Abstract:
A volute for a radial turbomachine is proposed. The turbomachine includes a radial compressor or a radial turbine. The volute is in particular for a radial compressor. The volute has a substantially annular cavity which is delimited at least by a first radial side surface. At least one substantially annularly circumferential groove is formed in the side surface.

Description:
CROSS REFERENCE TO RELATED APPLICATIONS 
     This application is the US National Stage of International Application No. PCT/EP2012/052160 filed Feb. 9, 2012 and claims benefit thereof, the entire content of which is hereby incorporated herein by reference. The International Application claims priority to the German application No. 10 2011 005 025.6 DE filed Mar. 3, 2011, the entire contents of which is hereby incorporated herein by reference. 
     FIELD OF INVENTION 
     The invention relates to a diffuser for a radial compressor, which diffuser has a substantially annular hollow chamber which is delimited at least by a first radial side surface. 
     BACKGROUND OF INVENTION 
     Radial compressors are known for example from EP 1 356 168 B1 or from EP 1 602 810 A1. 
     Such radial compressors are composed of a rotor or vane impeller which forms a compressor stage and which rotates about an axis of rotation and which has—with respect to the axis of rotation of the rotor—an axial inlet and a radial outlet. Gas to be compressed flows axially into the rotor of the compressor stage and is then diverted outward (radially, radial direction), wherein said gas exits the rotor at high speed. 
     Kinetic energy of the gas to be compressed which exits at high speed is then converted in a diffuser into potential energy in the form of pressure. 
     Such a diffuser is usually formed by two non-rotating rings which form an annular hollow chamber or an annular chamber, which annular chamber radially adjoins the rotor outlet, or which rings or annular walls/side surfaces adjoin the rotor outlet radially and are perpendicular to the axis of rotation or are at a highly obtuse angle with respect to said axis of rotation (radial annular chamber walls/radial side surfaces). 
     The gas exiting the rotor is conducted radially outward in said annular chamber between said two annular walls and passes to a collector. 
     Diffusers commonly have vanes, that is to say a blade arrangement, for diversion and improved control of the slowing of the flow. 
     It is also known that such radial compressors exhibit relatively high sound emissions or noise levels which constitute a (noise) disturbance in the surroundings of the radial compressor. Said sound emissions may furthermore also cause vibrations and structure-relevant malfunctions. 
     Dominant sound sources in a radial compressor are for example typically generated at the location of the vane impeller and of the diffuser inlet or any diffuser blades, owing to the high speed of the fluids flowing through said regions and owing to an interaction of rotor and stator components. 
     In particular, it is known here that radial compressors generate complex, transient, three-dimensional, rotating and/or pulsating pressure fields or sound fields at an outlet from the radial compressor (pressure side), for example at a pressure connection piece at said outlet, the sound waves of which pressure fields or sound fields propagate without disruption into the pipelines adjoining the pressure connection piece. 
     Here, in addition to the said noise disturbances, vibrations and structure-relevant malfunctions, pipeline vibrations may also occur which may lead to damage to the pipelines to the point of failure of the radial compressor or of the superordinate system that has the radial compressor. 
     The damping of such complex, transient, three-dimensional, rotating and/or pulsating pressure fields or sound fields is technically difficult. 
     Taking this as a starting point, efficient sound damping measures are necessary for sound-emission-generating radial compressors of said type. 
     Various “external” measures for limiting sound emissions, such as housings or casings, are known. Said noise reduction techniques may be relatively expensive, in particular if they are marketed as an “aftermarket” add-on product. 
     Furthermore, “internal” silencers for limiting sound emissions in the case of radial compressors are known. 
     Silencers in general are devices for preventing sound emissions. It is possible to distinguish between various types of silencers which reduce a generated acoustic power on the basis of different mechanisms. A distinction is made, for example, between adsorption and reflection/resonator silencers. 
     An absorption silencer such as is known for example from EP 1 602 810 A1 for a radial compressor comprises porous (adsorption) material, generally mineral wool, glass wool or glass fiber, which partially absorbs acoustic energy, that is to say converts said acoustic energy into heat. By means of absorption, primarily upper frequencies of the sound medium are damped in the silencer. 
     Similar fillings of a corresponding hollow chamber are also proposed in DE 603 10 663 T2, DE 601 20769 T2 and US 2009/229280 A1. DE 601 14 484 T2 discloses a circumferential groove whose depth is enlarged to more than 1.5 times the axial width of the cut-out portion of the compressor wheel. 
     Absorption silencers have the disadvantage that they are generally unsuitable for high pressures, because—owing to the high pressures—large amounts of energy act on the absorption material, or large amounts of heat must be absorbed by the absorption material, which can lead to damage to the porous material, such as for example a disintegration of the absorption material. 
     Resonator silencers or reflection silencers, which utilize the principle of acoustic reflection, generally comprise for this purpose multiple cavities or chambers which are passed by the sound medium, wherein reflections occur. As the sound medium passes repeatedly by interior spaces of the chambers, a reduction of acoustic pressure peaks of various frequencies occurs. Said reflections are—in structural terms—generated by impact walls, cross-sectional widenings and narrowings. By means of reflection, it is possible for any desired frequencies of the sound medium to be damped in the silencer. 
     Such a resonator silencer, based on a Helmholtz resonator principle, for a radial compressor is known from EP 1 356 168 B1 or from EP 1 443 217 A2. In said radial compressor, the diffuser therein has an acoustic lining in the form of an area with numerous bores which act as Helmholtz resonators. 
     In addition to a radial compressor of said type, a further known form of radial turbomachine is a radial turbine. 
     Such a radial turbine, such as is known for example from DE 44 38 611 C1, is based on a reversal of the physical principle of a radial compressor, and is accordingly—with corresponding components to those in a radial compressor—traversed by flow in a flow direction opposite to that in said radial compressor. 
     In radial turbines, too, the described emission problems arise in a corresponding manner. 
     For example, dominant sound sources in a radial turbine are typically generated at the location of the vane impeller or of a turbine wheel (both also referred to hereinafter for short as “rotor”) and of a turbine guide ring, or any guide ring blades, positioned upstream of the turbine wheel. 
     Here, too, it is possible for complex, transient, three-dimensional, rotating and/or pulsating pressure fields or sound fields to be generated at a suction side, that is to say at an inlet into the radial turbine, for example at a suction connection piece at said inlet, the sound waves of which pressure fields or sound fields propagate without disruption into the pipelines connected upstream to the suction connection piece. 
     Taking this as a starting point, efficient sound damping measures are necessary here, too, for sound-emission-generating radial turbines of said type. 
     SUMMARY OF INVENTION 
     The invention is based on the object of specifying a silencer which improves the disadvantages from the prior art, is simple to implement and is also simple to install—in a sound-emitting system or device such as a radial turbomachine—and which is suitable in particular for damping sound emissions in the case of a radial compressor or a radial turbine. 
     The object is achieved by means of a diffuser for a radial turbomachine, in particular a radial compressor, having the features of the independent patent claim. 
     Said diffuser has a substantially annular hollow chamber, an annular chamber, which is delimited at least by one first radial side surface. According to the invention, at least one substantially annularly encircling groove is formed in said side surface. 
     Here, said at least one substantially annularly encircling groove, which is open toward the annular chamber via a groove outlet (groove outlet opening), acts as an acoustic resonator, in particular a lambda/4 resonator—hereinafter also referred to for short as merely “resonator”—such that sound waves which pass the groove and which have the same frequency as an (acoustic) natural frequency or resonance frequency of said groove are reflected in a region of a groove outlet, and thus a sound propagation across the groove or across the resonator is reduced. 
     In this way, the sound propagation in the annular chamber can be reduced, and effective sound damping in the diffuser—and in the radial turbomachine or the radial compressor—can be attained. 
     By means of a selected geometry or dimensioning of the substantially annularly encircling groove, in particular by means of a depth of the groove, by means of a width/height of the groove or of the groove outlet, by means of a radial position of the groove in the radial side surface, the eigenform (eigenmode) or nodal diameter and natural frequency or resonance frequency of the groove is defined. 
     The configuration or the (three-dimensional) geometry of the circumferential groove are—in themselves—not subject to any limits as long as the encircling groove forms a cavity or a hollow chamber which acts as an acoustic resonator. 
     It is thus possible, for example, to realize circumferential grooves with any desired groove shapes, such as circumferential grooves with rectangular, V-shaped or trapezoidal cross section, circumferential grooves with an outwardly sloping wall and/or circumferential grooves in dovetail form and/or circumferential grooves with smooth and/or curved walls in regions or throughout and/or circumferential grooves with undercuts and/or with chambers. Undulating circumferential grooves or circumferential grooves with a stepped groove base are also possible. 
     Thus, if a sound wave passing the groove has the same eigenform as or an identical nodal diameter to an identical acoustic eigenform in the resonator or the groove, and/or if the sound wave passing the groove has the same natural frequency as the groove, the reflection is particularly effective. 
     That is to say, through suitable (three-dimensional) dimensioning of the groove, the acoustic natural frequency and the eigenform of the groove can be tuned to a sound wave to be reflected, that is to say to the frequency and eigenform thereof, and it is thus possible by means of the dimensioning of the groove for targeted frequencies to be damped. 
     In other words, owing to the three-dimensional nature of the substantially annularly encircling groove (also referred to as merely “circumferential groove”), the eigenforms thereof can be set by means of simple geometric parameters such that acoustic pressure patterns of a particular form passing the groove can be reflected in a particularly effective manner. 
     The form of said passing acoustic pressure patterns may for example be estimated using analytical relationships, for example using a formula according to Tyler &amp; Sofrin. 
     The geometry of a circumferential groove is simple to manufacture and, owing to the low number of free parameters, such as height, width, depth or shape, offers the possibility of inclusion in an optimization process. 
     Furthermore, the invention realizes a robust, maintenance-free (sound damping) solution which is not subject to wear even at high pressures and temperatures. The invention thus offers a considerable advantage in relation to approaches based on absorption material. 
     Since the “silencer” according to the invention is used close to the sound source (rotor and possibly bladed diffuser/possibly bladed guide ring), it is possible, with correct dimensioning, to also reduce the excitation of the rotor by acoustic pressure patterns. 
     With the use of the circumferential groove in the annular chamber, no further sound insulation measures are required, in particular in the pipeline system. Both a radiation of noise and also the excitation of pipeline vibrations can be considerably reduced. A considerable cost advantage is attained in relation to external silencer solutions. 
     Pressure losses to be expected are low, as shown both by numerical calculation and also by experiments. 
     Preferred refinements of the invention will emerge from the dependent claims. 
     In a preferred embodiment, the at least one first radial side surface has multiple substantially annularly encircling grooves which are in particular situated concentrically with respect to one another. The efficiency of the silencer can be increased by means of a plurality of such circumferential grooves. 
     Said circumferential grooves may particularly preferably be designed so as to have different dimensions in each case, in particular different depths and/or widths. For example, it may be provided here that the depth and the width of the circumferential grooves become smaller in each case with increasing radial distance to the outside in the annular hollow chamber or annular chamber. 
     In this way, that is to say by means of a plurality of circumferential grooves, it is possible in a targeted manner for multiple frequencies to be damped, to the point of wide-band sound damping of sound emissions in the radial turbomachine. For example, it is possible to realize a frequency band to be damped of 700 Hertz-2000 Hertz, 700 Hertz-4000 Hertz or 700 Hertz-6000 Hertz. 
     The efficiency of the “resonator silencer” may be further increased if the annular hollow chamber is delimited by a second radial side surface which is situated axially opposite the first radial side surface, which second radial side surface likewise has a substantially annularly encircling groove or—with a further increase in efficiency—multiple substantially annularly encircling grooves, which are in particular situated concentrically with respect to one another. 
     On this basis, it may be provided in a further preferred refinement that the one substantially annularly encircling groove of the first radial side surface is situated axially directly opposite, that is to say at the same radial height as, the one substantially annularly encircling groove of the second radial side surface. 
     Alternatively, however, it may also be provided that the one substantially annularly encircling groove of the first radial side surface is situated opposite the one substantially annularly encircling groove of the second radial side surface with a radial offset, that is to say at a different radial height. This may be advantageous in particular if, owing to elements, for example a blade arrangement, arranged in the annular hollow chamber or annular chamber, there is no space available for a “directly axially opposite arrangement” of the circumferential grooves. 
     Such a directly axially opposite arrangement, and also a radially offset arrangement of circumferential grooves, may also be provided in the case of in each case multiple substantially annularly encircling grooves, situated concentrically with respect to one another, in the two radial side surfaces. Here, too, the space conditions in the annular chamber (bladed annular chamber) may be decisive for the provision of radially offset circumferential grooves instead of a “directly axially opposite arrangement”. 
     In a further preferred refinement, the natural frequency of the at least one substantially annularly encircling groove is tuned to a frequency to be reflected. The frequency to be reflected may particularly preferably be a vane impeller rotational frequency (“blade passing frequency”) of a radial compressor or a second harmonic or third harmonic or fourth harmonic of the vane impeller rotational frequency of the radial compressor. The eigenform of the at least one substantially annularly encircling groove should preferably also be tuned to the eigenform of a sound wave to be reflected. 
     In a further preferred embodiment, the substantially annular hollow chamber has a blade arrangement. 
     This may have the result that the at least one substantially annularly encircling groove or multiple such circumferential grooves is or are arranged in a region of the blade arrangement in the annular chamber. 
     It may also be provided that the at least one substantially annularly encircling groove or multiple such circumferential grooves is or are arranged outside the region of the blade arrangement in the annular chamber. 
     It may also be provided that the at least one substantially annularly encircling groove has discontinuities. This may be provided for example if the annular chamber has a blade arrangement which prevents a fully encircling groove. 
     In a further preferred refinement, it is provided that the “(resonator) silencer” is used or realized in a radial compressor, as a diffuser therein. The “silencer” may also be used or realized in a radial turbine at a turbine guide ring positioned upstream of a turbine rotor of the radial turbine. 
     It may also be provided that—in the case of multiple substantially annularly encircling grooves—said grooves are formed such that a damping action is configured for a large rotational speed range of for example 50% to 105% of a nominal rotational speed of the radial turbomachine or of the radial compressor. 
    
    
     
       BRIEF DESCRIPTION OF DRAWINGS 
       Exemplary embodiments of the invention are illustrated in figures, which will be explained in more detail below. In the figures: 
         FIG. 1  is a sketched sectional illustration of a radial turbomachine, a radial compressor, having a resonator silencer as per one exemplary embodiment; 
         FIG. 2  is a sketched sectional illustration of a radial turbomachine, a radial compressor, having a resonator silencer as per a further embodiment; 
         FIG. 3  is a sketched sectional illustration of a radial turbomachine, a radial compressor, having a resonator silencer as per a further embodiment; 
         FIG. 4  shows, by way of example, an acoustic eigenmode in an annular groove in a radial compressor as per one embodiment. 
     
    
    
     DETAILED DESCRIPTION OF INVENTION 
     Exemplary embodiments: resonator silencer for a radial compressor 
       FIGS. 1 to 3  illustrate various configurations of radial compressors  100  each with a resonator silencer  1  realized or integrated in the diffuser. 
     Such radial compressors  100  have, as illustrated, a rotor  10  which rotates at high rotational speed about an axis  11 . The rotor  10  has a hub  12  and blades  13  that project radially from said hub. 
     The hub  12  has a first region  12   a  which is substantially cylindrical, a transition region  12   b  in which the hub radius widens, and an end region  12   c  which runs substantially perpendicular to the axis  11 . 
     The gas  2  that flows in axially in the flow direction  3  is set in rotation by the rotor  10  and exits the rotor  10  in the radial flow direction  3  with respect to the axis  11  and at an obtuse angle with respect to the axis  11 . 
     The blades  13  are fastened to a common backplate  14  of the hub  12 . The rotor  10  is situated in a housing  15 , the wall  16  of which is adapted to the outer contour of the rotor. The blower formed by the rotor  10  has an axial inlet  17  and a radial outlet  18  which extends over the circumference of the rotor  10 . 
     The outlet  18  is adjoined by the diffuser  20  which is fixedly connected to the housing  15  and which does not rotate. The diffuser  20  has a substantially radial supporting wall  21  to which there are attached vanes  22  (diffuser blade arrangement) which guide the flow passing the outlet  18 . 
     A further substantially radial wall  23  is situated axially opposite the radial supporting wall of the diffuser  20  at a distance therefrom, whereby the diffuser  20  forms an annular chamber, the annular chamber  30 , which is occupied by the blade arrangement  22 . 
     The vanes  22  run substantially radially with respect to the axis  11 . Between the vanes  22  there are formed diffuser ducts whose cross-sectional area increases from the inside to the outside. 
     It is the task of the diffuser  20  to slow the gas accelerated by the rotor  10 , which gas has high kinetic energy, and to convert the kinetic energy into pressure. 
     An outlet  26  of the diffuser  20  is adjoined—further downstream—by a pipeline system  29  (not illustrated in any more detail) (pressure side  27 ), which pipeline system is connected to the diffuser  20  via a pressure connection piece  28 . 
     Radial compressors  100  such as that illustrated generate high sound emissions which constitute a (noise) disturbance in the surroundings of the radial compressor  100  and which can cause vibrations, structure-relevant malfunctions and also pipeline vibrations in/on pipeline systems, which pipeline vibrations lead to damage to the pipelines, to the point of failure of the radial compressor  100 . 
     Dominant sound sources of such emissions are generated at the location of the vane impeller/rotor  10  and of the diffuser inlet  25  or any diffuser blades  22 , owing to the high speed of the fluids flowing through said regions. 
     In particular, complex, transient, three-dimensional, rotating and/or pulsating pressure fields or sound fields are generated at the pressure side  27  or at the pressure connection piece  28 , located there, of the radial compressor  100 , the sound waves of which pressure fields or sound fields can propagate without disruption into the pipelines  29  adjoining the pressure connection piece  28  and can cause the described damage there. 
     To prevent such damage, or as an effective sound insulation means, the radial compressors  100 —as shown in FIGS.  1  to  3 —provide in each case a resonator silencer  1  realized or integrated in the diffuser or in the annular chamber  30  there. 
     To prevent the propagation of the sound waves in the annular chamber  30  of the diffuser  20 , it is the case, as shown in  FIGS. 1 to 3 , that one or more circumferential/annular grooves  50  that extend annularly around the axis  11  are formed in the radial supporting wall  21  and/or in the radial wall  23 , which circumferential/annular grooves act as acoustic resonators, in particular as lambda/ 4  resonators. 
     Here, said circumferential grooves  50 —which run annularly and concentrically with respect to the axis  11 —may be formed in the annular chamber  30  on one side, for example on the radial supporting wall  21  or on the radial wall  23 , or else on both sides, that is to say both on the radial supporting wall  21  and also on the radial wall  23 . 
     Said circumferential grooves  50  may also be arranged either only in the region of the blade arrangement  22  of the diffuser or only in the region outside the blade arrangement  22  of the diffuser  20 , or else both in and outside the region of the blade arrangement  22  of the diffuser  20 . 
     Sound waves passing through the annular chamber  30  or passing by the circumferential/annular grooves  50 , which sound waves have the same frequency as one of the resonance frequencies of such a circumferential/annular groove  50 , are reflected and thus damped in the region of the resonator outlet  51 , that is to say of the groove opening or of the groove inlet  51 . 
       FIG. 1  shows an embodiment of said resonator silencer  1  which has two circumferential grooves  50  which run in each case annularly around the axis  11  concentrically with respect thereto. 
     One of the two circumferential grooves  50  is arranged on the radial supporting wall  21 . The second of the two circumferential grooves is arranged, at approximately the same radial distance from the axis  11 , in the radial wall  23 . The two circumferential grooves  50 , which are identical in terms of shape, width and depth and which have a U-shaped cross section, are accordingly situated axially directly opposite one another, that is to say at the same radial height. 
     The radial spacing of said circumferential grooves from the axis  11 , or the radial position thereof in the annular chamber  30 , is such that both circumferential grooves  50  are situated (radially) outside the bladed region  22  of the diffuser  20  or annular chamber  30 . 
       FIG. 2  shows a further embodiment of a resonator silencer  1  in the diffuser  20 , which resonator silencer has a multiplicity of annularly encircling circumferential grooves  50  which are in each case concentric with respect to the axis  11 . 
     A first proportion of said circumferential grooves  50 , in this case four circumferential grooves  50 , is arranged in the radial supporting wall  21  in the region of the blade arrangement  22  of the diffuser  20 . Directly axially opposite said circumferential grooves  50 , that is to say in each case at the same radial height or with the same radial spacing to the axis  11 , a second proportion of the circumferential grooves  50 , likewise four circumferential grooves  50 , is arranged on the radial wall  23 —and thus likewise in the bladed region  22  of the diffuser  20  or annular chamber  30 . 
     Mutually directly opposite circumferential grooves  50  are in this case in each case identical in terms of shape, width and depth. Here, the width and the depth of the circumferential grooves  50  decreases with increasing spacing from the axis  11 . In other words, with increasing radial spacing to the axis  11 , the circumferential grooves  50  become slimmer or narrower and shallower. All of the circumferential grooves  50  have a U-shaped cross section. 
       FIG. 3  shows a further embodiment of a resonator silencer  1  in the diffuser  20 , likewise with a multiplicity of annularly encircling circumferential grooves  50  which are in each case concentric with respect to the axis  11 . 
     In said embodiment as per  FIG. 3 , all of the circumferential grooves  50 , in this case four circumferential grooves  50 , are arranged, concentrically with respect to one another and concentrically with respect to the axis  11 , on the radial wall  23  in the region of the blade arrangement  22  of the diffuser  20 . With increasing radial spacing from the axis  11 , the width and the depth of the circumferential grooves  50  decrease. In other words, with increasing radial spacing from the axis  11 , the circumferential grooves  50  become slimmer or narrower and shallower. Here, too, all of the circumferential grooves  50  have a U-shaped cross section. 
       FIG. 4  shows, by way of example, an acoustic eigenmode  60  in an annular groove  50  of said type which acts as a resonator. 
       FIG. 4  shows  24  pressure maxima  61 . Said eigenmode or acoustic mode  60  is also characterized by 12 so-called nodal diameters  62  and a particular natural frequency. Sound waves which pass by the circumferential groove  50  and which are characterized by said natural frequency are reflected, and the sound propagation across or past the circumferential groove  50  is reduced. 
     If the sound wave passing by has the same nodal diameters  62  as the acoustic eigenform  60  in the circumferential groove  50  (resonator), the reflection process is particularly effective. 
     Resonator silencers  1  such as that described act with extremely high efficiency, in particular because they are used close to the sound source, the rotor  10  and (possibly bladed  22 ) diffuser  20 , such that further complex sound insulation measures, in particular for the entire pipeline system  29  of the radial compressor  100 , can be dispensed with.