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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
[email protected]; [email protected];
[email protected]
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.
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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].
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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
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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
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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
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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].
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- 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
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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
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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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