EP2198125B1

EP2198125B1 - Screw compressor pulsation damper

Info

Publication number: EP2198125B1
Application number: EP07839131.5A
Authority: EP
Inventors: Paul J. Flanigan; Bruce A. Fraser
Original assignee: Carrier Corp
Current assignee: Carrier Corp
Priority date: 2007-10-01
Filing date: 2007-10-01
Publication date: 2017-06-21
Anticipated expiration: 2027-10-01
Also published as: CN101809251B; WO2009045187A1; CN101809251A; EP2198125A4; US20100209280A1; ES2629981T3; EP2198125A1

Description

BACKGROUND

The present invention relates generally to screw compressors. Screw compressors typically comprise a pair of counter-rotating, mating male and female screws that have an intermeshing plurality of lands and channels, respectively, that narrow from an inlet end to a discharge end such that an effluent working fluid or gas, or some other such working matter, is reduced in volume as it is pushed through the screws. The discharged working matter is released in pulses as each mating land and channel pushes a volume of the working matter out of the compressor. Each pulse comprises a burst of wave energy that propagates through the working matter and the screw compressor as the working matter leaves the screws. The screw compressors are typically turned by motors operating at speeds such that the wave pulsations are discharged at a high frequency. The pulsations not only produce vibration of the screw compressor, but also produce noise that is amplified by the working matter and the compressor. Such vibration is undesirable as it wears components of the compressor and produces additional noise as the compressor vibrates. Noise from the discharging working matter and vibrating compressor is undesirable as it results in loud operating environments. Previous attempts to counteract these problems have involved mufflers, padded mounts and clamps that are mounted external to the screw compressor. These solutions do not address the underlying source of the noise and vibration and only provide after-the-fact countermeasures. In addition to adding cost and weight, such solutions provide only limited noise reduction and do not prevent wear on internal screw compressor components. Other solutions have proposed acoustic barriers that prevent pulsation damage to screw compressor components, but do not attenuate screw compressor noise or vibration. There is, therefore, a need for screw compressors having reduced effects from discharge pulsations.
US 3,146,720 discloses a positive displacement pump with sound attenuating means, and the features in the preamble to claim 1 are disclosed in US 2006/0165543 .

SUMMARY

The present invention provides a screw compressor in accordance with claim 1.
Exemplary embodiments of the invention include a screw compressor comprising a housing, a slide valve assembly and a pulsation damper. The housing receives a supply of working matter from a pair of intermeshing screw rotors, and comprises a slide recess, a pressure pocket, and a piston cylinder. The slide valve assembly regulates the capacity of the screw compressor, and comprises a slide valve axially movable within the slide recess and the pressure pocket, a piston head axially movable within the piston cylinder, and a piston shaft connecting the slide valve with the piston head. The pulsation damper comprises a flange for separating the pressure pocket from the piston cylinder, a bore for receiving the piston rod, and a damping channel extending through the flange for damping pressure pulsations in the working matter discharged from the pair of intermeshing screw rotors.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 shows a partially cutaway perspective view of a screw compressor in which the pulsation damper of the present invention is used.
FIG. 2 shows a schematic diagram of the screw compressor of FIG. 1 showing an outlet case incorporating the pulsation damper.
FIG. 3 shows a partially cutaway perspective view of the outlet case of FIG. 2 showing a plurality of damping channels comprising the pulsation damper.

DETAILED DESCRIPTION

FIG. 1 shows a partially cutaway perspective view of screw compressor 10, which compresses a working fluid or gas such as a refrigerant that is typically used in refrigeration or air conditioning systems. Screw compressor 10 includes rotor case 12, outlet case 14, slide case 16, male screw rotor 18, female screw rotor 20, drive motor 22 and slide valve assembly 23. Male screw rotor 18 and female screw rotor 20 are disposed within rotor case 12 and include shafting and bearings such that they can be rotationally driven by drive motor 22. For example, male screw rotor 18 includes shaft 24A (which extends axially through rotor case 12 and into motor 22 and rests on bearing 26A), and shaft 24B (which extends axially into outlet case 14 and rests in bearing 26B). Refrigerant is introduced into rotor case 12 at suction port 28, directed around motor 22 and into suction pocket 30 at the inlet of screw rotors 18 and 20. Male screw rotor 18 and female screw rotor 20 include meshing grooves and lands that form helical flow paths having decreasing cross sectional areas as the grooves and lands extend from suction pocket 30. Thus, the refrigerant is reduced in volume and pressurized as the refrigerant is directed into discharge pocket 32 by screw rotors 18 and 20, before being discharged at pressure port 34 and released to, for example, a condenser or evaporator of a cooling system. Slide valve assembly 23, which includes slide valve 36, piston rod 38, piston head 40 and spring assist 42, regulates the discharge capacity of screw compressor 10. In particular, piston head 38, piston rod 40 and spring assist 42, through a control system, translate slide valve 36 axially between rotors 18 and 20 to vary the volume of refrigerant compressed in the helical flow paths. Due to typically high speeds that motor 22 drives screw rotors 18 and 20, the multiple sets of meshing grooves and lands comprising the helical flow paths discharge the refrigerant into pressure pocket 32 in a series of high frequency pulsations, which effectuates undesirable noise and vibration. Outlet case 14 includes a pulsation damper that mitigates the pulsation effects of the discharged refrigerant. In the embodiment shown, screw compressor 10 comprises a two-screw compressor. However, in other embodiments, the present invention is readily applicable to compressors having three, four our more screw rotors that employ a reciprocating slide valve system.
FIG. 2 shows a schematic diagram of screw compressor 10 of FIG. 1, having pulsation damping means of the present invention. In particular, outlet case 14 includes damping channels 46A and 46B that attenuate the pulsation effects of refrigerant R within screw compressor 10. Screw compressor 10 also includes rotor case 12, slide case 16, female screw rotor 20, drive motor 22, slide valve assembly 23 (including slide valve 36, piston rod 38, piston head 40 and spring assist 42) and control system 48. Rotor case 12 includes slide recess 51, slide stop 52 and recirculation passage 53. Slide case 16 includes piston cylinder 54, and outlet case 14 includes rod flange 58. Together, rotor case 12, outlet case 14 and slide case 16 comprise a sealed flow path for directing refrigerant R through screw compressor 10. Refrigerant R is directed into rotor case 12 at suction port 28, and routed around motor 22 to suction pocket 30. Refrigerant R from suction pocket 30 is compressed by male screw rotor 18 (not shown) and female screw rotor 20 and discharged into pressure pocket 32. Female screw rotor 20 includes screw channels, or grooves, 50A - 50D that mesh with mating lands or lobes on male screw rotor 18 to form a sealed, decreasing-volume flow path. The sealed flow path decreases in volume such that refrigerant R is pushed and compressed as it moves from suction pocket 30 to discharge pocket 32. Accordingly, refrigerant R enters, for example, screw channel 50A at suction pocket 30 having pressure P₁ and is discharged from the same screw channel 50A at discharge pocket 32 having elevated pressure P₂. Thus, each screw channel delivers a small volume of refrigerant R to discharge pocket 32. As screw rotors 18 and 20 rotate, a series of discharge pulses of refrigerant R is released to discharge pocket 32, which causes undesirable noise and vibration of screw compressor 10. Outlet case 14 includes damping channels 46A and 46B, which act as pulsation dampers to reduce the noise and vibration effects of refrigerant R as it is discharged from screw rotors 18 and 20.
Outlet case 14, which includes discharge pocket 32, is disposed between rotor case 12 and slide case 16 such that it receives the high pressure side of screw rotors 18 and 20 at a first end, and piston rod 38 of slide assembly 23 at a second end. Slide valve 36 of slide assembly 23 is positioned within slide recess 51 of rotor case 12 such that it is disposed between male screw rotor 18 and female screw rotor 20. Slide valve 36 is connected with piston rod 38 and piston head 40 such that slide valve 36 can be axially withdrawn from slide recess 51 and extended into pressure pocket 32 to control the amount of pressurized refrigerant R entrained within screw channels 50A - 50D. For example, slide valve 36 can be extended to the fully-loaded position (to the left in FIG. 1) such that it abuts slide stop 52 and contacts the entire length of screw rotors 18 and 20. Thus, the capacity of screw compressor 10 is maximized by maximizing the amount of refrigerant R compressed in the lands and grooves of screw rotors 18 and 20. From the fully-loaded position, slide valve 36 is moved toward discharge pocket 32 (to the right in FIG. 1) to open recirculation passage 53, decreasing the discharge capacity of screw compressor 10.
Piston rod 38 extends through rod flange 58 to connect slide valve 36 within rotor case 12 to piston head 40 disposed within piston cylinder 54 of slide case 16. Piston head 40 includes first pressure side 56A, which is exposed to refrigerant R at pressure P₂, and second pressure side 56B, which is exposed to control oil at pressure P₃. Pressure P₂ is dictated by refrigerant R and screw rotors 18 and 20, while pressure P₃ is regulated by control system 48. Based on the loading (i.e. cooling demands) of the refrigerator or air conditioner to which screw compressor 10 is connected, control system 48, which comprises switches, valves, solenoids and the like, selectively provides control oil to piston cylinder 54. Control oil is admitted into piston cylinder 48 to increase pressure P₃ to exert a force on second pressure side 56B to move slide valve 36 toward slide stop 52 within slide recess 51. Pressure P₃ is reduced by removing control oil from piston cylinder 54 such that slide valve 36 can be withdrawn from slide recess 51. Spring assist 42 pushes on first pressure side 56A to assist in withdrawing slide valve 36 from slide recess 51. Piston head 40 is also in contact with refrigerant R, which exerts pressure P₂ on first pressure side 56A to push piston head 40 away from rod flange 58. Refrigerant R, is admitted into piston cylinder 54 through damping channels 46A and 46B disposed within rod flange 58. Damping channels 46A and 46B, piston cylinder 54 and rod flange 58 are configured to attenuate vibration and noise associated with the discharge of refrigerant R from screw rotors 18 and 20. Specifically, damping channels 46A and 46B act in concert with piston cylinder 54 to provide a Helmholtz resonator to absorb energy from the discharged pulses of refrigerant R.
FIG. 3 shows a partially cutaway perspective view of slide valve assembly 23 of FIG. 2, in which damping channels 46A - 46C of rod flange 58 are shown. Slide valve assembly 23 also includes slide valve 36, piston rod 38, piston head 40 and spring assist 42, which is omitted in FIG. 3 for clarity. Slide valve assembly 23 extends axially through rotor case 12, outlet case 14 and slide case 16 along an actuation path defined by slide recess 51, pressure pocket 32 and piston cylinder 54. Outlet case 14 is positioned within screw compressor 10 such that first end A connects with rotor case 12, and second end B connects with slide case 16. Slide valve 36 extends from slide recess 51 in rotor case 12 where it is disposed between rotor screws 18 and 20, and into pressure pocket 32 within outlet case 14. Piston rod 38 extends axially from slide valve 36 through central bore 60 in rod flange 58 of outlet case 14, and into piston cylinder 54 of slide case 16 where rod 38 connects with piston head 40.
Rod flange 58 comprises a collar positioned on second end B of outlet case 14 such that central bore 60 axially aligns with slide recess 51 (in which slide valve 36 translates within rotor case 12) and piston cylinder 54 (in which piston head 40 translates within slide case 16). In the embodiment shown, rod flange 58 is integrally cast or formed with outlet case 14 along second end B. Rod flange 58 separates piston cylinder 54 from slide recess 51 and pressure pocket 32 to form two separate chambers for refrigerant R. Rod flange 58 is provided with seal or bearing ring 62 and is attached to rod flange 58 with snap rings 64A and 64B, which are disposed within grooves in ring 62. In one embodiment, ring 62 comprises a seal and prevents refrigerant R from entering piston cylinder 54 between piston rod 38 and rod flange 58 at bore 60. In another embodiment, seal ring 62 comprises a bearing that assists in sliding of piston rod 38 though rod flange 58 as well as performing sealing functions. Damping channels 46A - 46C, however, permit refrigerant R to enter piston cylinder 54 within slide case 16.
Slide case 16 comprises piston cylinder 54, which forms an annular extension of outlet case 14 to accommodate piston rod 38 and piston head 40. Piston head 40 divides piston cylinder 54 into discharge side 54A and control side 54B. Piston head 40 includes seal 65 to prevent flow of control oil and refrigerant R past piston head 40. Piston cylinder 54, therefore, comprises a sealed canister for actuating piston head 40. Discharge side 54A of this sealed canister, however, also acts as a resonance chamber, that along with damping channels 46A - 46C, absorb some of the vibrational and acoustical effects of the pulsed discharges of refrigerant R.
As explained above, slide valve assembly 23 is connected with control system 48 (FIG. 2) to actuate the position of slide valve 36 along rotor screws 18 and 20. Slide valve 36 is translated to regulate the discharge capacity of refrigerant R from screw rotors 18 and 20. Control system 48 regulates flow of the control oil into control side 54B of piston cylinder 54 to vary pressure P₃. Refrigerant R flows into damping channels 46A - 46C into piston cylinder 54 within slide case 16 to pressurize discharge side 54A of piston cylinder 54 to pressure P₂. Refrigerant R is compressed to pressure P₂ between screw rotors 18 and released in pulsed discharges into pressure pocket 32 at slide valve 36 as screw rotors 18 and 20 counter-rotate to open and close the helical flow paths formed by the lobes and channels. The pulsed discharges of refrigerant R flow past rod flange 58 before being discharged from screw compressor 10 at pressure port 34 (FIG. 1). Damping channels 46A - 46C extend through rod flange 58 and permit refrigerant R to enter and pressurize piston cylinder 54 to pressure P₂.
In the embodiment shown in FIG. 3, rod flange 50 includes four damping channels: damping channels 46A - 46C, each disposed in a quadrant of rod flange 50, and a fourth damping channel omitted due to the section taken out of FIG. 3. Damping channels 46A - 46C comprise hollowed-out chambers extending through rod flange 58 of outlet case 14. The lengths of damping channels 46A - 46C are determined by the thickness of rod flange 58, but can be altered by inserting hollow damping tubes 66A - 66C into damping channels 46A - 46C. Damping tubes 66A - 66C are inserted into damping channels 46A - 46C such that they extend into piston cylinder 54 and into pressure pocket 32. As is illustrated in FIG. 3, damping tube 66A is inserted into damping channel 46A, and damping tube 66B is inserted into damping channel 46B. In the embodiment shown, damping tubes 66A - 66C each having the same length and the same diameter. In one embodiment, damping tubes 66A - 66C comprise stainless steel tubes press fit into damping channels 46A - 46C. The specific quantity and geometry of damping channels 46A - 46C and damping tubes 66A - 66C, however, is selected to dampen the acoustic and vibrational pulsation effects of refrigerant R, and can thus vary depending on the specific design parameters of screw compressor 10. Specifically, the number, length and diameter of damping tubes 66A - 66C are selected to extract the maximum amount of energy from refrigerant R as refrigerant R travels through tubes 66A - 66C into the resonance chamber formed by discharge side 54A.
Refrigerant R is discharged from screw rotors 18 and 20 in pulses at regular intervals having a frequency dictated by the speed at which motor 22 drives screw rotors 18 and 20. These pulses therefore produce undesirable sound waves that increase the noise generated by screw compressor 10. The energy contained in these sound waves, however, can be used to do work to attenuate the propagation of the sound waves from screw compressor 10. Outlet case 14 and slide case 16 are configured to function as a Helmholtz resonator, which comprises a container of fluid or gas having a necked opening, such as is produced by discharge side 54A, refrigerant R and channels 46A - 46C. A Helmholtz resonator utilizes the spring-like compressibility of the fluid or gas to extract energy from a wave oscillating at a given frequency. Refrigerant R fills discharge side 54A such that additional refrigerant attempting to enter discharge side 54A through channels 46A - 46C must compress the volume of refrigerant R already present within discharge side 54A. Thus, a pulsed wave of refrigerant R attempting to enter discharge side 54A, compresses refrigerant R until the crest of the wave is reached. Then, the pressurized refrigerant R within discharge side 54A will push back as the wave dissipates to the trough. As the pulsed wave propagates through crests and waves, the pressurized refrigerant R within discharge side 54A continues to compress and decompress, thus extracting energy from refrigerant R discharged from screw rotors 18 and 20. The energy extraction reduces the amplitude of the pulsation wave, thereby reducing noise and vibration generated by the pulsed discharges of refrigerant R.
A Helmholtz resonator extracts the maximum amount of energy from the fluid or gas when the frequency of the wave matches the natural or resonance frequency of the Helmholtz resonator. Thus, the resonance frequency of the Helmholtz resonator produced by discharge side 54A and damping channels 46A - 46C can be configured to match that of the pulsation discharges of refrigerant R as produced by motor 22. Equation (1) illustrates the resonance frequency of an elongate tube used in a Helmholtz resonator, where f_R is the resonance frequency of the tube, v is the speed of sound in the medium filling the tube, A₀ is the area of the tube, L is the length of the tube and V₀ is the volume of resonance chamber. $f_{R} = \frac{v}{2 π} \sqrt{\frac{A_{0}}{V_{0} L}}$
For the present invention, the tube or "necked opening" of the Helmholtz resonator comprises the aggregate of tubes 66A - 66C. Applying this equation to the embodiment of the present invention shown in FIG. 3, f_R is the resonance frequency of tubes 66A - 66C, v is the speed of sound in refrigerant R, A₀ is the total cross-sectional area of tubes 66A - 66C, L is the length of one of tubes 66A - 66C, and V₀ is the volume of discharge side 54A. The dimensions of tubes 66A - 66C are selected such that the frequency of the discharge pulses of refrigerant R from screw rotors 18 and 20 at a given capacity matches the resonance frequency of the tubes. For example, in one embodiment of the invention, screw compressor 10 is configured to operate at 3,600 RPM at full load. Volume V₀ , therefore, comprises the volume of discharge side 54A when piston head 40 is furthest away from rod flange 50 (all the way to the left in FIG. 3), and frequency f_R is 60 Hz. Thus, the areas and lengths of tubes 66A - 66C are selected based on other design requirements, such as dimensional constraints of rod flange 58 and slide case 16. Additionally, the number of tubes can be selected based on specific design considerations. In the embodiment shown, tubes 66A - 66C have the same lengths and diameters. Thus, screw compressor 10 is provided with a pulsation damper that is configured for damping pulsation effects of refrigerant R at a specific operating condition. However, in other embodiments, tubes 66A - 66C can have different geometries, such as different lengths and/or different diameters, such that the pulsation damper is tuned to one specific resonance frequency, or can attenuate vibration and acoustic effects over a range of frequencies. In other embodiments of the invention, rod flange 58 comprises a circular disk or annular ring that can be bolted or otherwise secured to piston cylinder 54 within slide case 16 such that pulsation dampers configured for different resonance frequencies can be interchangeably installed into screw compressor 10.
While the invention has been described with reference to an exemplary embodiment(s), It is intended that the invention is not be limited to the particular embodiment(s) disclosed, but that the invention will include all embodiments falling within the scope of the appended claims.

Claims

A screw compressor (10) comprising:
a housing for receiving a supply of working matter, the housing comprising:
a rotor case (12) have a suction pocket (30) and a slide recess (51);

a discharge case (14) having a discharge pocket (32) axially aligned with the slide recess (51); and

a slide case (16) having a piston cylinder (54) axially aligned with the

discharge pocket (32);

a pair of intermeshing screw rotors (18, 20) disposed within the rotor case (12) between the suction pocket (30) and the slide recess (51) for compressing the working matter and discharging the working matter into the discharge pocket (32);

a slide valve assembly (23) disposed adjacent the pair of intermeshing screw rotors (18, 20) and axially movable within the slide recess (51), discharge pocket (32) and piston cylinder (54) to regulate the capacity of the screw compressor (10);

the slide valve assembly (23) comprising:
a slide valve (36) axially movable within the slide recess (51) and the

discharge pocket (32);

a piston head (40) axially movable within the piston cylinder (54); and

a piston rod (38) connecting the slide valve (36) with the piston head (40); and

a pulsation damper carried by the discharge case (14) to dampen pressure pulsations of working matter discharged from the screw rotors (18, 20) to the discharge pocket (32) and passing into the piston cylinder (54); characterised in that:
the pulsation damper comprises:
a flange (58) for separating the discharge pocket (32) from the piston cylinder (54);

a bore (60) for receiving the piston rod (38); and

a damping channel (46A, 46B) extending through the flange (58).
The screw compressor of claim 1, wherein the pulsation damper comprises:
a resonance chamber enclosed within the piston cylinder (54) between the piston head (40) and the flange (58) such that working matter pressurizes the resonance chamber.
The screw compressor of claim 2, wherein the pulsation damper comprises a plurality of damping channels (46A, 46B, 46C).
The screw compressor of claim 3, wherein the lengths and diameters of the damping tubes (46A, 46B, 46C) are selected to produce a Helmholtz resonator having a natural frequency matching a frequency of the discharged working matter.
The screw compressor of claim 3 and further comprising a plurality of damping tubes (46A, 46B, 46C) inserted through the plurality of damping channels (46A, 46B, 46C).
The screw compressor of claim 1, wherein the pulsation damper further comprises a seal (62) positioned between the bore (60) and the flange (58).