EP0039459A1 - Silenced turbo-machine - Google Patents

Abstract

A method and apparatus for reducing unwanted pressure impulses in fluid moving devices comprising at least one resonator means (30) for selectively reducing part of the device's blade passing tone, said resonator means being positioned at the point of smallest separation between the casing (12) and the impeller (14).

Description

The invention relates to the construction and operation of turbo-working machines, for example pumps, blowers, compressors, turbines and the like, in which fluid is accelerated by a moving surface immersed in it and acting on it. In particular, the invention relates to turbo machines in which the generation of undesired pressure pulses is reduced or eliminated. Such pressure pulses arise when the moving surface or surfaces move close to fixed surfaces on or within the device. The undesired pressure pulses considerably increase the power requirement of the device for transporting the fluid. By arranging a suitably tuned acoustic resonator on or near the zone of the device in which the interaction of the moving and the fixed surface occurs, the undesired pressure pulses in this zone can be destroyed, whereby the Efficiency of the device is increased.

Excess energy based on the generation of unwanted pressure pulses is generally known as noise in turbo machines. The main component of the noise spectrum in axial or radial flow machines contains aerodynamically or hydrodynamically generated noise. In contrast, mechanical noise sources, such as bearings, drives or transmissions, play only a subordinate role. The aerodynamically or hydrodynamically generated sound can be split into two parts, namely the blade tone frequency and the noise. The blade tone frequency is caused by the interaction between the stationary machine housing and the fluid leaving the impeller at high speed. The main source of such interactions resulting from fluctuations in the _F luidgeschwindigkeit at the rotating _L aufradschaufeln nearest housing tongue. In radial turbo machines with bladed diffusers, the stationary guide vanes behind the impeller are the main acoustic source for the buzzer tone frequency. In this application, the fixed guide vanes are referred to as housing tongues. The spectrum of the bucket tone frequency is discrete. It has high amplitude peaks at the blade frequency, which is the product of the speed and number of blades. Additional acoustic energy peaks occur at harmonics of the blade frequency.

The random noise component of the audible spectrum arises from turbulent flow in the vicinity of the impeller and the stationary housing. Another contribution comes from the vortex separation at the rear edges of the impeller blades.

A reduction in the amplitude of the blade tone frequency is particularly important in turbomachinery because conventional noise reduction devices involve a considerable loss of energy and because sound components containing sound are much more unpleasant for the human ear than noise with a wide frequency spectrum. Many attempts have been made to reduce the noise output from centrifugal fans. The basic physical principles of these experiments are for the different types of radial turbo machines, e.g. Blowers, pumps, compressors, turbines and the like. essentially the same. It is e.g. known to increase the distance between the impeller and the housing tongues, thereby reducing the sound intensity. However, the most commonly used method requires an expensive enlarged fan housing and is therefore limited to applications where there is no space constraint.

It is also known to increase the radius of curvature of the housing tongue in order to reduce the share of the blade tone frequency of the noise spectrum. However, the reduction in blade frequency noise level resulting from this increase in radius is limited to about 3-5 dB for most impeller speeds.

It is also known to change the angle of inclination between the impeller blades and the housing tongue in order to obtain a sound intensity reduction of the order of 0-16 dB depending on the housing tongue spacing. However, inclined blades are difficult to produce economically and an inclined housing tongue requires a complicated geometry Fan outlet line, which results in flow losses. For this reason, previous attempts have proven to be not very practical, both in terms of design and mode of operation.

In the case of small fans, in which the two halves of a double inlet fan are manufactured separately and subsequently assembled, a staggering of the blades of double inlet or double row impellers is sometimes used to reduce the sound intensity. For large welded together blower units this type caused by the asymmetrical construction Allerding _s thermal stresses and deformations in the back plate of the impeller.

The introduction of transition grids on the front and rear edges of the impeller not only reduces the intensity of the sound, but also the part of the spectrum based on random fluctuations. However, there is a pressure drop due to the presence of the grilles, which results in a significant reduction in the blower output.

The noise spectrum assigned to the blade frequency can be broadened by irregularly arranging the blades around the hub. Although this is desirable from a subjective assessment of the noise, this design does not reduce the overall acoustic energy generated by the fan.

Attaching a _K eilriemens around the impeller reduces the sound intensity in a manner similar to the oblique design of the housing tongue. However, the triangular belt surrounding the impeller causes friction losses and lowers the blower output.

When a conventional volute casing replaced by a rectangular housing and the _G ehäusezunge is completely omitted, this results in a reduction in the sound intensity. Due to the insufficient flow guidance, this construction deteriorates the aerodynamic efficiency of the fan.

The application of an acoustic lining to the interior of the blower housing effectively reduces both the intensity of the sound and random noise. However, this approach has the same disadvantage as conventional absorbent damping agents. After a certain period of operation, the acoustic material is often torn away by the flow. Finally, if media loaded with liquid or solids are conveyed, the pro-surface of the lining is added.

There is also a mismatch between the acoustic impedance of the blower and duct system e.g. by adjusting the length of the fan outlet line for the purpose of significantly reducing the blade tone frequency emitted into the inlet line. However, this approach is not readily applicable to an existing fan because the influence of any modifications, e.g. of the blower inlet, also depends on the acoustic impedance of the existing pipe system at the blower outlet.

Finally, US Pat. No. 2,160,666 (McMahon) describes an "air cushion" behind the housing tongue as a means of reducing noise. McMahon does not, however, disclose that one maximum noise reduction can be achieved by tuning the cavity to the blade tone frequency and / or by suitable choice of the perforated cover of the resonator mouth.

All of the above-mentioned measures are based on the fact that it is known that the generation of the sound takes place in the interior of a concentrated area around the housing tongue of the blower housing. This phenomenon makes changing the sound generation mechanism directly at the source a promising option. The sound is caused by the interaction of the air flow leaving the impeller with the asymmetrical part of the fan housing which has the housing tongue. It is known that this sound generation is extremely sensitive to changes in the geometry of the housing tongue or the impeller.

The invention avoids the need for the change of the geometry of the fan and thus avoids ^mechanical limitations, difficulties, and efficiency decreases. The principle of the invention is to reduce the sound intensity by attaching an a / 4 or Helmholtz resonator in such a way that the mouth of the resonator forms the housing tongue of the centrifugal fan or compressor. If the resonator is tuned to the blade frequency or its harmonics, the pressures generated by the flow in the body tongue zone at that frequency are significantly reduced. Not much more space is required to add the resonator. The geometry of the housing tongue and the fan housing remain the same.

The object of the invention is to reduce the level of the bucket tone frequency in turbo-working machines without adversely affecting the characteristic curve, the efficiency or the size of the devices. In conjunction with, and in part as a result of, a lower level of the bucket tone frequency, an overall reduction in noise level is achieved, particularly with respect to those tones that are particularly disturbing to the human ear. A blower system to which the invention is applied has a higher overall performance than a similar blower system with conventional dampers or the like.

These objects are achieved in that the mouth of an acoustic resonator is arranged at or near the area of the device at which the interaction between moving and stationary surface parts occurs in order to generate the undesired pressure pulses. By suitable tuning of the resonator, very undesirable pressure pulses in this zone can be destroyed by the pressure pulses that arise at the mouth of the resonator. In terms of construction, the main source regions of the bucket sound are formed in the housing as resonators. These resonators are tuned to the blade frequency of the device or its harmonics. As a result, the pressure fluctuations generated by the current leaving the impeller or the fan blades can be destroyed or reduced directly at their source will. The main sources for the blade sound are the housing tongues in the case of bladeless diffusers and the guide blades in devices with bladed diffusers. In the context of this application, these two sources are referred to as "housing tongues" because they both protrude into the interior of the machine housing.

The basic idea of noise reduction according to the invention is applicable to fans, blowers, pumps, turbines and compressors in radial or axial construction. Experiments with centrifugal fans have shown that by attaching one or more resonators to or in the housing tongue, the noise level at the blade frequency or its harmonics can be reduced very effectively. Experimental reductions of up to 28 dB in the hearing range at the blade frequency were achieved.

One of the main advantages of the invention is that it does not require any intervention in the flow geometry and the optimal flow pattern. The geometry of the housing interior does not have to be changed, so that the mechanical characteristic and the efficiency of the device are not adversely affected by the presence of the resonators on the housing tongue. Only the internal geometry of the resonator itself has to be changeable if tunable resonators are desired.

The tuning to different frequencies is possible, for example, by changing the cavity volume by means of at least one movable side or rear wall. The resonator can be designed as a Helmholtz resonator, the resonance frequency of which depends on the volume of the cavity, or it can be an A / 4 resonator, the resonance frequency of which is significantly influenced by the depth of the cavity.

Since the basic frequency of the blade sound is directly proportional to the speed of the impeller, in one embodiment of the invention a control device responsive to the speed of the impeller can be provided, which changes the volume or the depth of the cavity as a direct function of the speed of the impeller. This ensures that the resonator is always "tuned" to the particular frequency that forms the main part of the noise spectrum of the machine.

In another embodiment of the invention, a plurality of resonators are arranged next to one another, each resonator being tuned to a different frequency. For example, one resonator can be tuned to the frequency of the fundamental oscillation and the other resonators can be tuned to harmonics of the fundamental oscillation frequency.

• Fig. 1 einen schematischen Querschnitt eines Zentrifugalgebläses mit einem an der Gehäusezunge angeordneten λ/4-Resonator,
• Fig. 2 eine vergrößerte Ansicht der Gehäusezunge mit einem λ/4-Resonator zur Veranschaulichung von Einzelheiten einer Ausführungsform der Resonatormündunq,
• Fig. 3 eine graphische Darstellung der Geräuschminderung, die mit dem Resonator erzielbar ist, der abgestimmt wurde, um eine maximale Verringerung der Schaufelfrequenz bie 7500 Upm, d.h. 750 Hz zu vermitteln,
• Fig. 4 eine ähnliche graphische Darstellung, wobei der Resonator abgestimmt ist, um eine maximale Verringerung der Schaufelfrequenz bei 6000 Upm, das heißt 600 Hz hervorzubringen,
• Fig. 5 ein Diagramm des Niveaus der Schaufelfrequenz als Funktion der Gebläsedrehzahl,
• Fig. 6 die Wirkung des auf 600 Hz abgestimmten λ/4-Resonators auf das gesamte Geräuschdruckniveau, das A-gewichtete Geräuschdruckniveau und die zweite _H.armonische der Schaufelfrequenz,
• Fig. 7 bis 12 eine Reihe von Versuchsdaten ähnlich denen gemäß Figur 5, jedoch mit dem Unterschied, daß die Resonatormündung mit einer perforierten Abdeckung versehen ist. Figur 7 zeigt zum Beispiel das Niveau der Schaufelfrequenz in bezug auf die Gebläsedrehzahl für verschiedene Resonatorzeiteinstellungen, wobei die Resonatormündung mit einer 20 % perforierten Abdeckung versehen ist, die mehrere Löcher mit 2,6 mm Durchmesser aufweist. Für das Diagramm nach Figur 8 wurde eine Resonatormündung mit 42 % Öffnungsbereich verwendet, der aus Löchern von 3,8 mm Durchmesser besteht, während gemäß Figur 9 die perforierte Resonatorabdeckung 58 % Öffnungsbereich mit einzelnen Lochgrößen von 4,4 mm Durchmesser aufweist. In jedem Falle war die Gesamtanzahl der Löcher in dem für die Figuren 7 bis 9 verwendeten Perforierungskörper konstant,
• Figuren 10 und 11 die Wirkung von Änderungen der Rate des öffnungsbereiches in dem Perforierungskörper bei Aufrechterhaltung eines konstanten Lochdurchmessers von 3,8 mm. Gemäß Figur 10 waren 5 von 43 Löchern geschlossen, während Figur 11 die Schaufeltonverminderung für den gleichen Perforierungskörper mit 16 geschlossenen von 43 Löchern zeigt,
• Figur 12 die Schaufeltonverminderung als Funktion der Gebläsedrehzahl für zwei Perforierungskörper der gleichen Rate des Öffnungsbereiches, jedoch mit verschiedenen Durchmessern und Anzahlen von Löchern,
• Fig. 13,14,15 Querschnittsansichten von Radialgebläsen mit verschiedenen Ausführungsformen eines abstimmbaren Resonators an der Gehäusezunge,
• Fig. 16 einen schematischen Querschnitt eines Axialgebläses mit einem abstimmbaren A/4-Resonator in einer Leitschaufel hinter den Gebläseschaufeln, und
• Fig. 17 eine Ansicht längs der Schnittlinien A-A in Figur 16.

It shows:

• 1 shows a schematic cross section of a centrifugal fan with a λ / 4 resonator arranged on the housing tongue,
• 2 is an enlarged view of the housing tongue with a λ / 4 resonator to illustrate details of an embodiment of the resonator mouth,
• 3 is a graphical representation of the noise reduction achievable with the resonator that has been tuned to provide a maximum reduction in blade frequency at 7500 rpm, ie 750 Hz.
• 4 is a similar graphical illustration with the resonator tuned to produce a maximum reduction in blade frequency at 6000 rpm, i.e. 600 Hz.
• 5 shows a diagram of the level of the blade frequency as a function of the fan speed,
• Fig. 6 shows the effect of the tuned to 600 Hz λ / 4-resonator on the entire sound pressure level, the A-weighted sound pressure level and the second _H .armonische the blade frequency,
• 7 to 12 show a series of test data similar to that according to FIG. 5, but with the difference that the resonator mouth is provided with a perforated cover. For example, FIG. 7 shows the level of the blade frequency in relation to the fan speed for various resonator time settings, the resonator mouth being provided with a 20% perforated cover which has a plurality of holes with a diameter of 2.6 mm. For the diagram according to FIG. 8, a resonator mouth with 42% opening area was used, which consists of holes with a diameter of 3.8 mm, while according to FIG. 9 the perforated resonator cover has 58% opening area with individual hole sizes with a diameter of 4.4 mm. In any case, the total number of holes in the perforation body used for FIGS. 7 to 9 was constant,
• FIGS. 10 and 11 show the effect of changes in the rate of the opening area in the perforation body while maintaining a constant hole diameter of 3.8 mm. According to FIG. 10, 5 of 43 holes were closed, while FIG. 11 shows the bucket tone reduction for the same perforation body with 16 closed of 43 holes,
• FIG. 12 shows the reduction in the bucket tone as a function of the fan speed for two perforation bodies with the same rate of the opening area, but with different diameters and numbers of holes,
• 13, 14, 15 cross-sectional views of radial fans with different embodiments of a tunable resonator on the housing tongue,
• 16 shows a schematic cross section of an axial blower with a tunable A / 4 resonator in a guide blade behind the blower blades, and
• 17 shows a view along the section lines AA in FIG. 16.

FIG. 1 shows a centrifugal fan 10, consisting of a fan housing 12 and a paddle wheel 14 accommodated therein. In the preferred embodiment according to the invention, which was used to carry out the experiments graphically illustrated in FIGS. 3 to 12, the paddle wheel 14 has six backward curved blades 17 provided and has a diameter of 140 mm. The blower housing 12 is designed as a logarithmic spiral with a very small distance between the housing tongue between the impeller 14 and the housing 12 in the housing tongue region 18. The housing tongue spacing 18 is only 4.4 mm, and the housing tongue radius 20 of the model blower 10 used for the tests is only 10 mm.

According to FIG. 1, an X / 4 resonator 30 is arranged at point 20 of the housing tongue. A piston 32, which can be displaced in a piston-like manner within the resonator 30, is used to tune the resonator for different frequencies.

The noise reduction effect of various designs of the resonator 30 was measured by placing a microphone in the fan outlet line 22. The microphone in the line was provided with a slotted tube to reduce the influence of turbulent pressure fluctuations on the microphone. The microphone signals were analyzed using a narrow-band filter, which could be tuned to the blade frequency or multiple harmonics using a synchronizing device. A real-time analyzer was used to record the frequency spectrum of the noise pressure in the line.

FIG. 2 shows an enlarged cross section of the housing tongue with the resonator 30, the mouth 34 on the housing tongue being shown as a perforated cover at the resonator input. The perforated cover 34 consists of 1 mm thick sheet metal, which was exchanged in the course of the tests for different configurations with different arrangement, size and number of openings 36. The cross-sectional area of the resonator 30 is rectangular and its largest side dimension corresponds to the width of the blower housing. The upper resonator wall 38 with a width of 46.1 mm is 16.5 mm above a lower resonator wall 41 of the same width.

To tune the resonator 30, the length of the resonator was changed via a movable Teflon plug 32, which formed an airtight seal with the resonator walls. Optimal tuning, i.e. maximum noise reduction at a given frequency, was obtained simply by observing the amplitude of the sound on the real time analyzer. All noise measurements were made at only one blower operating point located on the right hand side of the printhead volume flow characteristic, i.e. almost free outlet. In order to obtain reference data for comparison purposes, the resonator 30 was replaced in some experiments by a solid piece of wood with the same radius of curvature 20 and the same distance from the impeller 14.

FIGS. 3 and 4 show comparisons between the noise pressure spectrum measured in the fan outlet line for the fan with a conventional housing tongue design and with an integrated X / 4 resonator. The perforated cover plate 34 consists of 1 mm thick sheet metal, the holes 36 have a diameter of 3.1 mm and make up an overall opening area of 29%. For the spectrum shown in FIG Blade frequency at 7500 rpm, that is 750 Hz and corresponding to 6000 .Upm in Fig. 4 causes. In both Figures 3 and 4, the solid arrows illustrate the noise level of the paddle harmonics without a resonator, that is, the resonator 30 has been replaced by a solid piece of wood. The dashed arrows show the noise levels of the paddle harmonics, which were measured with the X / 4 resonator 30.

Correct tuning of the resonator not only reduces the blade frequency, but also its higher harmonics. However, it can also result in gain, as shown in Figure 4, where the third harmonic has been increased by 8 dB.

The reduction in the blade frequency is not the same in FIGS. 4 and 5, and the frequency spectrum measured at other fan speeds showed that the maximum possible reduction that can be achieved by carefully tuning the resonator for each case increases with the frequency.

This effect becomes clearer in FIG. 5, in which the level of the blade frequency is plotted as a function of the fan speed. The solid curve A applies to the blower with the usual housing tongue shape. The other lines in FIG. 5 were measured with the X / 4 resonator matched to the fan frequencies, that is to say 450, 500, 550, 600, 650, 700 or 750 Hz. The maximum possible reduction increases with the frequency in FIG. 5 and further tests have shown that this can be changed by using a different perforation body for the resonator mouth. The same perforation body as for FIGS. 3 and 4 was used for the data shown in FIG. For a given resonance frequency, the damping effect is not only present at that particular frequency or within a very narrow frequency band, but extends over a fairly wide frequency range. This is important because of the practical application of the noise control method, since it shows that a significant reduction in the bucket sound can be achieved even when the resonator is not properly tuned, be it due to an initial mismatch, or a change in fan speed due to fan load, or a change in fluid temperature that corresponds to a change in the speed of noise in the fluid being moved.

FIG. 6 shows the effect of the a / 4 resonator, which was tuned to 600 Hz over the entire noise pressure level "A", the A-weighted noise pressure level B and the second harmonic of the blade frequency C. The effectiveness of the X / 4 resonator with respect to the total or A-weighted noise level generally depends on the importance of the bucket sound. In the model blower used for experimental purposes, the overall noise level was reduced by 3.5 dB and the A-weighted level by 7 dB.

Another interesting result of FIG. 6 is that the second harmonic was mostly reduced at 3000 rpm, which corresponds to 600 Hz. However, the reduction is smaller than that of the blade frequency at the same resonance frequency, that is 6000 rpm (compare FIGS. 4 or 5). This shows that the impedance of the λ / 4 resonator changes as a result of the flow velocity along the housing tongue.

FIGS. 7 to 12 show a series of test data similar to that according to FIG. 5, the only difference being the perforation body 34 used to cover the resonator mouth. 7-9, the perforation rate was changed within 20% to 58% by increasing the diameter of the holes 36, that is, the total number of holes is the same in FIGS. 5 and 7 to 11. There are indications that the increase in the rate of the opening range shifts the range of maximum reduction of the blade frequency towards a lower frequency range.

A physical explanation of the damping phenomenon during the frequency shift can be developed in the following way. If the diameter of the holes. 36 is large, the oscillation of the air through the holes is little damped, so the reduction in the tone amplitude is large. It should be noted, however, that this reduction only occurs over a narrower bandwidth than that shown in Figure 9. Conversely, if the holes 36 are small, the damping rate of the resonator 30 increases, resulting in a lower reduction in the tone amplitude, but over a wider frequency range than shown in FIG. 7.

In FIGS. 10 and 11 an attempt was made to change the rate of the opening area and thereby the same diameter of the holes 36 as in FIG means keeping 3.8 mm. For this purpose, some holes 36 were sealed with putty. Figure 10 appears to indicate the same trend as before, that is the decrease in the rate of the opening area shifts the effective area to higher frequencies. However, a further reduction in the number of holes as in FIG. 11 does not support this observation.

Finally, in FIG. 12 two perforation bodies 34 are compared, the opening area of which is the same, but in which the diameter and the approach. Number of holes 36 differ from each other. This leads to the assumption that the rate of the opening area is the size that determines the frequency range of the maximum noise reduction and not the hole diameter.

One of the main advantages of the noise reduction method according to the invention is that the introduction of the resonator 30 requires only minor changes in the fan geometry. Therefore, as expected, the performance and flow rates for a fan 10 with or without a resonator 30 are very similar, if not practically the same. In order to empirically determine the difference in the aerodynamic performance of these two designs, the volume flow pumped through the test line 22 was measured. Table I lists the ratio of the volume flows with and without the X / 4 resonator 30 against the fan speed and the results show that the volume flow decreases by an average of 0.6%.

The above test results show that a substantial reduction in the blade frequency emitted by centrifugal fans can be achieved by installing a λ / 4 resonator on the housing tongue.

With an unchanged resonator shape, that is to say with a constant length and mouth geometry of the resonator 30, maximum attenuation of the blade frequency occurs at a frequency at which the noise wavelength is approximately four times the resonator length is. A relation for the effective length of the resonator is difficult to express due to the curved design and the perforations forming the mouthpiece 34 with construction parameters. In addition, the overflow holes 36 also act on the mass addition to the effective length. A somewhat lower damping of the blade frequency results within a frequency band in the region of the resonance frequency of the resonator. The width of this frequency band is determined by the damping rate of the resonator, which in turn depends on the size and number of holes 36 in the mouthpiece 34.

If the length of the resonator 30 is changed while maintaining the same resonator mouth shape 34, the resonance frequency changes, and therefore a reduction in the blade frequency can be achieved at other fan speeds. However, the maximum possible reduction changes with the frequency, which shows that there is an optimal impedance of the resonator 30 as a whole, which causes maximum attenuation. The overall impedance of resonator 30 appears to be very sensitive to changes in flow velocity at resonator mouth 34, and this helps explain why the amount of noise reduction achieved by properly tuning the resonator for a given fan speed will vary with frequency, or rather the fan speed changes. Trials with different cover plates 34 at the resonator mouth have also shown that this optimal impedance can be generated at different frequencies by simply changing the perforation rate at the resonator mouth 34.

Experiments with the described preferred embodiments of the invention showed that an X / 4 resonator on the housing tongue of a centrifugal fan is a highly effective and extremely simple means of reducing the blade tone. This method can be used to reduce noise in both new fan designs and existing systems. Since the use of an X / 4 resonator does not require a change in the geometry of the housing spiral, the aerodynamic efficiency of the fan is not impaired.

For variable speed fans, a λ / 4 resonator can still be used if the length of the resonator is changed in proportion to the fan speed by displacing a piston body in the resonator, and in a somewhat modified version of the preferred embodiment of the invention, therefore Piston body 32 is coupled to the impeller speed via any known suitable servomechanism.

Three different versions of a further preferred embodiment of the invention are shown schematically in FIGS. 13 to 15. In FIG. 13, for example, a centrifugal compressor 100, in the housing 112 of which an impeller 116 is accommodated, is equipped with a Helmholtz resonator 130 instead of an X / 4 resonator. The Helmholtz resonator 130 consists of a front wall 140, which runs essentially tangentially to the housing 112, but can follow the spiral shape of the housing in the vicinity of the housing tongue 120. Its rear wall consists of a main section 132 which is essentially parallel to the front wall 140 and whose lower part ends behind the housing tongue 120 at an articulation point 136. The end part 138 of the rear wall 132 is articulated under the articulation point 136 and is therefore freely adjustable within the resonance cavity 142 above the blower outlet 122. The resonance cavity 142 is in turn covered at the housing tongue end by means of a perforated cover 134. By moving the end flap 138 in the resonance cavity 142, the volume and thus the resonance frequency of the Helmholtz resonator 130 can be changed in order to be tuned for maximum blade frequency noise reduction.

In FIG. 14, the Helmholtz resonator 230 is curved so that it essentially corresponds to the spiral shape of the housing 212 of the blower 200. In addition to the housing tongue 220 of the impeller 216 and the housing 212, a resonance cavity 242 covered by a perforated body 234 at the mouth of the resonator 230. The front wall 240 of the resonator abuts the housing 212, while its rear wall 232 is flexible and resilient enough to be movable in the resonance cavity 242 at its lower edge. In this way, the volume and thus the resonance frequency of the Helmholtz resonator can in turn be adjusted to changes in the fan speed and the blade frequency.

Another design of a Helmholtz resonator is shown in FIG. 15. The Helmholtz resonator 330 is again curved so that its front wall 340 abuts and corresponds to the spiral housing 312. The mouth of the resonator 330 is in the vicinity of the housing tongue 320 of the impeller 316 and the housing 312 and is covered with a perforated metal sheet 334. The volume of the resonant cavity 342 above the fan outlet 322 is again variable, since the rear wall 332 is either articulated to the top wall 338 at 336 or, if rigidly attached at 338, is flexible enough to tune the resonator to achieve maximum blade frequency noise reduction 330 to allow.

Although the invention has been described primarily with reference to the use of an A / 4 or Helmholtz resonator within centrifugal fans, the same device and method can be used be used to reduce the noise level of the blade frequency of axial fans. For example, FIG. 16 shows a schematic cross section through an axial fan 400, an a / 4 resonator 430 being arranged in guide vanes 420 within the fan housing 410 and behind the axial rotor 416. A piston 432 within the vane. 420 allows the λ / 4 resonator 430 to be tuned to variable fan speeds as previously described.

Figure 17, which illustrates a front view of axial fan 400 along lines 17-17 of Figure 16, shows perforated cover 434 at the front end of vane 420 which forms the mouth of X / 4 resonator 430. The piston 432 can again be coupled directly to the fan speed via any servo control means, which ensures its automatic tuning to the correct resonance frequency for maximum efficiency under variable load conditions.

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

Claims (14)

1. Low-noise turbo machine, consisting of a stationary housing and a rotating centrifugal impeller accommodated therein, characterized by at least one resonator means (30) for selectively reducing a part of the blade tone spectrum, the resonator means (30) at the point (18) of the smallest separation is arranged between the housing (12) and the impeller (14). 2. Turbo machine which has in a stationary housing at least one rotating axial fan and at least one stationary guide blade behind it, characterized by the resonator means (30) for the selective _V ER- reduction of a portion at the blade Elton spectrum, wherein the resonator means (30) inside the stationary vanes is arranged behind the axial paddle wheel. 3. Turbo working machine according to claim 1 or 2, characterized in that the resonator means (30) instead of a fixed rear wall has a movable piston body (32) in the resonator cavity, whereby the resonator (30) is variably tunable to a quarter of the blade frequency. 4. Turbo working machine according to claim 1 or 2, characterized in that the resonator means (30) is provided over its front end with a perforated cover (34). 5. Turbo machine according to claim 1 or 2, characterized in that serve as resonator Helmholtz resonators (130) with at least one movable wall (138) for increasing or decreasing the volume of the resonator cavity (142) so that it can be tuned to different paddle frequencies . 6. Turbo working machine according to claim 1 or 2, characterized in that the inner walls of the resonators (30) are lined with sound-absorbing material. 7. Turbo working machine according to claim 3, characterized in that the resonator means has a plurality of individual X / 4 resonators (30) with movable piston bodies (32), each of which can be tuned for different absorption maxima of the blade tone spectrum. 8. Turbo working machine according to claim 5, characterized in that the resonator means has a plurality of Helmholtz resonators (130) with movable walls (138), each of which has different absorption maxima of the bucket tone spectrum is tunable. 9. A method for reducing the noise level of radial fans with a stationary housing and a rotating centrifugal impeller accommodated therein, characterized in that a resonator means is attached to the selective absorption of a part of the blade tone spectrum at the point of smallest separation between the housing and the impeller. 10. A method for reducing the noise level of axial fans with a stationary housing, at least one rotating axial impeller and at least one stationary guide vane behind the movable rotor blades, characterized in that a resonator means is attached to the selective absorption of a part of the vane tone spectrum within the stationary vane. 11. The method according to claim 9 or 10, characterized in that the resonator means are tuned to different absorption maxima of the blade tone spectrum, using a movable device for changing the volume of the resonator means. 12. The method according to claim 9 or 10, characterized in that when using a plurality of resonator means, each resonator means is tuned for different absorption maxima of the blade tone spectrum in order to change the volume of each resonator means for itself. 13. The method according to claim 9 or 10, characterized in that covering the inlet end of the resonator means with cover plates which have a plurality of holes of different sizes and arrangements. 14. The method according to claim 9 or 10, characterized in that the inner walls of the resonator are lined with sound-absorbing material.

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