Reciprocating piston compressor having improved noise attenuation

Information

  • Patent Grant
  • 6776589
  • Patent Number
    6,776,589
  • Date Filed
    Friday, January 24, 2003
    21 years ago
  • Date Issued
    Tuesday, August 17, 2004
    20 years ago
Abstract
A reciprocating piston compressor having a suction muffler and a pair of discharge mufflers to attenuate noise created by the primary pumping frequency in the primary pumping pulse. The suction muffler is disposed along a suction tube extending between the motor cap and the cylinder head of the compressor. The discharge mufflers are positioned in series within the compressor to receive discharge gases from the compression mechanism and are spaced one quarter of a wavelength from each other so as to sequentially diminish the problematic or noisy frequencies created during compressor operation. The motor/compressor assembly including the motor and compression mechanism is mounted to the interior surface of the compressor housing by spring mounts. These mounted are secured to the housing to define the position of the nodes and anti-nodes of the frequency created in the housing to reduce noise produced by natural frequencies during compressor operation.
Description




BACKGROUND OF THE INVENTION




The present invention relates to reciprocating piston fluid compression devices such as hermetic refrigerant compressors, particularly with regard to quieting same.




Fluid compression devices such as, for example, refrigerant compressors, receive a gas at a suction pressure and compress it to a relatively higher, discharge pressure. Depending on the type of compression device, the work exerted on the gas in compressing it is characterized by a series of intermittently exerted forces on the gas, the magnitude of these forces normally varying from zero to some maximum value. For example, in a cylinder of a reciprocating piston type compressor, this force ranges from zero at the piston's bottom dead center (BDC) position, to a maximum at or near the piston's top dead center (TDC) position, at which the pressure of the compressed gas is respectively at a minimum pressure (i.e., substantially suction pressure) and a maximum pressure (i.e., substantially discharge pressure). Some quantity of the gas is discharged from the cylinder as the piston assumes new positions as it advances from BDC to TDC, and thus the compressed gas flowing from the cylinder is not at a uniform pressure. Rather, the gas which flows from the cylinder, which is generally referred to as being at discharge pressure, actually has many different pressures.




Pulses of higher discharge pressure result in the compressed gas flowing from the cylinder, these pulses being in the portion of the flowing gas which leaves the cylinder as the piston approaches or reaches TDC. As the piston cycles in its cylinder, regular, equally distributed patterns of these pulses are created in the compressed gas which flows through a conduit, tube or line leading from the compression mechanism. The pulsating flow of compressed gas through this discharge line may be represented by sine waves of various frequencies and having amplitudes which may vary with changes in the quality of the refrigerant; these changes are effected by changes in refrigerant type, temperature or pressure. Pulsations at certain frequencies may be more noticeable, and thus more objectionable, than others.




Further, the nominal discharge pressure, i.e., the pressure at which the compressed gas is generally considered to be, will also vary with refrigerant quality. The frequency of these high pressure pulses in the compressed gas flowing through the discharge line, however, has a substantially constant frequency which directly correlates to the speed at which the gas is compressed in the cylinder, and the number of cylinders in operation. This frequency is referred to as the primary pumping frequency, and is generally the lowest frequency exhibited by the pressure pulsations in the compressed gas.




The amplitude of the pressure pulses at the primary pumping frequency tend to be the largest in the compressed gas flow. Because the primary pumping pulses are at low frequencies and large amplitudes, they are often the primary cause of objectionable noise or vibration characteristics in compressors or the refrigeration systems into which these compressors are incorporated. These systems normally also include at least two heat exchangers, a refrigerant expansion device, and associated refrigerant lines which link these components into a closed loop relationship. Pressure pulsations at other, higher frequencies have amplitudes which are relatively smaller, but certain of these pressure pulsations may also be objectionable. Further, some objectionable pressure pulsations may establish themselves in the conduits or lines which convey refrigerant substantially at suction pressure to the compression mechanism.




Substantial effort has been expended in attempting to quiet these pressure pulses in addressing noise or vibration concerns, and it is known to provide mufflers in the discharge or suction lines to help resolve these issues. These mufflers may be of the expansion chamber type, in which a first refrigerant line portion opens directly into a chamber, wherein the amplitude and/or frequency of at least one of the pulses may be altered, and from which the refrigerant exits through a second line portion. Further, it is known that the discharge chamber in the head of a reciprocating piston compressor can also serve as a type of expansion chamber muffler. An expansion chamber type muffler of any type is not entirely satisfactory, however, for it may cause a substantial pressure drop in the gas as it flows therethrough, resulting in compressor inefficiency. Further, such mufflers may not provide sufficient attenuation required by the application.




An alternative to an expansion chamber type of muffler is what is well known in the art as a Helmholtz resonator type of muffler wherein the wall of a portion of the discharge pressure line may be provided with a plurality of holes, that portion of the discharge line is sealably connected to a shell which defines a resonance chamber, the holes in the discharge line providing fluid communication between the interior of the discharge line and the resonance chamber. The size and/or quantity and/or axial spacing of these holes, and the volume of the resonance chamber, are variably sized to tune a Helmholtz resonator to a particular frequency, and the amplitude of pulses at that frequency are thereby attenuated. Compared to an expansion chamber type of muffler, a Helmholtz muffler provides the advantage of not causing so significant a pressure drop in the fluid flowing therethrough; thus compressor efficiency is not compromised to the same degree.




Although a Helmholtz resonator may be effective for attenuating the amplitude of fluid pulses having shorter wavelengths, in which case the resonator extends axially over at least a substantial portion of the pulse wavelength, prior Helmholtz resonator arrangements may not be effective for attenuating the amplitude of fluid pulses having longer wave lengths. As mentioned above, the primary pumping frequency tends to be rather low, the primary pumping pulses cyclically distributed over a rather long wavelength. By way of the example of a single-speed hermetic reciprocating piston type compressor, the motor thereof rotates at a speed which is directly correlated to the frequency of the alternating current (AC) electrical power which drives it. In the United States, AC power is provided at 60 cycles/second. The electrical current is directed through the windings of the motor stator, and electromagnetically imparts rotation to the rotor disposed inside the stator. The crankshaft of the compressor is rotatably fixed to the rotor and drives the reciprocating piston, which compresses the refrigerant. Thus the primary pumping frequency is at or near 60 cycles per second. The speed of sound in refrigerant gas at the discharge temperature and pressure of this example is 7200 inches per second. Thus, in accordance with the equation:






c/f=λ  (1)






where speed “c” is 7200 inches per second and frequency “f” is 60 cycles per second, for the above example wavelength “λ” of the primary pumping pulse is 120 inches. Notably, should the compressor be of the two cylinder variety, twice as many primary pumping pulses will be issued per revolution of the crankshaft; thus λ will then be 60 inches. It can be readily understood by those of ordinary skill in the art that simply providing a single Helmholtz resonator in the discharge line may be largely ineffective for attenuating the amplitude of a pulse which has such a long wavelength, for the point(s) of maximum pulse amplitude, which ought to be coincident with the resonator, may be too far separated. In order for a single Helmholtz resonator to quiet a pulse having such a long wavelength, the resonator would be far too long to facilitate easy packaging within the refrigerant system, let alone within the hermetic compressor housing.




What is needed is a noise attenuation system for a compression device which effectively addresses the noise and vibration issues associated with pressure pulses of relatively long wavelength, such as primary pumping pressure pulses, and which overcomes the above-mentioned limitations of previous muffler arrangements.




Typically, reciprocating piston compressors include a cylinder block having at least one cylinder bore in which is disposed a reciprocating piston. The piston is operatively coupled, normally through a connecting rod, to the eccentric portion of a rotating crankshaft. Rotation of the crankshaft, which may be operatively coupled to the rotor of an electric motor, induces reciprocation of the piston within the cylinder bore.




Covering an end of the cylinder bore, in abutting contact with the cylinder block directly or through a thin gasket member disposed therebetween, and in facing relation to the piston face, is a valve plate provided with suction and discharge ports which are both in fluid communication with the cylinder bore. Each of the suction and discharge ports are provided with a check valve through which gases are respectively drawn into and expelled from the cylinder bore by the reciprocating piston as the piston respectively retreats from or advances toward the valve plate.




The suction and discharge check valves are normally located adjacent and abut opposite planar sides of the valve plate and may, for example, be of a reed or leaf type which elastically deform under the influence of the gas pressure which acts thereon as the gas enters or leaves the cylinder bore through suction and discharge ports provided in the valve plate, and which are covered by the respective valves. The cylinder head is disposed on the side of the valve plate opposite that which faces the cylinder block, and in prior art compressors the head is in abutting contact with the valve plate, directly or perhaps through a thin gasket member disposed therebetween. Alternatively, the valve plate-interfacing surface of the head may be provided with a machined groove in which a seal is disposed, the seal compressed as the head is abutted to the interfacing valve plate surface.




The cylinder head is normally a die cast aluminum or cast iron component which at least partially defines separate suction and discharge chambers therein. Suction pressure gas is introduced into the head suction chamber through an inlet to the head; and the suction pressure gas is drawn by the retreating piston from the head suction chamber through the suction port of the valve plate, past the suction check valve, and into the cylinder bore, where the gas is compressed to substantially discharge pressure. The discharge check valve prevents gas in the discharge chamber from being drawn into the cylinder bore through the discharge port of the valve plate.




Discharge pressure gas in the cylinder bore is expelled through the discharge port of the valve plate, past the discharge check valve, and into the discharge chamber of the head, from which it is expelled through the outlet of the head. The suction check valve prevents gas in the cylinder bore from being expelled into the suction chamber of the head through the suction port of the valve plate. As noted above, the discharge chamber defined by the head of a reciprocating piston compressor may serve as a type of expansion chamber muffler. Enlarging the volume of this chamber by including such a spacer generally improves the head's ability to perform as an expansion type discharge muffler and better attenuate noise associated with pulses carried by the compressed gas.




Moreover, a problem experienced with some reciprocating compressors, particularly those in which the discharge gas is conveyed directly from the head discharge chamber through interconnected conduits to a heat exchanger, is that discharge pressure gas within the head discharge chamber does not readily exit the head, resulting in a pressure buildup in the head discharge chamber during compressor operation. Consequently, the cylinder bore may not be fully exhausted of discharge pressure gas at the end of the compression cycle because the buildup of gas within the head discharge chamber inhibits the accommodation therein of gas being exhausted thereinto from the cylinder. Because gas from the previous compression cycle has not been fully exhausted from the cylinder bore, less suction pressure gas can be drawn into the cylinder during the next compression cycle. Thus, the efficiency of the compressor is compromised. Moreover, the temperature of gas on the discharge side of the system, both within the head itself and the high side of the system, may become excessively high as more and more work is expended on the gas already at discharge pressure.




The previously preferred solution to this problem has been to enlarge the size of the head discharge chamber, thereby allowing gas which is exhausted from the cylinder bore to be more easily compressed into, and accommodated by, the head discharge chamber. As noted above, enlargement of this chamber usually also facilitates improvements in noise quality. One approach to enlarging the head's discharge chamber has been to retool the head. This solution carries with it attendant tooling costs which may not be insubstantial. Further, where a common head design is shared between different compressor models, a newly designed head which solves the problem for one model may not meet the needs (e.g., packaging requirements) of the other model(s), thereby requiring a plurality of head designs to be released and maintained in inventory.




Another approach to enlarging the head's discharge chamber is to provide a spacer between the valve plate and the existing head, which effectively enlarges the volume of the head discharge chamber (and the suction chamber as well). The spacer comprises a separate component which may be used in one compressor but not another, the two compressor models sharing a common head design. These spacers may be made of plastic or metal.




Previous plastic spacers have had coefficients of thermal expansion which differ substantially from those of the cylinder block and/or the head, and consequently may either shrink and thereby cause a leak across its sealing surfaces, or expand and be overly compressed between the valve plate and head, thereby placing considerable additional stress on the spacer, the head and bolts which extend through the spacer and attach the head and spacer to the cylinder block. If so stressed, the spacer may crack and consequently leak. Plastic spacers do, however, provide the benefits of being lightweight, and providing insulation against thermal conduction between the head and the cylinder block, thereby keeping the discharge gas somewhat cooler and thus reducing the capacity required of the heat exchanger which condenses the high pressure gas to a high pressure liquid. Plastic spacers are also made inexpensively by injection molding techniques.




Previous metal spacers, on the other hand, undesirably promote thermal conduction between the head and the cylinder block, weigh more, and usually are die cast and machined, resulting in a relatively more expensive part vis-a-vis a plastic spacer. A metal spacer, however, may have a coefficient of thermal expansion which avoids the above mentioned shrinkage and stress concerns attendant with plastic spacers. Further, prior plastic and metal spacers alike may require additional, separate gaskets to seal the opposite open spacer ends to the valve plate and head in order to provide a proper seal.




What is needed is an inexpensively produced head spacer for increasing the volume of the discharge chamber of the cylinder head, which provides seals between the head spacer and the valve plate, and between the head spacer and the cylinder head, without the need for additional seals.




Further, it is known to dispose an end cap over the end of the annular motor stator in a low-side hermetic compressor, the end cap covering both the stator end and the end of the motor rotor disposed inside the stator. It is also known to drawn suction pressure refrigerant gas from within the end cap through a suction tube extending therefrom which is in fluid communication with the inlet to a compression mechanism driven by the motor and disposed at the opposite end of the motor stator. Such a configuration is shown, for example, in U.S. Pat. Nos. 5,129,793 (Blass et al.) and 5,341,654 (Hewette et al.), and exemplified by the Model AV reciprocating compressors manufactured by the Tecumseh Products Company of Tecumseh, Mich. It is also known to provide suction mufflers in this tube intermediate the stator end cap and the compression mechanism, as taught by Blass et al. '793 and Hewette et al. '654.




A problem with such suction tube arrangements is that their lengths are fixed and particular to stators of a given height. A unique suction tube design must be provided for each different stator height in compressor assemblies which might otherwise be similar, resulting in part complexities and associated inventorying costs and efforts, and additional jigs and fixtures to produce different suction tube assemblies to accommodate these various stators. It would be desirably to provide a single suction tube assembly, with or without a muffler provided therein, which extends between the stator end cap and the inlet to the compression mechanism and can accommodate stators of different heights. Further, it may also be desirable to fix the distance of the muffler from the inlet to the compression mechanism to aid in properly tuning or packaging the muffler, while still accommodating these different stators.




Further still, it is known to resiliently support the motor/compressor assembly, which includes the motor and compression mechanism, within the hermetic shell or housing on a plurality of mounts affixed to the interior of the housing. Typically, these mounts are equally distributed about the interior circumference of the housing or otherwise placed thereabout in a manner which is merely convenient to attachment of the mounts to the motor/compressor assembly.




It is further understood by those of ordinary skill in the art that the housing has natural resonant frequencies that may produce loud, pure, undesirable tones when the housing is vibrated at or near those frequencies. Typically, equally distributing the mounts about the inner circumference of the housing may, at the points of contact therebetween, establish nodes which coincide with at least one of these natural frequencies. Similarly, placement of the mounts merely to facilitate convenient mounting of the motor/compressor assembly may also place these points of contact at nodes of natural frequencies which produce loud tones. Thus, previous compressors do not beneficially place the motor/compressor mounts on the housing in a manner which addresses the noise associated with excitation of these natural frequencies. To do so would reduce or eliminate the housing's natural resonant frequencies, and reduce the noise produced thereby.




SUMMARY OF THE INVENTION




One aspect of such a noise attenuation system for a compression device relates to an improved discharge pulse reduction system which comprises at least one muffler located in a discharge fluid line, the muffler spaced along the discharge line at a distance from a compressor discharge chamber or another upstream muffler which is a particular fraction or multiple of the wavelength of the primary pumping frequency. Thus, the amplitude of the primary pumping frequency, which may be reduced in the above-mentioned compressor discharge chamber or upstream muffler, is further reduced by the muffler placed at the above-mentioned distance therefrom, at which the already reduced amplitude reaches its new maximum value. Thus, the amplitude of the pulse at the primary pumping frequency is twice attenuated, improving the noise and vibration characteristics of the compressor and/or the refrigerant system into which it is incorporated. The muffler(s) may be of the Helmholtz or expansion chamber type.




Accordingly, the present invention provides a compressor assembly including a compression mechanism into which a gas is received substantially at a suction pressure and from which the gas is discharged substantially at a discharge pressure, the gas discharged from the compression mechanism carrying pressure pulses having a particular frequency and wavelength, these pressure pulses being of variable amplitude. A first muffler is provided through which the gas discharged from the compression mechanism flows, and a second muffler is provided in series communication with the first muffler and through which the gas having flowed through the first muffler flows. The first and second mufflers are spaced by a distance which is substantially equal to an odd multiple of one quarter of the wavelength, the amplitude being reduced in response to the gas having flowed through the second muffler.




The present invention also provides a compressor assembly including a compressor mechanism into which a gas is received substantially at a suction pressure and from which the gas is discharged substantially at a discharge pressure, the gas discharged from the compression mechanism carrying pressure pulses having a particular frequency and wavelength, these pressure pulses being of variable amplitude. Also provided is a conduit through which gas substantially at discharge pressure flows, and means for reducing the amplitude of the pressure pulses at locations at which the amplitudes reach their highest absolute values.




The present invention further provides a method for reducing the amplitude of pressure pulses having a particular wavelength in a fluid, including: flowing the pressure pulse-containing fluid through a conduit; attenuating the pressure pulse amplitude at a first location along the conduit; and further attenuating the pressure pulse amplitude at a second location along the conduit distanced from the first location a distance which is substantially equal to an odd multiple of one quarter of the wavelength.




A head spacer is provided for increasing the volume of a discharge chamber in the cylinder head assembly of a reciprocating piston compressor, in which the head spacer is disposed between a valve plate and a cylinder head, and has a plurality of substantially concentric, alternating ridges and valleys disposed around the periphery of first and second end surfaces of the head spacer. When the cylinder head is torqued down onto the cylinder block in response to a compressive load exerted on the cylinder head during the assembly of the cylinder head assembly, the tips of the ridges deform to form a continuous labyrinth seal between the head spacer and the cylinder head, and between the head spacer and the valve plate.




The head spacer may be made from an injection-molded plastic, and has a coefficient of thermal expansion which is substantially similar to the metal components of the cylinder head assembly, such that the head spacer may shrink and/or expand at the same rate as the cylinder block and cylinder head. Further, the plastic from which the head spacer is made provides insulation against thermal conduction between the valve plate and the cylinder head.




In one form thereof, a reciprocating piston compressor is provided, including cylinder block having a cylinder bore; a piston reciprocatingly disposed in the cylinder bore; a cylinder head connected to the cylinder block and partially defining a suction chamber into which gas is received and from which the gas exits into the cylinder bore substantially at a suction pressure, the cylinder head partially defining a discharge chamber into which gas is received from the cylinder bore and from which the gas exits substantially at a discharge pressure; a valve plate having a suction port through which the cylinder bore and the suction chamber fluidly communicate, and a discharge port through which the cylinder bore and the discharge chamber fluidly communicate; a suction check valve disposed over the suction port and past which gas flows from the suction chamber to the cylinder bore, flow from the cylinder bore to the suction chamber being inhibited by the suction check valve; a discharge check valve disposed over the discharge port and past which gas flows from the cylinder bore to the discharge chamber, flow from the discharge chamber to the cylinder bore being inhibited by the discharge check valve; and a spacer disposed between the valve plate and the cylinder head, the spacer having generally opposite first and second end surfaces, each of the first and second spacer and surfaces respectively abutting an interfacing surface of the valve plate and the cylinder head, the spacer partially defining the discharge chamber, a substantial portion of the volume of the discharge chamber located between spacer end surfaces; wherein the first and second spacer end surfaces are each provided with a plurality of substantially concentric ridges having tips, the ridge tips having one of a deformed state and an undeformed state, adjacent ones of the ridges separated by a valley, the ridge tips being placed in the deformed state in response to a compressive load exerted on the spacer between the valve plate and the cylinder head during assembly of the compressor, the deformed ridge tips providing a seal between the first spacer end surface and the valve plate, and between the second spacer end surface and the cylinder head.




In a further form thereof, a cylinder head spacer for a reciprocating piston compressor is provided, including a body portion made of a plastic material and having a substantially open interior extending between first and second end surfaces; and a plurality of substantially concentric, alternating ridges and valleys extending around a periphery of each of the first and second end surfaces, the ridges having one of a deformed state and an undeformed state, the ridges being placed in the deformed state in response to a compressive load exerted on the first and second end surfaces, such that the ridges extend into the valleys and contact adjacent ridges to form sealing surfaces, the sealing surfaces coplanar with the first and second end surfaces.




In another form thereof, a method of assembling a reciprocating piston compressor having a cylinder block with a cylinder bore opening, a valve plate, and a cylinder head, is provided, including the steps of providing a spacer having first and second end surfaces each provided with a plurality of substantially concentric ridges having tips, the ridge tips having one of a deformed state and an undeformed state, adjacent ones of the ridge tips separated by a valley; orienting the valve plate, the spacer, and the cylinder head in a stack arrangement over the cylinder bore opening; and exerting a compressive load on the ridge tips to deform the ridge tips to the deformed state, the deformed ridge tips providing sealing contact between the first spacer end surface and the valve plate, and between the second spacer end surface and the cylinder head.




In a still further form thereof, a method is provided of assembling a cylinder head assembly of a reciprocating piston compressor, the compressor having a cylinder block with a bolt hole therein, including the steps of providing a bolt, a suction leaf plate, a valve plate, and a cylinder head, each of which include a bolt hole therein; providing a spacer having a bolt hole, and first and second end surfaces each provided with a plurality of continuous, alternating ridges and valleys extending around a periphery of each of the first and second end surfaces, the ridges including tips having one of a deformed state and an undeformed state; positioning the suction leaf plate, the valve plate, the spacer, and the cylinder head, respectively, on the cylinder block such that the bolt holes are aligned; inserting the bolt through the bolt holes, and tightening the bolt to exert a compressive load on the ridge tips and deforming the ridge tips to the deformed state, the deformed ridge tips providing sealing contact between the first spacer end surface and the valve plate, and between the second spacer end surface and the cylinder head.




One advantage of the present head spacer is that it is inexpensively produced, and, because the head spacer comprises an individual component, the head spacer may used with existing compressor designs without retooling other components of the cylinder head assembly.




Another advantage is that the labyrinth seal produced by the deformation of the ridge tips of the head spacer obviates the need for additional seals between the head spacer and the valve plate, and between the head spacer and the cylinder head.




A further advantage is that the plastic material of the head spacer both provides insulation against thermal conduction between the cylinder block and the cylinder head, and has a coefficient of thermal expansion substantially similar to the other metal components of the cylinder head assembly to prevent the leakage due to the shrinkage and expansion which is observed with existing head spacers.




Another aspect of the inventive noise attenuation system for a compression device relates a suction tube assembly which extends between the stator end cap and the inlet to the compression mechanism, and may be telescoped in the general direction of the stator's longitudinal axis to accommodate stators of different heights. Certain embodiments of this suction tube assembly include a muffler, and this muffler may have a location which is fixed relative to the compression mechanism.




Accordingly, the present invention provides a compressor assembly including a compression mechanism having an inlet into which a gas substantially at suction pressure is received, and an outlet from which gas compressed by the compression mechanism is discharged substantially at a discharge pressure. A motor is also included which includes a rotor and a stator, the stator substantially surrounding the rotor and having an end, the rotor operatively coupled with the compression mechanism. An end cap is disposed over the stator end, the end cap having an interior in which is gas substantially at suction pressure. A suction tube of variable length is also provided through which the compression mechanism inlet and the end cap interior are in fluid communication, the suction tube comprising first and second tubes which are in sliding, telescoping engagement, whereby the length of the suction tube may be adjusted through relative axial movement of the first and second tubes.




The present invention also provides a compressor assembly including a compression mechanism having an inlet into which a gas substantially at suction pressure is received, and an outlet from which gas compressed by the compression mechanism is discharged substantially at a discharge pressure, and a motor having a rotor and a stator selected from a plurality of stators of differing heights. The stator substantially surrounds the rotor and has opposite ends distanced by the stator's height. The rotor is operatively coupled with the compression mechanism. An end cap is disposed over one of the stator ends and has an interior substantially at suction pressure, and first and second telescopingly engaged tubes defining a suction tube which extends axially over at least a portion of the stator height and through which the end cap interior and the compression mechanism inlet are in fluid communication. The suction tube has a length which is varied in response to the relative axial positions of the telescopingly engaged first and second tubes, whereby the suction tube length may be varied to accommodate a different stator alternatively selected from the plurality of stators.




Further, the present invention provides a compressor assembly including a compression mechanism having an inlet into which a gas substantially at suction pressure is received, and an outlet from which gas compressed by the compression mechanism is discharged substantially at a discharge pressure, and a motor having a rotor and a stator selected from a plurality of stators of differing heights. The stator substantially surrounds the rotor and has opposite ends distanced by the stator's height. The rotor is operatively coupled with the compression mechanism. An end cap is disposed over the stator and has an interior in which is gas substantially at suction pressure, the end cap being distanced from the compression mechanism inlet an amount dependent upon the stator's height. A tube assembly is provided through which gas is directed from the end cap interior to the compression mechanism inlet, the tube having means for adjusting its length, whereby the compressor assembly could alternatively comprise a different stator selected from the plurality of stators.




Still another aspect of the inventive noise attenuation system for a compression device relates to motor/compressor assembly mounts which are attached to the interior of the compressor housing in a manner which reduces or eliminates natural resonant frequencies of the housing. The mounts are distributed unequally about the inner circumference of the housing and attached thereto a positions which do not coincide with nodes of these frequencies. That is, the mounts are secured to the inside of the housing to interfere with the wave form produced by the natural frequencies in the compressor housing so as to reduce objectionable noise. Resonation of the housing at these natural frequencies is thus prevented, and the compressor quieted.




Accordingly, the present invention provides a compressor assembly including a housing having at least one natural frequency having a wave form with amplitude large enough for the housing, when vibrated at that frequency, to produce an objectionable noise. The natural frequency wave form has a plurality of natural nodes equally distributed about the circumference of the housing and natural anti-nodes located between adjacent natural nodes. A motor/compressor assembly is also provided which includes a compression mechanism in which gas is compressed from substantially a suction pressure to substantially a discharge pressure, and a motor operably engaged with the compressor mechanism. A plurality of mounts are unequally distributed about the circumference of the housing, the motor/compressor assembly being supported within the housing by the mounts. Each mount is attached to the housing at a first point, the first points not coinciding with the natural nodes of the natural frequency wave form. These first points define forced nodes on the circumference of the housing to which the nodes of the natural frequency wave form are forced, and the natural frequency wave form is altered in response to the natural nodes being forced to the forced nodes, whereby the housing is prevented from vibrating at the natural frequency.











BRIEF DESCRIPTION OF THE DRAWINGS




The above-mentioned and other features and advantages of this invention, and the manner of attaining them, will become more apparent and the invention itself will be better understood by reference to the following description of an embodiment of the invention taken in conjunction with the accompanying drawings, wherein:





FIG. 1

is a first longitudinal sectional view of a first embodiment of a compressor in accordance with the present invention;





FIG. 2

is a second longitudinal sectional view of the compressor shown in

FIG. 1

, along line


2





2


;





FIG. 3

is a sectional view of the compressor shown in

FIG. 1

, along line


3





3


;





FIG. 4

is a sectional view of the compressor shown in

FIG. 1

, along line


4





4


;





FIG. 5

is a sectional view of the compressor shown in

FIG. 1

, along line


5





5


;





FIG. 6

is a bottom view of the crankcase of the compressor shown in

FIG. 1

;





FIG. 7A

is a first side view of the suction muffler of the compressor shown in

FIG. 1

;





FIG. 7B

is a second side view of the suction muffler shown in

FIG. 7A

;





FIG. 7C

is a third side view of the suction muffler shown in

FIG. 7A

, in an alternative configuration in which the inlet tube thereof is shortened;





FIG. 8A

is an enlarged plan view of the valve assembly of the compressor shown in

FIG. 1

;





FIG. 8B

is an exploded side view of the valve assembly shown in

FIG. 8A

;





FIG. 9A

is a first plan view of a discharge tube of the compressor shown in

FIG. 1

, the discharge tube including a discharge muffler;





FIG. 9B

is a second plan view of the discharge tube of

FIG. 9A

;





FIG. 10

is a first longitudinal sectional view of a second embodiment of a compressor according to the present invention;





FIG. 11

is a second longitudinal sectional view of the compressor shown in

FIG. 10

, along line


11





11


;





FIG. 12

is a third longitudinal sectional view of the compressor shown in

FIG. 10

, along line


12





12


;





FIG. 13

is a sectional view of the compressor shown in

FIG. 10

, along line


13





13


;





FIG. 14

is a bottom view of the compressor shown in

FIG. 10

;





FIG. 15

is a sectional view of the compressor shown in

FIG. 10

, along line


15





15


, in which the motor and compression mechanism are not shown;





FIG. 16

is a sectional view of the compressor shown in

FIG. 10

, along line


16





16


, in which the motor, compression mechanism, bottom housing, and discharge tube are not shown;





FIG. 17

is a bottom view of the crankcase of the compressor shown in

FIG. 10

;





FIG. 18A

is a plan view of one embodiment of the head spacer included in the compressor shown in

FIG. 10

;





FIG. 18B

is a side view of the head spacer shown in

FIG. 18A

;





FIG. 18C

is a perspective view of an alternative embodiment of the head spacer included in the compressor shown in

FIG. 10

;





FIG. 18D

is a partial sectional view of the head spacer of

FIG. 18C

, showing the spacer prior to installation;





FIG. 18E

is a partial sectional view of the head spacer of

FIG. 18D

, showing the spacer installed;





FIG. 19A

is a side view of the suction muffler of the compressor shown in

FIG. 10

;





FIG. 19B

is a longitudinal sectional view of the suction muffler shown in

FIG. 19A

;





FIG. 20

is a longitudinal sectional view of the first discharge muffler of the compressor shown in

FIG. 10

;





FIG. 21

is a view of the discharge tube of the compressor shown in

FIG. 10

, the discharge tube including the second discharge muffler;





FIG. 22

is a longitudinal sectional view of the second discharge muffler of the compressor shown in

FIG. 10

;





FIG. 23A

is a schematic view of the primary pumping pulse in the discharge refrigerant in the compressor of

FIG. 10

for various distances between the first and second mufflers of that compressor;





FIG. 23B

is a schematic view of the amplitude of the primary pumping pulse in the discharge refrigerant in the compressor shown in

FIG. 10

, after passing through the first and second mufflers spaced a distance D;





FIG. 23C

is a schematic view of the amplitude of the primary pumping pulse in the discharge refrigerant in the compressor shown in

FIG. 10

, after passing through the first and second mufflers spaced a distance D′;





FIG. 24

is a perspective view of a compressor housing showing the formation of a vibration at a natural frequency; and





FIG. 25

is a sectional view of the compressor shown in

FIG. 5

, schematically illustrating a natural frequency wave form and a forced frequency wave form in the compressor housing.











Corresponding reference characters indicate corresponding parts throughout the several views. Although the drawings represent embodiments of the present invention, the drawings are not necessarily to scale and certain features may be exaggerated in order to better illustrate and explain the present invention.




DETAILED DESCRIPTION OF THE INVENTION




Referring to

FIGS. 1 and 2

there is shown a first embodiment of a reciprocating piston compressor assembly according to the present invention. Reciprocating piston compressor assembly


20


is a hermetic compressor assembly which may be part of a refrigeration or air-conditioning system (not shown). Compressor


20


is a 5-ton compressor having a displacement of approximately 5.6 cubic inches. Compressor assembly


20


comprises housing


22


having an interior surface to which mounts


24


are attached (FIGS.


1


-


5


). Mounts


24


include springs which resiliently support motor/compressor assembly


26


, to vibrationally isolate the motor/compressor assembly from housing


22


in a manner that will be described hereinbelow. Motor/compressor assembly


26


comprises motor


28


and compression mechanism


30


. In the depicted embodiment, compression mechanism


30


is of the reciprocating piston type, although it is to be understood that certain aspects of the present invention may be adapted to other types of compressor assemblies. Previous reciprocating piston compressors are described in U.S. Pat. No. 5,224,840 (Dreiman et al.) and U.S. Pat. No. 5,951,261 (Paczuski), the disclosures of which are expressly incorporated herein by reference. These incorporated patents are assigned to the assignee of the present invention.




Motor


28


comprises stator


32


which is provided with windings


33


, and rotor


34


as illustrated in FIG.


2


. Alternating current from an external power source (not shown) is directed through stator windings


33


via terminal cluster


35


(

FIGS. 3

,


4


and


5


) to electromagnetically induce rotation of rotor


34


. Crankshaft


36


extends longitudinally through central aperture


37


in rotor


34


to which it is rotatably attached to drive compression mechanism


30


. Shaft


36


is operably coupled to a pair of pistons


38


which are reciprocatively disposed in cylinder bores


40


formed in cylinder block


41


of cast-iron crankcase


42


, which is attached to the lower one of two opposite ends of the stator.




During compressor operation, refrigerant at suction pressure is drawn into housing


22


; compressor assembly


20


is a low-side compressor, motor


28


being in a low pressure and low temperature environment. The suction pressure refrigerant is drawn into housing


22


through inlet


45


which is held securely within aperture


47


located in the side of housing


22


by welding, brazing or the like (FIG.


3


). As illustrated in

FIG. 3

, inlet


45


is substantially aligned with suction inlet


46


located in one side of motor end cap


44


such that as suction pressure refrigerant is drawn into housing


22


, a portion of the fluid enters motor end cap


44


through inlet


46


. The remainder of the suction pressure fluid circulates within housing


22


. The suction pressure refrigerant which flows into motor end cap


44


, flows over the top of motor


28


to cool the top end thereof. The refrigerant exits motor end cap


44


through suction tube


48


which leads to inlet


50


of suction muffler


52


. Suction muffler


52


is a steel, expansion type muffler shown in

FIGS. 7A-7C

and includes expansion chamber


54


having a volume of 3.531 cubic inches. Alternatively, suction muffler


52


may be modified such that its expansion chamber


54


has a volume of 4.63 cubic inches. Suction muffler


52


has inlet


50


and outlet


56


which are sealingly connected to suction tubes


48


and


58


, respectively (FIGS.


7


A-


7


C). Suction tubes


48


and


58


have a diameter of approximately ⅞ inch and along with muffler


52


are constructed from a material such as steel. Although the openings suction tubes


48


and


58


are shown as being substantially offset within expansion chamber


54


(FIG.


7


A), muffler


52


may be modified to more closely align these openings so that fluid may flow more directly between them within chamber


54


. Moreover, those of ordinary skill in the art will recognize that the extent to which the ends of tubes


48


and


58


extend into chamber


54


may vary considerably depending on the frequency being attenuated within the muffler.




As shown in

FIGS. 1 and 2

, suction tube


58


is received in one end of suction plenum


60


which is secured at end


62


to cylinder head inlets


64


of cylinder head


66


. Suction plenum


60


is a plastic insert into which steel tube


58


is interference fitted and is held in place over suction chamber


61


in cylinder head


66


by strap


68


. Suction muffler


52


is tuned to attenuate noises created by suction check valves and pressure pulses having a frequency between 1000 and 1400 hertz.




Referring to

FIGS. 7A-7C

, in the shown embodiment of the present invention, suction tube


48


includes first tube


70


and second tube


72


in which the outer and inner diameters, respectively, are telescopically engaged. Suction tube


48


is constructed from steel, but may be constructed from any suitable material to withstand the compressor environment. First tube


70


has an outer diameter of ⅞ inch and is of a slightly smaller diameter than second tube


72


, which has an outer diameter of 1 inch. A sealing member such as O-ring


73


is disposed between first tube


70


and second tube


72


so as to sealingly engage the inner surface of second tube


72


with the outer surface of first tube


70


. First portion


70


is then telescopically movable within second tube


72


to provide an adjustable suction tube


48


having different lengths to accommodate different stator heights H (FIG.


2


), i.e., the distance between the opposite ends of a stator. The position of muffler


52


is such that tubes


70


and


72


are axially aligned along the general direction of the stator height.




As shown in

FIG. 2

, from suction muffler


52


, suction pressure gas is introduced into suction plenum


60


and into suction chamber


61


of cylinder head


66


from which the gas is drawn by the retreating pistons


38


through the suction check valves of valve assembly


74


(FIGS.


8


A and


8


B), and into cylinder bores


40


, wherein the gas is compressed to substantially discharge pressure. Cylinder head


66


is a material such as cast iron or aluminum. Once compressed, the discharge pressure gases flow past the discharge valve of valve assembly


74


and into discharge chamber


76


defined within cylinder head


66


. Discharge chamber


76


of this embodiment is of a size which is great enough to act as an expansion type muffler wherein the amplitude of the pressure wave of the compressed fluid is altered, thereby attenuating the noise created by the operation of the discharge valves and the primary pumping frequency. The volume of discharge chamber


76


is 6.93 cubic inches.




In the usual fashion, valve assembly


74


is provided between crankcase


42


and cylinder head


66


to direct the suction pressure and discharge pressure gases into and out of cylinder bores


40


. Valve assembly


74


is illustrated in

FIGS. 8A and 8B

and includes valve plate


78


having centrally located suction ports


80


and surrounding discharge ports


81


shown in dashed lines in FIG.


8


A. Discharge ports


81


are disposed beneath retaining plate


82


. Valve plate


78


and retaining plate


82


are constructed from a material such as steel. Between valve plate


78


and retaining plate


82


are discharge check valves


84


which open and close discharge ports


81


. Discharge check valves


84


are made from spring steel, as is well known in the art. Each discharge check valve


84


prevents gas in discharge chamber


76


from being drawn into a cylinder bore


40


through the associated discharge ports


81


of valve plate


78


. Discharge pressure gas in cylinder bores


40


is expelled through the discharge ports of valve plate


78


, past discharge check valves


84


, and into discharge chamber


76


, from which it is expelled through outlet


86


of cylinder head


66


into discharge tube


88


(FIGS.


1


and


3


).




Positioned on the opposite side of valve plate


78


are a pair of pins


90


which are aligned across suction ports


80


and fixed to valve plate


78


. Thin metal suction check valves


92


are constructed from spring steel as are discharge check valves


84


and include a pair of slots


93


, one being disposed at opposite ends of valves


92


(FIG.


8


B). Suction check valves


92


are positioned so that pins


90


are received within slots


93


to guide valves


92


between open and closed position. Suction valves


92


prevent gas in cylinder bores


40


from being expelled into suction chamber


61


in cylinder head


66


through suction ports


80


of valve plate


78


. In this particular compressor


20


, two valve assemblies


74


are provided on common plate


78


, one valve assembly


74


being disposed over each cylinder bore


40


.




The discharge pressure gases in discharge


76


are directed into a discharge tube which, as shown, may be comprised of multiple, series-connected tubes. The discharge tube extends from head


66


through aperture


94


in housing


22


, and is connected to the remainder of the refrigerant system (

FIGS. 1

,


3


, and


4


). This housing aperture is sealed about the discharge tube by any suitable manner. Referring now to

FIGS. 9A and 9B

there is shown discharge tube


96


which comprises part of the compressor discharge tube assembly. Discharge tube


96


is somewhat flexible in nature so that shocks associated with pressure pulses may be absorbed by the resilient flexing of tube


96


. Discharge tube


96


is secured to discharge tube


88


at


98


by any suitable method such as welding or brazing (FIG.


1


).




Located along discharge tube


96


is expansion type discharge muffler


100


which is a second muffler of compressor assembly


20


for further reducing the undesirable noise in the refrigerant gas. Flow of compressed refrigerant gas is directed along discharge tube


88


and discharge tube


96


in the direction of arrows


102


through muffler


100


(

FIGS. 1

,


9


A and


9


B). Both discharge tube


88


and discharge tube


96


are approximately ½ inch in diameter and are formed from a material such as steel. Muffler


100


is specifically spaced from discharge chamber


76


in cylinder head


66


in accordance with the present invention as will be described hereinbelow.




Referring to

FIGS. 10

,


11


and


12


, compressor assembly


104


is a second embodiment of a reciprocating piston compressor assembly according to the present invention. Compressor


104


is a 3-ton compressor having a displacement of approximately 3.5 cubic inches. Compressor assembly


104


is similar in structure and operation to compressor assembly


20


except as described herein. Suction pressure gases enter compressor housing


22


through inlet


45


which is held securely within aperture


47


located in the side of housing


22


by welding, brazing or the like. As illustrated in

FIGS. 10 and 13

, inlet


45


is substantially aligned with suction inlet


46


located in one side of motor end cap


44


such that as suction pressure refrigerant is drawn into housing


22


, a portion of the fluid enters inlet


46


into motor end cap


44


. The remainder of the suction pressure fluid circulates within housing


22


. The suction pressure refrigerant which flows into motor end cap


44


, flowing over the top of motor


28


to cool the top end thereof. The refrigerant exits motor end cap


44


through suction tube


48


which leads to inlet


50


of suction muffler


106


as shown in

FIGS. 12

,


19


A, and


19


B. Suction muffler


106


is of a Helmholtz type having tube


108


, which is part of suction tube


48


, provided with a plurality of axially-spaced hole arrangements


110


therealong. Tube


48


is constructed from a material such as steel or the like and has a diameter of approximately {fraction (


3


/


4


)} inch. Each arrangement of holes


10


comprises two pairs of holes


112


, the holes in each arrangement are cross drilled so that the holes are equally radially distributed about the circumference of tube


108


. Notably, each hole arrangement


110


is substantially equally spaced along the longitudinal axis of tube


108


. The number and size of holes


112


is dependant on the frequencies which are being attenuated. In this embodiment, holes


112


are formed in tube


108


by any suitable manner such as being punched or drilled and have a diameter of {fraction (3/16)} inch to attenuate noise created by the operation of valve arrangement


74


and the primary pumping frequency. Tube


108


is surrounded by shell


114


having ends


116


and


118


which are sealed to the exterior surface of tube


108


to create chamber


120


around hole arrangements


110


(

FIGS. 12

,


19


A, and


19


B). Shell


114


is made from any suitable material such as steel and has a volume of 1.16 cubic inches which is also dependant on the frequencies in the primary pumping pulse being attenuated.




As with compressor assembly


20


, the suction gas exits suction muffler


106


and enters cylinder head assembly


122


which includes cylinder head


66


covering a head spacer disposed between valve plate


78


and cylinder head


66


, as described in more detail below (FIG.


12


).




Cylinder head


66


and the head spacer together define enlarged suction chambers


126


and discharge chamber


128


therein which help to alleviate efficiency problems experienced with some reciprocating compressors. These problems include discharge pressure gas within discharge chamber


128


not readily exiting cylinder head


66


, resulting in a pressure buildup in discharge chamber


128


during compressor operation. Consequently, cylinder bore


40


may not be fully exhausted of discharge pressure gas at the end of the compression cycle because the buildup of gas within discharge chamber


128


inhibits the accommodation therein of gas being exhausted thereinto from cylinder


40


. Because gas from the previous compression cycle has not been fully exhausted from cylinder bore


40


, less suction pressure gas can be drawn into cylinder


40


during the next compression cycle. Thus, the efficiency of the compressor is compromised. Moreover, the temperature of gas on the discharge side of the system may become excessively high as more and more work is expended on the gas already at discharge pressure.




A first embodiment of head spacer


124


is shown in

FIGS. 18A and 18B

and provides means for enlarging suction chamber


126


and discharge chamber


128


(FIG.


12


). Spacer


124


includes body portion


130


having a substantially open interior extending between first planar end surface


132




a


and substantially parallel second planar end surface


132




b


with fastener apertures


134


therein. Head spacer


124


may be constructed from any suitable material including metal or plastic. Cylindrical portions


136


define suction passageways


138


therethrough, and are connected to body portion


130


by bridge portions


140


. The remainder of the substantially open interior of body portion


130


partially defines discharge chamber


128


, and cooperates with cylinder head


66


to form an enlarged discharge chamber


128


. Head spacer


124


thereby cooperates with cylinder head


66


to effectively increase the volume of discharge chamber


128


of cylinder head assembly


122


, in order to prevent the buildup of discharge pressure gas within discharge chamber


128


. Discharge chamber


128


therefore may accommodate a greater volume of discharge gas, allowing substantially all of the discharge gas to be exhausted from cylinder bores


40


during the operation of compressor


104


, improving the efficiency of compressor


104


. When assembling cylinder head assembly


122


, first and second end surfaces


132




a


,


132




b


of head spacer


124


are sealed with the adjacent surfaces of cylinder head


66


and valve assembly


78


, respectively, by gaskets (not shown). Compressor assembly


20


of the first embodiment is not provided with head spacer


124


due to a lack of clearance within housing


22


, however, if space were available, head spacer


124


would improve the efficiency of compressor


20


in the same manner as described above. As noted above, the discharge chamber within the head generally acts as an expansion chamber muffler, and enlargement of this chamber generally improves its effectiveness as such.




The second embodiment of head spacer


124


′, shown in

FIGS. 18C-18E

, which is provided with an alternative sealing method between surfaces


132




a


′ and


132




b


′ of spacer


124


′ and valve plate assembly


78


and cylinder head


66


. Spacer


124


′ may be formed of an injection-molded plastic. The plastic material has a coefficient of thermal expansion which is substantially similar to the metal components of cylinder assembly


122


, including cylinder block


41


and cylinder head


66


, such that head spacer


124


′ may shrink and/or expand at substantially the same rate as cylinder block


41


and cylinder head


66


. The plastic material of which head spacer


124


′ is formed provides insulation against thermal conduction between discharge chamber


128


and suction chamber


126


. One suitable plastic for head spacer


124


′ is PLENCO® a phenolic molding compound, Product No. 6553, available from Great Lakes Plastics, 7941 Salem Rd., Salem, Mich., which, after curing, has a coefficient of linear expansion of 12×10


−6


mm/mm/° C. (25° C. to 190° C.). (PLENCO® is a registered trademark of Plastics Engineering Co., 3518 Lakeshore Rd., Sheboygan, Wis.)




Referring to

FIGS. 18D and 18E

, the alternative method of accomplishing the above described sealing engagement of head spacer


124


′ includes providing a series or plurality of concentric, continuous ridges


142


on substantially parallel planar end surfaces


132




a


′,


132




b


′ of head spacer


124


′ disposed around the periphery of body portion


130


′ having corresponding and alternating ridge tips


144


and valleys


146


. As may be seen in

FIGS. 18C

, ridge tips


144


and valleys


146


are continuous, and circumferentially extend around the periphery of first and second end surfaces


132




a


′,


132




b


′ of head spacer


124


′. Referring again to

FIG. 18D

, ridges


142


are shown in an undeformed state, where tips


144


extend a first distance D


1


from each of first and second planar end surfaces


132




a


′,


132




b


′, and valleys


146


extend a second distance D


2


from each of planar first and second end surfaces


132




a


′,


132




b


′ opposite tips


144


. As shown in

FIG. 18D

, first distance D


1


is approximately twice the length of second distance D


2


, but may vary substantially. First and second end surfaces


132




a


′,


132




b


′ lie in planes perpendicular to a line L


1


-L


1


, which defines a central axis of head spacer


124


′. When head spacer


124


′ is placed between valve plate


78


and cylinder head


66


during assembly of cylinder head assembly


122


, and a compressive load is exerted upon cylinder head assembly


122


, for example, by torquing down fasteners such as bolts (not shown) to tighten cylinder head


66


, ridge tips


144


plastically deform to a deformed state as shown in FIG.


18


E.




In the deformed state shown in

FIG. 18E

, ridge tips


144


are deformed by the planar interfacing surfaces


148


,


150


of cylinder head


66


and valve plate


78


, respectively, into a generally mushroom shape in which portions of ridge tips


144


extend into adjacent valleys


146


, and portions of adjacent ridge tips


144


may contact one another to form sealing surface


152


between head spacer


144


and cylinder head


66


, as well as between head spacer


124


′ and valve plate


78


. Sealing surfaces


152


, created by the deformation of ridge tips


124


′, define labyrinth seals


154


. Labyrinth seals


154


are tortuous arrangements of deformed ridge tips


144


which seal discharge gas within discharge chamber


128


at the interface of head spacer


124


′ and cylinder head


66


, as well as at the interface of head spacer


124


′ and valve plate


78


. Labyrinth seals


154


sufficiently seal head spacer


124


′ between cylinder head


66


and valve plate


78


, obviating the need for additional seals. It may be seen from

FIG. 18E

that the interfacing surfaces of cylinder head


66


and valve plate


78


respectively lie in first and second planes which are respectively substantially coincident with the third and fourth planes defined by first and second end surfaces


132




a


′,


132




b


′ of head spacer


124


′, respectively, when the fasteners are tightened to torque cylinder head


66


down onto head spacer


124


′, valve plate


78


, and cylinder block


41


, and causing ridge tips


144


to deform to form labyrinth seals


154


.




Generally, during the assembly of compressor


104


and cylinder head assembly


122


, cylinder head


66


, valve plate


78


, and head spacer


124


are positioned respectively adjacent one another, in a stacked arrangement on cylinder block


41


, such that cylinder bores


40


are covered, and fastener apertures


134


in head spacer


124


and the foregoing components are aligned. Fasteners are then inserted through apertures


134


in cylinder head assembly


122


to engage cylinder block


41


and exert a compressive load on cylinder head assembly


122


. This tightens cylinder head assembly


122


down onto cylinder block


41


, which, seals adjacent surfaces


132




a


,


132




b


and cylinder head


66


and valve plate


78


, respectively. In the case of the alternative sealing method, ridge tips


144


of head spacer


124


′ are compressed from the undeformed state shown in

FIG. 18D

to the deformed state shown in

FIG. 18E

, providing sealing surfaces


152


and labyrinth seals


154


between head spacer


124


′ and cylinder head


66


, and between head spacer


124


′ and valve plate


78


.




The flow of gas through compressor assembly


104


is similar to that of compressor assembly


20


. The suction pressure gas flows into suction chamber


126


defined in cylinder head


66


and head spacer


124


. From chamber


126


, the suction pressure gas passes through suction ports


80


(

FIG. 8A

) of valve plate


78


into cylinder bores


40


where the refrigerant is compressed to a substantially higher discharge pressure. The compressed fluid flows through discharge ports


81


of valve plate


78


into discharge chamber


128


also defined by cylinder head


66


and head spacer


124


. The discharge pressure gas in chamber


128


exits cylinder head assembly


122


through discharge outlet


86


illustrated in FIG.


10


and enters first muffler


156


(

FIGS. 10

,


11


,


12


and


20


).




Referring now to

FIG. 20

, it can be seen that first muffler


156


comprises tube


160


having a diameter of approximately ⅝ inch, which may be a part of discharge tube


88


. Tube


160


extends through generally cylindrical shell


162


having first and second ends


164


and


166


. Shell ends


164


and


166


are sealed to the exterior surface of tube


160


and within shell


162


, tube


160


is provided with a plurality of hole arrangements


168


. Each arrangement of holes


168


comprises three pairs of holes


170


, the holes in each arrangement may be cross drilled so that the holes are equally radially distributed about the circumference of tube


160


. In this embodiment, each hole


170


is formed in the shape of an ellipse having an area of 0.0345 square inches. Notably, each arrangement of holes


168


are substantially equally spaced along the longitudinal axis of tube


160


. It is understood that holes


170


may be of any shape and size that adequately attenuate noise in the discharge pressure refrigerant.




As with compressor


20


, referring now to

FIG. 21

, there is shown discharge tube


96


which may be part of discharge tube


88


both of which being approximately ½ inch in diameter and constructed from steel. Located in discharge tube


96


is second muffler or resonator


158


as shown in greater detail in FIG.


22


. Like the first muffler


156


, second resonator


158


comprises part of a tube which extends through a shell, the tube within the shell having a plurality of spaced hole arrangements. As shown in

FIG. 22

, tube


171


extends through shell


172


which has first and second ends


174


and


176


. Ends


174


and


176


of the generally cylindrical shell


172


are sealed to the exterior surface of tube


171


. A plurality of hole arrangements


178


are axially spaced along tube


171


within shell


172


, each arrangement of holes


178


comprising a plurality of holes


180


. As described above, holes


180


may be cross drilled or punched through tube


171


, thereby equally radially distributing the holes about the circumference of the tube. Holes


180


are of similar size and shape to holes


170


of first muffler


156


. Second muffler


158


is spaced from first muffler


156


along discharge tube


96


a specific distance to better attenuate noises in the primary pumping pulse in the discharge pressure refrigerant as will be described hereinbelow.




It is to be noted that although first and second mufflers


156


and


158


depicted are of the Helmholtz type, it is to be understood that the present invention may be practiced using first and second mufflers which are merely expansion chambers. Such mufflers would not have a tube extending longitudinally through the muffler, but rather would have a tube which enters into the expansion chamber, which may be defined by shells


162


and


172


, and a tube which exits from the shell, the interior of the mufflers being open and hollow.




Compression devices such as hermetic compressors


20


and


104


(

FIGS. 1

,


2


, and


10


-


12


) are driven at a particular frequency which correlates directly with the speed at which driving motor


28


disposed within compressor shell or housing


22


rotates. As described above, motor


28


, which is well known in the art, has rotor


34


which is electromagnetically induced into rotation by current directed through windings


33


in stator


32


. Shaft


36


extending longitudinally through rotor


34


drives compression mechanism


30


. Thus, the frequency of the pressure pulses will be directly correlated to the speed of motor


28


. The speed of motor


28


in compressors


20


and


104


is approximately 3450 to 3500 rpm which directly correlated to the frequency of the alternating current which powers motor


28


. Thus, the frequency of the pulse which is associated with the frequency of the alternating current which powers motor


28


, can be predicted with accuracy because the cycle of the electrical power is a known quantity. For example, in the United States, electrical power of the alternating current type is normally provided at a 60 hertz cycle.




The cyclical pulsations in the refrigerant which result from its compression within compression mechanism


30


and which is directly and most elementally correlated to frequency of the electrical power which drives motor


28


, may be referred to as the primary pumping frequency within the primary pumping pulse. The primary pumping frequency will also be affected by the number of compression chambers which are compressing the fluid directed through discharge tube


88


. For example, a reciprocating piston type compressor may have a single cylinder and piston. Thus, the primary pumping frequency will be a factor of one times the frequency at which electrical power is provided to the motor. Similarly, as is the case with compressors


20


and


104


, a reciprocating compressor which has two cylinders


40


and pistons


38


driven off common shaft


36


will have a primary pumping frequency which is twice that of the single piston type compressor. Accordingly, a three piston type compressor will have a pumping frequency which is three times that of the single piston type compressor, and so on.




The primary pumping frequency wave form in the primary pumping pulse in the discharge pressure refrigerant has both a standing or nonmoving component as well as a traveling component, each of which having different amplitudes to produce different sounds or noises. The amplitude of the standing wave is much greater than the traveling wave and has fixed peaks and valleys as depicted in FIG.


23


B. The traveling wave (not shown) has a much smaller amplitude that produces much less noise during compressor operation than the standing wave. The amplitude of the traveling wave is reduced as the wave moves along a muffler or resonator, no specific placement of the muffler is required because the points of amplitude maximum absolute value (i.e., the points of lowest minimum or highest maximum amplitude) of the primary pumping frequency are not fixed. However, in order to effectively reduce the amplitude of the frequency of the standing wave, the muffler must be placed at the fixed points of amplitude maximum absolute value (i.e, the points of lowest minimum or highest maximum amplitude) of the primary pumping frequency wave form.




A single Helmholtz muffler is capable of reducing the amplitude of very specific frequencies, however, only in a narrow band width. Expansion mufflers are capable of reducing the amplitude of frequencies in a wide band width, however, the amplitudes attenuated are much lower than a Helmholtz resonator. In order to effectively reduce the noise produced during compressor operation by the primary pumping pulse, a single muffler and the compressor discharge chamber, or a pair of mufflers, are spaced along the discharge tube, at specifically calculated points in the primary pumping frequency wave form as is discussed below.




In accordance with the present invention the first and second mufflers of both compressors


20


and


104


are placed in series along the discharge tube assembly at a specific distance from one another, that distance corresponding to that distance between the expected minimum and maximum amplitudes of the primary pumping frequency wave form in the refrigerant. In compressors


20


and


104


, a problematic or noisy frequency is produced by a discharge pulse within the primary pumping pulse having a frequency of approximately 1400 hertz created by operation of discharge valve


84


of valve assembly


74


. Accordingly, discharge chamber


76


in cylinder head


66


and mufflers


100


,


156


and


158


are tuned and axially spaced along discharge tube


88


to reduce the amplitude of the discharge pulse at a frequency of 1400 hertz. It is understood that the mufflers are tuned for the use of refrigerant R22, if an alternative refrigerant were used in compressors


20


and


104


, the mufflers would have to be retuned.




The first muffler, which is essentially discharge chamber


76


in cylinder head


66


of compressor


20


, and first muffler


156


of compressor


104


, which may be positioned at any point downstream of head


66


, establish an initial point from which wavelength λ is measured. With reference now to

FIG. 23B

, wavelength λ is represented by a sine wave which begins at point A and ends at point B. Although

FIG. 23B

shows that point A coincides with a node or a point of minimum amplitude of the wave, it is to be understood that this placement of the first muffler need not be at such a node. In any case, the amplitude of the pressure wave exiting the first muffler will be reduced, at that frequency, relative to its amplitude prior to entering the first muffler. Thus wave form


182


extends for one complete wavelength λ between points A and B. As depicted in

FIG. 23B

, where wave form


182


has a node coinciding with point A, one half of wavelength λ also occurs at a node, as does the point of wave form


182


which coincides with point B. At one quarter and three quarters the length of wavelength λ from point A, it can be seen that wave form


182


has maximum amplitudes


184


and


186


. Those of ordinary skill in the art will recognize that at any other odd multiple of one quarter λ, wave form


182


will also be at a point of maximum absolute amplitude value. As shown in

FIG. 23B

, distance D is that distance from point A to the point of maximum amplitude


184


at one quarter λ and distance D′ is the distance between point A at maximum amplitude


186


at three quarter λ. These distances D and D′ correspond to the spacing between the first and second muffler as illustrated in FIGS.


23


A. The mufflers in FIG.


23


A are represented as mufflers


156


and


158


of compressor


104


, however, it is understood that mufflers


76


and


100


of compressor


20


could be represented in place of mufflers


156


and


158


, respectively. Wave form


182


demonstrates a frequency and general character of a pressure wave, the relationship between the wave form being that frequency of the primary pumping frequency. Thus the structure of the present invention can be established with help of the following equation:






c/f=λ






where c equals the speed of sound in the compressed refrigerant; f equals the primary pumping frequency; and λ is the wavelength.




The operating speed of compressors


20


and


104


running on a 60 hertz electrical input is 58 hertz. Compressors


20


and


104


being two cylinder type piston compressors, the primary pumping frequency is 2 times 58 hertz which approximately equals 116 hertz. This is incorporated into the above equation. The speed of sound in refrigerant is 7200 inches per second, however, this may vary with temperature and pressure.




The resulting λ is 62 inches. The point of maximum amplitude


184


at one quarter λ is thus 15½ inches. Thus, in order to further attenuate the amplitude of the pumping pulse in the discharge fluid, second muffler


100


or


158


should be located at a distance D of 15½ inches from first muffler


76


or


156


, respectively. Alternatively, second muffler


100


or


158


can be located at distance D′ from first muffler


76


or


156


, this distance corresponding to three quarters of the length λ or 46½ inches. Thus, by means of the present invention, the second muffler, by being placed at a particular distance corresponding to points of maximum amplitude of the pressure pulses in the primary pumping frequency, from the first muffler, the noise associated with the primary pumping frequency can be effectively and further attenuated vis-a-vis previous systems having but a single discharge muffler. The two mufflers of each compressor do not necessarily have to be precisely placed at 15½ inches from each other and may be placed a distance of approximately 12-20 inches apart before reaching a higher discharge pulse near a node.




With reference to mufflers


156


and


158


of the Helmholtz type, as shown in

FIG. 23A

, distances D and D′ shall be most effectively extended from the furthest downstream arrangement of holes


170


E in first muffler


156


and furthest upstream arrangements of holes


180


A in second muffler


158


. By so arranging the first and second Helmholtz type mufflers


156


and


158


, the greatest attenuation of the primary pumping pulse can be achieved by the first muffler, the second muffler having the greatest opportunity then to further attenuate the pumping pulse which reaches it.




Referring again to

FIG. 23B

as discussed above, wave form


182


represents a sine wave, which may be representative of the pressure pulse between the two mufflers, demonstrating the wavelength and points of maximum amplitude


184


and


186


along wavelength λ. The diminishing wave form is further shown in

FIGS. 23B

has a first amplitude A


1


before entering first mufflers


76


and


156


. After passing through the first mufflers, the amplitude of waveform


182


at point


184


is reduced at


188


to having an amplitude of A


2


(FIG.


23


B). With second mufflers


100


and


158


located at distance D from first mufflers


76


and


156


, respectively, it can be seen in

FIG. 23C

that wave form


182


will enter second mufflers


100


and


158


having an amplitude of A


2


and will be reduced as at


190


to having an amplitude of A


3


upon exiting the second mufflers. Similarly, with second mufflers


100


and


158


located at a distance D′ corresponding to point of maximum amplitude


186


, at three quarter λ, it can be seen that the amplitude A


2


of wave form


182


will be reduced as the refrigerant passes through second mufflers


100


and


158


, to a modified wave form shown at


190


having a reduced amplitude A


3


(FIG.


23


B).




Although compressors


20


and


104


depict that first muffler


76


,


156


and second mufflers


100


,


158


are packaged within housing


22


, it is to be understood that the separation of the first and second mufflers may be achieved in a discharge line external to housing


22


. The placement of the first and second mufflers along discharge tube


96


within housing


22


improves the packaging characteristics of compressors


20


and


104


, but is not a necessary aspect of the present invention.




During the operation of compressor assemblies


20


and


104


, the cylindrical shape of housing


22


has several natural resonant frequencies that produce loud, pure tones which are undesirable. In order to reduce or eliminate these frequencies, resilient mounts


24


illustrated in

FIGS. 5 and 16

are welded to housing


22


so as to span a node and an anti-node of the wave form. Mounts


24


are secured at


196


to crankcase


42


and at


198


to the inner surface of housing


22


by means such as weldment. The natural frequencies associated with housing


22


may have any number of nodes. The most problematic or noticeable frequency


193


is one in which there are six naturally occurring nodes


192


and anti-nodes


194


circumferentially spaced around housing


22


at equal distances (FIG.


24


).




To reduce the amount of noise produce by this natural frequency, the nodes and anti-nodes must be forced to an alternative position by specifically securing mounts


24


to housing


22


at points which are unequally distributed about the circumference of housing


22


and which do not coincide with naturally occurring nodes. The forced frequency


193


′ produced by mounts


24


is illustrated in FIG.


25


and is represented by dashed lines. It is critical that mounts


24


are unequally distributed about the circumference of housing


22


because if they were equally distributed, forced nodes


192


′ and anti-nodes


194


′ would fall on those of natural frequencies and thus the amplitude of the natural frequency would not be attenuated.




Referring to

FIG. 25

, one of ends


198


of each mount


24


is welded to the inside surface of housing


22


at positions offset from naturally occurring nodes


192


. The weld forces nodes


192


′, dampening the vibrations in housing


22


created by the natural frequency. The weld at opposite end


198


of mount


24


is then located so as to force anti-node


194


′ or points of maximum amplitude between two nodes. Forced anti-nodes


194


′ are then free to vibrate and cause tones which produce noise. These tones, however, are at a much lower amplitude which do not produce the same objectionable noise of the natural resonant frequencies.




While this invention has been described as having exemplary designs, the present invention may be further modified within the spirit and scope of this disclosure. Therefore, this application is intended to cover any variations, uses, or adaptations of the invention using its general principles. For example, aspects of the present invention may be applied to compressors other than reciprocating piston compressors. Further, this application is intended to cover such departures from the present disclosure as come within known or customary practice in the art to which this invention pertains.



Claims
  • 1. A reciprocating piston compressor, comprising:a cylinder block having a cylinder bore; a piston reciprocatingly disposed in said cylinder bore; a cylinder head connected to said cylinder block and partially defining a suction chamber into which gas is received and from which the gas exits into said cylinder bore substantially at a suction pressure, said cylinder head partially defining a discharge chamber into which gas is received from said cylinder bore and from which the gas exits substantially at a discharge pressure; a valve plate having a suction port through which said cylinder bore and said suction chamber fluidly communicate, and a discharge port through which said cylinder bore and said discharge chamber fluidly communicate; a suction check valve disposed over said suction port and past which gas flows from said suction chamber to said cylinder bore, flow from said cylinder bore to said suction chamber being inhibited by said suction check valve; a discharge check valve disposed over said discharge port and past which gas flows from said cylinder bore to said discharge chamber, flow from said discharge chamber to said cylinder bore being inhibited by said discharge check valve; and a spacer disposed between said valve plate and said cylinder head, said spacer having generally opposite first and second end surfaces, each of said first and second spacer end surfaces respectively abutting an interfacing surface of said valve plate and said cylinder head, said spacer partially defining said discharge chamber, a substantial portion of the volume of said discharge chamber located between said spacer end surfaces; wherein said first and second spacer end surfaces are each provided with a plurality of substantially concentric ridges having tips, said ridge tips having one of a deformed state and an undeformed state, adjacent ones of said ridges separated by a valley, said ridge tips being placed in said deformed state in response to a compressive load exerted on said spacer between said valve plate and said cylinder head during assembly of said compressor, said deformed ridge tips providing a seal between said first spacer end surface and said valve plate, and between said second spacer end surface and said cylinder head.
  • 2. The compressor assembly of claim 1, wherein said cylinder block and said spacer have substantially similar coefficients of thermal expansion.
  • 3. The compressor assembly of claim 1, wherein said cylinder head and said spacer have substantially similar coefficients of thermal expansion.
  • 4. The compressor assembly of claim 1, wherein said spacer-interfacing surfaces of said cylinder head and said valve plate respectively lie in first and second substantially parallel planes, said spacer extending between third and fourth substantially parallel planes which are substantially parallel to said first plane, said valley of said first and second spacer end surfaces located between said third and fourth planes, and, with said ridge tips in their said undeformed state, said ridge tips of said first and second spacer end surfaces are disposed on sides of said third and fourth planes opposite their respectively adjacent valleys.
  • 5. The compressor assembly of claim 4, wherein said first and third planes, and said second and fourth planes, respectively coincide.
  • 6. The compressor assembly of claim 5, wherein, with said ridge tips in their said deformed state, said ridge tips of said first and second spacer end surfaces lie on said third and fourth planes.
  • 7. The compressor assembly of claim 4, wherein each said valley of one of said first and second spacer end surfaces extend a first distance from the respective one of said third and fourth planes, and said ridge tips of said one of said first and second spacer end surfaces extend a second distance from the said respective one of said third and fourth planes, said first distance greater than said second distance.
  • 8. The compressor assembly of claim 7, wherein said first distance is approximately twice said second distance.
  • 9. A method of assembling a reciprocating piston compressor having a cylinder block with a cylinder bore opening, a valve plate, and a cylinder head, comprising the steps of:providing a spacer having first and second end surfaces each provided with a plurality of substantially concentric ridges having tips, the ridge tips having one of a deformed state and an undeformed state, adjacent ones of the ridge tips separated by a valley; orienting the valve plate, the spacer, and the cylinder head in a stacked arrangement over the cylinder bore opening; and exerting a compressive load on the ridge tips to deform the ridge tips to the deformed state, the deformed ridge tips providing sealing contact between the first spacer end surface and the valve plate, and between the second spacer end surface and the cylinder head.
CROSS REFERENCE TO RELATED APPLICATION

This application is a division of application Ser. No. 09/994,236, filed Nov. 27, 2001, (now U.S. Pat. No. 6,558,137) which claims the benefit under 35 U.S.C. §119(e) U.S. Provisional Patent Application Serial No. 60/250,709, filed Dec. 1, 2000.

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Provisional Applications (1)
Number Date Country
60/250709 Dec 2000 US