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This invention relates to the field of absorbing acoustic, aerodynamic and the combination of acoustic and aerodynamic pressure variations of fluctuations from
a fluid stream. More particularly for absorbing acoustic, aerodynamic and the combination of acoustic and aerodynamic pressure waves from the compressible fluid passing through a diffuser in a centrifugal compressor or other si&ilar
Centrifugal compressors are utilized by the refrigeration industry in most large installations where a
single large refrigeration. _ machine is used to provide
cooling, heating or both. Many methods have been attempted with varying degrees of success to limit the level of
loudness of the audible noise emitted by a centrifugal refrigeration machine. These methods have included encasing the motor and compressor (United States Patent No.
3,635,579); providing sound absorptive material at the in-let and outlet chambers-,of the compressor (United States
Patent N<3. 3,360,193); locating a baffle in the crossover pipe of a multi-stage compressor (United States Patent No.
3,676,012) and providing an annular muffler in the discharge line of the compressor.
Since large refrigeration installations consume high amounts of electrical energy every effort is made to! increase the efficiency of the refrigeration machine to
decrease the operating costs of the installation. The absorptive apparatus herein claimed is utilized to obtain
an overall efficiency increase in a refrigeration system having a centrifugal compressor.
The operational flow range of a centrifugal .^ REA(7"
OMPI compressor s norma y m e y e m n ma ow vo ume which can be produced without the occurrence of surge.
It is impractical to operate in surge due to pressure pulsations, dynamic and potentially dangerous thrust load variations and increased gas temperatures. When it is desirable to operate a centrifugal compressor under partial load it is necessary to operate the machine at sufficient flow volume to exceed the flow volume at surge notwithstanding that the partial load requirements could be met with a lesser flow rate. When operating at a flow rate which is higher than necessary to meet the load requirements, operating costs increase since the efficiency of the overall system is decreased. Even as surge is approached, aerodynamic instabilities arise introducing losses and lowering efficiency, so that operating costs increase as the surge line is approached. Hence, by decreasing the flow volume at which surge occurs the compressor can operate over a broader flow volume range and operate with a higher efficiency at flow ranges below the previously established surge volume.
Prior efforts to control the volume flow rate at which surge occurs have focused on the diffuser geometry and on providing vanes within, the diffuser to control the flow path of the fluid leaving the impeller. See
"Centrifugal .Compressors ... the Cause of the Curve" by

Donald C. Hallock from Air and Gas Engineering, Volume 1, Number 1, January' 1968.
It is an object of the present invention to re-duce the level of noise emitted from a centrifugal corn-compressor.
It is a further object of the present invention to increase the efficiency of a centrifugal compressor and to increase the operational range of a centrifugal com-pressor by lowering the flow volume at which surge occurs.
It is a further object of the present invention to reduce the level of the noise emitted from a moving flu stream.
It is another object of the present invention to reduce the noise level, increase the overall efficiency, and to increase the operational range and pressure rise of a centrifugal compressor without unduly impeding
fluid flow or creating severe boundary layer distortions within the fluid flow path.
It is yet another object of the present invention to provide absorbing apparatus which is adaptable to
existing centrifugal compressors with a minimum of struc- tural alterations.
It is still a further object of the present in- vention to absorb both acoustic and aerodynamic pressure variations within the fluid in communication with the ab- , sorbing apparatus.
It is also an object of the invention to prevent or delay back pressure and reverse fluid flow by means of acoustical and pressure absorbing material located in the diffusor o "a!' centrifugal compressor.
Other objects will be apparent from the description to follow and the appended claims.
The above objects are achieved according to a preferred embodiment of the invention by the provision of absorbing apparatus in communication with the fluid being compressed in the diffuser section of a compressor. A
porous absorbing material is mounted to form a portion of the wall surface of the diffuser. A resonant cavity is located on the opposite side of the absorbing material
from the fluid in such a manner that the fluid may flow through the absorbing material into the cavity. The
absorbing apparatus is annular in shape and the cavity is divided by concentric rings into a plurality of smaller cavities of by a single helical divider with periodic dams into a narrow elongated cavity or by a honeycomb or
similar divider into a multiplicity of cellular type cavities, Damping material such as fiberglass is inserted into the cavity to further aid in absorbing and damping pressure variations. The absorbing material is selected to have
" a flow resistance approximating the density of the fluid times the speed of sound in the fluid through the diffuser.

fugal compres

Figure 2 is an enlarged partial sectional view of the invention mounted to a portion of the diffuser
wall of a centrifugal compressor.
Figure 3 is a partial elevational end view taken along line 3-3 of Figure 1 of the invention in a centrifugal compressor showing the cavity divided into a narrow elongated cavity by a single helical divider and showing the location of the absorbing material.
Figure 4 is a graph of exit pressure from a centrifugal compressor versus exit volume from a centrifugal compressor shown with and without the claimed pressure variation absorber herein- and with the inlet vane angle of the compressor control vanes set at both 35 degrees and at 90 degrees.
Figure 5 is a graph of flow resistance versus absorption coefficient for air, R-ll, R-12 and R-22.
The following is a description of absorbing
apparatus mounted in communication with the fluid in a compressor to form portion of the diffuser wall of the diffuser within a centrifugal compressor and of a method of absorbing pressure waves within the fluid. I't is to be understood that the invention has like applicability to any moving fluid stream whether it be" in a centrifugal compressor, gas turbine or other dynamic head device, which converts increased dynamic head pressure created by moving blades or the like into increased static pressure. Furthermore it would be but a design expedient dependent on the design and operating characteristics of the compressor to select which wall of the diffuser or more than one wall of the diffuser upon which to mount this apparatus. If more than one wall is selected then the diffuser on each could be arranged to absorb different frequency pressure waves.

A multistage compressor could likewise utilize the present invention in one or more of the various compression stages.
It is to be further understood that the des-cription herein will refer to a refrigerant as the fluid _^-»

being compressed in a centrifugal compressor which is part, of an overall refrigeration machine. However, it is to be understood that the present invention will have like applicability to any compressible fluid be it a refrigerant, a gas or any other fluid. Since optimum porosity of the absorbing material is a function of
the gas or fluid properties, different gases or fluids will require absorbing material of varying porosity
to achieve optimum results.
Referring now to the drawings, it can be seen in

Figure 1 that in a typical centrifugal cαrpr-essor refrigerant enters the oorrpressor throug refrigerant inlet 42 then travels' along a refrigerant flow path through the control vanes 44 and into impeller chamber 46. Impeller 16 mounted on shaft 48 and driven by motor 15 then accelerates the refrigerant and discharges the refrigerant into diffuser 14. At the point of discharge, 15, from impeller blades 17, the refrigerant is traveling at a relatively high velocity and is at a relatively low static pressure. The refriger-ant then travels through diffuser 14 to collector 12 from which it is discharged into the remainder of the refrigeration The refrigerant leaving the diffusor and entering the collector is traveling at a relatively slow velocity as compared to when it entered the diffuser and is at a relatively high static pressure as compared to the pressure when it entered the diffuser. A pressure variation absorber 10, comprising absorbing material 20 and resonant cavity 13, is in communication with the refrigerant passing through diffuser 14. Volute casting 19 is shown in Figure 1 as structurally connecting the collector, the diffuser and the impeller chamber.
Referring now to Figure 2 which is a partial enlarged sectional view of the diffuser and the pressure variation absorber, it can be seen that absorbing material 20, a porous high flow resistance sheet of material, is mounted to form a portion of the surface of the diffuser wall. The absorbing material could likewise be mounted on the other diffuser wali; or on both walls. The absorbing material is mounted by means of screws 32 and by an adhesive (not shown) to a portion of volute casting 19.

On the opposite side of the absorbing material from the fluid is resonant cavity 18 defined by end dividers 23 and a backplate 26. As can be seen from Figure 3 pressur variation absorber 10 is annular in shape and each end divider forms a complete ring so that' resonant cavity 18 formed by the two end dividers, backplate 26 and the absorbing material 20 is annular in configuration althoug other configurations would be equally acceptable. It can be further seen in Figure 2 that the annular resonant cavity is divided into a series of smaller cavities by dividers 22. Dividers 22 may be a single helix with periodic solid flow barriers 33 as shown in- Figure 3 or ' comprise a series of concentric rings. A honeycomb or cellular type divider would also be satisfactory. No matter what the divider configuration a narrow cavity or series of cavities is provided. The pressure variation absorber is shown mounted within volute cavity 21 formed by various portions of volute 19. This particluar arrang merrt is structural and has no effect on the claimed inven If a narrow plurality of cavities were not provided the refrigerant flowing through the diffuser havin a relatively low static pressure at the end of the absorb closest to the impeller and a relatively high static pressure at the end of the absorber closest to the collec would enter the pressure variation absorber closest to the co tor and flew backwards towards the end of the pressure variation absorber closest to the impeller. This backward flo of refrigerant would then detract from the overall effici of the unit. By providing a single helical resonant cavi with periodic flow barriers back flow due to the pressure gradient is .sufficiently small relative to the high flow resistance of the absorbing material that overall machine efficiency is not substantially affected. Back flow can similarily be limited by concentric dividers or a multi-plicity of cellular type cavities so that the incremental pressure drop in each cavity is minimal.
It can be further seen that end dividers 23 and dividers 22 are sealed to prevent fluid flow between the separate cavities. Dividers 22 and end dividers 23 are mounted to absorbing material 20 and to backplate means of an epoxy type resin. The resonant cavity 18
of the pressure variation absorber is further filled with a damping material such as fiberglass to increase the
absorbing efficiency of the unit and to provide damping
of possible resonating pressure waves within cavity 18.
Figure 4 is an experimentally developed graph
of head pressure versus flow volume for a centrifugal
compressor equipped with and without the herein described invention. The graph shows both operation of the- com-pressor with the pressure variation absorber and without the pressure variation absorber. The dotted line, when
the machine ' is operated without the pressure variation
absorber, shows that surge occurs at a much higher flow
volume than with the present apparatus. Furthermore,
the graph shows the respective characteristics with the
control vanes set at a 35 degree angle and with the
control vanes set at a 90 degree angle. It can be seen
from the graph that the operational range between the
point when surge occurs and when head pressure is reduced below an operational value is greatly increased, especially at the lower flow volumes. In addition, the pressure rise of the compressor is increased particularly for
control vane settings below 90 degrees. The method of
utilizing this apparatus includes locating the pressure
variation absorber within the diffuser so that pressure
variations in the fluid within the diffuser are absorbed.
These variations include acoustical and aerodynamic waves generated by the impeller as the fluid is accelerated and those waves occurring as a result of surge and other aero-dynamic instabilities such as rotational stall as the fluid is pressurized and decelerated in the diffuser.
The precise mechanism which operates to improve the efficiency of the machine and to reduce the flow
volume at which surge occurs is not fully known. It has been discovered that an absorber designed to absorb
acoustic waves (which are pressure variations) and thereby reduce noise emitted by the machine also acts to absorb aerodynamic pressure variations which result from
surge and other aerodynamic instabilities affecting the overall efficiency of the machine. It is theorized tha^^TjRE^
r OMPI an absorber acts to restrict pressure variations resultin from either acoustic waves or aerodynamic instabilities. The efficiency improvement results from the elimination or reduction in severity of the aerodynamic instabilities A smooth flow without pressure variations not only result in a machine having a lower flow rate at which surge may occur and thereby having a greater operational range but also adds to the overall fficiency of the unit since the impeller is not forced to overcome these aerodynamic pressure fluctuations that the absorber is removing them from the system.
Random and periodic aerodyn-amic pressure vari- ' ations of unknown origin have also been detected within a centrifugal compressor. It is experimentally determine that the disclosed absorber also attenuates these variations further adding to the efficiency of the overall machine.
The absorbing material 20, such as "Feltmetal" or "Fibermetal" manufacturated by Brunswick Corporation of Muskegon, Michigan or "Rigimesh" manufactured by Aircraft Porous Media, Glen Cove, New York, is selected so that its flow resistance approximates the density of the fluid times the- speed of sound in the fluid across the absorbing material. Hence, the absorbing material is varied according to the fluid being used or more particularly according to the particular
refrigerant selected for the particular application. The table below shows various refrigerants, the various densities of the refrigerant leaving the impeller the various velocities of the speed of sound in the refrigerant and the consequent optimum flow resistance th absorbing material should have for each application.
(A conversion factor of 0.48823 is used to convert from

English to Metric units.)
D Deennssiittyy Speed of Sound Flow
Refrigerant Resistance.
Kσ/m3 M/Sec . Rayls Tegs)

Figure 5 is a graph of the maximum normal absorption coefficient versus flow resistance for Air, R-ll, R-12, and R-22 as measured in an acoustic impedance tube. This graph is a plot of values which shows that an absorption coefficient of approximately 1.0 is obtainable by selecting the proper flow resistance for the absorbing material. The graph confirms that material having the values set forth in the table is the optimum choice to absorb pressure variations for the particular refrigerant.
The resonant cavity backing the absorbing material is designed so that its depth is- one quarter the wave length of the wave length of the lowest frequency of sound that it is desired to absorb. For
example, if R-ll (trichloro-fluoromethane) is the
refrigerant being used in the machine and the pressure variation absorber is designed to eliminate acoustical noise at 300 Hertz and above, then, the cavity depth should be 12.7 centimeters; the velocity of the speed of --Sound of R-ll divided by four times the frequency.
Damping material is selected for the resonant cavity so that all frequencies greater than the frequency for which the cavity is designed will be absorbed or attenuated. The damping material helps to absorb the frequencies between the resonance peaks of the design frequency thereby providing an absorber which will absorb all frequencies from the minimum frequency increasing to the highest audible frequencies and beyond.
It can be seen from the above described
embodiment that there has been provided an acoustic and aerodynamic pressure variation absorber which has the capability of not only absorbing acoustic waves and thereby reducing the noise level emitted by a machine (e.g. a centrifugal compressor) and/or the fluid passing therethrough but also to absorb aerodynamic pressure variations so that the efficiency of the machine is increased and the overall operational range of the
machine is broadened.