JP2019529777A - Centrifugal compressor - Google Patents

Abstract

The centrifugal compressor (22) includes a casing (30), a first compression mechanism (23a), and a second compression mechanism (23b). The casing (30) has a first inlet portion (31a), a first outlet portion (33a), a second inlet portion (31b), and a second outlet portion (33b). The first compression mechanism (23a) includes a first inlet guide vane (32a) arranged at the first inlet portion (31a) and a first impeller (34a) arranged downstream of the first inlet guide vane (32a). And a first diffuser (36a) disposed at a first outlet portion (33a) downstream from the first impeller (34a), and a first motor (38a). The second compression mechanism (23b) includes a second inlet guide vane (32b) disposed at the second inlet portion (31b) and a second impeller (34b) disposed downstream of the second inlet guide vane (32b). And a second diffuser (36b) disposed at the second outlet portion (33b) downstream from the second impeller (34b), and a second motor (38b). The first and second motors (38a, 38b) are arranged to rotate the first and second shafts (42a, 42b) to rotate the first and second impellers (34a, 34b).

Description

  The present invention relates generally to centrifugal compressors. More specifically, the present invention relates to a centrifugal compressor including a pair of compressors and a pair of motors.

  A chiller system is a refrigeration machine or apparatus that removes heat from a medium. In general, a liquid such as water is used as a medium, and the chiller system operates in a vapor compression refrigeration cycle. This liquid can then circulate through the heat exchanger and cool the air or equipment as needed. As an inevitable by-product, waste heat is generated by freezing. This waste heat must be exhausted to the surroundings or must be recovered for heating purposes to improve efficiency. Conventional chiller systems often utilize a centrifugal compressor, often referred to as a turbo compressor. For this reason, such a chiller system is sometimes called a turbo chiller. Alternatively, other types of compressors may be utilized, for example screw compressors.

  In a conventional (turbo) chiller, refrigerant is compressed in a centrifugal compressor and sent to a heat exchanger. In the heat exchanger, heat exchange occurs between the refrigerant and the heat exchange medium (liquid). Since the refrigerant condenses in this heat exchanger, this heat exchanger is called a condenser. As a result, heat is transferred to the medium (liquid) and the medium is heated. The refrigerant leaving the condenser is expanded by an expansion valve and sent to another heat exchanger. In another heat exchanger, heat exchange takes place between the refrigerant and the heat exchange medium (liquid). Since the refrigerant is heated (vaporized) in this heat exchanger, this heat exchanger is called an evaporator. As a result, heat is transferred from the medium (liquid) to the refrigerant, and the liquid is cooled. Thereafter, the refrigerant from the evaporator is returned to the centrifugal compressor and this cycle is repeated. In many cases, the liquid utilized is water.

  A conventional centrifugal compressor basically includes a casing, an inlet guide vane, an impeller, a diffuser, a motor, various sensors, and a controller. The refrigerant flows in order through the inlet guide vane, the impeller, and the diffuser. Therefore, the inlet guide vane is connected to the gas intake port of the centrifugal compressor, and the diffuser is connected to the gas outlet port of the impeller. The inlet guide vane controls the flow rate of the refrigerant gas entering the impeller. The impeller increases the speed of the refrigerant gas. The diffuser functions to convert the speed (dynamic pressure) of the refrigerant gas provided by the impeller into pressure (static pressure). The motor rotates the impeller. The controller controls the motor, the inlet guide vane, and the expansion valve. In this way, the refrigerant is compressed in the conventional centrifugal compressor. Conventional centrifugal compressors can be one-stage or two-stage. One motor drives one or more impellers.

  See, for example, U.S. Patent No. 7,942,628 and U.S. Patent Application Publication No. 2010/0251750 as examples of the prior art.

  The chiller industry began proposing variable speed compressors in the 1990s seeking efficiency improvements.

  The chiller industry has also developed a two-stage centrifugal compressor in search of higher chiller efficiency.

  In the case of a two-stage centrifugal compressor structure, one motor is used to drive both impellers. This structure has been found to be problematic in (1) operating range and (2) efficiency.

  (1) Regarding the operating range, it has been discovered that there is a relationship between the operating range and the rotational speed of each impeller. With current technology, each impeller can only be rotated at the same speed as the other impellers, so trying to operate one of the impellers out of range can make the compressor difficult and / or inoperable Has been discovered.

  (2) Regarding efficiency, it has been discovered that if any impeller stops operating at the design point, the efficiency of the compressor is reduced.

  (1) With respect to the operating range, by rotating each impeller at a different rotational speed, there is a possibility that the impeller will not operate outside the operating range (that is, maintained within the operating range). Has been discovered.

  (2) It has been discovered that the efficiency of the entire compressor can be improved by adjusting the rotational speeds of the first stage and second stage impellers so that the efficiency of the impeller is increased.

  In the operation of current technology, the operating range of each compressor is governed by the operating range of the impeller. In a two-stage compressor, each stage may have its own boundary limits (choke flow limit, stall and surge limit, minimum unload limit), so its operating range performance may be further limited. Therefore, when any impeller operates out of range, the compressor cannot or should not operate.

  With current technology operation, if any impeller does not operate at its design point, the efficiency of the compressor is reduced. This is due to changes in the head coefficient and the flow coefficient. Once these values change, the compressor cannot operate at its design (maximum efficiency) point.

  Also, if the two stages do not match well, operating away from the design point, the advantage of the efficiency of the two-stage cycle over the one-stage cycle is less or less, and when moving away from the design and peak efficiency points, It has been found to cause a more serious loss of efficiency. Thus, it has been discovered that such chillers can be away from the “full load design point” for most of the operating time.

  Therefore, one subject of this invention is providing the centrifugal compressor which can maintain operation | movement within the operating range.

  Another object of the present invention is to provide a centrifugal compressor capable of improving efficiency.

  Yet another object of the present invention is to provide a centrifugal compressor that addresses one or more other discovered problems described above or known to those skilled in the art.

  One or more of the above-mentioned problems can be basically achieved by providing a centrifugal compressor including a casing, a first compression mechanism, and a second compression mechanism. The casing has a first inlet portion, a first outlet portion, a second inlet portion, and a second outlet portion. The first compression mechanism is disposed at a first inlet guide vane disposed at the first inlet portion, a first impeller disposed downstream of the first inlet guide vane, and a first outlet portion downstream from the first impeller. A first diffuser and a first motor. The first impeller is attached to the first shaft so as to be rotatable about the first rotation axis. The first motor is arranged to rotate the first shaft in order to rotate the first impeller. The second compression mechanism is disposed at a second inlet guide vane disposed at the second inlet portion, a second impeller disposed downstream of the second inlet guide vane, and a second outlet portion downstream from the second impeller. A second diffuser and a second motor. The second impeller is attached to the second shaft so as to be rotatable about the second rotation axis. The second motor is arranged to rotate the second shaft in order to rotate the second impeller.

  These and other objects, features, aspects and advantages of the present invention will become apparent to those skilled in the art from the following detailed description disclosing preferred embodiments in conjunction with the accompanying drawings.

  Reference is made to the accompanying drawings, which form a part hereof.

The schematic diagram which shows the two-stage chiller system (equipped with an economizer) which has the centrifugal compressor concerning one Embodiment of this invention.

The perspective view of the centrifugal compressor of the chiller system shown in FIG. For illustrative purposes, a portion is removed and shown in cross section.

The perspective view of the internal parts (for example, a shaft, an impeller, a magnetic bearing, and a motor) of the centrifugal compressor shown in FIG.1 and FIG.2.

FIG. 4 is an elevation view of internal parts (for example, a shaft, an impeller, a magnetic bearing, and a motor) of the centrifugal compressor illustrated in FIG. 3.

The partial longitudinal cross-section schematic diagram of the internal parts (for example, a shaft, an impeller, a magnetic bearing, and a motor) of the centrifugal compressor shown in FIG.3 and FIG.4. Additional parts such as sensors, backup bearings and coils are schematically shown in more detail.

The flowchart which shows the normal control of the centrifugal compressor shown in FIGS.

The graph which shows the operation | movement range of a two-stage compressor (operation | movement of the whole compressor). A represents an overall operating point outside the overall operating range.

The graph which shows the operation | movement range of the 1st stage impeller of the two-stage compressor shown to FIG. 7A. A1 represents the first stage operating point outside the first stage operating range.

The graph which shows the operation | movement range of the 2nd stage impeller of the 2 stage | paragraph compressor shown to FIG. 7A. A2 represents the second stage operating point within the second stage operating range.

The graph which shows the operation | movement range of a two-stage compressor (operation | movement of the whole compressor). A represents an entire operating point outside the entire operating range (similar to FIG. 7A), and B represents an operating point within the shifted entire operating range according to the present invention.

The graph which shows the operation | movement range of the 1st stage impeller of the 2 stage | paragraph compressor shown to FIG. 8A. A1 represents the first stage operating point outside the first stage operating range (similar to FIG. 7B), and B1 represents the first operating point shifted by lowering the rotational speed of the first stage impeller according to the present invention. To express.

The graph which shows the operation | movement range of the 2nd stage impeller of the 2 stage | paragraph compressor shown to FIG. 7A. A2 represents the second stage operating point (similar to FIG. 7C) within the second stage operating range.

The graph which shows the efficiency (efficiency of the whole compressor) of a two-stage compressor. E represents the design maximum efficiency point, and D and E represent the shifted low efficiency operating point.

FIG. 9B is a graph showing the efficiency of the first stage impeller of the two-stage compressor shown in FIG. 9A. E1 represents the design maximum efficiency point of the first stage, and D1 and E1 represent the low efficiency operating point of the shifted first stage.

The graph which shows the efficiency of the 2nd stage impeller of the 2 stage | paragraph compressor shown to FIG. 9A. E2 represents the design maximum efficiency point of the second stage, and D2 and E2 represent the low efficiency operating point of the shifted second stage.

The graph which shows the efficiency (efficiency of the whole compressor) of a two-stage compressor like FIG. 9A. E represents the design maximum efficiency point, and D and E represent the shifted low efficiency operating point.

FIG. 9B is a graph showing the efficiency of the first stage impeller of the two-stage compressor shown in FIG. 9A. E1 represents the design maximum efficiency point of the first stage, and D1 and F1 represent the low efficiency operating point of the shifted first stage. The arrows represent how efficiency can increase from point D1 by decreasing the speed of the first impeller or from point F1 by increasing.

The graph which shows the efficiency of the 2nd stage impeller of the 2 stage | paragraph compressor shown to FIG. 9A. E2 represents the design maximum efficiency point of the second stage, and D2 and F2 represent the low efficiency operating point of the shifted second stage. The arrows represent how efficiency can increase from point D2 by decreasing the speed of the second impeller or from point F2 by increasing.

  Selected embodiments will be described with reference to the drawings. It will be apparent to those skilled in the art from this disclosure that the following description of the embodiments is merely illustrative and is not intended to limit the invention as defined by the appended claims and their equivalents. Will.

  First, referring to FIG. 1, a chiller system 10 having a centrifugal compressor 22 according to an embodiment of the present invention is shown. Since the centrifugal compressor 22 in FIG. 1 is a two-stage compressor, the chiller system 10 in FIG. 1 is a two-stage chiller system. The two-stage chiller system of FIG. 1 also includes an economizer. FIG. 1 only shows an example of a chiller system in which the centrifugal compressor 22 according to the present invention can be used. For example, it will be apparent to those skilled in the art from this disclosure that the economizer of chiller system 10 can be omitted. However, in the illustrated embodiment, the economizer is present for the reasons described below.

  The chiller system 10 is conventional except for the centrifugal compressor 22 and the control method of the centrifugal compressor 22. Specifically, as will be described in more detail below, in an exemplary embodiment, the centrifugal compressor 22 includes a structure and is controlled to maintain operation within its operating range and to improve efficiency. The Therefore, here, the chiller system 10 will not be discussed and / or illustrated in detail except for the content related to the centrifugal compressor 22 and the control method of the centrifugal compressor 22. However, it will be apparent to those skilled in the art that conventional parts of the chiller system 10 can be constructed in various ways without departing from the scope of the present invention. In the illustrated embodiment, the chiller system 10 is preferably a water chiller that utilizes cooling water and chiller water in a conventional manner.

  The components of the chiller system 10 will be briefly described with continued reference to FIG. The chiller system 10 basically includes a chiller controller 20, a centrifugal compressor 22, a condenser 24, an expansion valve or orifice 25, an economizer 26, an expansion valve or orifice 27, and an evaporator 28. Are connected in series to form a loop refrigeration cycle. The economizer 26 is connected between the first compression mechanism and the second compression mechanism (for example, an intermediate stage) in the refrigerant circuit. Various sensors (not shown) are arranged throughout the circuit of the chiller system 10. Since the use of such a sensor and information for controlling the chiller system 10 from such a sensor is conventional, here, except for the contents relating to the control of the centrifugal compressor 22 according to the present invention, It will not be described and / or illustrated in detail. Thus, for the sake of brevity, it will be apparent to those skilled in the art from this disclosure that the normal operation of the chiller system 10 is omitted except for the contents related to the structure and operation of the centrifugal compressor 22.

  The compressor 22 will be described in more detail with reference to FIGS. In the illustrated embodiment, the compressor 22 is a two-stage centrifugal compressor. Therefore, the compressor 22 exemplified here includes two impellers. However, the compressor 22 may include three or more impellers (not shown). The two-stage centrifugal compressor 22 of the illustrated embodiment is conventional except that the compressor 22 includes a separate motor that is used to drive the two impellers. These motors are also controlled according to the present invention.

  Therefore, since the centrifugal compressor 22 includes the first stage and the second stage in fluid connection in series, the refrigerant compressed in the first stage is subsequently compressed in the second stage. The first stage includes a first stage inlet guide vane 32a, a first impeller 34a, a first diffuser / volute 36a, a first stage compressor motor 38a, and a first stage magnetic bearing 40a. Similarly, the second stage includes a second stage inlet guide vane 32b, a second impeller 34b, a second diffuser / volute 36b, a second stage compressor motor 38b, and a second stage magnetic bearing 40b. The centrifugal compressor 22 includes various conventional sensors (only a few are shown).

  Although magnetic bearings are described herein, it will be apparent to those skilled in the art from this disclosure that other types and forms of compressor bearings may be used in the present invention. The casing 30 covers other parts of the centrifugal compressor 22. The casing 30 includes an inlet portion 31 a and an outlet portion 33 a for the first stage of the compressor 22. The casing 30 also includes an inlet portion 31 b and an outlet portion 33 b for the second stage of the compressor 22. Thus, the casing 30 includes the first inlet portion 31a, the first outlet portion 33a, the second inlet portion 31b, and the second outlet portion 33b.

Therefore, in the illustrated embodiment, the centrifugal compressor 22 includes the casing 30, the first compression mechanism (first stage) 23a, and the second compression mechanism (second stage) 23b. The first compression mechanism 23a includes a first inlet guide vane 32a disposed in the first inlet portion 31a, a first impeller 34a disposed downstream of the first inlet guide vane 32a, and a first impeller 34a downstream from the first impeller 34a. It includes a first diffuser 36a disposed at one outlet portion 33a and a first motor 38a arranged to rotate the first shaft 42a in order to rotate the first impeller 34a. The first impeller 34a is mounted on the first shaft 42a rotatably to the first around the rotation axis X _1. The second compression mechanism 23b includes a second inlet guide vane 32b disposed at the second inlet portion 31b, a second impeller 34b disposed downstream of the second inlet guide vane 32b, and a second impeller downstream from the second impeller 34b. 2 includes a second diffuser 36b disposed in the outlet portion 33b and a second motor 38b arranged to rotate the second shaft 42b in order to rotate the second impeller 34b. Second impeller 34b is attached to the second shaft 42b rotatably in the second around the rotation axis X _2.

  As described in more detail below, the chiller controller 20 receives signals from various sensors and controls the inlet guide vanes 32a and 32b, the compressor motors 38a and 38b, and the magnetic bearings 40a and 40b. The refrigerant sequentially flows through the first stage inlet guide vane 32a, the first stage impeller 34a, the second stage inlet guide vane 32b, and the second stage impeller 34b. Inlet guide vanes 32a and 32b control the flow rate of refrigerant gas entering impellers 34a and 34b, respectively, in a conventional manner. The impellers 34a and 34b increase the speed of the refrigerant gas without substantially changing the pressure. The amount of increase in the speed of the refrigerant gas is determined by the speed of the motor. Diffuser / volutes 36a and 36b increase the pressure of the refrigerant.

  The diffuser / volutes 36 a and 36 b are fixed to the casing 30 so as not to move. The compressor motors 38a and 38b rotate the impellers 34a and 34b via the first and second shafts 42a and 42b, respectively. As best understood from FIG. 1, the first diffuser 36a is connected to the second impeller 34b so that the refrigerant compressed by the first compression mechanism 23a is further compressed by the second compression mechanism 23b. The first and second magnetic bearings 40a and 40b respectively support the shafts 42a and 42b so as to be magnetically rotatable. Alternatively, the bearing system may include roller elements, hydrodynamic bearings, hydrostatic bearings, and / or magnetic bearings, or a combination thereof. In this way, the refrigerant is compressed in the centrifugal compressor 22.

  During the operation of the chiller system 10, the first stage impeller 34a and the second stage impeller 34b of the compressor 22 rotate, and the low-pressure refrigerant in the chiller system 10 is sucked by the first stage impeller 34a. The flow rate of the refrigerant is adjusted by the inlet guide vane 32a. The refrigerant sucked by the first stage impeller 34a is compressed to an intermediate pressure. This refrigerant pressure is raised by the first diffuser / volute 36a. Thereafter, the refrigerant is introduced into the second stage impeller 34b. The flow rate of the refrigerant is adjusted by the inlet guide vane 32b. The second stage impeller 34b compresses the intermediate pressure refrigerant to a high pressure. This refrigerant pressure is raised by the second diffuser / volute 36b. Thereafter, the high-pressure gas refrigerant is discharged to the chiller system 10. In the illustrated embodiment, the impellers 34a and 34b are driven by separate motors 38a or 38b so that the rotational speeds of the first and second impellers 34a and 34b can be changed independently.

  The first and second magnetic bearings 40a and 40b will be described in more detail with reference to FIGS. The first and second magnetic bearings 40a and 40b are conventional except for the contents described here. Therefore, here, the first and second magnetic bearings 40a and 40b will not be discussed and / or illustrated in detail except for the contents relating to the present invention. Rather, it will be apparent to those skilled in the art that any suitable magnetic bearing can be used without departing from the invention. The first magnetic bearing 40a preferably includes a first impeller radial magnetic bearing 44a, a first far radial magnetic bearing 46a, and a first axial (thrust) magnetic bearing 48a. Similarly, the second magnetic bearing 40b preferably includes a second impeller radial magnetic bearing 44b, a second far radial magnetic bearing 46b, and a second axial (thrust) magnetic bearing 48b.

At least one radial magnetic bearing 44a or 46a rotatably supports the first shaft 42a, and at least one radial magnetic bearing 44b or 46b rotatably supports the second shaft 42b. Thrust magnetic bearing 48a, by acting on the first thrust disc 45a, which supports the first shaft 42a in the axial direction along the first rotational axis X _1. The thrust magnetic bearing 48a includes a thrust disk 45 attached to the first shaft 42a. Similarly, the thrust magnetic bearing 48b, by acting on the second thrust disc 45b, which supports the second shaft 42b in the axial direction along the second rotation axis X _2. The thrust magnetic bearing 48b includes a thrust disk 45b attached to the second shaft 42b.

The first thrust disc 45a from the first shaft 42a, extending radially in a direction perpendicular to the first axis of rotation X _1, is fixed to the first shaft 42a. Second thrust disk 45b from the second shaft 42b, extending radially in a direction perpendicular to the second rotation axis X _2, is fixed to the second shaft 42b. First along the rotation axis X ₁ position of the first shaft 42a (axial position) is controlled by the axial position of the first thrust disk 45a. Similarly, first along the rotation axis X ₂ position of the second shaft 42b (axial position) is controlled by the axial position of the second thrust disc 45b. The first radial magnetic bearings 44a and 46a are disposed at both axial ends of the first compressor motor 38a. The second radial magnetic bearings 44b and 46b are disposed at both axial ends of the second compressor motor 38b. In the illustrated embodiment, the first and second rotation axes X ₁ and X ₂ coincide with each other. Further, in the illustrated embodiment, the first and second rotation axes X ₁ and X ₂ are parallel.

  With continued reference to FIGS. 2-5, the various sensors detect the radial and axial positions of shafts 42a and 42b relative to magnetic bearings 44a, 44b, 46a, 46b, 48a and 48b in a conventional manner to provide a chiller. A signal is sent to the controller 20. The chiller controller 20 then controls the current sent to the magnetic bearings 44a, 44b, 46a, 46b, 48a and 48b in a conventional manner to maintain the shafts 42a and 42 in the correct position. Thus, the magnetic bearing 40a is preferably a combination of active magnetic bearings 44a, 46a and 48a, which uses gap sensors 54a, 56a and 58a (FIG. 5) to monitor the position of the shaft, A signal indicating the position of the shaft is sent to the chiller controller 20. Thus, each of the magnetic bearings 44a, 46a and 48a is preferably an active magnetic bearing. Similarly, the magnetic bearing 40b is preferably a combination of active magnetic bearings 44b, 46b and 48b, which uses gap sensors 54b, 56b and 58b (FIG. 5) to monitor the position of the shaft. , A signal indicating the position of the shaft is sent to the chiller controller 20. Thus, each of the magnetic bearings 44b, 46b and 48b is preferably an active magnetic bearing.

  Therefore, the centrifugal compressor 22 includes a first magnetic bearing 40a that rotatably supports the first shaft 42a and a second magnetic bearing 40b that rotatably supports the second shaft 42b. The first shaft 42a has a first inlet end on which the first impeller 34a is mounted and a first far end. The first motor 38a is mounted on the first shaft 42a between the first impeller 34a and the first far end. The second shaft 42b has a second inlet end on which the second impeller 34b is mounted and a second far end. The second motor 38b is mounted on the second shaft 42b between the second impeller 34b and the second far end.

  As described above, the first and second magnetic bearings 40a and 40b include a combination of a radial magnetic bearing and an axial magnetic bearing. Specifically, the magnetic bearing 44a is a first impeller-side radial magnetic bearing disposed between the first impeller 34a and the first motor 38 in the axial direction, and the magnetic bearing 44b is a second radial bearing in the axial direction. This is a second impeller-side radial magnetic bearing disposed between the impeller 34b and the second motor 38b. The magnetic bearing 46a is a first far-side radial magnetic bearing disposed on the side of the first motor 38a that is far from the side on which the first impeller 34a is mounted in the axial direction. The magnetic bearing 46b is a second far-side radial magnetic bearing disposed in the axial direction on the side of the second motor 38b that is far from the side on which the second impeller 34b is mounted in the axial direction. In any case, the first magnetic bearing 40a includes at least one first radial magnetic bearing 44a or 44b and at least one first axial thrust magnetic bearing 48a, and the second magnetic bearing 40b includes at least one first magnetic bearing 40a. 2 radial magnetic bearings 44b or 46b and at least one second axial magnetic bearing 48b.

  In the illustrated embodiment, the first axial thrust magnetic bearing 48a is disposed adjacent to the first far side radial bearing 46a in the axial direction, and the second axial thrust magnetic bearing 48b in the axial direction is the second far side radial bearing. 46b is disposed adjacent to 46b. Therefore, the first axial thrust magnetic bearing 48a is disposed at the first far end of the first shaft 42a in the axial direction, and the second axial thrust magnetic bearing 48b is disposed at the second far end of the second shaft 42b in the axial direction. Is done. The first far end (of the shaft 42a) and the second far end (of the shaft 42b) are axially separated from each other, and the first axial thrust magnetic bearing 48a and the second axial thrust magnetic bearing 48b are axially separated from each other. And a gap is formed between them.

  The gap sensors 54a, 54b, 56a, 56b, 58a and 58b are schematically shown only in FIG. Similarly, as shown only in FIG. 5, in the illustrated embodiment, backup bearings (not labeled) are provided at each end of each shaft 42a, 42b. It will be apparent to those skilled in the art from this disclosure that backup bearings (unsigned) can be omitted. Similarly, it will be apparent to those skilled in the art from this disclosure that one or more gap sensors may be omitted to simplify the magnetic bearings 40a and 40b. Furthermore, it will be apparent to those skilled in the art from this disclosure that the magnetic bearing can be passively controlled by the chiller controller 20 when the gap sensor is illuminated.

  The motors 38a and 38b according to the present invention will be described in more detail with reference to FIGS. The first motor 38a includes a first stator 60a and a first rotor 62a. Similarly, the second motor 38b includes a second stator 60b and a second rotor 62b. The stator 60a is fixed to the inner surface of the casing 30, and the rotor 62a is fixed to the shaft 42a. Similarly, the stator 60b is fixed to the inner surface of the casing 30, and the rotor 62b is fixed to the shaft 42b. Stators 60a and 60b and rotors 62a and 62b are conventional. Therefore, when electric power is sent to the stator 60a, the rotor 62a is rotated at a speed corresponding to the supplied electric power. Further, when electric power is sent to the stator 60b, the rotor 62b is rotated at a speed corresponding to the supplied electric power. Since the rotor is fixed to the shaft, the shaft is also rotated, so that the impellers 34a and 34b are also rotated.

  As described above, since two separate motors 38a and 38b are provided to rotate the first and second impellers 34a and 34b, the first and second impellers 34a and 34b can be independently operated at different speeds. Can be rotated. More specifically, motors 38a and 38b preferably receive power from separate variable frequency drives (VFD) 64a and 64b, respectively. The variable frequency drives (VFD) 64a and 64b receive control signals from the chiller controller 20 and independently control the rotation speeds of the first and second impellers 34a and 34b, respectively. A control method of the variable frequency drives (VFD) 64a and 64b will be described below with reference to the control flowchart shown in FIG. 6 and the graphs shown in FIGS. 7A to 10C.

  The independent control of the motors 38a and 38b for independently controlling the rotation speeds of the first and second impellers 34a and 34b will be described in more detail with reference to FIG. As described above, the rotation speeds of the first and second motors 38a and 38b can be controlled independently. Specifically, the controller 20 is programmed to independently control the rotation speeds of the first and second motors 38a and 38b according to the flowchart of FIG. The loop shown in FIG. 6 is started and repeated by the following triggers. (1) The compressor discharge pressure changes more than 10% per minute and / or (2) The compressor suction pressure changes more than 10% per minute. However, if the customer changes the chiller setting (ie, leaving / exiting water temperature setting), this can also trigger the trigger of FIG. The first variable frequency drive (VFD) 64a is connected to the first motor 38a and the controller 20 so as to control the rotation speed of the first motor 38a in accordance with FIG. The second variable frequency drive (VFD) 64b is connected to the second motor 38b and the controller 20 so as to control the rotation speed of the second motor 38b in accordance with FIG.

  The start / repeat point in FIG. 6 is step S1, and the end / repeat point is step S14. Steps S2 to S4 are steps used to calculate the current efficiency and determine whether or not the first stage compression mechanism 23a is operating at the most efficient point (for example, see FIG. 9B). If it is determined in step S4 that the first stage compression mechanism 23a is operating at the most efficient point, the controller 20 proceeds to step S5. Otherwise, the controller 20 proceeds to step S8. In steps S8 to S10, the controller 20 adjusts the speed of the first variable frequency drive (VFD) and the inlet guide vane 32a to improve the efficiency of the first compression mechanism 23a (see, for example, FIG. 10B). After these changes, if it is determined in step S10 that the first stage compression mechanism 23a is operating at the most efficient point, the controller 20 proceeds to step S5. Otherwise, the controller 20 returns to * A and repeats the above determination and control.

  When the controller 20 proceeds to step S5, the same logic as in the previous paragraph is repeated for the second stage compression mechanism. Steps S5 to S7 are steps used to calculate the current efficiency and determine whether or not the second stage compression mechanism 23b is operating at the most efficient point (see, for example, FIG. 9C). When it is determined in step S7 that the second stage compression mechanism 23b is operating at the most efficient point, the controller 20 proceeds to step S14. Otherwise, the controller 20 proceeds to step S11. In steps S11 to S13, the controller 20 adjusts the speed of the second variable frequency drive (VFD) and the inlet guide vane 32b to improve the efficiency of the second compression mechanism 23b (see, for example, FIG. 10C). After these changes, if it is determined in step S13 that the second stage compression mechanism 23b is operating at the most efficient point, the controller 20 proceeds to step S14. Otherwise, the controller 20 returns to * B and repeats the above determination and control.

  In addition to the control shown in FIG. 6, the rotational speeds of the variable frequency drives (VFD) 64a and 64b can be controlled to keep the first and second compression mechanisms 23a and 23b within their operating ranges. However, this is not illustrated in the flowchart because the logic of FIG. 6 can be used except that “efficiency” is replaced by “operation range”. This type of control can also be understood from FIGS. 7A-8C described below.

  Details of the operation of the centrifugal compressor according to the present invention will be described. In the illustrated embodiment, since the two stages are distinct, each stage (each impeller) operates with independent speed control. As described in the previous paragraph, the operating range of each impeller can be maintained within the boundary limits of each impeller by changing the speed of each impeller separately. Also, as disclosed herein, if the impeller speed can be changed independently, it may be possible to vary the economizer flow rate over a wide range of operating conditions. In addition, each stage (each impeller) can be operated by independent speed control, that is, by changing the speed of each impeller separately, the boundary limits are adjusted to better each of the two stages. Since the matching and expansion of the operating range of the two-stage compressor can be achieved, the mass flow rate between each stage can be controlled by independent speed control, especially when considering changing the flow rate of the economizer over a wide range of operating conditions. It is possible to better balance work input.

  Impeller matching for shape-configurable products that can be sold to a wide variety of customers with a wide variety of operating conditions, since the most restrictive operating boundary limit of any stage becomes the limit of the two-stage compressor (Selecting compatible impellers) can be important. If the matching is poor, the operation range becomes useless (it works smoothly at or near the design point, but it cannot or does not work well away from the design point) , Efficiency and cost are not comparable to single stage designs). Only in the best scenario for impeller matching can the operating range be improved by applying this new concept. Significant design challenge to find the “optimal” design because the mass flow of the second stage impeller changes significantly as the sidestream from the economizer steam enters the second stage inlet over a wide range of operating conditions Occurs.

  There is a relationship between the operating range and the rotational speed of each impeller. Current technology (a normal two-stage compressor with one motor and two impellers) can only rotate each impeller at the same speed, so if any impeller operates out of range, the compressor will It becomes impossible to operate. In addition, in the current technology (a normal two-stage compressor including one motor and two impellers), if any impeller stops operating at the design point, the efficiency of the compressor decreases.

  The technique of the exemplary embodiment can improve the operating range and efficiency of the compressor because the new structure allows each compressor to rotate at different speeds. Specifically, by rotating each impeller at a different rotational speed, the impeller does not operate outside the operating range. Moreover, if the efficiency of the rotational speeds of the first stage and the second stage impellers are adjusted and their efficiency is increased, the efficiency of the entire compressor is improved.

  FIG. 7A is a graph showing the operating range of the two-stage compressor (the entire compressor operation), where A represents the entire operating point outside the entire operating range. FIG. 7B is a graph showing the operating range of the first stage impeller, and A1 represents the first stage operating point outside the first stage operating range. FIG. 7C is a graph showing the operating range of the second stage impeller, and A2 represents the second stage operating point within the second stage operating range.

  In current state of the art operation (a typical two-stage compressor with one motor and two impellers), the operating range of each compressor is governed by the operating range of the impeller. Therefore, if any impeller operates outside the range, the compressor cannot be operated. As shown in FIGS. 7A-7C, the second stage impeller can operate at A2 (FIG. 7C), but the first stage impeller cannot operate at A1 (FIG. 7B). As a result, the compressor will not operate at A (FIG. 7A).

  FIG. 8A is a graph showing the operating range of the two-stage compressor (the operation of the entire compressor), A represents the entire operating point (similar to FIG. 7A) outside the entire operating range, and B represents the present invention. This represents an operating point within the shifted overall operating range. In FIG. 8B, A1 represents the first stage operating point (similar to FIG. 7B) outside the first stage operating range, and B1 is shifted by lowering the rotational speed of the first stage impeller according to the present invention. Represents one operating point. FIG. 8C is similar to FIG. 7C, and A2 represents the second stage operating point within the second stage operating range.

  By rotating each impeller at a different speed, both impellers can be operated within range. As shown in FIG. 8B, the operating point of the first stage impeller moves from A1 to B1 by decreasing the rotational speed. As a result, the operating point of the entire compressor moves from A to B in FIG. 8A. The operating point of the second stage impeller does not change because the second stage impeller is already operating within the range.

  FIG. 9A is a graph showing the efficiency of the two-stage compressor (the efficiency of the entire compressor), where E represents the design maximum efficiency point, and D and E represent the shifted low efficiency operating points. In FIG. 9B, E1 represents the first stage design maximum efficiency point, and D1 and E1 represent the shifted first stage low efficiency operating point. Similarly, in FIG. 9C, E2 represents the second stage maximum design efficiency point, and D2 and E2 represent the shifted second stage low efficiency operating point.

  In the operation of the current technology (a normal two-stage compressor with one motor and two impellers), if any impeller does not operate at the design point, the efficiency of the compressor is reduced. This is due to changes in the head coefficient and the flow coefficient. Once these values change, the compressor cannot operate at its design (maximum efficiency) point. See FIGS. 9A-9C.

  FIG. 10A is a graph showing the efficiency of the two-stage compressor (efficiency of the whole compressor), as in FIG. 9A, where E represents the design maximum efficiency point, and D and E are the shifted low efficiency operating points. Represents. In FIG. 10B, E1 represents the first stage design maximum efficiency point, and D1 and F1 represent the shifted first stage low efficiency operating point. The arrows represent how efficiency can be increased from point D1 by decreasing the speed of the first impeller or from point F1 by increasing. FIG. 10C is the same as FIG. 10b, but for the second stage.

  By rotating each impeller at a different speed, each impeller can be operated at a point close to the design point. When the flow coefficient is low (such as D1 or D2), the impeller speed is reduced to increase efficiency. On the other hand, when the flow coefficient is high (points such as F1 and F2), the impeller speed is increased to increase efficiency. Thus, the overall compressor efficiency approaches the maximum efficiency. As the rotational speed changes, the position of the inlet guide vane (IGV) also needs to be changed. This is because the flow coefficient and the head coefficient are changed by changing the rotation speed. Specifically, when the number of revolutions per minute (RPM) of the impeller increases, it is necessary to close the inlet guide vane (IGV) to reduce the inlet flow rate. On the other hand, when the number of revolutions per minute (RPM) of the impeller decreases, it is necessary to open the inlet guide vane (IGV) and increase the inlet flow rate.

  1-6, the chiller controller 20 may include a number of control sections programmed to control conventional parts in a conventional manner. For example, a conventional magnetic bearing control section, a conventional compressor variable frequency drive, a conventional compressor motor control section, a conventional inlet guide vane control section, and a conventional expansion valve control section. These sections may be separate or combined sections.

  In the exemplary embodiment, the control section is a section of chiller controller 20 that is programmed to perform control of the parts of chiller 10 according to FIG. 6 and as described and illustrated herein. However, as long as one or more controllers are programmed to perform control of the parts of the chiller system 10 as described herein, the exact number, location and / or and / or number of control sections, portions and / or chiller controllers 20 It will be apparent to those skilled in the art from this disclosure that the structure can be changed without departing from the invention.

  Since the chiller controller 20 is conventional, disclosed herein is at least one microprocessor or CPU, input / output (I / O) interface, random access memory (RAM), and read only memory (ROM). And a storage device (temporary or permanent) that forms a computer readable medium programmed to execute one or more control programs for controlling the chiller system 10. The chiller controller 20 may selectively include an input interface for accepting input from the user such as a keypad, and a display device used to display various parameters to the user. Since the parts and programming are conventional except for the contents described here, they are not described in detail here except for the contents necessary for understanding the embodiment.

<General interpretation of terms>
In understanding the scope of the present invention, the term “comprising” and its derivatives, as used herein, identify the presence of the described feature, element, component, group, number, and / or step. Is intended to be a non-limiting term that does not exclude the presence of other undescribed features, elements, components, groups, values and / or steps. The aforementioned matters also apply to words having similar meanings such as the terms “include”, “have” and their derivatives. Also, the terms “parts”, “sections”, “parts”, “members” or “elements” may have both singular and plural meanings when used in the singular.

  The term “detect” as used herein to describe an operation or function performed by a component, section, or device includes a component, section, or device that does not require physical detection, Rather, it includes determinations, measurements, modeling, predictions, calculations, etc. to perform an operation or function.

  The term “configured” as used herein to describe a device component, section or part includes hardware and / or software constructed and / or programmed to perform a desired function.

  As used herein, terms such as “substantially”, “about” and “approximately” refer to a reasonable amount of modification of the modified word that does not significantly change the final result.

  While only specific embodiments have been selected to describe the present invention, various changes and modifications can be made herein without departing from the scope of the invention as defined in the appended claims. Will be apparent to those skilled in the art from this disclosure. For example, the size, shape, location or orientation of the various components can be varied as needed and / or desired. Components that are shown to be directly connected to or in contact with each other may have intermediate structures disposed therebetween. The function of one element can be performed by two elements and vice versa. The structure and function of one embodiment may be employed in other embodiments. Not all advantages need to be present simultaneously in a particular embodiment. All features that are unique compared to the prior art, whether alone or in combination with other features, include the structural and / or functional concepts embodied by such features, and are separate from further inventions by the applicant. Should be regarded as a description of Accordingly, the foregoing description of the embodiments of the invention is merely exemplary and is not intended to limit the invention as defined by the appended claims and their equivalents.

Claims (16)

A casing having a first inlet portion, a first outlet portion, a second inlet portion, and a second outlet portion;
A first inlet guide vane disposed in the first inlet portion;
A first impeller disposed downstream of the first inlet guide vane and attached to the first shaft so as to be rotatable about a first rotation axis;
A first diffuser disposed at the first outlet portion downstream from the first impeller, and a first motor arranged to rotate the first shaft to rotate the first impeller. A first compression mechanism;
A second inlet guide vane disposed in the second inlet portion;
A second impeller disposed downstream of the second inlet guide vane and attached to the second shaft so as to be rotatable about a second rotation axis;
A second diffuser disposed at the second outlet portion downstream from the second impeller, and a second motor arranged to rotate the second shaft to rotate the second impeller. A second compression mechanism;
A centrifugal compressor. The rotational speeds of the first motor and the second motor can be controlled independently.
The centrifugal compressor according to claim 1. A controller programmed to independently control the rotational speeds of the first motor and the second motor;
The centrifugal compressor according to claim 2. A first variable frequency drive (VFD) connected to the first motor and the controller to control the rotation speed of the first motor to be changeable;
A second variable frequency drive (VFD) connected to the second motor and the controller to control the rotation speed of the second motor to be changeable;
Further comprising
The centrifugal compressor according to claim 3. A first magnetic bearing that rotatably supports the first shaft;
A second magnetic bearing for rotatably supporting the second shaft;
Further comprising
The centrifugal compressor according to any one of claims 1 to 4. The first rotation axis and the second rotation axis coincide with each other;
The centrifugal compressor according to claim 5. The first shaft has a first inlet end on which the first impeller is mounted, and a first far end, and the first motor is between the first impeller and the first far end. Is attached to the first shaft,
The second shaft has a second inlet end on which the second impeller is mounted and a second far end, and the second motor is between the second impeller and the second far end. Is attached to the second shaft,
The first magnetic bearing includes a first impeller side radial magnetic bearing disposed between the first impeller and the first motor in the axial direction,
The second magnetic bearing includes a second impeller radial magnetic bearing disposed between the second impeller and the second motor in the axial direction.
The centrifugal compressor according to claim 5 or 6. The first magnetic bearing includes a first far-side radial magnetic bearing disposed on a side portion of the first motor that is far from a side portion on which the first impeller is mounted in the axial direction,
The second magnetic bearing includes a second far-side radial magnetic bearing disposed on a side portion of the second motor that is far from a side portion on which the second impeller is mounted in the axial direction.
The centrifugal compressor according to claim 7. The first magnetic bearing includes a first axial thrust magnetic bearing,
The second magnetic bearing includes a second axial thrust magnetic bearing,
The centrifugal compressor according to any one of claims 5 to 8. The first axial thrust magnetic bearing is disposed on a side portion of the first motor that is distant from a side portion on which the first impeller is mounted in the axial direction.
The second axial thrust magnetic bearing is disposed on a side portion of the second motor that is remote from a side portion on which the second impeller is mounted in the axial direction.
The centrifugal compressor according to claim 9. The first shaft has a first inlet end on which the first impeller is mounted and a first far end, and the first motor is between the first impeller and the first far end. Is attached to the first shaft,
The second shaft has a second inlet end on which the second impeller is mounted and a second far end, and the second motor is between the second impeller and the second far end. Is attached to the second shaft,
The first magnetic bearing includes a first far-side radial magnetic bearing disposed on a side portion of the first motor that is far from a side portion on which the first impeller is mounted in the axial direction,
The second magnetic bearing includes a second far-side radial magnetic bearing disposed on a side portion of the second motor that is far from a side portion on which the second impeller is mounted in the axial direction.
The centrifugal compressor according to claim 5 or 6. The first magnetic bearing includes at least one first radial magnetic bearing and at least one first axial thrust magnetic bearing;
The second magnetic bearing includes at least one second radial magnetic bearing and at least one second axial thrust magnetic bearing.
The centrifugal compressor according to claim 5 or 6. The first axial thrust magnetic bearing is disposed at the first far end in the axial direction;
The second axial thrust magnetic bearing is disposed at the second far end in the axial direction;
The first far end and the second far end are separated from each other, and the first axial thrust magnetic bearing and the second axial thrust magnetic bearing are separated from each other to form a gap therebetween.
The centrifugal compressor according to claim 12. The first diffuser is connected to the second impeller so that the refrigerant compressed by the first compression mechanism is further compressed by the second compression mechanism.
The centrifugal compressor according to any one of claims 1 to 13. A chiller system including the centrifugal compressor according to any one of claims 1 to 14,
An evaporator,
A condenser,
An expansion device;
Further comprising
The compressor, the evaporator, the condenser, and the expansion mechanism are connected to each other to form a refrigerant circuit;
Chiller system. The refrigerant circuit further comprises an economizer connected between the first compression mechanism and the second compression mechanism.
The chiller system according to claim 15.

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