Use of Dissipative Silencers for Fan Noise Control

Alexandre Luiz Amarante Mesquitaa, Andr Luiz Amarante Mesquitab, Ernesto Arthur Monteiro Filhoc

a,b Mechanical Engineering Department, Federal University of Par, Belm, PA, 66075-110,

Brazil c Solve Engenharia, Belm, PA,66110-010, Brazil

aalexmesq@ufpa.br; bandream@ufpa.br; csolve@solveengenharia.com.br

Abstract. Axial and centrifugal fans are very used in industries in general. These equipments have great applicability in the product development as well as ambient comfort. Among the operational problems in these equipments, the noise frequently arises as principal causes. The fundamental approach is the utilization of absorptive, parallel, or circular baffle-type silencer. The features of this type of silencer are good high-frequency attenuation and minimal aerodynamic pressure loss. In this context, this work presents a review of the common noise sources in fans and the procedures for noise attenuation. Finally, an application case is presented to illustrate the use of dissipative silencer.

1. INTRODUCTION

Industrial fans are very used in industries in mine and metallurgical Amazon region, in Brazil, where they are used to move large volumes of air for ventilation, dust collection, drying operations, etc. Among the operational problems in these equipments, the vibration and noise frequently arise as principal causes, which as consequence, these problems can result in low productivity and discomfort. The vibration is caused due mechanical problems and the noise is generated due aerodynamic interactions and due mechanical problems also, as a consequence of vibration. The noise control can be achieved through of actions at the source of the sound waves, modifications on the path or isolating the receiver. In this context, this work reviews briefly the commons causes of noise and vibration in fans and how to eliminate them. Then, this paper discusses the traditional way of noise control on the path, i.e., through the use of absorptive, parallel baffle-type silencer, known as dissipative silencers. An application case to illustrate the use of dissipative silencer is presented.

2. INDUSTRIAL FANS

In the broad sense, fans are generally understood to be air-moving devices using a centrifugal or axial-flow type of air propulsion. Fans are divided into two general classifications, centrifugal and axial. In centrifugal fans the flow through the impeller is essentially radial outward from an axis of rotation; centrifugal force causes a flow and compression of the mass of air through the rotor. There are three basic types of centrifugal fans, curved backward or forward blades and radial blade, as illustrated in Fig. 1. There are numerous types designed for a wide variety of applications; however, they usually can be considered variations and/or combinations of these basic types [1].

(a) (b) (c)

Figure 1: Examples of rotors of centrifugal fans: (a) Backward curve; (b) Forward curve; (c) Radial Blade. Axial fans take their name from the fact that the airflow is along the axis of the fan. Shown in Fig. 2 is an example of an axial fan. To avoid a circular flow pattern and to increase performance, guide vanes are usually installed downstream of the rotor. Axial fans with exit guide vanes are called vane axial and those without, tube axial [1].

Figure 2: Example of an axial fan [2]. The sound power generated by a fan varies between the fifth and sixth power of the fan tip speed [3]. A reduction in speed can make a considerable difference to the fan noise. It is important to understand that for any given delivery volume, delivery pressure and fan type there is one speed and one diameter at which the efficiency is a maximum; if one runs the fan at any other speed, the efficiency falls and the noise increases. Then, a large fan running at slower speed is not necessarily quieter. If the fan operates other than at its peak efficiency, a further addition has to be made to the sound power level as in Table 1, however manufactures should have available curves of efficiency and the corresponding noise levels [3, 4].

Table 1: Fan efficiency adjustment, i.e., the number of decibels by which the sound power level of a fan should be increased because of its operation at other than peak efficiency.

Airfoil centrifugal and vaneaxial fan

Backward-curved centrifugal fan Forward-curved centrifugal fan

Efficiency %

Increase dB

Efficiency %

Increase DB

Efficiency %

Increase dB

80 to 72 0 75 to 67 0 65 to 58 0 71 to 68 3 66 to 64 3 57 to 55 3 67 to 60 6 63 to 56 6 54 to 49 6 59 to 52 9 55 to 49 9 48 to 42 9 51 to 44 12 48 to 41 12 41 to 36 12

3. FAN NOISE SOURCES

The noise generated by fans can be classified in aerodynamic and non-aerodynamic noise. The non-aerodynamic noise is caused by defects in mechanical components of the machine or due to structural resonances. The most common causes of noise due mechanical problems (non-aerodynamic noise) are: (i) Fan Unbalance: Unbalance is one of the leading causes of vibration in rotating

machinery. Unbalance is simply an unequal distribution of rotor weight along the shaft axis. Some common causes of irregular mass distribution are porosity in casting, non uniform density of material, manufacturing tolerances, gain or loss of material during operation, maintenance actions, etc. Because of these irregularities the actual axis of rotation does not coincide with one of the principal axes of inertia of the body, and variable disturbing forces are produced which result in vibrations and consequently in noise. In order to remove these vibrations and establish proper operation, balancing becomes necessary. The forces generated due to an unbalance are proportional to the rotating speed of the rotor squared. Therefore, the balancing of high-speed equipment is especially important. For a quietest operation, vibration-isolation mounts should be used as well as flexible connections to duct work (see Figure 5). The more perfect the fan balance, the less the likelihood of noise generation from this source.

(ii) Bearing Noise: Well-lubricated sleeve bearings are somewhat quieter than ball or roller bearings. Precision antifriction bearings can be obtained and in the larger units where the fan noise is higher, antifriction bearing are quite satisfactory. Where the ball or roller bearings are damaged or the raceways pitted, a high-frequency noise is usually present and may be detected by a vibration analysis.

(iii) Motor Noise: Noise of magnetic origin may be radiated by the fan if the impeller is mounted directly on the motor shaft. In some low-speed, very quiet installations, the fan is isolated from the fan shaft to reduce this possibility [5]. The usual precautions for isolating the motor feet should be observed. For higher-speed, higher-pressure fans, such a mount is less practical and less important, since the sound emitted is of lower intensity than the fan noise. A frequent source of noise that may give trouble is that of the built-in motor cooling fan. Longer blades of the backward-curved blade type will help in this respect if the direction of rotation is fixed [5].

(iv) Structural Resonance: A wide range of frequencies is present in most fan noise. If the energy in a given band is high and corresponds to the natural frequency of some part of

the fan (generally flat panels) the resulting noise may be radiated efficiently. Added bracing can be used to raise the natural frequency of the part to some higher value or damping material may be applied to it in order to reduce the noise radiation.

The aerodynamic noise generated by fans comprises broadband noise resulting from vortex generation and intake turbulence, on which is superimposed pure tone components related to fan geometry and rotational speed. The non-harmonic aerodynamic noise is related to generation of vortices due the turbulent airflow on solid surfaces, mainly on the blades. When a blade move through the air a pressure gradient is built up across the blade in the direction of its thickness. If the air flow close to the blade is steady, or laminar, this pressure gradient is essentially constant and little noise results. However, with an incorrect designed blade profile, the flow may separate from the suction side of the blade, thus giving rise to rather large eddies. Moreover, this point of separation is variable. Hence, the pressure pattern and eddy formation fluctuate rapidly and cause considerable noise. Also, Von Karman vortices will be shed from the trailing edge of the blade, forming the wake, since this edge must have a finite thickness. Since they are random in size and point of release from the blade, a broadband noise spectrum results. For axial fans the noise due to such vortices increases with the thickness of the trailing edge. For centrifugal fans, this is true only if the air completely fills the space between the blades [5]. Figure 3 illustrates the vortex noise generation. This basic mechanism described previously by which any surface in a flow generates noise will occur to some degree or other in a fan, even if the entry conditions to the impeller are perfect. However, it is common the case where the intake flow to the impeller is itself turbulent due existence of obstacles against this airflow. This kind of flow will increase the turbulence generated from blades and consequently the noise will be increased. Fans should b therefore be placed well downstream of obstacles, valves, corners, and changes of cross-section. Figure 4 shows examples of this type of noise generation [6].

Figure 3: Airflow on a blade. Figure 4: Examples of fan installations producing turbulence[6].

The harmonic noise, commonly called blade noise, is basic to all types of fans. Every time a blade passes a given point, the air at that point receives an impulse. The repetition rate of this impulse the blade-passing frequency determines the fundamental tone of this type of noise. For symmetrically spaced blades, the fundamental frequency is determined by the product of the number of blades and the rpm. Also, multiples of this frequency will be present. For axial fans, increased blade width will, in general, reduce the intensity of the

harmonics. If a vaneaxial fan has the same number of guide vanes as blades, this will accentuate the noise at blade frequency and its harmonics, especially if the vanes are close to the blades [5]. For minimum noise the vanes should not equal the number of blades and the vanes should be spaced as far as practical from the blades. In centrifugal fans the origin of the discrete tones has another source: as the blades pass the cut off point in the scroll, abrupt pressure changes or pulses also occur at the blade passing frequency and higher integer-ordered harmonics. In order to have minimum noise some guidelines are available: (i) a clearance of 5% to 10% of the wheel diameter is considered optimum by most manufacturers; (ii) backward-inclined blades are generally quieter than forward-inclined blades [1].

4. SILENCERS

The noise generated by air/ gas handling/consuming equipment, such as fans, blowers, and internal combustion engines, is controlled in their trajectory through the use of two types of devices [7]:

(i) Active noise control silencers whose noise cancellation features are controlled by various electromechanical feed-forward and feedback techniques;

(ii) Passive silencers and lined ducts whose performance is a function of the geometric and sound-absorbing properties of their components. The passive silencers can be classified in reactive and dissipative silencers

The reactive silencers consist typically of several pipe segments that interconnect with a number of larger-diameter chambers. These silencers reduce the radiated acoustical power primarily through the use of cross-sectional discontinuities that reflect the sound back toward the source. These devices contain no absorbing material but depend on the reflection or expansion of the sound waves with corresponding self-destruction as the basic noise reduction mechanism. The dissipative silencers are the most widely used devices to attenuate the noise in ducts through which fluid flows and in which the broadband sound attenuation must be achieved. They are frequently used in the intake and exhaust ducts connected to industrial equipments such as fans, blowers, etc., and also the ventilation and access openings of acoustical enclosures. They have an allowed pressure drop that typically ranges 125 to 1500 Pa (0.5-6 in. of water) [7]. These devices contain fibrous or porous materials and depend on absorptive dissipation of the acoustical energy. This paper discusses the application of this type of silencer. First, some guidelines are given in order to design such a silencer and then, an application case is presented.

4.1. Dissipative Silencers

The use of dissipative (or absorptive) silencer is the classical solution for fan noise attenuation. These devices transform acoustic energy into heat (i.e., dissipate the acoustic energy) through the use of sound absorbing material in the internal walls. The principal advantages of these devices are: provides good absorption at medium and high frequencies and useful for narrow and broadband noise; however the disadvantages are: performance falls off at low frequencies (i.e., attenuation is strongly frequency dependent) and absorptive material can disintegrate under harsh conditions (protective facing material will reduce this problem). The most common configurations of dissipative silencers include parallel-baffle

silencers, round silencers, and lined ducts [7]. Figure 5 illustrates the installation of these silencers in a centrifugal fan. This paper focuses the discussion on the parallel-baffle silencers (Figure 5 and Figure 6).

Figure 5: Example of devices for fan noise attenuation [1]. Figure 6: Parallel-baffle silencer [2].

4.2. Guidelines for Design of Parallel Baffle Dissipative Silencers

The acoustical performance of parallel baffles depends on primarily of three parameters: length of baffles; thickness of the absorbing material and the spacing of baffles. The acoustical performance of a silencer is directly proportional to its length. The thicker the absorbent materials, the lower the frequencies that can be absorbed. For higher frequencies though, thinner absorbent layers are effective, but the large gap allows noise to pas directly along. This layers and narrows passages are therefore more effective at high frequencies. For good absorption over the widest frequency range, thick absorbent and narrow passages are best [6] (see Fig. 7).

Figure 7: Example of thickness of absorbing material and the spacing of baffles [6].

Other guidelines for design of parallel baffle dissipative silencers are:

- The plane wave motion presents essentially a grazing incidence to the absorbing treatment, and hence little sound is absorbed for this type of sound wave motion. The performance of absorptive silencers can be sharply improved if the line of sight through the silencer is blocked or eliminated, but care must be taken in relation to pressure drop. Various curves and staggered patterns have been design and are commercially available [1].

- In order to attenuate the sound at the low end of the frequency spectrum, the baffle thickness B (see Fig. 8) must not be greater than the wavelength of the frequency under consideration. To provide reasonable attenuation at the high end of the frequency spectrum, the air passage between layers C (see Fig. 8) must be smaller than the wavelength of the frequency under consideration [7].

Figure 8: Silencer with absorbent material in the walls and only one baffle in its interior.

An empirical formula for estimating the linear attenuation is given [8]:

Attenuation = [dB/ft] )/( 6.12 4.1SP (1)

where P is the perimeter of the internal revetment acoustical [in], S is the open cross-sectional area of the duct [in2], and is the Sabine absorption coefficient of the absorbent material [dimensionless].

5. APLICATION CASE

In this section, it is shown an application case of design of a parallel baffle dissipative silencer. This silencer was design to be placed in the intake duct of an axial fan (see Fig. 9). The global sound pressure level generate at a distance of 1m from the intake duct was higher than the permissible level dictate for Brazilian laws. Table 2 shows the frequency spectrum of the sound pressure level. The overall sound pressure level generate was above 100 dB.

Figure 9: Dimensions of the intake duct of an axial fan.

Table 2: Values of sound pressure in [dB] as function of frequency [Hz].

Hz 31.5 63 125 250 500 1k 2k 4k 8k 16k dB 61.6 64.3 73.2 89.5 95.4 95.6 93.6 90.9 88.8 70.4

According with procedures and guidelines given in previous section, the value of baffle thickness B is defined to be 20cm, for mineral wool fibbers used as absorbent material. Again, according with the guidelines, the air passage between layers C has to be smaller than 17.46 cm. The effective width of the silencer is 3674mm minus two times the thickness of the steel walls. Therefore, considering the width of the silencer equal to 3664.475mm and thickness B=20cm, the next step is to determine the distance C and the number and layout of the baffles. After several simulations, the dimensions found are listed in Table 3.

Table 3: Number and dimensions of absorbent layers. Width of the air passage

(C) Thickness of the baffles

(B) Thickness of the absorbent in

the walls (B/2) Number of baffles

16.20 cm 20 cm 10 cm 09 The layout of the baffles in order to block the line of sight through the silencer is shown in Fig. 10. The total length of the baffles is 2.05 m.

Figure 10: layout of the baffles. After the installation of the silencer, the sound pressure level was measured again. The resulting spectrum of frequencies is shown in table 4. The new overall sound pressure level was measured be 80.2 dB, which is according with the Brazilian laws for 8 hours of noise exposure.

Table 4: Values of sound pressure in [dB] as function of frequency [Hz].

Frequency[Hz] 31,5 63 125 250 500 1k 2 k 4k 8 k Sound Pressure (dBA) 53,5 61,4 62,6 72,8 76,3 74 69,7 66,2 61,1

6. CONCLUDING REMARKS

Among the operational problems in industrial fans, the vibration and noise frequently arise as main causes, resulting in low productivity and discomfort. In this context, this work reviews briefly the commons causes of noise and vibration in fans and how to eliminate them. Then, this paper discusses the traditional way of noise control through the use of absorptive, parallel baffle-type silencer, known as dissipative silencers. Some guidelines are given in

order to design correctly a dissipative silencer. In the last part of the paper a successful application case of use of dissipative silencer is presented

REFERENCES

[1] L.H. Bell and D.H. Bell, Industrial Noise Control - Fundamentals and Applications, Marcel Dekker, Inc., New York, 1993.

[2] S. Gerges, Rudo Fundamentos e Controle, NR Editora, 2000.

[3] A. Barber, Handbook of Noise and Vibration Control, Elsevier Science, 1992.

[4] D. Whicker et al., Noise & Vibration Control in Mechanical Systems, ASHRAE SMACNA & MCA Workshop, 1994.

[5] C. Harris, Handbook of Noise Control, McGraw Hill Book Company; New York, 1957.

[6] S. Ingemansson, Noise Control Principles and Practice, Brel & Kjaer, 1986.

[7] L. Beranek and L. Vr, Noise and Vibration Control Engineering Principles and Applications, John Wiley & Sons, 1992.

[8] H. Sabine, The Absorption of Noise in Ventilating Ducts, J. Acoust. Soc. Am., 1940.

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