Chiller turbo compressor assembly

Abstract

The present invention relates to a turbo-compressor assembly of a chiller, comprising a compressor I, an electric motor III and a turbo-expander II arranged in a common housing, the common housing comprising a compressor body 19, an electric motor body 21, a turbo-expander body 20 and a shaft 13 mounted in the common housing in gas bearings opening the supply of working gas from the turbo-compressor assembly, the compressor impeller 1, the turbo-expander turbine impeller 5 and the electric motor rotor 8 being fixedly mounted on the shaft 13, the compressor impeller 1 and the turbine impeller 5 being arranged at opposite ends of the common shaft and the electric motor rotor 8 being arranged between them. Between the compressor impeller 1 and the rotor 8 of the electric motor, a shaft 13 is mounted in a first radial dynamic pressure gas bearing 101; between the turbine impeller 5 and the rotor 8 of the electric motor, a shaft 13 is mounted in a second radial dynamic pressure gas bearing 102; between the second radial dynamic pressure gas bearing 102 and the turbine impeller 5, a shaft is mounted in a thrust dynamic pressure gas bearing 11; an inlet for a secondary flow S2 of working gas of the turbo-compressor assembly to the first radial bearing 101 is formed between the inner wall of the body 19 of the compressor I and the shaft 13; and an inlet for a secondary flow S3 of working gas to the second radial bearing 102 and the thrust bearing 11 is formed between the inner wall of the body 20 of the turbo-expander II and the shaft 13.

Description

[Technical Field]

[0001] The present invention relates to a chiller turbo-compressor assembly comprising a compressor, an electric motor and a turbo-expander arranged in a common housing, the common housing comprising a compressor body, an electric motor body, a turbo-expander body and a shaft mounted in the common housing in gas bearings opening the supply of working gas from the turbo-compressor assembly, the compressor impeller, the turbo-expander turbine impeller and the electric motor rotor being fixedly mounted on the shaft, the compressor impeller and the turbine impeller being arranged at opposite ends of the common shaft and the electric motor rotor being arranged therebetween. [Background technology]

[0002] Industrial chillers, with outputs ranging from several tens of kilowatts to megawatts, typically consist of a compressor, an electric motor, and an expansion turbine. Air or other cooling medium is first compressed by the compressor driven by the electric motor, then supplied to the turbine where it expands and cools. Energy transferred to the reaction turbine as the compressed medium expands is used to drive the compressor, thereby reducing the chiller's energy consumption. The cold air from the turbine is then heated in a heat exchanger or cooled in a cooling chamber and returned to the compressor inlet. This type of chiller is disclosed, for example, in Czech Patent No. 308332 (B6).

[0003] In recent years, with increasing demands for energy efficiency, ease of maintenance, and reliability in chillers, machines in which the refrigeration system comprises a compressor in a common housing, a motor, and an expansion turbine mounted on a common shaft are becoming increasingly common. The shaft of such chillers is usually housed within the housing in two radial bearings and one thrust bearing, whereby the radial bearings support the rotating shaft and damp its vibrations, while the thrust bearing fixes the shaft relative to the machine body, damping shocks of the shaft caused by changes in axial forces acting on the shaft of the rotating machine and providing a single fixation point for the shaft relative to the machine body during relative axial displacement of the shaft portion towards the housing portion caused by axial forces and differential thermal expansion of different machine parts, particularly when the machine is started and stopped.

[0004] For the above purposes, several commonly available types of bearings have been used, such as rolling bearings, plain bearings, hydrodynamic bearings, or hydrostatic gas bearings. However, the use of these types of bearings has various drawbacks. U.S. Patent No. 8,347,648 (B2) discloses a turbine for use in a refrigeration cycle, which includes a common turbine body, a single-wheel compressor mounted on a common shaft, a single-wheel expansion turbine, an electric motor, a thrust magnetic bearing, and a radial rolling-contact bearing, with the compressor wheel and turbine wheel extending from opposite ends of the common shaft. This design of a turbine for producing cold air has the drawback that rolling-contact bearings do not allow the machine to operate at energy-efficient high speeds.

[0005] The use of hydrostatic bearings for rotary machines, such as steam turbines, is not suitable for the cooled turbocompressors mentioned above. This is because hydrostatic bearings require the installation and sealing of an oil system separate from the machine components, through which the working cooling medium (usually air or other gas) flows. This makes fixing with bearings located within the machine's common housing difficult, potentially making the entire machine disproportionately expensive. Similarly, the use of hydrostatic bearings, which utilize pressurized gas or fluid as the medium, requires a separate system to supply the high-pressure medium—for example, a high-pressure hydraulic system pressurized and driven by an external energy source—which increases costs and reduces the machine's reliability.

[0006] The above drawbacks of conventional bearings have led some refrigerator manufacturers to use both thrust and radial magnetic bearings. U.S. Patent No. 7,322,207 (B2) discloses an air refrigerant cooling system comprising a single-wheel compressor, a single-wheel expansion turbine, an electric motor, an axial magnetic bearing, and two radial magnetic bearings mounted in a common housing on a common shaft, with the compressor wheel and turbine wheel extending from opposite ends of the common shaft. The stator and rotor of the machine's electric motor are cooled by cooling air circulating in a separate cooling system, separate from the cooling system's working gas. Leakage of the working medium is suctioned separately through separate vents—one outlet pipe for compressor leakage and another vent for turbine leakage—so that the leaked working medium does not pass through the electric motor rotor. The functionality of magnetic bearings depends on the supply of electricity, and a backup power source is required to prevent serious damage to the system in the event of a sudden power outage. Another significant drawback of magnetic bearings is their low stiffness, which does not ensure sufficient damping of vibrations that occur when the machine is running or stopped, and when transitioning from critical speeds. This causes undesirable vibrations of the machine at speeds that correspond to multiples of the machine's natural frequency. When magnetic bearings are used for larger and more powerful machines, they must be backed up with additional plain bearings due to the low stiffness. However, adding plain bearings to a common machine housing increases the axial length of the machine, increasing manufacturing costs. Furthermore, it extends the length of the shaft between the bearings and, in some cases, the length of the protruding end of the shaft behind the bearing. It also adversely affects the dynamics of the machine shaft.

[0007] U.S. Patent Publication No. 2019-170190(A1) discloses a rotor system including a rotating shaft, in which a first rotating impeller, a first thrust air bearing, a thrust disk, a second thrust air bearing, a first journal air bearing, a stator for supplying rotational force to the rotating shaft structure, a second journal air bearing, and a second rotating impeller heavier than the first rotating impeller are arranged axially on the shaft. The temperature of the air flowing into the second rotating impeller is higher than the temperature of the air flowing into the first rotating impeller. This arrangement of the rotor system minimizes thermal deformation of the disk, which is advantageous in applications where the working air / gas is at a high temperature, such as turbochargers and combustion gas turbines. 。

[0008] It is therefore an object of the present invention to eliminate or reduce the above-mentioned drawbacks of known technical solutions by means of a turbocompressor assembly for a chiller comprising a hydrodynamic gas bearing using the working gas of the chiller. [Prior art documents] [Patent documents]

[0009] [Patent Document 1] Czech Patent No. 308332(B6) [Patent Document 2] US Patent No. 8347648(B2) [Patent Document 3] US Patent No. 7322207(B2) [Patent Document 4] U.S. Patent Publication No. 2019-170190(A1) Summary of the Invention

[0010] The object of the present invention is achieved by a compound turbo-compressor assembly of a chiller comprising three machines, namely a compressor, an electric motor and a turbo-expander, arranged in a common housing of a turbo-compressor assembly mounted on a common shaft.

[0011] The shaft is mounted in a common housing / body of the assembly in two radial dynamic pressure gas bearings and one thrust dynamic pressure gas bearing. A supply of working gas from the turbocompressor assembly is routed into these bearings. Compared to other types of bearings, such as magnetic bearings, rolling bearings, plain bearings, dynamic pressure bearings, or hydrostatic gas bearings, the advantages of dynamic pressure gas bearings include a small installation size that reduces the axial length of the assembly, sufficient rigidity to damp unwanted vibrations of the assembly, no need for an external energy source, either electrical or pressure, high operational reliability, and low maintenance and throughput requirements for the working gas cooling system. The throughput rate of the working gas cooling system along the entire shaft of the turbocompressor assembly, including through the bearings, allows the working gas to be utilized to cool the shaft and balance the axial forces of the compressor turbine.

[0012] A first radial dynamic pressure gas bearing with an assembly shaft mounted thereon is disposed within the assembly housing between the compressor diffuser and the electric motor stator, with the shaft mounted within the first radial bearing between the compressor impeller and the electric motor rotor. A second radial dynamic pressure gas bearing is disposed within the assembly housing between the spiral chamber of the gas inlet to the turboexpander and the electric motor stator, with the shaft mounted within the second radial bearing between the turboexpander turbine impeller and the electric motor rotor. A thrust dynamic pressure gas bearing with an assembly shaft mounted thereon is disposed within the assembly housing between the spiral chamber of the gas inlet to the turboexpander and the second radial dynamic pressure gas bearing, with the bearing disk, which is part of the shaft, mounted within the thrust bearing between the turboexpander turbine impeller and the second radial bearing. A secondary flow supply of working gas to the turbocompressor assembly for the first radial bearing is formed between the inner wall of the compressor body and the shaft, and a secondary flow supply of working gas to the second radial bearing and thrust bearing is formed between the inner wall of the turboexpander body and the shaft.

[0013] Locating the radial bearings as close as possible to the compressor impeller and the turboexpander turbine impeller is advantageous in damping unwanted vibrations of the assembly's common shaft, which occur when the machine exceeds a critical speed while in operation or at rest, when the rotational frequency of the machine is a multiple of the natural frequency of the common shaft.

[0014] The turbocompressor assembly for the chiller according to the present invention operates at full capacity with a temperature difference of up to 300°C, when the gas temperature at the outlet of the turboexpander diffuser drops to minus 150°C and the compressed air temperature in the compressor diffuser reaches plus 150°C. The thrust bearing is the axially stable point of the relative position of the shaft and the assembly body. Placing the thrust bearing as close as possible to the cold side of the assembly, where the material expansion is least, and correspondingly, locating the thrust bearing disk on the shaft as close as possible to the turboexpander turbine impeller, is advantageous in terms of designing the relative positions of the individual parts of the assembly housing and the assembly shaft, as well as the clearances between them, taking into account the different thermal expansions of the individual parts of the assembly during machine operation at various temperatures and fluctuating temperature gradients. Another advantage of locating the thrust bearing on the cold side of the assembly near the turbocompressor turbine impeller is that the relatively cool air from the turboexpander, which has a higher density than the relatively warm air from another part of the turbocompressor assembly, is used as the working gas for the gas dynamic thrust bearing. The use of cooler, more dense gas improves the stiffness of the thrust dynamic gas bearing, resulting in more reliable operation.

[0015] The rotating impeller of the compressor, which draws gas in substantially axially and expels it substantially radially into the compressor diffuser, is subjected to a reaction force from the compressed gas, the axial component of which acts on the assembly's common shaft in the direction of the shaft's axis relative to the direction of working gas flow into the compressor diffusor. Such axial forces and their rapid changes can damage thrust bearings and cause machine seizure. To balance these unwanted axial forces, a piston formed on the shaft is used. The working gas pressure acts on a surface of the piston perpendicular to the shaft axis, creating an axial force that compensates for the undesired axial force caused by the gas pressure on the compressor impeller. The piston is preferably positioned between the first radial bearing and the electric motor rotor, thereby maintaining the advantageous position of the first radial bearing as close as possible to the compressor impeller. The advantage of a radial dynamic gas bearing is that the working gas exerting pressure on the piston can pass freely through the bearing, eliminating the need for other means to supply the piston. The space between the shaft and the housing, located between the piston and the rotor of the electric motor, is sealed by a shaft seal, which restricts the passage of the working gas flow further towards the rotor of the electric motor and maintains a gas pressure in the piston space high enough to apply a sufficiently large force to the piston.

[0016] In the present invention of a turbo-compressor assembly for a chiller, a so-called pseudo-labyrinth type labyrinth seal is used to seal the space behind the piston to balance the axial force, and several radially aligned teeth seal against an essentially smooth opposing cylindrical surface. It is preferred to use at least five sealing teeth against an opposing substantially smooth cylindrical surface to effectively seal, achieve a greater pressure drop, minimize associated working medium leakage, and increase the efficiency of the associated chiller.

[0017] So-called pseudo-labyrinth seals are preferably used in the present invention to facilitate assembly of pot-type bodies of assemblies when one of the sealing cylindrical surfaces is substantially smooth. In a pot-type body / housing arrangement of an assembly, the individual components of the assembly, including the shaft, are inserted into the pot-shaped assembly housing / body, which consists of a sub-body having the shape of a pot or a substantially hollow cylinder, and are connected and sealed to each other. Such turbocompressor housing / body does not have joints at the dividing surfaces parallel to the axis of the rotary machine. During operation, there are large temperature differences between the various axially arranged parts of the assembly, and the assembly is heated / cooled by the working medium supplied to / discharged from the machine substantially radially. There is also a relatively large temperature difference between the top and bottom of the machine, resulting in different temperature deformations at the top and bottom of the machine. If the joints of the body parts are parallel to the axis of the assembly, these joints will deform and lose their airtightness.

[0018] To achieve a sufficiently small radial gap between the sealing blade and the opposing, substantially cylindrical sealing surface, and to achieve sufficient gas-tightness of the aforementioned shaft seal without the need for more than ten sealing teeth, which would unnecessarily extend the axial length of the shaft and the entire machine, it is advantageous for the cylindrical sealing surface to be part of the assembly housing and also function as the carrier for the radial hydrodynamic gas bearing. The design of the bearing carrier, which is integrated with the housing portion forming the surface that is part of the shaft seal and the portion forming the surface connecting the bearing carrier and the sealing portion to the piston, reduces the number of body parts that would otherwise be manufactured separately and attached separately to the machine body. This not only saves time but, in particular, reduces assembly tolerances, resulting in a smaller sealing gap between the labyrinth seal teeth and the opposing sealing surface, which can be achieved with greater precision. In this preferred embodiment, the first radial gas bearing is mounted on a carrier that has a first inner cylindrical surface for mounting the bearing casing and a second inner cylindrical surface that is part of the shaft seal. Both of these cylindrical surfaces are substantially parallel to the axis of the shaft, and the carrier also has an annular third surface located between the two cylindrical surfaces, substantially perpendicular to them, and positioned in contact with the piston when assembled.

[0019] In the above-described preferred embodiment of the radial bearing carrier with integrated seal, the bearing carrier forms a common portion of the compressor housing and the electric motor housing, the carrier extends partially into the compressor housing, the compressor housing on which the radial hydrodynamic gas bearing is mounted and the seal extends partially into the electric motor housing, where the carrier is surrounded by the compressor body in which it is housed, and where the carrier extends into the electric motor housing, the carrier is surrounded by the inner body of the electric motor in which it is housed.

[0020] The thrust dynamic pressure gas bearing disk is fixedly mounted on the shaft between a portion of the shaft attached to the second thrust dynamic pressure gas bearing and the blade impeller of the turboexpander turbine. The thrust dynamic pressure gas bearing disk has a first substantially flat side axially oriented toward the rotor of the electric motor and a second substantially flat side axially opposite the first side of the disk within the shaft axis. The first side of the disk abuts the shaft to which the disk is fixedly connected, for example, such that the first side of the disk has a protrusion inserted into a cavity formed in the shaft and / or the disk is fixed to the shaft by a screw, and the disk and shaft are radially fixed against rotation with respect to each other by screws or pins fixed in holes in the disk and shaft. The turboexpander turbine impeller abuts and is fixedly connected to a substantially flat second side of the disk, axially opposite the side on which the impeller blades are located, so that, for example, one of the adjacent sides has a protrusion inserted into the opposite side, and / or the impeller is fixed to the disk by screws, and the impeller and disk are radially fixed against each other's rotation by screws or pins received in holes in the disk and the shaft. An axial gap in the form of a radial groove formed on the shaft surface between the disk and the turboexpander impeller is preferably formed between the second side of the disk and its associated turboexpander turbine impeller. This axial gap reduces the contact area between the disk and the relatively cool turboexpander impeller, thus partially blocking the thermal bridge through which heat is transferred from the hot rotor of the electric motor to the cooler impeller of the turboexpander via the common shaft. Another advantage of the above technical solution, in which the disks of the thrust bearing and the shaft of the assembly are not made of one piece of material but are joined together as described above, is the use of thinner shafts on the part of the radial bearing and the associated saving of material on the shaft and on the body of the turboexpander to which it is attached.To mount the shaft in the pot body without a horizontal dividing surface, a shaft having a diameter smaller than that of the shaft on which the rotor of the electric motor is mounted is used, in the portion housed in the turboexpander body and mounted in a radial hydrodynamic gas bearing arranged between the rotor of the electric motor and the impeller of the turboexpander. During assembly, the shaft passes through the internal cavity of the cylindrical body of the turboexpander or through the internal cavity of the cylindrical body of a bearing carrier housed in the turboexpander body, with the portion housed in the radial bearing, and a thrust bearing disk is mounted on the shaft. During installation, the portion of the shaft having a diameter larger than that of the portion of the shaft on which the rotor of the electric motor is mounted and mounted in the radial bearing is not inserted through the cylindrical body of the turboexpander.

[0021] The stator of the electric motor is unnecessarily heated during assembly and is advantageously cooled by a cooling medium different from the working gas of the refrigerator, and the leakage flows through the rotor of the electric motor but does not have a sufficient flow rate to sufficiently cool the entire electric motor or the temperature of the inner part of the electric motor is not low enough. In a preferred embodiment of the turbo compressor assembly of the refrigerator, the body of the electric motor has an inner body in which the stator of the electric motor is housed and an outer body in which the inner body of the electric motor is housed. The inner body and the outer body are pot-type bodies as described above, and the inner body is inserted into the outer body, which completely surrounds the inner body radially and at least partially surrounds the inner body axially. In order to efficiently flow the cooling medium near the stator of the electric motor, a common channel for the flow of the cooling medium of the electric motor is formed in the inner and outer bodies. An embodiment of the cooling channel uses a potted arrangement of the electric motor body, and the channel is preferably formed in the electric motor body such that the portion passing through the inner body is formed as at least one spiral-shaped groove on the outer surface of the inner body and is surrounded along its entire length by the adjacent outer body except for two passages into the outer body where the medium enters and exits the groove.

[0022] The turbocompressor assembly according to the invention is suitable for use in known chillers, for example chillers according to Czech patent no. 308332 (B6), and is particularly suitable for the industrial production of cold air.

[0023] The invention will be explained with reference to an exemplary embodiment of a chiller turbocompressor assembly, which is represented diagrammatically in the drawings. [Brief explanation of the drawings]

[0024] [Figure 1] 1 shows an axial section through a turbocompressor assembly showing the components essential to explaining the principles of the present invention. [Figure 2] 2 illustrates an axial cross section through the turbo-compressor assembly shown in FIG. 1 showing the flow of cooler medium through the turbo-compressor assembly. [Figure 3] 1 illustrates schematically details of one embodiment of a portion of a turboexpander. DETAILED DESCRIPTION OF THE INVENTION

[0025] A turbocompressor assembly (FIGS. 1 and 2) for generating cold air for a cooling system includes a compressor I, a turboexpander II, and an electric motor III arranged between the compressor I and the turboexpander II. The turbocompressor assembly includes a common assembly housing including a body 19 of the compressor I, a body 20 of the turboexpander II, a body 21 of the electric motor III, and a shaft 13 common to the compressor I, the turboexpander II, and the electric motor III. The radial blade impeller 1 of the compressor I, the rotor 8 of the electric motor with permanent magnets, and the radial blade impeller 5 of the turboexpander II are arranged on the common compressor shaft 13 on an axis 131, which is also the axis of the assembly. Thus, the rotor 8 of the electric motor III is arranged between the impeller 1 of the compressor I and the impeller 5 of the turboexpander II, which are attached to both ends of the shaft 13; in the illustrated embodiment, they are attached to the shaft surface. The shaft 13 is mounted in two radial dynamic pressure gas bearings 10, 101, 102 and one thrust dynamic pressure gas bearing 11 in the common housing of the turbo assembly.

[0026] In the embodiment of the invention shown in Figures 1 and 2, the motor of the assembly is a synchronous electric motor III connected to a known control unit (not shown), but in other embodiments of the invention, any other type of motor, for example, an asynchronous electric motor, can be used. The motor of the assembly includes a rotor 8 of the electric motor III mounted on the assembly's common shaft 13, a stator 9, and a body 21 including an inner body 211 and an outer body 212. The stator 9 is housed within the inner body 211, which in turn is housed within the outer body 212 such that the outer body 212 is mounted on the outer cylindrical surface of the inner body 211. Thus, the inner cylindrical body 211 is inserted into the pot of the outer cylindrical body 211, and the outer body 212 is partially surrounded axially by the inner body 211. The inner and outer bodies 211, 212 of the electric motor III include channels 12 for carrying a coolant flow S5 for cooling the stator 9 of the electric motor III. The cooling channel 12 comprises an inner cooling channel 121 formed in the inner body 211 and an outer cooling channel 122 formed in the outer body 212. In this embodiment of the present invention, the inner cooling channel 122 is designed such that the spiral-shaped groove 121 of the inner body 211 is milled on the outer cylindrical surface of the inner body 211, and the groove 121 is covered by the inner cylindrical surface of the outer body 212 when the body 21 of the electric motor III is assembled. The outer cooling channel 122 is designed in the outer body 212 to pass through the outer body 212 at two locations to connect the inner cylindrical surface of the outer body 212 to the outer cylindrical surface of the outer body 211, thus connecting the inner cooling channel 121 to the cooling medium supply / discharge pipes (not shown) of the stator of the electric motor. A flow S5 (FIG. 2), such as a gas or fluid, flows through the first outer channel 122 in the body 9 of the electric motor III and is discharged from the body 9 through the second outer channel 122.

[0027] Compressor I (Figs. 1 and 2) comprises a confuser 4 at the inlet to the compressor I, a radial blade impeller 1 mounted on a common shaft 13 of the assembly, and a body 19 of the compressor I. In the body 19, a diffuser 2 is formed, to which a spiral outlet chamber 3 is connected.

[0028] In this embodiment of the invention, the first radial bearing 101, which is a hydrodynamic gas bearing, such as a foil bearing with curved foils or another known hydrodynamic gas bearing, is mounted within a carrier 16 that is partially surrounded by the body 19 of the compressor I, partially extends within the electric motor III, and is surrounded by the inner body 211 of the electric motor III. The carrier 16 of the first radial bearing 101 has three inner surfaces. The first inner cylindrical surface 161 of the carrier 16 is substantially parallel to the cylindrical surface of the shaft 13 and thus to the axis 131 of the shaft 13, and the first radial bearing 101 is disposed thereon. The casing of the first radial bearing 101 is inserted into this portion of the carrier 16. The second inner cylindrical surface 163, just like the first inner cylindrical surface 161, is substantially parallel to the axis 131 of the shaft. The second cylindrical surface 163 has a larger inner diameter than the first cylindrical surface 161 and is part of the shaft seal 14. In the assembled state of the assembly, the first cylindrical surface 161 of the carrier 16 of the bearing 101 is disposed between the second inner cylindrical surface 163 and the impeller 1 of the compressor I. A third inner surface 162, which is substantially perpendicular to the axis 131 of the shaft, is disposed annularly between the first inner cylindrical surface 161 and the second inner cylindrical surface 163 of the carrier 16 and, in the assembled assembly, is disposed axially relative to the piston 15, which is formed on the shaft 13 and serves to balance the axial force acting on the common shaft 13 of the assembly by the impeller 1 of the compressor I.

[0029] In an exemplary embodiment of the present invention, the shaft seal 14 is a pseudo-labyrinth seal, with 7 to 9 circular sealing teeth arranged on the cylindrical surface of the shaft 13, perpendicular to the shaft axis 131. The circular sealing teeth seal against a substantially flat second inner surface 163 of the carrier 16, which they do not contact. The number of sealing teeth depends on the pressure of the working medium before entering the seal 14 and the required pressure drop across the seal 14, with the optimum number of teeth ranging from 5 to 10 depending on the characteristics of the chiller in which the turbocompressor is used.

[0030] A common suction channel 22 is arranged downstream of the shaft seal 14 in the inner body 211 of the electric motor III in the direction towards the rotor 8 of the electric motor III, by which gaseous working medium leakages, such as air passing from the compressor I through the shaft seal 14 and working medium leakages passing from the turbine of the turboexpander II through the electric motor III, are sucked out. The working medium leakages are discharged from the common suction channel 22 through a pipe 23 (FIG. 2) into a confuser 4 at the inlet to the compressor I.

[0031] In this embodiment, the teeth of the shaft seal 14 are fixedly mounted on the rotating shaft 13 of the assembly and, when the assembly is assembled, are positioned against the second inner surface 163 of the bearing carrier 16, which is a non-moving part of the assembly housing. In this embodiment, the teeth of the seal 14 are milled directly into the material of the shaft 13; however, in other embodiments, strips of sheet metal caulked into grooves can be used instead of milled teeth. In another embodiment of the seal 14, the surface 163 comprises a thermally sprayed material that is softer than the material of the carrier 16 and / or a metal felt in which the labyrinth teeth form sealing grooves during machine operation. In another embodiment of the seal 14, a labyrinth seal is used in conjunction with blades on the second inner surface 163 of the carrier 16 that seal against the substantially smooth cylindrical surface of the shaft 13. In yet another embodiment of the seal 14, other known types of suitable seals, such as brush seals or leaf seals, or combinations of these types of seals with the labyrinth seal described above, are used.

[0032] The turboexpander II comprises a blade impeller 5 of the expansion turbine of the turboexpander II mounted on a common shaft 13 of the assembly, and a body 20 of the turboexpander II in which an inlet spiral chamber 6 and an outlet diffuser 7 are formed. In the body 20, a second radial dynamic pressure gas bearing 102 and a thrust dynamic pressure gas bearing 11, for example a foil bearing with curved foils or other known suitable dynamic pressure gas bearings, are mounted.

[0033] In this embodiment of the turbocompressor assembly, the thrust bearing 11 is mounted between the second radial bearing 102 and the blade impeller 5 of the turbine of the turboexpander II, which is located on a relatively cooler portion of the shaft 13 of the turbocompressor assembly. In another embodiment, the second radial bearing 102 may be mounted between the thrust bearing 11 and the blade impeller 5. The thrust bearing 11 constitutes a stable point at which the shaft 13 is axially fixed against displacement within the housing of the assembly. The disk 111 of the thrust bearing 11 includes a disk first side and a disk second side, which together define a body of the disk 111 fixedly mounted on the shaft 13. The disk 111 mounted on the shaft 13 is oriented such that the disk first side is closer to the rotor 8 on the shaft 13 than the second disk side 111. The first and second sides of the disk 111 have a substantially circular shape. The disk 111 (FIG. 3) further comprises an inner part 111b that is substantially concentric with the axis 131 and serves mainly to secure the disk 111 to the shaft 13, and an outer part 111a that radially surrounds the inner part 111b and is attached to the thrust bearing 11. On both axially facing sides of the inner part 111b of the disk 111, protrusions 112, 113 are arranged, which are fixedly connected to the inner part 111a of the disk of the bearing 11. The disk 111 of the thrust bearing 11 is mounted on the shaft 13 so that the first side on which the protrusion 112 is arranged is recessed by the protrusion 112 in the shaft 13 and fixed by bolts 18, and is fixed by fixing means such as bolts or pins that are fixedly attached axially in the shaft 13 and in the disk 111.

[0034] The impeller 5 is mounted on the shaft 13 such that the side of the impeller 5 axially opposite the side on which the blades 51 are disposed is mounted on a second side of the disk 111 axially opposite the first side of the disk 111 at the axis 131 of the shaft 13. The impeller 5 is mounted on the protrusion 113 and attached to the disk 111 by bolts 18 and secured by fastening means, such as screws or pins, axially secured within the impeller 5 and the disk 111. The impeller 5 mounted on the disk 111 abuts the disk 111 so as to contact only the protrusion 113, thereby forming a radial groove 17 between the inner portion 111b of the disk 111 and the impeller 5. This fixed mounting arrangement of the impeller 5 on the shaft 13 reduces the contact area between the impeller 5 on the shaft 13 and the disk 111, thereby reducing heat transfer between the cold and hot sections of the turbocompressor assembly. Furthermore, the groove 17 is cooled by the leakage working medium flow S3.

[0035] The working medium flow of the chiller turbocompressor according to the invention is shown diagrammatically in FIG. 2, which shows the main working medium flow S1, the working medium leakage flows S2, S3, and S4, and the cooling medium flow S5 of the body 21 of the electric motor III during machine operation. A gas flow S1 having a pressure of 1 to 5 bar is fed into the single-wheel radial compressor I through the inlet opening of the condenser 4 and compressed in the diffuser 2 at a compression ratio of 1.6 to 2.2. The compressed working gas main flow S1 having a pressure of 1.6 to 11 bar and a temperature of up to 150 °C is further fed to a known heat exchanger (not shown) and from there to the inlet spiral chamber 6 of the turboexpander II, from where the flow S1 is fed to the impeller 5 of the single-stage radial expansion turbine, where it is expanded at an expansion ratio of 1.5 to 2.1 and simultaneously cooled to a set temperature of down to -150 °C. The cooled flow S1 of working gas is fed from the diffuser 7 to a known heat exchanger and / or cooling chamber (not shown) where it is heated and from there returned to the confuser 4 at the inlet of the compressor I. A secondary flow S2 of compressed working gas, having a pressure lower than that of the gas in the diffuser 2 and attaining a pressure value of 1.5 to 10 bar during machine operation, is fed from the compressor I through the first radial dynamic gas bearing 101 to the piston 15, the working surface of the piston 15 adjacent to face 162 being subjected to an axial force opposite to the axial component of the force exerted on the shaft 13 by the impeller 1 of the compressor I; the action of this axial force is balanced and relieved by the thrust bearing 11. Stream S2 is then fed from piston 15 through shaft seal 14 to a collecting channel 22 which extracts the working medium leakage, from where it is channeled in stream S4 together with the working gas leakage, which is fed by stream S3 from turboexpander II, passing through thrust bearing 11, radial bearing 102 and rotor 8 of electric motor III, through duct 23 into the inlet opening of confuser 4 of compressor I. Stream S3, into which the relatively cold working medium leakage from turboexpander II is channeled, cools shaft 13 and rotor 8 of electric motor II.The stator 9 of electric motor II, which is attached to the body 21, 211, is cooled by a cooling medium, such as a fluid or gas, such that the cooling medium is supplied by a flow S5 from a known source, not shown, to the outer cooling channel 122, through which it is supplied to the inner channel 121, cooling the body 21 with the stator 9, and flow S5 is supplied to the outer channel 122, from where it is supplied to a known source, not shown, of cooling medium. The cooling medium flow S5 of the stator 9 of electric motor III is not in fluid communication with any of the working medium flows S1, S2, S3, S4 of the cooler. [Explanation of symbols]

[0036] I Compressor II Turbo Expander III Electric motor 1-blade compressor impeller 2 Compressor diffuser 3 Compressor spiral chamber 4 Confuser at the compressor inlet 5 Turboexpander turbine blade impeller 51 Turboexpander turbine blades and impeller blades 6 Turboexpander spiral chamber 7 Diffuser at the outlet of the turbo expander 8 Rotor of an electric motor with permanent magnets 9 Electric motor stator 10 Radial dynamic pressure gas bearing 101 First radial dynamic pressure gas bearing 102 Second radial dynamic pressure gas bearing 11 Thrust dynamic pressure gas bearing 111 Thrust dynamic pressure gas bearing disk 111a Outer part of the disk of the thrust dynamic pressure gas bearing 111b Inner part of the disk of the thrust dynamic pressure gas bearing 112 First protrusion of the disk of the thrust dynamic pressure gas bearing 113 Second protrusion of the disk of the thrust dynamic pressure gas bearing 12. Channel for cooling medium of electric motor 121 Inner channel for cooling medium of electric motor 122 Outer channel for cooling medium of electric motor 13 Shaft 131 Shaft axis 14 Shaft seal 15 Piston for balancing axial forces 16 Radial bearing carrier 161 First inner cylindrical surface of radial bearing carrier 162 Third inner surface of radial bearing carrier 163 Second cylindrical surface of radial bearing carrier 17 Radial groove between the disk of a thrust bearing and the blade impeller of a turboexpander turbine 18 Screws for attaching the turbo expander turbine blade impeller 19 Compressor body 20 Turbo expander body 21 Electric motor body 211 Inner body of electric motor 212 outer body of electric motor 22 Common leakage suction channel for working medium 23 Common leakage extraction duct for working medium S1 Main working gas S2 Secondary flow of working gas from the compressor S3 Secondary flow of working gas from turboexpander S4 Common flow of working gas leakage from compressor and turboexpander S5 Coolant flow in the electric motor body

Claims

1. A turbo-compressor assembly for a chiller, comprising a compressor (I), an electric motor (III) and a turbo-expander (II) arranged in a common housing, the common housing comprising the body (19) of the compressor, the body (21) of the electric motor, the body (20) of the turbo-expander and a shaft (13) mounted in the common housing in a gas bearing opening the supply of working gas from the turbo-compressor assembly, the impeller (1) of the compressor, the impeller (5) of the turbine of the turbo-expander and the electric motor. In a turbo compressor assembly in which a rotor (8) of an electric motor is fixedly attached to the shaft (13), the compressor impeller (1) and the turbine impeller (5) are disposed on opposite ends of the common shaft, and the rotor (8) of the electric motor is disposed therebetween, the shaft (13) is mounted in a first radial dynamic pressure gas bearing (101) between the compressor impeller (1) and the rotor (8) of the electric motor, and the shaft (13) is mounted in a second radial dynamic pressure gas bearing (101) between the turbine impeller (5) and the rotor (8) of the electric motor. The shaft is mounted in a thrust dynamic pressure gas bearing (11) between the second radial dynamic pressure gas bearing (102) and the turbine impeller (5), and a supply portion of the secondary flow (S2) of the working gas of the turbo compressor assembly to the first radial dynamic pressure gas bearing (101) is formed between the inner wall of the body (19) of the compressor (I) and the shaft (13), and a supply portion of the secondary flow (S3) of the working gas to the second radial dynamic pressure gas bearing (102) and the thrust dynamic pressure gas bearing (11) is formed between the inner wall of the body (19) of the compressor (I) and the shaft (13). a first radial dynamic pressure gas bearing (101) formed between an inner wall of the body (20) of the turbo-expander (II) and the shaft (13), the first radial dynamic pressure gas bearing (101) being sealed towards the rotor (8) of the electric motor (III) by a shaft seal (14), and an annular surface formed on the shaft between the first radial bearing (101) and the shaft seal (14), perpendicular to the axis (131) of the shaft (13), serving as a piston (15) for balancing the axial force acting on the shaft (13).

2. 2. The turbocompressor assembly of claim 1, wherein the shaft seal (14) between the first radial dynamic pressure gas bearing (101) and the rotor (8) is a pseudo-labyrinth formed by at least five teeth sealing against a flat cylindrical surface.

3. 3. The turbocompressor assembly of claim 1, wherein the first radial dynamic pressure gas bearing is mounted on a carrier having a first inner cylindrical surface for mounting a housing of the radial bearing parallel to the axis of the shaft, a second inner cylindrical surface parallel to the axis of the shaft and part of the shaft seal, and an annular third inner surface substantially perpendicular to the axis of the shaft and disposed between the first inner cylindrical surface and the second inner cylindrical surface, wherein a diameter of the cylindrical cavity enclosed by the second inner cylindrical surface is larger than a diameter of the cylindrical cavity enclosed by the first inner cylindrical surface.

4. 4. The turbocompressor assembly according to claim 3, characterized in that the carrier (16) of the first radial dynamic pressure gas bearing (101) is made of a single material and is surrounded in a first part by the body (19) of the compressor (I) in which it is housed and in a second part by the inner body (211) of the electric motor (III) in which it is housed.

5. 5. The turbocompressor assembly according to claim 1, wherein a disk (111) of the thrust hydrodynamic gas bearing (11) is mounted on the shaft (13), the disk being oriented with a first side of the disk (11) towards the rotor (8) and a second side of the disk towards the impeller (5) of the turbine of the turboexpander (II), and radial grooves (17) are formed in the shaft (13) between the disk (11) and the impeller (5).

6. 6. The turbocompressor assembly according to claim 1, wherein the body (21) of the electric motor (III) comprises an inner body (211) that houses a stator (9) of the electric motor (III) and an outer body (21) that houses the inner body (211), and wherein the inner and outer bodies (211, 212) define a common cooling channel (12) that serves for the flow of a cooling medium for cooling the stator (9) of the electric motor.

Images