Throughout industry, the need to move large quantities of air or convey products is ever present. To meet the air and product handling requirements, a variety of fans and blowers is most often employed. It is rare not to find one or more fans or blowers in each department of an industrial or manufacturing complex.
In many ways fans and blower noise is one of the easiest and most straight-forward acoustical problems to solve. We shall now develop a systematic approach to controlling fan and blower noise. Fans will be considered first, followed by blowers. In this text, blowers will refer to the high-pressure rotary positive displacement type which can better be described as compressors.
The fans we shall focus attention on are those used to move large volumes of air for ventilation, dust or oil mist collection, drying operations, etc., which are relatively low-speed low-static-pressure units. The majority of fans can be classified as either axial or centrifugal and we shall deal with each on an individual basis.
There are two basic types of centrifugal fans, backward or forward curved and radial. 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. Now the noise from centrifugal fans is dominantly a superposition of discrete tones at the impeller or blade passing frequency and broadband aerodynamic noise.
The origin of the discrete tones is from two sources. First, each time a blade passes a point in space, a pressure fluctuation is created due to the displacement of air and/or aerodynamic lift if the blade is of an airfoil configuration. Second, as the blades pass the cutoff point in the scroll abrupt pressure changes or pulses also occur at the blade passing frequency and higher integer-ordered harmonics.
The amplitude of the discrete noise is difficult to predict; however, some design guidelines are available:
1. The cutoff clearance is critical, i.e., the clearance between the blade and scroll at the cutoff point. A clearance of 5 per cent to 10 per cent of the wheel diameter is considered optimum by most manufacturers.
2. Backward-inclined blades are generally quieter than forward-inclined blades.
The broadband aerodynamic noise originates from vortices created at the leading and/or trailing edge of the blades and turbulence imparted to the fluid, usually in the form of eddy like flow. Here again the accurate prediction of noise levels for these fans is at best very difficult, but an empirical approximation which provides good first order results for the average sound power level in the range of 500 to 4000 Hz. Spectral sound power corrections for estimating fan noise is given in Table 20.1.
The sound power of centrifugal fans varies dynamically with operating efficiency or performance, wheel speed, scroll design, etc. Fortunately, rather reliable spectral sound power levels are provided today by most of the manufacturers with elaborate correction tables for performance, speeds, etc. As such the estimation method just outlined should be used only in the absence of commercially tested or measured data.
Effective noise control of centrifugal fans is not complicated by a variety of measures or approaches. Centrifugal fans are basically low-pressure high-flow-volume devices. Therefore, the fundamental approach is the utilisation of absorptive, parallel or circular baffle type silencers.
The features of this type of silencer are good high-frequency attenuation and minimal aerodynamic pressure loss. Often the problem of pressure loss is more difficult to resolve than the noise reduction because of the extreme sensitivity of centrifugal fans to upstream or downstream flow restriction.
Illustrated in Fig. 20.3 is a simple approach which is most often taken in industrial installations. Here a tubular silencer is installed on the inlet of the fan through an adapter section and a parallel baffle duct type silencer is installed on the exhaust. A flexible coupling of dense material adapts the silencers to the fan and provides vibration isolation between the fan and the ductwork. Naturally, vibration mounts are provided to isolate the fan from the floor or support platform.
In air-conditioning or ventilating installations, the fan is often installed in a mechanical room. Here the intake silencer is a modular assembly of parallel baffles which extends from the floor to the ceiling of the room. The fresh air, after passing through the silencers, enters the heat exchanger and into the inlet of the fan.
The exhaust treatment is again typically a parallel baffle duct silencer. Note again that the flexible couplings and vibration isolators should be used to mechanically decouple the rotating machinery from the duct system and floor.
Axial fans take their name from the fact that the airflow is along the axis of the 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.
Axial fans generally operate at higher pressures than centrifugal fans and usually are considerably noisier. Common applications include heating, ventilating and air-conditioning systems and forced or induced draft fans for steam boilers.
Because of the number of blades (8 to 30 are typical) and relatively high rotational speeds, the noise from axial fans is generally characterised by strong discrete blade passing tones. The procedure for estimating the average power level follows identically the method as given for centrifugal fans and hence will not be repeated.
Further, to obtain a spectral power level for design purposes, the corrections for the octave bands are given in Table 20.2. Again, the method of calculation of octave band power levels proceeds identically as outlined for centrifugal fans.
It must be emphasised again that these estimates of sound power should be considered, at best, first order and the use of commercially tested or manufacturer’s data is strongly preferred.
Another question that frequently arises is ‘What is the change in sound power level if a fan speed is changed’?
First-order estimates can be obtained in conjunction from the following fan laws:
1. Volume flow (Q) varies linearly as fan speed.
2. Static pressure (Pt) varies as the square of fan speed.
In addition to the aerodynamic and discrete blade noise, another noise mechanism due to the blade-vane wake interaction may also be present in axial fans. Before discussing the interaction noise, it is essential to understand a few fundamentals of the aerodynamics and noise generation mechanism of the rotor- only noise. To visualise these phenomena, consider first a single-stage axial-type fan which has four blades and is rotating in a homogeneous uniform fluid medium such as air.
The aerodynamic forces or pressures associated with each blade can be resolved into two components:
(i) A drag force due primarily to the thickness of the blades;
(ii) A lift force due to the aerodynamic (airfoil) shape of the blades.
Now, as the fan turns, a rotating pressure pattern is formed with pressure lobes at each blade. If a microphone is placed near the rotor, say a few inches upstream, it is easy to see that periodic pressure fluctuations would be felt each time a blade passed.
From the periodic nature of the disturbance, the rotating lobed pressure pattern can be visualised as a superposition of harmonically related sinusoidal patterns rotating at shaft speed. The rotor-only pressure pattern model is now established. With a narrow-band analyser, the spectral fourier components can be separated into a fundamental tone and integer-order higher harmonics.
If one were to be positioned on a blade of this fan, it would be easy to see that no dynamic pressure fluctuations would be felt during rotation. However, if the airflow to the fan were distorted, say due to the presence of an upstream guide vane or bearing support, and then every time a blade passed through the wake, a pressure fluctuation would be felt by the riding observer. This fluctuation is often erroneously thought of as a ‘slap’ as the blade impacts the wake.
Actually, the wake presents velocity gradients which affect the aerodynamic performance of the blade. That is, there is a fluctuation of the lift force. To be sure, the magnitude of the fluctuations depends on the size, proximity and aerodynamic influence of the obstacle, but it has the common property of recurring every time the blade completes a revolution. From this conceptual model, one might expect that the resultant interaction noise would be discrete tones at the vane or wake passing frequency. Such is not so.
It can, however, be shown that the resultant noise is at the usual blade passing frequency and integer harmonics thereof. In short, no new tones are produced. As such, it is difficult to separate and determine the magnitude of the rotor-only or interaction noise.
Qualitatively, however, it is important to note that the resultant interaction pressure patterns are also lobe type in character, with rotation related to shaft speed in such a way that discrete tones are generated only at the blade passing frequency and integer harmonics.
The most important conclusion of this discussion can now be stated. Certain combinations of blades and vanes produce rotating patterns which rotate at speeds many times the original shaft rotational speed. Now it can be shown that as the interaction pattern phase velocity increases, the radiated sound power typically increases.
In fact, as the circumferential linear speed of a pattern approaches Mach 1, a rapid increase in noise level is frequently observed. At this critical point, the interaction pattern is said to be at cutoff and generally maximum acoustical power is radiated. This phenomenon is well documented for axial flow compressors and propellers and recent investigations have shown similar effects on vane axial fans.
Another noise mechanism which is a dominant source of noise in axial machines is Karman vortex noise. The mechanism behind this source of noise has its origin in one of the earliest recorded references to sound. As the wind blew through the pillars of the ancient Greek temples, eerie discrete tones or whistles were produced. The Greeks felt that these whistles were the voice of the God of the wind, Aeolus and hence the tones were called Aeolian tones. Every sailor knows these tones well as the wind blows through the rigging.
The origin of the periodic sound can be visualised by considering the flow of air over a cylinder. At a given velocity, the flow pattern behind the cylinder (downstream) exhibits an oscillatory motion as vortices are shed alternately on one side and then the other. The trail of ‘eddies’ forms what is called a Karman vortex street which possesses strong periodic components.
An adjacent segment farther out radially would also produce a pure tone but at a sightly higher frequency due to the higher linear velocity. Continuing this reasoning over the entire blade to the tip and for each blade of the fan, it is easy to see that the resultant noise would be broadband in character. In actual practice, vortex noise from fan blades is broadband, but the bulk of acoustical energy is concentrated in a narrow frequency range. This is especially true for relatively low-speed fans.
Since axial fans are mounted in ducts with cylindrical geometry, one basic approach to noise control is generally followed. A tubular absorptive silencer is installed on the inlet and exhaust. The silencer selection methods follow those guidelines presented dealing with silencers and will not be repeated except to say that the fan should be decoupled from the ductwork and the floor, ceiling or platform, etc. In that axial fans generally operate at higher rotational speeds than centrifugal fans, vibration isolation is inherently easier.
Silencer selection guidelines with respect to aerodynamic constraints (pressure loss) can be summarised simply as follows:
1. Calculate the volume flow or face velocity and compare it to the silencer manufacturer’s pressure loss specifications.
2. A good rule of thumb is to size the silencer such that the open flow-through cross section is 1.25 to 1.5 times the fan duct cross-sectional area.
Manufacturers make no distinction between fans and blowers. However, in industry the term blower is usually reserved for high-pressure devices most often used for conveying materials or products.
There are two basic types:
1. Centrifugal fans with a radial wheel construction operated at high speeds.
2. Rotary positive displacement blowers.
The second type is distinctly different and we shall deal with these machines in detail.
Basically rotary displacement blowers consist of two counter rotational lobelike gear impellers.
As each impeller lobe passes the blower inlet, it traps a definite volume of air or gas and carries it around the case to the blower outlet. This cycle repeats itself 4 times with every complete revolution of the driving shaft. Hence, the character of the noise is dominantly periodic with discrete tones at the compression frequency and integer-ordered fourier harmonics thereof.
For the two-impeller configuration, the frequencies of the discrete tones fn can be calculated from-
N = shaft rotational speed (rpm)
The pressure pulses from these blowers are quite severe and sound level measurements near the inlet and exhaust ports often exceed 140 dB. Hence, the character of the noise is dominantly periodic with discrete tones at the compression frequency and integer-ordered Fourier harmonics thereof.
Because of the discrete and dominantly low-frequency character of blower noise, reactive chamber- type silencers are especially effective. In most installations, a silencer will be required on both the inlet and exhaust.
In addition, an inlet filter may also be a requirement. Fortunately, many filter manufacturers offer a line of combination filter-silencers. As such, including a filter-silencer in series with the inlet silencer provides additional noise reduction.
Well-designed reactive silencers properly sized to the blower will have pressure losses on the order of a few inches of H2O. These losses are usually trivial since most blowers provide static pressures on the order of 10 to 20 psi. It is extremely important to provide vibration isolation between the blower and pipe ducts and also between the blower-motor unit and the floor or support platform.
Isolation between the blower and the connecting pipe ducts is best achieved by including a high-pressure reinforced flexible coupling in the duct 3 to 5 diameters from the blower. A note of caution here: Frequently the flexible coupling provides a major acoustical leak since the transmission loss through the coupling is lower than the heavy walled pipe. However, this problem can be overcome by simply wrapping the coupling area with a composite of 1 inch polyurethane foam and 1-lb/ft2 dense vinyl.
Finally, even with all the noise reduction measures included as outlined, noise levels in close proximity will likely be in the range of 85 to 95 dBA. The resultant levels are due to general mechanical noise associated with gears and bearings and noise transmitted through the blower case itself.
Therefore, for critical noise sensitive installations, serious consideration should be given to enclosing the blowers. Since they generally require little monitoring or maintenance, a room type enclosure is best, especially when several blowers (often the case) must be similarly enclosed.
In final summary, it should be emphasised that control of fan and blower noise is relatively straightforward. With the design guidelines presented here, even the most stringent design criteria can be met or exceeded with readily available commercial materials and equipment.
It should also be noted that the leading manufacturers of fans and blowers are members of or subscribe to the policies of the Air Moving and Conditioning Association (AMCA) or the American Society of Heating, Refrigerating and Air Conditioning Engineers (ASHRAE). In both organisations, noise measurement techniques for fans and blowers are strictly controlled by standard and as such, reliable sound power data are usually available.