THRUST BEARING PLACEMENT FOR COMPRESSOR

Abstract

A compressor includes a shaft, a motor configured to drive the shaft into rotation, and a thrust bearing configured to permit rotation of the shaft and support an axial load of the shaft. The thrust bearing is positioned about the shaft and between the motor and an impeller of the compressor.

Description

CROSS-REFERENCE TO RELATED APPLICATION

This application claims priority to and the benefit of U.S. Provisional Application No. 62/611,722, entitled “THRUST BEARING PLACEMENT FOR COMPRESSOR,” filed Dec. 29, 2017, which is hereby incorporated by reference in its entirety for all purposes.

BACKGROUND

This application relates generally to vapor compression systems such as chillers, and more specifically to a compressor of a chiller.

Vapor compression systems (e.g., chillers) utilize a working fluid, typically referred to as a refrigerant, which changes phase between vapor, liquid, and combinations thereof in response to being subjected to different temperatures and pressures associated with operation of the vapor compression system. For example, a heating, ventilation, air conditioning, and refrigeration (HVAC&R) system may include a chiller, which is a type of vapor compression system that cycles a refrigerant to remove heat from, or cool, a flow of water traversing tubes that extend through a chiller evaporator. The chilled water flow may be directed to nearby structures to absorb heat, or provide cooling, before being cycled back to the chiller evaporator to be cooled once again.

Chillers utilize compressors, such as centrifugal compressors, in order to pump or otherwise move the refrigerant about the chiller. A traditional centrifugal compressor may include a motor which rotates a shaft in order to operate the traditional centrifugal compressor. Unfortunately, certain operating and/or ambient conditions may negatively impact the shaft and associated components, reducing an efficiency of the compressor and corresponding chiller. Accordingly, improved compressors for chillers and/or other vapor compression systems may be desired.

SUMMARY

An embodiment includes a compressor having a shaft, a motor configured to drive the shaft into rotation, and a thrust bearing configured to permit rotation of the shaft and support an axial load of the shaft. The thrust bearing is positioned about the shaft and between the motor and an impeller of the compressor.

Another embodiment includes a heating, ventilation, air conditioning, and refrigerant (HVAC&R) system having a compressor. The compressor includes a shaft, a motor configured to drive the shaft into rotation, an impeller coupled to the shaft and configured to be driven into rotation by the shaft, and a thrust bearing configured to permit rotation of the shaft and to support an axial load of the shaft. The thrust bearing is positioned about the shaft and between the motor and the impeller of the compressor.

Another embodiment includes a centrifugal compressor having a shaft, a motor configured to drive the shaft into rotation, and a thrust bearing configured to permit rotation of the shaft and to support an axial load of the shaft. The thrust bearing is configured to be disposed within a bearing cavity of the centrifugal compressor, about the shaft, and between the motor and an impeller of the centrifugal compressor.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a perspective view of a building that may utilize an embodiment of a heating, ventilation, air conditioning, and refrigeration (HVAC&R) system in a commercial setting, in accordance with an aspect of the present disclosure;

FIG. 2 is a perspective view of an embodiment of a vapor compression system, in accordance with an aspect of the present disclosure;

FIG. 3 is a schematic illustration of an embodiment of the vapor compression system of FIG. 2, in accordance with an aspect of the present disclosure;

FIG. 4 is a schematic illustration of another embodiment of the vapor compression system of FIG. 2, in accordance with an aspect of the present disclosure;

FIG. 5 is a cross-sectional side view of an embodiment of a compressor for use in the vapor compression system of FIG. 2, and having a thrust bearing positioned between a motor of the compressor and an impeller of the compressor, in accordance with an aspect of the present disclosure; and

FIG. 6 is a cross-sectional side view of another embodiment of a compressor for use in the vapor compression system of FIG. 2, and having a thrust bearing positioned between a motor of the compressor and an impeller of the compressor, in accordance with an aspect of the present disclosure.

DETAILED DESCRIPTION

One or more specific embodiments will be described below. In an effort to provide a concise description of these embodiments, not all features of an actual implementation are described in the specification. It should be appreciated that in the development of any such actual implementation, as in any engineering or design project, numerous implementation-specific decisions must be made to achieve the developers' specific goals, such as compliance with system-related and business-related constraints, which may vary from one implementation to another. Moreover, it should be appreciated that such a development effort might be complex and time consuming, but would nevertheless be a routine undertaking of design, fabrication, and manufacture for those of ordinary skill having the benefit of this disclosure.

As set forth above, traditional chillers utilize compressors, such as centrifugal compressors, in order to pump refrigerant about the chiller. A traditional centrifugal compressor may include a motor which rotates a shaft of the compressor in order to operate the traditional centrifugal compressor. Unfortunately, certain operating and/or ambient conditions, such as certain temperatures and/or pressures, may negatively impact the shaft and associated components, reducing an efficiency of the compressor and corresponding chiller. For example, as the operating temperature of the centrifugal compressor increases, the shaft of the centrifugal compressor may thermally expand in an axial direction. Shaft growth due to thermal expansion may impact the axial location of the impeller within a diffuser passage of the compressor. Because appropriate axial location of the impeller within the diffuser passage improves efficiency of the compressor, changes to the axial location of the impeller within the diffuser passage due to temperature increase and corresponding thermal expansion of the shaft may reduce efficiency of the compressor.

In accordance with present embodiments, a thrust bearing is positioned on an impeller-side of the motor, and about the shaft. In other words, the thrust bearing is positioned between a motor of the compressor and an impeller of the compressor, as opposed to an opposing non-impeller-side of the motor. The thrust bearing is configured to permit rotation of the shaft, and to support an axial load of the shaft against the thrust bearing. In doing so, the axial position of the thrust bearing with respect to axially static components of the compressor does not change. Positioning the thrust bearing on the impeller-side of the motor (i.e., between the motor and the impeller) and about the shaft causes a shorter shaft length between the thrust bearing and the impeller, compared to positioning the thrust bearing on the opposing side (i.e., non-impeller-side) of the motor. By reducing the shaft length between the thrust bearing and the impeller, axial displacement of the shaft toward the impeller (i.e., between the thrust bearing and the impeller) caused by temperature changes (i.e., thermal expansion) is reduced, as most of the axial displacement of the shaft caused by thermal expansion occurs on the opposing side of the thrust bearing. Accordingly, axial displacement of the impeller within the diffuser passage is reduced compared to embodiments having no thrust bearing, or a thrust bearing positioned farther away from the impeller. Reducing the axial displacement of the impeller improves efficiency of the compressor, as described above, at least in part because it reduces or negates misalignment of the impeller with respect to the diffuser passage. These and other features will be described in detail below with respect to the drawings.

Turning now to the drawings, FIG. 1 is a perspective view of an embodiment of a heating, ventilation, air conditioning, and refrigeration (HVAC&R) system 10 of a building 12 for a typical commercial setting. The HVAC&R system 10 may include a vapor compression system 14 that supplies a chilled liquid, which may be used to cool the building 12. The HVAC&R system 10 may also include a boiler 16 to supply warm liquid to heat the building 12, and an air distribution system which circulates air through the building 12. The air distribution system can also include an air return duct 18, an air supply duct 20, and/or an air handler 22. In some embodiments, the air handler 22 may include a heat exchanger that is connected to the boiler 16 and the vapor compression system 14 by conduits 24. The heat exchanger in the air handler 22 may receive either heated liquid from the boiler 16 or chilled liquid from the vapor compression system 14, depending on the mode of operation of the HVAC&R system 10. The HVAC&R system 10 is shown with a separate air handler on each floor of building 12, but in other embodiments, the HVAC&R system 10 may include air handlers 22 and/or other components that may be shared between or among floors.

FIGS. 2 and 3 are embodiments of the vapor compression system 14 that can be used in the HVAC&R system 10. The vapor compression system 14 may circulate a refrigerant through a circuit starting with a compressor 32. The circuit may also include a condenser 34, an expansion valve(s) or device(s) 36, and a liquid chiller or an evaporator 38. The vapor compression system 14 may further include a control panel 40 that has an analog to digital (A/D) converter 42, a microprocessor 44, a non-volatile memory 46, and/or an interface board 48.

Some examples of fluids that may be used as refrigerants in the vapor compression system 14 are hydrofluorocarbon (HFC) based refrigerants, for example, R-410A, R-407, R-134a, hydrofluoro olefin (HFO), “natural” refrigerants like ammonia (NH₃), R-717, carbon dioxide (CO₂), R-744, or hydrocarbon based refrigerants, water vapor, or any other suitable refrigerant. In some embodiments, the vapor compression system 14 may be configured to efficiently utilize refrigerants having a normal boiling point of about 19 degrees Celsius (66 degrees Fahrenheit) at one atmosphere of pressure, also referred to as low pressure refrigerants, versus a medium pressure refrigerant, such as R-134a. As used herein, “normal boiling point” may refer to a boiling point temperature measured at one atmosphere of pressure.

In some embodiments, the vapor compression system 14 may use one or more of a variable speed drive (VSDs) 52, a motor 50, the compressor 32, the condenser 34, the expansion valve or device 36, and/or the evaporator 38. The motor 50 may drive a shaft of the compressor 32, and may be powered by a variable speed drive (VSD) 52. The VSD 52 receives alternating current (AC) power having a particular fixed line voltage and fixed line frequency from an AC power source, and provides power having a variable voltage and frequency to the motor 50. In other embodiments, the motor 50 may be powered directly from an AC or direct current (DC) power source. The motor 50 may include any type of electric motor that can be powered by a VSD or directly from an AC or DC power source, such as a switched reluctance motor, an induction motor, an electronically commutated permanent magnet motor, or another suitable motor. The motor 50, the VSD 52, or both may be separate from the compressor 32, or may be partially or fully integrated with the compressor 32. It should be noted that, in certain embodiments, the motor 50 and/or the VSD 52 may be integral with the compressor 32. For example, the motor 50 may be partially or entirely contained within a casing of the compressor 32.

The compressor 32 compresses a refrigerant vapor and delivers the vapor to the condenser 34 through a discharge passage. In some embodiments, the compressor 32 may be a centrifugal compressor. The refrigerant vapor delivered by the compressor 32 to the condenser 34 may transfer heat to a cooling fluid (e.g., water or air) in the condenser 34. The refrigerant vapor may condense to a refrigerant liquid in the condenser 34 as a result of thermal heat transfer with the cooling fluid. The liquid refrigerant from the condenser 34 may flow through the expansion device 36 to the evaporator 38. In the illustrated embodiment of FIG. 3, the condenser 34 is water cooled and includes a tube bundle 54 connected to a cooling tower 56, which supplies the cooling fluid to the condenser.

The liquid refrigerant delivered to the evaporator 38 may absorb heat from another cooling fluid, which may or may not be the same cooling fluid used in the condenser 34. The liquid refrigerant in the evaporator 38 may undergo a phase change from the liquid refrigerant to a refrigerant vapor. As shown in the illustrated embodiment of FIG. 3, the evaporator 38 may include a tube bundle 58 having a supply line 60S and a return line 60R connected to a cooling load 62. The cooling fluid of the evaporator 38 (e.g., water, ethylene glycol, calcium chloride brine, sodium chloride brine, or any other suitable fluid) enters the evaporator 38 via return line 60R and exits the evaporator 38 via supply line 60S. The evaporator 38 may reduce the temperature of the cooling fluid in the tube bundle 58 via thermal heat transfer with the refrigerant. The tube bundle 58 in the evaporator 38 can include a plurality of tubes and/or a plurality of tube bundles. In any case, the vapor refrigerant exits the evaporator 38 and returns to the compressor 32 by a suction line to complete the cycle.

FIG. 4 is a schematic of the vapor compression system 14 with an intermediate circuit 64 incorporated between condenser 34 and the expansion device 36. The intermediate circuit 64 may have an inlet line 68 that is directly fluidly connected to the condenser 34. In other embodiments, the inlet line 68 may be indirectly fluidly coupled to the condenser 34. As shown in the illustrated embodiment of FIG. 4, the inlet line 68 includes a first expansion device 66 positioned upstream of an intermediate vessel 70. In some embodiments, the intermediate vessel 70 may be a flash tank (e.g., a flash intercooler). In other embodiments, the intermediate vessel 70 may be configured as a heat exchanger or a “surface economizer.” In the illustrated embodiment of FIG. 4, the intermediate vessel 70 is used as a flash tank, and the first expansion device 66 is configured to lower the pressure of (e.g., expand) the liquid refrigerant received from the condenser 34. During the expansion process, a portion of the liquid may vaporize, and thus, the intermediate vessel 70 may be used to separate the vapor from the liquid received from the first expansion device 66. Additionally, the intermediate vessel 70 may provide for further expansion of the liquid refrigerant because of a pressure drop experienced by the liquid refrigerant when entering the intermediate vessel 70 (e.g., due to a rapid increase in volume experienced when entering the intermediate vessel 70). The vapor in the intermediate vessel 70 may be drawn by the compressor 32 through a suction line 74 of the compressor 32. In other embodiments, the vapor in the intermediate vessel may be drawn to an intermediate stage of the compressor 32 (e.g., not the suction stage). The liquid that collects in the intermediate vessel 70 may be at a lower enthalpy than the liquid refrigerant exiting the condenser 34 because of the expansion in the expansion device 66 and/or the intermediate vessel 70. The liquid from intermediate vessel 70 may then flow in line 72 through a second expansion device 36 to the evaporator 38.

In accordance with the present disclosure, the compressor 32 in the embodiments illustrated in, and described with respect to, FIGS. 1-4 may include a thrust bearing positioned about a shaft of the compressor 32. The thrust bearing is configured to permit rotation of the shaft, and to support an axial load of the shaft against the thrust bearing. In doing so, the axial distance between the thrust bearing and other axially static components of the compressor 32 does not change. Positioning the thrust bearing on the impeller-side of the motor 50 (i.e., between the motor 50 and the impeller of the compressor 32), in accordance with the present disclosure, causes a shorter shaft length between the thrust bearing and the impeller compared to, for example, positioning the thrust bearing on the opposing side (i.e., non-impeller-side) of the motor. By reducing the shaft length between the impeller and the thrust bearing, which maintains its axial position with respect to axially static components of the compressor 32, axial displacement of the shaft caused by thermal expansion of the shaft toward the impeller (i.e., between the thrust bearing and the impeller) is reduced. In other words, the majority of axial growth of the shaft, if any, due to thermal expansion occurs not within the short shaft length between the thrust bearing and the impeller, but instead within the longer shaft length extending from the other end of the thrust bearing (e.g., toward the motor 50, through the motor 50, and on the non-impeller-side of the motor 50). Accordingly, displacement of the impeller within the diffuser passage (due to temperature changes that cause thermal expansion of the shaft) is reduced compared to embodiments having no thrust bearing, or having a thrust bearing positioned farther from the impeller (e.g., on the non-impeller-side of the motor). By reducing displacement of the impeller, efficiency of the compressor is improved. These and other features will be described in detail below with respect to the drawings.

FIG. 5 is a side view of an embodiment of the compressor 32 for use in the vapor compression system 14 of FIG. 2. The compressor 32 includes a thrust bearing 100 positioned about a shaft 101 of the compressor 32, positioned between the motor 50 of the compressor 32 and an impeller 102 of the compressor 32 with respect to an axial direction 104 (e.g., along a longitudinal axis 115 of the compressor 32). The thrust bearing 100, the shaft 101, the impeller 102, the motor 50, and other features of the compressor 32 may be contained within a casing 103 of the compressor 32.

As previously described, the motor 50 may be integral with the compressor 32. The motor 50 may be configured to rotate the shaft 101 to cause compression of refrigerant passing through the compressor 32 (and to move the refrigerant through a corresponding chiller or vapor compression system). For example, the shaft 101 may cause rotation of the impeller 102, which includes a set of vanes and/or blades that gradually increases the energy of the refrigerant and directs the refrigerant toward a diffuser passage 112 of the compressor 32. In general, the diffuser passage 112 of the compressor 32 may include a vane, vaneless, or variable geometry diffuser, which operates to diffuse the high-energy refrigerant gas. That is, the diffuser passage 112 and corresponding diffuser may convert the kinetic energy of the refrigerant gas into pressure by gradually reducing a velocity thereof. The refrigerant gas may then flow through a collector 113 downstream of the diffuser passage 112. The impeller 102 may be axially positioned (e.g., with respect to direction 104 along the longitudinal axis 115) such that the impeller 102 guides the refrigerant toward appropriate locations of the diffuser passage 112 and corresponding diffuser. By reducing displacement of the impeller 102, misalignment of the impeller 102 with respect to the diffuser passage 112, which would otherwise reduce an efficiency of the compressor 32, is reduced or negated.

As operating and/or ambient temperatures increase during operation of the compressor 32, the shaft 101 may thermally expand along an axial direction 104 (e.g., along the longitudinal axis 115 of the compressor 32). However, the thrust bearing 100 is configured to permit rotation of the shaft 101 while supporting an axial load of the shaft 101. In other words, the thrust bearing 100 maintains its axial position with respect to, for example, the motor 50.

In accordance with present embodiments, the thrust bearing 100 is positioned between the motor 50 of the compressor 32 and the impeller 102 of the compressor 32. In other words, the thrust bearing 100 is positioned on an impeller-side 106 of the motor 50, as opposed to a non-impeller-side 108 of the motor 50. By positioning the thrust bearing 100 on the impeller-side 106 of the motor 50, a shaft length 110 between the thrust bearing 100 and the impeller 102 is less than if the thrust bearing 100 were disposed on the non-impeller-side 108 of the motor 50. Thus, the available shaft length 110 that can thermally expand between the axially static thrust bearing 100 and the impeller 102 is small compared to embodiments having a thrust bearing positioned farther away from the impeller 102. Indeed, as shown, an additional shaft length 111 extending from the thrust bearing 100, through the motor 50, and through the non-impeller-side 108 of the motor 50 is significantly larger than the illustrated shaft length 110 between the thrust bearing 100 and the impeller 102. By reducing the available shaft length 110 that can thermally expand in the axial direction 104 into the impeller 102 as shown and described, an axial displacement of the impeller 102 is reduced. Because appropriate axial location of the impeller 102 with respect to a diffuser passage 112 of the compressor 32 improves efficiency of the compressor 32, the disclosed thrust bearing 100 and corresponding axial location along the shaft 101 (e.g., on the impeller-side 106 of the motor 50, as opposed to the non-impeller-side 108 of the motor 50) improves efficiency of the compressor 32. It should be noted that the compressors 32 of FIGS. 5 and 6 may be centrifugal compressors, hermetic compressors, overhung compressors, or any combination thereof (e.g., hermetic overhung centrifugal compressor). Further, in FIG. 5, the thrust bearing 100 is a lubricated hydrodynamic bearing, whereas in FIG. 6, the thrust bearing 100 is a magnetic bearing. The above-described effects associated with the disclosed positioning the thrust bearing 100 between the motor 50 and the impeller 102 (i.e., along the impeller-side 106 of the motor 50) is applicable to magnetic bearings, anti-friction bearings (e.g., oil or refrigerant lubricated anti-friction bearings such as roller bearings or ball bearings), lubricated hydrodynamic thrust bearings, and any other suitable thrust bearings. It should be noted that the thrust bearing 100 may be positioned within a cavity 105 (e.g., bearing cavity) formed within the compressor 32 (e.g., formed inside the casing 103 and/or with respect to other features of the compressor 32). The cavity 105 may be configured (e.g., sized and/or shaped) to receive the thrust bearing 100, and the thrust bearing 100 may be configured (e.g., sized and/or shaped) to be disposed in the cavity 105.

FIG. 6 is cross-sectional side view of another embodiment of the compressor 32 for use in the vapor compression system 14 of FIG. 2. Similar to the embodiment illustrated in FIG. 5, the compressor 32 illustrated in FIG. 6 includes the thrust bearing 100 positioned about the shaft 101 and on the impeller-side 106 of the motor 50. Accordingly, the shaft length 110 between the thrust bearing 100 and the impeller 102 is reduced compared to embodiments having no thrust bearing or a thrust bearing positioned farther from the impeller (e.g., on the non-impeller-side 108 of the motor 50). By reducing the shaft length 110 between the axially static thrust bearing 100 and the impeller 102, axial displacement of the shaft length 110 between the thrust bearing 100 and the impeller 102 (e.g., caused by increased temperature) is reduced compared to traditional embodiments, thereby reducing an axial displacement of the impeller 102 with respect to the diffuser passage 112. By improving maintenance of the position of the impeller 102 along the axial direction 104, efficiency of the compressor 32 is improved.

As previously described, in FIG. 5, the thrust bearing 100 is a lubricated hydrodynamic bearing, whereas in FIG. 6, the thrust bearing 100 is a magnetic bearing. However, any suitable thrust bearing may be utilized in FIG. 5 and in FIG. 6. For example, the above-described technical effects, which relate to a position of the thrust bearing 100 and corresponding reduction of displacement of the impeller 102 with respect to the diffuser passage 112, may be applicable in embodiments utilizing magnetic bearings, anti-friction bearings (e.g., oil or refrigerant lubricated anti-friction bearings such as roller bearings or ball bearings), lubricated hydrodynamic bearings, and any other suitable thrust bearings.

While only certain features and embodiments have been illustrated and described, many modifications and changes may occur to those skilled in the art (e.g., variations in sizes, dimensions, structures, shapes and proportions of the various elements, values of parameters (e.g., temperatures, pressures, etc.), mounting arrangements, use of materials, colors, orientations, etc.) without materially departing from the novel teachings and advantages of the subject matter recited in the claims. The order or sequence of any process or method steps may be varied or re-sequenced according to alternative embodiments. It is, therefore, to be understood that the appended claims are intended to cover all such modifications and changes as fall within the true spirit of the disclosure. Furthermore, in an effort to provide a concise description of the exemplary embodiments, all features of an actual implementation may not have been described (i.e., those unrelated to the presently contemplated best mode of carrying out the disclosure, or those unrelated to enabling the claimed disclosure). It should be appreciated that in the development of any such actual implementation, as in any engineering or design project, numerous implementation specific decisions may be made. Such a development effort might be complex and time consuming, but would nevertheless be a routine undertaking of design, fabrication, and manufacture for those of ordinary skill having the benefit of this disclosure, without undue experimentation.

Claims

1. A compressor, comprising:

a shaft;

a motor configured to drive the shaft into rotation; and

a thrust bearing configured to permit rotation of the shaft and to support an axial load of the shaft, wherein the thrust bearing is positioned about the shaft and between the motor and an impeller of the compressor.

2. The compressor of claim 1, wherein the compressor comprises a centrifugal compressor.

3. The compressor of claim 1, wherein the compressor comprises a hermetic overhung centrifugal compressor.

4. The compressor of claim 1, wherein the shaft extends within the motor and beyond opposing sides of the motor.

5. The compressor of claim 1, comprising the impeller and a diffuser passage, wherein the impeller is axially aligned with the diffuser passage such that the impeller directs refrigerant toward and into the diffuser passage.

6. The compressor of claim 1, comprising a casing in which the shaft, the motor, and the thrust bearing are disposed.

7. The compressor of claim 6, comprising a bearing cavity formed internal to the casing and in which the thrust bearing is disposed.

8. The compressor of claim 1, wherein the thrust bearing comprises a magnetic bearing.

9. The compressor of claim 1, wherein the thrust bearing comprises an anti-friction bearing.

10. The compressor of claim 1, wherein the thrust bearing comprises a lubricated hydrodynamic thrust bearing.

11. A heating, ventilation, air conditioning, and refrigerant (HVAC&R) system comprising a compressor, wherein the compressor comprises:

a shaft;

a motor configured to drive the shaft into rotation;

an impeller coupled to the shaft and configured to be driven into rotation by the shaft; and

a thrust bearing configured to permit rotation of the shaft and to support an axial load of the shaft, wherein the thrust bearing is positioned about the shaft and between the motor and the impeller of the compressor.

12. The HVAC&R system of claim 11, comprising an evaporator from which the compressor is configured to receive refrigerant.

13. The HVAC&R system of claim 11, comprising a condenser configured to receive refrigerant from the compressor.

14. The HVAC&R system of claim 11, wherein the compressor comprises a diffuser passage axially aligned with the impeller such that the impeller directs refrigerant toward and into the diffuser passage.

15. The HVAC&R system of claim 11, wherein the thrust bearing comprises a magnetic bearing, an anti-friction bearing, or a lubricated hydrodynamic thrust bearing.

16. A centrifugal compressor, comprising:

a shaft;

a motor configured to drive the shaft into rotation; and

a thrust bearing configured to permit rotation of the shaft and to support an axial load of the shaft, wherein the thrust bearing is configured to be disposed within a bearing cavity of the centrifugal compressor, about the shaft, and between the motor and an impeller of the centrifugal compressor.

17. The centrifugal compressor of claim 16, wherein the centrifugal compressor comprises a hermetic overhung centrifugal compressor.

18. The centrifugal compressor of claim 16, comprising a casing in which the shaft, the motor, and the thrust bearing are configured to be disposed.

19. The centrifugal compressor of claim 18, wherein the bearing cavity is formed within the casing.

20. The centrifugal compressor of claim 16, wherein the thrust bearing comprises a magnetic bearing, an anti-friction bearing, or a lubricated hydrodynamic thrust bearing.