EP0050621A1

EP0050621A1 - Noise reduction system

Info

Publication number: EP0050621A1
Application number: EP19810900348
Authority: EP
Inventors: Gary H. Koopmann; Wolfgang Neise
Original assignee: Individual
Current assignee: Individual
Priority date: 1980-04-28
Filing date: 1980-04-28
Publication date: 1982-05-05
Also published as: DE3161106D1; WO1981003201A1; EP0039459A1; EP0039459B1

Abstract

Procede et appareil de reduction d'impulsions de pression indesiree dans des dispositifs deplacant des fluides comprenant au moins un resonateur (30) pour reduire selectivement une partie de l'intensite acoustique passant par la pale du dispositif, ce resonateur etant place au point de la plus petite separation entre l'enveloppe (12) et l'impulseur (14).Method and apparatus for reducing pulses of unwanted pressure in devices displacing fluids comprising at least one resonator (30) for selectively reducing a part of the acoustic intensity passing through the blade of the device, this resonator being placed at the point of the smallest separation between the casing (12) and the impeller (14).

Description

NOISE REDUCTION SYSTEM

TECHNICAL FIELD

This invention relates to the construction and operation of fluid moving devices (pumps, fans, blowers, turbines, etc.) where fluid is accelerated by a moving surface immersed in, and acting against, that fluid. More particularly the invention relates to fluid moving devices in which the generation of unwanted pressure impulses is reduced or nullified. Such pressure impulses are generated when the moving surface or surfaces pass near fixed surfaces on or within the device. Such unwanted pressure impulses substantially increase the amount of power the device requires to move the fluid. By positioning an appropriately tuned acous¬ tic resonator on or near the region of the device where the moving and fixed surface interact, the unwanted pressure impulses in this region can be cancelled, thus increasing device efficiency.

BACKGROUND ART

Wasted power in fluid moving devices due to generation of undesired pressure impulses is generically known as noise. The dominant component of the noise spectrum of axial or radial flow machines comprises aerodynamically or hydrodynamically generated noise. Mechanical sources of noise, i.e., bearings, drive trains or transmissions, are not very significant. Aerodynamically or hydrodynamically generated sounds may be divided into two compo¬ nents; namely, the blade passing frequency tone and the background noise. The blade passing frequency tone is produced by the interaction of the stationary casing of the machine with the high velocity fluid leaving the impeller. The dominant source of the fluid velocity fluctuations produced by such interactions is located at the cutoff tongue nearest the rotating impeller blades. In the case of radial fluid moving devices with vaned diffusers, the stationary guide vanes behind the impeller are the main acoustic source of the blade passing frequency tone. Stationary guide vanes are referred to as cutoff tongues in this application. The blade passing frequency tone spectrum is discrete. It has large amplitude peaks at the blade passing frequency, which is the product of rotational frequency and number of blades. Additional peaks of acloustical energy occur at harmonics of the blade passing frequency.

The random noise component of the audio spectrum is caused by turbulent flow in the vicinity of the impeller and the stationary casing. Another component is contributed by vortex shedding from the trailing edges of the impeller blades.

Amplitude reduction of the blade passing frequency tone of fluid moving devices is of particular importance both because con¬ ventional noise reducing apparatus constitute a waste of significant fan power and because noise containing specific tonal components is much more displeasing to the human ear than noise having a broad frequency spectrum. A large number of attempts have been made to reduce the noise output of centrifugal fans. The underlying physical principles of these attempts are basically the same for the various types of radial fluid moving devices, such as fans, pumps, blowers, turbines and the like. For example, it is old in the art to in¬ crease the distance between the impeller and the cut-off, thereby reducing the tone intensity. This most frequently used method does, however, require an expensive enlarged fan casing and is thus restricted to applications where no space limitations exist.

Similarly, it is old in the art to increase the radius of curvature of the cut-off tongue in order to decrease the blade passing frequency tone part of the noise spectrum. The reduction in blade passing frequency noise level resulting from this increase in radius is limited, however, to approximately 3 to 5 dB, for most impeller speeds.

Altering the angle of inclination between the impeller blades and the cut-off is also known to result in a tone intensity, reduction in the order of 0-16 dB, depending on the cut-off clear¬ ance. However, inclined blades are difficult to manufacture eco¬ nomically and an inclined cut-off requires a complicated geometry of the fan outlet duct which results in flow losses. Thus past attempts have proven not to be very practical from both design and operating considerations.

In the case of small fans, e.g., where the two halves of a double inlet fan are manufactured separately and mounted together afterwards, staggering the blades of double inlet or double row impellers is sometimes used to reduce the tone intensity. In cases of large, welded fan units, however, this type of asymmetric con¬ struction causes thermal stresses and deformations in the backplate of the impeller.

The introduction of transition meshes at the leading and trailing edge of the impeller reduces not only the tone intensity but also the part of the spectrum associated with random fluctuations. A pressure drop results from the presence of the meshes, however, which causes fan efficiency to decrease substantially.

By spacing the blades in an irregular pattern around the hub, the spectrum of the noise associated with the blade passing frequency can be broadened. Although this is desirable from a subjective assessment of the noise, the overall acoustical energy generated by the fan is not reduced by this configuration.

Placement of a wedge-shaped guide belt around the impeller reduces tone intensity in a manner similar to that of the inclined cut-off design. However, the triangular guide belt surrounding the impeller causes frictional losses and decreases fan efficiency.

When a. conventional spiral casing is replaced with a rectangular one, and the cut-off is omitted altogether, the tone intensity is reduced. Due to the insufficient guidance of the flow, however, this design degrades the aerodynamic performance of the fan.

Applying acoustical liners to the interior of the fan casing effectively reduces both the tone intensity and the random noise. However, the main disadvantage of this method is the same as in the case of conventional absorptive silencers. After a certain time of operation, the acoustic material is often removed by the flow. Also, if media loaded with liquid or solid particles are handled, the porous surface of the liner eventually becomes blocked. A mismatch between the acoustic impedances of fan and ducted system, such as by manipulating the length of the fan outlet duct is also known to result in a substantial reduction of the blade passing frequency tone radiated into the inlet duct. This procedure, however, cannot be a standard one for a given fan, because the effect of any modification of say, the fan inlet, also depends on the acoustic impedance of the given duct system at the fan outlet.

Lastly, in U.S. Patent Number 2,160,666 to McMahon, an "air cushion" behind the cut-off was suggested as a means to reduce noise. However, McMahon does not suggest that maximum noise reduction can be accomplished by tuning the cavity to the blade passing frequency tone and/or by suitable selection of the perforate cover of the resonator mouth.

All of the above discussed methods are based on the fact that the production of the tone is known to originate from within a concentrated region around the cut-off of the fan casing. This feature makes the prospect for altering the tone producing mechanism directly at the source an attractive possibility. The tone is produced by the interaction of the mean air flow leaving the impeller with the asymmetric part of the fan casing comprising the cut-off. It is well-known that this mechanism of tone production is extremely sensitive to changes in the geometry of either the cut-off or the impeller.

The present invention obviates the necessity to alter the geometry of the fan, thus eliminating the mechanical restrictions, complications and efficiency reduction of approaches. The principle of the present invention consists of reducing the. tone intensity by placing a quarter-wavelength or Helmholtz-resonator, such that the mouth of the resonators forms the cut-off of the centrifugal fan or compressor. If the resonator is tuned to the blade passing frequency or its harmonics, the pressures generated by the flow at the cut-off region at this frequency are substantially reduced. Addition of the resonator does not require significantly more space. The geometry of the cut-off and fan casing remain the same.

CMPI _ Y.^'?0 DISCLOSURE OF THE INVENTION

An object of the present invention is to reduce the level of the blade frequency tone of fluid moving devices without nega¬ tively affecting the characteristic curve, efficiency or size of the devices. In conjunction with, and partly as a result of, a lower level of the blade frequency tone, an overall reduction in noise level is achieved, particularly with respect to those tones most displeasing to the human ear. A fan system using the present inven¬ tion has a higher overall efficiency than a similar fan system using conventional silencers or attenuators. These objectives are achieved by positioning the mouth of an acoustic resonator on or near the region of the device where moving and the fixed surfaces interact to produce the undesirable pressure pulses. By appropriately tuning the resonator, most undesirable pressure impulses in this region can be cancelled by the pressure impulses generated at the mouth of the resonator. Structurally, the dominant source regions of the blade passage tone within the casing are constructed as resonators. These resonators are tuned to the device's blade passage frequency or its harmonics. This allows the pressure fluctuations generated by the flow leaving the impeller or fan blades to be nullified or reduced directly at their source. The dominant sources for the blade passage tone are the cut-off tongues in the case of vane!ess dif- fusers and the guide vanes in the devices with vaned diffusers. For the purposes of this application both of these sources shall be referred to as "cut-off tongues" since they are both protrusions into the interior of the device's casing.

The principle of noise reduction of the present invention is applicable to blowers, fans, pumps, turbines and compressors of either the radial or the axial type. Experiments with centrifugal blowers have proven that by positioning one or more resonators in the vicinity of or inside the cut-off tongue, the noise level can be very effectively reduced at the blade passage frequency of its harmonics. Experimental reduction of up to 28dB in the audio level at the blade passage frequency have been achieved. One of the major advantages of the present invention consists of obviating the necessity for interfering with the flow geometry and optimal flow design of the device. The geometry of the interior of the device's casing need not be modified and thus the mechanical characteristics and the efficiency coefficient of the device are not adversely affected by the presence of the cut-off tongue resonators. Only the internal geometry of the resonator itself need be variable, if tunable resonators are desired. Tuning to different frequencies may be achieved, for example, by varying the volume of the cavity by means of at least one movable sidewall or rearwall. Thus, the resonator may be constructed as a Helmholtz resonator whose resonance frequency is a function of the volume of the cavity or, alternatively, as a quarter wavelength resonator the resonance frequency of which is essentially a function of the length of the cavity.

Since the fundamental frequency of the blade passage tone is directly proportional to the rotational frequency of the impeller, in one embodiment of the present invention a regulator responsive to the impeller's rotational velocity is provided that varies the volume or the length of cavity, respectively, in as a direct function of the impeller's angular velocity. This allows the resonator to always be "tuned" to the particular frequency that constitutes the dominant part of the noise spectrum for the device.

In another embodiment of the invention, a plurality of resonators are provided next to each other, each of which is tuned to a different frequency, e.g., one to the fundamental and the others to harmonics of the fundamental frequency.

BRIEF DESCRIPTION OF THE DRAWINGS

FIGURE 1 is a schematic cross-sectional view of a centri¬ fugal fan with a tunable quarter wavelength resonator mounted at the cut-off.

FIGURE 2 is an enlarged view of the cut-off with a quarter wavelength resonator showing details of one configuration of the resonator mouth.

^~:π FIGURE 3 is a graph of the noise reduction attainable with the resonator tuned to give a maximum reduction of the blade passing frequency at 7500 rpm, i.e., 750 Hz.

FIGURE 4 is a simil r graph with the resonator tuned to give a maximum reduction of the blade passing frequency at 6000 rpm, i.e., 600 Hz.

FIGURE 5 is a plot of the level of the blade passing frequency as a function of the fan speed.

FIGURE 6 shows the effect of the quarter wavelength reson¬ ator tuned at 600 Hz upon the overall sound pressure level, the A-weighted sound pressure level and the second harmonic of the blade passage frequency.

. FIGURES 7 to 12 show a series of experimental data similar to those in FIGURE 5, the only difference being the perforate used to cover the resonator mouth. FIGURE 7, for example, shows the level of the blade passing frequency plotted against fan speed for various resonator time settings with the resonator mouth being covered by a 20% perforate cover, containing a number of 2.6 mm diameter holes. In plotting FIGURE 8 a resonator mouth with 42% open area made up holes of 3.8 mm diameter was used, while in FIGURE 9 the resonator perforate employed had 58% open area with individual hole sizes of 4.4 mm diameter. In each case, the total number of holes in the perforate used for plotting FIGURES 7 to 9 was constant.

FIGURES 10 and 11 show the effect of variations in the rate of open area in the perforate while maintaining a constant hole diameter of 3.8 mm. In FIGURE 10, five out of the 43 holes were closed, while FIGURE 11 shows the blade passing tone reduction for the same perforate with 16 of its 43 holes closed.

FIGURE 12 shows the blade passing tone reduction as a function of fan speed for two perforate of the same rate of open area but having different hole diameters and numbers of holes.

FIGURES 13, 14 and 15 are cross-sectional views of radial with different embodiments of a tunable resonator at the cut-off mouth. FIGURE 16 is a systematic cross-sectional view of an axial vane fan with a tunable quarter wavelength resonator installed in a guide vane behind the fan blades.

FIGURE 17 is a view taken along section lines A-A of FIGURE 16.

BEST MODE FOR CARRYING OUT THE INVENTION

Referring now to FIGURE 1, a centrifugal fan 10 is shown, consisting of a blower casing 12 and an impeller 14 housed therein. In the preferred embodiment of the present invention which was used to conduct the tests summarized graphically in FIGURES 3-12, impeller 14 possesses six backward curved blades 17 and is 140 mm in diameter. Casing 12 is of the logon^*thmic spiral type with a very small cut-off distance between impeller 14 and casing 12 in the cut-off region 18. The cut-off clearance 18 is only 4.4 mm and the cut-off radius 20 of the model fan 10 used for test purposes is only 10 mm.

A quarter wavelength resonator 30 is shown in FIGURE 1, positioned at the cut-off point 20. A piston-like slidable plug 32 inside resonator 30 is used to tune the resonator to various fre¬ quencies.

The noise reducing effect of various configurations of resonator 30 was measured by positioning a microphone in the fan outlet duct 22. The microphone in the duct was equipped with a slit-tube to minimize the effect of turbulent pressure fluctuations on the microphone. The microphone signals were analyzed by using a narrow band filter which could be tuned to the blade passing fre¬ quency or multiple harmonics thereof by means of a tracking unit. A real time analyzer was used for taking frequency spectra of the sound pressure in the duct.

Referring now to FIGURE 2, an enlarged cross-sectional view of the cut-off with resonator 30 is shown with cut-off mound 34 being shown as a perforated cover at the resonator entrance. Perforate 34 is a 1 mm sheet metal cover which, in the course of the experiments, was exchanged for various embodiments with differing arrangement, sizes and number of openings 36. The cross-sectional area of reson¬ ator 30 is rectangular with its largest lateral dimension being equal

C P to the width of the fan casing. Upper resonator wall 38, being 46.1 mm wide is 16.5 mm above a lower resonator wall 41 of equal width.

Tuning of resonator 30 was achieved by changing the length of the resonator via a moveable Teflon plug 32 which formed an air¬ tight seal with the resonator walls. Optimum tuning, i.e., maximum noise reduction at a given frequency, was made easy by observing the amplitude of the tone on the realtime analyzer. All sound measure¬ ments were made at only one operating point of the fan which was on the very righthand side of the pressure-head volume-flow charac¬ teristic, i.e., almost at free delivery. To obtain reference data for purposes of comparison, resonator 30 was in some experiments, replaced by a solid piece of wood having the same radius of curvature 20 and the same distance from impeller 16.

Experimental Results

FIGURES 3 and 4 show comparisons between the sound pressure spectra measured in the fan outlet duct for the fan with a conven¬ tional cut-off configuration and with the quarter wavelength resonator mounted. Perforate cover plate 34 consisted of a 1 mm thick sheet metal with holes 36 being 3.1 mm diameter holes totaling up to a 29% open area. For the spectra shown in FIGURE 3, the resonator was tuned to give a maximum reduction of the blade passing frequency at 7500 rpm, i.e., 750 Hz, and, accordingly, at 6000 rpm in FIGURE 4. In both FIGURE 3 and 4 the solid arrows represent the noise levels of the blade passage tone harmonics without a resonator, i.e., with resonator 30 being replaced by a solid piece of wood. The dashed arrows represent the noise levels of the blade passage tone harmonics measured with quarter wavelength resonator 30.

Proper tuning of the resonator does not reduce only the blade passing frequency but also its higher harmonics. However, it may also result in an amplification, as is the case of FIGURE 4 where the third harmonic is increased by 8dB.

The reduction of the blade passing frequency is not the same in FIGURES 4 and 5, and frequency spectra measured at other fan speeds revealed that the maximum possible reduction obtainable by carefully tuning the resonator for each case increased with fre¬ quency.

This effect becomes more evident in.FIGURE 5 where the level of the blade passing frequency is plotted as a function of the fan speed. The sol d curve "A" is for the fan with the conventional cut-off configuration. The other lines shown in FIGURE 5 were mea¬ sured with the quarter wavelength resonator tuned to fan frequencies, i.e. 450, 500, 550, 600, 650, 700 and 750 Hz, respectively. The maximum possible reduction increases with frequency in FIGURE 5, and further tests showed that this can be changed by using a different perforate for the resonator mouth. For the data presented in FIGURE 5, the same perforate was used as in FIGURES 3 and 4. For a given resonance frequency, the damping effect is present not only at this particular frequency or within a very small frequency band but is extended over a fairly broad frequency range. This is important in view of the practical application of this noise control method because it demonstrates that a substantial reduction of the blade passage tone can be accomplished even when the resonator is out of tune, be it due to initial mistuning or due to a change in fan speed as an effect of fan loading or to change in fluid temperature, which corresponds to a change in the speed of sound in the fluid being moved.

FIGURE 6 shows the effect of the quarter wavelength reson¬ ator tuned at 600 Hz upon the overall sound pressure level "A", the A-weighted sound pressure level "B", and the second harmonic of the blade passage frequency "C". The effectiveness of the quarter wavelength resonator with respect to overall or A-weighted noise levels will depend in general on the relative importance of the blade passing tone. In the case of the model fan used for test purposes, the overall sound pressure level was reduced by 3.5 dB, and the A-weighted level by 7 dB.

Another interesting result from FIGURE 6 is that the second harmonic is reduced most at 3000 rpm which corresponds to 600 Hz. However, the reduction is less than that of the blade passing frequenc at the same resonance frequency, i.e., 6000 rpm (compare FIGURE 4 or 5). This indicates that the impedance of the quarter wavelength resonator is changed as a result of the flow velocity along the cut-off.

FIGURES 7 to 12 show a series of experimental data similar to those in FIGURE 5, the only difference being the perforate 34 used to cover the resonator mouth. From FIGURES 7 to 9, the rate of perforation was changed within 20% to 58% by increasing the diameter of holes 36, i.e., the total number of holes is the same in FIGURE 5 and 7 to 11. There is an indication, that increasing the rate of open area shifts the range of maximum reduction of the blade passing frequency towards a lower frequency range.

A physical explanation of the damping association with the frequency shift can be developed in the following way. When the diameter of holes 36 is large, the oscillation of air through the holes is only slightly damped, hence the reduction in the tone amplitude is large. Note however, that this reduction occurs only over a narrower band-width as shown in FIGURE 9. Conversely, when holes 36 are small, the damping rate of resonator 30 is increased which results in a lower tone amplitude reduction, but over a wider frequency range as shown in FIGURE 7.

In FIGURES 10 and 11 an attempt was made to vary the rate of open area while maintaining the same diameter of holes 36 as in FIGURE 8, i.e., 3.8 mm. This was done by closing a number of holes 36 with putty. FIGURE 10 seems to indicate the same trend as before, i.e., reducing the rate of open area shifts the effective range to higher frequencies. However, further reducing the numbers of holes, as in FIGURE 11, did not support this observation.

In FIGURE 12, finally, two perforates 34 are compared which are of the same rate of open area but have different diameter holes 36 and different numbers of holes. This result suggests that the rate of open area is the quantity which determines the frequency range of maximum noise reduction, rather than the hole diameter.

One of the main advantages of the noise reduction method of the present invention consists in the fact that the introduction of resonator 30 requires only minimal changes in fan geometry. As a result, efficiency and flow rates for a fan 10 with and without resonator 30 are expected to be very similar, if not practically identical. In order to empirically determine the difference in aerodynamic performance for these two configurations, the volume flow pumped through test duct 22 was measured. In Table I the ratio of the volume flows delivered with and without quarter wavelength resonator 30 is listed against fan speed, and the results show that the volume flow is diminished by 0.6% on average.-

Table I

Ratio of Volume Flow Delivered by Fan With

Conventional Cut-Off to that Delivered By

Fan With Quarter Wavelength Resonator Mounted

Fan Speed Ratio of

(rpm) Volume Flow

4500 0.9874

5500 0.9941

6000 0.9976

6500 0.9933

7000 0.9975

7500 0.9970

AVERAGE 0.9945

The above outlined test results show that a substantial reduction of the blade passage frequency tone radiated from centri¬ fugal fans can be accomplished by mounting a quarter wavelength resonator at the cut-off.

For a fixed resonator configuration, i.e., keeping the length and the mouth geometry of resonator 30 constant, maximum attenuation of the blade passage frequency occurs at a frequency at which the sound wavelength equals approximately four times the resonator length. A relation for the effective length of the re¬ sonators in terms of the design parameters is difficult to determine because of the curved geometry and perforations making up mouthpiece 34. In addition, the flow over holes 36 also has an effect on the

O PI added mass contribution to the effective length. Attenuation of the blade passage frequency to somewhat lesser degree occurs within a frequency band about the resonance frequency of the resonator. The width of this frequency band is determined by the damping rate of the resonator which itself depends on the size and number of holes 36 in mouthpiece 34.

If the length of resonator 30 is altered, while maintaining the same resonator mouth configuration 34, the resonance frequency is changed, and hence reduction of the blade passage frequency can be achieved at different fan speeds. However, the maximum possible reduction varies with frequency, which indicates that there is an optimum impedance of resonator 30 as a whole which produces a maximum attenuation. The overall impedence of resonator 30 appears to be very sensitive to changes in the flow velocity over resonator mouth 34, and this helps to explain why the amount of noise reduction obtained by proper tuning of the resonator for a given fan speed varies with frequency, or better fan speed. Experiments with differ¬ ent cover plates 34 of the resonator mouth have also shown that this optimum impedance can be generated at different frequencies simply by changing the rate of perforation of resonator mouth 34.

Experiments with the above described preferred embodiments of the present invention have shown that a quarter wavelength reson¬ ator mounted at the cut-off of a centrifugal fan is a highly ef¬ ficient and extremely simple means to reduce the blade passage tone. This method can be used for new fan constructions as well as for reducing the noise in existing installations. Since the use of a quarter wavelength resonator does not require a change in the geo¬ metry of the scroll, the aerodynamic performance of the fan is not affected.

In cases of fans with variable speed, a quarter wavelength resonator can still be used, provided the length of the resonator is changed in proportion with the fan speed, by moving a plug in the resonator, in a slightly different version of the preferred embodi¬ ment of the present invention, plug 32 is, therefore, coupled to the impeller speed by any appropriate servo-mechanical means well known in the art. f eferrexL Three different versions of yet another retiffled embodiment of the present invention are shown schematically in FIGURES 13-15. In FIGURE 13, for example, centrifugal compressor 100, housing impeller 1.16 within casing 112, is equipped with a Hel holtz re¬ sonator 130 in lieu of a quarter wavelength resonator. Helmholtz resonator 130 consists of a front wall 140 substantially tangential to casing 112 but adapted to follow the Tatter's spiral form near the cut-off 120. Its rear wall is formed by a main section 132 sub¬ stantially parallel to front wall 140 which, in its lower portion rear cut-off 120, terminates in a hinge point 136. The end portion 138 of rear wall 130 is ovably hinged below hinge point 136 and thus freely displaceable within resonance cavity 142 housed atop fan outlet 122. Resonance cavity 142, in turn, is covered at the cut-off end by a perforate cover 134. By moving end flap 138 within reso¬ nance cavity 142 the volume and hence the resonance frequency of Helmholz resonator 130 can be altered so as to be tuned for maximum blade passage frequency noise reduction.

In FIGURE 14, Helmholtz resonator 230 is curved so as to substantially conform to the spiral shape of casing 212 of fan 200. Adjacent to the cut-off 220 of impeller 216 and casing 212 resonance cavity 242 is covered by perforate 234 at the mouth of resonator 230. Resonator front wall 240 abuts casing 212 while its rear wall 232, in flexible and resilient enough to be oveable at its bottom tip within resonance cavity 242. In that fashion, the volume and, therefore, the resonance frequency of Helmholtz resonator is again adjustable to variations in fan speed and blade passage frequency.

Yet another configuration of a Helmholtz resonator is shown in FIGURE 15. Helmholtz resonator 330 is again curved such that its front wall 340 abuts and conforms to the spiral casing 312. Mouth of resonator 330 is positioned adjacent the cut-off 320 of impeller 316 and casing 312 is again covered with a perforated metal sheet 334. The resonance cavity 342 positioned atop fan outlet 322 is again variable in volume since rear wall 332 is either hinged at 336 to top wall 338 or, if fixedly attached at 338, is flexible enough to permit tuning of resonator 330 to cause maximum blade passage frequency noise reduction.

CM Although the invention has been described primarily in terms of the application of quarter wavelength or Helmholtz reson¬ ators within centrifugal fans, the same apparatus and method may be used for reducing the noise level of the blade passage frequency of axial fans. In FIGURE 16, for example, a schematic cross-section through axial fan 400 shows the application of a quarter wavelength resonator 430 within guide vanes 420 inside fan housing 410 and behind axial rotor 416. A piston 432 inside guide vane 420 permits tuning of quarter wavelength resonator 430 to variable fan speeds as previously described. FIGURE 17, a front view of axial fan 400 taken along lines 17-17 of FIGURE 16, shows the perforate cover 434 at the front end of guide vane 420 which forms the mouth of quarter wavelength resonator 430. Piston 432 may again be coupled directly by any convenient servo-control means to fan speed, thus assuring its automatic tuning to the correct resonance frequency for maximum effect under variable load conditions.

Various other embodiments and modifications will also be apparent from the foregoing description. The invention is therefore not limited to the specific embodiments disclosed, but extends to every embodiment within the scope of the appended claims.

Claims

CLAIMS:

1. Low noise fluid moving apparatus having a stationary casing and a movable centrifugal impeller housed therein, wherein the improvement comprises: at least one resonator means for selectively reducing part of the blade passing tone spectrum, said resonator means being positioned at the point of smallest separation between the casing and the impeller.

2. Low noise fluid moving apparatus having, within a station¬ ary casing, one or more movable axial vane impellers with one or more stationary guide vanes behind them, wherein the improvement com¬ prises: resonator means for selectively reducing part of the blade passing tone spectrum, said resonator means being housed within the stationary guide vanes behind the impeller.

3. The apparatus of Claim 1 or 2 wherein the resonator means further include a movable plug within said resonator cavity in lieu of a fixed rear wall, thus rendering said resonator variably tunable to one-fourth the blade passing frequency.

4. The apparatus of claim 1 or 2 wherein said resonator means further include a perforated cover over the front end of said re¬ sonator.

5. The apparatus of Claim 1 or 2 wherein said resonator means are Helmholtz resonators having at least one movable wall for in¬ creasing or decreasing the volume of the resonator cavity, thus being tunable to various blade passing tone frequencies.

6. The apparatus of Claim 1 or 2 wherein said resonator means are further characterized by the fact that the interior walls of the resonators are lined with sound absorbing material.

7. The apparatus of Claim 3 wherein said resonator means further include a plurality of separate quarter wavelength resonators having movable plugs, each one being separately tunable to different absorption maxima of the blade passing tone spectrum.

8. The apparatus of Claim 5 wherein said resonator means further include a plurality of Helmholtz resonators having movable walls, each one being separately tunable to different absorption maxima of the blade passing tone spectrum.

9. A method for reducing the noise level of radial fans having a stationary casing and a movable centrifugal impeller disposed therein, comprising the step of: positioning resonator means at the point of smallest separation between casing and impeller for selectively absorbing part of the blade passing tone spectrum.

10. A method for reducing the noise level of axial fans having a stationary casing housing one or more movable axial vane impellers and one or more stationary guide vanes behind the movable rotor vanes, comprising the step of: positioning resonator means within the stationary guide vanes for selectively absorbing part of the blade passing tone spectrum.

11. The method of Claim 9 or 10 further including the step of: tuning said resonator means to various absorption maxima of the blade passing tone spectrum using movable apparatus^*for varying the volume of said resonator means.

12. The method of Claim 9 or 10 wherein a plurality of said resonator means are employed and further including the step of: tuning each of said resonator means separately to different absorption maxima of the blade passing tone spectrum using separately movable apparatus for separately varying the volume of each of said resonator means.

13. The method of Claims 9 or 10, further including the step of: covering the inlet end of said resonator means with cover plates containing a plurality of holes of various sizes and arrangements.

14. The method of Claim 9 or 10 further including the step of: lining the interior walls of said resonator means with sound absorbing material.