating of the reading every 0.4 s. An average of the readings is computed every 5 s using an A/D converter and a PC. Fig. 3 shows the total suction head distribution of the radial blade impeller with and without straighteners at the design point. From this figure, it can be seen that the total head at the straighteners exit is approximately constant. Compared with the free inlet, a drop in the suction head is detected by the presence of the straighteners that increases as the diameter of tubes decreases.
0.00 -0.05 -0.10
Free d=2.5 mm
-0.15
d=4 mm d=15 mm
-0.20 -0.25 0.0
0.2
0.4
0.6
0.8
1.0
r/ro
Fig. 3. Total suction head distribution at the inlet of the radial fan with and without straighteners at design point.
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Table 1 Characteristics of the different impellers Parameter
Original impeller
Impeller I
Impeller II
Impeller III
Impeller IV
Outlet angle, b2 (deg) Inlet angle, b1 (deg) Blade length (mm)
90 90 80
60 25 86
75 60 84
105 125 84
120 150 86
Delivery Static Head (m water)
The shaft power of the fan is measured by a digital wattmeter with accuracy ±0.09%, while the rotational speed is measured by a digital tachometer, model Lutron Dt-2236, with accuracy ±0.05%. The noise level in dB, measured by sound pressure level of the fans with different inlet configurations, is determined. A portable sound level meter equipped with a special stand and set to A-weighting (slow response) is used. Three different near field measuring locations have been chosen at a standard distance equal to twice the impeller housing diameter in accordance with DIN 45635: at fan inlet, near the delivery duct exit and behind the fan motor. The noise level was always found to be maximum near the exit of the delivery duct, and therefore, measurements were recorded and are presented only at this station. The effect of the straighteners on fans with different exit blade angles has also been investigated. Accordingly, four new impellers, two backward and two forward facing, with different exit blade angles have been
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V (m3/s)
Static Efficiency%
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Shaft Power (kW)
2.0 1.5 1.0 0.5 0.0 0.00
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V (m3/s)
Fig. 4. Fan performance for radial impeller (b2 = 90°) with different straighteners.
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constructed using the original scroll housing. More details about the different impellers are tabulated in Table 1. 3. Experimental results and discussion
Delivery Static Head (m water)
The characteristic curves of the five different tested impellers are shown in Figs. 4–8. The measured delivery static head, together with the calculated static efficiency and the shaft power, are plotted versus the volume flow rate for each fan with the different types of inlet straighteners of L/D = 1.4. Comparisons with the free inlet condition were performed on the same plots. The performance of the radial fan using different straighteners is shown in Fig. 4. At large flow rate, a remarkable increase in static head due to the straighteners can be noticed after 0.15 m3/s. This is accompanied by an appreciable improvement in the fan efficiency. By decreasing the flow rate, the effect of the straighteners vanishes. As the diameter of the straightener tubes increases, a more flattened efficiency curve is observed. An improvement of 5 points in efficiency, corresponding to a relative increase of 18%, is obtained with straighteners of 15 mm diameter, and a corresponding decrease in shaft power is noticeable. This is due to the good guidance of the flow provided by the straighteners at the impeller inlet. Furthermore, the unstable operating range of the fan extends farther due to the straighteners. It is useful to note that during the experiments, the surge point was detected by fluctuations in the pressure readings in addition to the high audible noise. The 0.30
β2=60º
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Free d=2.5 mm
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d=4 mm
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d=15 mm
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Zigzag
0.05 0.00 0.00
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V (m3/s)
Static Efficiency %
40.0 30.0 20.0 10.0 0.0 0.00
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V (m3/s)
Shaft Power (kW)
2.0 1.5 1.0 0.5 0.0 0.00
0.05
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0.20
V (m3/s)
Fig. 5. Fan performance for backward impeller (b2 = 60°) with different straighteners.
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Delivery Static Head (m water)
results of the noise test in Fig. 9, expressed in sound pressure levels in dB, (A) show that the use of straighteners decreases the noise level along the operating range by approximately 5 dB at the high flow rate. In practice, an overall sound power increase of 3 dB is just perceptible to the human ear and 5 dB is clearly louder. The performance of the backward fans with 60° and 75° exit blade angle is shown in Figs. 5 and 6, respectively. The first impeller exhibits improvement that could reach about 6 points in efficiencies employing the zigzag straighteners. This represents a relative increase of 21% in efficiency. However, a small drop in the static head associated with an increase in shaft power is observed. A reduction of about 3 dB in noise level is the result, Fig. 9. It worth noting that the noise level increases as the blade angle increases, which was also detected by Liberman [15]. As the blade angle increases to 75°, a lower efficiency is obtained all over the operating range at the free inlet condition. This is due to the high incidence losses resulting from the corresponding large inlet blade angle. Using the inlet straighteners leads to a further drop in static head as well as in efficiency. However, the use of straighteners with very small tube diameter increases, obviously, the efficiency but on account of the low delivery head. This is associated with a noticeable decrease in the maximum flow rate, choke point. However, the onset of the surge point shifts to a lower flow rate. Figs. 7 and 8 indicate the performance of the forward fans with exit blade angle 105° and 120°. The use of inlet straighteners result in small increases in static efficiency for 105° on account of the appreciable drop in static head, whereas, for the impeller with 120°, a deterioration in efficiency as well as in delivery head is noticed. In this case, it is worth noting that at the free inlet condition, the fan efficiency is already very 0.30 0.25
β2=75º Free
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d=2.5 mm
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V (m3/s)
Static Efficiency %
40.0 30.0 20.0 10.0 0.0 0.00
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V (m3/s)
Shaft Power (kW)
2.0 1.5 1.0 0.5 0.0 0.00
0.05
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0.15
0.20
V (m3/s)
Fig. 6. Fan performance for backward impeller (b2 = 75°) with different straighteners.
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β2=105
0.30
Free
0.25
d=2.5 mm
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d=4 mm
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d=15 mm
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Zigzag
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V(m3/s)
Static Efficiency %
40.0 30.0 20.0 10.0 0.0 0.00
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V(m3/s)
Shaft Power (kW)
2.0 1.5 1.0 0.5 0.0 0.00
0.05
0.10
0.15
0.20
V(m3/s)
Fig. 7. Fan performance for forward impeller (b2 = 105°) with different straighteners.
low. This is probably due to separation of the flow inside the impeller passages as the number of blades is much lower than usual for forward facing blades. This is associated with a higher noise level compared to radial and backward facing blades. This is in agreement with the results of Liberman [15]. To assess the effect of the straighteners on the fan operation, some parameters should be taken into consideration. These parameters are the flow margin and the surge margin. The flow margin is defined as V sp Flow margin ðFMÞ ¼ 1 100% V max where Vsp and Vmax are the volume flow rate at the surge point and the maximum flow rate, respectively. The surge margin is calculated from the definition given by Cumpsty [16] as " P !# Surge margin ðSMÞ ¼
PV sp 1 V op
where P is the static pressure ratio and the suffix op indicates operating point corresponding to the condition at maximum efficiency. Figs. 10 and 11 show the calculated flow margin and the surge margin, respectively, for the different fans with different straighteners compared to the free inlet condition. The results show that the flow margin may be
Delivery Static Head (m water)
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β2=120º Free
0.25
d=2.5 mm
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d=4 mm
0.15
d=15 mm
0.10
Zigzag
0.05 0.00 0.00
0.05
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V(m3/s)
Static Efficiency %
30.0
20.0
10.0
0.0 0.00
0.05
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0.20
V(m3/s)
Shaft Power (kW)
2.0 1.5 1.0 0.5 0.0 0.00
0.05
0.10
0.15
0.20
V(m3/s)
Fig. 8. Fan performance for forward impeller (b2 = 120°) with different straighteners.
arbitrarily increased up to 12% by using inlet straighteners for the backward and radial impellers. For the forward blade, this reduces the flow margin, especially with small diameter. Great improvements in the surge margin are depicted for the different blade angles when using the different straighteners, Fig. 11. From the previous results, it can be deduced that the effect of the straighteners on the fan performance varies according to the exit blade angle. Accordingly, the most suitable straightener for each impeller type can be selected. Straighteners with tube diameter 15 mm conform well to the radial blades, whereas the zigzag type is advisable to be used with backward blades. It follows from the results analysis that inlet straighteners are not convenient for forward blades. The effect of straightener length on the fan performance has been studied. Samples of the results obtained using short straighteners of L/D = 1.125 are presented in Fig. 12 compared to the longer one taking into consideration the most efficient inlet configuration for each fan. It can be noted that for the radial and backward impeller with blade angle 60°, decreasing the length of the straighteners weakened the performance of the fan, whereas for blade angle 75°, the shorter zigzag type of straighteners improves the fan performance. 4. Conclusion The present paper investigates the effects of inlet straighteners on the performance, operating range and instantaneous surge of a centrifugal fan. Experimental investigations concerning different types and sizes of
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Free d=2.5 mm
85
d=4 mm d=15 mm Zigzag
80 0.00
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V (m3/s)
Noise (dB)
90
β2=60º 85
80 0.00
0.05
0.10
0.15
0.20
V (m3/s)
Fig. 9. Noise level for different straighteners for radial and backward (60°) impellers.
Fig. 10. Comparison of the flow margin for different blade angles with different straighteners.
inlet straighteners for radial, backward and forward fans were conducted. The following conclusions can be drawn: 1. The effect of straighteners on the fan performance depends mainly on the exit blade angle. More flattened efficiency curves are obtained by increasing the straightener tube diameters. An improvement of 5 points in efficiency corresponding to 18% relative increase in efficiency is obtained using circular tube straighteners with 15 mm diameter and L/D = 1.4 for a radial impeller. A relative increase of 21% in the efficiency of a backward fan of 60° blade angle associated with a small drop in delivery head is obtained when using straighteners of zigzag type. A bad effect for the different straighteners is observed on the fan performance for the forward impeller. A slight effect is noted on the maximum permissible flow rate (choke point) for the radial fan, while for backward and forward blades, it decreases by using straighteners. 2. The flow margin increases up to 12% for backward and radial impellers. 3. Improvements of surge margin are depicted for the different blade angles.
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Fig. 11. Comparison of the surge margin for different blade angles with different straighteners.
Circular type (d=15 mm)
β2=90º
Deivery Static Head (m Water)
0.30
Zigzag type
β2=60º
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Free
0.05
0.05
L/D=1.4
L/D=1.125
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Zigzag type β2=75º
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Fig. 12. The effect of straighteners length on fan performance.
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4. The use of straighteners decreases the noise level by approximately 3–5 dB at high flow rate compared with the free inlet condition for radial and backward impellers. 5. The effect of straighteners length varies with exit blade angles. For the backward impeller with blade angle 60°, as well as for radial fans, the longer zigzag and circular straighteners, respectively, give better performance, whereas for blade angle 75°, the shorter zigzag is the best.
References [1] Ariga I, Kasai N, Masuda S, Watanabe Y, Watanabe I. The effect of inlet distortion on the performance characteristics of a centrifugal compressor. Trans ASME, J Eng Power 1983;105:223–30. [2] Ariga I, Masuda S, Okita A. Inducer stall in a centrifugal compressor with inlet distortion. ASME paper 86-GT-139, 1986. [3] Graber EJ, Braithwaite WM. Summary of recent investigations of inlet distortion effects on engine stability. AIAA paper no. 74-236, 1974. [4] Greitzer EEM. The stability of pumping systems. Trans ASME, J Fluid Eng 1981;103:193–242. [5] Baghdadi S, Lueke JE. Compressor stability analysis. Trans ASME, J Fluid Eng 1982;104:242–9. [6] Wright T, Madhavan S, Di Re J. Centrifugal fan performance with distorted inflows. Trans ASME, J Eng Gas Turb Power 1984;106(October):895–900. [7] Madhavan S, Wright T. Rotating stall caused by pressure surface flow separation on centrifugal fan blades. ASME paper 84-GT-35, 1984. [8] Madhavan S, Wright T. Rotating stall caused by pressure surface flow separation on centrifugal fan blades. Trans ASME, J Eng Gas Turb Power 1985;107(July):775–81. [9] Chen P, Soundra-Nayagam M, Bolton AN, Simpson HC. Unstable flow in centrifugal fans. Trans ASME, J Fluids Eng 1996; 118(March):128–33. [10] Montazerin N, Damangir A, Mirian S. A new concept for squirrel-cage fan inlet. Proc Instn Mech Eng 1998;212(Part A):343–9. [11] Kassens L, Rautenberg M. Flow measurements behind the inlet guide vane of a centrifugal compressor. ASME paper 98-GT-86, 1998. [12] Coppinger M, Swain E. Performance prediction of an industrial centrifugal compressor inlet guide vane system. IMechE, Part A 2000;214:153–64. [13] Jack BE. Fan selection and to reduce inefficiency and low frequency noise generation. In: Fan noise 2003, International symposium Senlis, 23–25 September. [14] Bhope DV, Padole PM. Experimental and theoretical analysis of stresses, noise and flow in centrifugal fan impeller. Mech Mach Theory 2004;39(12):1257–71. [15] Liberman MYu. Investigation of noise characteristics for centrifugal ship fans. In: XI session of the Russian acoustical society, Moscow, November 19–23, 2001. p. 572–5. [16] Cumpsty NA. Compressor aerodynamics. New York: John Wiley and Sons, Inc.; 1989.
Effect of inlet straighteners on centrifugal fan performance N.N. Bayomi a
a,*
, A. Abdel Hafiz a, A.M. Osman
b
Faculty of Engineering, Mataria, Helwan University, 11718 Masaken, El-Helmia, Cairo, Egypt b Faculty of Engineering, Shoubra, Zagazig University, Cairo, Egypt Received 31 July 2005; accepted 30 January 2006 Available online 3 April 2006
Abstract The use of straighteners in the inlet duct of centrifugal fans is suggested for eliminating any inlet distortion. An experimental investigation was performed to study the effect of inlet straighteners on the performance characteristics of centrifugal fans. Two types of straighteners were used, circular tubes and zigzag cross section, with different lengths. Circular tubes with different diameters have been investigated. The study was conducted on three types of fans, namely radial, backward with exit blade angles 60° and 75° and forward with 105° and 120°. The results confirm that the inlet straighteners exhibit different effects on the fan performance for the different blade angles. Accordingly, the results indicate the selection of long circular tube straighteners with large diameter for radial blades, long zigzag type for backward 60° blade angle and short zigzag type for backward 75° blade angle. Generally, good improvements in efficiency are observed for radial and backward blades on account of a slight drop in static head. In addition, an increase in the flow margin up to 12% and a decrease in the noise level from 3 to 5 dB are indicated compared to the free inlet condition. On the contrary, unfavorable influences are exerted on the forward fan performance. Ó 2006 Published by Elsevier Ltd. Keywords: Centrifugal fan; Straighteners; Noise; Distortion
1. Introduction In the traditional market of centrifugal fans for industrial, commercial and utility applications, strong emphasis has long been placed on the initial cost of these fans. Considerations of ease of manufacturing and installation and maintenance of the equipment in the field have tempered any improvements in performance. The growing breadth of fan applications causes variations in inlet duct configuration due to spatial restrictions. Flow non-uniformity is frequently generated at the impeller inlet, and consequently, deterioration of fan performance is expected. Generally, this is known as the inlet distortion. Deviations from a steady uniform distribution of the flow properties can include variations in swirl, velocity, turbulence, total and static pressures, velocity, temperature, flow angle and fluid density. Non-uniform inlet profiles are created in industrial fans or in ventilation systems using a 90° bend directly upstream of the inlet due to mechanical *
Corresponding author. Tel.: +20 2 4838988; fax: +20 2 2735437. E-mail addresses: [email protected], [email protected] (N.N. Bayomi).
0196-8904/$ - see front matter Ó 2006 Published by Elsevier Ltd. doi:10.1016/j.enconman.2006.01.003
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Nomenclature D D FM H L Psh R ro SM V b1 b2 DHst c g P
inlet duct diameter straighteners tube diameter flow margin straightener zigzag height straightener length shaft power radius outer radius of inlet duct surge margin volume flow rate inlet blade angle exit blade angle static head difference specific weight efficiency = cVDHst/Psh static pressure ratio
Subscripts Max maximum Op operating point Sp surge point
limitations, which dictate the radial entry to the machine. Generally, the air separates at the top surface of the bend and generates secondary flow within the cross sections of the inlet. The swirl generated by the secondary flow and the separation results in a distortion of the flow field at the fan entry. Ariga et al. [1] divided the inlet distortion for compressors into two dominant forms, radial and circumferential distortions. The former one is subdivided into tip and hub distortion. Hub distortion occurs when an axisymmetric obstacle is used at the center portion of the inlet fan such as a tachometer pick up and hub cover. Tip distortion happens when axisymmetric boundary layers of an inlet duct exist or axisymmetric obstacles such as an orifice plate are used. Circumferential distortion happens from non-axisymmetric obstacles such as struts or a bending duct. Although the non-uniformity of inlet flow appears frequently in centrifugal fans, only a few data about inlet distortions are found in the literature, and most of them are for centrifugal compressors. Field measurements of compressor performance indicated that both efficiency and pressure rise were several percentage points lower than the expected performance [1,2]. The onset of stall was influenced or magnified by severely distorted inflows [3–5]. Similarly, for centrifugal fans, Wright et al. [6] showed significant degradation in efficiency and pressure rise, as much as 10–15%, resulting from moderately to severely distorted inflow patterns. The existence of inlet distortion is considered to cause partial flow separation at the entrance of the fan compared to non-distorted conditions. Moreover, the flow range becomes narrower due to the fact that the beginning of the instability of the flow, such as rotating stall and surge, in centrifugal fans is affected by seriously distorted inflows. Consequently, it is necessary that the distorted flow be rectified before it enters the impeller. This can be done by different ways based on a mechanism in which secondary vortices are counteracted by vortices generated in the opposite sense of the secondary flow by additional vortex generators. Inlet guide vanes were employed by Madhavan and Wright [7,8], Chen et al. [9], Montazerin et al. [10], Kassens and Rautenberg [11] and Coppinger and Swain [12]. Unfortunately, additional inlet vortices occur in fans with inlet vane control, causing unstable flow at the entrance of the impeller, which further complicates the situation. This instability causes unfavorable effects on the stall point and increases noise and vibration levels, which can lead to fatigue cracks in inlet ducts as well as in the rotor [9]. Also, Jack [13] found that centrifugal fans that operate at
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inefficient lower volumes are subject to rotating stall or surge, which wastes power and generates excessive low frequency noise. Bhope and Padole [14] investigated the noise level and fluid flow in a centrifugal fan impeller. The present paper suggests the fitting of annular straighteners at the entrance of the impeller in order to rectify the non-uniformity of the flow and to eliminate the vortices that are generated by the existence of inlet distortion. The main objective of this work is to assess the use of these straighteners on the fan performance. For this purpose, two different types of straighteners, circular and corrugated (zigzag), are considered with different sizes. The investigation is conducted on five different impellers with different exit blade angles. Comparisons with free inlet fans are performed. Measurements of static head, shaft power and noise levels at different loads are conducted for the different cases. The analysis of these measurements gives some information concerning the operating range and surge margin for these types of fans. 2. Test facilities and instrumentation Experimental investigation were conducted in the Turbomachine Laboratory of the Mataria Faculty. The test rig consists of a low pressure commercial centrifugal fan of the radial type, a test inlet duct and a delivery duct. The fan wheel is comprised of 16 straight blades of 3 mm thickness with constant blade width of 60 mm welded to a back plate and a shroud. The impeller inner and outer diameters are 215 mm and 394 mm, respectively. The scroll casing is of constant rectangular width. The fan is driven by an electric motor of shaft power 3 hp at constant speed of 2800 rpm. The test inlet duct is 160 mm in diameter and 300 mm long. The exit circular duct of 100 mm diameter is connected to the rectangular outlet of the fan through a conical connection and fitted at the end with a throttle valve. Fig. 1 illustrates the test rig layout equipped with the measuring devices. In this investigation, the suggested straighteners for overcoming any tip or circumferential distortion are located in the inlet duct at a distance of 30 mm from the impeller entrance. Two types of straighteners are designed, both with constant annular cross section of inner and exit diameters 45 mm and 160 mm, respectively. One type consists of annular bundles of plastic or PVC tubes with different diameters, 2.5, 4 and 15 mm. The other type (the zigzag type) is manufactured with the same process used for catalytic converters for car exhaust. Hardened paper foil is corrugated and wound together with non-corrugated foil, making a triangular cross section of height 10 mm. The various layers of corrugated and non-corrugated foils are glued to each other, making the annular shape. The length of the straighteners, considered as a parameter, has been
Circular tube
Static taps
Centrifugal fan
Prandtl probe
Spherical valve
Flow Static taps
Pitot tube Straight blade
Computer Sound level meter
Micromanometer
Multi-channel Static and/or total pressure Switch
Wattmeter
(a) Test rig. Impeller
Straightener
Impeller
Air inlet
(b) Setting location of straighteners. Bellmouth
Fig. 1. Test rig and measuring devices layout.
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L
D
h=10 mm
d=2.5 mm
d=4 mm
d=15 mm
2640 tube
1030 tube
74 tube
Zigzag type
Circular type
Fig. 2. A schematic drawing of the different shape of straighteners.
Total suction head (m water)
varied from 225 mm to 180 mm, making a length to duct diameter ratio, L/D, of 1.4 and 1.125. Schematic drawings show the different shapes of the straighteners in Fig. 2. The average static pressures at the inlet and exit of the fan are measured through four taps equally distributed circumferentially, Fig. 1. The flow velocity distribution across the delivery duct diameter was measured using a standard cylindrical Prandtl probe with inner diameter 2 mm mounted on a traverse mechanism with accuracy ±0.1%. The probe is located at ten diameters from the delivery duct inlet to ensure uniformity of the flow. The flow through the fan is controlled by a spherical regulator valve located at the end of the delivery pipe. In order to check the flow uniformity at the fan inlet downstream of the straighteners, the total pressure distribution is measured by a shielded Pitot tube using a traversing mechanism with accuracy ±0.1%. All the pressures were measured through a multi-channel switch by a digital micro-manometer, model Yokogawa 2655, with resolution of 0.1 Pa and updating of the reading every 0.4 s. An average of the readings is computed every 5 s using an A/D converter and a PC. Fig. 3 shows the total suction head distribution of the radial blade impeller with and without straighteners at the design point. From this figure, it can be seen that the total head at the straighteners exit is approximately constant. Compared with the free inlet, a drop in the suction head is detected by the presence of the straighteners that increases as the diameter of tubes decreases.
0.00 -0.05 -0.10
Free d=2.5 mm
-0.15
d=4 mm d=15 mm
-0.20 -0.25 0.0
0.2
0.4
0.6
0.8
1.0
r/ro
Fig. 3. Total suction head distribution at the inlet of the radial fan with and without straighteners at design point.
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Table 1 Characteristics of the different impellers Parameter
Original impeller
Impeller I
Impeller II
Impeller III
Impeller IV
Outlet angle, b2 (deg) Inlet angle, b1 (deg) Blade length (mm)
90 90 80
60 25 86
75 60 84
105 125 84
120 150 86
Delivery Static Head (m water)
The shaft power of the fan is measured by a digital wattmeter with accuracy ±0.09%, while the rotational speed is measured by a digital tachometer, model Lutron Dt-2236, with accuracy ±0.05%. The noise level in dB, measured by sound pressure level of the fans with different inlet configurations, is determined. A portable sound level meter equipped with a special stand and set to A-weighting (slow response) is used. Three different near field measuring locations have been chosen at a standard distance equal to twice the impeller housing diameter in accordance with DIN 45635: at fan inlet, near the delivery duct exit and behind the fan motor. The noise level was always found to be maximum near the exit of the delivery duct, and therefore, measurements were recorded and are presented only at this station. The effect of the straighteners on fans with different exit blade angles has also been investigated. Accordingly, four new impellers, two backward and two forward facing, with different exit blade angles have been
0.30
β2=90º
0.25
Free
0.20
d=2.5 mm
0.15
d=4 mm d=15mm
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Zigzag
0.05 0.00 0.00
0.05
0.10
0.15
0.20
0.25
0.30
0.35
0.25
0.30
0.35
0.25
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0.35
V (m3/s)
Static Efficiency%
40.0 30.0 20.0 10.0 0.0 0.00
0.05
0.10
0.15
0.20
V (m3/s)
Shaft Power (kW)
2.0 1.5 1.0 0.5 0.0 0.00
0.05
0.10
0.15
0.20
V (m3/s)
Fig. 4. Fan performance for radial impeller (b2 = 90°) with different straighteners.
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constructed using the original scroll housing. More details about the different impellers are tabulated in Table 1. 3. Experimental results and discussion
Delivery Static Head (m water)
The characteristic curves of the five different tested impellers are shown in Figs. 4–8. The measured delivery static head, together with the calculated static efficiency and the shaft power, are plotted versus the volume flow rate for each fan with the different types of inlet straighteners of L/D = 1.4. Comparisons with the free inlet condition were performed on the same plots. The performance of the radial fan using different straighteners is shown in Fig. 4. At large flow rate, a remarkable increase in static head due to the straighteners can be noticed after 0.15 m3/s. This is accompanied by an appreciable improvement in the fan efficiency. By decreasing the flow rate, the effect of the straighteners vanishes. As the diameter of the straightener tubes increases, a more flattened efficiency curve is observed. An improvement of 5 points in efficiency, corresponding to a relative increase of 18%, is obtained with straighteners of 15 mm diameter, and a corresponding decrease in shaft power is noticeable. This is due to the good guidance of the flow provided by the straighteners at the impeller inlet. Furthermore, the unstable operating range of the fan extends farther due to the straighteners. It is useful to note that during the experiments, the surge point was detected by fluctuations in the pressure readings in addition to the high audible noise. The 0.30
β2=60º
0.25
Free d=2.5 mm
0.20
d=4 mm
0.15
d=15 mm
0.10
Zigzag
0.05 0.00 0.00
0.05
0.10
0.15
0.20
0.25
0.30
0.35
0.25
0.30
0. 35
0.25
0.30
0.35
V (m3/s)
Static Efficiency %
40.0 30.0 20.0 10.0 0.0 0.00
0.05
0.10
0.15
0.20
V (m3/s)
Shaft Power (kW)
2.0 1.5 1.0 0.5 0.0 0.00
0.05
0.10
0.15
0.20
V (m3/s)
Fig. 5. Fan performance for backward impeller (b2 = 60°) with different straighteners.
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Delivery Static Head (m water)
results of the noise test in Fig. 9, expressed in sound pressure levels in dB, (A) show that the use of straighteners decreases the noise level along the operating range by approximately 5 dB at the high flow rate. In practice, an overall sound power increase of 3 dB is just perceptible to the human ear and 5 dB is clearly louder. The performance of the backward fans with 60° and 75° exit blade angle is shown in Figs. 5 and 6, respectively. The first impeller exhibits improvement that could reach about 6 points in efficiencies employing the zigzag straighteners. This represents a relative increase of 21% in efficiency. However, a small drop in the static head associated with an increase in shaft power is observed. A reduction of about 3 dB in noise level is the result, Fig. 9. It worth noting that the noise level increases as the blade angle increases, which was also detected by Liberman [15]. As the blade angle increases to 75°, a lower efficiency is obtained all over the operating range at the free inlet condition. This is due to the high incidence losses resulting from the corresponding large inlet blade angle. Using the inlet straighteners leads to a further drop in static head as well as in efficiency. However, the use of straighteners with very small tube diameter increases, obviously, the efficiency but on account of the low delivery head. This is associated with a noticeable decrease in the maximum flow rate, choke point. However, the onset of the surge point shifts to a lower flow rate. Figs. 7 and 8 indicate the performance of the forward fans with exit blade angle 105° and 120°. The use of inlet straighteners result in small increases in static efficiency for 105° on account of the appreciable drop in static head, whereas, for the impeller with 120°, a deterioration in efficiency as well as in delivery head is noticed. In this case, it is worth noting that at the free inlet condition, the fan efficiency is already very 0.30 0.25
β2=75º Free
0.20
d=2.5 mm
0.15
d=4 mm
0.10
d=15 mm
0.05
Zigzag
0.00 0.00
0.05
0.10
0.15
0.20
0.25
0.30
0.35
0.25
0.30
0.35
0.25
0.30
0.35
V (m3/s)
Static Efficiency %
40.0 30.0 20.0 10.0 0.0 0.00
0.05
0.10
0.15
0.20
V (m3/s)
Shaft Power (kW)
2.0 1.5 1.0 0.5 0.0 0.00
0.05
0.10
0.15
0.20
V (m3/s)
Fig. 6. Fan performance for backward impeller (b2 = 75°) with different straighteners.
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3314
β2=105
0.30
Free
0.25
d=2.5 mm
0.20
d=4 mm
0.15
d=15 mm
0.10
Zigzag
0.05 0.00 0.00
0.05
0.10
0.15
0.20
0.25
0.30
0.35
0.25
0.30
0.35
0.25
0.30
0.35
V(m3/s)
Static Efficiency %
40.0 30.0 20.0 10.0 0.0 0.00
0.05
0.10
0.15
0.20
V(m3/s)
Shaft Power (kW)
2.0 1.5 1.0 0.5 0.0 0.00
0.05
0.10
0.15
0.20
V(m3/s)
Fig. 7. Fan performance for forward impeller (b2 = 105°) with different straighteners.
low. This is probably due to separation of the flow inside the impeller passages as the number of blades is much lower than usual for forward facing blades. This is associated with a higher noise level compared to radial and backward facing blades. This is in agreement with the results of Liberman [15]. To assess the effect of the straighteners on the fan operation, some parameters should be taken into consideration. These parameters are the flow margin and the surge margin. The flow margin is defined as V sp Flow margin ðFMÞ ¼ 1 100% V max where Vsp and Vmax are the volume flow rate at the surge point and the maximum flow rate, respectively. The surge margin is calculated from the definition given by Cumpsty [16] as " P !# Surge margin ðSMÞ ¼
PV sp 1 V op
where P is the static pressure ratio and the suffix op indicates operating point corresponding to the condition at maximum efficiency. Figs. 10 and 11 show the calculated flow margin and the surge margin, respectively, for the different fans with different straighteners compared to the free inlet condition. The results show that the flow margin may be
Delivery Static Head (m water)
N.N. Bayomi et al. / Energy Conversion and Management 47 (2006) 3307–3318 0.30
3315
β2=120º Free
0.25
d=2.5 mm
0.20
d=4 mm
0.15
d=15 mm
0.10
Zigzag
0.05 0.00 0.00
0.05
0.10
0.15
0.20
0.25
0.30
0.35
0.25
0.30
0.35
0.25
0.30
0.35
V(m3/s)
Static Efficiency %
30.0
20.0
10.0
0.0 0.00
0.05
0.10
0.15
0.20
V(m3/s)
Shaft Power (kW)
2.0 1.5 1.0 0.5 0.0 0.00
0.05
0.10
0.15
0.20
V(m3/s)
Fig. 8. Fan performance for forward impeller (b2 = 120°) with different straighteners.
arbitrarily increased up to 12% by using inlet straighteners for the backward and radial impellers. For the forward blade, this reduces the flow margin, especially with small diameter. Great improvements in the surge margin are depicted for the different blade angles when using the different straighteners, Fig. 11. From the previous results, it can be deduced that the effect of the straighteners on the fan performance varies according to the exit blade angle. Accordingly, the most suitable straightener for each impeller type can be selected. Straighteners with tube diameter 15 mm conform well to the radial blades, whereas the zigzag type is advisable to be used with backward blades. It follows from the results analysis that inlet straighteners are not convenient for forward blades. The effect of straightener length on the fan performance has been studied. Samples of the results obtained using short straighteners of L/D = 1.125 are presented in Fig. 12 compared to the longer one taking into consideration the most efficient inlet configuration for each fan. It can be noted that for the radial and backward impeller with blade angle 60°, decreasing the length of the straighteners weakened the performance of the fan, whereas for blade angle 75°, the shorter zigzag type of straighteners improves the fan performance. 4. Conclusion The present paper investigates the effects of inlet straighteners on the performance, operating range and instantaneous surge of a centrifugal fan. Experimental investigations concerning different types and sizes of
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Free d=2.5 mm
85
d=4 mm d=15 mm Zigzag
80 0.00
0.05
0.10
0.15
0.20
0.25
0.30
0.35
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V (m3/s)
Noise (dB)
90
β2=60º 85
80 0.00
0.05
0.10
0.15
0.20
V (m3/s)
Fig. 9. Noise level for different straighteners for radial and backward (60°) impellers.
Fig. 10. Comparison of the flow margin for different blade angles with different straighteners.
inlet straighteners for radial, backward and forward fans were conducted. The following conclusions can be drawn: 1. The effect of straighteners on the fan performance depends mainly on the exit blade angle. More flattened efficiency curves are obtained by increasing the straightener tube diameters. An improvement of 5 points in efficiency corresponding to 18% relative increase in efficiency is obtained using circular tube straighteners with 15 mm diameter and L/D = 1.4 for a radial impeller. A relative increase of 21% in the efficiency of a backward fan of 60° blade angle associated with a small drop in delivery head is obtained when using straighteners of zigzag type. A bad effect for the different straighteners is observed on the fan performance for the forward impeller. A slight effect is noted on the maximum permissible flow rate (choke point) for the radial fan, while for backward and forward blades, it decreases by using straighteners. 2. The flow margin increases up to 12% for backward and radial impellers. 3. Improvements of surge margin are depicted for the different blade angles.
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Fig. 11. Comparison of the surge margin for different blade angles with different straighteners.
Circular type (d=15 mm)
β2=90º
Deivery Static Head (m Water)
0.30
Zigzag type
β2=60º
0.30
0.25
0.25
0.25
0.20
0.20
0.20
0.15
0.15
0.15
0.10
0.10
0.10 Free
Free
0.05
0.05
L/D=1.4
L/D=1.125
0.00 0.1 0.2 V(m3/s)
0.00 0.0
0.3
0.1 0.2 V(m3/s)
0.0
0.3
40.0
40.0
40.0
30.0
30.0
30.0
20.0
20.0
20.0
Free
10.0
Free
10.0
L/D=1.4
L/D=1.4
L/D=1.125
L/D=1.125
0.0
10.0
0.1 0.2 V(m3/s)
0.3
0.1 0.2 V(m3/s)
0.3
Free L/D=1.4 L/D=1.125
0.0
0.0 0.0
L/D=1.4
L/D=1.125
0.00
Static Efficiency %
Free
0.05
L/D=1.4
L/D=1.125
0.0
Zigzag type β2=75º
0.30
0.0
0.1 0.2 V(m3/s)
0.3
0.0
0.1 0.2 V(m3/s)
Fig. 12. The effect of straighteners length on fan performance.
0.3
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4. The use of straighteners decreases the noise level by approximately 3–5 dB at high flow rate compared with the free inlet condition for radial and backward impellers. 5. The effect of straighteners length varies with exit blade angles. For the backward impeller with blade angle 60°, as well as for radial fans, the longer zigzag and circular straighteners, respectively, give better performance, whereas for blade angle 75°, the shorter zigzag is the best.
References [1] Ariga I, Kasai N, Masuda S, Watanabe Y, Watanabe I. The effect of inlet distortion on the performance characteristics of a centrifugal compressor. Trans ASME, J Eng Power 1983;105:223–30. [2] Ariga I, Masuda S, Okita A. Inducer stall in a centrifugal compressor with inlet distortion. ASME paper 86-GT-139, 1986. [3] Graber EJ, Braithwaite WM. Summary of recent investigations of inlet distortion effects on engine stability. AIAA paper no. 74-236, 1974. [4] Greitzer EEM. The stability of pumping systems. Trans ASME, J Fluid Eng 1981;103:193–242. [5] Baghdadi S, Lueke JE. Compressor stability analysis. Trans ASME, J Fluid Eng 1982;104:242–9. [6] Wright T, Madhavan S, Di Re J. Centrifugal fan performance with distorted inflows. Trans ASME, J Eng Gas Turb Power 1984;106(October):895–900. [7] Madhavan S, Wright T. Rotating stall caused by pressure surface flow separation on centrifugal fan blades. ASME paper 84-GT-35, 1984. [8] Madhavan S, Wright T. Rotating stall caused by pressure surface flow separation on centrifugal fan blades. Trans ASME, J Eng Gas Turb Power 1985;107(July):775–81. [9] Chen P, Soundra-Nayagam M, Bolton AN, Simpson HC. Unstable flow in centrifugal fans. Trans ASME, J Fluids Eng 1996; 118(March):128–33. [10] Montazerin N, Damangir A, Mirian S. A new concept for squirrel-cage fan inlet. Proc Instn Mech Eng 1998;212(Part A):343–9. [11] Kassens L, Rautenberg M. Flow measurements behind the inlet guide vane of a centrifugal compressor. ASME paper 98-GT-86, 1998. [12] Coppinger M, Swain E. Performance prediction of an industrial centrifugal compressor inlet guide vane system. IMechE, Part A 2000;214:153–64. [13] Jack BE. Fan selection and to reduce inefficiency and low frequency noise generation. In: Fan noise 2003, International symposium Senlis, 23–25 September. [14] Bhope DV, Padole PM. Experimental and theoretical analysis of stresses, noise and flow in centrifugal fan impeller. Mech Mach Theory 2004;39(12):1257–71. [15] Liberman MYu. Investigation of noise characteristics for centrifugal ship fans. In: XI session of the Russian acoustical society, Moscow, November 19–23, 2001. p. 572–5. [16] Cumpsty NA. Compressor aerodynamics. New York: John Wiley and Sons, Inc.; 1989.